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HANDBOOK OF AIR CONDITIONING AND REFRIGERATION
Shan K. Wang
Second Edition
McGraw-Hill
New York San Francisco Washington, D.C. Auckland Bogot?
Caracas Lisbon London Madrid Mexico City Milan
Montreal New Delhi San Juan Singapore
Sydney Tokyo Toronto
Library of Congress Cataloging-in-Publication Data
Wang, Shan K. (Shan Kuo)
CONTENTS
Preface to Second Edition xi
Preface to First Edition xiii
Chapter 1. Introduction 1.1
Chapter 2. Psychrometrics 2.1
Chapter 3. Heat and Moisture Transfer through Building Envelope 3.1
Chapter 4. Indoor and Outdoor Design Conditions 4.1
Chapter 5. Energy Management and Control Systems 5.1
Chapter 6. Load Calculations 6.1
Chapter 7. Water Systems 7.1
Chapter 8. Heating Systems, Furnaces, and Boilers 8.1
Chapter 9. Refrigerants, Refrigeration Cycles, and Refrigeration
Systems 9.1
Chapter 10. Refrigeration Systems: Components 10.1
Chapter 11. Refrigeration Systems: Reciprocating, Rotary, Scroll, and Screw 11.1 vii
Chapter 12. Heat Pumps, Heat Recovery, Gas Cooling, and Cogeneration
Systems 12.1
Chapter 13. Refrigeration Systems: Centrifugal 13.1
Chapter 14. Refrigeration Systems: Absorption 14.1
Chapter 15. Air Systems: ComponentsFans, Coils, Filters, and Humidifiers 15.1
Chapter 16. Air Systems: EquipmentAir-Handling Units and Packaged
Units 16.1
Chapter 17. Air Systems: Air Duct Design 17.1
Chapter 18. Air Systems: Space Air Diffusion 18.1
Chapter 19. Sound Control 19.1
Chapter 20. Air Systems: Basics and Constant-Volume Systems 20.1
Chapter 21. Air Systems: Variable-Air-Volume Systems 21.1
Chapter 22. Air Systems: VAV SystemsFan Combination, System Pressure,
and Smoke Control 22.1
Chapter 23. Air Systems: Minimum Ventilation and VAV System Controls 23.1
Chapter 24. Improving Indoor Air Quality 24.1
Chapter 25. Energy Management and Global Warming 25.1
Chapter 26. Air Conditioning Systems: System Classification, Selection,
and Individual Systems 26.1
viii CONTENTS
Chapter 27. Air Conditioning Systems: Evaporative Cooling Systems
and Evaporative Coolers 27.1
Chapter 28. Air Conditioning Systems: Space Conditioning Systems 28.1
Chapter 29. Air Conditioning Systems: Packaged Systems
and Desiccant-Based Systems 29.1
Chapter 30. Air Conditioning Systems: Central Systems and Clean-Room
Systems 30.1
Chapter 31. Air Conditioning Systems: Thermal Storage Systems 31.1
Chapter 32. Commissioning and Maintenance 32.1
Appendix A. Nomenclature and Abbreviations A.1
Appendix B. Psychrometric Chart, Tables of Properties, and I-P Units to
SI Units Conversion B.1
Index follows Appendix B
CONTENTS ix
CHAPTER 1
INTRODUCTION
1.1
1.1 AIR CONDITIONING 1.1
1.2 COMFORT AND PROCESSING AIR
CONDITIONING SYSTEMS 1.2
Air Conditioning Systems 1.2
Comfort Air Conditioning Systems 1.2
Process Air Conditioning Systems 1.3
1.3 CLASSIFICATION OF AIR
CONDITIONING SYSTEMS ACCORDING
TO CONSTRUCTION AND OPERATING
CHARACTERISTICS 1.3
Individual Room Air Conditioning
Systems 1.4
Evaporative-Cooling Air Conditioning
Systems 1.4
Desiccant-Based Air Conditioning
Systems 1.4
Thermal Storage Air Conditioning
Systems 1.5
Clean-Room Air Conditioning
Systems 1.5
Space Conditioning Air Conditioning
Systems 1.5
Unitary Packaged Air Conditioning
Systems 1.6
1.4 CENTRAL HYDRONIC AIR
CONDITIONING SYSTEMS 1.6
Air System 1.6
Water System 1.8
Central Plant 1.8
Control System 1.9
Air, Water, Refrigeration, and Heating
Systems 1.10
1.5 DISTRIBUTION OF SYSTEMS
USAGE 1.10
1.6 HISTORICAL DEVELOPMENT 1.11
Central Air Conditioning Systems 1.11
Unitary Packaged Systems 1.12
Refrigeration Systems 1.12
1.7 POTENTIALS AND CHALLENGES 1.13
Providing a Healthy and Comfortable
Indoor Environment 1.13
The Cleanest, Quietest, and Most
Precise and Humid Processing
Environment 1.13
Energy Use and Energy Efficiency 1.13
Environmental ProblemsCFCs and
Global Warming 1.15
Air Conditioning or HVAC&R Industry
1.15
1.8 AIR CONDITIONING PROJECT
DEVELOPMENT 1.16
Basic Steps in Development 1.16
Design-Bid and Design-Build 1.17
The GoalAn Environmentally Friendlier,
Energy-Efficient, and Cost-Effective
HVAC&R System 1.17
Major HVAC&R Problems 1.17
1.9 DESIGN FOR AIR CONDITIONING
SYSTEM 1.18
Engineering Responsibilities 1.18
Coordination between Air Conditioning
and Other Trades, Teamwork 1.19
Retrofit, Remodeling, and
Replacement 1.19
Engineers Quality Control 1.20
Design of the Control System 1.20
Field Experience 1.21
New Design Technologies 1.21
1.10 DESIGN DOCUMENTS 1.21
Drawings 1.22
Specifications 1.22
1.11 CODES AND STANDARDS 1.23
1.12 COMPUTER-AIDED DESIGN AND
DRAFTING (CADD) 1.25
Features of CADD 1.25
Computer-Aided Design 1.25
Computer-Aided Drafting (CAD) 1.26
Software Requirements 1.26
REFERENCES 1.26
1.1 AIR CONDITIONING
Air conditioning is a combined process that performs many functions simultaneously. It conditions
the air, transports it, and introduces it to the conditioned space. It provides heating and cooling from
its central plant or rooftop units. It also controls and maintains the temperature, humidity, air
movement, air cleanliness, sound level, and pressure differential in a space within predetermined
1.2 CHAPTER ONE
limits for the comfort and health of the occupants of the conditioned space or for the purpose of
product processing.
The term HVAC&R is an abbreviation of heating, ventilating, air conditioning, and refrigerating.
The combination of processes in this commonly adopted term is equivalent to the current definition
of air conditioning. Because all these individual component processes were developed prior to the
more complete concept of air conditioning, the term HVAC&R is often used by the industry.
1.2 COMFORT AND PROCESSING AIR CONDITIONING
SYSTEMS
Air Conditioning Systems
An air conditioning, or HVAC&R, system is composed of components and equipment arranged in
sequence to condition the air, to transport it to the conditioned space, and to control the indoor environmental
parameters of a specific space within required limits.
Most air conditioning systems perform the following functions:
1. Provide the cooling and heating energy required
2. Condition the supply air, that is, heat or cool, humidify or dehumidify, clean and purify, and
attenuate any objectionable noise produced by the HVAC&R equipment
3. Distribute the conditioned air, containing sufficient outdoor air, to the conditioned space
4. Control and maintain the indoor environmental parameterssuch as temperature, humidity,
cleanliness, air movement, sound level, and pressure differential between the conditioned space
and surroundingswithin predetermined limits
Parameters such as the size and the occupancy of the conditioned space, the indoor environmental
parameters to be controlled, the quality and the effectiveness of control, and the cost involved determine
the various types and arrangements of components used to provide appropriate characteristics.
Air conditioning systems can be classified according to their applications as (1) comfort air
conditioning systems and (2) process air conditioning systems.
Comfort Air Conditioning Systems
Comfort air conditioning systems provide occupants with a comfortable and healthy indoor environment
in which to carry out their activities. The various sectors of the economy using comfort air
conditioning systems are as follows:
1. The commercial sector includes office buildings, supermarkets, department stores, shopping
centers, restaurants, and others. Many high-rise office buildings, including such structures as the
World Trade Center in New York City and the Sears Tower in Chicago, use complicated air conditioning
systems to satisfy multiple-tenant requirements. In light commercial buildings, the air conditioning
system serves the conditioned space of only a single-zone or comparatively smaller area.
For shopping malls and restaurants, air conditioning is necessary to attract customers.
2. The institutional sector includes such applications as schools, colleges, universities, libraries,
museums, indoor stadiums, cinemas, theaters, concert halls, and recreation centers. For example,
one of the large indoor stadiums, the Superdome in New Orleans, Louisiana, can seat 78,000 people.
3. The residential and lodging sector consists of hotels, motels, apartment houses, and private
homes. Many systems serving the lodging industry and apartment houses are operated continuously,
on a 24-hour, 7-day-a-week schedule, since they can be occupied at any time.
4. The health care sector encompasses hospitals, nursing homes, and convalescent care facilities.
Special air filters are generally used in hospitals to remove bacteria and particulates of submicrometer
size from areas such as operating rooms, nurseries, and intensive care units. The relative humidity in a
general clinical area is often maintained at a minimum of 30 percent in winter.
5. The transportation sector includes aircraft, automobiles, railroad cars, buses, and cruising
ships. Passengers increasingly demand ease and environmental comfort, especially for longdistance
travel. Modern airplanes flying at high altitudes may require a pressure differential of
about 5 psi between the cabin and the outside atmosphere. According to the Commercial Buildings
Characteristics (1994), in 1992 in the United States, among 4,806,000 commercial buildings having
67.876 billion ft2 (6.31 billion m2) of floor area, 84.0 percent were cooled, and 91.3 percent
were heated.
Process Air Conditioning Systems
Process air conditioning systems provide needed indoor environmental control for manufacturing,
product storage, or other research and development processes. The following areas are examples of
process air conditioning systems:
1. In textile mills, natural fibers and manufactured fibers are hygroscopic. Proper control of humidity
increases the strength of the yarn and fabric during processing. For many textile manufacturing
processes, too high a value for the space relative humidity can cause problems in the spinning
process. On the other hand, a lower relative humidity may induce static electricity that is harmful
for the production processes.
2. Many electronic products require clean rooms for manufacturing such things as integrated circuits,
since their quality is adversely affected by airborne particles. Relative-humidity control is
also needed to prevent corrosion and condensation and to eliminate static electricity. Temperature
control maintains materials and instruments at stable condition and is also required for workers who
wear dust-free garments. For example, a class 100 clean room in an electronic factory requires a
temperature of 72  2F (22.2  1.1C), a relative humidity at 45  5 percent, and a count of dust
particles of 0.5-m (1.97  105 in.) diameter or larger not to exceed 100 particles/ ft3 (3531 particles
/m3).
3. Precision manufacturers always need precise temperature control during production of precision
instruments, tools, and equipment. Bausch and Lomb successfully constructed a constanttemperature
control room of 68  0.1F (20  0.56C) to produce light grating products in the
1950s.
4. Pharmaceutical products require temperature, humidity, and air cleanliness control. For instance,
liver extracts require a temperature of 75F (23.9C) and a relative humidity of 35 percent.
If the temperature exceeds 80F (26.7C), the extracts tend to deteriorate. High-efficiency air filters
must be installed for most of the areas in pharmaceutical factories to prevent contamination.
5. Modern refrigerated warehouses not only store commodities in coolers at temperatures of
27 to 32F (2.8 to 0C) and frozen foods at 10 to 20F (23 to 29C), but also provide
relative-humidity control for perishable foods between 90 and 100 percent. Refrigerated storage
is used to prevent deterioration. Temperature control can be performed by refrigeration systems
only, but the simultaneous control of both temperature and relative humidity in the space can only
be performed by process air conditioning systems.
1.3 CLASSIFICATION OF AIR CONDITIONING SYSTEMS
ACCORDING TO CONSTRUCTION AND OPERATING
CHARACTERISTICS
Air conditioning systems can also be classified according to their construction and operating
characteristics as follows.
INTRODUCTION 1.3
Individual Room Air Conditioning Systems
Individual room, or simply individual air conditioning systems employ a single, self-contained
room air conditioner, a packaged terminal, a separated indoor-outdoor split unit, or a heat pump. A
heat pump extracts heat from a heat source and rejects heat to air or water at a higher temperature
for heating. Unlike other systems, these systems normally use a totally independent unit or units in
each room. Individual air conditioning systems can be classified into two categories:
 Room air conditioner (window-mounted)
 Packaged terminal air conditioner (PTAC), installed in a sleeve through the outside wall
The major components in a factory-assembled and ready-for-use room air conditioner include
the following: An evaporator fan pressurizes and supplies the conditioned air to the space. In tubeand-
fin coil, the refrigerant evaporates, expands directly inside the tubes, and absorbs the heat energy
from the ambient air during the cooling season; it is called a direct expansion (DX) coil. When
the hot refrigerant releases heat energy to the conditioned space during the heating season, it acts as
a heat pump. An air filter removes airborne particulates. A compressor compresses the refrigerant
from a lower evaporating pressure to a higher condensing pressure. A condenser liquefies refrigerant
from hot gas to liquid and rejects heat through a coil and a condenser fan. A temperature control
system senses the space air temperature (sensor) and starts or stops the compressor to control its
cooling and heating capacity through a thermostat (refer to Chap. 26).
The difference between a room air conditioner and a room heat pump, and a packaged terminal
air conditioner and a packaged terminal heat pump, is that a four-way reversing valve is added to all
room heat pumps. Sometimes room air conditioners are separated into two split units: an outdoor
condensing unit with compressor and condenser, and an indoor air handler in order to have the air
handler in a more advantageous location and to reduce the compressor noise indoors.
Individual air conditioning systems are characterized by the use of a DX coil for a single room.
This is the simplest and most direct way of cooling the air. Most of the individual systems do not
employ connecting ductwork. Outdoor air is introduced through an opening or through a small air
damper. Individual systems are usually used only for the perimeter zone of the building.
Evaporative-Cooling Air Conditioning Systems
Evaporative-cooling air conditioning systems use the cooling effect of the evaporation of liquid
water to cool an airstream directly or indirectly. It could be a factory-assembled packaged unit or a
field-built system. When an evaporative cooler provides only a portion of the cooling effect, then it
becomes a component of a central hydronic or a packaged unit system.
An evaporative-cooling system consists of an intake chamber, filter(s), supply fan, direct-contact
or indirect-contact heat exchanger, exhaust fan, water sprays, recirculating water pump, and water
sump. Evaporative-cooling systems are characterized by low energy use compared with refrigeration
cooling. They produce cool and humid air and are widely used in southwest arid areas in the
United States (refer to Chap. 27).
Desiccant-Based Air Conditioning Systems
A desiccant-based air conditioning system is a system in which latent cooling is performed by
desiccant dehumidification and sensible cooling by evaporative cooling or refrigeration. Thus, a
considerable part of expensive vapor compression refrigeration is replaced by inexpensive evaporative
cooling. A desiccant-based air conditioning system is usually a hybrid system of dehumidification,
evaporative cooling, refrigeration, and regeneration of desiccant (refer to Chap. 29).
There are two airstreams in a desiccant-based air conditioning system: a process airstream and a
regenerative airstream. Process air can be all outdoor air or a mixture of outdoor and recirculating
1.4 CHAPTER ONE
air. Process air is also conditioned air supplied directly to the conditioned space or enclosed manufacturing
process, or to the air-handling unit (AHU), packaged unit (PU), or terminal for further
treatment. Regenerative airstream is a high-temperature airstream used to reactivate the desiccant.
A desiccant-based air conditioned system consists of the following components: rotary desiccant
dehumidifiers, heat pipe heat exchangers, direct or indirect evaporative coolers, DX coils and vapor
compression unit or water cooling coils and chillers, fans, pumps, filters, controls, ducts, and piping.
Thermal Storage Air Conditioning Systems
In a thermal storage air conditioning system or simply thermal storage system, the electricity-driven
refrigeration compressors are operated during off-peak hours. Stored chilled water or stored ice in
tanks is used to provide cooling in buildings during peak hours when high electric demand charges
and electric energy rates are in effect. A thermal storage system reduces high electric demand for
HVAC&R and partially or fully shifts the high electric energy rates from peak hours to off-peak hours.
A thermal storage air conditioning system is always a central air conditioning system using
chilled water as the cooling medium. In addition to the air, water, and refrigeration control systems,
there are chilled-water tanks or ice storage tanks, storage circulating pumps, and controls (refer to
Chap. 31).
Clean-Room Air Conditioning Systems
Clean-room or clean-space air conditioning systems serve spaces where there is a need for critical
control of particulates, temperature, relative humidity, ventilation, noise, vibration, and space pressurization.
In a clean-space air conditioning system, the quality of indoor environmental control
directly affects the quality of the products produced in the clean space.
A clean-space air conditioning system consists of a recirculating air unit and a makeup air
unitboth include dampers, prefilters, coils, fans, high-efficiency particulate air (HEPA) filters,
ductwork, piping work, pumps, refrigeration systems, and related controls except for a humidifier in
the makeup unit (refer to Chap. 30).
Space Conditioning Air Conditioning Systems
Space conditioning air conditioning systems are also called space air conditioning systems. They
have cooling, dehumidification, heating, and filtration performed predominately by fan coils, watersource
heat pumps, or other devices within or above the conditioned space, or very near it. A fan
coil consists of a small fan and a coil. A water-source heat pump usually consists of a fan, a finned
coil to condition the air, and a water coil to reject heat to a water loop during cooling, or to extract
heat from the same water loop during heating. Single or multiple fan coils are always used to serve
a single conditioned room. Usually, a small console water-source heat pump is used for each control
zone in the perimeter zone of a building, and a large water-source heat pump may serve several
rooms with ducts in the core of the building (interior zone, refer to Chap. 28).
Space air conditioning systems normally have only short supply ducts within the conditioned
space, and there are no return ducts except the large core water-source heat pumps. The pressure
drop required for the recirculation of conditioned space air is often equal to or less than 0.6 in. water
column (WC) (150 Pa). Most of the energy needed to transport return and recirculating
air is saved in a space air conditioning system, compared to a unitary packaged or a central
hydronic air conditioning system. Space air conditioning systems are usually employed with a
dedicated (separate) outdoor ventilation air system to provide outdoor air for the occupants in the
conditioned space.
Space air conditioning systems often have comparatively higher noise level and need more
periodic maintenance inside the conditioned space.
INTRODUCTION 1.5
1.6 CHAPTER ONE
Unitary Packaged Air Conditioning Systems
Unitary packaged air conditioning systems can be called, in brief, packaged air conditioning systems
or packaged systems. These systems employ either a single, self-contained packaged unit or
two split units. A single packaged unit contains fans, filters, DX coils, compressors, condensers,
and other accessories. In the split system, the indoor air handler comprises controls and the air system,
containing mainly fans, filters, and DX coils; and the outdoor condensing unit is the refrigeration
system, composed of compressors and condensers. Rooftop packaged systems are most widely
used (refer to Chap. 29).
Packaged air conditioning systems can be used to serve either a single room or multiple rooms.
A supply duct is often installed for the distribution of conditioned air, and a DX coil is used to cool
it. Other components can be added to these systems for operation of a heat pump system; i.e., a centralized
system is used to reject heat during the cooling season and to condense heat for heating
during the heating season. Sometimes perimeter baseboard heaters or unit heaters are added as a
part of a unitary packaged system to provide heating required in the perimeter zone.
Packaged air conditioning systems that employ large unitary packaged units are central systems
by nature because of the centralized air distributing ductwork or centralized heat rejection systems.
Packaged air conditioning systems are characterized by the use of integrated, factory-assembled,
and ready-to-use packaged units as the primary equipment as well as DX coils for cooling, compared
to chilled water in central hydronic air conditioning systems. Modern large rooftop packaged
units have many complicated components and controls which can perform similar functions to the
central hydronic systems in many applications.
1.4 CENTRAL HYDRONIC AIR CONDITIONING SYSTEMS
Central hydronic air conditioning systems are also called central air conditioning systems. In a central
hydronic air conditioning system, air is cooled or heated by coils filled with chilled or hot water
distributed from a central cooling or heating plant. It is mostly applied to large-area buildings with
many zones of conditioned space or to separate buildings.
Water has a far greater heat capacity than air. The following is a comparison of these two media
for carrying heat energy at 68F (20C):
The heat capacity per cubic foot (meter) of water is 3466 times greater than that of air. Transporting
heating and cooling energy from a central plant to remote air-handling units in fan rooms is
far more efficient using water than conditioned air in a large air conditioning project. However, an
additional water system lowers the evaporating temperature of the refrigerating system and makes a
small- or medium-size project more complicated and expensive.
A central hydronic system of a high-rise office building, the NBC Tower in Chicago, is illustrated
in Fig. 1.1. A central hydronic air conditioning system consists of an air system, a water
system, a central heating /cooling plant, and a control system.
Air System
An air system is sometimes called the air-handling system. The function of an air system is to
condition, to transport, to distribute the conditioned, recirculating, outdoor, and exhaust air, and to
control the indoor environment according to requirements. The major components of an air system
Air Water
Specific heat, Btu/ lb F 0.243 1.0
Density, at 68F, lb/ ft3 0.075 62.4
Heat capacity of fluid at 68F, Btu / ft3 F 0.018 62.4
1.7
FIGURE 1.1 Schematic diagram of the central hydronic air conditioning
system in NBC Tower.
1.8 CHAPTER ONE
are the air-handling units, supply/return ductwork, fan-powered boxes, space diffusion devices, and
exhaust systems.
An air-handling unit (AHU) usually consists of supply fan(s), filter(s), a cooling coil, a heating
coil, a mixing box, and other accessories. It is the primary equipment of the air system. An AHU
conditions the outdoor/ recirculating air, supplies the conditioned air to the conditioned space, and
extracts the returned air from the space through ductwork and space diffusion devices.
A fan-powered variable-air-volume (VAV) box, often abbreviated as fan-powered box, employs
a small fan with or without a heating coil. It draws the return air from the ceiling plenum, mixes it
with the conditioned air from the air-handling unit, and supplies the mixture to the conditioned
space.
Space diffusion devices include slot diffusers mounted in the suspended ceiling; their purpose is
to distribute the conditioned air evenly over the entire space according to requirements. The return
air enters the ceiling plenum through many scattered return slots.
Exhaust systems have exhaust fan(s) and ductwork to exhaust air from the lavatories, mechanical
rooms, and electrical rooms.
The NBC Tower in Chicago is a 37-story high-rise office complex constructed in the late 1980s.
It has a total air conditioned area of about 900,000 ft2 (83,600 m2). Of this, 256,840 ft2 (23,870 m2)
is used by NBC studios and other departments, and 626,670 ft2 (58,240 m2) is rental offices located
on upper floors. Special air conditioning systems are employed for NBC studios and departments at
the lower level.
For the rental office floors, four air-handling units are located on the 21st floor. Outdoor air
either is mixed with the recirculating air or enters directly into the air-handling unit as shown in
Fig. 1.2. The mixture is filtrated at the filter and is then cooled and dehumidified at the cooling coil
during cooling season. After that, the conditioned air is supplied to the typical floor through the
supply fan, the riser, and the supply duct; and to the conditioned space through the fan-powered box
and slot diffusers.
Water System
The water system includes chilled and hot water systems, chilled and hot water pumps, condenser
water system, and condenser water pumps. The purpose of the water system is (1) to transport
chilled water and hot water from the central plant to the air-handling units, fan-coil units, and fanpowered
boxes and (2) to transport the condenser water from the cooling tower, well water, or other
sources to the condenser inside the central plant.
In Figs. 1.1 and 1.2, the chilled water is cooled in three centrifugal chillers and then is distributed
to the cooling coils of various air-handling units located on the 21st floor. The temperature of
the chilled water leaving the coil increases after absorbing heat from the airstream flowing over the
coil. Chilled water is then returned to the centrifugal chillers for recooling through the chilled water
pumps.
After the condenser water has been cooled in the cooling tower, it flows back to the condenser of
the centrifugal chillers on lower level 3. The temperature of the condenser water again rises owing
to the absorption of the condensing heat from the refrigerant in the condenser. After that, the condenser
water is pumped to the cooling towers by the condenser water pumps.
Central Plant
The refrigeration system in a central plant is usually in the form of a chiller package. Chiller packages
cool the chilled water and act as a cold source in the central hydronic system. The boiler plant,
consisting of boilers and accessories, is the heat source of the heating system. Either hot water is
heated or steam is generated in the boilers.
In the NBC Tower, the refrigeration system has three centrifugal chillers located in lower level 3
of the basement. Three cooling towers are on the roof of the building. Chilled water cools from 58
to 42F (14.4 to 5.6C) in the evaporator when the refrigerant is evaporated. The refrigerant is then
compressed to the condensing pressure in the centrifugal compressor and is condensed in liquid
form in the condenser, ready for evaporation in the evaporator.
There is no boiler in the central plant of the NBC Tower. To compensate heat loss in the perimeter
zone, heat energy is provided by the warm plenum air and the electric heating coils in the fanpowered
boxes.
Control System
Modern air conditioning control systems for the air and water systems and for the central plant consist
of electronic sensors, microprocessor-operated and -controlled modules that can analyze and perform
calculations from both digital and analog input signals, i.e., in the form of a continuous variable.
Control systems using digital signals compatible with the microprocessor are called direct digital
control (DDC) systems. Outputs from the control modules often actuate dampers, valves, and relays
by means of pneumatic actuators in large buildings and by means of electric actuators for small
projects.
In the NBC Tower, the HVAC&R system is monitored and controlled by a microprocessor-based
DDC system. The DDC controllers regulate the air-handling units and the terminals. Both communicate
with the central operating station through interface modules. In case of emergency, the
INTRODUCTION 1.9
FIGURE 1.2 Schematic drawing of air system for a typical floor of offices in the NBC Tower.
1.10 CHAPTER ONE
fire protection system detects alarm conditions. The central operating station gives emergency
directions to the occupants, operates the HVAC&R system in a smoke control mode, and actuates
the sprinkler water system.
Air, Water, Refrigeration, and Heating Systems
Air, water, refrigeration, heating, and control systems are the subsystems of an air conditioning or
HVAC&R system. Air systems are often called secondary systems. Heating and refrigeration systems
are sometimes called primary systems.
Central hydronic and space conditioning air conditioning systems both have air, water, refrigeration,
heating, and control systems. The water system in a space conditioning system may be a
chilled /hot water system. It also could be a centralized water system to absorb heat from the condenser
during cooling, or provide heat for the evaporator during heating.
For a unitary packaged system, it consists of mainly air, refrigeration, and control systems. The
heating system is usually one of the components in the air system. Sometimes a separate baseboard
hot water heating system is employed in the perimeter zone.
A evaporative-cooling system always has an air system, a water system, and a control system. A
separate heating system is often employed for winter heating.
In an individual room air conditioning system, air and refrigeration systems are installed in indoor
and outdoor compartments with their own control systems. The heating system is often a component
of the supply air chamber in the room air conditioner. It can be also a centralized hot water
heating system in a PTAC system.
Air conditioning or HVAC&R systems are therefore often first described and analyzed through
their subsystems and main components: such as air, water, heating, cooling/ refrigeration, and control
systems. Air conditioning system classification, system operating characteristics, and system
selection must take into account the whole system.
Among the air, water, and refrigeration systems, the air system conditions the air, controls and
maintains the required indoor environment, and has direct contact with the occupants and the manufacturing
processes. These are the reasons why the operating characteristics of an air conditioning
system are esssentially represented by its air system.
1.5 DISTRIBUTION OF SYSTEMS USAGE
According to surveys conducted in 1995 by the Department of Energy/Energy Information
Administration (DOE/EIA) of the United States, for a total floor space of 58,772 million ft2 (5462
million m2) in commercial buildings in 1995 and for a total of 96.6 million homes in 1993 (among
these, 74.1 million homes were air conditioned), the floor space, in million square feet, and the
number of homes using various types of air conditioning systems are as follows:
Percent of Million Percent
Million ft2 floor space homes of homes
Individual room systems 12,494 22 33.1 45
Evaporative-cooling systems 2,451 4
Space conditioning systems (estimated) (8)
Unitary packaged systems (including air-source heat
pump as well as desiccant-based systems) 26,628 48 41.0 55
Central hydronic systems (including thermal storage
and clean-room systems) 13,586 24
Others 949 2
Much of the floor space may be included in more than one air conditioning system. Given the
possibility that the floor space may be counted repeatedly, the original data listed in the DOE /EIA
publication were modified. The 8 percent of the space system includes part in central hydronic systems
and part in unitary packaged systems. Among the air conditioned homes in 1993, the unitary
packaged system is the predominate air conditioning system in U.S. homes.
1.6 HISTORICAL DEVELOPMENT
The historical development of air conditioning can be summarized briefly.
Central Air Conditioning Systems
As part of a heating system using fans and coils, the first rudimentary ice system in the United
States, designed by McKin, Mead, and White, was installed in New York Citys Madison Square
Garden in 1880. The system delivered air at openings under the seats. In the 1890s, a leading consulting
engineer in New York City, Alfred R. Wolf, used ice at the outside air intake of the heating
and ventilating system in Carnegie Hall. Another central ice system in the 1890s was installed in
the Auditorium Hotel in Chicago by Buffalo Forge Company of Buffalo, New York. Early central
heating and ventilating systems used steam-engine-driven fans. The mixture of outdoor air and return
air was discharged into a chamber. In the top part of the chamber, pipe coils heat the mixture
with steam. In the bottom part is a bypass passage with damper to mix conditioned air and bypass
air according to the requirements.
Air conditioning was first systematically developed by Willis H. Carrier, who is recognized as
the father of air conditioning. In 1902, Carrier discovered the relationship between temperature and
humidity and how to control them. In 1904, he developed the air washer, a chamber installed with
several banks of water sprays for air humidification and cleaning. His method of temperature and
humidity regulation, achieved by controlling the dew point of supply air, is still used in many industrial
applications, such as lithographic printing plants and textile mills.
Perhaps the first air-conditioned office was the Larkin Administration Building, designed by
Frank L. Wright and completed in 1906. Ducts handled air that was drawn in and exhausted at roof
level. Wright specified a refrigeration plant which distributed 10C cooling water to air-cooling
coils in air-handling systems.
The U.S. Capitol was air-conditioned by 1929. Conditioned air was supplied from overhead
diffusers to maintain a temperature of 75F (23.9C) and a relative humidity of 40 percent during
summer, and 80F (26.7C) and 50 percent during winter. The volume of supply air was controlled
by a pressure regulator to prevent cold drafts in the occupied zone.
Perhaps the first fully air conditioned office building was the Milan Building in San Antonio,
Texas, which was designed by George Willis in 1928. This air conditioning system consisted of one
centralized plant to serve the lower floors and many small units to serve the top office floors.
In 1937, Carrier developed the conduit induction system for multiroom buildings, in which recirculation
of space air is induced through a heating/cooling coil by a high-velocity discharging
airstream. This system supplies only a limited amount of outdoor air for the occupants.
The variable-air-volume (VAV) systems reduce the volume flow rate of supply air at reduced loads
instead of varying the supply air temperature as in constant-volume systems. These systems were introduced
in the early 1950s and gained wide acceptance after the energy crisis of 1973 as a result of
their lower energy consumption in comparison with constant-volume systems. With many variations,
VAV systems are in common use for new high-rise office buildings in the United States today.
Because of the rapid development of space technology after the 1960s, air conditioning systems
for clean rooms were developed into sophisticated arrangements with extremely effective air
filters. Central air conditioning systems always will provide a more precisely controlled, healthy,
and safe indoor environment for high-rise buildings, large commercial complexes, and precisionmanufacturing
areas.
INTRODUCTION 1.11
Unitary Packaged Systems
The first room cooler developed by Frigidaire was installed about in 1928 or 1929, and the Atmospheric
Cabinet developed by the Carrier Engineering Company was first installed in May 1931.
The first self-contained room air conditioner was developed by General Electric in 1930. It was a
console-type unit with a hermetically sealed motor-compressor (an arrangement in which the motor
and compressor are encased together to reduce the leakage of refrigerant) and water-cooled condenser,
using sulfur dioxide as the refrigerant. Thirty of this type of room air conditioner, were built
and sold in 1931.
Early room air conditioners were rather bulky and heavy. They also required a drainage connection
for the municipal water used for condensing. During the postwar period the air-cooled model
was developed. It used outdoor air to absorb condensing heat, and the size and weight were greatly
reduced. Annual sales of room air conditioners have exceeded 100,000 units since 1950.
Self-contained unitary packages for commercial applications, initially called store coolers, were
introduced by the Airtemp Division of Chrysler Corporation in 1936. The early models had a refrigeration
capacity of 3 to 5 tons and used a water-cooled condenser. Air-cooled unitary packages
gained wide acceptance in the 1950s, and many were split systems incorporating an indoor air handler
and an outdoor condensing unit.
Packaged units have been developed since the 1950s, from indoor to rooftops, from constantvolume
to variable-air-volume, and from few to many functions. Currently, packaged units enjoy
better performance and efficiency with better control of capacity to match the space load. Computerized
direct digital control, one of the important reasons for this improvement, places unitary packaged
systems in a better position to compete with central hydronic systems.
Refrigeration Systems
In 1844, Dr. John Gorrie designed the first commercial reciprocating refrigerating machine in the
United States. The hermetically sealed motor-compressor was first developed by General Electric
Company for domestic refrigerators and sold in 1924.
Carrier invented the first open-type gear-driven factory-assembled, packaged centrifugal chiller
in 1922 in which the compressor was manufactured in Germany; and the hermetic centrifugal
chiller, with a hermetically sealed motor-compressor assembly, in 1934. The direct-driven hermetic
centrifugal chiller was introduced in 1938 by The Trane Company. Up to 1937, the capacity of centrifugal
chillers had increased to 700 tons.
During the 1930s, one of the outstanding developments in refrigeration was the discovery by
Midgely and Hene of the nontoxic, nonflammable, fluorinated hydrocarbon refrigerant family
called Freon in 1931. Refrigerant-11 and refrigerant-12, the chlorofluorocarbons (CFCs), became
widely adopted commercial products in reciprocating and centrifugal compressors. Now, new refrigerants
have been developed by chemical manufacturers such as DuPont to replace CFCs, so as
to prevent the depletion of the ozone layer.
The first aqueous-ammonia absorption refrigeration system was invented in 1815 in Europe. In
1940, Servel introduced a unit using water as refrigerant and lithium bromide as the absorbing solution.
The capacities of these units ranged from 15 to 35 tons (52 to 123 kW). Not until 1945 did
Carrier introduce the first large commercial lithium bromide absorption chillers. These units were
developed with 100 to 700 tons (352 to 2460 kW) of capacity, using low-pressure steam as the heat
source.
Positive-displacement screw compressors have been developed in the United States since the
1950s and scroll compressors since the 1970s because of their higher efficiency and smoother rotary
motion than reciprocating compressors. Now, the scroll compressors gradually replace the reciprocating
compressors in small and medium-size refrigeration systems. Another trend is the development of
more energy-efficient centrifugal and absorption chillers for energy conservation. The energy consumption
per ton of refrigeration of a new centrifugal chiller dropped from 0.80 kW/ton (4.4 COPref)
in the late 1970s to 0.50 kW/ton (7.0 COPref) in the 1990s. A series of rotary motion refrigeration
1.12 CHAPTER ONE
compressors with small, medium, to large capacity and using scroll, screw, or centrifugal compressors
will be manufactured from now on.
1.7 POTENTIALS AND CHALLENGES
Air conditioning or HVAC&R is an industry of many potentials and challenges, including the
following.
Providing a Healthy and Comfortable Indoor Environment
Nowadays, people in the United States spend most of their time indoors. A healthy and comfortable
indoor environment provided by air conditioning is a necessity for people staying indoors, no matter
how hot or cold and dry or humid the outside climate might be. According to the American
Housing Survey conducted by the U.S. Census Bureau in 1991, of 92.3 million homes, 66 million,
or 71 percent of the total, were air conditioned, and 81.9 million, or nearly 89 percent of the total,
were heated. According to the Energy Information Administration (EIA), in 1992, for a total floor
area of 67.8 billion ft2 (6.3 billion m2) of commercial buildings, 84 percent were cooled and 91 percent
were heated.
The Cleanest, Quietest, and Most Precise and Humid
Processing Environment
A class 1 clean room to manufacture integrated circuits in a semiconductor factory may be the
cleanest processing environment that is provided in the semiconductor industry in the 1990s. The
dust particle count does not exceed 1 particle/ ft3 (35 particles /m3) of a size of 0.5m and larger,
with no particle exceeding 5m. A constant-temperature room of 68F  0.1F (20  0.56C) is
always surrounded by other constant-temperature rooms or by a buffer area of lower tolerance to
maintain the fluctuation of its sensed temperature within  0.05F (0.028C) during working
hours.
A recording studio in a television broadcasting station often needs a noise criteria (NC) curve
(refer to Chap. 4) of NC 15 to 20, in which a sound level less than that of a buzzing insect can be
heard. A refrigerated warehouse that stores vegetables such as cabbage, carrots, and celery needs a
temperature of 32F (0C) and a relative humidity of 98 to 100 percent to prevent deterioration and
loss of water.
Air conditioning or HVAC&R systems will provide not only the cleanest, precisest, quietest, and
most humid environment with fluctuations of the controlled variable within required limits, but also
at optimum energy use and first cost.
Energy Use and Energy Efficiency
Based on the data published in the Annual Energy Review in 1993, the total energy use in 1992 in
the United States was 82.14 quad Btu or 1015 Btu (86.66 EJ, or 1018 J). The United States alone
consumed about one-fourth of the worlds total production. Of the total energy use of 82.14 quad
Btu in the United States, the residential /commercial sector consumed about 36 percent of the total,
the industrial sector consumed another 36 percent, and transportation consumed the rest28 percent.
Petroleum, natural gas, and coal were the three main sources, providing more than 85 percent
of the energy supply in 1992 in the United States.
According to DOE/EIA energy consumption survey 0321 for residential buildings (1993) and
survey 0318 for commercial buildings (1995), the average annual energy use of HVAC&R systems
INTRODUCTION 1.13
1.14 CHAPTER ONE
in the residential /commercial sector was about 45 percent of the total building energy consumption.
Also assuming that the annual energy use of HVAC&R systems was about 1 percent of the total in
both industrial and transportation sectors, then the estimate of annual energy use of the HVAC&R
systems in 1992 in the United States was about 17 percent of the total national energy use, or onesixth
of the total national energy use.
The world energy resources of petroleum, natural gas, and coal are limited. The population of
the United States in 1992 was only about one-twentieth of the worlds total population; however,
we consumed nearly one-fifth of the worlds total energy produced. Energy use must be reduced.
After the energy crisis in 1973, the U.S. Congress enacted the Energy Policy and Conservation
Act of 1975 and the National Energy Policy Act of 1992. The enactment of energy efficiency legislation
by federal and state governments and the establishment of the Department of Energy (DOE)
in 1977 had a definite impact on the implementation of energy efficiency in United States.
In 1975, the American Society of Heating, Refrigerating and Air Conditioning Engineers
(ASHRAE) published Standard 90-75, Energy Conservation in New Building Design. This standard
was revised and cosponsored by the Illuminating Engineering Society of North America as
ASHRAE/IES Standard 90.1-1989, Energy Efficient Design of New Buildings Except New Low-
Rise Residential Buildings, in 1989; and it was revised again as current ASHRAE/IESNA Standard
90.1-1999, Energy Standard for Buildings Except Low-Rise Residential Buildings, in 1999. Many
other organizations also offered valuable contributions for energy conservation. All these events
started a new era in which energy efficiency has become one of the important goals of HVAC&R
system design and operation.
Because of all these efforts, the increase in net annual energy use for residential and commercial
buildings, i.e., the total energy use minus the electrical system energy loss in the power plant, from
1972 to 1992 in the United States was only 1.3 percent, as shown in Fig. 1.3. However, during the
same period, the increase in floor area of commercial buildings in the United States was about 83
Commercial building
floor space
Energy
use
U. S. commercial buildings, billion ft2
Residential and commercial buildings energy use, 1015 Btu 0
10
20
30
40
50
60
70
80
15.89
1950 1960 1970 1980 1990
16.09
0
5
10
15
FIGURE 1.3 Residential and commercial building energy use and commercial building floor space increase between
1992 and 1950 in the United States.
percent. Energy efficiency will be a challenge to everyone in the HVAC&R industry in this generation
as well as in many, many generations to come.
Environmental ProblemsCFCs and Global Warming
The surface of the earth is surrounded by a layer of air, called the atmosphere. The lower atmosphere
is called the homosphere, and the upper atmosphere is called the stratosphere.
In the mid-1980s, chlorofluorocarbons (CFCs) were widely used as refrigerants in mechanical
refrigeration systems, to produce thermal insulation foam, and to produce aerosol propellants for
many household consumer products. CFC-11 (CCl3F) and CFC-12 (CCl2F2) are commonly used
CFCs. They are very stable. Halons are also halogenated hydrocarbons. If CFCs and halons leak or
are discharged from a refrigeration system during operation or repair to the lower atmosphere, they
will migrate to the upper stratosphere and decompose under the action of ultraviolet rays throughout
their decades or centuries of life. The free chlorine atoms will react with oxygen atoms of the
ozone layer in the upper stratosphere and will cause a depletion of this layer. The theory of the depletion
of the ozone layer was proposed in the early and middle-1970s. The ozone layer filters out
harmful ultraviolet rays, which may cause skin cancer and are a serious threat to human beings.
Furthermore, changes in the ozone layer may significantly influence weather patterns. Since 1996,
actions have been taken to ban the production of CFCs and halons, before it is too late.
A cloudless homosphere is mainly transparent to short-wave solar radiation but is quite opaque
to long-wave infrared rays emitted from the surface of the earth. Carbon dioxide (CO2) has the
greatest blocking effect of all; water vapor and synthetic CFCs also play important roles in blocking
the direct escape of infrared energy. The phenomenon of transparency to incoming solar radiation
and blanketing of outgoing infrared rays is called the greenhouse effect. The increase of the
CO2, water vapor, CFCs, and other gases, often called greenhouse gases (GHGs), eventually will
result in a rise in air temperature near the earths surface. This is known as the global warming
effect.
Over the past 100 years, global warming has caused an increase of 1F. For the same period,
there was a 25 percent increase of CO2. During the 1980s, the release of CO2 to the atmosphere
was responsible for about 50 percent of the increase in global warming that was attributable to
human activity, and the release of CFCs had a 20 percent share. Some scientists have predicted an
accelerated global warming in the coming 50 years because of the increase in the worlds annual
energy use. Further increases in global temperatures may lead to lower rainfalls, drop in soil moisture,
more extensive forest fires, more flood, etc.
CFC production in developed countries has been banned since January 1, 1996. Carbon dioxide
is the product of many combustion processes. Alternative refrigerants to replace CFCs must also
have a low global warming potential. Designers and operators of the HVAC&R systems can reduce
the production of CO2 through energy efficiency and the replacement of coal, petroleum, and natural
gas power by hydroelectric, solar, and nuclear energy, etc. More studies and research are
needed to clarify the theory and actual effect of global warming.
Air Conditioning or HVAC&R Industry
The air conditioning or HVAC&R industry in the United States is an expanding and progressing industry.
In 1995, the installed value of nonresidential air conditioning hit $20 billion. According to
American Refrigeration Institute (ARI) and Heating/Piping/Air Conditioning data, from 1985 to
1995, the annual rate of increase of installed value of air conditioning systems is 8.7 percent.
Because of the replacement of the old chillers using CFCs, in 1994, the installed value of retrofit,
remodeling, and replacement accounted for up to two-thirds of all HVAC&R expenditures. This
trend may continue in the beginning of the new century.
Based on data from ARI, in 1995, shipment of packaged units reached a record figure of 5
million products, of which heat pumps comprised a one-fifth share. Centrifugal and screw chiller
INTRODUCTION 1.15
1.16 CHAPTER ONE
shipments were 7500 units, absorption chillers 502 units, and reciprocating chillers, at a
significantly smaller capacity, were 14,000 units.
1.8 AIR CONDITIONING PROJECT DEVELOPMENT
Basic Steps in Development
The basic steps in the development and use of a large air conditioning system are the design, construction,
commissioning, operation, energy efficiency upgrading, and maintenance. Figure 1.4 is a
diagram which outlines the relationship between these steps and the parties involved. The owner
sets the criteria and the requirements. Design professionals in mechanical engineering consulting
FIGURE 1.4 Steps in the development and use of air conditioning systems in buildings.
firms design the air conditioning system and prepare the design documents. Manufacturers supply
the equipment, instruments, and materials. Contractors install and construct the air conditioning
system. After construction, the air conditioning system is commissioned by a team, and then it is
handed over to the operation and maintenance group of the property management for daily operation.
Following a certain period of operation, an energy service company (ESCO) may often be
required to upgrade the energy efficiency of the HVAC&R system (energy retrofit).
Design-Bid and Design-Build
There are two types of project development: design-bid and design-build. A design-bid project
separates the design and installation responsibilities, whereas in a design-build project, engineering
is done by the installing contractor. Some reasons for a design-build are that the project is too small
to retain an engineering consultant, or that there is insufficient time to go through the normal
design-bid procedures.
According to Bengard (1999), the main advantages of design-build include established firm
price early, single-source responsibility, accelerated project delivery, and performance guarantees.
The market has experienced nearly 300 percent domestic growth since 1986.
The GoalAn Environmentally Friendlier, Energy-Efficient, and
Cost-Effective HVAC&R System
The goal is to provide an HVAC&R system which is environmentally friendlier, energy-efficient,
and cost-effective as follows:
 Effectively control indoor environmental parameters, usually to keep temperature and humidity
within required limits.
 Provide an adequate amount of outdoor ventilation air and an acceptable indoor air quality.
 Use energy-efficient equipment and HVAC&R systems.
 Minimize ozone depletion and the global warming effect.
 Select cost-effective components and systems.
 Ensure proper maintenance, easy after-hour access, and necessary fire protection and smoke control
systems.
Major HVAC&R Problems
In Coads paper (1985), a study by Wagner-Hohn-Ingles Inc. revealed that many large new buildings
constructed in the early 1980s in the United States suffer from complaints and major defects.
High on the list of these problems is the HVAC&R system. The major problems are these:
1. Poor indoor air quality (IAQ)sick building syndrome. Poor indoor air quality causes the
sick building syndrome. The National Institute of Occupational Safety and Health (NIOSH) in 1990
reported that between 1971 and 1988, 529 field investigations found that lack of outdoor air, improper
use, and poor operation and maintenance of HVAC&R systems were responsible for more
than one-half of sick building syndrome incidents. Field investigations found that 20 to 30 percent
of the buildings had poor-air-quality problems. Sick building syndrome is covered in detail in
Chap. 4.
2. Updated technology. In recent years, there has been a rapid change in the technology of air
conditioning. Various types of VAV systems, air and water economizer, heat recovery, thermal storage,
desiccant dehumidification, variable-speed drives, and DDC devices have become more effective
and more advanced for energy efficiency. Many HVAC&R designers and operators are not
INTRODUCTION 1.17
1.18 CHAPTER ONE
properly equipped to apply and use these systems. Unfortunately, these sophisticated systems are
managed, constructed, and operated under the same budget and schedule constraints as the less
sophisticated systems. After years of operation, most HVAC&R equipment and controls need to be
upgraded for energy efficiency.
3. Insufficient communication between design professionals, construction groups, and operators.
Effective operation requires a knowledgeable operator to make adjustments if necessary. The operator
will operate the system at his or her level of understanding. If adequate operating and maintenance
documents are not provided by the designer and the contractor for the operator, the
HVAC&R system may not be operated according to the designers intentions.
4. Overlooked commissioning. Commissioning means testing and balancing all systems, functional
testing and adjusting of components and the integrated system, and adjusting and tuning the
direct digital controls. An air conditioning system is different from the manufacturing products having
models and prototypes. All the defects and errors of the prototypes can be checked and corrected
during their individual tests, but the more complicated HVAC&R system, as constructed and
installed, is the end product. Therefore, proper commissioning, which permits the system to perform
as specified in the design documents, is extremely important. Unfortunately, the specifications
seldom clearly designate competent technicians for the responsibility of commissioning the entire
integrated system. Commissioning is covered in detail in Chap. 32.
5. Reluctant to try innovative approaches. In addition to Coads paper, a survey was conducted
by the American Consulting Engineers Council (ACEC) in 1995. For the 985 engineering firms that
responded out of about 5500 firms total, there were 522 legal claims filed against them, 49 percent
of suits resolved without payment. Among the 522 legal claims, 9 percent were filed against mechanical
engineering firms. Fifty-two percent of 1995 legal claims were filed by owners, and
13 percent were filed by contractors or subcontractors.
Because of the legal claims, insurance companies discourage innovation, and engineering
consulting firms are reluctant to try innovative techniques. Twenty-one percent of these firms said
they were very much reluctant, 61 percent said they were somewhat affected, and 25 percent said
they were a little affected. Support and encouragement must be given to engineering firms for carefully
developed innovative approaches to projects. In addition to the ASHRAE Technology Awards,
more HVAC&R innovative awards should be established in large cities in the United States to
promote innovative approaches.
1.9 DESIGN FOR AIR CONDITIONING SYSTEM
System design determines the basic characteristics. After an air conditioning system is constructed
according to the design, it is difficult and expensive to change the design concept.
Engineering Responsibilities
The normal procedure in a design-bid project includes the following steps and requirements:
1. Initiation of a construction project by owner or developer
2. Selection of design team
3. Setting of the design criteria and indoor environmental parameters
4. Selection of conceptual alternatives for systems and subsystems; preparation of schematic layouts
of HVAC&R
5. Preparation of contract documents, working drawings, specifications, materials and construction
methods, commissioning guidelines
6. Competitive bidding by contractors
7. Evaluation of bids; negotiations and modification of contract documents
8. Advice on awarding of contract
9. Review of shop drawings and commissioning schedule, operating and maintenance manuals
10. Monitoring, supervision, and inspection of construction
11. Supervision of commissioning: testing and balancing; functional performance tests
12. Modification of drawings to the as-built condition and the finalization of the operation and
maintenance manual
13. Acceptance
Construction work starts at contract award following the bidding and negotiating and ends at the
acceptance of the project after commissioning.
It is necessary for the designer to select among the available alternatives for optimum comfort,
economics, energy conservation, noise, safety, flexibility, reliability, convenience, and maintainability.
Experience, education, and judgment all enter into the selection process. If both a complicated
system and a simple system yield the same performance, the simple system is preferred for its reliability,
operator convenience, and lower cost.
Coordination between Air Conditioning and Other Trades, Teamwork
Air conditioning, plumbing, and fire protection systems are mechanical systems in a building. Both
mechanical and electrical systems provide building services for the occupants and goods inside the
building. Coordination between the mechanical engineer for HVAC&R and the architect, as well as
between mechanical and structural or electrical engineers, or teamwork, becomes important.
Factors requiring input from both the architect and the mechanical engineer include the following:
1. Shape and the orientation of the building
2. Thermal characteristics of the building envelope, especially the type and size of the windows
and the construction of external walls and roofs
3. Location of the ductwork and piping to avoid interference with each other, or with other trades
4. Layout of the diffusers and supply and return grilles
5. Minimum clearance provided between the structural members and the suspended ceiling for the
installation of ductwork and piping
6. Location and size of the rooms for central plant, fan rooms, duct and pipe shafts
If the architect makes a decision that is thermally unsound, the HVAC&R engineer must offset
the additional loads by increasing the HVAC&R system capacity. Lack of such coordination results
in greater energy consumption.
HVAC&R and other building services must coordinate the following:
1. Utilization of daylight and the type of artificial light to be installed
2. The layout of diffusers, grilles, return inlets, and light troffers in the suspended ceiling
3. Integration with fire alarm and smoke control systems
4. Electric power and plumbing requirements for the HVAC&R equipment and lighting for equipment
rooms
5. Coordination of the layout of the ductwork, piping, electric cables, etc.
Retrofit, Remodeling, and Replacement
Retrofit of an HVAC&R system must be tailored to the existing building and integrated with existing
systems. Each retrofit project has a motivation which may be related to environment, safety and
health, indoor environment, energy conservation, change of use, etc. Because of ozone depletion,
INTRODUCTION 1.19
the production of CFCs ceased in the United States on January 1, 1996. All refrigeration systems
using CFC refrigerants will be replaced by or converted to non-CFC refrigerant systems. For the
sake of safety, smoke control systems and stairwell pressurization are items to be considered. For
the sake of the occupants health, an increase in the amount of outdoor air may be the right choice.
If energy conservation is the first priority, the following items are among the many that may be considered:
efficiency of lighting; energy consumption of the fans, pumps, chillers, and boilers; insulation
of the building envelope; and the remodeling or replacement of fenestration. For improving the
thermal comfort of the occupant, an increase of the supply volume flow rate, an increase of cooling
and heating capacity, and the installation of appropriate controls may be considered.
Engineers Quality Control
In 1988, the American Consulting Engineers Council chose People, Professionalism, Profits: A
Focus on Quality as the theme of its annual convention in New York City. Quality does not mean
perfection. In HVAC&R system design, quality means functionally effective, for health, comfort,
safety, energy conservation, and cost. Quality also implies the best job, using accepted professional
practices and talent. Better quality always means fewer complaints and less litigation. Of course, it
also requires additional time and higher design cost. This fact must be recognized by the owner and
developer.
The use of safety factors allows for the unpredictable in design, installation, and operation. For
example, a safety factor of 1.1 for the calculated cooling load and of 1.1 to 1.15 for the calculated
total pressure drop in ductwork is often used to take into account unexpected inferiority in fabrication
and installation. An HVAC&R system should not be overdesigned by using a greater safety factor
than is actually required. The initial cost of an overdesigned HVAC&R system is always higher,
and it is energy-inefficient.
When an HVAC&R system design is under pressure to reduce the initial cost, some avenues to
be considered are as follows:
1. Select an optimum safety factor.
2. Minimize redundancy, such as standby units.
3. Conduct a detailed economic analysis for the selection of a better alternative.
4. Calculate the space load, the capacity of the system, and the equipment requirements more
precisely.
5. Adopt optimum diversity factors based on actual experience data observed from similar buildings.
A diversity factor is defined as the ratio of the simultaneous maximum load of a system to the
sum of the individual maximum loads of the subdivisions of a system. It is also called the simultaneous-
use factor. For example, in load calculations for a coil, block load is used rather than the sum
of zone maxima for sizing coils for AHUs installed with modulating control for the chilled water
flow rate.
The need for a higher-quality design requires that engineers have a better understanding of the
basic principles, practical aspects, and updated technology of HVAC&R systems in order to avoid
overdesign or underdesign and to produce a satisfactory product.
Design of the Control System
Controlling and maintaining the indoor environmental parameters within predetermined limits
depends mainly on adequate equipment capacity and the quality of the control system. Energy can
be saved when the systems are operated at part load with the equipments capacity following the
system load accurately by means of capacity control.
1.20 CHAPTER ONE
Because of the recent rapid change of HVAC&R controls from conventional systems to energy
management systems, to DDC with microprocessor intelligence, and then to open-protocol BACnet,
many designers have not kept pace. In 1982, Haines did a survey on HVAC&R control system
design and found that many designers preferred to prepare a conceptual design and a sequence of
operations and then to ask the representative of the control manufacturer to design the control system.
Only one-third of the designers designed the control system themselves and asked the representative
of the control manufacturer to comment on it.
HVAC&R system control is a decisive factor in system performance. Many of the troubles with
HVAC&R arise from inadequate controls and/or their improper use. The designer should keep pace
with the development of new control technology. He or she should be able to prepare the sequence
of operations and select the best-fit control sequences for the controllers from a variety of the manufacturers
that offer equipment in the HVAC&R field. The designer may not be a specialist in the
details of construction or of wiring diagrams of controllers or DDC modules, but he or she should
be quite clear about the function and sequence of the desired operation, as well as the criteria for
the sensors, controllers, DDC modules, and controlled devices.
If the HVAC&R system designer does not perform these duties personally, preparation of a systems
operation and maintenance manual with clear instructions would be difficult. It would also be
difficult for the operator to understand the designers intention and to operate the HVAC&R system
satisfactorily.
Field Experience
It is helpful for the designer to visit similar projects that have been operated for more than 2 years
and talk with the operator before initiating the design process. Such practice has many advantages.
1. The designer can investigate the actual performance and effectiveness of the air conditioning
and control systems that he or she intends to design.
2. According to the actual operating records, the designer can judge whether the system is overdesigned,
underdesigned, or the exact right choice.
3. Any complaints or problems that can be corrected may be identified.
4. The results of energy conservation measures can be evaluated from actual performance instead
of theoretical calculations.
5. The designer can accumulate valuable practical experience from the visit, even from the defi-
ciencies.
New Design Technologies
Computer-aided design and drafting (CADD) and knowledge-based system (KBS) or expert system
assisted design are prospective new design technologies that have been used in HVAC&R design.
CADD is covered in detail in Sec. 1.12, and KBS is covered in Chap. 5.
1.10 DESIGN DOCUMENTS
In the United States, the construction (installation) of air conditioning or HVAC&R systems is
usually performed according to a contract between the owner or developer and the installer, the
contractor. The owner specifies the work or job that is committed to be accomplished within a time
period. The contractor shall furnish and install equipment, ductwork, piping, instruments, and the
related material of the HVAC&R system for a given compensation. The construction of an air conditioning
or HVAC&R system is usually a part of the construction of a building.
INTRODUCTION 1.21
1.22 CHAPTER ONE
Both drawings and specifications are legal documents. They are legal because the designer
conveys the requirements of the owner or developer to the contractor through these documents.
Drawings and specifications define the work to be done by the contractor. They should be clear and
should precisely and completely show the work to be accomplished. Drawings and specifications
are complements to each other. Things to be defined should be shown or specified only once, either
in drawing or in specification. Repetition often causes ambiguity and error.
Drawings
The layout of an HVAC&R system and the locations and dimensions of its equipment, instruments,
ducts, pipes, etc., are best shown and illustrated by drawings. HVAC&R drawings consist of mainly
the following:
 Floor plans. System layout including plant room, fan rooms, mechanical room, ductwork, and
pipelines is always illustrated on floor plans. Each floor has at least one floor plan. HVAC&R
floor plans are always drawn over the same floor plan of the architectural drawing.
 Detail drawings. These drawings show the details of a certain section of an HVAC&R system, or
the detail of the installation of certain equipment, or the connection between equipment and ductwork
or pipeline. Standard details are often used to save time.
 Sections and elevations. Sectional drawings are helpful to show the inner part of a section of a
system, a piece of equipment, or a device. They are especially useful for places such as the plant
room, fan room, and mechanical room where lots of equipment, ductwork, and pipelines are
found. Elevations often show clearly the relationship between the HVAC&R components and the
building structure.
 Piping diagram. This diagram shows the piping layout of the water system(s) and the flow of
water from the central plant to the HVAC&R equipment on each floor.
 Air duct diagram. This diagram illustrates the air duct layout as well as the airflow from the airhandling
unit or packaged unit to the conditioned spaces on each floor through space diffusion
devices.
 Control diagrams. These diagrams show the zone level control systems, each type of functional
control system for air-handling units or packaged units, water systems, outdoor air ventilation
systems, sequencing of compressors, network communication, etc.
 Equipment schedule. This schedule provides the quatity and performance characteristics of the
equipment or device in tabulated form as drawings. Drawings are always available to the installer
at any time but the specifications are not. This is why the equipment schedule should appear on
drawings instead of inside the specification.
 Legends. Symbols and abbreviations are often defined in a legend. ASHRAE has proposed a set
of symbols in the Fundamentals handbook.
Sometimes, three-dimensional isometric drawings are necessay for piping and air duct diagrams.
For a floor plan, a scale of 1/8 in. 	 1 ft (1:100) is often used. The size of the drawings should be
selected according to the size of the project. Drawing sheet sizes of 24  36 in., 30  42 in., and
36  48 in. (610  915 mm, 762  1067 mm, and 915  1219 mm) are widely adopted for large
projects. After completion, every drawing should be carefully checked for errors and omissions.
Specifications
Detailed descriptions of equipment, instruments, ductwork, and pipelines, as well as performances,
operating characteristics, and control sequences are better defined in specifications. Specifications
usually consist of the legal contract between the owner and the contractor, installer, or vendor, and
the technical specifications that specify in detail the equipment and material to be used and how
they are installed.
The Construction Specifications Institute (CSI) has developed a format, called the
Masterformat, for specifications. This Masterformat is widely adopted by most HVAC&R construction
projects. Masterformat promotes standardization and thereby facilitates the retrieval of information
and improves construction communication. The 1988 edition of Masterformat consists of 16
divisions: 01000 General Requirements, 02000 Site Work, 03000 Concrete, . . ., 14000 Conveying
Systems, 15000 Mechanical, and 16000 Electrical. In mechanical, it is subdivided into the following
major sections:
15050 Basic Mechanical Materials and Methods
15250 Mechanical Insulation
15300 Fire Protection
15400 Plumbing
15500 Heating, Ventilating, and Air Conditioning
15550 Heat Generation
15650 Refrigeration
15750 Heat Transfer
15850 Air Handling
15880 Air Distribution
15950 Controls
15990 Testing, Adjusting, and Balancing
Each of the above-mentioned sections may contain: general considerations, equipment and material,
and field installation.
According to whether the wanted vendor is specified or not, specifications can be classified into
two categories: (1) performance specification, which depends only on the performance criteria, and
(2) or-equal specification, which specifies the wanted vendor.
Here are some recommendations for writing an HVAC&R specification:
 The required indoor environmental parameters to be maintained in the conditioned space during
summer and winter outdoor design conditions, such as temperature, humidity, outdoor ventilation
air rates, air cleanliness, sound level, and space pressure, shall be clearly specified in the general
consideration of section 15500.
 Use simple, direct, and clear language without repetition.
 All the terms must be well defined and written in a consistent manner.
 Dont write specifications or refer to other works without having personal knowledge of the content
or even understanding its meaning.
 All the specifications must be tailored to fit the designed HVAC&R system. Never list an item that
is not listed in the specified HVAC&R system, such as the return ducts in a fan-coil system.
1.11 CODES AND STANDARDS
Codes are generally mandatory state or city laws or regulations that force the designer to create the design
without violating human safety and welfare. State and city codes concerning structural integrity,
electrical safety, fire protection, and prevention of explosion of pressure vessels must be followed.
Standards describe consistent methods of testing, specify confirmed design guidelines, and recommend
standard practices. Conformance to standards is usually voluntary. However, if design
INTRODUCTION 1.23
criteria or system performance is not covered in local codes, ASHRAE Standards become the vital
document to assess the indoor air quality in a lawsuit against designer or contractor.
Since the energy crisis of 1973, the U. S. federal government and Congress have published legislation
which profoundly affects the design and operation of HVAC&R systems. As mentioned
before, in 1975, the Energy Policy and Conservation Act and in 1978 the National Energy Conservation
Policy Act were enacted for energy conservation. Also in 1992, the Clean Air Act Amendments
phase out the use of CFCs and recycling of refrigerants, to protect the ozone layer as well as
to provide an acceptable indoor air quality.
Among the HVAC&R related standards published by various institutions, the following by
ASHRAE and some other institutions have greatly influenced the design and operation of air conditioning
systems:
 ASHRAE/IES Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential
Buildings, which is often called the Energy Standard. This standard is under continuous maintenance
by a standing standard committee (SSPC) for regular publication of addenda or revisions.
 ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality, which is often called
the Indoor Air Quality Standard. This standard is under continuous maintenance.
 ANSI/ASHRAE Standard 55-1992, Thermal Environmental Conditions for Human Occupancy
 ASHRAE Standard 15-1992, Safety Code for Mechanical Refrigeration.
 ANSI/ASHRAE Standard 34-1997, Number Designation and Safety Classification of Refrigerants.
This standard is under continuous maintenance.
 ANSI/ASHRAE Standard 135-1995, BACnet: A Data Communication Protocol for Building Automation
and Control Networks.
 Air Conditioning and Refrigeration Institute, Standards for Unitary Equipment.
 National Fire Protection Association NFPA 90A, Standard for the Installation of Air Conditioning
and Ventilating Systems.
 Sheet Metal and Air Conditioning Contractors National Association, HVAC Duct Construction
StandardsMetal and Flexible.
These standards are covered in detail in later chapters.
Recently, many HVAC&R consulting firms seek compliance with the International Organization
for Standardization (ISO) 9000 for quality control. ISO is a specialized international agency for
standardization, at present comprising the national standards institutions of about 90 countries. The
American National Standards Institue (ANSI) is a member that represents the United States.
The ISO 9000 series has now become the quality mangement standard worldwide. It set the guidelines
for management participation, design control, review of specifications, supplier monitoring, and
services provided. The goal of ISO 9000 is to guarantee that products or the services provided by a
manufacturer or an engineering consulting firm are appropriate to the specifications and are within
the tolerances agreed upon. The ISO 9000 series consists of five quality system standards:
ISO 9000 Guidelines of the Selection and the Application of Quality Management and Quality
Assurance Standards
ISO 9001 Modeling of Quality Assurance in Design Development, Manufacturing, Installation
and Servicing
ISO 9002 Modeling of Quality Assurance in Production and Installation
ISO 9003 Modeling of Quality Assurance in Final Inspection and Testing
ISO 9004 Guidelines for the Implementation of Quality Management and to Provide Quality
Systems
There are no legal requirements to be registered to ISO 9000 standards. Meeting ISO 9000 standards
does not mean that the firm provides better service or products than those of nonregistered
companies.
1.24 CHAPTER ONE
Many companies are assembling quality assurance programs which show that they are accepting
compliance to ISO 9000 standards instead of registration. Such an arrangement will identify the
ISO 9000 standard that can be best applied to your operation and also support your customers ISO
9000 quality audit.
1.12 COMPUTER-AIDED DESIGN AND DRAFTING (CADD)
Because personal computers (PCs) provide a low-cost tool for computations and graphics and
owing to the availability of lots of design computer software during the 1980s, computer-aided
design and drafting (CADD) for air conditioning systems has grown quickly in recent years. Today,
computer software is often used to develop design documents, drawings, and specifications in common
use in engineering consulting firms. According to the 1994 CADD Application and User
Survey of design firms in Engineering Systems (June 1994), . . . firms with high productivity
reported that they perform 95 percent of projects on CADD.
In addition to the CADD, there is software available for computer-aided facilities management
(CAFM) for operation and maintenance.
Features of CADD
Compared with conventional design and drafting, CADD of air conditioning systems has the following
features:
 Different trade designers such as those in architecture, HVAC&R, plumbing, fire protection, and
electrical engineers can merge their work in plot output files to produce composite drawings in
different layers. Mechanical system design can integrate with equipment selection programs.
 Graphical model construction of air conditioning systems and subsystems in two-dimensional
(2D) or three-dimensional (3D) presentation uses architectural and structural models as backgrounds.
 CADD links the engineering design models based on calculation and the graphical model based
on drafting.
 It provides the ability to develop and compare the alternate design schemes quickly as well as the
capability to redesign or to match the changes during construction promptly.
 CADD establishes database libraries for design and graphical models.
 It develops design documents including as-built drawings and equipment schedules.
A saving of design time up to 40 percent has been claimed by some engineering consulting firms.
Computer-Aided Design
Current CADD for HVAC&R systems can be classified into two categories: (1) engineering design
and simulation, and (2) drafting graphical model construction. The software for engineering design
can again be subdivided into the following groups:
 Load calculations, energy simulation, and psychrometric analysis. Software for space load calculations
could be a part of the energy simulation software. However, load calculations are often
mainly used to determine the peak or block load at design conditions. These loads are the inputs
for equipment selection. However, the software for energy simulation is used mainly to determine
the optimum selection from different alternatives. Psychrometric analysis is sometimes a useful
tool for load calculation and energy simulation.
INTRODUCTION 1.25
1.26 CHAPTER ONE
 Equipment selection. All large HVAC&R equipment manufacturers offer software, often called an
electronic catalog, to help customers select their product according to the capacity, size, configuration,
and performance.
 Equipment schedule and specifications. Software generates equipment schedules or CSI masterformat
for specifications.
 Price list. Software reports the price of base unit, accessories, and materials.
Computer-Aided Drafting (CAD)
 Software to reproduce architectural drawings is the foundation of CADD. Graphical models of
items such as ducts, pipes, equipment, and system components in HVAC&R systems are developed
against backgrounds of architecture and structural models. Automated computer-aided drafting
(AutoCAD) is the leading PC-based drafting software used by most design firms to assist the
designer during design with automated drawings, databases, and layering schemes.
 Software applications to develop duct and piping work layouts are the two primary CADs used in
air conditioning system design. They can interface with architectural, electrical, and plumbing
drawings through AutoCAD. CAD for duct and piping work can also have hydraulic modeling capacities.
Tags and an HVAC&R equipment schedule including components and accessories can be
produced as well. CAD for ducts and piping work is covered in detail in corresponding chapters.
 Software is also available to produce details of the refrigerating plant, heating plant, and fan room
with accessories, duct, and piping layout. Many manufacturers also supply libraries of files for
AutoCAD users to input the details of their products into CAD drawings.
 HVAC&R CAD for ducts and piping work layout and details often uses graphical model construction.
The elements of the graphical model are linked to its associated information and stored
in the databases. Model construction starts by locating an end device such as a diffuser or terminal.
Add the supply and return pipes from different layers. Software for graphical model construction
has the features of copying, repeating, moving, rotating, and mirror-imaging which enable
the designer to construct the graphical models quickly. After supply and return main ducts, equipment,
and components are added, this completes the supply and return duct systems of this floor.
Duct and piping layouts can also be accomplished by drawing a single-line layout on the CAD,
and the software will convert to two-dimensional or three-dimensional drawings.
Software Requirements
 A CADD software tailored to a specific HVAC&R design or drafting purpose is needed.
 Built-in error checking and error finding are necessary features.
 Transferability of drawings and data among different CADD systems is important. ASHRAE
Handbook 1995, HVAC Applications, recommended that software interface to the most prominent
formats: the Initial Graphics Exchange Specification (IGES), the Standard Exchange Format
(SEF), and the Data Exchange Format (DXF).
 While using the many available computer software programs for loads, energy, and hydraulic
property calculations, one should verify and calibrate the results against field-measured values in
order to improve the effectiveness of the computer software when the conditions and affecting
factors may be varied.
REFERENCES
Amistadi, H., Design and Drawing: Software Review, Engineering Systems, June 1993, pp. 1829.
Amistadi, H., HVAC Product Software, Engineering Systems, January 1994, pp. 5662.
INTRODUCTION 1.27
Amistadi, H., Software Review: HVACR Product Software, Engineering Systems, January 1995, pp. 6069.
Arnold, D., The Evolution of Modern Office Buildings and Air Conditioning, ASHRAE Journal, no. 6, 1999,
pp. 4054.
ASHRAE, ASHRAE Handbook 1995, HVAC Applications, ASHRAE Inc., Atlanta, GA, 1995.
Bengard, M., The Future of Design-Build, Engineered Systems, no. 12, 1999, pp. 6064.
Census Bureau, Snapshot Portrays Profile of Heating and Cooling Units 1991, AC, Heating & Refrigeration
News, Aug. 28, 1995, pp. 2324.
Coad,W. J., Courses and Cures for Building System Defects, Heating/Piping/Air Conditioning, February 1985,
pp. 98100.
Coad,W. J., Safety Factors in HVAC Design, Heating/Piping/Air Conditioning, January 1985, pp. 199203.
Crandell, M. S., NIOSH Indoor Air Quality Investigations 1971 through 1987, Proceedings of the Air Pollution
Control Association Annual Meeting, Dallas, 1987.
Denny, R. J., The CFC Footprint, ASHRAE Journal, November 1987, pp. 2428.
DiIorio, E., and Jennett, Jr., E. J., High Rise Opts for High Tech, Heating/Piping/Air Conditioning, January
1989, pp. 8387.
Department of Energy/Energy Information Administration (DOE/EIA), Household Energy Consumption and
Expenditures 1993, Part I: National Data, DOE/EIA-0321 Washington, DC, 1995.
Department of Energy/Energy Information Administration (DOE/EIA), Nonresidential Building Energy Consumption
Survey: Characteristics of Commercial Buildings, 1995, DOE/EIA-0246(95),Washington, DC, 1997.
Department of Energy/Energy Information Administration (DOE/EIA), Nonresidential Building Energy Consumption
Survey: Commercial Buildings Consumption and Expenditures 1995, DOE/EIA-0318(95),Washington,
DC, 1998.
Energy Information Administration, Commercial Buildings Characteristics 1992 Commercial Building Energy
Consumption Survey,Washington, D.C., 1994.
Faust, F. H., The Early Development of Self-Contained and Packaged Air Conditioner, ASHRAE Transactions,
vol. 92, 1986, Part II B, pp. 353360.
Grant,W. A., From 36 to 56Air Conditioning Comes of Age, ASHVE Transactions, vol. 63, 1957,
pp. 69110.
Guedes, P., Encyclopedia of Architectural Technology, section 4, 1st ed., McGraw-Hill, New York, 1979.
Haines, R.W., How Are Control Systems Designed? Heating/Piping/Air Conditioning, February 1982, p. 94.
Haines, R.W., Operating HVAC Systems, Heating/Piping/Air Conditioning, July 1984, p. 106.
Haines, R.W., and Wilson, C. L., HVAC Systems Design Handbook, McGraw-Hill, New York, 1994.
Houghten, F. C., Blackshaw, J. L., Pugh, E. M., and McDermott, P., Heat Transmission as Influenced by Heat
Capacity and Solar Radiation, ASHVE Transactions, vol. 38, 1932, pp. 231284.
Kohloss, F. H., The Engineers Liability in Avoiding Air Conditioning System Overdesign, ASHRAE
Transactions, vol. 87, 1983, Part I B, pp. 155162.
Korte, B., Existing Buildings: Vast HVAC Resource, Heating/ Piping/ Air Conditioning, March 1989,
pp. 5763.
Korte, B., The Health of the Industry, Heating/Piping/and Air Conditioning, January 1996, pp. 6970.
Lewis, L. L., and Stacey, Jr., A. E., Air Conditioning the Hall of Congress, ASHVE Transactions, vol. 36, 1930,
pp. 333346.
MacCraken, C. D., The Greenhouse Effect on ASHRAE, ASHRAE Journal, June 1989, pp. 5254.
Maynich, P., and Bettano, M., CADD Aids Fast-Track Hospital Expansion, Engineering Systems, June 1994,
pp. 3840.
McClive, J. R., Early Development in Air Conditioning and Heat Transfer, ASHRAE Transactions, vol. 92,
1986, Part II B, pp. 361365.
Miller, A., Thompson, J. C., Peterson, R. E., and Haragan, D.R., Elements of Meteorology, 4th ed., Bell &
Howell Co., Columbus, OH, 1983.
Nagengast, B., The First Centrifugal Chiller: The German Connection, ASHRAE Journal, no. 1, 1998,
pp. 1819.
Penny, T., and Althof, J., Trends in Commercial Buildings, Heating/Piping/Air Conditioning, September 1992,
pp. 5966.
Rowland, F. S., The CFC Controversy: Issues and Answers, ASHRAE Journal, December 1992, pp. 2027.
Simens, J., A Case for Unitary Systems, Heating/Piping/Air Conditioning, May 1982, pp. 60, 6873.
Thomas, V. C., Mechanical Engineering Design Computer Programs for Buildings, ASHRAE Transactions,
1991, Part II, pp. 701710.
Turner, F., Industrial News: Study Shows 49% of Suits Resolved without Payment, ASHRAE Journal, April
1996, pp. 1011.
Whalen, J. M., An Organized Approach to Energy Management, Heating/Piping/Air Conditioning, September
1985, pp. 95102.
Wilson, L., A Case for Central Systems, Heating/Piping/Air Conditioning, May 1982, pp. 6167.
1.28 CHAPTER ONE
CHAPTER 2
PSYCHROMETRICS
2.1
2.1 PSYCHROMETRICS 2.1
Moist Air 2.1
Equation of State of an Ideal Gas 2.2
Equation of State of a Real Gas 2.2
Calculation of the Properties of
Moist Air 2.3
2.2 DALTONS LAW AND THE GIBBSDALTON
LAW 2.3
2.3 AIR TEMPERATURE 2.4
Temperature and Temperature
Scales 2.4
Thermodynamic Temperature Scale 2.5
Temperature Measurements 2.6
2.4 HUMIDITY 2.7
Humidity Ratio 2.7
Relative Humidity 2.7
Degree of Saturation 2.8
2.5 PROPERTIES OF MOIST AIR 2.8
Enthalpy 2.8
Moist Volume 2.9
Density 2.10
Sensible Heat and Latent Heat 2.10
Specific Heat of Moist Air at
Constant Pressure 2.10
Dew-Point Temperature 2.11
2.6 THERMODYNAMIC WET-BULB
TEMPERATURE AND THE WET-BULB
TEMPERATURE 2.11
Ideal Adiabatic Saturation Process 2.11
Thermodynamic Wet-Bulb
Temperature 2.12
Heat Balance of an Ideal Adiabatic
Saturation Process 2.12
Psychrometer 2.12
Wet-Bulb Temperature 2.12
Relationship between Wet-Bulb
Temperature and Thermodynamic
Wet-Bulb Temperature 2.14
2.7 SLING AND ASPIRATION
PSYCHROMETERS 2.14
2.8 HUMIDITY MEASUREMENTS 2.16
Mechanical Hygrometers 2.16
Electronic Hygrometers 2.16
Comparison of Various Methods 2.19
2.9 PSYCHROMETRIC CHARTS 2.19
2.10 DETERMINATION OF
THERMODYNAMIC PROPERTIES ON
PSYCHROMETRIC CHARTS 2.22
Computer-Aided Psychrometric
Calculation and Analysis 2.25
REFERENCES 2.25
2.1 PSYCHROMETRICS
Psychrometrics is the study of the thermodynamic properties of moist air. It is used extensively to
illustrate and analyze the characteristics of various air conditioning processes and cycles.
Moist Air
The surface of the earth is surrounded by a layer of air called the atmosphere, or atmospheric air.
From the point of view of psychrometrics, the lower atmosphere, or homosphere, is a mixture of
dry air (including various contaminants) and water vapor, often known as moist air.
The composition of dry air is comparatively stable. It varies slightly according to geographic
location and from time to time. The approximate composition of dry air by volume percent is the
following:
The amount of water vapor present in moist air at a temperature range of 0 to 100F (17.8 to
37.8C) varies from 0.05 to 3 percent by mass. It has a significant influence on the characteristics of
moist air.
Water vapor is lighter than air. A cloud in the sky is composed of microscopic beads of liquid
water that are surrounded by a thin layer of water vapor. These layers give the cloud the needed
buoyancy to float in the air.
Equation of State of an Ideal Gas
The equation of state of an ideal gas indicates the relationship between its thermodynamic properties,
or
pv  RTR (2.1)
where p  pressure of gas, psf (Pa)
v  specific volume of gas, ft3 /lb (m3 /kg)
R  gas constant, ft  lbf / lbmR (J/kgK)
TR  absolute temperature of gas, R (K)
Since v  V/m, then Eq. (2.1) becomes
pv  mRTR (2.2)
where V  total volume of gas, ft3 (m3)
m  mass of gas, lb (kg)
Using the relationship m  nM, and R  Ro /M, we can write Eq. (2.2) as
pv  nRoTR (2.3)
where n  number of moles, mol
M  molecular weight
Ro  universal gas constant, ft  lbf / lbmR (J/mol K)
Equation of State of a Real Gas
A modified form of the equation of state for a real gas can be expressed as
(2.4)
where A, B, C,   virial coefficients and Z  compressibility factor. The compressibility factor
Z illustrates the degree of deviation of the behavior of the real gas, moist air, from the ideal gas
due to the following:
1. Effect of air dissolved in water
pv
RTR
 1  Ap  Bp2  Cp3      Z
Nitrogen 78.08
Oxygen 20.95
Argon 0.93
Carbon dioxide 0.03
Other gases such as neon, sulfur dioxide, etc. 0.01
2.2 CHAPTER TWO
2. Variation of the properties of water vapor attributable to the effect of pressure
3. Effect of intermolecular forces on the properties of water vapor itself
For an ideal gas, Z  1. According to the information published by the former National Bureau
of Standards of the United States, for dry air at standard atmospheric pressure (29.92 in. Hg, or 760
mm Hg) and a temperature of 32 to 100F (0 to 37.8C) the maximum deviation is about 0.12
percent. For water vapor in moist air under saturated conditions at a temperature of 32 to 100F
(0 to 37.8C), the maximum deviation is about 0.5 percent.
Calculation of the Properties of Moist Air
The most exact calculation of the thermodynamic properties of moist air is based on the formulations
developed by Hyland and Wexler of the U.S. National Bureau of Standards. The psychrometric
chart and tables of ASHRAE are constructed and calculated from these formulations.
Calculations based on the ideal gas equations are the simplest and can be easily formulated. According
to the analysis of Nelson and Pate, at a temperature between 0 and 100F (17.8 and 37.8C),
calculations of enthalpy and specific volume using ideal gas equations show a maximum deviation of
0.5 percent from the exact calculations by Hyland and Wexler. Therefore, ideal gas equations will be
used in this text for the formulation and calculation of the thermodynamic properties of moist air.
Although air contaminants may seriously affect the health of occupants of the air conditioned
space, they have little effect on the thermodynamic properties of moist air since their mass concentration
is low. For simplicity, moist air is always considered as a binary mixture of dry air and water
vapor during the analysis and calculation of its properties.
2.2 DALTONS LAW AND THE GIBBS-DALTON LAW
Daltons law shows that for a mixture of gases occupying a given volume at a certain temperature,
the total pressure of the mixture is equal to the sum of the partial pressures of the constituents of the
mixture, i.e.,
pm  p1  p2    (2.5)
where pm  total pressure of mixture, psia (Pa)
p1, p2, . . .  partial pressure of constituents 1, 2, . . . , psia (Pa)
The partial pressure exerted by each constituent in the mixture is independent of the existence of
other gases in the mixture. Figure 2.1 shows the variation of mass and pressure of dry air and water
vapor, at an atmospheric pressure of 14.697 psia (101,325 Pa) and a temperature of 75F (23.9C).
The principle of conservation of mass for nonnuclear processes gives the following relationship:
mm  ma  mw (2.6)
where mm  mass of moist air, lb (kg)
ma  mass of dry air, lb (kg)
mw  mass of water vapor, lb (kg)
Applying Daltons law for moist air, we have
pat  pa  pw (2.7)
where pat  atmospheric pressure or pressure of the outdoor moist air, psia (Pa)
pa  partial pressure of dry air, psia (Pa)
pw  partial pressure of water vapor, psia (Pa)
PSYCHROMETRICS 2.3
Daltons law is based on experimental results. It is more accurate for gases at low pressures.
Daltons law can be further extended to state the relationship of the internal energy, enthalpy, and
entropy of the gases in a mixture as the Gibbs-Dalton law:
mmum  m1u1  m2u2    
mmhm  m1h1  m2h2     (2.8)
mmsm  m1s1  m2 s2    
where mm  mass of gaseous mixture, lb (kg)
m1, m2, . . .  mass of the constituents, lb (kg)
um  specific internal energy of gaseous mixture, Btu/ lb (kJ /kg)
u1, u2, . . .  specific internal energy of constituents, Btu/ lb (kJ /kg)
hm  specific enthalpy of gaseous mixture, Btu/ lb (kJ /kg)
h1, h2, . . .  specific enthalpy of constituents, Btu/ lb (kJ /kg)
sm  specific entropy of gaseous mixture, Btu/ lb R (kJ/kgK)
s1, s2, . . .  specific entropy of constituents, Btu/ lb R (kJ/kgK)
2.3 AIR TEMPERATURE
Temperature and Temperature Scales
The temperature of a substance is a measure of how hot or cold it is. Two systems are said to have
equal temperatures only if there is no change in any of their observable thermal characteristics
when they are brought into contact with each other. Various temperature scales commonly used to
measure the temperature of various substances are illustrated in Fig. 2.2.
In conventional inch-pound (I-P) units, at a standard atmospheric pressure of 14.697 psia
(101,325 Pa), the Fahrenheit scale has a freezing point of 32F (0C) at the ice point, and a boiling
point of 212F (100C). For the triple point with a pressure of 0.08864 psia (611.2 Pa), the magnitude
on the Fahrenheit scale is 32.018F (0.01C). There are 180 divisions, or degrees, between the
boiling and freezing points in the Fahrenheit scale. In the International System of Units (SI units),
the Celsius or Centigrade scale has a freezing point of 0C and a boiling point of 100C. There are
2.4 CHAPTER TWO
FIGURE 2.1 Mass and pressure of dry air, water vapor, and moist air.
100 divisions between these points. The triple point is at 0.01C. The conversion from Celsius scale
to Fahrenheit scale is as follows:
F  1.8(C)  32 (2.9)
For an ideal gas, at TR  0, the gas would have a vanishing specific volume. Actually, a real gas
has a negligible molecular volume when TR approaches absolute zero. A temperature scale that includes
absolute zero is called an absolute temperature scale. The Kelvin absolute scale has the same
boiling-freezing point division as the Celsius scale. At the freezing point, the Kelvin scale is 273.15 K.
Absolute zero on the Celsius scale is 273.15C. The Rankine absolute scale division is equal to
that of the Fahrenheit scale. The freezing point is 491.67R. Similarly, absolute zero is 459.67F
on the Fahrenheit scale.
Conversions between Rankine and Fahrenheit and between Kelvin and Celsius systems are
R  459.67  F (2.10)
K  273.15  C (2.11)
Thermodynamic Temperature Scale
On the basis of the second law of thermodynamics, one can establish a temperature scale that is
independent of the working substance and that provides an absolute zero of temperature; this is
called a thermodynamic temperature scale. The thermodynamic temperature T must satisfy the
following relationship:
(2.12)
where Q  heat absorbed by reversible engine, Btu/h (kW)
Qo  heat rejected by reversible engine, Btu/h (kW)
TR  temperature of heat source of reversible engine, R (K)
TRo  temperature of heat sink of reversible engine, R (K)
Two of the ASHRAE basic tables, Thermodynamic Properties of Moist Air and Thermodynamic
Properties of Water at Saturation, in ASHRAE Handbook 1993, Fundamentals, are based on
the thermodynamic temperature scale.
TR
TRo

Q
Qo
PSYCHROMETRICS 2.5
FIGURE 2.2 Commonly used temperature scales.
Temperature Measurements
During the measurement of air temperatures, it is important to recognize the meaning of the terms
accuracy, precision, and sensitivity.
1. Accuracy is the ability of an instrument to indicate or to record the true value of the measured
quantity. The error indicates the degree of accuracy.
2. Precision is the ability of an instrument to give the same reading repeatedly under the same conditions.
3. Sensitivity is the ability of an instrument to indicate change of the measured quantity.
Liquid-in-glass instruments, such as mercury or alcohol thermometers, were commonly used in
the early days for air temperature measurements. In recent years, many liquid-in-glass thermometers
have been replaced by remote temperature monitoring and indication systems, made possible
by sophisticated control systems. A typical air temperature indication system includes sensors, amplifiers,
and an indicator.
Sensors. Air temperature sensors needing higher accuracy are usually made from resistance temperature
detectors (RTDs) made of platinum, palladium, nickel, or copper wires. The electrical
resistance of these resistance thermometers characteristically increases when the sensed ambient
air temperature is raised; i.e., they have a positive temperature coefficient . In many engineering
applications, the relationship between the resistance and temperature can be given by
(2.13)
where R  electric resistance, 	
R32, R212  electric resistance, at 32 and 212F (0 and 100C), respectively, 	
T  temperature, F (C)
The mean temperature coefficient  for several types of metal wires often used as RTDs is shown
below:
Many air temperature sensors are made from thermistors of sintered metallic oxides, i.e.,
semiconductors. They are available in a large variety of types: beads, disks, washers, rods, etc. Thermistors
have a negative temperature coefficient. Their resistance decreases when the sensed air temperature
increases. The resistance of a thermistor may drop from approximately 3800 to 3250 	 when
the sensed air temperature increases from 68 to 77F (20 to 25C). Recently developed high-quality
thermistors are accurate, stable, and reliable. Within their operating range, commercially available
thermistors will match a resistance-temperature curve within approximately 0.1F (0.056C). Some
manufacturers of thermistors can supply them with a stability of 0.05F (0.028C) per year. For direct
digital control (DDC) systems, the same sensor is used for both temperature indication, or monitoring,
and temperature control. In DDC systems, RTDs with positive temperature coefficient are
widely used.
Measuring range,F , 	/F
Platinum 400 to 1350 0.00218
Palladium 400 to 1100 0.00209
Nickel 150 to 570 0.0038
Copper 150 to 400 0.0038


R212  R32
180 R32
R  R32(1 
T )
2.6 CHAPTER TWO
Amplifier(s). The measured electric signal from the temperature sensor is amplified at the solid
state amplifier to produce an output for indication. The number of amplifiers is matched with the
number of the sensors used in the temperature indication system.
Indicator. An analog-type indicator, one based on directly measurable quantities, is usually a
moving coil instrument. For a digital-type indicator, the signal from the amplifier is compared with
an internal reference voltage and converted for indication through an analog-digital transducer.
2.4 HUMIDITY
Humidity Ratio
The humidity ratio of moist air w is the ratio of the mass of water vapor mw to the mass of dry air ma
contained in the mixture of the moist air, in lb / lb (kg/kg). The humidity ratio can be calculated as
(2.14)
Since dry air and water vapor can occupy the same volume at the same temperature, we can apply the
ideal gas equation and Daltons law for dry air and water vapor. Equation (2.14) can be rewritten as
(2.15)
where Ra, Rw  gas constant for dry air and water vapor, respectively, ftlbf / lbmR(J/kgK). Equation
(2.15) is expressed in the form of the ratio of pressures; therefore, pw and pat must have the
same units, either psia or psf (Pa).
For moist air at saturation, Eq. (2.15) becomes
(2.16)
where pws  pressure of water vapor of moist air at saturation, psia or psf (Pa).
Relative Humidity
The relative humidity  of moist air, or RH, is defined as the ratio of the mole fraction of water vapor
xw in a moist air sample to the mole fraction of the water vapor in a saturated moist air sample
xws at the same temperature and pressure. This relationship can be expressed as
(2.17)
And, by definition, the following expressions may be written:
(2.18)
(2.19) xws 
nws
na  nws
xw 
nw
na  nw


xw
xws
T,p
ws  0.62198
pws
pat  pws

53.352
85.778
pw
pat  pw
 0.62198
pw
pat  pw
w 
mw
ma

pwVRaTR
PaVRwTR

Ra
Rw
pw
pat  pw
w 
mw
ma
PSYCHROMETRICS 2.7
where na  number of moles of dry air, mol
nw  number of moles of water vapor in moist air sample, mol
nws  number of moles of water vapor in saturated moist air sample, mol
Moist air is a binary mixture of dry air and water vapor; therefore, we find that the sum of the mole
fractions of dry air xa and water vapor xw is equal to 1, that is,
xa  xw  1 (2.20)
If we apply ideal gas equations pwV  nwRoTR and paV  naRoTR, by substituting them into
Eq. (2.19), then the relative humidity can also be expressed as
(2.21)
The water vapor pressure of saturated moist air pws is a function of temperature T and pressure p,
which is slightly different from the saturation pressure of water vapor ps. Here ps is a function of
temperature T only. Since the difference between pws and ps is small, it is usually ignored.
Degree of Saturation
The degree of saturation 	 is defined as the ratio of the humidity ratio of moist air w to the humidity
ratio of the saturated moist air ws at the same temperature and pressure. This relationship can be
expressed as
(2.22)
Since from Eqs. (2.15), (2.20), and (2.21) w  0.62198 xw/xa and ws  0.62198 xws /xa, Eqs. (2.20),
(2.21), and (2.22) can be combined, so that
(2.23)
In Eq. (2.23), pws pat ; therefore, the difference between and 	is small. Usually, the maximum
difference is less than 2 percent.
2.5 PROPERTIES OF MOIST AIR
Enthalpy
The difference in specific enthalpy 
h for an ideal gas, in Btu/ lb (kJ / kg), at a constant pressure can
be defined as

h  cp (T2  T1) (2.24)
where cp  specific heat at constant pressure, Btu/ lb F (kJ/kgK)
T1, T2  temperature of ideal gas at points 1 and 2, F (C)
As moist air is approximately a binary mixture of dry air and water vapor, the enthalpy of the moist
air can be evaluated as
h  ha  Hw (2.25)


	
1  (1 
	)xws

	
1  (1 
	)( pws /pat )
	

w
ws
T,p


pw
pws
T,p
2.8 CHAPTER TWO
where ha and Hw are, respectively, enthalpy of dry air and total enthalpy of water vapor, in Btu/ lb
(kJ/ kg). The following assumptions are made for the enthalpy calculations of moist air:
1. The ideal gas equation and the Gibbs-Dalton law are valid.
2. The enthalpy of dry air is equal to zero at 0F (17.8C).
3. All water vapor contained in the moist air is vaporized at 0F (17.8C).
4. The enthalpy of saturated water vapor at 0F (17.8C) is 1061 Btu/lb (2468 kJ/ kg).
5. For convenience in calculation, the enthalpy of moist air is taken to be equal to the enthalpy of a mixture
of dry air and water vapor in which the amount of dry air is exactly equal to 1 lb (0.454 kg).
Based on the preceeding assumptions, the enthalpy h of moist air can be calculated as
h  ha  whw (2.26)
where hw  specific enthalpy of water vapor, Btu / lb (kJ /kg). In a temperature range of 0 to 100F
(17.8 to 37.8C), the mean value for the specific heat of dry air can be taken as 0.240 Btu/ lb F
(1.005 kJ/kg K). Then the specific enthalpy of dry air ha is given by
ha  cpd T  0.240 T (2.27)
where cpd  specific heat of dry air at constant pressure, Btu/ lb F (kJ/kgK)
T  temperature of dry air, F (C)
The specific enthalpy of water vapor hw at constant pressure can be approximated as
hw  hg0  cpsT (2.28)
where hg0  specific enthalpy of saturated water vapor at 0F (17.8C)its value can be taken
as 1061 Btu/lb (2468 kJ/kg)
cps  specific heat of water vapor at constant pressure, Btu/ lb F (kJ/kgK)
In a temperature range of 0 to 100F (17.8 to 37.8C), its value can be taken as 0.444 Btu/ lb F
(1.859 kJ/kg K). Then the enthalpy of moist air can be evaluated as
h  cpdT  w(hg0  cpsT)  0.240 T  w(1061  0.444 T) (2.29)
Here, the unit of h is Btu/ lb of dry air (kJ / kg of dry air). For simplicity, it is often expressed as
Btu/ lb (kJ / kg).
Moist Volume
The moist volume of moist air v, ft3 /lb (m3 / kg), is defined as the volume of the mixture of the dry
air and water vapor when the mass of the dry air is exactly equal to 1 lb (1 kg), that is,
(2.30)
where V  total volume of mixture, ft3 (m3)
ma  mass of dry air, lb (kg)
In a moist air sample, the dry air, water vapor, and moist air occupy the same volume. If we apply
the ideal gas equation, then
(2.31) v 
V
ma

RaTR
pat  pw
v 
V
ma
PSYCHROMETRICS 2.9
where pat and pw are both in psf (Pa). From Eq. (2.15), pw  patw/(w  0.62198). Substituting this
expression into Eq. (2.31) gives
(2.32)
According to Eq. (2.32), the volume of 1 lb (1 kg) of dry air is always smaller than the volume of
the moist air when both are at the same temperature and the same atmospheric pressure.
Density
Since the enthalpy and humidity ratio are always related to a unit mass of dry air, for the sake of
consistency, air density a, in lb / ft3 (kg/m3), should be defined as the ratio of the mass of dry air to
the total volume of the mixture, i.e., the reciprocal of moist volume, or
(2.33)
Sensible Heat and Latent Heat
Sensible heat is that heat energy associated with the change of air temperature between two state
points. In Eq. (2.29), the enthalpy of moist air calculated at a datum state 0F (17.8C) can be
divided into two parts:
h  (cpd  wcps)T  whg0 (2.34)
The first term on the right-hand side of Eq. (2.34) indicates the sensible heat. It depends on the temperature
T above the datum 0F (17.8C).
Latent heat hfg (sometimes called hig) is the heat energy associated with the change of the state
of water vapor. The latent heat of vaporization denotes the latent heat required to vaporize liquid
water into water vapor. Also, the latent heat of condensation indicates the latent heat to be removed
in the condensation of water vapor into liquid water. When moisture is added to or removed from a
process or a space, a corresponding amount of latent heat is always involved, to vaporize the water
or to condense it.
In Eq. (2.34), the second term on the right-hand side, whg0 , denotes latent heat. Both sensible
and latent heat are expressed in Btu/ lb (kJ / kg) of dry air.
Specific Heat of Moist Air at Constant Pressure
The specific heat of moist air at constant pressure cpa is defined as the heat required to raise its temperature
1F (0.56C) at constant pressure. In (inch-pound) I-P units, it is expressed as Btu/lbF (in SI
units, as J/kgK). In Eq. (2.34), the sensible heat of moist air qsen, Btu/h (W), is represented by
(2.35)
where mass flow rate of moist air, lb/h (kg/ s). Apparently
cpa  cpd  wcps (2.36)
Since cpd and cps are both a function of temperature, cpa is also a function of temperature and, in addition,
a function of the humidity ratio.
For a temperature range of 0 to 100F (17.8 to 37.8C), cpd can be taken as 0.240 Btu/ lb F
(1005 J/kgK) and cps as 0.444 Btu/ lb F (1859 J/kg K). Most of the calculations of cpa(T2  T1)
m?a 
qsen  m?a(cpd  wcps)T  m?acpaT
a 
ma
V

1
v
v 
RaTR(1  1.6078 w)
Pat
2.10 CHAPTER TWO
have a range of w between 0.005 and 0.010 lb/lb (kg/ kg). Taking a mean value of w  0.0075
lb / lb (kg / kg), we find that
cpa  0.240  0.0075 0.444  0.243 Btu/ lbF (1020 J/kgK)
Dew-Point Temperature
The dew-point temperature Tdew is the temperature of saturated moist air of the same moist air sample,
having the same humidity ratio, and at the same atmospheric pressure of the mixture pat. Two
moist air samples at the same Tdew will have the same humidity ratio w and the same partial pressure
of water vapor pw. The dew-point temperature is related to the humidity ratio by
ws ( pat, Tdew)  w (2.37)
where ws  humidity ratio of saturated moist air, lb/ lb (kg / kg). At a specific atmospheric pressure,
the dew-point temperature determines the humidity ratio w and the water vapor pressure pw of the
moist air.
2.6 THERMODYNAMIC WET-BULB TEMPERATURE AND
THE WET-BULB TEMPERATURE
Ideal Adiabatic Saturation Process
If moist air at an initial temperature T1, humidity ratio w1, enthalpy h1, and pressure p flows over a
water surface of infinite length in a well-insulated chamber, as shown in Fig. 2.3, liquid water will
evaporate into water vapor and will disperse in the air. The humidity ratio of the moist air will gradually
increase until the air can absorb no more moisture.
As there is no heat transfer between this insulated chamber and the surroundings, the latent heat
required for the evaporation of water will come from the sensible heat released by the moist air.
This process results in a drop in temperature of the moist air. At the end of this evaporation process,
the moist air is always saturated. Such a process is called an ideal adiabatic saturation process,
where an adiabatic process is defined as a process without heat transfer to or from the process.
PSYCHROMETRICS 2.11
FIGURE 2.3 Ideal adiabatic saturation process.
Thermodynamic Wet-Bulb Temperature
For any state of moist air, there exists a thermodynamic wet-bulb temperature T* that exactly
equals the saturated temperature of the moist air at the end of the ideal adiabatic saturation process
at constant pressure. Applying a steady flow energy equation, we have
(2.38)
where enthalpy of moist air at initial state and enthalpy of saturated air at end of ideal adiabatic
saturation process, Btu/ lb (kJ /kg)
 humidity ratio of moist air at initial state and humidity ratio of saturated air at end
of ideal adiabatic saturation process, lb / lb (kg /kg)
 enthalpy of water as it is added to chamber at a temperature T*, Btu/ lb (kJ /kg)
The thermodynamic wet-bulb temperature T*, F (C), is a unique property of a given moist air
sample that depends only on the initial properties of the moist airw1, h1 and p. It is also a fictitious
property that only hypothetically exists at the end of an ideal adiabatic saturation process.
Heat Balance of an Ideal Adiabatic Saturation Process
When water is supplied to the insulation chamber at a temperature T* in an ideal adiabatic saturation
process, then the decrease in sensible heat due to the drop in temperature of the moist air is just
equal to the latent heat required for the evaporation of water added to the moist air. This relationship
is given by
(2.39)
where T1  temperature of moist air at initial state of ideal adiabatic saturation process, F (C)
hfg *  latent heat of vaporization at thermodynamic wet-bulb temperature, Btu/ lb (J /kg)
Since cpa  cpd  w1cps, we find, by rearranging the terms in Eq. (2.39),
(2.40)
Also
(2.41)
Psychrometer
A psychrometer is an instrument that permits one to determine the relative humidity of a moist air
sample by measuring its dry-bulb and wet-bulb temperatures. Figure 2.4 shows a psychrometer,
which consists of two thermometers. The sensing bulb of one of the thermometers is always kept
dry. The temperature reading of the dry bulb is called the dry-bulb temperature. The sensing bulb of
the other thermometer is wrapped with a piece of cotton wick, one end of which dips into a cup of
distilled water. The surface of this bulb is always wet; therefore, the temperature that this bulb
measures is called the wet-bulb temperature. The dry bulb is separated from the wet bulb by a radiation-
shielding plate. Both dry and wet bulbs are cylindrical.
Wet-Bulb Temperature
When unsaturated moist air flows over the wet bulb of the psychrometer, liquid water on the surface
of the cotton wick evaporates, and as a result, the temperature of the cotton wick and the wet bulb
drops. This depressed wet-bulb reading is called the wet-bulb temperature T, and the difference between
the dry-bulb and wet-bulb temperatures is called the wet-bulb depression.
Let us neglect the conduction along the thermometer stems to the dry and wet bulbs and also
assume that the temperature of the water on the cotton wick is equal to the wet-bulb temperature of
the moist air. Since the heat transfer from the moist air to the cotton wick is exactly equal to the
latent heat required for vaporization, then, at steady state, the heat and mass transfer per unit area of
the wet-bulb surface can be calculated as
(2.42)
where hc, hr  mean convective and radiative heat transfer coefficients, Btu/h ft2 F (W/m2 K)
hd  mean convective mass-transfer coefficient, lb /h ft2 (kg/ s m2)
T  temperature of undisturbed moist air at a distance from wet bulb, F (C)
T  wet-bulb temperature, F (C)
Tra  mean radiant temperature, F (C)
w1, ws  humidity ratio of moist air and saturated air film at surface of cotton wick, lb / lb
(kg/kg)
latent heat of vaporization at wet-bulb temperature, Btu/ lb (J /kg)
Based on the correlation of cross-flow forced convective heat transfer for a cylinder, NuD  C Ren
Pr0.333, and based on the analogy between convective heat transfer and convective mass transfer, the
following relationship holds: hd  hc/(cpaLe0.6667). Here, Nu is the Nusselt number, Re the
Reynolds number, and Le the Lewis number. Also C is a constant.
Substituting this relationship into Eq. (2.42), we have
(2.43)

The term T  T indicates the wet-bulb depression.
Combining Eqs. (2.43) and (2.44) then gives
(2.45)
Relationship between Wet-Bulb Temperature and Thermodynamic
Wet-Bulb Temperature
Wet-bulb temperature is a function not only of the initial state of moist air, but also of the rate of
heat and mass transfer at the wet bulb. Comparing Eq. (2.40) with Eq. (2.45), we find that the wetbulb
temperature measured by using a psychrometer is equal to thermodynamic wet-bulb temperature
only when the following relationship holds:
(2.46)
2.7 SLING AND ASPIRATION PSYCHROMETERS
Sling and aspiration psychrometers determine the relative humidity through the measuring of the
dry- and wet-bulb temperatures.
A sling psychrometer with two bulbs, one dry and the other wet, is shown in Fig. 2.5a. Both dry
and wet bulbs can be rotated around a spindle to produce an airflow over the surfaces of the dry and
wet bulbs at an air velocity of 400 to 600 fpm (2 to 3 m/ s). Also a shield plate separates the dry
and wet bulbs and partly protects the wet bulb against surrounding radiation.
An aspiration psychrometer that uses a small motor-driven fan to produce an air current flowing
over the dry and wet bulbs is illustrated in Fig. 2.5b. The air velocity over the bulbs is usually kept
at 400 to 800 fpm (2 to 4 m/ s). The dry and wet bulbs are located in separate compartments and are
entirely shielded from the surrounding radiation. When the space dry-bulb temperature is within a
range of 75 to 80F (24 to 27C) and the space wet-bulb temperature is between 65 and 70F (18
and 21C), the following wet-bulb constants K can be used to calculate the humidity ratio of the
moist air:
Aspiration psychrometer K  0.000206 1/F
Sling psychrometer K  0.000218 1/F
After the psychrometer has measured the dry- and wet-bulb temperatures of the moist air, the
humidity ratio w can be calculated by Eq. (2.43). Since the saturated water vapor pressure can be
found from the psychrometric table, the relative humidity of moist air can be evaluated through
Eq. (2.15).
According to the analysis of Threlkeld (1970), for a wet-bulb diameter of 0.1 in. (2.5 mm) and
an air velocity flowing over the wet bulb of 400 fpm (2 m/ s), if the dry-bulb temperature is 90F
(32.2C) and the wet-bulb temperature is 70F (21.1C ), then (T  T*)/(T  T) is about 2.5
percent. Under the same conditions, if a sling psychrometer is used, then the deviation may be
reduced to about 1 percent. If the air velocity flowing over the wet bulb exceeds 400 fpm (2 m/ s),
there is no significant reduction in the deviation.
Distilled water must be used to wet the cotton wick for both sling and aspiration psychrometers.
Because dusts contaminate them, cotton wicks should be replaced regularly to provide a clean surface
for evaporation.

2.8 HUMIDITY MEASUREMENTS
Humidity sensors used in HVAC&R for direct humidity indication or operating controls are separated
into the following categories: mechanical hygrometers and electronic hygrometers.
Mechanical Hygrometers
Mechanical hygrometers operate on the principle that hygroscopic materials expand when they
absorb water vapor or moisture from the ambient air. They contract when they release moisture
to the surrounding air. Such hygroscopic materials include human and animal hairs, plastic polymers
like nylon ribbon, natural fibers, wood, etc. When these materials are linked to mechanical
linkages or electric transducers that sense the change in size and convert it into electric signals, the
results in these devices can be calibrated to yield direct relative-humidity measurements of the
ambient air.
Electronic Hygrometers
There are three types of electronic hygrometers: Dunmore resistance hygrometers, ion-exchange
resistance hygrometers, and capacitance hygrometers.
Dunmore Resistance Hygrometer. In 1938, Dunmore of the National Bureau of Standards developed
the first lithium chloride resistance electric hygrometer in the United States. This instrument
depends on the change in resistance between two electrodes mounted on a hygroscopic material.
Figure. 2.6a shows a Dunmore resistance sensor. The electrodes could be, e.g., a double-threaded
winding of noble-metal wire mounted on a plastic cylinder coated with hygroscopic material. The
wires can also be in a grid-type arrangement with a thin film of hygroscopic material bridging the
gap between the electrodes.
At a specific temperature, electric resistance decreases with increasing humidity. Because the
response is significantly influenced by temperature, the results are often indicated by a series of
isothermal curves. Relative humidity is generally used as the humidity parameter, for it must
be controlled in the indoor environment. Also the electrical response is more nearly a function of
relative humidity than of the humidity ratio.
The time response to accomplish a 50 percent change in relative humidity varies directly according
to the air velocity flowing over the sensor and also is inversely proportional to the saturated
vapor pressure. If a sensor has a response time of 10 s at 70F (21C), it might need a response time
of 100 s at 10F (12C).
Because of the steep variation of resistance corresponding to a change in relative humidity, each
of the Dunmore sensors only covers a certain range of relative-humidity measurements. A set of
several Dunmore sensors is usually needed to measure relative humidity between 1 percent and 100
percent.
A curve for output, in direct-current (dc) volts, versus relative humidity is shown in Fig. 2.6b
for a typical Dunmore sensor. It covers a measuring range of 10 to 80 percent, which is usually
sufficient for the indication of relative humidity for a comfort air conditioned space. This typical
Dunmore sensor has an accuracy of 3 percent when the relative humidity varies between 10 and
60 percent at a temperature of 70F (21C, see Fig. 2.6b). Its accuracy reduces to 4 percent when
the relative humidity is in a range between 60 and 80 percent at the same temperature.
In addition to lithium chloride, lithium bromide is sometimes used as the sensor.
Ion-Exchange Resistance Hygrometer. The sensor of a ion-exchange resistance electric hygrometer
is composed of electrodes mounted on a baseplate and a high-polymer resin film, used as a
humidity-sensing material, cross-linking the electrodes as shown in Fig. 2.7a and b. Humidity is
measured by the change in resistance between the electrodes. When the salt contained in the humidity-
sensitive material bridging the electrodes becomes ion-conductive because of the presence of
water vapor in the ambient air, mobile ions in the polymer film are formed. The higher the relative
humidity of the ambient air, the greater the ionization becomes, and therefore, the greater the
concentration of mobile ions. On the other hand, lower relative humidity reduces the ionization and
results in a lower concentration of mobile ions. The resistance of the humidity-sensing material
reflects the change of the relative humidity of the ambient air.
In Fig. 2.7b, the characteristic curves of an ion-exchange resistance electric hygrometer show
that there is a nonlinear relationship between resistance R and relative humidity I/>. These sensors
cover a wider range than Dunmore sensors, from 20 to 90 percent relative humidity.
Capacitance Hygrometer. The commonly used capacitance sensor consists of a thin-film plastic
foil. A very thin gold coating covers both sides of the film as electrodes, and the film is mounted
2.18 CHAPTER TWO
inside a capsule. The golden electrodes and the dividing plastic foil form a capacitor. Water vapor
penetrates the gold layer, which is affected by the vapor pressure of the ambient air and, therefore,
the ambient relative humidity. The number of water molecules absorbed on the plastic foil determines
the capacitance and the resistance between the electrodes.
FIGURE 2.7 Ion-exchange resistance-type electric hygrometer. (a)
Front and side view of sensor; (b) characteristic curve of R versus .
(Reprinted by permission of General Eastern Instruments.)
Comparison of Various Methods
The following table summarizes the sensor characteristics of various methods to be used within a
temperature range of 32 to 120F (0 to 50C) and a range of 10 to 95 percent RH:
Psychrometers are simple and comparatively low in cost. They suffer no irreversible damage at 100
percent RH, as do the sensors of electric hygrometers. Unfortunately, complete wet-bulb depression
readings of psychrometers become difficult when relative humidity drops below 20 percent or when
the temperature is below the freezing point. For remote monitoring, it is difficult to keep sufficient
water in the water reservoir. Therefore, psychrometers are sometimes used to check the temperature
and relative humidity in the air conditioned space manually.
Mechanical hygrometers directly indicate the relative humidity of the moist air. They are also
simple and relatively inexpensive. Their main drawbacks are their lack of precision over an extensive
period and their lack of accuracy at extreme high and low relative humidities. Electronic
hygrometers, especially the polymer film resistance and the capacitance types, are commonly used
for remote monitoring and for controls in many air conditioning systems. Both the electronic and
mechanical hygrometers need regular calibration. Initial calibrations are usually performed either
with precision humidity generators using two-pressure, two-temperature, and divided-flow systems
or with secondary standards during manufacturing (refer to ASHRAE Standard 41.6-1982, Standard
Method for Measurement of Moist Air Properties). Regular calibrations can be done with a
precision aspiration psychrometer or with chilled mirror dew-point devices.
Air contamination has significant influence on the performance of the sensor of electronic and
mechanical hygrometers. This is one of the reasons why they need regular calibration.
2.9 PSYCHROMETRIC CHARTS
Psychrometric charts provide a graphical representation of the thermodynamic properties of moist
air, various air conditioning processes, and air conditioning cycles. The charts are very helpful during
the calculation, analysis, and solution of the complicated problems encountered in air conditioning
processes and cycles.
Basic Coordinates. The currently used psychrometric charts have two types of coordinates:
 h-w chart. Enthalpy h and humidity ratio w are basic coordinates. The psychrometric charts published
by ASHRAE and the Chartered Institution of Building Services Engineering (CIBSE) are
h-w charts.
 T-w chart. Temperature T and humidity ratio w are basic coordinates. Most of the psychrometric
charts published by the large manufacturers in the United States are T-w charts.
For an atmospheric pressure of 29.92 in. Hg (760 mm Hg), an air temperature of 84F (28.9C),
and a relative humidity of 100 percent, the humidity ratios and enthalpies found from the psychrometric
charts published by ASHRAE and Carrier International Corporation are shown below:
Operating method Accuracy, % RH
Psychrometer Wet-bulb depression 3
Mechanical Dimensional change 3 to 5
Dunmore Electric resistance 1.5
Ion-exchange Electric resistance 2 to 5
Capacitance Electric capacitance and resistance 3 to 5
PSYCHROMETRICS 2.19
The last digit for humidity ratios and for enthalpies read from ASHRAEs chart is an approximation.
Nevertheless, the differences between the two charts are less than 1 percent, and these are considered
negligible.
In this handbook, for manual psychrometric calculations and analyses, ASHRAEs chart will be
used.
Temperature Range and Barometric Pressure. ASHRAEs psychrometric charts are constructed
for various temperature ranges and altitudes. In Appendix B only the one for normal temperature,
that is, 32 to 120F (0 to 50C), and a standard barometric pressure at sea level, 29.92 in. Hg
(760 mm Hg), is shown. The skeleton of ASHRAEs chart is shown in Fig. 2.8.
Enthalpy Lines. For ASHRAEs chart, the molar enthalpy of moist air is calculated from the
formulation recommended by Hyland and Wexler (1983) in their paper Formulations for the Thermodynamic
Properties of Dry Air from 173.15 K to 473.15 K, and of Saturated Moist Air from
173.15 K to 473.15 K, at Pressures to 5 MPa. For ASHRAEs chart, the enthalpy h lines incline at an
angle of 25 to the horizontal lines. The scale factor for the enthalpy lines Ch, Btu/lb ft (kJ/kgm), is
(2.47)
where Lh  shortest distance between enthalpy lines h2 and h1, ft (m).

Humidity Ratio Lines. In ASHRAEs chart, the humidity ratio w lines are horizontal. They form
the ordinate of the psychrometric chart. The scale factor Cw, lb/lb ft (kg /kgm), for w lines in
ASHRAEs chart is
(2.48)
where Lw  vertical distance between w2 and w1, ft (m). For ASHRAEs chart, the humidity ratio w
can be calculated by Eq. (2.15).
Constant-Temperature Lines. For ASHRAEs chart, since enthalpy is one of the coordinates,
only the 120F constant-temperature T line is a true vertical. All the other constant-temperature
lines incline slightly to the left at the top. From Eq. (2.27), T  ha /cpd; therefore, one end of the T
line in ASHRAES chart can be determined from the enthalpy scale at w  0. The other end can be
determined by locating the saturated humidity ratio ws on the saturation curve.
Saturation Curve. A saturation curve is a locus representing a series of state points of saturated
moist air. For ASHRAEs chart, the enthalpy of saturation vapor over liquid water or over ice at a
certain temperature is calculated by the formula recommended in the Hyland and Wexler paper. The
humidity ratios of the saturated moist air ws between 0 and 100F (17.8 and 37.8C) in the psychrometric
chart can also be calculated by the following simpler polynomial:
(2.49)
where Ts  saturated temperature of moist air, F (C)
a1  0.00080264
a2  0.000024525
a3  2.5420 106
a4   2.5855 108
a5  4.038 1010
If we use Eq. (2.49) to calculate ws, the error is most probably less than 0.000043 lb/lb (kg / kg). It
is far smaller than the value that can be identified on the psychrometric chart, and therefore the calculated
ws is acceptable.
Relative-Humidity and Moist Volume Lines. For ASHRAEs chart, relative-humidity  lines,
thermodynamic wet-bulb T* lines, and moist volume v lines all are calculated and determined
based on the formulations in Hyland and Wexlers paper.
Thermodynamic Wet-Bulb Lines. For ASHRAEs chart, only thermodynamic wet-bulb T* lines
are shown. Since the dry-bulb and the thermodynamic wet-bulb temperatures coincide with each
other on the saturation curve, one end of the T* line is determined. The other end of the T* line can
be plotted on the w line where w  0. Let the state point of the other end of the T* line be represented
by 1. 
Cooling and Dehumidifying Curves. The two cooling and dehumidifying curves plotted
on ASHRAEs chart are based on data on coil performance published in the catalogs of U.S.
manufacturers. These curves are very helpful in describing the actual locus of a cooling and dehumidifying
process as well as determining the state points of air leaving the cooling coil.
2.10 DETERMINATION OF THERMODYNAMIC PROPERTIES ON
PSYCHROMETRIC CHARTS
There are seven thermodynamic properties or property groups of moist air shown on a psychrometric
chart:
1. Enthalpy h
2. Relative humidity 
3. Thermodynamic wet-bulb temperature T*
4. Barometric or atmospheric pressure pat
5. Temperature T and saturation water vapor pressure pws
6. Density  and moist volume v
7. Humidity ratio w, water vapor pressure pw, and dew-point temperature Tdew
The fifth, sixth, and seventh are thermodynamic property groups. These properties or properties
groups are independent of each other except that the difference in slope between the enthalpy h line
and thermodynamic wet-bulb temperature T* line is small, and it is hard to determine their intersection.
Usually, atmospheric pressure pat is a known value based on the altitude of the location. Then, in
the fifth property group, pws is a function of temperature T only. In the sixth property group, according
to Eq. (2.33), a  1/v; that is, air density and moist volume are dependent on each other. In the
seventh property group, for a given value of pat, properties w, pw, and Tdew are all dependent on each
other.
When pat is a known value, and if the moist air is not saturated, then any two known independent
thermodynamic properties can determine the magnitudes of the remaining unknown properties. If
the moist air is saturated, then any independent property will determine the remaining magnitudes.
Example 2.1. The design indoor air temperature and relative humidity of an air conditioned space
at sea level are 75F (23.9C) and 50 percent. Find the humidity ratio, the enthalpy, and the density
of the indoor moist air
1. By using the ASHRAE chart
2. By calculation
Determine also the dew-point and thermodynamic wet-bulb temperatures of the moist air. The
following information is required for the calculations:

1. Plot the space point r on ASHRAEs chart by first finding the space temperature Tr  75F on
the abscissa and then following along the 75F constant-temperature line up to a relative humidity
 50 percent, as shown in Fig. 2.9.
Draw a horizontal line from the space point r. This line meets the ordinate, humidity ratio w, at
a value of wr  0.00927 lb/lb (kg/ kg). This is the humidity ratio of the indoor space air. Draw a
line parallel to the enthalpy line from the space point r. This line meets the enthalpy scale line at
a value of hr  28.1 Btu/ lb. This is the enthalpy of the indoor space air. Draw a horizontal line
from the space point r to the left. This line meets the saturation curve at a dew-point temperature
of 55F (12.8C). Draw a line parallel to the thermodynamic wet-bulb temperature lines through
the space point r. The perpendicular scale to this line shows a thermodynamic wet-bulb temperature
 62.5F (16.9C). Draw a line parallel to the moist volume lines through the space point r.
The perpendicular scale to this line shows a moist volume vr  13.68 ft3/lb (0.853 m3 / kg).
2. The calculations of the humidity ratio, enthalpy, and moist volume are as follows:
From Eq. (2.49), the humidity ratio of the saturated air at the dry-bulb temperature is
ws  0.00080264  0.000024525T  2.542e-06T 2  2.5855e-08T 3  4.038e-10T 4
 0.00080264  0.0018394  0.014299  0.010908  0.012776
 0.018809 lb/ lb
According to Eq. (2.16), the saturated water vapor pressure of the indoor air is
T*r
PSYCHROMETRICS 2.23
FIGURE 2.9 Thermodynamic properties determined from ASHRAEs psychrometric chart.
From Eq. (2.21), the water vapor pressure of indoor air is
pw  pws  0.5 0.4314  0.2157 psia
Then, from Eq. (2.15), the humidity ratio of the indoor air is
From Eq. (2.29), the enthalpy of the indoor moist air is
hr  cpdTF  w (hg  cpsT)
hr  0.240 75  0.009264(1061  0.444 75)
 28.14 Btu/ lb
From Eq. (2.32), the moist volume of the indoor air is
Then, from Eq. (2.33), the density of the indoor moist air is
Comparison of the thermodynamic properties read directly from ASHRAEs chart and the calculated
values is as follows:
Apparently, the differences between the readings from ASHRAEs chart and the calculated values
are very small.
Example 2.2. An HVAC&R operator measured the dry- and wet-bulb temperatures in an air
conditioned space as 75F (23.9C) and 63F (17.2C), respectively. Find the relative humidity
of this air conditioned space by using ASHRAEs chart and by calculation. The humidity ratios
of the saturated air at temperatures of 75 and 63F are 0.018809 and 0.012355 lb/lb (kg/ kg),
respectively.
Solution. It is assumed that the difference between the wet-bulb temperature as measured by
sling or aspiration psychrometer and the thermodynamic wet-bulb temperature is negligible. At a
measured dry-bulb temperature of 75F (23.9C) and a wet-bulb temperature of 63F (17.2C), the
relative humidity read directly from ASHRAEs chart is about 51.8 percent. From Sec. 2.6, the wetbulb
constant K for a sling psychrometer is 0.000218 1/F. Then, from Eq. (2.43), the humidity
ratio of the space air can be calculated as

From Eq. (2.15), the vapor pressure of the space air is
And from Eq. (2.16), the saturated vapor pressure at a space temperature of 75F is
Hence, from Eq. (2.21), the calculated relative humidity of the space air is
The difference between the value read directly from ASHRAEs chart and the calculated one is
52.53 percent  51.8 percent  0.7 percent.
Computer-Aided Psychrometric Calculation and Analysis
There are two kinds of psychrometric computer-aided software available on the market: psychrometric
calculations and psychrometric graphics. Most of the psychrometric software is Windowsbased
computer programs.
Software for psychrometric calculations can determine any one of the thermodynamic properties
of the moist air if two of the independent properties are known. Psychrometric calculation software
usually also finds the thermodynamic property of the mixture of airstreams and provides altitude
effect adjustments. Software programs for psychrometric graphics are far more powerful tools than
psychrometric calculations. On the computer screen, the following is shown: a psychrometic chart,
any number of labeled air conditioned state points, the corresponding air conditioned processes and
the air conditioning cycle, and the spreadsheet that lists the thermodynamic properties, airflow in
cubic feet per minute, and the heat transfer during the air conditioning processes. The thermodynamic
properties of any of the state points and therefore the characteristics of the air conditioning
process and cycle can be varied. As a result, the operation of the air system at either full load or part
load, under cooling or heating modes, can be investigated and analyzed.

Aslam, S., Charmchi, M., and Gaggioli, R. A., Psychrometric Analysis for Arbitrary Dry-Gas Mixtures and
Pressures Using Microcomputers, ASHRAE Transactions, 1986, Part I B, pp. 448460.
Hedlin, C. P., Humidity Measurement with Dunmore Type Sensors, Symposium at ASHRAE Semiannual
Meeting, ASHRAE Inc., New York, February 1968.
Hyland, R. W., and Wexler, A., Formulations for the Thermodynamic Properties of Dry Air from 173.15 K to
473.15 K, and of Saturated Moist Air from 173.15 K to 372.15 K, at Pressures to 5 MPa, ASHRAE Transactions,
1983, Part II A, pp. 520535.
Hyland, R. W., and Wexler, A., Formulations for the Thermodynamic Properties of the Saturated Phases of H2O
from 173.15 K to 473.15 K, ASHRAE Transactions, 1983, Part II A, pp. 500519.
Kamm, V., New Psychrometric Software Offer Free on the Internet, The Air Conditioning, Heating and Refrigeraion
News , July 22, 1996, pp. 1819.
McGee, T. D., Principles and Methods of Temperature Measurements, 1st ed.,Wiley, New York, 1988.
Nelson, R. M., and Pate, M., A Comparison of Three Moist Air Property Formulations for Computer Applications,
ASHRAE Transactions, 1986, Part I B, pp. 435447.
Stewart, R. B., Jacobsen, R. T., and Becker, J. H., Formulations for Thermodynamic Properties of Moist Air at
Low Pressures as Used for Construction of New ASHRAE SI Unit Psychrometric Charts, ASHRAE Transactions,
1983, Part II A, pp. 536548.
Threlkeld, J. L., Thermal Environmental Engineering, 2d ed., Prentice-Hall, Englewood Cliffs, NJ, 1970.
The Trane Company, Psychrometry, La Crosse, WI, 1979.
Wang, S. K., Air Conditioning, vol. 1, 1st ed., Hong Kong Polytechnic, Hong Kong, 1987.
2.26 CHAPTER TWO
CHAPTER 3
HEAT AND MOISTURE
TRANSFER THROUGH
BUILDING ENVELOPE
3.1
3.1 BUILDING ENVELOPE 3.2
3.2 HEAT-TRANSFER
FUNDAMENTALS 3.2
Conductive Heat Transfer 3.3
Convective Heat Transfer 3.4
Radiant Heat Transfer 3.5
Overall Heat Transfer 3.6
Heat Capacity 3.8
3.3 HEAT-TRANSFER COEFFICIENTS 3.8
Coefficients for Radiant Heat
Transfer 3.8
Coefficients for Forced Convection 3.9
Coefficients for Natural Convection 3.10
Surface Heat-Transfer Coefficients 3.10
3.4 MOISTURE TRANSFER 3.11
Sorption Isotherm 3.11
Moisture-Solid Relationship 3.12
Moisture Migration in Building
Materials 3.13
Moisture Transfer from the Surface of
the Building Envelope 3.14
Convective Mass-Transfer
Coefficients 3.15
Moisture Transfer in Building
Envelopes 3.16
3.5 CONDENSATION IN BUILDINGS 3.17
Visible Surface Condensation 3.17
Concealed Condensation within the
Building Envelope 3.18
3.6 THERMAL INSULATION 3.18
Basic Materials and Thermal
Properties 3.19
Moisture Content of Insulation
Material 3.19
Economic Thickness 3.21
Thermal Resistance of Airspaces 3.21
3.7 SOLAR ANGLES 3.22
Basic Solar Angles 3.22
Hour Angle and Apparent Solar
Time 3.24
Solar Angle Relationships 3.24
Angle of Incidence and Solar
Intensity 3.24
3.8 SOLAR RADIATION 3.25
Solar Radiation for a Clear Sky 3.26
Solar Radiation for a Cloudy Sky 3.28
3.9 FENESTRATION 3.29
Types of Window Glass (Glazing) 3.29
Optical Properties of Sunlit Glazing 3.30
3.10 HEAT ADMITTED THROUGH
WINDOWS 3.32
Heat Gain for Single Glazing 3.32
Heat Gain for Double Glazing 3.34
Shading Coefficients 3.36
Solar Heat Gain Factors and Total
Shortwave Irradiance 3.37
Selection of Glazing 3.39
3.11 SHADING OF GLASS 3.40
Indoor Shading Devices 3.40
External Shading Devices 3.42
Shading from Adjacent Buildings 3.43
3.12 HEAT EXCHANGE BETWEEN THE
OUTER BUILDING SURFACE AND ITS
SURROUNDINGS 3.46
Sol-Air Temperature 3.47
3.13 COMPLIANCE WITH ASHRAE/IESNA
STANDARD 90.1-1999 FOR BUILDING
ENVELOPE 3.48
General Requirements 3.48
Mandatory Provisions 3.49
Prescriptive Building Envelope
Option 3.49
Building Envelope Trade-Off
Option 3.50
3.14 ENERGY-EFFICIENT AND COSTEFFECTIVE
MEASURES FOR BUILDING
ENVELOPE 3.50
Exterior Walls 3.50
Windows 3.50
Infiltration 3.51
Energy-Efficient Measures for
Commercial Buildings in
the United States 3.51
REFERENCES 3.51
3.1 BUILDING ENVELOPE
Building envelope consists of the building components that enclose conditioned spaces. Heat,
moisture, and contaminants may be transferred to or from the outdoors or unconditioned spaces and
therefore affect the indoor environment of the conditioned space.
Building envelope used for air conditioned space in buildings consists of mainly walls, roofs,
windows, ceilings, and floors. There are two types of partitions: exterior partitions and demising
partitions. An exterior partition is an opaque, translucent, or transparent solid barrier that separates
conditioned space from outdoors or space which is not enclosed. A demising partition is a solid
barrier that separates conditioned space from enclosed unconditioned space.
 An exterior wall is a solid exterior partition which separates conditioned space from the outdoors.
Exterior walls are usually made from composite layers including any of the following: stucco,
bricks, concrete, concrete blocks, wood, thermal insulation, vapor barrier, airspace, and interior finish.
The gross exterior wall area is the sum of the window area, door area, and exterior wall area.
 A partition wall is an interior solid barrier which separates a conditioned space from others. Partition
walls are usually made from interior surface finishes on two sides, wooden studs and boards,
concrete blocks, concrete, bricks, and thermal insulation. A demising partition wall separates a
conditioned space from an enclosed unconditioned space.
 Wall below grade is a solid barrier below ground level which often separates a basement or a
crawl space from soil.
 A roof is an exterior partition that has a slope less than 60 from horizontal and has a conditioned
space below directly or through a ceiling indirectly. Roofs are usually made from clay tile, waterproof
membrane, concrete and lightweight concrete, wood, and thermal insulation.
 A ceiling is an interior partition that separates the conditioned space from a ceiling plenum. The
ceiling plenum may or may not be air conditioned. Ceilings are usually made from acoustic tile or
boards, thermal insulation, and interior surface finishes.
 An exterior floor is a horizontal exterior partition under conditioned space. A floor placed over a
ventilated basement or a parking space is an exterior floor. Exterior floors are usually made from
wood, concrete, thermal insulation, and face tiles.
 Slab on grade is a concrete floor slab on the ground. There is usually a vapor barrier, thermal
insulation, and gravel and sand fill between the concrete slab and the ground.
 A window is glazing of any transparent or translucent material plus sash, frame, mullions, and
dividers in the building envelope. Glazing is usually made from glass and transparent plastics.
Frames are often made from wood, aluminium, and steel. The window area is the area of the
surface of glazing plus the area of frame, sash, and mullions. A fenestration is any area on the
exterior building envelope which admits light indoors. Fenestrations include windows, glass
doors, and skylights.
 A skylight is glazing having a slope of less than 60 from the horizontal. There is often a conditioned
space below skylight(s).
3.2 HEAT-TRANSFER FUNDAMENTALS
Heat transfer between two bodies, two materials, or two regions is the result of temperature
difference. The science of heat transfer has provided calculations and analyses to predict rates of
heat transfer. The design of an air conditioning system must include estimates of heat transfer
between the conditioned space, its contents, and its surroundings, to determine cooling and heating
loads. Heat-transfer analysis can be described in three modes: conduction, convection, and
radiation.
3.2 CHAPTER THREE
Conductive Heat Transfer
Conduction is the mechanism of heat transfer in opaque solid media, such as through walls and
roofs. For one-dimensional steady-state heat conduction qk, Btu /h (W), Fouriers law gives the following
relationship:
(3.1)
where k  thermal conductivity, Btu /h ft F (W/mC)
A  cross-sectional area normal to heat flow, ft2 (m2)
T  temperature, F (C)
x  coordinate dimension along heat flow, ft (m)
Equation (3.1) shows that the rate of heat transfer is directly proportional to the temperature gradient
dT/dx, the thermal conductivity k, and the cross-sectional area A. The minus sign indicates that
the heat must flow in the direction of decreasing temperature; i.e., if the temperature decreases as x
increases, the gradient dT/dx is negative, so that heat conduction is a positive quantity.
For steady-state heat conduction through a plane composite wall with perfect thermal contact
between each layer, as shown in Fig. 3.1, the rate of heat transfer through each section of the
composite wall must be the same. From Fouriers law of conduction,
(3.2)
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.3
FIGURE 3.1 Steady-state one-dimensional heat conduction through a
composite wall.
kA, kB, kC  thermal conductivity of layers A, B, and C, respectively, of composite wall,
Btu/h ft F (W/mC)
Eliminating T2 and T3, we have
(3.3)
For a multilayer composite wall of n layers in perfect thermal contact, the rate of conduction heat
transfer is given as
(3.4)
Subscript n indicates the nth layer of the composite wall.
In Eq. (3.2), conduction heat transfer of any of the layers can be written as
(3.5)
where R*  thermal resistance, hF/Btu (C/W). In Eq. (3.5), an analogy can be seen between
heat flow and Ohms law for an electric circuit. Here the temperature difference T  T1  T2 indicates
thermal potential, analogous to electric potential. Thermal resistance R* is analogous to electric
resistance, and heat flow qk is analogous to electric current.
The total conductive thermal resistance of a composite wall of n layers RT, hF/Btu (C/W),
can be calculated as
(3.6)
where  thermal resistances of layers 1, 2,   , n layer of the composite wall,
hF/Btu (C/W). The thermal circuit of a composite wall of three layers is shown in the lower
part of Fig. 3.1.
Convective Heat Transfer
Convective heat transfer occurs when a fluid comes in contact with a surface at a different temperature,
such as the heat transfer taking place between the airstream in a duct and the duct wall.
Convective heat transfer can be divided into two types: forced convection and natural or free
convection. When a fluid is forced to move along the surface by an outside motive force, heat is
transferred by forced convection. When the motion of the fluid is caused by the density difference
of the two streams in the fluid as a product of contacting a surface at a different temperature, the
result is called natural or free convection.
The rate of convective heat transfer qc, Btu /h (W), can be expressed in the form of Newtons
law of cooling as
qc  hcA(Ts  T) (3.7)
where hc  average convective heat-transfer coefficient, Btu/h ft2 F (W/m2 C)
Ts  surface temperature, F (C)
T  temperature of fluid away from surface, F (C)
In Eq. (3.7), the convective heat-transfer coefficient hc is usually determined empirically. It is
related to a dimensionless group of fluid properties, such as the correlation of flat-plate forced
R*1 , R*2 ,    R*n
R*T  R*1 	 R*2 	    	 R*n
R* 
L
k A
qk 
k A
L
T 
T
R*
qk 
T1  Tn	1
L1 /(k1 A) 	 L2 /(k2 A) 	    	 Ln /(kn A)
qk 
T1  T4
LA /(kA A) 	 LB /(kB A) 	 LC /(kC A)
3.4 CHAPTER THREE
convection, as shown in the following equation:
(3.8)
where C  constant and k  thermal conductivity, Btu /h ft F (W/mC). In Eq. (3.8), NuL is the
Nusselt number, which is based on a characteristic length L. Characteristic length can be the length
of the plate, the diameter of the tube, or the distance between two plates, in feet (meters). The
Reynolds number ReL is represented by the following dimensionless group:
(3.9)
where  density of fluid, lb/ ft3 (kg/m3)
v  velocity of fluid, ft/s (m/s)
	 absolute viscosity of the fluid, lb/ ft  s (kg /m s)
 kinematic viscosity of fluid, ft2/s (m2/s)
The Prandtl number Pr is represented by
(3.10)
where cp  specific heat at constant pressure of the fluid, Btu / lbF (J/kgC). Fluid properties
used to calculate dimensionless groups are usually related to the fluid temperature Tf . That is,
(3.11)
Convective heat transfer can also be considered analogous to an electric circuit. From Eq. (3.7),
the convective thermal resistance Rc*, hF/Btu (C/W), is given as
(3.12)
Radiant Heat Transfer
In radiant heat transfer, heat is transported in the form of electromagnetic waves traveling at the
speed of light. The net rate of radiant transfer qr, Btu /h (W), between a gray body at absolute
temperature TR1 and a black surrounding enclosure at absolute temperature TR2 (for example, the
approximate radiation exchange between occupant and surroundings in a conditioned space) can be
calculated as
(3.13)
where 
 Stefan-Boltzmann constant  0.1714  108 Btu/h ft2 R4 (W/m2 K4 )
A1  area of gray body, ft2 (m2 )
1  emissivity of surface of gray body
TR1, TR2  absolute temperature of surfaces 1 and 2, R (K)
If we multiply the right-hand side of Eq. (3.13) by (TR1  TR2)/(TR1  TR2), and let the thermal
resistance
(3.14) R*r 
TR1  TR2

A1 1(T R1
4  T R2
4 )

1
hr A1
qr 

A1 1(T R1
4  T R2
4 )
Rc* 
1
hc A
Tf 
Ts  T
2
Pr 
3600 	cp
k
ReL 
vL
	

vL

NuL 
hc L
k
 C ReL n Prm
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.5
where hr  radiant heat-transfer coefficient, Btu/h ft2 F (W/m2 C), we find that as a
consequence,
(3.15)
where T1, T2  temperature of surfaces 1 and 2, F (C).
If either of the two surfaces is a black surface or can be approximated as a black surface, as was
previously assumed for the conditioned space surroundings, the net rate of radiant heat transfer
between surfaces 1 and 2 can be evaluated as
(3.16)
where F12  shape factor for a diffuse emitting area A1 and a receiving area A2. The thermal resistance
can be similarly calculated.
Overall Heat Transfer
In actual practice, many calculations of heat-transfer rates are combinations of conduction, convection,
and radiation.
Consider the composite wall shown in Fig. 3.1; in addition to the conduction through the
wall, convection and radiation occur at inside and outside surfaces 1 and 4 of the composite wall.
At the inside surface of the composite wall, the rate of heat transfer qi, Btu /h (W), consists of convec-
tive heat transfer between fluid, the air, and solid surface qc and the radiant heat transfer qr, as
follows:
(3.17)
where Ti  indoor temperature, F (C). From Eq. (3.17), the inside surface heat-transfer coefficient
hi at the liquid-to-solid interface, Btu /h ft2 F (W/m2 C) is
and the thermal resistance Ri* of the inner surface due to convection and radiation, hF/Btu
(C/W), is
(3.18)
Similarly, at the outside surface of the composite wall, the rate of heat-transfer qo, Btu/h ft2 F
(W/m2 C), is
(3.19)
where ho  outside surface heat-transfer coefficient at fluid-to-solid interface, Btu /h ft2 F
(W/m2 C)
A4  area of surface 4, ft2 (m2 )
T4  temperature of surface 4, F (C)
To  outdoor temperature, F (C)
The outer thermal resistance Ro*, hF/Btu (C/W), is
For one-dimensional steady-state heat transfer, the overall heat-transfer rate of the composite
R*o 
1
ho A4
qo  ho A4(T4  To)
Ri* 
1
hi A1
hi  hc 	 hr
 hi A1(Ti  T1)
qi  qc 	 qr  hc A1(Ti  T1) 	 hr A1(Ti  T1)
qr 

A1F1 2(T R1
4  T R2
4 )
qr  hr A1(T1  T2)
3.6 CHAPTER THREE
wall q, Btu /h (W), can be calculated as
(3.20)
where U  overall heat-transfer coefficient, often called the U value, Btu/h ft2 F (W/m2 C)
A  surface area perpendicular to heat flow, ft2 (m2 )
RT*  overall thermal resistance of composite wall, hF/Btu (C/W)
and
(3.21)
Also, the thermal resistances can be written as
Therefore, the overall heat transfer coefficient U is given as
(3.22)
For plane surfaces, area A  A1  A2  A3  A4. For cylindrical surfaces, because the inside
and outside surface areas are different, and because UA  1/RT*, it must be clarified whether the
area is based on the inside surface area Ai, outside surface area Ao, or any chosen surface area.
For convenient HVAC&R heat-transfer calculations, the reciprocal of the overall heat-transfer
coefficient, often called the overall R value RT, h ft2 F/Btu (m2 C/W), is used. So RT can be
expressed as
(3.23a)
where Ri, Ro  R values of inside and outside surfaces of composite wall, h ft2 F/Btu
(m2 C/W)
R1, R2,   , Rn  R values of components 1, 2, . . ., n, h ft2 F/Btu (m2 C/W)
and
(3.23b)
Sometimes, for convenience, the unit of R is often omitted; for example, R-10 means the R value
equals 10 h ft2 F/Btu (m2 C/W).
A building envelope assembly, or a building shell assembly, includes the exterior wall assembly
(i.e., walls, windows, and doors), the roof and ceiling assembly, and the floor assembly. The areaweighted
average overall heat-transfer coefficient of an envelope assembly Uav, Btu/h ft2 F
(W/m2 C), can be calculated as
(3.24)
where A1, A2 . . ., An  area of individual elements 1, 2, . . ., n of envelope assembly, ft2 (m2 )
Ao  gross area of envelope assembly, ft2 (m2)
U1, U2, . . ., Un  overall heat-transfer coefficient of individual paths 1, 2, . . ., n of the
envelope assembly, such as paths through windows, paths through walls,
and paths through roof, Btu /h ft2 F (W/m2 C)
Uav 
U1 A1 	 U2 A2 	    	 Un An
Ao
Ri 
1
hi
R1 
L1
k1
R2 
L2
k2
   Rn 
Ln
kn
Ro 
1
ho
RT 
1
U
 Ri 	 R1 	 R2 	    	 Rn 	 Ro
U 
1
1/hi 	 LA / kA 	 LB / kB 	 LC / kC 	 1/ho
R*A 
LA
kA A R*B 
LB
kB B R*C 
LC
kCC
R*T  R*i 	 R*A 	 R*B 	 R*C 	 R*o 
1
UA
q  qi  qk  qo  UA(Ti  To) 
Ti  To
R*T
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.7
Heat Capacity
The heat capacity (HC) per square foot (meter) of an element or component of a building envelope
or other structure depends on its mass and specific heat. Heat capacity HC, Btu/ ft2 F (kJ/m2 C),
can be calculated as
(3.25)
where m  mass of building material, lb (kg)
c  specific heat of building material, Btu/ lbF (kJ/kgC)
A  area of building material, ft2 (m2)
 density of building material, lb / ft3 (kg/m3 )
L  thickness or height of building material, ft (m)
3.3 HEAT-TRANSFER COEFFICIENTS
Determination of heat-transfer coefficients to be used for load calculations or year-round energy
estimates is complicated by the following types of variables:
 Building envelopes, exterior wall, roof, glass, partition wall, ceiling, or floor
 Fluid flow, turbulent flow, or laminar flow, forced or free convection
 Heat flow, horizontal heat flow in a vertical surface, or an upward or downward heat flow in a
horizontal surface
 Space air diffusion, ceiling or sidewall inlet, or others
 Time of operationsummer, winter, or other seasons
Among the three modes of heat transfer, convection processes and their related coefficients are the
least understood, making analysis difficult.
Coefficients for Radiant Heat Transfer
For a radiant exchange between the inner surface of an exterior wall and the surrounding surfaces
(such as the surfaces of partition walls, ceilings, and floors) in an air conditioned room, the sum of
the shape factors F1 n can be considered as unity. If all surfaces are assumed to be black, then the
radiative heat-transfer coefficient hr, Btu/h ft2 F (W/m2 C), can be calculated as
(3.26)
where TRis  absolute temperature of inner surface of exterior wall, roof, or external window glass,
R (K)
TRrad  absolute mean radiant temperature of surrounding surfaces, R (K)
Often TRrad is approximately equal to the air temperature of the conditioned space TRr when both are
expressed in degrees Rankine (kelvins).
Radiant heat transfer coefficients calculated according Eq. (3.26) for various surface temperatures
and temperature differences between the inner surface of any building envelope and the surrounded
surfaces are presented in Table 3.1. From Table 3.1, hr depends on the absolute temperature of inner
surface TRis and the temperature difference Tis  Trad.
hr 

 (T Ris
4  T Rrad
4 )
TRis  TRrad
HC
A

Lc
3.8 CHAPTER THREE
Coefficients for Forced Convection
Before one can select the mean convective heat-transfer coefficient hc in order to calculate the rate
of convective heat transfer or to determine the overall heat-transfer coefficient U value during cooling
load calculations, the type of convection (forced or natural) must be clarified. Forced convection
between the surfaces of the building envelope and the conditioned space air occurs when the airhandling
system is operating or there is a wind over the outside surface. In an indoor space, free
convection is always assumed when the air-handling system is shut off, or there is no forced-air
motion over the surface involved.
The convective heat-transfer coefficient hc also depends on the air velocity v flowing over the
surface, the configuration of the surface, the type of space air diffusion, and the properties of the
fluid. According to Kays and Crawford (1980), a linear or nearly linear relationship between hc and
v holds. Even though the air velocity v in the occupied zone may be only 30 fpm (0.15 m/ s) or even
lower when the air-handling system is operating, however, the mode of heat transfer is still considered
as forced convection and can be expressed as
(3.27)
where n  exponential index, usually between 0.8 and 1
v  bulky air velocity of fluid 0.5 to 1 ft (0.15 to 0.30 m) from surface, fpm (m/min)
A, B  constants
On the basis of test data from Wong (1990), Sato et al. (1972), and Spitler et al. (1991), as well as
many widely used energy estimation computer programs, the convective heat-transfer coefficient
for forced convection hc, Btu/h ft2 F (W/m2 C), for indoor surfaces can be determined as
(3.28)
For outside surfaces, the surface heat-transfer coefficient ho  hc 	 hr, Btu/h ft2 F (W/m2 C),
can be calculated as
(3.29)
where vwind  wind speed, fpm.
ho  1.8 	 0.004vwind
hc  1.0 	 0.008v
hc  A 	 Bv
n
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.9
TABLE 3.1 Radiative Heat Transfer Coefficients hr
Temperature
difference
Inner surface temperature Tis, F
TisTrad, 60 70 75 80 85
F (520R) (530R) (535R) (540R) (545R)
Btu/h ft2 F
1 0.871 0.909 0.935 0.968 0.990
2 0.866 0.910 0.936 0.966 0.990
3 0.862 0.908 0.935 0.963 0.989
5 0.856 0.904 0.930 0.958 0.984
10 0.844 0.892 0.918 0.945 0.971
20 0.819 0.868 0.893 0.919 0.945
Note: Calculations made by assuming that Trad  Tr and the emissivity of inner
surface i  1.
Coefficients for Natural Convection
For natural convection, the empirical relationship between the dimensionless groups containing the
convective heat-transfer coefficient hc can be expressed as follows:
(3.30)
In Eq. (3.30), GrL, called the Grashof number, is based on the characteristic length L.
(3.31)
where  coefficient of volume expansion of fluid, 1/ R (1/K)
g  acceleration of gravity, ft / s2 (m/s2)
 density of fluid, lb/ ft3 (kg/m3)
C  constant
And Ra is called the Rayleigh number, and Ra  GrPr.
Natural convective heat transfer is difficult to evaluate because of the complexity of the recirculating
convective stream of room air that is the result of the temperature distribution of the surfaces
and the variation of temperature profile of the stream. Computer programs using numerical techniques
have been developed recently. In actual practice, simplified calculations are often adopted.
Altmayer et al. (1983) in their recent experiments and analyses found that The ASHRAE free
convection correlations provide a fair prediction of the heat flux to the room air from cold and hot
surfaces. The errors are mainly caused by the variation of temperature of the convective airstreams
after contact with cold or hot surfaces. In simplified calculations, this variation can only be included
in the calculation of the mean space air temperature T.
Many rooms have a vertical wall height of about 9 ft (2.7 m). If the temperature difference
Ts T  1F (0.56C), the natural convection flow is often turbulent. If Ts  T  1F (0.56C),
as in many cooling load calculations between a partition wall and space air, the flow is laminar.
Based on data published in ASHRAE Handbook 1997, Fundamentals, the natural convection
coefficients hc, Btu/h ft2 F (W/m2 C), are given as follows:
Vertical plates:
Large plates, turbulent flow hc  0.19(Tsa)0.33 (3.32)
Small plates, laminar flow (3.33)
Horizontal plates, facing upward when air is heated or facing downward when air is cooled:
Large plates, turbulent flow hc  0.22(Tsa)0.33 (3.34)
Small plates, laminar flow (3.35)
Horizontal plates, facing upward when air is cooled or facing downward when air is heated:
Small plates (3.36)
Here, Tsa indicates the temperature difference between the surface and the air.
Surface Heat-Transfer Coefficients
The surface heat-transfer coefficient h, sometimes called the surface conductance, Btu/h ft2 F
(W/m2 C), is the combination of convective and radiant heat-transfer coefficients; that is,
hc  0.12Tsa
L 0.25
hc  0.27Tsa
L 0.25
hc  0.29Tsa
L 0.25
GrL  g 2(Ts  T )
	2
NuL  C(GrLPr)n  C RaL n
3.10 CHAPTER THREE
h  hc 	 hr. Table 3.2 lists the h values for various surface types at Tsa  10F (5.6C) and
Tsa  1F (0.56C) during summer and winter design conditions. The values are based upon the
following:
 Trad  Ta, where Ta is the air temperature, F (C).
 Tsa indicates temperature difference between surface and air, F (C).
 Emissivity of the surface  0.9 and  0.2.
3.4 MOISTURE TRANSFER
Moisture is water in the vapor, liquid, and solid states. Building materials exposed to excessive
moisture may degrade or deteriorate as a result of physical changes, chemical changes, and biological
processes. Moisture accumulated inside the insulating layer also increases the rate of heat
transfer through the building envelope. Moisture transfer between the building envelope and the
conditioned space air has a significant influence on the cooling load calculations in areas with hot
and humid climates.
Sorption Isotherm
Moisture content X, which is dimensionless or else in percentage, is defined as the ratio of the mass
of moisture contained in a solid to the mass of the bone-dry solid. An absorption isotherm is a
constant-temperature curve for a material in which moisture content is plotted against an increased
ambient relative humidity during an equilibrium state; i.e., the rate of condensation of water vapor
on the surface of the material is equal to the rate of evaporation of water vapor from the material. A
desorption isotherm is also a constant-temperature curve for a material. It is a plot of moisture content
versus a decreased ambient relative humidity during equilibrium state.
Many building materials show different absorption and desorption isotherms. The difference in
moisture content at a specific relative humidity between the absorption and desorption isotherms is
called hysteresis. Figure 3.2 shows absorption and desorption isotherms of a building material.
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.11
TABLE 3.2 Surface Heat-Transfer Coefficients h, Btu/h ft2 F
Surface emissivity 
0.90 0.20
Indoor surface Indoor surface
Tsa  10F Tsa  10F
Direction of Outdoor
Description heat flow Summer Winter Tsa  1F surface Summer Winter Tsa  1F
Forced convection
30 fpm 2.21 2.11 2.16 1.46 1.43 1.44
50 fpm 2.37 2.27 2.32 1.62 1.59 1.60
660 fpm (7.5 mph) 4.44
1320 fpm (15 mph) 7.08
Free convection
Horizontal surface Upward 1.36 1.27 1.17 0.62 0.60 0.44
Vertical surface 1.42 1.33 1.15 0.68 0.66 0.30
Horizontal surface Downward 1.03
Note: Assume space temperature Tr  74F year-round and Tr  Trad; here Trad indicates the mean radiant temperature of the surroundings.
Temperature also has an influence on the moisture content of many building materials. At a constant
relative humidity in ambient air, the moisture content of a building material will be lower at a
higher temperature.
When a building material absorbs moisture, heat as heat of sorption is evolved. If liquid water is
absorbed by the material, an amount of heat ql, Btu / lb (kJ /kg) of water absorbed, similar to the
heat of solution, is released. This heat results from the attractive forces between the water molecules
and the molecules of the building material. If water vapor is absorbed, then the heat released
qv, Btu / lb (kJ / kg), is given by
(3.37)
where hfg  latent heat of condensation, Btu/ lb (kJ /kg). Heat of sorption of liquid water ql varies
with equilibrium moisture content for a given material. The lower the X and the  of ambient air,
the higher will be the value of ql. For pine, ql may vary from 450 Btu/lb (1047 kJ/kg) for nearly
bone-dry wood to 20 Btu/ lb (46.5 kJ /kg) at a moisture content of 20 percent. Many building materials
have very low ql values compared with hfg, such as a ql of 40 Btu/ lb (93 kJ / kg) for sand.
Moisture-Solid Relationship
Many building materials have numerous interstices and microcapillaries of radius less than 4 
106 in. (0.1 	m), which may or may not be interconnected. These interstices and microcapillaries
provide large surface areas to absorb water molecules. Moisture can be bound to the solid surfaces
by retention in the capillaries and interstices, or by dissolution into the cellular structures of fibrous
materials.
When the relative humidity  of ambient air is less than 20 percent, moisture is tightly bound to
individual sites in the monomolecular layer (region A, in Fig. 3.2). Moisture moves by vapor diffusion.
The binding energy is affected by the characteristics of the surface, the chemical structure of
the material, and the properties of water.
qv  ql 	 hfg
3.12 CHAPTER THREE
FIGURE 3.2 Sorption isotherms.
When relative humidity  is 20 to 60 percent (region B, Fig. 3.2), moisture is bound to the
surface of the material in multimolecular layers and is held in microcapillaries. Moisture begins to
migrate in liquid phase, and its total pressure is reduced by the presence of moisture in microcapillaries.
The binding energy involved is mainly the latent heat of condensation.
When   60 percent, i.e., in region C of Fig. 3.2, moisture is retained in large capillaries. It is
relatively free for removal and chemical reactions. The vapor pressure of the moisture is influenced
only moderately because of the moisture absorbed in regions A and B. Moisture moves mainly in a
liquid phase.
Because the density of the liquid is much greater than that of water vapor, the moisture content
in building materials is mostly liquid. Unbound moisture can be trapped in interstices having a
radius greater than 4  105 in. (1 	m), without significantly lowering the vapor pressure. Free
moisture is the moisture in excess of the equilibrium moisture content in a solid. For insulating
materials in which interstices are interconnected, air penetrates through these open-cell structures,
and, moisture can be accumulated in the form of condensate and be retained in the large capillaries
and pores.
Moisture Migration in Building Materials
Building envelopes are not constructed only with open-cell materials. The airstream and its associated
water vapor cannot penetrate building envelopes. Air leakage can only squeeze through the
cracks and gaps around windows and joints. However, all building materials are moisturepermeable;
in other words, moisture can migrate across a building envelope because of differences
in moisture content or other driving potentials.
According to Wong and Wang (1990), many theories have been proposed by scientists to predict
the migration of moisture in solids. The currently accepted model of moisture flow in solids is
based upon the following assumptions:
 Moisture migrates in solids in both liquid and vapor states. Liquid flow is induced by capillary
flow and concentration gradients; vapor diffusion is induced by vapor pressure gradients.
 During the transport process, the moisture content, the vapor pressure, and the temperature are
always in equilibrium at any location within the building material.
 Heat transfer within the building material is in the conduction mode. It is also affected by latent
heat from phase changes.
 Vapor pressure gradients can be determined from moisture contents by means of sorption
isotherms.
 Ficks law is applicable.
 Only one-dimensional flow across the building envelope is considered. Building materials are
homogeneous.
If the temperature gradient is small, for simplicity, the mass flux for one-dimensional
flow, lb/h ft2 (kg/hm2), can be expressed as
(3.38)
where Dlv  mass diffusivity of liquid and vapor, ft2 /h (m2/h)
X  moisture content, lb / lb (kg / kg) dry solid
w  density of water, lb/ ft3 (kg/m3 )
x  coordinate dimension along moisture flow, ft (m)
A  area of building envelope perpendicular to moisture flow, ft2 (m2 )
Mass diffusivities of some building materials as a function of moisture content are shown in
Fig. 3.3.
m?w /A  wDlv
dX
dx
m? w /A
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.13
Moisture Transfer from the Surface of the Building Envelope
At a certain time instant, moisture migrating from any part in the building envelope to its surface
must be balanced by convective moisture transfer from the surface of the building envelope to the
ambient air and the change of the moisture content as well as the corresponding mass concentration
at the surface of the building envelope. Such a convective moisture transfer is often a part of the
space latent cooling load. Analogous to the rate of convective heat transfer [Eq. (3.7)], the rate of
convective moisture transfer , lb/h (kg/ h), can be calculated as
(3.39)
where hm  convective mass-transfer coefficient, ft /h (m/h)
Aw  contact area between moisture and ambient air, ft2 (m2 )
Cws, Cwr  mass concentration of moisture at surface of building envelope and of space air, lb/ ft3
(kg/m3)
It is more convenient to express the mass concentration difference Cws  Cwr in terms of a
humidity ratio difference ws  wr. Here ws and wr represent the humidity ratio at the surface of the
building envelope and of the space air, respectively, lb/lb (kg / kg). In terms of mass concentration
we can write
(3.40)
(3.41)
Then the rate of convective moisture transfer can be rewritten as
(3.42)
In Eq. (3.42), the surface humidity ratio ws depends on the moisture content Xs, the temperature Ts,
and therefore the vapor pressure pws in the interstices of the surface of the building envelope. From
the known Xs, Ts, and the sorption isotherm, the corresponding relative humidity s at the surface
can be determined. If the difference between relative humidity and the degree of saturation 	is
m?w 
ahmAw(ws  wr)
Cwr 
awr
Cws 
aws
m? w  hm Aw (Cws  Cwr)
m? w
3.14 CHAPTER THREE
FIGURE 3.3 Mass diffusivity Dlv of some building materials. (Abridged
with permission from Bruin et al. Advances in Drying, Vol. 1, 1979.)
ignored, then
(3.43)
where 	s  degree of saturation at surface. In Eq. (3.43), wss represents the humidity ratio of the
saturated air. It can be determined from Eq. (2.49) since Ts is a known value.
In Eq. (3.42), the contact area Aw between the liquid water at the surface of the building envelope
and the space air is a function of moisture content Xs. A precise calculation of Aw is very complicated,
but if the surface area of the building material is As, then a rough estimate can be made
from
(3.44)
Convective Mass-Transfer Coefficients
The Chilton-Colburn analogy relates the heat and mass transfer in these forms:
(3.45)
where v  air velocity remote from surface, ft / s (m/ s)
cpa  specific heat of moist air, Btu / lbF (J /kgC)
hm  convective mass-transfer coefficient, ft /h (m/ s)
 kinematic viscosity, ft2/s (m2/s)
 thermal diffusivity of air, ft2/s (m2/s)
Daw  mass diffusivity for water vapor diffusing through air, ft2/s (m2/s)
Sc  Schmidt number
Subscript a indicates dry air and w, water vapor.
For air at 77F (25C), cpa  0.243 Btu/ lbF (1020 kJ/kgC), va  1.74  104 ft2/s
(1.62  105 m2 / s),  2.44  104 ft2 / s (2.27  105 m2 / s), Daw  2.83 ft2 / s (0.263 m2 / s),
and a  0.0719 lb/ ft3 (1.15 kg/m3 ), we can show that
and that
At a space temperature of 77F (25C), therefore,

hc (0.713)2/3
0.0719  0.243  (0.614)2/3  63.2hc
hm 
hc Pr 2/3
a cpa Sc2/3
Sc 
a
Daw

1.74  104
2.83  104  0.614
Pr 
a


1.74  104
2.44  104
 0.713
hc
acpa
a
2/3
 hm  vw
Daw2/3
hc
a cpa
Pr2/3 
hm
v
Sc2/3
jH  jD
Aw  Xs As
ws 
	swss 
swss
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.15
Moisture Transfer in Building Envelopes
Moisture transfer in building envelopes of a typical residential building can proceed along two
paths, as shown in Fig. 3.4:
1. Moisture migrates inside the building material mainly in the form of liquid if the relative humidity
of the ambient air is more than 50 percent. It will be transported to the indoor or outdoor air
by convective mass transfer. The driving potentials are the moisture content of the building material,
the vapor pressure gradient inside the building material, and the humidity ratio gradient between
the surface and the ambient air.
2. Air leakage and the associated water vapor infiltrates or exfiltrates through the cracks and
gaps around the windows, doors, fixtures, outlets, and between the joints. Air and moisture enter the
cavities and the airspace in the building envelope. If the sheathing is not airtight, air leakage and its
water vapor penetrate the perforated insulating board and discharge to the atmosphere, as shown in
Fig. 3.4. If the sheathing is a closed-cell, airtight insulating board, then the airstream may infiltrate
through gaps between the joints of the insulating board or through cracks between the window sill
and the external wall and discharge to the atmosphere.
The driving potential of the air leakage and the associated water vapor is the total pressure
differential between indoor and outdoor air across the building envelope due to wind effects, stack
effect, mechanical ventilation, or a combination of these. Moisture is moving in the vapor form.
In leaky buildings, the moisture transfer by means of air leakage is often far greater than the
moisture migration through solids. For better-sealed commercial buildings, moisture migration
through the building envelope may be important.
3.16 CHAPTER THREE
FIGURE 3.4 Moisture transfer in building envelope along two paths.
Installation of vapor retarders is an effective means to block or to reduce the moisture transfer in
the building envelope. Vapor retarders are covered in greater detail later.
3.5 CONDENSATION IN BUILDINGS
When moist air contacts a solid surface whose temperature is lower than the dew point of the moist
air, condensation occurs on the surface in the form of liquid water, or sometimes frost. Condensation
can damage the surface finish, deteriorate the material and cause objectionable odors, stains,
corrosion, and mold growth; reduce the quality of the building envelope with dripping water; and
fog windows. Two types of condensation predominate in buildings:
1. Visible surface condensation on the interior surfaces of external window glass, below-grade
walls, floor slabs on grade, and cold surfaces of inside equipment and pipes
2. Concealed condensation within the building envelope
Visible Surface Condensation
To avoid visible surface condensation, either the dew point of the indoor air must be reduced to a
temperature below that of the surface, or the indoor surface temperature must be raised to a level
higher than the dew point of the indoor air. The indoor surface temperature of the building envelope,
such as T1 of the composite exterior wall in Fig. 3.1, or any cold surface where condensation
may occur can be calculated from Eqs. (3.17) and (3.20) as
(3.46)
where q  rate of heat transfer through building envelope, Btu /h (W)
hi  inside surface heat-transfer coefficient, Btu/h ft2 F (W/m2 C)
Increasing the thermal insulation is always an effective and economical way to prevent condensation
because it saves energy and raises the surface temperature of the building envelope and other
cold surfaces.
The dew point temperature of the indoor air Tr  dew is a function of humidity ratio wr of the space
air. At a specific space temperature Tr, the lower the relative humidity, the lower the Tr  dew. There
are several ways to lower the dew point of the indoor space air:
 By blocking and controlling infiltration of hot, humid outside air
 By reducing indoor moisture generation
 By using a vapor retarder, such as polyethylene film or asphalt layer, to prevent or decrease the
migration of moisture from the soil and the outside wetted siding and sheathing
 By introducing dry outdoor air through mechanical or natural ventilation when doing so is not in
conflict with room humidity criteria
 By using dehumidifiers to reduce the humidity ratio of the indoor air
During winter, any interior surface temperature of the wall, roof, or glass is always lower than
the indoor space temperature. Better insulation and multiple glazing increase the interior surface
temperature and, at the same time, reduce the heat loss, providing a more comfortable indoor environment
than is possible by decreasing the relative humidity. To reduce excessive indoor humidity
in many industrial applications, a local exhausting booth that encloses the moisture-generating
source is usually the remedy of first choice.
T1  Ti 
U(Ti  To)
hi
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.17
During summer, the outer surface of chilled water pipes and refrigerant pipes (even the cold
supply duct) is at a lower temperature than the indoor air temperature. Sufficient pipe and duct
insulation is needed to prevent surface condensation. When an air-handling system is shut down,
the heavy construction mass often remains at a comparatively lower surface temperature as the indoor
humidity rises. To avoid indoor surface condensation, the enclosure must be tight enough to
prevent the infiltration of hot, humid air from outdoors.
Concealed Condensation within the Building Envelope
Usually, concealed condensation within building materials does not accumulate as a result of moisture
migration. Excessive free moisture, usually in the form of liquid or frost, at any location inside
the building envelope results in a higher moisture content than in surrounding areas. Therefore, it
produces a moisture migration outward from that location rather than into it.
During cold seasons, concealed condensation in building envelopes is mainly caused by the
warm indoor air, usually at a dew-point temperature of 32F (0C) and above. It leaks outward
through the cracks and openings and enters the cavities, the gaps between component layers, or
even the penetrable insulating material. It ultimately contacts a surface at a temperature lower than
the dew point of the indoor air. Donnelly et al. (1976) discovered that a residential stud wall panel
with a poor vapor retarder accumulated about 3 lb(14.6 kg) of moisture per square foot (meter) of
wall area during a period of 31 days in winter. This moisture accumulated in the mineral fiber insulating
board adjacent to the cold side of the sheathing and at the interface between the insulating
board and the sheathing.
The rate of condensation , lb/h (kg/ h), can be calculated as
(3.47a)
where  volume flow rate of air leakage, cfm (m3 /min)
r  density of indoor air, lb/ ft3 (kg/m3 )
ws con  saturated humidity ratio corresponding to temperature of surface upon which condensation
occurs, lb / lb (kg /kg)
Vapor retarders are effective for reducing moisture transfer through the building envelope. They
not only decrease the moisture migration in the building material significantly, but also block air
leakage effectively if their joints are properly sealed. Vapor retarders are classified as rigid, flexible,
and coating types. Rigid type includes reinforced plastic, aluminum, and other metal sheets. Flexible
type includes metal foils, coated films, and plastic films. Coatings are mastics, paints, or fusible
sheets, composed of asphalt, resin, or polymeric layers. The vapor retarder is generally located on
the warm side of the insulation layer during winter heating.
In an area where summer cooling is dominant, a vapor retarder to prevent condensation in the
external brick wall and insulation layer would often be located on the outside of the insulation. The
absorptive brick wall takes substantial amounts of water during rainfall. The subsequent sunny
period drives the water vapor farther inside the wall and into the insulation layer where the water
vapor condenses.
3.6 THERMAL INSULATION
Thermal insulation materials, usually in the form of boards, slabs, blocks, films, or blankets, retard
the rate of heat transfer in conductive, convective, and radiant transfer modes. They are used within
building envelopes or applied over the surfaces of equipment, piping, or ductwork to achieve the
following benefits:
 Savings of energy by reducing heat loss and heat gain from the surroundings
V?
lk
m? con  60V?lk
r (wr  ws con )
m? con
3.18 CHAPTER THREE
 Prevention of surface condensation by increasing the surface temperature above the dew point of
the ambient air
 Reduction of temperature difference between the inside surface and the space air for the thermal
comfort of the occupants, when radiant heating or cooling is not desired
 Protection of the occupant from injury due to contact with hot pipes and equipment
Basic Materials and Thermal Properties
Basic materials in the manufacture of thermal insulation for building envelopes or air conditioning
systems include
 Fibrous materials such as glass fiber, mineral wool, wood, cane, or other vegetable fibers
 Cellular materials such as cellular glass, foam rubber, polystyrene, and polyurethane
 Metallic reflective membranes
Most insulating materials consist of numerous airspaces, either closed cells (i.e., airtight cellular
materials) or open cells (i.e., fibrous materials, penetrable by air).
Thermal conductivity k, an important physical property of insulating material, indicates the rate of
heat transfer by means of a combination of gas and solid conduction, radiation, and convection
through an insulating material, expressed in Btu in./h ft2 F or Btu/h ftF (W/mC). The thermal
conductivity of an insulating material depends on its physical structure (such as cell size and shape or
diameter of the fibrous materials), density, temperature, and type and amount of binders and additives.
Most of the currently used thermal insulating materials have thermal conductivities within a
range of 0.15 to 0.4 Btu in./h ft2 F (0.021 to 0.058 W/mC). Thermal properties of some building
and insulating materials, based on data published in the ASHRAE Handbook 1993, Fundamentals,
are listed in Table 3.3.
The thermal conductivity of many cellular insulating materials made from polymers like
polyurethane and extruded polystyrene is not significantly affected by change in density . Other
insulating materials have a density at which the thermal conductivity is minimum. Deviating from
this , conductivity k increases always, whether  is greater or smaller in magnitude. The thermal
conductivity of an insulating material apparently increases as its mean temperature rises. For
polystyrene, k increases from 0.14 to 0.32 Btu in./h ft2 F (0.020 to 0.046 W/mC) when its
mean temperature is raised from 300 to 570F (150 to 300C).
Moisture Content of Insulation Material
From the point of view of moisture transfer, penetrability is an important characteristic. Closed-cell
airtight board or block normally cannot have concealed condensation except at the gap between the
joints and at the interface of two layers. Only when air and its associated water vapor penetrate an
open-cell insulating material can concealed condensation form if they contact the surfaces of the
interstices and pores at a temperature lower than the dew point of the penetrating air. Concealed
condensation might also accumulate in open-cell insulating materials. If there is excess free moisture
in an insulating material, the thermal insulation may be degraded. The increase of moisture in
the thermal insulation layer is often due to the absorption from the ambient air, space air, or moisture
transfer from adjacent layer, ground, or wetted surface.
Smolenski (1996) brought forward the question of how much is too much. In roofing insulation,
many consider the answer to be the moisture content to produce a more than 20 percent loss in thermal
efficiency or a thermal resistance ratio (TRR) of less than 80 percent. The thermal resistance
ratio is defined as the ratio of wet to dry thermal resistance of the thermal insulation, in percent, or
(3.47b) TRR 
(wet thermal resistance)  100
dry thermal resistance
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.19
The equilibrium moisture content of most commomly used insulation material at 90 percent relative
humidity ambient air, as well as the moisture content of insulation material at 80 percent TRR by
weight (percent of dry weight) and by volume (percent of volume of insulating material) are listed
below:
Density,
Moisture content, %
Insulation material lb/ft3  90% 80% TRR, by weight 80% TRR, by volume
Cellular glass 8.4 0.2 23 3.1
Expanded polystyrene 1.0 2.0 383 6.1
Glass fiber 9.2 1.1 42 6.2
Urethane 2.1 6.0 262 8.8
Phenolic foam 2.6 23.4 25 1.0
For instance, for cellular glass when in contact with an ambient air of 90 percent relative humidity
at normal room temperaure that reaches an equilibrium state during moisture absorption, it has
3.20 CHAPTER THREE
TABLE 3.3 Thermal Properties of Selected Materials
Thermal Specific
Density, conductivity, heat,
lb / ft3 (Btu/h ft F) (Btu/ lb F) Emissivity
Aluminum (alloy 1100) 171 128 0.214 0.09
Asbestos: insulation 120 0.092 0.20 0.93
Asphalt 132 0.43 0.22
Brick, building 123 0.4 0.2 0.93
Brass (65% Cu, 35% Zn) 519 69 0.09 0.033 Highly polished
Concrete (stone) 144 0.54 0.156
Copper (electrolytic) 556 227 0.092 0.072 Shiny
Glass: crown (soda-lime) 154 0.59 0.18 0.94 Smooth
Glass wool 3.25 0.022 0.157
Gypsum 78 0.25 0.259 0.903 Smooth plate
Ice (32F) 57.5 1.3 0.487 0.95
Iron: cast 450 27.6 0.12 0.435 Freshly turned
Mineral fiberboard:
acoustic tile, wet-molded 23 0.035 0.14
wet-felted 21 0.031 0.19
Paper 58 0.075 0.32 0.92
Polystyrene, expanded, molded beads 1.25 0.021 0.29
Polyurethane, cellular 1.5 0.013 0.38
Plaster, cement and sand 132 0.43 0.91 Rough
Platinum 1340 39.9 0.032 0.054 Polished
Rubber: vulcanized, soft 68.6 0.08 0.48 0.86 Rough
hard 74.3 0.092 0.95 Glossy
Sand 94.6 0.19 0.191
Steel (mild) 489 26.2 0.12
Tin 455 37.5 0.056 0.06 Bright
Wood: fir, white 27 0.068 0.33
oak, white 47 0.102 0.57 0.90 Planed
plywood, Douglas fir 34 0.07 0.29
Wool: fabric 20.6 0.037
Source: Adapted with permission from ASHRAE Handbook 1989, Fundamentals.
only a moisture content of 0.2 percent by dry weight, which is far lower than the moisture content
of 23 percent by weight at 80 percent TRR. Most above-listed insulating materials have a far lower
equilibrium moisture content of dry weight at an ambient air of 90 percent relative humidity than
moisture content at 80 percent TRR. Also, a closed-cell structure cellular glass requires months or
years to reach an equilibrium moisture content by volume corrresponding to 80 percent TRR
whereas mineral wool and calcium silica absorb moisture in only hours.
Economic Thickness
The economic thickness of insulation is the thickness with the minimum owning and operating
costs. Owning cost is the net investment cost of the installed insulation Cin, in dollars, less any capital
investment that can be made as a result of lower heat loss or gain Cpt, in dollars. Theoretically,
for a new plant, some small savings might be made because of the reduction of the size of a central
plant; but in actual practice, this is seldom considered. For an existing plant where add-on insulation
is being considered, Cpt is zero because the plant investment has already been made.
Operating cost Cen includes the annualized cost of energy over the life of a new plant, or the
remaining life of an existing plant, in n years. It can also be taken as the number of years over
which an owner wishes to have a total return of the net investment, considering both the interest and
the fuel escalation rate. Total cost of insulation Ct, in dollars, for any given thickness is
(3.48)
When the thickness of the insulating material increases, the quantity CinCpt also increases, as
shown in Fig. 3.5, and Cen decreases. As a result, Ct first decreases, drops to a minimum, and then
increases. The optimum economic thickness occurs when Ct drops to a minimum. The closest commercially
available thickness is the optimum thickness.
Thermal Resistance of Airspaces
Thermal resistance of an enclosed airspace Ra has a significant effect on the total thermal resistance
RT of the building envelope, especially when the value of RT is low. Thermal resistance Ra depends
on the characteristic of the surface (reflective or nonreflective), the mean temperature, the temperature
difference of the surfaces perpendicular to heat flow, the width across the airspace along the heat flow,
Ct  Cin  Cpt 	 Cen
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.21
FIGURE 3.5 Optimum thickness of insulation material.
and the direction of airflow. The R values of the enclosed airspaces Ra, h ft2 F/Btu (m2 C/W),
abridged from data published in ASHRAE Handbook 1989, Fundamentals, are presented in Table 3.4.
3.7 SOLAR ANGLES
Basic Solar Angles
The basic solar angles between the suns rays and a specific surface under consideration are shown
in Figs. 3.6 and 3.7.
3.22 CHAPTER THREE
FIGURE 3.6 Basic solar angles and position of suns rays at summer solstice.
TABLE 3.4 R Values of Enclosed Airspace Ra, h ft2 F/Btu
Mean Temperature Type of Direction Width of
Emissivity E
temperature, F difference, F surface of heat flow airspace, in. 0.05 0.2 0.82
Summer, 90 10 Horizontal Upward 0.5 2.03 1.51 0.73
3.5 2.66 1.83 0.80
Vertical Horizontal 0.5 2.34 1.67 0.77
3.5 3.40 2.15 0.85
Horizontal Downward 0.5 2.34 1.67 0.77
3.5 8.19 3.41 1.00
Winter, 50 10 Horizontal Upward 0.5 2.05 1.60 0.84
3.5 2.66 1.95 0.93
Vertical Horizontal 0.5 2.54 1.88 0.91
3.5 3.40 2.32 1.01
Horizontal Downward 0.5 2.55 1.89 0.92
3.5 9.27 4.09 1.24
Emissity E  1/(1/e1 	 1/e2  1), where e1, e2 indicate the emittances on two sides of the airspace.
Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.
3.23
FIGURE 3.7 Solar intensity and angle of incidence.
 Solar altitude angle (Fig. 3.7a and b) is the angle ROQ on a vertical plane between the suns ray
OR and its projection on a horizontal plane on the surface of the earth.
 Solar azimuth  (Fig. 3.7a) is the angle SOQ on a horizontal plane between the due-south direction
line OS and the horizontal projection of the suns ray OQ.
 Solar declination angle  (Fig. 3.6) is the angle between the earth-sun line and the equatorial
plane. Solar declination changes with the times of the year. It is shown in Fig. 3.6 on June 21.
 Surface-solar azimuth  (Fig. 3.7a and c) is the angle POQ on a horizontal plane between the
normal to a vertical surface OP and the horizontal projection of the suns rays OQ.
 Surface azimuth  (Fig. 3.7a) is the angle POS on a horizontal plane between OP and the direction
line SN.
 Latitude angle L (Fig. 3.6) is the angle SOO on the longitudinal plane between the equatorial
plane and the line OO that connects the point of incidence of the suns ray on the surface of earth
O and the center of the earth O.
Hour Angle and Apparent Solar Time
Hour angle H (Fig. 3.6) is the angle SOQ on a horizontal plane between the line OS indicating the
noon of local solar time tls and the horizontal projection of the suns ray OQ. The values of the hour
angle H before noon are taken to be positive. At 12 noon, H is equal to 0. After 12 noon, H is negative.
Hour angle H, in degrees, can be calculated as
H  0.25 (time in minutes from local solar noon) (3.49)
The relationship between apparent solar time tas, as determined by a sundial and expressed in apparent
solar time, and local standard time tst, both in minutes, is as follows:
(3.50)
where M  local standard time meridian and G  local longitude, both in degrees. In Eq. (3.50), teq,
in minutes, indicates the difference in time between the mean time indicated by a clock running at a
uniform rate and the solar time due to the variation of the earths orbital velocity throughout the year.
Solar Angle Relationships
The relationship among the solar angles is given as
sin  sin  sinL 	 cos  cosH cosL (3.51)
and
(3.52)
Angle of Incidence and Solar Intensity
The angle of incidence  (Fig. 3.7a) is the angle between the suns rays radiating on a surface and
the line normal to this surface. For a horizontal surface, the angle of incidence H is ROV; for a vertical
surface, the angle of incidence V is ROP; and for a tilted surface, the angle of incidence  is
ROU. Here,  is the angle between the tilted surface ABCD and the horizontal surface.
Let IDN be the intensity of direct normal radiation, or solar intensity, on a surface normal to the
suns ray, in Btu/h ft2 (W/m2). In Fig. 3.7b, IDN is resolved into the vertical component RQ  IDN
cos  
sin
 sin L  sin 
cos  cos L
tas  teq 	 tst  4(M  G)
3.24 CHAPTER THREE
sin  and the horizontal component OQ  IDN cos . In Fig. 3.7c, for the right triangle OPQ, angle
POQ  , and hence OP  IDN cos  cos . In Fig. 3.7d, OPL is a right triangle. From point P, a
line PT can be drawn perpendicular to the line normal OU, and hence, two right triangles are
formed: OTP and LTP. In OTP, because angle OPT  , the horizontal component of IDN along the
line normal of the tilting surface OU is OT  IDN cos  cos  sin. In Fig. 3.7e, draw line OU
parallel to line OU. Again, line QM can be drawn from point Q perpendicular to OU. Then the
right triangles PTL and QMR are similar. Angles PLT and QRM are both equal to . In the right
triangle QMR, the component of RQ that is parallel to the line normal to the tilting surface is MR 
IDN sin cos.
The intensity of solar rays normal to a tilted surface I, Btu/h ft2 (W/m2) is the vector sum of
the components of the line normal to the tilted surface, or
(3.53)
3.8 SOLAR RADIATION
Solar radiation provides most of the energy required for the earths occupants, either directly or
indirectly. It is the source of indoor daylight and helps to maintain a suitable indoor temperature
during the cold seasons. At the same time, its influence on the indoor environment must be reduced
and controlled during hot weather. The sun is located at a mean distance of 92,900,000 mi,
(149,500,000 km) from the earth, and it has a surface temperature of about 10,800F (6000C). It
emits electromagnetic waves at wavelengths of 0.29 to 3.5 	m (micrometers). Visible light has
wavelengths of 0.4 to 0.7 	m and is responsible for 38 percent of the total energy striking the earth.
The infrared region contains 53 percent. At the outer edge of the atmosphere at a mean earth-sun
distance, the solar intensity, called the solar constant Isc, is 434.6 Btu/h ft2 (1371 W/m2 ). The
extraterrestrial intensity Io, Btu/h ft2 (W/m2) varies as the earth-sun distance changes during the
earths orbit.
Based on the data from Miller et al. (1983), the breakdown of solar radiation reaching the
earths surface and absorbed by the earth is listed in Table 3.5. As listed in Table 3.5, only 50 percent
of the solar radiation that reaches the outer edge of the earths atmosphere is absorbed by the
clouds and the earths surface. At any specific location, the absorption, reflection, and scattering of
solar radiation depend on the composition of the atmosphere and the path length of the suns rays
through the atmosphere, expressed in terms of the air mass m. When the sun is directly overhead,
m  1.
 IDN (cos
 cos  sin  	 sin  cos 
I  IDN cos 
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.25
TABLE 3.5 Components of Solar Radiation That Traverse the Earths
Atmosphere
Components Breakdowns
Scattered by air 11% Reflected to space 6%
Scattered to earth 5
Absorbed by water vapor,
dust, etc. 16
Intercepted by clouds 45 Reflected to space 20
Absorbed by clouds 4
Diffused through clouds
and absorbed by earth 21
Traversed through air 28 Absorbed by earth 24
Reflected by earth 4
In Table 3.5, the part of solar radiation that gets through the atmosphere and reaches the earths
surface, in a direction that varies with solar angles over time, is called direct radiation. The part that
is diffused by air molecules and dust, arriving at the earths surface in all directions, is called diffuse
radiation. The magnitude of solar radiation depends on the composition of the atmosphere, especially
on the cloudiness of the sky. Therefore, different models are used to calculate the solar radiation
reaching the surface of building envelopes.
Solar Radiation for a Clear Sky
ASHRAE recommends use of the following relationships to calculate the solar radiation for a clear
sky. The solar intensity of direct normal radiation IDN, Btu/h ft2 (W/m2 ), can be calculated as
(3.54)
where A  apparent solar radiation when air mass m  0 (its magnitudes are listed in Table 3.6),
Btu/hft (W/m2 )
B  atmospheric extinction coefficient (Table 3.6), which depends mainly on the amount of
water vapor contained in the atmosphere
In Eq. (3.54), the term 1/sin  in exponential form denotes the length of the direct radiation path
through the atmosphere. The term Cn is the clearness number of the sky; Cn takes into account the
dryness of the atmosphere and the dust contained in the air at a geographic location. Estimated Cn
values for nonindustrial locations in the United States are shown in Fig. 3.8.
From Eq. (3.53), the direct radiation radiated onto a horizontal surface through a clear sky IDH,
Btu/h ft2 (W/m2), is the vertical component of IDN, that is,
(3.55)
and the direct radiation irradiated onto a vertical surface for clear sky IDV, Btu/h ft2 (W/m2 ), is the
horizontal component of IDN, or
(3.56) IDV  IDN cos V  IDN cos  cos 
IDH  IDN cos H  IDN sin 
IDN 
ACn
exp (B/sin )
3.26 CHAPTER THREE
TABLE 3.6 Extraterrestrial Solar Radiation and Related Data for 21st Day of Each Month,
Base Year 1964
A B C
Equation of Declination, (Dimensionless
I0, Btu/h ft2 time, min deg Btu /h ft2 ratios)
January 448.8 11.2 20.0 390 0.142 0.058
Febuary 444.2 13.9 10.8 385 0.144 0.060
March 437.7 7.5 0.0 376 0.156 0.071
April 429.9 	1.1 	11.6 360 0.180 0.097
May 423.6 	3.3 	20.0 350 0.196 0.121
June 420.2 1.4 	23.45 345 0.205 0.134
July 420.3 6.2 	20.6 344 0.207 0.136
August 424.1 2.4 	12.3 351 0.201 0.122
September 430.7 	7.5 0.0 365 0.177 0.092
October 437.3 	15.4 10.5 378 0.160 0.073
November 445.3 	13.8 19.8 387 0.149 0.063
December 449.1 	1.6 23.45 391 0.142 0.057
A: apparent solar radiation; B: atmospheric extinction coefficient; C: diffuse radiation factor.
Source: ASHRAE Handbook 1997, Fundamentals. Reprinted with permission.
From Eq. (3.53), for the direct radiation radiated onto a tilted surface at an angle  with the horizontal
plane through clear sky ID, Btu/h ft2 (W/m2 ), can be evaluated as
(3.57)
Most of the ultraviolet solar radiation is absorbed by the ozone layer in the upper atmosphere. Direct
solar radiation, through an air mass m  2, arrives on the earths surface at sea level during a
clear day with a spectrum of  3 percent in the ultraviolet, 44 percent in the visible, and 53 percent
in the infrared.
The diffuse radiation Id, Btu/h ft2 (W/m2 ), is proportional to IDN on cloudless days and can be
approximately calculated as
(3.58)
where C  diffuse radiation factor, as listed in Table 3.6, and Cn  clearness number of sky from
Fig. 3.8. In Eq. (3.58), Fss indicates the shape factor between the surface and the sky, or the fraction
of shortwave radiation transmitted through the sky that reaches the surface. For a vertical surface
Fss  0.5, for a horizontal surface Fss  1, and for any tilted surface with an angle 
(3.59)
The total or global radiation on a horizontal plane IG, Btu/h ft2 (W/m2), recorded by the U.S.
Fss 
1.0 	 cos 
2
Id 
CIDN Fss
Cn 2
ID  I  IDN cos   IDN (cos
 cos
 sin  	 sin
 cos)
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.27
FIGURE 3.8 Estimated atmospheric clearness numbers in the United States for nonindustrial localities.
National Climatic Data Center (NCDC), can be calculated as
(3.60)
The reflection of solar radiation from any surface Iref, Btu/h ft2 (W/m2 ), is given as
(3.61)
where s  reflectance of the surface and Fsr  shape factor between the receiving surface and the
reflecting surface. The ground-reflected diffuse radiation falling on any surface Isg, Btu/h ft2
(W/m2), can be expressed as
Isg  g Fsg IG (3.62)
where Fsg  shape factor between the surface and the ground and g  reflectance of the ground.
For concrete, g  0.23, and for bitumen and gravel, g  0.14. A mean reflectance g  0.2 is
usually used for ground.
The total intensity of solar radiation It, Btu/h ft2 (W/m2 ), falling on a surface at a direction
normal to the surface on clear days, is given by
(3.63)
where I  component of reflected solar intensity in direction normal to surface, Btu /h ft2 (W/m2 )
Iref.DN  component of reflected solar intensity in direction of sun ray, Btu/h ft2 (W/m2 )
Solar Radiation for a Cloudy Sky
For cloudy skies, the global horizontal irradiation IG*, Btu/h ft2 (W/m2), usually can be obtained
from the NCDC. If it is not available, then it can be predicted from the following relationship:
(3.64a)
Here Ccc indicates the cloud cover, on a scale of 0 to 10, and can be calculated by
(3.64b)
where CT  total cloud amount
Ccir. j  clouds covered by cirriforms, including cirrostratus, cirrocumulus, and cirrus, in j  1
to 4 layers
Both CT and Ccir. j values can be obtained from the major weather stations. The values of coefficients
P, Q, and R, according to Kimura and Stephenson (1969), are listed below:
P Q R
Spring 1.06 0.012  0.0084
Summer 0.96 0.033  0.0106
Autumn 0.95 0.030  0.0108
Winter 1.14 0.003  0.0082
Ccc  CT  0.5 4
j1
Ccir. j
I*G  1 	
Ccc Q
P
	
Ccc
2 R
P IG
 (IDN 	 Iref.DN) cos 
	 Id
It  ID 	 Id 	 Iref. S
Iref  sFsr (ID 	 Id)
IG  ID 	 Id  IDN sin 
	
C
Cn 2
3.28 CHAPTER THREE
The direct radiation for a cloudy sky ID*, Btu/h ft2 (W/m2), can be calculated as
(3.65)
The diffuse radiation for a cloudy sky Id*, Btu/h ft2 (W/m2 ), is calculated as
Id*  IG*  ID* (3.66)
3.9 FENESTRATION
Fenestration is the term used for assemblies containing glass or light-transmitting plastic, including
appurtenances such as framing, mullions, dividers, and internal, external, and between-glass shading
devices, as shown in Fig. 3.12a. The purposes of fenestrations are to (1) provide a view of the
outside world, (2) permit entry of daylight, (3) admit solar heat as a heating supplement in winter,
(4) act as an emergency exit for single-story buildings, and (5) add to aesthetics.
Solar radiation admitted through a glass or window pane can be an important heat gain for
commercial buildings, with greater energy impact in the sun belt. HVAC&R designers are asked to
control this solar load while providing the required visibility, daylight, and winter heating benefits
as well as fire protection and safety features.
Types of Window Glass (Glazing)
Most window glasses, or glazing, are vitreous silicate consisting of silicon dioxide, sodium oxide,
calcium oxide, and sodium carbonate. They can be classified as follows:
 Clear plate or sheet glass or plastic. Clear plate glass permits good visibility and transmits more
solar radiation than other types.
 Tinted heat-absorbing glass. Tinted heat-absorbing glass is fabricated by adding small amounts of
selenium, nickel, iron, or tin oxides. These produce colors from pink to green, including gray or
bluish green, all of which absorb infrared solar heat and release a portion of this to the outside
atmosphere through outer surface convection and radiation. Heat-absorbing glass also reduces
visible light transmission.
 Insulating glass. Insulating glass consists of two panesan outer plate and a inner plateor
three panes separated by metal, foam, or rubber spacers around the edges and hermetically sealed
in a stainless-steel or aluminum-alloy structure. The dehydrated space between the glass panes
usually has a thickness of 0.125 to 0.75 in. (3.219 mm) and is filled with air, argon, or other
inert gas. Air- or gas-filled space increases the thermal resistance of the fenestration.
 Reflective coated glass. Reflective glass has a microscopically thin layer of metallic or ceramic
coating on one surface of the glass, usually the inner surface of a single-pane glazing or the outer
surface of the inner plate for an insulating glass. For a single pane, the coating is often protected
by a layer of transparent polyester. The chromium and other metallic coatings give excellent re-
flectivity in the infrared regions but reduced transmission of visible light compared to clear plate
and heat-absorbing glass. Reflections from buildings with highly reflective glass may blind drivers,
or even kill grass in neighboring yards.
 Low-emissivity (low-E) glass coatings. Glazing coated with low-emissivity, or low-E, films has
been in use since 1978. It is widely used in retrofit applications. A low-emissivity film is usually
a vacuum-deposited metallic coating, usually aluminum, on a polyester film, at a thickness of
about 4  107 in. (0.01 	m). Because of the fragility of the metal coating, protection by another
polyester film against abrasion and chemical corrosion must be provided. Recently, copper and
I*D 
I*G sin  (1  Ccc /10)
sin 
	 C/Cn 2
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.29
silver coatings on polyethylene and polypropylene film for protection have been used for better
optical transmission. A low-E film coating reduces the U value about 25 to 30 percent for single
panes. When combined with other solar control devices, these films can reduce solar heat gain
further.
Optical Properties of Sunlit Glazing
When solar radiation inpinges on the outer surface of a plate of glass with an intensity of I at an
angle of incidence , as shown in Fig. 3.9, a portion of the solar radiation is transmitted, another
portion is reflected from both inner and outer surfaces, and the remaining portion is absorbed.
Let r indicate the portion reflected and a the portion absorbed. Also let  be the angle of incidence.
In Fig. 3.9, it can be seen that the portion of the solar radiation transmitted through the glass
is actually the sum of the transmittals after successive multiple reflections from the outer and inner
surfaces. The decimal portion of I transmitted through the glass, or , represents the transmittance
of the window glass. Similarly, the portion of solar radiation reflected from the window glass is the
sum of the successive reflections from the outer surfaces after multiple reflections and absorptions,
3.30 CHAPTER THREE
FIGURE 3.9 Simplified representation of multiple transmissions, reflections, and absorptions of solar radiation
at glass surfaces.
and it is identified as , the reflectance. The portion absorbed is the sum of the successive absorptions
within the glass , or its absorptance.
Figure 3.10 illustrates the spectral transmittance of several types of glazing. All are transparent
for shortwave solar radiation at a wavelength between 0.29 and 3	m and are opaque to longwave
radiation in the infrared range with a wavelength greater than 3	m. Most interior furnishings,
equipment, and appliances have an outer surface temperature lower than 250F (120C), emitting
almost all longwave radiation. At such temperatures, glass is opaque to longwave radiation emitted
from inside surfaces and lets only shortwave radiation through. This trapping of longwave radiation
indoors is called the greenhouse effect.
In Fig. 3.10 one can also see that clear plate glass has a high transmittance  0.87 for visible
light, and that  0.8 for infrared from 0.7 to 3	m. Heat-absorbing glass has a lower  and higher
absorptance  for both visible light and infrared radiation. Bluish-green heat-absorbing glass has a
higher  in the visible light range and a lower  in the infrared range than gray heat-absorbing
glass. Some reflective glazing has a high reflectance and a significantly lower  in the visible light
range and is opaque to radiation at wavelengths greater than 2	m. Such characteristics for heatabsorbing
and reflective glasses are effective for reducing the amount of solar heat entering the
conditioned space during cooling as well as heating seasons. This fact sometimes presents a
dilemma for the designer, who must finally compromise to get the optimum combination of con-
flicting properties.
Another important property of glass is that both  and  decrease and  increases sharply as the
incident angle  increases from 60 to 90. At 90,  and  are 0 and  is equal to 1. That is why the
solar radiation transmitted through vertical glass declines sharply at noon during summer with solar
altitudes   70.
For all types of plate glass, the sum of these radiation components is always equal to 1, that is,
	 	  1
and
I  I 	 I 	 I (3.67)
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.31
FIGURE 3.10 Spectral transmittance for various types of window glasses.
(Source: ASHRAE Handbook 1989, Fundamentals. Reprinted by permission.)
3.10 HEAT ADMITTED THROUGH WINDOWS
For external glazing without shading, the heat gain admitted into the conditioned space through
each square foot of sunlit area As of window Qwi /As, Btu/h ft2 (W/m
2), can be calculated as follows:
Heat gain through each ft2

solar radiation transmitted
	
inward heat flow from glass
of sunlit window through window glass inner surface into conditioned space
That is,
(3.68)
where QRCi  inward heat flow from the inner surface of an unshaded sunlit window, Btu/h (W).
Heat Gain for Single Glazing
For an external, sunlit single-glazed window without shading, the inward heat flow from the inner
surface of the glass, as shown in Fig. 3.11, can be evaluated as
QRCi  inward absorbed radiation 	 conductive heat transfer
(3.69)  UAs It
ho
	 To  Ti
Qwi
As

It 	 QRCi
As
3.32 CHAPTER THREE
FIGURE 3.11 Heat admitted through a single-glazing window
glass.
where ho  heat-transfer coefficient for outdoor surface of window glass, Btu/h ft2 F (W/m2 C)
To  outdoor air temperature, F (C)
Ti  indoor air temperature,F (C)
Heat admitted through a unit area of the single-glazing window glass is
(3.70)
Solar heat gain coefficient (SHGC) is the ratio of solar heat gain entering the space through the
window glass to the incident solar radiation, total shortwave irradiance for a single-glazed window
is given as
(3.71)
In these last three equations, U indicates the overall heat-transfer coefficient of the window, in
Btu/h ft2 F (W/m2 C), and can be calculated as
(3.72)
In Eq. (3.72), A represents area, in square feet (square meters); subscript wg indicates glass, and eg
signifies the edge of the glass including the sealer and spacer of the insulating glass. The edge of
glass has a width of about 2.5 in. (64 mm). The subscript f means the frame of the window. Some U
values for various types of windows at winter design conditions are listed in Table 3.7. For summer
U 
Uwg Awg 	 Ueg Aeg 	 Uf Af
Awg 	 Aeg 	 Af


	
U
ho
SHGC  Qws / It As
 SHGCIt 	 U(To  Ti)
Qwi
As

It 	 U 
It
ho
	 To  Ti
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.33
TABLE 3.7 Overall Heat-Transfer Coefficient U Values for Windows at Winter Conditions* with Commercial Type of Frame,
Btu/h ft2 F)
Aluminum
frame with Wood
Gas Space Emittance Aluminum thermal or vinyl
between between of low-E frame of break, with frame of
Type glasses glasses, in. film Glass Edge Uf  1.9 Uf  1.0 Uf  0.41
Overall coefficient Btu/h ft2 F
Single glass 1.11 1.23 1.10 0.98
Double glass Air 3/8 0.52 0.62 0.74 0.60 0.51
Double glass Air 3/8 0.40 0.43 0.55 0.67 0.54 0.45
Double glass Air 3/8 0.15 0.36 0.51 0.62 0.48 0.39
Double glass Argon 3/8 0.15 0.30 0.48 0.57 0.43 0.34
Triple glass Air 3/8 0.34 0.50 0.60 0.46 0.38
Triple glass Air 3/8 0.40 0.30 0.48 0.57 0.44 0.35
Triple or double
glass with polyester
film suspended
in between Argon 3/8 0.15 0.17 0.43 0.47 0.34 0.25
*Winter conditions means 70F indoor, 0F outdoor temperature and a wind speed of 15 mph.
Low-E film can be applied to surface 2 for double glass (see Fig. 3.12b) or surface 2, 3, 4, or 5 for triple glass (any surface other than outer and
inner surfaces).
Source: Abridged with permission from ASHRAE Handbook, 1989, Fundamentals.
design conditions at 7.5 mph (3.3 m/s) wind speed, the listed U values should be multiplied by a
factor of 0.92. The U values of windows depend on the construction of the window, the emissivity
of the surfaces of glass or plastic sheets, and the air velocity flowing over the outdoor and indoor
surfaces. Qws indicates the solar heat gain entering the space, in Btu/h (W).
As shown in Fig. 3.11, solar radiation having a total intensity I, Btu/h ft2 (W/m2), when it is radiated
on the outer surface of a vertical pane with an angle of incidence v, the line normal indicating
total shortwave irradiance It, in Btu/h ft2 (W/m2) actually consists of
(3.73)
Heat Gain for Double Glazing
For a double-glazed window, the inward heat flow per square foot of the inner surface of the glass,
as shown in Fig. 3.12b, is
(3.74)  U 
Io
ho
	  1
ho
	
1
ha  Ii 	 To  Ti 	
QRCi
As
 Nio Io 	 Nii Ii 	 U(To  Ti)
It  I cos v
I  IDN 	 Id 	 Iref. DN
3.34 CHAPTER THREE
FIGURE 3.12 Heat flow through an insulating glass (double-pane) window. (a) Construction of a typical insulating glass; (b) heat flow
and temperature profiles.
where Nio  inward fraction of solar radiation absorbed by outdoor glass
Nii  inward fraction of solar radiation absorbed by indoor glass
Io, Ii  solar intensity irradiated on outdoor and indoor glass, Btu/h ft2 (W/m2)
The solar radiation absorbed by the outdoor and indoor panes, in Btu/h ft2 (W/m2), is
Io  o It Ii  i It (3.75)
Absorptance of the outdoor panes o is
(3.76)
Absorptance of the indoor pane is
(3.77)
Subscript o indicates outdoor; i, indoor; and 1, 2, 3, and 4, the surfaces of the panes as shown in
Fig. 3.12b.
The airspace heat-transfer coefficient ha, Btu/h ft2 F (W/m2 C), is the reciprocal of the R
value of the airspace Ra listed in Table 3.4, that is,
(3.78)
The transmittance of solar radiation through both outdoor and indoor panes oi can be calculated
as
(3.79)
Then the heat admitted per square foot (square meter) through a double-glass window Qwoi /As
Btu/h ft2 (W/m2), can be calculated as
(3.80)
The SHGCoi for a double-glass window can be calculated as
(3.81)
Because a glass plate is usually between 0.125 and 0.25 in. (3 and 6 mm) thick, with a thermal
conductivity k of about 0.5 Btu/h ft F (W/mC), there is only a small temperature difference
between the inner and outer surfaces of the plate when solar radiation is absorbed. For the sake
of simplicity, it is assumed that the temperature of the plate Tg is the same in the direction of heat
flow.
For a double-glazed window, the glass temperature of the outdoor pane Tgo, F (C), can be
calculated as
(3.82) Tgo  To 	 Io 	
Ii 
QRCi
As  1
ho
	
Rgo
2 
SHGCoi 
oi 	 U o
ho
	
i
ha
 SHGCoi It 	 U(To  Ti)
Qwoi
As

oi It 	 Uo
h

	
i
h

	
i
ha 	 It 	 U(To  Ti)
oi 
o i
1 
2 3
ha 
1
Ra
i 
3 o
1 
2 3
o 
1 	
2 oi 3
1 
2 3
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.35
The temperature of the indoor pane Tgi, F (C), is calculated as
(3.83)
where Rgo, Rgi  R values of outdoor and indoor panes, h ft2 F/Btu (m2 C/W)
hi  heat-transfer coefficient of inside surface 4 (see Fig. 3.12b) of inside plate,
Btu/h ft2 F (W/m2 C)
Shading Coefficients
The shading coefficient is defined as the ratio of solar heat gain of a glazing assembly of specific
construction and shading devices at a summer design solar intensity and outdoor and indoor
temperatures, to the solar heat gain of a reference glass at the same solar intensity and outdoor and
indoor temperatures. The reference glass is double-strength sheet glass (DSA) with transmittance
  0.86, reflectance  0.08, absorptance  0.06, and FDSA  0.87 under summer design
conditions. The shading coefficient SC is an indication of the characteristics of a glazing and the
associated shading devices, and it can be expressed as
(3.84)
where SHGCw  solar heat gain coefficient of specific type of window glass
SHGCDSA  solar heat gain coefficient of standard reference double-strength sheet glass
Shading coefficients of various types of glazing and shading devices are presented in Table 3.8.

SHGCw
SHGCDSA

SHGCw
0.87
 1.15 SHGCwi
SC 
solar heat gain of specific type of window glass
solar heat gain of double-strength sheet glass
Tgi  Tr 	 QRCi  1
hi
	
Rgi
2 
3.36 CHAPTER THREE
TABLE 3.8 Shading Coefficients for Window Glass with Indoor Shading Devices
Venetian blinds Roller shade Draperies
Thickness Solar
Type of glass of glass, in. transmittance Glass Med.* Light Opaque, white Translucent Med. Light
Clear 0.87 to 0.79 0.74 0.67 0.39 0.44 0.62 0.52
Heat-absorbing or 0.46 0.57 0.53 0.30 0.36 0.46 0.44
0.34 0.54 0.52 0.28 0.32
0.24 0.42 0.40 0.28 0.31 0.38 0.36
Reflective-coated 0.30 0.25 0.23
0.40 0.33 0.29
0.50 0.42 0.38
Insulating glass Outer Inner
Clear out , or 0.87 0.87 0.62 0.58 0.35 0.40
Clear in
Heat-absorbing out 0.46 0.80 0.39 0.36 0.22 0.30
Clear in
Reflective glass 0.20 0.19 0.18
0.30 0.27 0.26
0.40 0.34 0.33
*Med. indicates medium color.
 Light indicates light color.
Draperies Med. represents draperies of medium color with a fabric openness of 0.10 to 0.25 and yarn reflectance of 0.25 to 0.50.
 Draperies Light repesents draperies of light color with a fabric openness below 0.10 and yarn reflectance over 0.50.
Source: Adapated with permission from ASHRAE Handbook 1989, Fundamentals.
1
4
1
8
3
32
3
8
1
4
3
16
3
32
Solar Heat Gain Factors and Total Shortwave Irradiance
The solar heat gain factor (SHGF), Btu/h ft2 (W/m2), is designated as the average solar heat gain
during cloudless days through DSA glass. In the ASHRAE Handbook 1993, Fundamentals are tabulated
SHGF value for various latitudes, solar times, and orientations for load and energy calculations.
For calculating the summer cooling peak load, the concept of maximum SHGF has been
introduced. This is the maximum value of SHGF on the 21st of each month for a specific latitude,
as listed in Table 3.9. For high elevations and on very clear days, the actual SHGF may be 15
percent higher than the value, listed in Table 3.9. In dusty industrial areas or at very humid locations,
the actual SHGF may be lower.
According to ASHRAE Handbook 1997, Fundamentals, Gueymard (1995) provides a comprehensive
model for calculating the spectral and broadband total shortwave irradiance It, in Btu/h ft2
(W/m2) for cloudless sky conditions and allows the input of the concentrations of a variety atmospheric
constituents.
Example 6.1. A double-glass window of a commercial building facing west consists of an outdoor
clear plate glass of 0.125-in. and an indoor reflective glass of 0.25-in. thickness with a reflective
film on the outer surface of the indoor glass, as shown in Fig. 3.12b. This building is located at
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.37
TABLE 3.9 Maximum Solar Heat Gain Factors (Max SHGF)
Max SHGF, Btu /h ft2
N
(shade) NNE/NNW NE/NW ENE/WNW E/W ESE/WSW SE/SW SSE/SSW S HOR*
North latitude, 40
Jan. 20 20 20 74 154 205 241 252 254 133
Feb. 24 24 50 129 186 234 246 244 241 180
Mar. 29 29 93 169 218 238 236 216 206 223
Apr. 34 71 140 190 224 223 203 170 154 252
May 37 102 165 202 220 208 175 133 113 265
June 48 113 172 205 216 199 161 116 95 267
July 38 102 163 198 216 203 170 129 109 262
Aug. 35 71 135 185 216 214 196 165 149 247
Sept. 30 30 87 160 203 227 226 209 200 215
Oct. 25 25 49 123 180 225 238 236 234 177
Nov. 20 20 20 73 151 201 237 248 250 132
Dec. 18 18 18 60 135 188 232 249 253 113
North latitude, 32
Jan. 24 24 29 105 175 229 249 250 246 176
Feb. 27 27 65 149 205 242 248 232 221 217
Mar. 32 37 107 183 227 237 227 195 176 252
Apr. 36 80 146 200 227 219 187 141 115 271
May 38 111 170 208 220 199 155 99 74 277
June 44 122 176 208 214 189 139 83 60 276
July 40 111 167 204 215 194 150 96 72 273
Aug. 37 79 141 195 219 210 181 136 111 265
Sept. 33 35 103 173 215 227 218 189 171 244
Oct. 28 28 63 143 195 234 239 225 215 213
Nov. 24 24 29 103 173 225 245 246 243 175
Dec. 22 22 22 84 162 218 246 252 252 158
*Horizontal surface
Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.
40 north latitude. The detailed optical properties of their surfaces are as follows:
The R value of the 0.25-in. (6-mm) thickness indoor glass is Rg  0.035 h ft2 F/Btu (0.0063
m2 C/W), and that of the enclosed airspace is Ra  1.75 h ft2 F/Btu (0.32 m2 C/W).
At 4 P.M. on July 21, the outdoor temperature at this location is 100F (37.8C), the indoor temperature
is 76F (24.4C), and the total solar intensity at a direction normal to this west window is
248 Btu/h ft2 (782 W/m2). The outdoor surface heat-transfer coefficient ho  4.44 Btu/h ft2 F
(25.2 W/m2C), and the heat-transfer coefficient of the inner surface of the indoor glass hi  2.21
Btu/h ft2 F (12.5 W/m2 C). Calculate the following:
1. The inward heat flow of this window
2. The temperatures of the outdoor and indoor glasses
3. The shading coefficient of this double-glass window
4. The total heat gain admitted through this window
Solution
1. From Eq. (3.76), the absorption coefficient of the outdoor glass can be calculated as
And from Eq. (3.77), the absorption coefficient of the indoor glass is
From Eq. (3.75), the heat absorbed by the outdoor glass is
Also, the heat absorbed by the indoor glass is calculated as
And then, from Eq. (3.22), the overall heat-transfer coefficient of this double-glass window is

1
1/2.21 	 0.035 	 1.75 	 0.035 	 1/4.44
 0.400 Btu / h ft2F
U 
1
1/hi 	 Rg 	 Ra 	 Rg 	 1/ho
Ii  aiIt  0.135  248  33.5 Btu / h ft2
Io 
o It  0.189  248  46.9 Btu / h ft2
i 
3 o
1 
2 3

0.16  0.8
1  0.08  0.68
 0.135
 0.12 	
0.12  0.8  0.68
1  0.08  0.68
 0.189
o 
1 	
2 o 3
1 
2 3
 0.80 1  0.08 1  0.12 e1  0.84
2  0.08 2  0.12 e2  0.84
i  0.16 3  0.68 3  0.16 e3  0.15
4  0.08 4  0.76 e4  0.84
3.38 CHAPTER THREE
The inward heat flow from the inner surface of the indoor glass, from Eq. (3.74), is given as
2. From Eq. (3.82), the temperature of the outdoor glass is
and the temperature of the indoor glass is
3. From Eq. (3.79), the transmittance for both panes is
and from Eq. (3.81), the solar heat gain coefficient for the double-glass window is
Then, from Eq. (3.84), the shading coefficient of this double-glass window is given by
4. From Eq. (3.80), the total heat gain admitted through this double-glass window per square
foot of sunlit area is
Selection of Glazing
During the selection of glazing, the following factors should be considered: visual communication,
use of daylight, thermal comfort, summer and winter solar heat gain, street noise attenuation, safety
 64.2 	 9.6  73.8 Btu / h ft2 (233 W/ m2)
Qwi
As
 SHGCoi It 	 U(To  Ti)  0.259  248 	 0.40(100  76)
SC 
SHGCoi
SHGCDSA

0.259
0.87
 0.297
 0.135 	
0.40  0.189
4.44
	 0.40
4.44
	 0.40  1.75 0.135  0.259
SHGCoi 
oi 	
Uo
ho
	 U
ho
	
U
ha i
oi 
o i
1 
2 3

0.8  0.16
1  0.08  0.68
 0.135
 76 	 40.29  1
2.21
	
0.035
2   94.9F (34.9C)
Tgi  Ti 	
QRCi
As
1
hi
	
Rg
2 
 100 	 (46.9 	 33.5  40.29) 1
4.44
	
0.035
2   109.7F
Tgo  To 	 o It 	
i It 
QRCi
As
1
ho
	
Rg
2 
 0.400  46.9
4.44
	 33.5
1
4.44
	 1.75 	 100  76	  40.29 Btu / h ft2 (127.1 W/ m2)
QRCi
As
 U 
o It
ho
	
i It
1
ho
	
1
ha
	 To  Tr	
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.39
and fire protection, and life-cycle cost analysis. Energy conservation considerations, including the
control of solar heat with the optimum combination of absorbing and reflective glass and various
shading devices, are covered in the next section.
To reduce the heat loss through glass during winter, one can install double or triple glazing,
storm windows, or low-emission film coating on the surface of the glass. Elmahdy (1996) and de
Abreu et al. (1996) tested the thermal performance of seven insulating glass units. If a clear,
double-glazed insulating glass unit with a silicone foam spacer of 0.5 in. (13 mm) between two
panes is taken as the base unit, the U values of these seven insulating glass units are as follows:
Low-e coating reduces the U value by about 30 percent, and the triple-glazing drops about onethird
compared with a clear, double-glazed insulating glass unit. If the width of the airspace is
decreased to 0.25 in. (6.5 mm), its U value will increase 15 percent. When the airspace is wider
than 0.5 in. (13 mm), regardless of whether a metal or silicone foam spacer is adopted, neither has
any significant effect on U value of the insulation glass unit.
3.11 SHADING OF GLASS
Shading projected over the surface of glass significantly reduces its sunlit area. Many shading
devices increase the reflectance of the incident radiation. There are two types of shading: deliberately
installed shading devices, which include indoor and external shading devices, and shading
from adjacent buildings.
Indoor Shading Devices
Indoor shading devices not only provide privacy but also are usually effective in reflecting part of
the solar radiation back to the outdoors. They also raise the air temperature of the space between
the shading device and the window glass, which in turn reduces the conductive heat gain in summer.
Indoor shading devices are easier to operate and to maintain and are more flexible in operation
than external shading devices. Three types of indoor shading devices are commonly used: venetian
blinds, draperies, and roller shades.
Venetian Blinds. Most horizontal venetian blinds are made of plastic or aluminum slats, spaced 1
to 2 in. (25 to 50 mm) apart, and some are made of rigid woven cloth. The ratio of slat width to slat
spacing is generally 1.15 to 1.25. For light-colored metallic or plastic slats at a 45 angle, typical
optical properties are  0.05,  0.55, and  0.40. Vertical venetian blinds with wider slats
are widely used in commercial buildings.
Consider a single-glazed window combined with indoor venetian blinds at a slat angle   45,
as shown in Fig. 3.13. Let the subscripts g, v, and a represent the glass, the venetian blinds, and the
U value,
Unit Glazing Airspace, in. Spacer Btu/h ft2 F
1 Clear, double-glazed 0.5 Foam 0.51
2 Clear, double-glazed 0.5 Aluminum 0.51
3 Clear, double-glazed 0.25 Foam 0.58
4 Clear, double-glazed 0.75 Foam 0.51
5 Low-e, double-glazed 0.5 Foam 0.36
6 Clear, Triple-glazed 0.5 Foam 0.32
7 Clear, Triple-glazed 0.25 Foam 0.39
3.40 CHAPTER THREE
air between the venetian blinds and the glass. Also let o indicate the outward direction and i the
inward direction. Of the solar radiation transmitted through the glass and radiating on the surface of
the slats g It,
 A fraction vo g It is reflected outward from the surface of each slat.
 Another fraction vi g It is reflected from the slat surface into the conditioned space.
 A third fraction absorbed by the slat is either convected away by the space air or reflected from
the surface to the indoor surroundings in the form of longwave radiation.
If the glass has a high transmittance, the slat temperature Tv1 will be higher, as shown by the lower
solid temperature curve in Fig. 3.13. If the glass has a high absorptance, its temperature Tg2 will be
higher. In Table 3.8 the shading coefficients of various combinations of venetian blinds and glazing
are listed.
According to results of a field survey by Inoue et al. in four office buildings in Japan in 1988,
60 percent of the venetian blinds were not operated during the daytime. The incident angle of the
direct solar radiation had greater influence on operation of the blinds than the intensity. Automatic
control of slat angle and of raising and lowering the blinds is sometimes used and may become
more popular in the future.
Draperies. These are fabrics made of cotton, regenerated cellulose (such as rayon), or synthetic
fibers. Usually they are loosely hung, wider than the window, and pleated; and they can be drawn
open or closed as required. Drapery-glass combinations reduce the solar heat gain in summer and
increase the thermal resistance in winter. Reflectance of the fabric is the dominant factor in the
reduction of heat gain, and visibility is a function of the openness of the weave.
Roller Shades. These are sheets made of treated fabric or plastic that can be pulled down to cover
the window or rolled up. Roller shades have a lower SC than do venetian blinds and draperies.
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.41
FIGURE 3.13 Heat transfer through a single-glazed window
combined with venetian blinds.
When glass is covered with shades, any outdoor visual communication is blocked, and the visible
light transmittance is less than that of other indoor shading devices.
External Shading Devices
External shading devices include overhangs, side fins, egg-crate louvers, and pattern grilles, as
shown in Fig. 3.14. They are effective in reducing the solar heat gain by decreasing the sunlit area.
However, the external shading devices do not always fit into the architectural requirements and are
less flexible and more difficult to maintain. Pattern grilles impair visibility significantly.
Figure 3.15 shows the shaded area of a glass pane constructed with both overhang and side fin.
The profile angle  is defined as the angle between a horizontal plane and a tilted plane that includes
the suns rays. We see that tan   UQ/OR  tan / cos . Let SW be the width of the
shadow and SH be the height of the shadow projected on the plane of the glass by direct solar radiation,
in feet (meters). Also, let W be the width of the glass and H be the height. Then the shadow
width on the plane of the glass is
SW  PV tan  (3.85)
and the shadow height is
(3.86)
where PV  projection of side fin plus mullion and reveal, ft (m)
PH  horizontal projection of overhang plus transom and reveal, ft (m)
The projection factor due to the external shading device Fpro can be calculated as
(3.87) Fpro 
PH
H
SH  PH tan   PH
tan 
cos 
3.42 CHAPTER THREE
FIGURE 3.14 Types of external shades. (a) Overhang; (b) egg-crate
louver; (c) pattern grilles.
The net sunlit area of the glass As, ft2 (m2), can then be calculated as
As  (W  SW)(H  SH) (3.88a)
and the shaded area of the glass Ash, ft2 (m2), is given by
(3.88b)
where Ag  area of the glass, ft2 (m2).
Shading from Adjacent Buildings
Shadows on glass cast by adjacent buildings significantly reduce the sunlit area of the glass.
For example, in Fig. 3.16 we see the area on building A shaded because of the presence of
building B. Let two sides of the shaded building A coincide with the X and Y axes on the plan
view shown in Fig. 3.16. In the elevation view of Fig. 3.16, the shadow height on the facade of
building A is
(3.89)
and the shadow width SW on the facade of building A is
(3.90) SW  WOB 	 WB  LAB tan 
SH  HB  LAB tan 
Ash  Ag  As  WH  (W  SW)(H  SH)
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.43
FIGURE 3.15 Shaded area of a window glass constructed with an overhang and side fin.
where HB  height of building B, ft (m)
WB  width of building B, ft (m)
LAB  distance between two buildings along X axis, ft (m)
WOB  distance between X axis and building B, ft (m)
Subscript A indicates the shaded building and B the shading building.
Because of the suns varying position, the solar altitude  and the surface solar azimuth  change
their values from time to time. Table 3.10 lists the solar altitude  and solar azimuth  at north
latitudes 32 and 40. To evaluate the shaded area of the outer surface of a building, a computer program
can determine which of several hundred representative points on this outer surface are sunlit or
shaded at specific times of the day. Then a ratio of shaded area to total area can be calculated.
At a certain time instant, a window glass area can be shaded due to the effect of the overhang
and vertical fins, or the effect of the adjacent building, or both. For the calculation of the combined
shading effect of overhangs, vertical fins, and adjacent building, it is recommended that the number
of sunlit and shaded windows of an outer surface of a building under the effect of the adjacent
building be calculated first. According to Eq. (3.88a), the total sunlit area can be calculated as the
product of the net sunlit area for each window because of the overhang and vertical fins and the
number of windows in the sunlit area of that outer surface.
3.44 CHAPTER THREE
FIGURE 3.16 Shading from adjacent building.
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.45
TABLE 3.10 Solar Altitude (ALT) and Solar Azimuth (AZ)
North latitude
32 40
Solar time Solar position Solar time Solar time Solar position Solar time
Date AM ALT AZ PM AM ALT AZ PM
Dec. 8 10 54 4 8 5 53 4
9 20 44 3 9 14 42 3
10 28 31 2 10 21 29 2
11 33 16 1 11 25 15 1
12 35 0 12 12 27 0 12
Jan. 7 1 65 5 8 8 55 4
	 8 13 56 4 9 17 44 3
Nov. 9 22 46 3 10 24 31 2
10 31 33 2 11 28 16 1
11 36 18 1 12 30 0 12
12 38 0 12
Feb. 7 7 73 5 7 4 72 5
	 8 18 64 4 8 15 62 4
Oct. 9 29 53 3 9 24 50 3
10 38 39 2 10 32 35 2
11 45 21 1 11 37 19 1
12 47 0 12 12 39 0 12
Mar. 7 13 82 5 7 11 80 5
	 8 25 73 4 8 23 70 4
Sept. 9 37 62 3 9 33 57 3
10 47 47 2 10 42 42 2
11 55 27 1 11 48 23 1
12 58 0 12 12 50 0 12
Apr. 6 6 100 6 6 7 99 6
	 7 19 92 5 7 19 89 5
Aug. 8 31 84 4 8 30 79 4
9 44 74 3 9 41 67 3
10 56 60 2 10 51 51 2
11 65 37 1 11 59 29 1
12 70 0 12 12 62 0 12
May 6 10 107 6 5 2 115 7
	 7 23 100 5 6 13 106 6
July 8 35 93 4 7 24 97 5
9 48 85 3 8 35 87 4
10 61 73 2 9 47 76 3
11 72 52 1 10 57 61 2
12 78 0 12 11 66 37 1
12 70 0 12
June 5 1 118 7 5 4 117 7
6 12 110 6 6 15 108 6
7 24 103 5 7 26 100 5
8 37 97 4 8 37 91 4
9 50 89 3 9 49 80 3
10 62 80 2 10 60 66 2
11 74 61 1 11 69 42 1
12 81 0 12 12 73 0 12
Source: ASHRAE Handbook 1981, Fundamentals. Reprinted with permission.
3.12 HEAT EXCHANGE BETWEEN THE OUTER BUILDING
SURFACE AND ITS SURROUNDINGS
Because atmospheric temperature is lower at high altitudes, there is always a radiant heat loss from
the outer surface of the building to the sky vault without clouds. However, it may be offset or partly
offset by reflected solar radiation from the ground on a sunny day. Radiant heat loss from the building
needs to be calculated during nighttime and included in year-round energy estimation.
In commercial and institutional buildings using glass, concrete, or face brick on the outside surface
of the building envelope, the migration of moisture through the glass pane is rather small.
Because of the heavy mass of the concrete wall, the influence of the diurnal cyclic variation of the
relative humidity of outdoor air on moisture transfer through the building envelope is also small.
Therefore, for simplicity, the moisture transfer between the building envelope and the outside air
can be ignored.
The heat balance at the outer building surface, as shown in Fig. 3.17, can be expressed as
follows:
(3.91)
In Eq. (3.91), Qsol represents the solar radiation absorbed by the outer surface of the building envelope,
in Btu/ h (W). It can be calculated as
(3.92)
where os  absorptance of the outer surface of the building envelope. From Eq. (3.61), the reflection
of solar radiation from any reflecting surface to the outer surface of the building and absorbed
by it, or qref , Btu /h (W), is given by
(3.93)
where A  total area of the outer surface of the building envelope, . The
term qos indicates the convective heat transfer from the outer surface of the building outward, in
A  As 	 Ash, ft2 (m2)
Qref  A os Iref  A s os Fsr (ID 	 Id )
Qsol 
os [As (ID 	 Id ) 	 Ash Id]
Qsol 	 Qref  Qos 	 (Qrad  Qat) 	 Qoi
3.46 CHAPTER THREE
FIGURE 3.17 Heat balance at the outer surface of a building.
Btu/h (W). From Eq. (3.7), it can be calculated as
(3.94)
where Tos , To  outer surface temperature of the building and the outdoor air temperature, respectively,
F (C)
hc  convective heat-transfer coefficient of outer surface, Btu /h ft2 F (W/m2 C)
The term Qoi denotes the inward heat flow from the outer building surface, in Btu/h (W). The term
Qrad  Qat indicates the net heat emitted from the outer surface of the building because of the radiation
exchange between the surface and the atmosphere, in Btu/h (W). Here, Qrad represents
the longwave radiation emitted from the surface, and Qat indicates the atmospheric radiation to the
surface.
Kimura (1977) found that atmospheric radiation can be expressed as
(3.95)
where Ccc  cloud cover factor, which can be obtained from local climate records, dimensionless
Kcc  cloudy reduction factor
Usually, the smaller the value of Ccc, the higher the clouds and the smaller the Kcc value. For simplicity,
it can be calculated as
(3.96)
And Br is actually the emissivity of the atmosphere, and it can be expressed by an empirical formula
developed by Brunt:
(3.97)
where pw  water vapor pressure, psia. In Eq. (3.95), TRg represents the absolute ground temperature,
R (K), and TRo the absolute outdoor temperature, R (K). Then, radiant heat loss can be written as
(3.98)
where os  emissivity of outer surface of building
Fsat  shape factor between outer surface and atmosphere
TRos  absolute outer surface temperature of building, R (K)
Sol-Air Temperature
For a sunlit outer surface of a building, if Qref is mainly from ground-reflected solar radiation, it is
offset by and Eq. (3.91) becomes
Let . Substituting for Qoi, then, gives
and
(3.99)
In Eq. (3.99), Tsol is called sol-air temperature, in F (C). It is a fictitious outdoor temperature that
combines the effect of the solar radiation radiated on the outer surface of the building and the inward
heat transfer due to the outdoorindoor temperature difference.
Tsol  To 	
os It
ho
os It  ho(Tos  To) 	 ho(Tsol  Tos)
Qoi  ho A(Tsol  Tos)
Qsol  Qos 	 Qoi
Qrad  Qat,
Qrad  Qat 
os AFsat {T Ros
4  [(1  CccKcc )T Ro
4
Br 	 Ccc Kcc T Rg
4 ]}
Br  0.51 	 0.55?pw
Kcc  0.83  0.4Ccc
Qat  (1  CccKcc )  T Ro
4 Br 	 CccKcc
 T Rg
4
Qos  hc A(Tos  To)
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.47
Example 3.2. At midnight of July 21, the outdoor conditions of an office building are as follows:
Outdoor temperature 75F (23.9C)
Water vapor pressure of outdoor air 0.215 psia
Ground temperature 72F (22.2C)
Cloud cover factor Ccc 0.1
Emissivity of the outer surface 0.90
Shape factor between the surface and sky 0.5
If the outer surface temperature of this building is 76F (24.4C) and the Stefan-Boltzmann constant

  0.1714  108 Btu/h ft2 R4 (5.67  108 W/m2 K4), find the radiant heat loss from
each square foot (square meter) of the vertical outer surface of this building.
Solution. From Eq. (3.97) and given data, we can see
And from Eq. (3.96), the cloudy reduction factor is
From the given, TRos  76 	 460  536R, TRo  75 	 460  535R, and TRg  72 	 460 
532R; then, from Eq. (3.98), the radiant heat loss from the outer surface of the building due to
radiant exchange between the surface and the atmosphere is
3.13 COMPLIANCE WITH ASHRAE/IESNA STANDARD 90.1-1999
FOR BUILDING ENVELOPE
An energy-efficient and cost-effective building envelope should meet the requirements and design
criteria in ASHRAE/IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise
Residential Buildings; the DOE Code of Federal Regulations, Title10, Part 435, Subpart A, Performance
Standard for New Commercial and Multi-Family High-Rise Residential Buildings; and local
energy codes. The DOE Code is very similar to Standard 90.1.
The design and selection of the building envelope are generally the responsibility of an architect
with the assistance of a mechanical engineer or contractor. Building envelopes are usually designed,
or even constructed, before the HVAC&R system is designed. A speculative building is built of
known use and type of occupancy, but the exact tenants are unknown. A shell building is built before
the use and occupancy are determined.
Compliance with the ASHRAE/IES Standard 90.1-1999 for building envelopes includes the following.
Refer to Standard 90.1-1999 for exceptions and details.
General Requirements
The requirements apply to the exterior building envelope which separates conditioned space from
the outdoors and the semiexterior building envelope which separates the conditioned space from
semiheated space or unconditioned space, or separated semiheated space from the unconditioned
space or outdoors.
Semiheated space is an enclosed space, but not a conditioned space, within a building that is
heated by a heating system whose output capacity is greater than or equal to 3.4 Btu/h ft (10.71
	 0.1  0.79  (5.32)4]}  14.27 Btu / h ft2 (45.02 W/ m2)
 0.90  0.5  0.174{(5.36)4  [(1  0.1  0.79)(5.35)4  0.765
Qrad  Qat 
os Fsat

{T Ros
4  [(1  CccKcc )T Ro
4 Br 	 CccKcc T Rg
4 ]}
Kcc  0.83  0.4Ccc  0.83  0.4(0.1)  0.79
Br  0.51 	 0.55?pw  0.51 	 0.55?0.215  0.765
3.48 CHAPTER THREE
W/m2). Compliance includes mandatory provisions, the prescriptive building envelope option, and
the building envelope trade-off option.
Mandatory Provisions
The standard mandates that insulation materials shall be installed in accordance with the manufacturers
recommendation to achieve rated R value of insulation. Rated R value of insulation is the
thermal resistance of the insulation alone in units of h  ft2 F/Btu (m2 C/W) at a mean temperature
of 75F (23.9C). Rated R value refers to the thermal resistance of the added insulation in
framing cavities or insulated sheathing only and does not include the thermal resistance of other
building materials or air films.
Insulation shall be installed in a permanent manner in substantial contact with the inside surface.
The roof insulation shall not be installed on a suspended ceiling with removable ceiling panels. Insulation
outdoors shall be covered with a protective material to prevent damage from sunlight,
moisture, landscaping operations, equipment maintenance, and wind. Insulation materials contacts
with the ground shall have a water absoption rate no greater than 0.3 percent.
Fenestration performance shall be determined from production line units or representative units purchased.
U factors shall be determined in accordance with National Fenestration Rating Council (NFRC)
100. Solar heat gain coefficient (SHGC) for the overall fenestration area shall be determined in accordance
with NFRC 200. Visible light transmittance shall be determined in accordance with NFRC 300.
The following areas of building envelope shall be sealed, caulked, gasketed, or weather-stripped
until air leakage is minimal:
 Joints around fenestration and door frames
 Junctions between walls and foundations, between wall corners, between walls and roofs, wall
and floors, or walls and panels
 Openings because of penetrations of utility services through roofs, walls, and floors
 Fenestrations and doors built at site
 Building assemblies used as ducts and plenums
 Joints, seams, and penetrations due to vapor retarders
 Other openings in the building envelope
Air leakage of fenestration and doors shall be determined in accordance with NFRC 400. Air
leakage shall not exceed 1.0 cfm/ft2 (5.0 L/s m2) for glazed swinging entrance doors and revolving
doors and also 0.4 cfm/ft2 (2.0 L/sm2) for all other products. A door that separates conditioned
space from the exterior shall be protected with an enclosed vestibule with all doors into and out of
the vestibule installed with self-closing devices.
All mandatory provisions in Standard 90.1-1999 will be presented in the form standard mandates
that, whereas for nonmandatory provisions, only standard specifics that will be used.
Prescriptive Building Envelope Option
The exterior building envelope shall comply with the requirements for the appropriate conditioned
space in Table 5.3 for various climate (located in Normative Appendix B of Standard 90.1-1999). In
Table B1 to B26 of I-P edition:
 There are 26 tables, each of them has a number of heating degree days of 65F (HDD65), and a
number of cooling degree days of 50F (CDD 50). Select a table with number of HDD 65 and
CDD 50 equal to or most nearly equal to the values where the proposed building locates.
 There are three kinds of conditioned space: nonresidential, residential, semiheated.
 Total vertical fenestration area, including both fixed and operable fenestration shall be less than
50 percent of the gross wall area. The total skylight area, including glass and plastic skylights
with or without a curb, shall be less than 5 percent of the roof area.
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.49
Fenestration, including fixed and operable vertical fenestration, shall have a U factor not greater
than that specified in Table 5.3, 90.1-1999. Vertical fenestration shall have an SHGC not greater
than that specified for all orientations in Table 5.3, 90.1-1999. There are only visible light transmittance
criteria in the Building Envelope Trade-Off Option.
 All roofs shall have a rated R value of insulation not less than that specified in Table 5.3,90.1-
1999. Skylight curbs shall be insulated to the level of the roofs with the insulation entirely above
the deck or R-5, whichever is less.
 All above-grade walls shall have a rated value of insulation not less than that specified in
Table 5.3, 90.1-1999. Mass wall heating capacity shall be determined from Table A-6 or A-7 in
Standard 90.1-1999. Below-grade walls shall have a rated R value of insulation not less than that
specified in Table 5.3, 90.1-1999.
 All floors and heated or unheated slab-on-grade floors shall have a rated R value of insulation not
less than that specified in Table 5.3, 90.1-1999. Slab-on-grade floor insulation shall be installed
around the perimeter of the slab-on-grade floor to the distance specified.
 All opaque doors shall have a U factor not greater than that specified in Table 5.3, 90.1-1999.
Building Envelope Trade-Off Option
The building envelope complies with the standard if the proposed building satisfied the provisions
of 5.1 and 5.2 of Standard 90.1-1999 and the envelope performance factor of the proposed building
is less than or equal to the envelope performance factor of the budget building. The envelope performance
factor considers only the building envelope components. Schedules of operation, lighting
power input, equipment power input, occupant density, and mechanical systems shall be the same
for both the proposed building and budget building. Envelope performance factors shall be calculated
using the procedures of Normative Appendix C, Standard 90.1-1999. Refer to Standard 90.1-
1999 for details.
3.14 ENERGY-EFFICIENT AND COST-EFFECTIVE MEASURES FOR
BUILDING ENVELOPE
According to the economic parametric analysis of the thermal design in Johnson et al. (1989), the
following are energy-efficient and cost-effective measures for the design of building envelopes for
office buildings.
Exterior Walls
An increase in the mass of the exterior wall, i.e., its thermal capacitance, reduces only the peak
heating and cooling loads. The increase in electrical usage and capital investment often offsets the
benefit of the decrease of peak loads. The variation of life-cycle costing is often negligible.
Increasing the insulation to more than 2-in. (50-mm) thick decreases only the natural gas usage
for heating. This is cost-effective for areas with very cold winters. In some cases, the increase in the
annual cost due to the increase in the capital investments of the insulation may balance the reduction
of gas usage. A careful analysis is required.
Windows
The vertical fenestration area ratio (VFR) is the single parameter that most influences the building
life-cycle costing among building envelopes for high-rise buildings. For many office buildings, VFR
lies between 0.2 and 0.3.
3.50 CHAPTER THREE
Heat-absorbing and -reflective glasses produce a significant building cost savings in areas where
solar heat control is important in summer. Double-panes, triple-panes, and low-emission films are
effective in reducing the U value of the window assembly and therefore, the heating and cooling
loads. Indoor shading devices with window management systems are cost-effective. For example,
an indoor shading device can be turned on when solar heat gain exceeds 20 Btu/h ft2 (63 W/m2).
Although overhangs reduce the cooling loads, they may increase the need for electric lighting for
daylit buildings. The effect of overhang usage on building life-cycle costing is not significant in
many instances.
Infiltration
Infiltration has a significant influence on heating and cooling loads. Indoor air quality must be guaranteed
by mechanical ventilation systems and sufficient outdoor air intake through these systems. It
is desirable that windows and cracks in joints be well sealed. In multiple-story buildings, infiltration
through elevator shafts, pipe shafts, and duct shafts should be reduced.
Energy-Efficient Measures for Commercial Buildings in the United States
According to the EIAs Commercial Buildings Characteristics, in 1992, the breakdown of
energy conservation features for commercial buildings for a total area of 67,876 million ft2 (6308
million m2) in the United States is as follows:
Roof or ceiling insulation 74%
Wall insulation 49%
Storm or multiple glazing 44%
Tinted, reflective, or shading glass 37%
Exterior or interior shading devices 50%
Windows that are openable 43%
REFERENCES
de Abreu, P. F., Fraser, R. A., Sullivan, H. F., and Wright, J. L., A Study of Insulated Glazing Unit Surface
Temperature Profiles Using Two-Dimensional Computer Simulation, ASHRAE Transactions, 1996, Part II,
pp. 497507.
Altmayer, E. F., Gadgil, A. J., Bauman, F. S., and Kammerud, R. C., Correlations for Convective Heat Transfer
from Room Surfaces, ASHRAE Transactions, 1983 Part II A, pp. 6177.
ASHRAE, ASHRAE Handbook 1989, Fundamentals, Atlanta, GA, 1989.
ASHRAE, ASHRAE Handbook 1997, Fundamentals, ASHRAE Inc., Atlanta, GA, 1997.
ASHRAE, ASHRAE/IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential
Buildings, Atlanta, GA, 1999.
ASHRAE, Procedure for Determining Heating and Cooling Loads for Computerizing Energy Calculations, Algorithms
for Building Heat Transfer Subroutines, Atlanta, GA, 1976.
Bauman, F., Gadgil, A., Kammerud, R., Altmayer, E., and Nansteel, M., Convective Heat Transfer in Buildings:
Recent Research Results, ASHRAE Transactions, 1983, Part I A, pp. 215233.
Chandra, S., and Kerestecioglu, A. A., Heat Transfer in Naturally Ventilated Rooms: Data from Full Scale
Measurements, ASHRAE Transactions, 1984, Part I B, pp. 211225.
Dahlen, R. R., Low-E Films for Window Energy Control, ASHRAE Transactions, 1987, Part I, pp. 15171524.
Deringer, J. J., An Overview of Standard 90.1: Building Envelope, ASHRAE Journal, no. 2, 1990, pp. 3034.
HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.51
Donnelly, R. G., Tennery, V. J., McElroy, D. L., Godfrey, T. G., and Kolb, J. O., Industrial Thermal Insulation,
An Assessment, Oak Ridge National Laboratory Report TM-5283, TM-5515, and TID-27120, 1976.
Energy Information Administration, Commercial Buildings Characteristics 1992, Commercial Buildings Energy
Consumption Survey April 1994,Washington, 1994.
Elmahdy, H., Surface Temperature Measurement of Insulating Glass Units Using Infrared Thermography,
ASHRAE Transactions, 1996, Part II, pp. 489496.
Fairey, P. W., and Kerestecioglu, A. A., Dynamic Modeling of Combined Thermal and Moisture Transport in
Buildings: Effect on Cooling Loads and Space Conditions, ASHRAE Transactions, 1985 Part II A,
pp. 461472.
Galanis, N., and Chatigny, R., A Critical Review of the ASHRAE Solar Radiation Model, ASHRAE Transactions,
1986, Part I A, pp. 410419.
Glicksman, L. R., and Katsennelenbogen, S., A Study of Water Vapor Transmission Through Insulation under
Steady State and Transient Conditions, ASHRAE Transactions, 1983, Part II A, pp. 483499.
Gueymard, C. A., A Simple Model of the Atmospheric Radiative Transfer of Sunshine: Algorithms and Performance
Assessment, Report FSEC-PF-270-95, Florida Solar Energy Center, Cocoa, Fla., 1995.
Hagentoft, C-E., Moisture Conditions in a North-Facing Wall with Cellulose Loose-Fill Insulation: Construction
with and without a Vapor Retarder and Air Leakage, ASHRAE Transactions, 1995, Part I, pp. 639646.
Handegrod, G. O. P., Prediction of the Moisture Performance of Walls, ASHRAE Transactions, 1985, Part II B,
pp. 15011509.
Inoue, T., Kawase, T., Ibamoto, T., Takakusa, S., and Matsuo,Y., The Development of an Optimal Control System
for Window Shading Devices Based on Investigations in Office Buildings, ASHRAE Transactions, 1988,
Part II, pp. 10341049.
Johnson, C. A., Besent, R. W., and Schoenau, G. J., An Economic Parametric Analysis of the Thermal Design of a
Large Office Building under Different Climatic Zones and Different Billing Schedules, ASHRAE Transactions,
1989, Part I, pp. 355369.
Kays, M. M., and Crawford, M. E., Convective Heat and Mass Transfer, 2d ed., McGraw-Hill, New York, 1980.
Kimura, K., Scientific Basis of Air Conditioning, Applied Science Publishers, London, 1977.
Kimura, K., and Stephenson, D. G., Solar Radiation on Cloudy Days, ASHRAE Transaction, 75(1), 1969, pp.
18.
Miller, A., Thompson, J. C., Peterson, R. E., and Haragan, D. R., Elements of Meteorology, 4th ed., Bell and
Howell Co., Columbus, Ohio, 1983.
Robertson, D. K., and Christian, J. E., Comparison of Four Computer Models with Experimental Data from
Test Buildings in Northern New Mexico, ASHRAE Transactions, 1985, Part II B, pp. 591607.
Sato, A., Eto, N., Kimura, K., and Oka, J., Research on the Wind Variation in the Urban Area and Its Effects in
Environmental Engineering No. 7 and No. 8Study on Convective Heat Transfer on Exterior Surface of
Buildings, Transactions of Architectural Institute of Japan, no. 191, January 1972.
Smolenski, C. P., Absorption in Thermal Insulation: How Much Is Too Much? HPAC no. 11, 1996, pp. 4958.
Spitler, J. D., Pedersen, C. O., and Fisher, D. E., Interior Convective Heat Transfer in Buildings with Large Ventilative
Flow Rates, ASHRAE Transactions, 1991, Part I, pp. 505514.
Stewart,W. E., Effect of Air Pressure Differential on Vapor Flow through Sample Building Walls, ASHRAE
Transactions, 1998, Part II, pp. 1724.
Verschoor, J. D., Measurement of Water Vapor Migration and Storage in Composite Building Construction,
ASHRAE Transactions, 1985, Part II A, pp. 390403.
Wang, S. K., Air Conditioning, vol. 1, Hong Kong Polytechnic, Hong Kong, 1987.
Wong, S. P. W., Simulation of Simultaneous Heat and Moisture Transfer by Using the Finite Difference Method
and Verified Tests in a Test Chamber, ASHRAE Transactions, 1990, Part I, pp. 472485.
Wong, S. P .W., and Wang, S. K., Fundamentals of Simultaneous Heat and Moisture Transfer between the
Building Envelope and the Conditioned Space Air, ASHRAE Transactions, 1990, Part II, pp. 7383.
3.52 CHAPTER THREE
4.1
CHAPTER 4
INDOOR AND OUTDOOR
DESIGN CONDITIONS
4.1 INDOOR DESIGN CONDITIONS 4.1
4.2 HEAT EXCHANGE BETWEEN
HUMAN BODY AND INDOOR
ENVIRONMENT 4.2
Two-Node Model of Thermal
Interaction 4.2
Steady-State Thermal Equilibrium 4.3
Transient Energy Balance 4.3
4.3 METABOLIC RATE AND SENSIBLE
HEAT LOSSES FROM HUMAN BODY 4.4
Metabolic Rate 4.4
Mechanical Work 4.4
Sensible Heat Exchange 4.5
Clothing Insulation 4.7
4.4 EVAPORATIVE HEAT LOSSES 4.7
Respiration Losses 4.7
Evaporative Heat Loss from Skin
Surface 4.7
Maximum Evaporative Heat Loss due to
Regulatory Sweating 4.7
Diffusion Evaporative Heat Loss and
Total Skin Wetness 4.8
4.5 MEAN RADIANT TEMPERATURE
AND EFFECTIVE TEMPERATURE 4.9
Mean Radiant Temperature 4.9
Effective Temperature 4.14
4.6 FACTORS AFFECTING THERMAL
COMFORT 4.14
4.7 THERMAL COMFORT 4.15
Fangers Comfort Equation 4.15
ASHRAE Comfort Zones 4.17
Comfort-Discomfort Diagrams 4.17
4.8 INDOOR AIR TEMPERATURE AND AIR
MOVEMENTS 4.20
Comfort Air Conditioning Systems 4.20
Design Considerations 4.21
Indoor Design Temperatures for Comfort
Air Conditioning 4.21
Process Air Conditioning Systems 4.23
4.9 HUMIDITY 4.23
Comfort Air Conditioning Systems 4.23
Process Air Conditioning Systems 4.24
4.10 SICK BUILDING SYNDROME AND
INDOOR AIR QUALITY 4.27
Indoor Air Contaminants 4.27
Basic Strategies to Improve Indoor Air
Quality 4.29
Outdoor Air Requirements for
Occupants 4.30
4.11 AIR CLEANLINESS 4.31
4.12 SOUND LEVEL 4.32
Sound and Sound Level 4.32
Sound Power Level and Sound Pressure
Level 4.32
Octave Bands 4.33
Addition of Sound Levels 4.33
Human Response and Design
Criteria 4.34
4.13 SPACE PRESSURE
DIFFERENTIAL 4.37
4.14 OUTDOOR DESIGN
CONDITIONS 4.38
Summer and Winter Outdoor Design
Conditions 4.39
The Use of Outdoor Weather Data in
Design 4.39
Outdoor Weather Characteristics and
Their Influence 4.42
REFERENCES 4.42
4.1 INDOOR DESIGN CONDITIONS
Indoor design parameters are those that the air conditioning system influences directly to produce a
required conditioned indoor environment in buildings. They are shown below and grouped as follows:
1. Basic design parameters
 Indoor air temperature and air movements
 Indoor relative humidity
2. Indoor air quality
 Air contaminants
 Outdoor ventilation rate provided
 Air cleanliness for processing
3. Specific design parameters
 Sound level
 Pressure differential between the space and surroundings
The indoor design parameters to be maintained in an air conditioned space are specified in the
design document and become the targets to be achieved during operation. In specifying the indoor
design parameters, the following points need to be considered:
1. Not all the parameters already mentioned need to be specified in every design project. Except
for the indoor air temperature which is always an indoor design parameter in comfort air conditioning,
it is necessary to specify only the parameters which are essential to the particular situation
concerned.
2. Even for process air conditioning systems, the thermal comfort of the workers should also be
considered. Therefore, the indoor design parameters regarding health and thermal comfort for the
occupants form the basis of design criteria.
3. When one is specifying indoor design parameters, economic strategies of initial investment
and energy consumption of the HVAC&R systems must be carefully investigated. Design criteria
should not be set too high or too low. If the design criteria are too high, the result will be an excessively
high investment and energy cost. Design criteria that are too low may produce a poor indoor
air quality, resulting in complaints from the occupants, causing low-quality products, and possibly
leading to expensive system alternations.
4. Each specified indoor design parameter is usually associated with a tolerance indicated
as a  sign, or as an upper or lower limit. Sometimes there is a traditional tolerance understood
by both the designers and the owners of the building. For instance, although the summer
indoor design temperature of a comfort air conditioning system is specified at 75 or 78F
(23.9 or 25.6C), in practice a tolerance of 23F (1.1  1.7C) is often considered
acceptable.
5. In process air conditioning systems, sometimes a stable indoor environment is more important
than the absolute value of the indoor parameter to be maintained. For example, it may not be necessary
to maintain 68F (20C) for all areas in precision machinery manufacturing. More often, a
72F (22.2C) or even a still higher indoor temperature with appropriate tolerance will be more
suitable and economical.
4.2 HEAT EXCHANGE BETWEEN HUMAN BODY AND
INDOOR ENVIRONMENT
Two-Node Model of Thermal Interaction
In 1971, Gagge et al. recommended a two-node model of human thermal interaction. In this model,
the human body is composed of two compartments: an inner body core, including skeleton, muscle
and internal organs; and an outer shell of skin surface. The temperatures of the body core and the
surface skin are each assumed to be uniform and independent. Metabolic heat production, external
mechanical work, and respiratory losses occur only in the body core. Heat exchange between the
body core and the skin surface depends on heat conduction from direct contact and the peripheral
blood flow of the thermoregulatory mechanism of the human body.
4.2 CHAPTER FOUR
Steady-State Thermal Equilibrium
When the human body is maintained at a steady-state thermal equilibrium, i.e., the heat storage
at the body core and skin surface is approximately equal to zero, then the heat exchange between
the human body and the indoor environment can be expressed by the following heat balance
equation:
M  W  C  R  Esk  Eres (4.1)
where M  metabolic rate, Btu /h ft2 (W/m2)
W  mechanical work performed, Btu/h ft2 (W/m2)
C  R  convective and radiative, or sensible heat loss from skin surface,
Btu/h ft2 (W/m2)
Esk  evaporative heat loss from skin surface, Btu /h ft2 (W/m2)
Eres  evaporative heat loss from respiration, Btu/h ft2 (W/m2)
In Eq. (4.1), the ft2 in the unit Btu /h ft2 applies to the skin surface area. The skin surface area of a
naked human body can be approximated by an empirical formula proposed by Dubois in 1916
AD  0.657mb
0.425 Hb
0.725 (4.2)
where AD  Dubois surface area of naked body, ft2 (m2)
mb  mass of human body, lb (kg)
Hb  height of human body, ft (m)
In an air conditioned space, a steady-state thermal equilibrium is usually maintained between the
human body and the indoor environment.
Transient Energy Balance
When there is a transient energy balance between the human body and the indoor environment, the
thermal interaction of the body core, skin surface, and indoor environment forms a rate of positive
or negative heat storage both in the body core and on the skin surface.
The human body needs energy for physical and mental activity. This energy comes from the oxidation
of the food taken into the human body. The heat released from this oxidation process is
called metabolic heat. It dissipates from the skin surface of the human body into the surroundings.
In a cold environment, the thermoregulatory mechanism reduces the rate of peripheral blood circulation,
lowering the temperature of the skin and preventing any greater heat loss from the human
body. However, if the heat loss and the mechanical work performed are greater than the rate of
metabolic heat produced, then the temperatures of both the body core and the skin surface fall, and
shivering or other spontaneous activities occur to increase the production of heat energy within the
human body.
On the other hand, in a hot environment, if a large amount of heat energy needs to be dissipated
from the human body, the physiological control mechanism increases the blood flow to the skin
surface. This raises the skin temperature. If the heat produced is still greater than the heat actually
dissipated and the temperature of the body core increased from its normal temperature of about
97.6 to about 98.6F (36.4 to about 37.0C), then liquid water is released from the sweat glands for
evaporative cooling.
For a transient state of energy balance between the human body and the indoor environment, the
rate of heat storage in the body core Scr and the skin surface Ssk, both in Btu/h ft2 (W/m2), can be
calculated as
Scr  Ssk  M  W  (C  R)  Esk  Eres (4.3)
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.3
4.3 METABOLIC RATE AND SENSIBLE HEAT LOSSES FROM
HUMAN BODY
Metabolic Rate
The metabolic rate M is the rate of energy release per unit area of skin surface as a result of the
oxidative processes in the living cells. Metabolic rate depends mainly on the intensity of the physical
activities performed by the human body. The unit of metabolic rate is called the met. One met is
defined as 18.46 Btu/h ft2 (58.24 W/m2) of metabolic heat produced in the body core. In Table 4.1
are listed the metabolic rates of various activities.
Mechanical Work
Some of the energy released from the oxidative processes within the body core can be partly transformed
to external mechanical work through the action of the muscles. Mechanical work W is
usually expressed as a fraction of the metabolic rate and can be calculated as
W  	M (4.4)
where 	 mechanical efficiency. For most office work, mechanical efficiency is 	 less than 0.05.
Only when there is a large amount of physical activity such as bicycling, lifting and carrying, or
walking on a slope may increase 	 to a value of 0.2 to 0.24.
4.4 CHAPTER FOUR
TABLE 4.1 Metabolic Rate for Various Activities
Metabolic rate
Activity level Met Btu/h ft2
Resting
Sleeping 0.7 13
Seated, quiet 1.0 18
Office work
Reading, seated 1.0 18
Typing 1.1 20
Teaching 1.6 30
Domestic work
Cooking 1.62.0 2937
House cleaning 2.03.4 3763
Walking
Speed 2 mph 2.0 37
4 mph 3.8 70
Machine work
Light 2.02.4 3744
Heavy 4.0 74
Dancing, social 2.44.4 4481
Sports
Tennis, singles 3.64.0 6674
Basketball 5.07.6 90140
Wrestling 7.08.7 130160
Source: Adapted with permission from ASHRAE
Handbook 1989, Fundamentals.
Sensible Heat Exchange
Sensible heat loss or, occasionally, sensible heat gain R  C represents the heat exchange between
the human body and the indoor environment through convective and radiative heat transfer.
Figure 4.1 shows the sensible heat exchange between the human body and the environment. The
combined convective and radiative heat transfer can be calculated as
(4.5)
where Tcl  mean surface temperature of clothing, F (C). The operative temperature To is defined
as the weighted average of the mean radiant temperature Trad and indoor air temperature Ta, both in
F (C), that is,
(4.6)
The surface heat-transfer coefficient is defined as
h  hc  hr (4.7)
and the ratio of the clothed surface area to the naked surface area is
(4.8)
where hc, hr  convective and radiative heat-transfer coefficient, Btu/h ft2 F (W/m2 C)
Acl  surface area of clothed body, ft2 (m2)
The mean radiant temperature Trad is discussed in greater detail in later sections.
According to Seppenan et al. (1972), the convective heat-transfer coefficient hc for a person
standing in moving air, when the air velocity is 30 v 300 fpm (0.15 v 1.5 m/ s), is
hc  0.0681v0.69 (4.9)
fcl 
Acl
AD
To 
hrTrad  hcTa
hr  hc
 fclh(Tcl  To)
C  R  fclhc (Tcl  Ta)  fclhr(Tcl  Trad)
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.5
FIGURE 4.1 Sensible heat exchange between the human body and the
indoor environment.
When the air velocity v	30 fpm (0.15 m/ s), hc  0.7 Btu/h ft2 F (4 W/m2 C). For typical
indoor temperatures and a clothing emissivity nearly equal to unity, the linearized radiative heattransfer
coefficient hr  0.83 Btu/h ft2 F (4.7 W/m2 C). Let Rcl be the R value of clothing, in
h ft2 F/Btu (m2 C/W). Then
(4.10)
where Tsk  mean skin surface temperature, F (C). If the human body is able to maintain a thermal
equilibrium with very little evaporative loss from the skin surface, then the skin temperature
Tsk.n will be around 93F (33.9C).
Combining Eqs. (4.5) and (4.10), and eliminating Tcl, we find
(4.11)
In this equation, the dimensionless clothing efficiency Fcl is defined as
(4.12)
In Eq. (4.11), if To Tsk, then C  R could be negative, i.e., there could be a sensible heat gain.
Fcl 
1
Rcl fcl h  1

Tcl  To
Tsk  To
 Fcl fcl h(Tsk  To)
C  R 
fcl h(Tsk  To)
Rcl fcl h  1
C  R 
Tsk  Tcl
Rcl
4.6 CHAPTER FOUR
TABLE 4.2 Insulation Values for Clothing Ensembles*
Rcl,
h ft2 F
Ensemble description Rcl, clo Btu fcl
Walking shorts, short-sleeve shirt 0.41 0.36 1.11
Fitted trousers, short-sleeve shirt 0.50 0.44 1.14
Fitted trousers, long-sleeve shirt 0.62 0.55 1.19
Same as above, plus suit jacket 0.96 0.85 1.23
Loose trousers, long-sleeve shirt,
long-sleeve sweater, T-shirt 1.01 0.89 1.28
Sweat pants, sweatshirt 0.77 0.68 1.19
Knee-length skirt, short-sleeve shirt,
pantyhose (no socks), sandals 0.54 0.48 1.26
Knee-length skirt, long-sleeve shirt,
full slip, pantyhose (no socks) 0.67 0.59 1.29
Long-sleeve coveralls, T-shirt 0.72 0.63 1.23
Overalls, long-sleeve shirts, long
underwear tops and bottoms,
flannel long-sleeve shirt 1.00 0.88 1.28
*For mean radiant temperature equal to an air temperature and air velocity less
than 40 fpm.
Unless otherwise noted, all ensembles included briefs or panties, shoes,
and socks.
Source: Adapted from McCullough and Jones (1984). Reprinted with permission.
Clothing Insulation
Clothing insulation Rcl can be determined through measurements on a heated manikin, a model of
the human body for laboratory experiments. After C  R is measured from the thermal manikin in
a controlled indoor environment, Rcl can be calculated from Eq. (4.11) since fcl, Tsk, To, and h are
also known values. Clothing insulation Rcl can be expressed either in h  ft2 F/Btu (m2 C/W) or
in a new unit called clo, where 1 clo  0.88 h ft2 F/Btu (0.16 m2 C/W).
Clothing insulation Rcl values and area ratios fcl for typical clothing ensembles, taken from
McCullough and Jones (1984), are listed in Table 4.2.
4.4 EVAPORATIVE HEAT LOSSES
Evaporative heat loss E is heat loss due to the evaporation of sweat from the skin surface Esk and
respiration losses Eres. Actually, metabolic heat is mainly dissipated to the indoor air through the
evaporation of sweat when the indoor air temperature is nearly equal to the skin temperature.
Respiration Losses
During respiration, there is convective heat loss Cres that results from the temperature of the
inhaled air being increased to the exhaled air temperature, or about 93F (33.9C). There is also a
latent heat loss Lres due to evaporation of liquid water into water vapor inside the body core.
The amount of respiration loss depends mainly on the metabolic rate. In summer, at an indoor
temperature of 75F (23.9C) and a relative humidity of 50 percent, respiratory losses Eres  Cres
 Lres are approximately equal to 0.09M. In winter, Eres is slightly greater. For simplicity, let
Eres  0.1M.
Evaporative Heat Loss from Skin Surface
Evaporative heat loss from the skin surface Esk consists of (1) the evaporation of sweat as a result of
thermoregulatory mechanisms of the human body Ersw and (2) the direct diffusion of liquid water
from the skin surface Edif. Evaporative heat loss due to regulatory sweating Ersw, Btu/h ft2 (W/m2),
is directly proportional to the mass of the sweat produced, i.e.,
(4.13)
where mass flow rate of sweat produced, lb/h ft2 (kg/ s m2)
hfg  latent heat of vaporization at 93F (33.9C), Btu/ lb (J /kg)
Maximum Evaporative Heat Loss due to Regulatory Sweating
The wetted portion of the human body needed for the evaporation of a given quantity of sweat wrsw
is
(4.14)
In Eq. (4.14), Emax represents the maximum evaporative heat loss due to regulatory sweating when
the skin surface of the human body is entirely wet. Its magnitude is directly proportional to the
vapor pressure difference between the wetted skin surface and the indoor ambient air, and it can be
wrsw 
Ersw
Emax
m? rsw 
Ersw
m? rswhfg
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.7
calculated as
Emax  he,c(psk, s  pa) (4.15)
where psk, s  saturated water vapor pressure at skin surface temperature, psia (kPa)
pa  water vapor pressure of ambient air, psia (kPa )
he,c  overall evaporative heat-transfer coefficient of clothed body, in Btu/h ft2psi
(W/m2 kPa)
Woodcock (1962) proposed the following relationship between he, c and hs, the overall sensible heat
transfer coefficient, in Btu/h ft2 psi (W/m2 kPa):
(4.16)
The moisture permeability index im denotes the moisture permeability of the clothing and is dimensionless.
Clothing ensembles worn indoors usually have an im  0.3 to 0.5. The moisture permeability
indexes im of some clothing ensembles are presented in Table 4.3.
The Lewis relation LR in Eq. (4.16) relates the evaporative heat-transfer coefficient he and the
convective heat-transfer coefficient hc, both in Btu /h ft2 F(W/m2C). LR  f (he / hc) has a magnitude
of 205F/psi (16.5C/kPa). In Eq. (4.16), the overall sensible heat-transfer coefficient hs can
be calculated as
(4.17)
where Rt  total resistance to sensible heat transfer between the skin and the indoor environment,
h ft2 F/Btu (m2C/W).
Diffusion Evaporative Heat Loss and Total Skin Wetness
The minimum level of evaporative heat loss from the skin surface occurs when there is no regulatory
sweating and the skin wetness due to direct diffusion Edf, min, Btu/h ft2 (W/m2), is approximately
equal to 0.06Emax under normal conditions, or
Edf, min  0.06Emax (4.18)
When there is a heat loss from regulatory sweating Ersw, the diffusion evaporative heat loss
Edif, Btu/h ft2 (W/m2), for the portion of skin surface that is not covered with sweat can be
hs 
1
Rt

fcl h
fcl hRcl  1
im LR 
he,c
hs
4.8 CHAPTER FOUR
TABLE 4.3 Moisture Permeability of Clothing Ensembles
Ensemble description im
Cotton/ polyester long-sleeve shirt, long
trousers, street shoes, socks, briefs 0.385
Cotton short-sleeve shirt, long trousers, work
boots, socks, briefs, cotton gloves 0.41
Cotton/nylon long-sleeve shirt, cotton/nylon
trousers, combat boots, socks, helmet liner
(army battle dress uniform) 0.36
*Measured with Trad  Ta, and air velocity  40 fpm.
Source: Adapted with permission from ASHRAE Handbook
1989, Fundamentals.
calculated as
Edif  (1  wrsw)0.06Emax (4.19)
Therefore, the total evaporative heat loss from the skin surface is
Esk  Ersw  Edif
 wrswEmax  (1  wrsw)0.06Emax  (0.06  0.94wrsw)Emax
 wskEmax (4.20)
In Eq. (4.20), wsk is called the total skin wetness; it is dimensionless, and it can be calculated as
(4.21)
4.5 MEAN RADIANT TEMPERATURE AND EFFECTIVE
TEMPERATURE
Mean Radiant Temperature
Mean radiant temperature TRad is defined as the temperature of a uniform black enclosure in which
an occupant would have the same amount of radiative heat exchange as in an actual indoor environment.
Mean radiant temperature TRad, R(K), can be calculated by the expression
T 4
Rad  T 4R1F0  1  T 4
R2 F0 2  T 4
RnF0  n (4.22)
where TR1, TR2,   , TRn  absolute temperature of surrounding surfaces of indoor
environment, R(K)
F0  1  shape factor denoting fraction of total radiant energy leaving surface
of occupants clothing 0 and arriving on the surface 1
F0  2  fraction of total radiant energy leaving surface 0 and arriving on
surface 2, etc.
Shape factors F0  n depend on the position and orientation of the occupant as well as the dimensions
of the enclosure. One can use Figs. 4.2 and 4.3 to estimate the mean value of the shape factor between
a seated person and rectangular surfaces. The sum of the shape factors of all the surfaces
with respect to the seated occupant in an enclosure is unity.
The temperature measured by a globe thermometer, called the globe temperature, is often used
to estimate the mean radiant temperature. The globe thermometer consists of a copper hollow
sphere of 6-in. (152-mm) diameter that is coated with black paint on the outer surface. A precision
thermometer or thermocouple is inserted inside the globe with the sensing bulb or the thermojunction
located at the center of the sphere. Because the net radiant heat received at the globe surface is
balanced by the convective heat transfer from the globe surface in reaching a thermal equilibrium,
according to Bedford and Warmer (1935), such a relationship gives
T 4
Rad  T 4
Rg  0.247  109 v0.5(TRg  TRa) (4.22a)
where TRad  absolute mean radiant temperature, R (K)
TRg  absolute globe temperature, R (K)
v  ambient air velocity, fpm (m/s)
TRa  absolute air temperature, R (K)
After TRg, v, and TRa are measured, the mean radiant temperature TRad can be calculated from Eq.
(4.22). The mean radiant temperature indicates the effect, due to the radiant energy from the
wsk 
Esk
Emax
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.9
4.10 CHAPTER FOUR
FIGURE 4.2 Mean value of shape factor between a sedentary person and a horizontal plane. (Source: P. O. Fanger,
Thermal Comfort Analysis and Applications in Environmental Engineering, 1972. Reprinted with permission.)
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.11
FIGURE 4.3 Mean value of shape factor between a seated person and a vertical plane. (Source: P. O. Fanger,
Thermal Comfort Analysis and Applications in Environmental Engineering, 1970. Reprinted with permission.)
surroundings, on radiant exchange between an occupant or any substance and the enclosure. Such
an influence may be significant if the mean radiant temperature is several degrees higher than the
temperature of the indoor air.
Example 5.1. The dimensions of a private office and the location of a person seated within it are
shown in Fig. 4.4. The surface temperatures of the enclosure are as follows:
West window 88F (31.1C)
West wall 80F (26.7C)
North partition wall 75F (23.9C)
East partition wall 75F (23.9C)
South partition wall 75F (23.9C)
Floor 78F (25.6C)
Ceiling 77F (25C)
Calculate the mean radiant temperature of the enclosure that surrounds this office. The orientation
of the seated occupant is unknown.
Solution
1. Regarding the north partition wall, the shape factor denotes the fraction of the total radiant energy
that leaves the outer surface of the clothing of the occupant (surface 0) and arrives directly on
the north partition wall (surfaces 1, 2, 3, and 4) and is given by
F0  1,2,3,4  F0  1  F0  2  F0  3  F0  4
Here F0  1 is the shape factor for the radiation from surface 0 to surface 1 of the north partition wall.
Based on the curves in Fig. 4.3, for a ratio of b/L  1.8/3  0.6 and a ratio of a/L  4.5/3  1.5,
4.12 CHAPTER FOUR
FIGURE 4.4 Dimension, in feet, of a private office.
the shape factor F0 1  0.04. Here L is the horizontal distance from the occupant to the north partition
wall. The shape factors F0  2, F0  3, and F0  4 can be calculated in the same manner.
2. To determine the shape factor F0  5 for the radiation from the occupant (surface 0) to the floor
(surface 5), we note that the vertical distance L from the center of the seated occupant to the floor is
1.8 ft. From the curves in Fig. 4.2, the ratio b/L  3/1.8  1.67, and the ratio a/L  4.5/1.8 
2.5; thus the shape factor F0  5  0.068. All the remaining shape factors can be determined in the
same manner as listed in Table 4.4.
3. For the north partition wall (surface 1),
T4
R1F0  1  (75  460)4  0.04  32.77  108
Other products T 4
Rn F0  n can be similarly calculated, as listed in Table 4.4. The sum of the products
T4
Rn F0 n  832.56  108. Therefore
T 4
Rad  832.56  108
That is,
TRad  537.2R or Trad  77.2F(25.1C)
4. The sum of the shape factors F0  n  0.994 is nearly equal to 1.
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.13
TABLE 4.4 Values of F0  n and T4
RnF0  n in Example 4.1
Surface Shape
Surface temperature, R factor b/L a/L F0 n T 4
RnF0 n  108
North partition wall 535 F0  1 0.6 1.5 0.04 32.77
F0  2 0.6 3.5 0.045 36.87
F0  3 2.2 1.5 0.07 57.35
F0  4 2.2 3.5 0.087 71.27
East partition wall 535 F0  17 0.63 0.29 0.014 11.47
F0  18 0.63 0.86 0.03 24.58
F0  19 0.17 0.29 0.004 3.28
F0  20 0.17 0.86 0.009 7.37
South partition wall 535 F0  9 0.73 0.5 0.023 18.84
F0  10 0.73 1.17 0.038 31.13
F0  11 0.2 0.5 0.008 6.55
F0  12 0.2 1.17 0.013 10.65
West wall 540 F0  13 0.4 0.67 0.018 15.31
540 F0  14 0.4 2 0.03 25.51
West window 548 F0  15 1.6 0.67 0.04 36.07
548 F0  16 1.6 2 0.07 63.13
Floor 538 F0  5 1.67 2.5 0.068 56.97
F0  6 1.67 5.8 0.073 61.16
F0  7 5 2.5 0.087 72.89
F0  8 5 5.8 0.102 85.45
Ceiling 537 F0  21 1.36 0.68 0.033 27.44
F0  22 1.36 1.59 0.052 43.24
F0  23 0.45 0.68 0.015 12.47
F0  24 0.45 1.59 0.025 20.79
 0.994 8.32.56  108
Effective Temperature
The effective temperature ET* is the temperature of an environment that causes the same total heat
loss from the skin surface as in an actual environment of an operative temperature equal to ET* and
at a relative humidity of 50 percent. And ET* can be calculated as
ET*  To  wsk im LR (0.5pET, s) (4.23)
where pET, s  saturated water vapor pressure at ET*, psia (kPa abs.). The right-hand side of
Eq. (4.23) describes the conditions of the indoor air regarding the total heat loss from the human
body. The same value of the combination To  wsk imLR(0.5pET, s ) results in the same amount of total
heat loss from the skin surface, if other parameters remain the same. Theoretically, total skin
wetness wsk and clothing permeability index im are constants for a specific ET* line.
Because the effective temperature is based on the operative temperature To, it is a combined index
of Ta, Trad, and pa. In an indoor air temperature below 77F (25C), the constant-ET* lines are
nearly parallel to the skin temperature lines for sedentary occupants with a clothing insulation of
0.6 clo; therefore, ET* values are reliable indexes to indicate thermal sensations at normal indoor
air temperature during low activity levels. The term effective temperature was originally proposed
by Houghton and Yaglou in 1923. A new definition of ET* and its mathematical expression were
developed by Gagge et al. in 1971. It is the environmental index commonly used in specifying and
assessing thermal comfort requirements.
4.6 FACTORS AFFECTING THERMAL COMFORT
Daily experience and many laboratory experiments all show that thermal comfort occurs only under
these conditions:
1. There is a steady-state thermal equilibrium between the human body and the environment; i.e.,
heat storage of the body core Scr and the skin surface Ssk are both equal to zero.
2. Regulatory sweating is maintained at a low level.
From the heat balance equation at steady-state thermal equilibrium Eq. (4.1) we have
M  W  C  R  Esk  Eres
Let Eres  0.1M and the mechanical efficiency 	 0.05M. From Eq. (4.11), C  R can be determined.
Also, from Eqs. (4.15), (4.16), and (4.20), Esk is a known value. If we substitute into
Eq. (4.1), the heat balance equation at steady-state thermal equilibrium can be expressed as
M(1  0.05  0.1)  Fcl fclh(Tsk  To)  wsk im LR hs (psk, s  pa)
or
0.85M  Fcl fcl h(Tsk  To)  wsk imLRhs ( psk, s  pa) (4.24)
In Eq. (4.24), the physiological and environmental factors that affect the balancethe metabolic
rate and the heat losses on the two sides of the equationare as follows:
1. Metabolic rate M determines the magnitude of the heat energy that must be released from the
human body, i.e., the left-hand side of the equation.
2. Indoor air temperature Ta is a weighted component of the operating temperature To. It also
affects the sensible heat loss and the vapor pressure of indoor air pa in the calculation of the
evaporative loss from the skin surface.
4.14 CHAPTER FOUR
3. Mean radiant temperature Trad is another weighted component of the operating temperature To. It
affects the sensible heat loss from the human body.
4. Relative humidity of the ambient air a is the dominating factor that determines the difference
psk, s  pa in the evaporative loss from the skin surface. Air relative humidity becomes important
when the evaporative heat loss due to regulatory sweating is the dominating heat loss from the
human body.
5. Air velocity va influences the heat transfer coefficient h and the clothing efficiency Fcl in the
term in Eq. (4.24) for the sensible heat loss from the human body. It also affects the overall sensible
heat transfer coefficient hs in the evaporative heat loss term and the clothing permeability
im term in Eq. (4.24).
6. Clothing insulation Rcl affects the clothing efficiency Fcl, the area ratio fcl, the heat transfer coef-
ficient h, the clothing permeability index im, and the overall sensible heat-transfer coefficient hs.
4.7 THERMAL COMFORT
Thermal comfort is defined as the state of mind in which one acknowledges satisfaction with regard
to the thermal environment. In terms of sensations, thermal comfort is described as a thermal sensation
of being neither too warm nor too cold, defined by the following seven-point thermal sensation
scale proposed by ASHRAE:
3  cold
2  cool
1  slightly cool
0  neutral
1  slightly warm
2  warm
3  hot
Fangers Comfort Equation
A steady-state energy balance is a necessary condition for thermal comfort, but is not sufficient by
itself to establish thermal comfort. Fanger (1970) calculated the heat losses for a comfortable person,
experiencing a neutral sensation, with corresponding skin temperature Tsk and regulatory
sweating Ersw. The calculated heat losses L are then compared with the metabolic rate M. If L  M,
the occupant feels comfortable. If L M, then this person feels cool; and if L 	 M, then this person
feels warm.
Using the responses of 1396 persons during laboratory experiments at Kansas State University
of the United States and Technical University of Denmark, Fanger developed the following equation
to calculate the predicted mean vote (PMV) in the seven-point thermal sensation scale:
PMV  (0.303e0.036M  0.276)(M  L) (4.25)
In Eq. (4.25), the metabolic rate M and heat losses L are both in Btu/h ft2 (W/m2). According to
Fangers analysis, the predicted percentage of dissatisfied (PPD) vote for thermal comfort at a
PMV  0 is 5 percent, and at a PMV  1 is about 27 percent.
Tables of PMV and comfort charts including various combinations of operating temperature To,
air velocity v, metabolic rate M, and clothing insulation Rcl have been prepared to determine comfortable
conditions conveniently. Fangers comfort charts also include relative humidity. Six of his
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.15
4.16
FIGURE 4.5 Fangers comfort charts. (Abridged with permission from ASHRAE Handbook 1981, Fundamentals.)
comfort charts at various activity levels, wet-bulb temperatures, relative humidities, and air velocities
are shown in Fig. 4.5. In the first four charts, the air temperatures are equal to the mean radiant
temperatures. Three of these four have a clothing insulation of 0.5 clo. The other is for 1-met
activity level and 1 clo, because one rarely finds an occupant with such heavy outwear at an activity
level of 2 or 3 met. In the fifth and sixth charts, the air temperature could be different from the
mean radiant temperature with a constant relative humidity of 50 percent. These comfort variations
clearly show that all six factorsair temperature Ta , mean radiant temperature Trad, relative humidity
, air velocity v, metabolic rate M, and clothing insulation Rclseriously affect the thermal
comfort.
For example, from Fangers comfort chart, a sedentary occupant at an activity level of 1.0 met,
with a clothing insulation of 0.5 clo, in an air conditioned space at a relative humidity of 50 percent
and an air velocity less than 20 fpm (0.1 m/ s), feels comfortable with an air temperature equal to
the mean radiant temperature of 78F (25.6C). If all values were identical except for a 2-met activity
level, the temperature would need to be 67F (19.4C) for the same level of comfort.
Another factor, the duration of exposure to the indoor thermal environment, should be discussed
here. If an indoor environment can provide thermal comfort for the occupant, the duration of the
exposure has no significant influence upon the physiological responses of the persons thermal regulatory
mechanism. If the indoor environment is uncomfortable, subjecting the subject to a certain
degree of heat or cold stress, the time exposure will influence the persons physiological response.
ASHRAE Comfort Zones
Based on results of research conducted at Kansas State University and at other institutions,
ANSI/ASHRAE Standard 55-1992 specified winter and summer comfort zones to provide for the
selection of the indoor parameters for thermal comfort (see Fig. 4.6). This chart is based upon an
occupant activity level of 1.2 met (69.8 W/m2). For summer, typical clothing insulation is 0.5 clo,
that is, light slacks and short-sleeve shirt or comparable ensemble; there is no minimum air speed
that is necessary for thermal comfort. Standard 55-1992 recommended a summer comfort zone with
an effective boundary temperature ET*  73 to 79F (22.5 to 26C) at 68F (20C) wet-bulb as its
upper-slanting boundary and dew-point temperature 36F (2.2C) as its bottom flat boundary. If the
clothing insulation is 0.1 clo higher, the boundary temperatures both should be shifted 1F (0.6C)
lower. Rohles et al. (1974) and Spain (1986) suggested that the upper boundary of the summer
comfort zone can be extended to 85 or 86F (29.4 or 30C) ET* if the air velocity of the indoor air
can be increased to 200 fpm (1 m/ s) by a ceiling fan or other means.
The winter comfort zone is based upon a 0.9-clo insulation including heavy slacks, long-sleeve
shirt, and sweater or jacket at an air velocity of less than 30 fpm (0.15 m/s). Standard 55-1992 recommended
a winter comfort zone with an effective boundary temperature ET*  68 to 74F (20 to
23.3C) at 64F (17.8C) wet-bulb as its slanting upper boundary and at dew-point 36F (2.2C) as
its bottom flat boundary.
Indoor air parameters should be fairly uniform in order to avoid local discomfort. According to
Holzle et al., 75 to 89 percent of the subjects tested found the environment within this summer
comfort zone to be thermally acceptable.
ASHRAE comfort zones recommend only the optimal and boundary ET* for the determination
of the winter and summer indoor parameters. For clothing insulation, activity levels, and indoor air
velocities close to the values specified in Standard 55-1992, a wide range of indoor design conditions
are available.
Comfort-Discomfort Diagrams
A comfort diagram provides a graphical presentation of the total heat loss from the human body at
various operative or air temperatures and indoor relative humidities when the activity level, level of
clothing insulation, and air velocity are specified. The abscissa of the comfort diagram is the opera-
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.17
tive temperature To or ambient air temperature Ta, whereas the ordinate can be either water vapor
pressure p or humidity ratio w. Figure 4.7 shows a comfort diagram with a sedentary activity level,
a clothing insulation value of 0.6 clo, and still-air conditions, i.e., an air velocity v 	 20 fpm
(0.1 m/ s). The curved lines represent relative humidity, and the straight lines represent effective
temperature ET*. The short dash curves diverging from the ET* lines are total skin wetness wsk
lines. The figure is based upon To  Ta.
4.18 CHAPTER FOUR
Winter
Summer
w, lb/lb
Dew-point temperature Tdew, F
68 F wet bulb
64 F wet bulb
Operative temperature T0, F
70
60
50
40
30
ET* 68 F
73
74
79
20
10
60 70 80 90
0
0.005
0.010
0.015
100%
50%
30%
70%
60%
FIGURE 4.6 ASHRAE comfort zones. (Adapted with permission from ANSI/ASHRAE Standard, 551992.)
Effective temperature ET* lines are calculated according to Eq. (4.23). Because the total skin
wetness wsk is a constant for a specific ET* below 79F ET*, the ET* lines and wsk lines coincide
with one other. At higher ET* values, wsk lines curve to the left at high relative humidities. For low
ambient air temperatures, evaporative heat loss from the skin surface Esk is small; therefore, ET*
and wsk lines are nearly vertical. As Esk becomes greater and greater, the slopes of the ET* and wsk
lines decrease accordingly.
The comfort diagram is divided into five zones by the ET* and wsk lines:
1. Body cooling zone. For the condition given in Fig. 4.7, if the effective temperature ET* 	
73F (22.8C), the occupant will feel cold in this zone. Because the heat losses exceed the net metabolic
rate, the skin and body core temperatures tend to drop gradually.
2. Comfort zone. This is the zone between the lower boundary ET*  73F (22.8C), and wsk 
0.06, and the higher boundary ET*  86F (30C), and wsk  0.25. Steady-state thermal equilibrium
is maintained between the occupant and the environment, and regulatory sweating is at a
rather low level. The occupant will feel comfortable in this zone, and the heart rate (HR) is between
76 and 87 beats per minute.
ASHRAEs winter and summer comfort zones are a part of this zone. The lower boundary of the
ASHRAE winter comfort zone forms the lower boundary of this comfort zone. The reason that the
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.19
FIGURE 4.7 Comfort-discomfort diagram.
lower boundary in this diagram is ET*  73F (22.8C) whereas ET*  68F (20C) in
ASHRAEs winter comfort zone is that a lower clothing insulation of 0.6 clo is used here.
3. Uncomfortable zone. In this zone, 86F 	 ET*  95F (30C  ET*  35C) and 0.25 	
wsk  0.45. Thermal equilibrium also exists between the occupant and the environment, and the
evaporative heat loss due to regulatory sweating dominates. Heart rate shows a range between 87
and 100. The occupant feels uncomfortable, i.e., warm or hot, when his or her physiological parameters
are in these ranges.
4. Very uncomfortable zone. In this zone, 95F 	 ET* 	 106.5F (35C 	 ET* 	 41.4C) and
0.45 	 wsk 	 1. Although thermal equilibrium is still maintained with zero heat storage at the skin
and the body core, there is a danger of a heat stroke when ET* 95F (35C). Toward the upper
boundary of this zone, the skin surface is nearly entirely wet, and the heart rate exceeds 120. Under
these conditions, the occupant will feel very hot and very uncomfortable.
5. Body heating zone. When ET*  106.5F (41.4C) and wsk  1, thermal regulation by evaporation
fails. At a higher ET* or wsk, the environment is intolerable, and the temperatures of the
body core and skin tend to rise gradually.
For an air conditioned space with an occupant at low activity levels (M 	 2 met), the indoor
environment is usually maintained within the comfort zone, and the physiological and thermal responses
of the occupant are also in the comfort zone. Only at higher activity levels do the thermal
responses occasionally fall into the discomfort zone.
4.8 INDOOR AIR TEMPERATURE AND AIR MOVEMENTS
Comfort Air Conditioning Systems
For comfort air conditioning systems, most occupants have a metabolic rate of 1.0 to 1.5 met. The
indoor clothing insulation in summer is usually 0.35 to 0.6 clo, and in winter it is 0.8 to 1.2 clo.
Relative humidity has a lesser influence on thermal comfort, and will be discussed in the next section,
but indoor air temperature and air velocity are discussed here.
Many researchers have conducted tests to determine the effects of airspeed on the preferred indoor
air temperature and the thermal comfort of occupants. The relationship between the preferred
indoor air temperatures and various airspeeds is presented in Fig. 4.8. Most of the data were taken
under these conditions: metabolic rate M  400 Btu/h (117 W), clothing insulation Rcl  0.6 clo,
Trad  Ta, and relative humidity of the indoor air   50 percent. The one exception is the students
in Holzles experiments, who had 0.54 clo for summer and 0.95 clo for winter. Examination of
Fig. 4.8 shows the following:
 Higher indoor air temperature requires greater indoor air velocity to provide thermal comfort.
 Variation of airspeed has a greater influence on preferred indoor air temperature at lower air
temperatures.
ANSI/ASHRAE Standard 55-1992 recommended that within the thermally acceptable temperature
ranges in the ASHRAE summer and winter comfort zones discussed in Sec. 4.7, there be no minimum
airspeed (nondirectional) that is necessary for thermal comfort.
If temperature is increased above the level allowed for the comfort zone, means must be provided
to elevate the airspeed. For instance, when indoor air temperature Ta  Trad (mean radiant
temperature), given that the airspeed for a summer comfort zone of 79F (26C) is 40 fpm (0.2
m/ s), if Ta has an increase of 2F (1.1C) from 79 to 81F (26 to 27.2C), according
to ANSI/ASHRAE Standard 55-1992, there must be a relevant increase of airspeed of about
70 fpm (0.35 m/ s).
Before specifying Ta for summer conditions, one needs to determine whether occupants are
likely to wear suit jackets, such as members of a church congregation or guests in a multipurpose
4.20 CHAPTER FOUR
hall. In such cases, a reduction of 4F of summer optimal ET* may be necessary because of the increase
in clothing insulation of about 0.4 clo.
Design Considerations
When one is specifying indoor design conditions, thermal comfort must be provided at optimum
cost while using energy efficiently. These principles should be considered:
1. To determine the optimum summer and winter indoor design temperatures, consider the local
clothing habits and the upper and lower acceptable limits on clothing insulation at various operative
temperatures To, as shown in Fig. 4.9.
2. It is always more energy-efficient to use different indoor design temperatures for summer and
winter than a year-round constant value. An unoccupied-period setback during winter always
saves energy. There are also buildings in which a constant indoor temperature is required for the
health and comfort of the occupants, such as in many health care facilities.
3. For short-term occupancies, or when the metabolic rate is higher than 1.2 met, a strategy of using
a lower energy-use ceiling fan or a wall-mounted fan to provide higher air velocity may be
considered. Thus a higher indoor design temperature within the extended summer comfort zone
may be acceptable for occupants, especially in industrial settings.
Indoor Design Temperatures for Comfort Air Conditioning
According to ANSI/ASHRAE Standard 55-1992, Thermal Environmental Conditions for Human
Occupancy, and ASHRAE/IES 90.1-1999, Energy-Standard for Buildings Except Low-Rise
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.21
FIGURE 4.8 Preferred indoor air temperatures at various air velocities.
Residential Buildings, the following indoor design temperature and air speed apply for comfort air
conditioning systems when the activity level is 1.2 met, there is a relative humidity of 50 percent in
summer, mean airspeed  30 fpm (0.15 m/ s), and Ta  Trad:
According to ANSI/ASHRAE Standard 55-1992, within the thermally acceptable temperature
ranges, there is no minimum air speed that is necessary for thermal comfort. If the summer indoor
temperature is 79F (26C), an airspeed of 40 fpm (0.2 m/s) is recommended. If Ta 79F (26C),
a relevant increase of airspeed in the indoor occupied zone should be considered. Refer to
ANSI/ASHRAE Standard 55-1992.
If the space relative humidity can be lowered to 35 to 40 percent in summer, then the higher
limit 78F (25.5C) is often specified.
Indoor badminton and table tennis tournament arenas should have air velocities below 30 fpm
(0.15 m/ s). To avoid nonuniformity and to prevent local discomfort, the air temperature difference
between 4 in. from the floor and 67 in. above the floor should not exceed 5F (3C). The radiant
temperature asymmetry in the vertical direction should be less than 9F (5C), and in the horizontal
direction less than 18F (10C).
To determine whether the specified air temperature and airspeed are met, they should be measured
at 4-, 24-, and 43-in. (0.1-, 0.6-, and 1.1-m) levels for sedentary occupants, and 4-, 43-, and
67-in. (0.1-, 1.1-, and 1.7-m) levels for standing activity. The duration for determining the mean
value of the air movement should be 3 min. or 30 times the 90 percent response time of the measuring
instrument, whichever is greater.
Typical clothing Optimum operative Indoor design
insulation, clo temperature temperature range
Winter 0.9 71F (22C) 6974F (20.523.5C)
Summer 0.5 76F (24.5C) 7479F (23.526C)
4.22 CHAPTER FOUR
FIGURE 4.9 Relationship between clothing insulation Rcl and operating
temperature To. (Source: ASHRAE Transactions 1983, Part I B. Reptinted
with permission.)
Process Air Conditioning Systems
For process air conditioning systems, indoor design temperature is usually based on previous experiences.
For precision manufacturing projects, a basic temperature plus a tolerance, such as 72 
2F (22.2  1.1C) for a precision machinery assembling workshop, is often specified. First, the
tolerance should be neither too tight nor too loose. Second, either the temperature fluctuation at
various times within the working period or the temperature variation within the working space,
or both, should be included in this tolerance.
In unidirectional-flow clean rooms, the velocity of the airstream in the working area is often
specified as 90  20 fpm (0.45  0.1 m/ s) to prevent contamination of the products.
4.9 HUMIDITY
Comfort Air Conditioning Systems
According to ANSI/ASHRAE Standard 55-1992, for the zone occupied by people engaged in light,
primarily sedentary activity (1.2 met), the relative humidity should conform with the limits of
ASHRAE winter and summer comfort zones, as shown in Fig. 4.6. These limits are intended to
maintain acceptable thermal conditions for the occupants comfort.
ASHRAE/IESNA Standard 90.1-1999 mandates that where a zone is served by system(s) with
both humidification or dehumidification capacity, means shall be provided to prevent simultaneous
operation of humidification and dehumidification equipment. Standard 90.1-1999 also specifies that
where humidistatic controls are provided, such controls shall be capable of preventing reheating,
mixing of hot and cold air streams, and simultaneous heating and cooling. Refer to Section 29.12
for more details.
The following results should be considered during the design and evaluation of the performance
of comfort air conditioning systems:
 Maintaining an indoor space relative humidity r between 20 and 30 percent in winter prevents or
reduces the condensation at the inner side of the window glass.
 In high-occupancy applications, it may be economical and is still comfortable to specify the summer
indoor relative humidity at 55 to 60 percent if the indoor temperature is within the summer
comfort zone.
 The indoor relative humidity at part load may be considerably higher than at full load in some air
conditioning systems in summer.
 When the indoor relative humidity is below 25 percent, the incidence of respiratory infections increases
significantly. If, simultaneously, indoor temperatures are low, such as below 70F (21C),
the induced static electricity in carpeted rooms may cause uncomfortable shocks to occupants
contacting metal furniture or decorations.
 Because the increase in the outdoor ventilation rates in ASHRAE Standard 62-1999 for the air
conditioning system serves the building located in areas where the humidity ratio of outdoor air is
very low during winter, the system has greater difficulty maintaining an indoor relative humidty
of 20 to 30 percent without winter humidification.
 If a humidifier is installed, its humidifying capacity should not exceed the actual humidifying
requirements so that wet surfaces do not occur inside the air-handling unit, packaged unit, and
supply ducts. Wet surface and dirt cause the growth of microorganisms and poor indoor air quality.
Therefore, for comfort air conditioning systems, the recommended indoor relative humidity
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.23
levels are as follows:
Process Air Conditioning Systems
Humidity affects the physical properties of many materials and, therefore, their manufacturing
processes.
Moisture Content. Relative humidity has a marked influence on the moisture content of hygroscopic
materials such as natural textile fibers, paper, wood, leather, and foodstuffs. Moisture content
affects the weight of the products and sometimes their strength, appearance, and quality.
Dimensional Variation. Hygroscopic materials often extend at higher relative humidity and contract
at lower humidity. A 2 percent increase in moisture content may result in a 0.2 percent increase in
dimension of paper. That is why lithographic printing requires a relative humidity of 45  2 percent.
Corrosion and Rust. Corrosion is an electrochemical process. Moisture encourages the formation
of electrolytes and therefore the corrosion process. A relative humidity greater than 50 percent may
affect the smooth operation of bearings in precision instruments. When indoor relative humidity
exceeds 70 percent, rust may be visible on the surface of the machinery and on parts made of steel
and iron.
Static Electricity. Static electricity may cause minute particles to repel or attract one another,
which is detrimental to many manufacturing processes. Static electricity charges minute dust particles,
in the air, causing them to cling to equipment and work surfaces. Static electricity exists in an
indoor environment at normal air temperatures when relative humidity is less than about 40 percent.
Loss of Water. Vegetables and fruits lose water vapor through evaporation from their surfaces
during storage. Low temperatures and high relative humidities, such as  90 to 98 percent, may
reduce water loss and delay desiccation.
It is important to specify the exact relative humidity required for product quality and cost
control. For process air conditioning systems, the specified relative humidity is either a year-round
single value or a range. A strict relative-humidity requirement always includes a basic value and a
tolerance, such as the relative humidity for lithographic printing mentioned before. When temperature
and relative humidity controls are both required, they should be specified as a combination.
Consider this example:
Case Study 4.1. A factory workshop has the following environmental parameters during summer:
Indoor air temperature 79F (26.1C)
Indoor air relative humidity 50 percent
Temperature, F (C) Relative humidity, %
Clean room 72  2 (22.2  1.1) 45  5
Relative humidity, %
Summer 3065
Winter
Commercial and public buildings 2060
Health care buildings 3060
4.24 CHAPTER FOUR
Space air velocity 20 fpm (0.1 m/s)
Clothing insulation of the workers 0.5 clo
Activity level 3 met
As a result of comfort complaints by personnnel, you are asked to recommend effective and economical
corrective measures to improve the indoor environment. The following table includes the
information required during analysis:
Area ratio of the clothed body fcl 1.2
Permeability index of clothing im 0.4
Skin surface temperature,Tsk 92.7F (33.7C)
Saturation water vapor pressure at 92.7F 0.764 psia (5.27 kPa abs.)
at 79F 0.491 psia (3.39 kPa abs.)
Solution
1. When the space air velocity v  20 fpm, the convective heat-transfer coefficient can be calculated,
from Eq. (4.9), as
hc  0.0681v0.69  0.0681  200.69  0.538 Btu/h ft2 F
Because at normal indoor conditions the radiative heat-transfer coefficient hr  0.83 Btu/h ft2 F,
the surface heat-transfer coefficient h is
h  hc  hr  0.538  0.83  1.368 Btu/h ft2 F
The clothing insulation Rcl  0.5  0.88  0.44 h ft2 F/Btu. From Eq. (4.12), the clothing effi-
ciency is found to be
Therefore, from Eq. (4.11), we find the sensible heat loss from the skin surface of the worker
C  R  Fcl fclh(Tsk  To)  0.5806  1.2  1.368(92.7  79)
 13.06 Btu/h ft2
From Eq. (4.17), the overall sensible heat-transfer coefficient hs is
From Eq. (4.16), the overall evaporative heat-transfer coefficient he,c is
he,c  hs im LR  0.9531  0.4  205  78.16 Btu/h ft2  psi
The maximum evaporative heat loss Emax due to regulatory sweating can be calculated from
Eq. (4.15) as
Emax  he,c(psk, s  pa)  78.16(0.764  0.5  0.491)
 40.24 Btu/h ft2

1.2  1.368
1.2  1.368  0.44  1
 0.9531 Btu/h ft2F
hs 
fclh
fclhRcl  1
Fcl 
1
Rcl fclh  1

1
0.44  1.2  1.368  1
 0.5806
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.25
In Eq. (4.24),
0.85M  0.85  3  18.46  47.07 Btu/h ft2
When 	 0.05 and Eres  0.1M, the total evaporative loss from the skin surface of the worker,
from Eq. (4.1), is
Esk  0.85M  (C  R)  47.07  13.06  34.01 Btu/h ft2 (107.3 W/m2)
Therefore, the total skin wetness is
In Fig. 4.7, when wsk  0.845, a person is very uncomfortable.
2. From Fangers comfort chart, shown in Fig. 4.5, for a person with an activity level of 3 met
and a clothing insulation of 0.5 clo, at a relative humidity of 50 percent and an air velocity of 20
fpm and with Ta  Trad for neutral thermal sensation, the indoor air temperature should be 56F
(13.3C). Obviously, this is not economical because too much refrigeration is required.
3. Let us analyze the results if ceiling fans or wall fans are used to increase the space air velocity
v to 300 fpm (1.5 m/ s). Then the convective heat-transfer coefficient hc is
hc  0.681  3000.69  3.486 Btu/h ft2 F
The surface heat-transfer coefficient is
h  3.486  0.83  4.316 Btu/h ft2 F
Also the clothing efficiency is calculated as
Then the sensible heat loss is equal to
C  R  0.305  1.2  4.316(92.7  79)  21.64 Btu/h ft2
The overall sensible heat-transfer coefficient is
The overall evaporative heat-transfer coefficient he,c can be shown to be
he,c  1.58  0.4  205  129.5 Btu/h ft2  psi
Then the maximum evaporative heat loss due to regulatory sweating is
Emax  129.5(0.7604  0.5  0.491)  66.68 Btu/h ft2 (210.3 W/m2)
The total skin wetness is
wsk  47.07  21.64/66.68  0.381
That is, wsk has been greatly reduced compared with the value at v  20 fpm (0.1 m/ s). In Fig. 4.7,
wsk  0.381 is in the uncomfortable zone. Workers will feel warm, but the indoor environment has
been considerably improved. This may be the most cost-effective solution.
hs 
1.2  4.316
1.2  4.136  0.44  1
 1.58 Btu /h ft2F
Fcl 
1
0.44  1.2  4.316  1
 0.305
wsk 
Esk
Emax

34.01
40.24
 0.845
4.26 CHAPTER FOUR
4.10 SICK BUILDING SYNDROME AND INDOOR AIR QUALITY
Sick building syndrome is a kind of building-related illness that has received public attention since
the 1970s. ASHRAE (1987) defined the sick building as . . . a building in which a significant
number (more than 20 percent) of building occupants report illness perceived as being building related.
This phenomenon, also known as sick building syndrome is characterized by a range of
symptoms including, but not limited to, eye, nose, and throat irritation, dryness of mucous membranes
and skin, nose bleeds, skin rash, mental fatigue, headache, cough, hoarseness, wheezing,
nausea, and dizziness. Within a given building, there usually will be some commonality among the
symptoms manifested as well as temporal association between occupancy in the building and appearance
of symptoms. If there are signs of actual illnesses, these illness are classified as buildingrelated
illnesses.
Poor indoor air quality (IAQ) is the dominant factor that causes sick building syndrome. Indoor
air quality is defined as an indication of harmful concentrations of the indoor air contaminants that
affect the health of the occupants or the degree of satisfaction of a substantial majority (80 percent
or more) of occupants exposed to such an indoor environment. Poor control of the indoor air temperature
and relative humidity are causes of discomfort symptoms. They may also increase the
indoor air contaminants. However, unsatisfactory indoor temperature and indoor relative humidity
are only indirect causes of poor indoor air quality.
National Institute for Occupational Safety and Health (NIOSH) of the United States (1989),
according to the results of 529 building investigations between 1971 and 1988, and Health and
Welfare Canada (HWC), according to the results of 1362 building investigations between 1984
and 1989, classified the reasons for sick building syndrome as follows:
Inadequate ventilation includes lack of outdoor air, poor air distribution, poor thermal control, and
inadequate maintenance; and it is the primary cause of indoor air quality. The survey found 70 to 80
percent of the investigated buildings had no known problems. Effective operation and control of the
HVAC&R system will be discussed in later chapters.
In the United States, most people spend about 90 percent of their time indoors. The purpose of
specifying the indoor design conditions in the design documents is to provide the occupants with a
satisfactory indoor environment at optimum cost.
After the energy crisis in 1973, a lower outdoor ventilation rate, a tighter building shell, and the
use of variable-air-volume (VAV) systems at part-load operation may reduce the amount of outdoor
air intake significantly. Indoor air quality therefore has become one of the critical HVAC&R problems
especially in commercial buildings since the 1980s.
Indoor Air Contaminants
Based on the results of the field investigations of three office buildings by Bayer and Black in 1988,
the indoor air contaminants that relate to indoor air quality and the symptoms of the sick building
syndrome are mainly the following:
NIOSH, 529 Buildings HWC, 1362 Buildings
No. of buildings Percent No. of buildings Percent
Inadequate ventilation 280 53 710 52
Indoor contaminants 80 15 165 12
Outdoor contaminants 53 10 125 9
Biological contaminant 27 5 6 0.4
Building fabric contamination 21 4 27 2
Unknown sources 68 13 329 24
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.27
1. Total particulate concentration. This parameter includes particulates from building materials,
combustion products, and mineral and synthetic fibers. In February 1989, the U.S. Environmental
Protection Agency (EPA) specified the allowable indoor concentration level of particulates of
10 	m and less in diameter (which can penetrate deeply into the lungs, becoming hazardous to
health) as follows:
50 	g/m3 (0.000022 gr/ ft3): 1 year
150 	g/m3 (0.000066 gr/ ft3): 24 h
According to ASHRAE Handbook 1997, Fundamentals, particles less than 2 	m in diameter are
most likely retained in the lungs, and particles less than 0.1 to 0.5 	m in diameter may leave the
lungs with the exhaled breath. Particles larger than 8 to 10 	m in diameter are separated and retained
in the upper respiratory tract. Particles between 2 and 8 	m in diameter are deposited mainly
in the conducting airways of the lungs and are swallowed or coughed out quickly.
2. Combustion products. Carbon monoxide (CO) is a colorless, odorless gas, a product of incomplete
combustion. CO interferes with the delivery of oxygen throughout the body. NO2 is a
combustion product from gas stoves and other sources. There is growing evidence that NO2 may
cause respiratory disease. Indoor concentrations for CO and NO2 are the same as specified in the
National Primary Ambient-Air Quality Standard by the EPA later in this section.
3. Volatile organic compounds (VOCs). These include formaldehyde and a variety of aliphatic,
aromatic, oxygenated, and chlorinated compounds. Mucous membrane irritation caused by
formaldehyde is well established. U.S. Department of Housing and Urban Development (HUD)
specifies a target level of indoor concentration of formaldehyde for manufacturing homes of
0.4 ppm.
4. Nicotine. Environmental tobacco smoke is clearly a discomfort factor to many adults who do
not smoke. Nicotine and other components of tobacco smoke are also a health risk for human beings.
5. Radon. Radon is a colorless, odorless, inert radiative gas widely found in soil, rocks, and water,
created by the decay of the radium and uranium. It travels through the pores of rock and soil and in-
filtrates into a house along cracks and other openings in the basement slab or walls; pressure-driven
radon containing soil gas is caused by thermal stack, wind, and the mechanical ventilation system.
The annual average concentration of radon in residential buildings in the United States is about
1.25 pCi/L. Only about 6 percent of U.S. homes have an annual average radon concentration exceeding
the EPA recommended annual average indoor concentration of 4 pCi/L for residential and
school occupancies. Pressure-driven flow of radon containing soil gas is the primary source for elevated
concentrations. At various locations in the United States, the indoor radon concentrations may
vary hourly, daily, and seasonally, sometimes by as much as a factor of 10 to 20 on a daily basis.
The radiative decay of radon produces a series of radioactive isotopes called progeny. These progeny
are chemically active. They can deposit directly into the lung, or attach to airborne particles and
then deposit into the lung. Some of the progeny are alpha particle emitters and may lead to cellular
changes and initiate lung cancer.
6. Occupant-generated contaminants and odors. These include odors and emissions (bioeffluents)
from the human body, particulates, and other contaminants.
7. Bioaerosols. These contaminants include bacteria, mold and midew, viruses, and pollens.
Bacteria and viruses are airborne, carried by dust or transmitted by people and animals; standing
water (wet surface) and dirt (nutrients) can become the breeding ground for mold, mildew, and
other biological contaminants. Pollens originate from plants.
In addition to the preceding indoor air contaminants, others such as sulfur dioxide and ozone
can have a significant effect on occupants. Carbon dioxide (CO2) is a kind of gas released from
human beings and is not an indoor contaminant at the concentrations found in most buildings.
ASHRAE Standard 62-1999 specifies guidelines for indoor concentration for ozone during continuous
exposure time as 100 	g/m3.
4.28 CHAPTER FOUR
Basic Strategies to Improve Indoor Air Quality
There are three basic strategies to improve indoor air quality: control the contaminated source,
remove air contaminants from the indoor air by air cleaner, and use outdoor ventilation air to dilute
the concentrations of indoor air contaminants.
To eliminate or to reduce the emissions of air contaminants from the contaminated source is often
the most effective way to improve the IAQ as well as periodically cleaning the ducts interior
surface and coils condensate pan and using building materials and carpets that do not release or release
only negligible volatile organic compounds and dust. However, emissions and odor released
from occupants are difficult to eliminate or to reduce, and smoking is now prohibited in many public
places and limited to specified areas in many commercial buildings.
The volatile organic compounds and combustion products contain numerous minute particles of
size between 0.003 and 1 	m. Only high-efficiency air filters and activated carbon filters can remove
these minute particles and odors from the airstream effectively. High efficiency air filters and
carbon filters are expensive to install, operate, and maintain.
Adequate outdoor ventilation air to dilute the air contaminations in practice has been proved an
essential, practical, and cost-effective means to improve the indoor air quality. ASHRAE Standard
62-1999, Ventilation for Acceptable Indoor Air Quality, specifies two alternative procedures to obtain
acceptable IAQ: the ventilation rate procedure and indoor air quality procedure.
 In the ventilation rate procedure, acceptable indoor air quality is achieved by providing ventilation
air of specified quality and quantity to the space.
 In the IAQ procedure, acceptable air quality is achieved within the space by controlling known
and specifiable contaminants.
The ventilation rate and IAQ procedurs are discused again in Sec. 23.2.
ASHRAE Standard 62-1999 defined acceptable indoor air quality as air in which there are no
known contaminants at harmful concentrations as determined by cognizant authorities and with
which a substantial majority (80 percent or more) of the people exposed do not express dissatisfaction.
ASHRAE Standard 62-1999 defines ventilation air as that portion of supply air that is outdoor
air plus any recirculated air that has been treated for the purpose of maintaining acceptable indoor
air quality. Ventilation is the process of supplying and removing ventilation by natural and mechanical
means, and the ventilation rate means the rate of ventilation air supplied to the conditioned
space through the air system.
If outdoor air is used to dilute the concentration of indoor contaminants, its quality must meet
the National Primary Ambient-Air Quality Standard provided by the EPA. Part of the time average
concentrations are shown below:
Long-term concentration Short-term concentration
Average period Average period
Pollutants 	g/m3 ppm of exposure 	g/m3 ppm of exposure
Particulate matter 50 1 year 150 24 hours
Sulfur oxides 80 0.03 1 year 365 0.14 24 hours
Carbon monoxide 40,000 35 1 hour
10,000 9 8 hours
Nitrogen dioxide 100 0.055 1 year
Oxidants (ozone) 235 0.12 1 hour
Lead 1.5 3 months
Only particulate matter is expressed in annual geometric means; the other two are annual arithmetic
means. For carbon monoxide and ozone, both values are not to be exceeded more than once a year.
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.29
Outdoor Air Requirements for Occupants
For both comfort and process air conditioning systems, outdoor air is required to do the following:
 To meet metabolic requirements of the occupants
 To dilute the indoor air contaminants, odors, and pollutants to maintain an acceptable indoor air
quality
 To support any combustion process or replace the amount of exhaust air required in laboratories,
manufacturing processes, or restrooms
 To provide makeup for the amount of exfiltrated air required when a positive pressure is to be
maintained in a conditioned space
The amount of outdoor air required for metabolic oxidation processes for occupants is actually
rather small. ASHRAE Standard 62-1999 noted that where only dilution ventilation is used
to control indoor air quality, a CO2 indoor-to-outdoor differential concentration is
not greater than about 700 ppm, and the CO2 production of a sedentary occupant who is eating
a normal diet is 0.0106 cfm (0.3 L / min). The amount of outdoor air required for each indoor
occupant can be calculated as
(4.26)
Thayer (1982), using data from different authors, developed a dilution index that indicates
15 cfm (7 L/s) of outdoor air per person will satisfy more than 80 percent of the occupants in
the space.
Usually, for comfort air conditioning systems, the same outdoor air used for the dilution of the
concentration of air contaminants including human bioeffluents is sufficient for the metabolic oxygen
requirement, for exhausting air from restrooms, and for replacing exfiltrated air lost from the
conditioned space as a result of positive pressure. If the outdoor air supply is used to dilute the concentration
of a specific indoor air contaminant, the rate of outdoor air supply in cfm (L/min),
can be calculated as
(4.27)
where  rate of generation of contaminants in space, mg/ s
Ci, Co  concentrations of air contaminants indoors and outdoors, respectively, mg/m3
Values of Co can be found from the EPA National Primary Ambient-Air Quality Standards. The indoor
concentration of CO2 and other contaminants should meet the specified value as stated before.
Some of the outdoor air requirements for ventilation, often called the ventilation rate, specified in
ASHRAE Standard 62-1999 are indicated in Table 4.5. For clean rooms, Federal Standard 209B
specifies the rate of outdoor air, or makeup air, to be 5 to 20 percent of the supply air.
When only dilution ventilation is used to control indoor air quality and CO2 is used as an indicator
of human bioeffluents, ASHRAE Standard 62-1999 noted that an indoor-outdoor differential
concentration not greater than about 700 ppm of CO2 indicates that comfort (odor) criteria related
to bioeffluents are likely to be satisfied. The CO2 concentrations in outdoor air typically are between
300 to 350 ppm. Using CO2 as an indicator of bioeffluents does not elimnate the need for
consideration of other contaminants.
The refrigeration capacity required to cool and dehumidify the outdoor air can be a major porm
? par
V?
o 
2118m? par
Ci  Co
V?
o,
 15 cfm (7 L / s)
V?
o,oc 
V?
CO2
Ci,CO2  Co

0.0106
0.0007
V?
o, oc
V?
CO2
(Ci,CO2  Co)
4.30 CHAPTER FOUR
tion of the total refrigeration requirement during summer, depending upon the occupant density and
the amount of exhaust air. The amount of infiltration depends on the wind speed and direction, as
well as the outdoor and indoor temperatures and pressure differences, which are variable. Therefore,
infiltration is not a reliable source of outdoor air supply. Infiltration cannot replace specified
outdoor air ventilation requirements.
4.11 AIR CLEANLINESS
The manufacturing process of semiconductors, pharmaceutical, aerospace, and operating rooms in
health care facilities need clean indoor environment, clean spaces, and clean rooms. A clean room is
a constructed enclosed area in which air cleanliness is expressed in terms of particle count of air
contaminants and in which the associated temperature, humidity, air pressure, and lighting are controlled
within specific limits. A clean space is a defined area in which air cleanliness and environmental
conditions are controlled within specific limits.The quality of their manufactured products is
closely related to the size and number of particulates contained in the space air. An indoor air quality
of allowable total particulate annual-average concentration of 50 	g/m3 cannot meet the air
cleanliness requirements. Therefore, Federal Standard (FS) 209E specifies the following classes for
clean spaces and clean rooms:
Class 1. Particle count not to exceed 1 particle/ ft3 (35 particles /m3) of a size of 0.5 	m and
larger, with no particle exceeding 5 	m.
Class 10. Particle count not to exceed 10 particles/ ft3 (353 particles /m3) of a size of 0.5 	m and
larger, with no particle exceeding 5 	m.
Class 100. Particle count not to exceed 100 particles/ ft3 (3531 particles /m3) of a size of 0.5 	m
and larger.
Class 1000. Particle count not to exceed 1000 particles/ ft3 (35,315 particles/m3) of a size of
0.5 	m and larger.
Class 10,000. Particle count not to exceed 10,000 particles/ ft3 (353,150 particles/m3) of a size
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.31
TABLE 4.5 Outdoor Air Requirements
Recommended by ASHRAE Standard 621999
Application cfm/person
Dining room 20
Bar and cocktail lounges 30
Conference rooms 20
Office spaces 20
Office conference rooms 20
Retail stores 0.20.3*
Beauty shops 25
Ballrooms and discos 25
Spectator areas 15
Theater auditoriums 15
Transportation waiting rooms 15
Classrooms 15
Hospital patient rooms 25
Residences 0.35
Smoking lounges 60
*cfm/ft2 floor area
Air changes/h
Source: Abridged with permission from ASHRAE
Standard 62-1999.
of 0.5 	m and larger or 65 particles / ft3 (2295 particles / ft3) of a size 5.0 	m and larger.
Class 100,000. Particle count not to exceed 100,000 particles/ ft3 (3,531,500 particles/m3) of
a size of 0.5 	m and larger or 700 particles / ft3 (24,720 particles / ft3) of a size 5.0 	m and larger.
Refer to Federal Standard 209E for more details.
Since workers in these clean rooms wear protective gowns and hats, in order to provide these air
cleanliness classes, a year-round constant temperature and associated relative humidity and a speci-
fied unidirectional airflow should be maintained. High-efficiency particulate air (HEPA) filters and
ultra-low-penetration air (ULPA) filters should be installed in the air conditioning systems. Special
building materials with hard and clean surfaces should be used as the building envelope. More details
are covered in Chap. 30.
4.12 SOUND LEVEL
Sound and Sound Level
Sound can be defined as a variation in pressure due to vibration in an elastic medium such as air. A
vibrating body generates pressure waves, which spread by alternate compression and rarefaction of
the molecules within the transmitting medium. Airborne sound is a variation of air pressure, with
atmospheric pressure as the mean value. Because sound is transmitted by compression and expansion
of molecules, it cannot travel in a vacuum. The denser the material, the faster the traveling
speed of a sound wave. The velocity of a sound wave in air is approximately 1130 fps. In water, it is
about 4500 fps and in steel 15,000 fps.
Noise is any unwanted sound. In air systems, noise should be compensated for, either by attenuation
(the process of reducing the amount of sound that reaches the space) or by masking it with
other, less objectionable sounds.
Sound Power Level and Sound Pressure Level
Sound power is the ability to radiate power from a sound source excited by an energy input. The intensity
of sound power is the power output from a sound source expressed in watts (W). Because of
the wide variation of sound outputfrom the threshold hearing level of 1012 W to a level of
108 W, generated by the launching of a Saturn rocket, a ratio of 1020 to 1it is more appropriate
and convenient to use a logarithmic expression to define sound power level, i.e.,
(4.28)
where Lw  sound power level, dB
w  sound source power output,W
Here 1012W, or 1 pW (picowatt), is the international reference base, and re indicates the reference
base.
The human ear and microphones are pressure-sensitive. Analogous to the sound power level, the
sound pressure level is defined as
(4.29)
where Lp  sound pressure level, dB
p  sound pressure, Pa
Lp  20 log
p
2  105 Pa
re 20 	Pa
Lw  10 log
w
1012 W
re 1 pW
4.32 CHAPTER FOUR
Here the reference sound pressure level is 2  105 Pa (pascal), or 20 	Pa, corresponding to the
hearing threshold. Because sound power is proportional to the square of the sound pressure, 10 log
p2  20 log p. Sound pressure levels of various sources are listed in Table 4.6.
The sound power level of a specific source is a fixed output. It cannot be measured directly and
can only be calculated through the measurement of the sound pressure level. On the other hand,
sound pressure level is the sound level measured at any one point and is a function of distance from
the source and characteristics of the surroundings.
Octave Bands
Sound waves, like other waves, are characterized by the relationship between wavelength, speed,
and frequency:
(4.30)
People can hear frequencies from 20 Hz to 20 kHz. To study and analyze sound, we must break it down
into components. A convenient way is to subdivide the audible range into eight octave bands or sometimes
- octave bands. An octave is a frequency band in which the frequency of the upper band limit
is double the frequency of the lower band limit. The center frequency of an octave or a octave band is
the geometric mean of its upper and lower band limits. An octave or octave band is represented by its
center frequency. The eight octave bands and their center frequencies are listed in Table 4.7.
Addition of Sound Levels
Because sound levels, in dB, are in logarithmic units, two sound levels cannot be added arithmetically.
If sound levels A, B, C, . . ., in dB, are added, the combined overall sound level L can be
calculated as
L  10 log (100.1A  100.1B  100.1C    ) (4.31)
1
3-
1
3-
24
13
Wavelength 
speed
frequency
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.33
TABLE 4.6 Typical Sound Pressure Levels
Sound Sound pressure Subjective
Source pressure, Pa level dB re 20 	Pa reaction
Military jet takeoff at 100 ft 200 140 Extreme
Passengers ramp at jet Threshold
airliner (peak) 20 120 of pain
Platform of subway station
(steel wheels) 2 100
Computer printout room* 0.2 80
Conversational speech at 3 ft 0.02 60
Window air conditioner* 0.006 50 Moderate
Quiet residential area* 0.002 40
Whispered conversation at 6 ft 0.0006 30
Buzzing insect at 3 ft 0.0002 20
Threshold of good hearing 0.00006 10 Faint
Threshold of excellent youthful Threshold
hearing 0.00002 0 of hearing
*Ambient.
Source: Abridged from ASHRAE Handbook 1989, Fundamentals. Reprinted by permission.
Human Response and Design Criteria
The human brain does not respond in the same way to lower frequencies as to higher frequencies.
At lower sound pressure levels, it judges a 20-dB sound at 1000 Hz to have the same loudness as a
52-dB sound at 50 Hz. However, at high sound pressure levels, a 100-dB sound at 1000 Hz seems
as loud as 110 dB at 50 Hz.
The purpose of noise control in an air conditioned space is to provide a background sound low
enough to avoid interference with the acoustical requirements of the occupants. The distribution of
the background sound should be balanced over a wide range of frequencies, without whistle, hum,
rumble, or audible beats. Three types of criteria for sound control are currently used in indoor system
design:
1. A-weighted sound level dBA. The A-weighted sound level dBA tries to simulate the response
of the human ear to sound at low sound pressure levels. An electronic weighting network automatically
simulates the lower sensitivity of the human ear to lower-frequency sounds by subtracting a
certain number of decibels at various octave bands, such as approximately 27 dB in the first octave
band, 16 dB in the second, 8 dB in the third, and 4 dB in the fourth. The A-weighted sound level
gives a single value. It is simple and also takes into consideration the human judgment of relative
loudness at low sound pressure levels. Its main drawback is its failure to consider the frequency
spectrum or the subjective quality of sound.
2. Noise criteria, or NC, curves. NC curves represent actual human reactions during tests. The
shape of NC curves is similar to the equal loudness contour representing the response of the human
ear, as shown in Fig. 4.10. NC curves are intended to indicate the permissible sound pressure level
of a broadband noise at various octave bands by a single NC curve sound level rating. NC curves
are practical. They also consider the frequency spectrum of the broadband noise. The main problem
with NC curves is that the shape of the curve does not approach a balanced, bland-sounding spectrum
that is neither rumbly nor hissy.
3. Room criteria, or RC, curves. RC curves, as shown in Fig. 4.11, are similar to NC curves
except that the shape of an RC curve is a close approximation of a balanced, bland-sounding
spectrum.
ASHRAE recommends the indoor design RC or NC ranges presented in Table 4.8. For sounds
containing significant pure tones or impulsive sounds, a 5- to 10-dB lower value should be specified.
Noise is always an annoying element and source of complaints in indoor environments. Attenuation
4.34 CHAPTER FOUR
TABLE 4.7 Octave Bands and Their Center
Frequencies
Band
Band frequency, H2
number Lower Center Upper
22.4 31.5 45
1 45 63 90
2 90 125 180
3 180 250 355
4 355 500 710
5 710 1,000 1,400
6 1,400 2,000 2,800
7 2,800 4,000 5,600
8 5,600 8,000 11,200
Source: Abridged with permission from ASHRAE
Handbook 1989, Fundamentals with permission.
to achieve an NC or RC goal less than 30 dB for a central air conditioning system or a packaged system
is very expensive. In actual practice, the NC or RC criteria range for private residences and
apartments varies greatly because of personal requirements and economic considerations.
To meet the listed design criteria range, the measured sound pressure levels of at least three of the
four octave bands between 250 and 2000 Hz should be within the listed NC or RC range. In
industrial workshops with machinery and equipment, Occupational Safety and Health Administration
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.35
FIGURE 4.10 Noise criteria (NC) curves. (Source: ASHRAE Handbook 1989, Fundamentals. Reprinted with
permission.)
(OSHA) Standard Part 1910.95, published by the U.S. Department of Labor, specifies: Feasible
administrative or engineering controls shall be utilized when employees are subjected to sound
levels exceeding those in Table G-16. If such controls fail to reduce sound levels to below those in
Table G-16, personal protective equipment shall be provided and used. Exposure to impulsive and
impact noise should not exceed 140 dB peak sound pressure level. Table G-16 is reproduced in this
text as Table 4.9.
4.36 CHAPTER FOUR
FIGURE 4.11 Room criteria (RC) curves. (Source: ASHRAE Handbook 1989, Fundamentals. Reprinted with
permission.)
4.13 SPACE PRESSURE DIFFERENTIAL
Most air conditioning systems strive to maintain a slightly higher indoor pressure than that of the
surroundings. This positive pressure tends to eliminate or reduce infiltration, the entry of untreated
air to the space. Negative space pressure may cause space high humidity levels, mold and mildew
growth, combustion equipment backdraft, and entry of sewer gas. For rooms where toxic,
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.37
TABLE 4.8 Recommended Criteria for Indoor Design RC
or NC Range
Recommended criteria
Type of area for RC or NC range
Private residence 2530
Apartments 2530
Hotels/motels
Individual rooms or suites 3035
Meeting/banquet rooms 2530
Halls, corridors, lobbies 3540
Service/support areas 4045
Offices
Executive 2530
Conference 2530
Private 3035
Open-plan 3035
Computer equipment rooms 4045
Public circulation 4045
Hospitals and clinics
Private rooms 2530
Wards 3035
Operating rooms 3540
Corridors 3540
Public areas 3540
Churches 2530
Schools
Lecture and classrooms 2530
Open-plan classrooms 3035
Libraries 3540
Concert halls 
Legitimate theaters 
Recording studios 
Movie theaters 3035
Note: These are for unoccupied spaces, with all systems
operating.
*Design goals can be increased by 5 dB when dictated by budget
constraints or when intrusion from other sources represents a limiting
condition.
An acoustical expert should be consulted for guidance on these
critical spaces.
Source: ASHRAE Handbook 1987, HVAC Systems and Applications.
Reprinted with permission.
TABLE 4.9 Occupational Noise Exposure
(Occupational Health and Safety Administration
Table G-16)
Duration Sound level
per day, h dBA slow response
8 90
6 92
4 95
3 97
2 100
1 102
1 105
110
(or less) 115
Note: If the variations in noise level involve
maxima at intervals of 1 s or less, it is to be considered
continuous.
In all cases where the sound levels exceed the
values shown herein, a continuing, effective hearing
conservation program shall be administered.
Source: Occupational Safety and Health Standard,
Part 1910.95. Reprinted with permission.
1
4
1
2
1
2
hazardous, contaminated, or objectionable gases or substances are produced, a slightly lower
pressure, or an appropriate negative room pressure, should be maintained to prevent the diffusion of
these substances to the surroundings and at the same time prevent and reduce the damage of the
uncontrolled outdoor airflow.
The magnitude of the positive or negative pressure to be maintained in the space must be carefully
determined. A higher positive pressure always means a greater amount of exfiltrated air,
which creates requirements for greater volumes of outdoor air intake. A higher negative pressure
means a greater infiltration. Both result in higher initial and operating costs. Space pressurization
is closely related to the amount of effective leakage area, or tightness of the building. Cummings
et al. (1996) reported that the space pressure differentials between indoor and outdoor air and
building tightness vary widely for seven field-surveyed restaurants. For comfort systems in lowrise
buildings, the recommended pressure differential is 0.005 to 0.03 in. WC (1.25 to
7.5 Pa), which is often adopted for average building tightness. For a 44-story high-rise
building, the pressure differential between the entrance lobby outdoors and rooftop floor because
of the stack effect during winter may exceed 0.3 in. WC (75 Pa). The phrase in. WC represents
the pressure at the bottom of a water column that has a height of the specified number of
inches.
Another type of space pressurization control for the health and comfort of passengers in an aircraft
is cabin pressurization control. If an airplane is flying at an altitude of 28,000 ft (8540 m) with
an ambient pressure of 4.8 psia (33.1 kPa abs.), the minimum cabin pressure required is 10.8 psia
(75 kPa abs.); and a cabin pressurization control system must be installed to maintain a pressure
differential of 10.8  4.8  6 psi (41 kPa) between the cabin and the ambient air.
Process air conditioning systems, such as for clean rooms, need properly specified space
pressurization to prevent contaminated air from entering the clean, uncontaminated area from surrounding
contaminated or semicontaminated areas. According to Federal Standard 209B, Clean
Rooms and Work Station Requirements, Controlled Environment, published by the U.S. government
in 1973, the minimum positive pressure differential between the clean room and any adjacent area
of less clean requirements should be 0.05 in. WC (12.5 Pa) with all entryways closed. When the entryways
are open, an outward flow of air is to be maintained to minimize migration of contaminants
into the room.
Space pressurization is discussed in Chaps. 20 and 23.
4.14 OUTDOOR DESIGN CONDITIONS
In principle, the capacity of air conditioning equipment is selected so that indoor design conditions
can be maintained when the outdoor weather does not exceed the design values. The outdoor
weather affects the space cooling load and the capacity of the air system to condition the required
amount of outdoor air. It is not economical to choose the annual maximum or annual minimum values
as the design data. Outdoor design conditions are usually determined according to statistical
analysis of the weather data of the previous 30 years so that either 99.6 or 99 percent of the winter
indoor design conditions can then be attained annually, or only 0.4, 1.0, or 2.0 percent of the time
annually the summer indoor design conditions are exceeded. The greater the need for strict control
of indoor environmental parameters, the greater the percentage coverage of the winter outdoor design
hours or the smaller the percentage occurrence of the total hours that exceed the outdoor
design values.
The recommended design values are based on data from the National Climatic Data Center
(NCDC) from 1961 to 1991, or from 1982 to 1993, and Canadian Weather Energy and Engineering
Data Sets (CWEEDS) from 1953 through 1993. Per ASHRAE Handbook 1997, Fundamentals, and
ASHRAE/IESNA Standard 90.1-1999, the design conditions recommended for large cities in the
United States are listed in Table 4.10.
4.38 CHAPTER FOUR
Summer and Winter Outdoor Design Conditions
In Table 4.10, note the following:
 Summer outdoor design dry-bulb temperature for a specific locality To,s, in F (C), is the rounded
higher integral number of the statistically determined summer design temperature To,ss such that
the average number of hours of occurrence of outdoor temperatures To higher than To,ss annually
is on average 0.4 percent (35 h), 1 percent (88 h), or 2 percent (175 h).
 The summer outdoor mean coincident design wet-bulb temperature To, s, in F (C), is the mean
of all wet-bulb temperatures occurring at the specific summer outdoor design dry-bulb temperature
To,s during the summer.
 The 1.0 percent summer design wet-bulb temperature To1 is the design value having an average
annual occurrence of To  To1 of 88 h. Variable To1 is used for evaporative cooling systems,
cooling towers, and evaporative condensers, which are covered in later chapters.
 The mean daily range, in F (C), indicates the mean of the difference between the daily maximum
and minimum temperatures for the warmest month.
 The winter outdoor design dry-bulb temperature To,w, in F, is the rounded lower integral value of
the statistically determined winter outdoor design temperature To,ws, such that the annual average
number of hours of occurrence of outdoor temperature at values To  To,ws should be equaled or
exceeded 99.6 or 99 percent of the total number of annual hours. The annual average number of
hours in winter when To 	 0.99To,ws is 35 h, and To 	 0.99To,ws is 88 h.
 The number of degree-days is the difference between a base temperature and the mean daily outdoor
air temperature for any one day Tbase  To,m, both in F. Annually, the total number of heating
degree-days with a base temperature of 65F, or HDD65, is
(4.32)
where n  number of days whose To,m 	 65F per annum. The total number of cooling degreedays
with a base temperature of 50F, or CDD50, is
(4.33)
where m  number of days whose To,m 50F per annum. Heating and cooling degree-days with
different base temperatures have been used as climatic parameters to calculate the energy flux
through building envelope, or to determine the U value and the configuration of the building
envelope.
 In the last two columns, MWS/MWD to DB 99.6% indicates mean coincident wind speed
(MWS)/mean coincident wind direction (MWD), i.e., most frequently occurring with the heating
dry-bulb 99.6 percent. In MWD, wind direction is expressed in degrees; 270 represents west, and
180 south.
The Use of Outdoor Weather Data in Design
During the design of air conditioning systems, the following parameters are often adopted:
1. Indoor and outdoor design conditions are used to calculate the space cooling and heating loads.
2. Summer outdoor dry-bulb and coincident wet-bulb temperatures are necessary to evaluate the
coil load. The summer outdoor wet-bulb temperature is used to determine the capacity of the
evaporative coolers, cooling towers, and evaporative condensers.
CDD50  
m1
(To,m  50)
HDD65  
n1
(65  To,m)
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.39
4.40 TABLE 4.10 Climatic Conditions for the United States and Canada
Winter Summer
Annual Annual Annual
average daily cooling heating
Design incident degree- degree-
Mean wet- solar days days
daily bulb, energy on base base
Lat, Elevation, range, 1% east or west, 50F, 65F,
City State degree ft F F F F F F F Btu / ft2  day CDD50 HDD65 MWS MWD
Albuquerque New Mexico 35.05 5,315 13 18 96/60 93/60 91/60 25.4 64 1,105 3,908 4,425 8 360
Anchorage Alaska 61.17 131 14 9 71/59 68/57 65/56 12.6 58 538 10,570 8 290
Atlanta Georgia 33.65 1,033 18 23 93/75 91/74 88/73 17.3 76 807 5,038 2,991 12 320
Atlantic
City CO New Jersey 39.45 66 8 13 91/74 88/73 86/72 18.1 76 9 310
Baltimore Maryland 39.18 154 11 15 93/75 91/74 88/73 18.8 76 739 3,709 4,707 10 290
Billings Montana 45.80 3,570 13 7 93/63 90/62 87/61 25.8 64 814 2,466 7,164 10 230
Birmingham Alabama 33.57 630 18 23 94/75 92/75 90/74 18.7 77 789 5,206 2,918 7 340
Bismarck North Dakota 46.77 1,660 21 16 93/68 90/67 86/66 26.5 70 766 2,144 8,968 7 290
Boise Idaho 43.57 2,867 2 9 96/63 94/63 91/62 30.3 64 916 2,807 5,861 6 130
Boston Massachusetts 42.37 30 7 12 91/73 87/71 84/70 15.3 74 659 2,897 5,641 17 320
Bridgeport Connecticut 41.17 16 8 12 86/73 84/72 82/71 14.1 74 2,997 5,537 14 320
Buffalo New York 42.93 705 2 5 86/70 84/69 81/68 17.7 72 609 2,468 6,747 12 270
Burlington Vermont 44.47 341 11 6 87/71 84/69 82/68 20.4 72 698 2,228 7,771 6 70
Caribou Maine 46.87 623 14 10 85/69 82/67 79/66 19.5 70 649 1,470 9,651 10 270
Casper Wyoming 42.92 5,289 13 5 92/59 89/58 86/58 30.4 61 961 2,082 7,682 9 260
Charleston South Carolina 32.90 49 25 28 94/78 92/77 90/77 16.2 79 796 6,188 2,013 7 20
Charleston West Virginia 38.37 981 6 11 91/73 88/73 86/71 19.1 75 667 3,655 4,646 7 250
Charlotte North Carolina 35.22 768 18 23 94/74 91/74 89/73 17.8 76 809 4,704 6 50
Chicago, Illinois 41.98 673 6 1 91/74 88/73 86/71 19.6 75 729 2,941 6,536 10 270
Cincinnati CO Ohio 39.10 482 5 12 93/74 90/75 88/73 20.0 76 3,733 4,988 9 260
Cleveland Ohio 41.42 804 1 6 89/73 86/72 84/71 18.6 74 2,755 6,201 12 230
Concord New Hampshire 43.20 344 8 2 90/71 87/70 84/68 24.1 73 630 2,087 7,554 4 320
Dallas/
Fort Worth Texas 32.90 597 17 24 100/74 98/74 96/74 20.3 77 875 6,587 2,259 13 350
Denver Colorado 39.75 5,331 3 3 93/60 90/59 87/59 26.9 63 971 2,732 6,020 6 180
Des Moines Iowa 41.53 965 9 4 93/76 90/74 87/73 18.5 76 788 3,371 6,497 11 320
Detroit Michigan 42.23 663 0 5 90/73 87/72 84/70 20.4 74 676 3,046 6,167 11 240
Honolulu Hawaii 21.35 16 61 63 89/73 88/73 87/73 12.2 75 953 9,949 0 5 320
Houston Texas 29.97 108 27 31 96/77 94/77 92/77 18.2 79 805 6,876 1,599 8 340
Indianapolis Indiana 39.73 807 3 3 91/75 88/74 86/73 18.9 77 692 3,453 5,615 8 230
Jackson Mississippi 32.32 331 21 25 95/77 93/76 92/76 19.2 79 833 5,900 2,467 7 340
Kansas City Missouri 39.32 1,024 1 4 96/75 93/75 90/74 18.8 77 3,852 5,393 10 320
Lansing Michigan 42.77 873 3 2 89/73 86/72 84/70 21.7 74 2,449 7,101 8 290
Las Vegas Nevada 36.08 2,178 27 30 108/66 106/66 103/65 24.8 70 1,136 6,745 2,407 7 250
Design
dry-bulb/
Design mean coincident
dry-bulb wet-bulb
99.6% 99% 0.4% 1% 2%
MWS/MWD
to DB
99.6%
4.41
Lincoln CO Nebraska 40.85 1,188 7 2 97/74 94/74 91/73 22.3 76 3,455 6,278 9 350
Little Rock Arkansas 34.92 312 16 21 97/77 95/77 92/76 19.5 79 831 5,299 3,155 9 360
Los Angeles CO California 33.93 105 43 45 85/64 81/64 78/64 10.9 69 962 4,777 1,458 6 70
Louisville Kentucky 38.18 489 6 12 93/76 90/75 88/74 18.2 77 727 4,000 4,514 10 290
Memphis Tennessee 35.05 285 16 21 96/78 94/77 92/77 16.8 79 806 5,467 3,082 10 20
Miami Florida 25.82 13 46 50 91/77 90/77 89/77 11.4 79 874 9,474 200 10 340
Milwaukee Wisconsin 42.95 692 7 2 89/74 86/72 83/70 16.6 74 724 2,388 7,324 13 290
Minneapolis/
St. Paul Minnesota 44.88 837 16 11 91/73 88/71 85/70 19.1 74 709 2,680 7,981
New Orleans Louisiana 29.98 30 30 34 93/79 92/78 90/78 15.5 80 838 6,910 1,513 7 340
New York New York 40.65 23 11 15 91/74 88/72 85/71 13.9 75 650 3,342 5,027 17 320
Norfolk Virginia 36.90 30 20 24 93/77 91/76 88/75 15.3 77 792 1,586 3,609 12 340
Phoenix Arizona 33.43 1,106 34 37 110/70 108/70 106/70 23 75 1,116 8,425 1,350 5 90
Pittsburgh Pennsylvania 40.50 1,224 2 7 89/72 86/70 84/69 19.5 73 642 2,836 5,968 10 260
Portland CO Oregon 45.60 39 22 27 90/67 86/66 83/64 21.6 67 647 2,517 4,522 13 120
Providence Rhode Island 41.73 62 5 10 89/73 86/71 83/70 17.4 74 677 2,743 5,884 12 340
Rapid City South Dakota 44.05 3,169 11 5 95/65 91/65 88/64 25.3 68 819 2,412 7,301 9 350
St. Louis Missouri 38.75 564 2 8 95/76 93/75 90/74 18.3 78 797 4,283 4,758 12 290
Salt Lake City Utah 40.78 4,226 6 11 96/62 94/62 92/61 27.7 65 975 3,276 5,765 7 160
San Antonio Texas 29.53 794 26 30 98/73 96/73 94/74 19.1 77 878 7,142 1,644 10 350
San Diego California 32.73 30 44 46 85/67 81/67 79/67 8.9 71 950 5,223 1,256 3 70
San Francisco California 37.62 16 37 39 83/63 78/62 74/61 16.7 63 941 2,883 3,016 5 160
CO
Seattle CO Washington 47.45 449 23 28 85/65 81/64 78/62 18.3 65 621 2,120 4,611 10 10
Shreveport Louisiana 32.47 259 22 26 97/77 95/77 93/76 19.1 79 843 6,166 2,264 9 360
Spokane Washington 47.62 2,461 1 7 92/62 89/61 85/60 26.1 63 758 2,032 6,842 7 50
Syracuse New York 43.12 407 3 2 88/72 85/71 83/70 20.3 73 611 2,399 6,834 7 90
Tucson Arizona 32.12 2,556 31 34 104/65 102/65 100/65 29.4 71 1,112 6,921 1,678 7 140
Tulsa Oklahoma 36.20 676 9 14 100/76 97/76 94/75 19.5 78 820 5,150 3,691 11 360
Washington DC 38.85 66 15 20 95/76 92/76 89/74 16.6 78 724 11 340
Wichita Kansas 37.65 1,339 2 8 100/73 97/73 94/73 22.2 76 4,351 4,791 13 360
Wilmington Delaware 39.68 79 10 14 91/75 89/74 86/73 17.0 76 726 3,557 4,937 11 290
Canada
Calgary Alberta 51.12 3,556 22 17 83/60 80/59 77/57 22.0 61 7 0
Montreal Quebec 45.47 118 12 7 85/71 83/70 80/68 17.6 72 7 250
Regina Saskatchewan 50.43 1,893 29 24 89/64 85/64 82/62 23.6 66 9 270
Toronto Ontario 43.67 568 4 1 87/71 84/70 81/68 20.2 72 9 340
Vancouver British Columbia 49.18 7 18 24 76/65 74/64 71/62 14.0 64 6 90
Winnipeg Manitoba 49.90 784 27 23 87/68 84/67 81/66 20.5 70 7 320
Note: CO designates locations within an urban area. Most of the data are taken from city airport temperature observations. Some semirural data are comparable to airport data.
Source: Abridged with permission from ASHRAE Handbook 1997, Fundamentals and ASHRAE/IES Standard 90.1-1999.
3. Outdoor weather data presented consecutively for a whole year of 8760 h, or other simplified
form, are sometimes used for year-round energy estimations.
4. Outdoor climate has a significant influence on the selection of an air conditioning system and its
components.
Outdoor Weather Characteristics and Their Influence
The following are characteristics of outdoor climate in the United States:
1. For clear days during summer, the daily maximum temperature occurs between 2 and 4 P.M. In
winter, the daily minimum temperature usually occurs before sunrise, between 6 and 8A.M.
2. The maximum combined influence of outdoor temperature and solar radiation on load calculations
usually occurs in July or August. January is often the coldest month in many locations.
3. The mean daily temperature range of the warmest month varies widely between different locations.
The smallest mean daily range, 10F (5.6C), occurs at Galveston, Texas. The greatest,
45F (25C), occurs at Reno, Nevada. Many cities have a mean daily range of 15 to 25F (8.3 to
13.9C). Usually, coastal areas have smaller mean daily ranges, and continental areas and areas
of high elevation have greater values.
4. Extremes in the difference between 1 percent design dry-bulb temperature in summer and the 99
percent design temperature in winter in the United States are
 A maximum of 119F (66C) at Bettles, Alaska
 A minimum of 18F (10C) at Kaneohe, MCAS, Hawaii
5. Snelling (1985) studied outdoor design temperatures for different locations all over the country.
He found that extremely cold temperatures may last for 3 to 5 days. Extremely hot temperatures
seldom last more than 24 h.
6. Among the 538 locations listed in Table 4.10, 138 locations have a summer mean coincident
wet-bulb temperature corresponding to 1 percent dry-bulb, Tos  68F (20C). There it is advantageous
to use evaporative cooling systems to replace part of or all the refrigeration capacity.
7. When commercial buildings are only occupied after 10 A.M. or even after 12 noon, a winter outdoor
design temperature higher than the 99.0 percent design dry-bulb temperature should be
considered.
REFERENCES
ANSI/ASHRAE, ANSI/ASHRAE Standard 55-1992, Thermal Environmental Conditions for Human Occupancy,
Atlanta, GA, 1992.
ASHRAE, ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality, Atlanta, GA, 1999.
ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, Atlanta, GA, 1996.
ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, GA, 1997.
ASHRAE Evironmental Health Committee 1987, Indoor Air Quality Position Paper, Atlanta, GA, 1987.
Bayer, C. W., and Black, M. S., IAQ Evaluations of Three Office Buildings, ASHRAE Journal, July 1988,
pp. 4852.
Bedford, T., and Warmer, C. G., The Globe Thermometer in Studies of Heating and Ventilation, Journal of the
Institution of Heating and Ventilating Engineers, vol. 2, 1935, p. 544.
Berglund, L., Mathematic Models for Predicting the Thermal Comfort Response of Building Occupants,
ASHRAE Transactions, 1978, Part I, pp. 735749.
Cena, K., Spotila, J. R., and Avery, H. W., Thermal Comfort of the Elderly Is Affected by Clothing, Activity, and
Psychological Adjustment, ASHRAE Transactions, 1986, Part II A, pp. 329342.
4.42 CHAPTER FOUR
Cummings, J. B.,Withers, Jr., C. R., Moyer, N. A., Fairey, P. W., and McKendry, B. B., Field Measurement of
Uncontrolled Airflow and Depressurization in Restaurants, ASHRAE Transactions, 1996, Part I, pp. 859869.
DeBat, R. J., Humidity: The Great Equalizer, HPAC, no. 10, 1996, pp. 6671.
Eade, R., Humidification Looming Larger in IAQ, Engineered Systems, no 1, 1996, pp. 4958.
Fanger, P. O., Thermal Comfort Analysis and Applications in Environmental Engineering, McGraw-Hill, New
York, 1970.
Federal Supply Service, Federal Standard No. 209B, Clean Rooms and Work Station Requirements, Controlled
Environment, General Services Administration,Washington, 1973.
Gagge, A. P., Stolwijk, J. A. J., and Nishi,Y., An Effective Temperature Scale Based on a Simple Model of
Human Physiological Regulatory Response, ASHRAE Transactions, 1971, Part I, p. 247.
George, A. C., Measurement of Sources and Air Concentrations of Radon and Radon Daughters in Residential
Buildings, ASHRAE Transactions, 1985, Part II B, pp. 19451953.
Holzle, A. M., Munson, D. M., and McCullough, E. A., A Validation Study of the ASHRAE Summer Comfort
Envelope, ASHRAE Transactions, 1983, Part I B, pp. 126138.
Janssen, J. E., Ventilation for Acceptable Indoor Air Quality, ASHRAE Journal, October 1989, pp. 4046.
Kirkbride, J., Lee, H. K., and Moore, C., Health and Welfare Canadas Experience in Indoor Air Quality Investigation,
Indoor Air 90, vol. 5, D. S. Walkinshaw, ed., Ottawa: International Conference on Indoor Air and
Climate, 1990, pp. 99106.
McCullough, E. A., and Jones, B. W., A Comprehensive Data Base for Estimating Clothing Insulation, IER
Technical Report 8401, Kansas State University, 1984.
National Institute for Occupational Safety and Health, Indoor Air Quality: Selected Reference, Division of
Standards Development and Technology Transfer, Cincinnati, Ohio, 1989.
OSullivan, P., Energy and IAQ Can Be Complementary, Heating/Piping/Air Conditioning, February 1989,
pp. 3742.
Persily, A., Ventilation Rates in Office Buildings, ASHRAE Journal, July 1989, pp. 5254.
Rohles, Jr., F. H.,Woods, J. E., and Nevins, R. G., The Effect of Air Movement and Temperature on the Thermal
Sensations of Sedentary Man, ASHRAE Transactions, 1974, Part I, pp. 101119.
Seppanen, O., McNall, P. E., Munson, D. M., and Sprague, C. H., Thermal Insulating Values for Typical Clothing
Ensembles, ASHRAE Transactions, 1972, Part I, pp. 120130.
Snelling, H. J., Duration Study for Heating and Air Conditoning Design Temperature, ASHRAE Transactions,
1985, Part II, pp. 242249.
Spain, S., The Upper Limit of Human Comfort from Measured and Calculated PMV Values in a National
Bureau of Standards Test House, ASHRAE Transactions, 1986, Part I B, pp. 2737.
Sterling, E. M., Arundel, A., and Sterling, T. D., Criteria for Human Exposure to Humidity in Occupied Buildings,
ASHRAE Transactions, 1985, Part I B, pp. 611622.
Thayer,W. W., Tobacco Smoke Dilution Recommendations for Comfort Ventilation, ASHRAE Transactions,
1982, Part II, pp. 291304.
Woodcock, A. H., Moisture Transfer in Textile Systems, Textile Research Journal, vol. 8, 1962, pp. 628633.
INDOOR AND OUTDOOR DESIGN CONDITIONS 4.43
CHAPTER 5
ENERGY MANAGEMENT AND
CONTROL SYSTEMS
5.1
5.1 AUTOMATIC CONTROL SYSTEM AND
HISTORICAL DEVELOPMENT 5.2
Air Conditioning Automatic Control
System 5.2
Historical Development 5.2
Distribution of Energy Management and
Control Systems 5.3
5.2 CONTROL LOOP AND CONTROL
METHODS 5.5
Control Loop 5.5
Sequence of Operations 5.5
Control Methods 5.7
Comparison of Control Methods
5.3 CONTROL MODES 5.9
Two-Position Control 5.9
Step Control and Modulating
Control 5.10
Floating Control 5.11
Proportional Control 5.11
Proportional plus Integral (PI)
Control 5.13
Proportional-Integral-Derivative (PID)
Control 5.14
Compensation Control, or Reset 5.15
Applications of Various Control
Modes 5.15
5.4 SENSORS AND TRANSDUCERS 5.16
Temperature Sensors 5.18
Humidity Sensors 5.18
Pressure Sensors 5.19
Flow Sensors 5.19
Carbon Dioxide and Air Quality
Sensors 5.20
Occupancy Sensors 5.20
Wireless Zone Sensors and Intelligent
Network Sensors 5.21
Transducers or Transmitters 5.21
5.5 CONTROLLERS 5.21
Direct-Acting or Reverse-Acting 5.21
Normally Closed or Normally Open 5.22
Pneumatic Controllers 5.22
Electric and Electronic Controllers 5.23
Direct Digital Controllers 5.23
5.6 WATER CONTROL VALVES AND
VALVE ACTUATORS 5.26
Valve Actuators 5.26
Types of Control Valves 5.27
Valve Characteristics and Ratings 5.27
Valve Selection 5.29
Valve Sizing 5.30
5.7 DAMPERS AND DAMPER
ACTUATORS 5.32
Types of Volume Control Dampers 5.32
Damper Actuators (Motors) 5.33
Volume Flow Control between Various
Airflow Paths 5.33
Flow Characteristics of Opposed- and
Parallel-Blade Dampers 5.35
Damper Selection 5.37
Damper Sizing 5.37
5.8 SYSTEM ARCHITECTURE 5.38
Architecture of a Typical EMCS with
DDC 5.38
System Characteristics 5.40
Future Development 5.41
5.9 INTEROPERABILIITY AND OPEN
PROTOCOL BACnet 5.41
Interoperability 5.41
BACnetOpen Data Communication
Protocol 5.41
Application Layer 5.42
Network Layer 5.43
Data Link/Physical LayerNetwork
Technology 5.43
Connection between BACnet and
Proprietary Network 5.44
LonTalk Protocol 5.44
5.10 CONTROL LOGIC AND ARTIFICIAL
INTELLIGENCE 5.45
Fuzzy Logic 5.45
Knowledge-Based Systems and
Expert Systems 5.47
Artificial Neural Networks 5.51
5.11 PROGRAMMING FOR DDC
SYSTEMS 5.53
Evolution of DDC Programming 5.53
Graphical Programming 5.53
Templates 5.54
Graphical Programming for Mechanical
Cooling Control 5.55
5.12 TUNING DDC UNITS 5.55
Tuning PI Controllers 5.55
Self-Tuning PI and PID Controllers 5.55
Adaptive Control 5.56
5.13 FACTORS AFFECTING CONTROL
PROCESSES 5.56
Load 5.56
Climate Change 5.56
System Capacity 5.57
Performance of Control Processes 5.57
Thermal Capacitance 5.58
5.14 FUNCTIONAL CONTROLS 5.58
Generic Controls 5.59
5.1 AUTOMATIC CONTROL SYSTEM AND HISTORICAL
DEVELOPMENT
Air Conditioning Automatic Control System
An air conditioning automatic control system or simply a control system, primarily modulates
the capacity of the air conditioning equipment to maintain predetermined parameter(s) within an
enclosure or for the fluid entering or leaving the equipment to meet the load and climate changes at
optimum energy consumption and safe operation. The predetermined parameter or variable to
be controlled is called the controlled variable. In air conditioning or in HVAC&R, the controlled
variable can be temperature, relative humidity, pressure, enthalpy, fluid flow, contaminant concentration,
etc.
Because of the variation of the space load and the outdoor climate, a control system is one of the
decisive factors for an air conditioning, or an HVAC&R, system to achieve its goal: to effectively
control the indoor environmental parameters, provide an adequate amount of outdoor air, be energy-
efficient, and provide better security and safety. Today nearly all HVAC&R systems are installed
with a control system to provide effective operation and energy conservation.
Historical Development
Early controls for comfort air conditioning systems used in the Capitol building since 1928 were
pneumatic controls that included high- and low-limit thermostats in the supply air discharge of the
air-handling unit. These controls were used to maintain the desired supply air temperature. Dewpoint
control of the supply air maintained room humidity within a desirable range. A thermostat in
the return-air passage was used to control the air volume flow rate passing through the cooling coil,
and a 16-point recorder connected with resistance thermometers was used to record the room temperature
at various locations.
Shavit (1995) stated that the use of thermocouples offers the possibility of remote monitoring of
the space temperature. In coordination with pneumatic local control systems, the first centralized
monitoring system with a central panel and remote set point change was installed in the White
House in 1950. The concept of a building automation system (BAS) which centralizes the monitoring
management of building services was devised and developed in 1950s. The primary improvement
of the pneumatic control system occurred in 1963. The vapor and bulb-based sensors were
replaced by a remote pneumatic sensor using a bimetal strip and a nozzle flapper to regulate the
compressed air pressure between 3 and 15 psi (20.7 and 103.4 kPa).
Shavit also indicated that the first on-line computer was introduced in 1967 in One Main Place
in Dallas, Texas. This system had the first set energy conservation software including enthalpy control,
demand limit, optimum start/ stop, reset according to the zone of highest demand, chiller optimization,
and night purge. In 1970, the introduction of all-electronic, solid-state centralized control
system was another significant milestone. Solid-state components improved the scanning process
and serial transmission as well as reduced many wires in the trunk wiring to a single pair.
5.2 CHAPTER FIVE
Specific Controls 5.60
Commissioning and Maintenance 5.61
5.15 FAULT DETECTION AND
DIAGNOSTICS 5.61
Basics 5.61
Expert System Rule-Based
Diagnostics 5.62
ARX and ANN Model-Based
Diagnosis 5.63
5.16 CONTROLS IN ASHRAE/IESNA
STANDARD 90.1-1999 5.66
General 5.66
Off-Hour Controls 5.66
REFERENCES 5.67
The 1973 energy crisis greatly boosted the control industry. Many energy management systems
(EMSs) were installed for the purpose of saving energy. The microprocessor-based direct digital
control (DDC) was first introduced in 1981 and the DDC unit controller in 19861987. By means
of the incorporated hardware and software, it tremendously increased the control functions, developed
the control logic, and accelerated the data processing and analysis.
Distribution of Energy Management and Control Systems
A building automation system is a centralized monitoring, operation, and management system of the
building services in a building including air conditioning (HVAC&R), lighting, fire protection, and security.
Only air conditioning control systems and smoke control systems are covered in this handbook.
In the United States, a single-zone, two-position electric control system, as shown in Fig. 5.1, is
now widely used in residential and light commercial buildings that adopt DX packaged systems.
Single-zone means that the load characteristics of the whole conditioned space are similar, it is
controlled by one controller, and two-position means on-off control. This kind of control system is
simple, easy to manage, and low in cost. For a large high-rise building constructed in the 1990s, an
energy management and control system (EMCS), as shown in Fig. 5.2, is often the choice. Variable
air volume indicates that airflow is varied to match the variation of the system load. Today, an
EMCS is a microprocessor-based system with direct digital control which optimizes the operation
and the indoor environmental parameters of an HVAC&R system in order to maintain a healthy and
comfortable environment at optimum energy use. Today, an EMCS is an advanced-generation energy
management and control system and is also a part of a building automation system.
According to the surveyed results in Commercial Buildings Characteristics 1992 by EIA, the
percentages of the floor area served by the installed EMCS among the total conditioned area in
commercial buildings are as follows:
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.3
DX-coil
Compressor
Thermostat
Furnace
Condensing
unit
Refrigerant
pipes
Supply
fan
FIGURE 5.1 A single-zone, two-position electric control system.
Year built Building with EMCS, million ft2 Percentage of floor space
Before 1970 5431 18
19701979 3313 29
19801989 4343 36
19901992 1236 58
5.4 CHAPTER FIVE
3
1
1 1 5 6 4 2
3
P3
P1
M
SD
P
T
SD T2
T1
T4
2 1
T3
P2 DM4
DM2
DM3
MPS
DM1
Outdoor
air damper
Air-handler
Smoke detector
Pressure sensor
Temperature sensor
M
DM
MPS
Multicompensator
Damper actuator
Manual positioning switch
Relief
air damper
Relief
fan
14.0
70
60
50%
20%
13.5
50
40
40
0
S
S ru r
r ru
sf
sf cc
m
50 60 70
T, F
w, lb/lb
80 90 100
0.020
0
0.004
0.008
0.012
0.016
0
Cooling
coil
Outdoor air
reference
Conditioned
space
r
Recirculating air damper
Inlet
vanes
sf
S
ru
C
cc
FIGURE 5.2 Control diagram and psychrometric analysis of an energy management and control system for
a single-zone, VAV cooling packaged system.
The later the building is constructed and the larger the conditioned area of the commercial building,
the greater is the chance to install an energy management and control system (EMS).
5.2 CONTROL LOOP AND CONTROL METHODS
Control Loop
The basic element of a control system is a control loop. A control loop often consists of a sensor
(such as T2 in Fig. 5.2), which senses and measures the controlled variable of recirculating air; a
DDC unit controller which compares the sensed input signal with the predetermined condition (the
set point) and sends an output signal to actuate the third element, a controlled device or control element
(such as a damper or a valve in Fig. 5.2). Modulation, minute adjustments, or on-off control of
dampers and valves will change the controlled devices position or operating status, which affects
the controlled variable by changing the airflow and water flow, electric power supply, and so on.
The controlled variable is thus varied toward the predetermined value, the set point. Air and water
are control media or control agents. Airflow and water flow are the manipulated variables. The
equipment that varies the output capacity by changing the opening position of the dampers and
valves is called the process plant.
There are two types of control loops: open and closed. An open-loop system assumes a fixed
relationship between the controlled variable and the input signal being received. The sensed variable
is not the controlled variable, so there is no feedback. An example of an open-loop system
would be a ventilating fan that turns on when the outdoor temperature exceeds a specified set point.
The sensed variable is the outdoor air temperature, and the controlled variable is the state of the fan
(on or off).
A closed-loop system depends on sensing the controlled variable to vary the controller output
and modulate the controlled device. In Fig. 5.2, temperature sensor T2 senses the controlled
variable, the recirculating air temperature entering the air handler Tru; the DDC unit controller receives
this sensed signal input and produces an output according to the software stored to modulate
the fan speed, or the position of the inlet vanes of the supply fan. As the supply fan speed changes,
or inlet vanes open and close, the supply air volume flow rate varies and the space temperature Tr
and recirculating temperature Tru change accordingly. This change in Tru is sensed again by T2 and
fed back to the controller for further modulation of fan speed or inlet vanes to maintain values of
Tru that approach the set points. These components form a closed-loop system.
Figure 5.3 shows a block diagram of this closed-loop system. It shows a secondary input to the
controller, such as the outdoor air temperature To, which may reset the set point in the controller to
provide better and more economical control. The disturbances that affect the controlled variable
are load variations and changes in the outdoor weather. After the controller senses the signal feedback,
it sends a corrective signal to the controlled device based on the difference between the
sensed controlled variable and the set point. Thus, Tru is under continuous comparison and correction.
A control system or its component, control subsystem, used to control the controlled variable(s)
in a conditioned space, or within a mechanical device or equipment, may contain only one control
loop; or it may contain two or more loops.
Sequence of Operations
The sequence of operations is a description of the sequential order of the functional operations that
a control system is supposed to perform, which plans and guides the operation and control of an air
conditioning system. For a single-zone variable-air-volume (VAV) cooling packaged system shown
in Fig. 5.2, when cold air supply is required during full occupancy, the sequence of operations of
cooling mode is as follows:
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.5
 The supply fan is started and stopped by the scheduling software stored in the DDC unit
controller. Manual override is possible. When the time schedule puts this system in cooling
occupy mode, the microprocessor-based controller goes through a short initiation period, such as
a 2-min period. During this period, dampers are driven to fully open, minimum open, and fully
closed positions. These determine the effective range of the economizer potentiometer range. The
smoke detector in the return air or the low-temperature limit sensor of the mixed air will stop the
supply fan if necessary. The supply fan status (on or off) is determined by the pressure differential
switch across the fan. When the initiation period is completed, the supply fan is turned on.
 The control system tends to maintain the recirculating temperature Tru at around the cold set
point, and it uses the 100 percent all outdoor air free cooling economizer cycle as the first-stage
cooling. If the outdoor temperature To  Tru, the outdoor air damper is fully opened and the recirculating
air damper closed. If the outdoor air temperature To  Tru, the outdoor air damper is
closed to a minimum position to provide required outdoor ventilation air, and the recirculating
damper is fully opened.
 When the initiation period is completed, the supply fan is turned on from its zero speed. If Tru,
sensed by the temperature sensor T2, is at a value above the set point, Tru  Tc, set, also To 
Tru, the speed of the supply fan is gradually increased by the variable-speed drive inverter, resulting
in a higher supply volume flow rate. When the fan speed is raised to its upper limit, the
supply volume flow rate is then at its maximum value. A still higher space load further raises
Tru, and it exceeds the cold set point Tc, set, and if To  Tru, chilled water starts to flow to the
cooling coil to cool the air simultaneously with the outdoor air free cooling, in order to make
Tru  Tc, set.
 Only when the outdoor air damper is fully open and the static pressure difference between the
space air and outdoor air is greater than a preset value, such as ps  o  0.03 in. WC (7.5 Pa), will
the relief fan be energized. Its speed is modulated to maintain ps  o  0.03 in. WC (7.5 Pa).
 When Tru drops, the chilled water flow to the cooling coil is reduced first, and then the outdoor air
free cooling, outdoor air volume flow rate.
Refer to Sec. 21.2 for the details of the sequence of operations of single-zone VAV cooling
systems. In the design and operation of an EMCS system, the necessary documentation includes the
sequence of operation, control diagrams, specifications, operation, and maintenance manual.
5.6 CHAPTER FIVE
FIGURE 5.3 Block diagram for a closed-loop feedback control system.
Control Methods
According to the types of control signal and the different kinds of energy used to transmit the
signals, as well as whether a software is used during control operation, control methods can be
classified as direct digital, pneumatic, electric, and electronic.
Analog and Digital. There are two types of control signals: analog and digital. An analog signal
is in the form of a continuous variable. It often uses the magnitude of electric voltage or pneumatic
pressure to represent the air temperature. A digital signal is a series of on and off pulses used to
transmit information.
A conventional analog controller receives a continuous analog signal, such as a voltage or a
pneumatic signal, that is proportional to the magnitude of the sensed variable. The controller compares
the signal received from the sensor to the desired value (i.e., the set point) and sends a signal
to the actuator in proportion to the difference between the sensed value and the set point.
A digital controller, or microprocessor-based controller, receives an electric signal from
sensor(s). It converts the electric signal to digital pulses of different time intervals to represent the
signals values. The microprocessor of the digital controller performs the mathematical operations
and knowledge processing on these values. The output from the microprocessor can be either in
digital form to actuate relays or converted to an analog signal (say, a voltage or a pneumatic pressure)
to operate the actuator(s).
Direct Digital Control (DDC). A control system using DDC involves adopting a microprocessorbased
digital controller to perform mathematical operations and knowledge processing according to
the predetermined control algorithms or computer programs. The key element of DDC compared to
analog control is the software and hardware contained in the direct digital controller which expands
the control functions tremendously and adopts recently developed control logic. ADDC unit usually
has more precise sensors and uses the same type of controlled devices as other control methods.
Figure 1.2 shows an energy management and control system using DDC for an air-handling unit
in a typical floor of the NBC Tower, and Fig. 5.2 shows an EMS with DDC for a single-zone VAV
system.
Pneumatic Control. In a control system using pneumatic control, compressed air is used to operate
the sensors, controllers, and actuators and to transmit the signals. It consists of: a compressed
air supply and distribution system, sensors, controllers, and actuators. Figure 5.4 shows a typical
pneumatic control system. In Fig. 5.4, a filter is used to remove the dust particles, including submicrometer-
size particles, contained in the air. The function of the pressure-reducing valve is to reduce
the pressure of compressed air discharged from the air compressor to the required value in the
main supply line. The discharged compressed air is usually at a gauge pressure of 18 to 25 psig
(124 to 172 kPag), and the pressure signal required to actuate the valve or damper actuator is 3 to
13 psig (20.7 to 89.6 kPa g).
The advantages of pneumatic control are as follows:
 The compressed air itself is inherently a proportional control signal.
 The cost of modulating actuators is low, especially for large valves and dampers.
 Pneumatic controls require less maintenance and have fewer problems.
 Pneumatic controls are explosionproof.
The disadvantages stem mainly from comparatively fewer control functions, the high cost of
sophisticated pneumatic controllers, and the comparatively higher first cost of a clean and dry,
compressed air supply for small projects.
Electric or Electronic Control. Both types of control use electric energy as the energy source for
the sensors and controllers. A control system using electric control often offers two-position on-off
control, as shown in Fig. 5.1. Switches, relays, contactors, and electromechanical devices are
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.7
system components for electric control systems. They are generally used for low-cost, small, and
simpler control systems. In addition to the switches and relays, a control system using electronic
control has transistors, diodes, capacitors, and printed-circuit boards as system components. Electronic
control systems always have more accurate sensors and solid-state controllers with sophisticated
functions and can be easily interfaced with the building automation system. Electronic control
has a faster response and more accurate processing of data than electric control systems. Electronic
control systems cost more and need skilled personnel for maintenance and troubleshooting.
Comparison of Control Methods
Because of the increasing demand for more complicated controllers to satisfy the needs of better
indoor environmental control, satisfactory indoor air quality, improved energy saving, lower cost,
and greater reliability, the recent trend is to use EMCS with direct digital control for more demanding
and large projects. A modern DDC system consists of electronic sensors, microprocessor-based
controllers incorporated with electronic components, and electronic or pneumatic actuators. It is often
more effective and cheaper to operate a large pneumatic actuator than to use a large electronic
actuator in a DDC system.
An EMCS using DDC offers the following advantages:
1. Flexibility of providing required control functions and the ability to coordinate multiple functions
directly from complicated software programs (it can even mimic a human expert within a
certain knowledge domain and offer artificial intelligence)
2. More precise and faster-response control actions provided by the microprocessor-based controller
3. The possibility of using high-level self-checking and self-tuning system components, which
increases the system reliability
5.8 CHAPTER FIVE
FIGURE 5.4 A typical pneumatic control system.
One of the main disadvantages of DDC is a still-higher first cost. However, the cost per control
action is comparatively lower. As more and more DDC systems are installed, DDC will become
familiar to us. The cost of the microprocessor-based controllers and other DDC system components
will drop further in the future. In the 1990s, the trend in HVAC&R control for new and retrofit
projects is to use EMCS with DDC systems except for small projects.
5.3 CONTROL MODES
Control modes describe how the corrective action of the controller takes place as well as its effect
on the controlled variable. For applications in HVAC&R, control modes can be classified as twoposition,
step, floating, proportional, proportional-integral, and proportional-integral-derivative.
Before we discuss the control modes in detail, the term lag, or time lag, should be introduced.
According to ASHRAE terminology, lag is (1) the time delay required for the sensing element to
reach equilibrium with the controlled variable; or (2) any retardation of an output with respect to
the causal input, including the delay, because of the transport of material or the propagation of a
signal.
Two-Position Control
In two-position control, the controller controls the final control element at one of two positions:
maximum or minimum (except during the short period when it changes position). Examples of twoposition
control include starting and stopping the motor of a fan, pump, or compressor and turning
on or off an electric heater. Sometimes it is called on-off control.
Figure 5.5 shows a two-position control mode for an electric heater installed in a branch duct. In
the middle diagram, the ordinate indicates the controlled variable, the space temperature Tr; in the
lower diagram, the ordinate denotes the output capacity of the final control element, the electric
heater. If the controller turns on the electric heater when the sensor senses a space temperature
69.5F (20.8C), and turns off at 70.5F (21.4C), the result is a cyclic operation of the electric
heater and the rise and fall of Tr toward the two positions of 69.5F (20.8C) and 70.5F (21.4C).
The thermal storage effect of the electric heater, the branch duct, the building envelope that surrounds
the space, and the sensor itself will have a time lag effect on Tr. The rise of Tr is a convex
curve with an overshoot higher than 70.5F (21.4C), and the drop of Tr is a concave curve with an
undershoot lower than 69.5F (20.8C). The difference between the two points, on and off, is called
the differential.
If the heating capacity of the electric heater that results from an on-and-off cyclic operation is
represented by Qe,t, and if the actual space heating load , the slope of the rising Tr curve
will then be greater than the slope of the falling Tr curve. As a consequence, the on period will be
shorter than the off period. Such a cyclic Tr curve is shown in the upper left corner of Fig. 5.5. If
, the condition will be reversed (see the upper right corner of Fig. 5.5). The two-position
control varies the ratio of on and off periods to meet any variation in the space heating load.
To reduce the overshoot and undershoot of the controlled variable in a two-position control
mode, a modification called timed two-position control has been developed. A small heating
element attached to the temperature sensor is energized during on periods. This additional heating
effect on the sensor shortens the on timing. During off periods, the heating element is deenergized.
The differential of two-position control, the overshoot, and the undershoot all result in a greater
fluctuation of the controlled variable. A suitable differential is always desirable in two-position control
in order to prevent very short cycling, which causes hunting, a phenomenon of short cycling of
the controlled variable. Two-position control is not suitable for precise control of the controlled
variable, but it is often used for status control, such as opening or closing a damper, turning a small
single piece of equipment on or off for capacity control, etc., and for lower-cost control systems.
Qrh  1
2Qe,t
Qrh 	 1
2Qe,t
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.9
Step Control and Modulating Control
In step control, the controller operates the relays or switches in sequence to vary the output capacity
of the process plant in steps or stages. The greater the deviation of the controlled variable from the
set point and the faster the rate of change of the controlled variable, the higher will be the output
capacity of the process plant. In modulating control, the controller activates the control device continuously
and thus the change of the output capacity gradually.
During the earlier stage of step control, there must be a differential of controlled variable between
two on or two off points of a particular piece of equipment of the process plant. The result is a
greater fluctuation of the controlled variable. In the DDC controller-activated new generation of step
control, the software of the DDC unit determines how fast each capacity step is added or subtracted
according to the deviation from the set point and the rate of change of the controlled variable.
Figure 5.6 shows step control of the cooling capacity of refrigeration compressors using DDC to
maintain the discharge air temperature Tdis within predetermined limits in a packaged unit. When
the air economizer alone can no longer balance the DX coils load, and Tdis floats to the upper limit
of the control band, point 1, the first stage of cooling is energized. If the cycling and the continuously
energized first-stage cooling capacity still cannot balance the coils load, then Tdis continues to
rise until it reaches point 2, which is 1F (0.56C) higher than the upper limit of the control band.
The first-stage cooling is then locked on, and the second stage is energized to cycling. When Tdis
rises to point 2, all the present cooling stages will be locked on, and an additional cooling stage is
added for cycling. When Tdis drops to point 7, the current cycling cooling stage will be deenergized,
and the one cooling stage next to the deenergized stage will be cycling. It always needs a time
delay, say at least 4 min, to turn on or off again in order to prevent hunting and other damage.
5.10 CHAPTER FIVE
FIGURE 5.5 A typical two-position control mode.
Step control is another kind of on-and-off control having smaller varying capacity and thus a
fluctuating controlled variable. It has been widely used for the capacity control of refrigeration
compressors and the electric heaters.
Floating Control
In a floating control mode, the controller moves the control device by means of the actuator toward
the set point only when the control point is out of the differential, or dead band, as shown in
Fig. 5.7. A control point is the actual value of the controlled variable at a certain time instant.
Figure 5.7 represents a duct static pressure control system using a floating control mode by opening
and closing the inlet vanes of a supply fan. The control deviceinlet vanescan be moved in either
an opening or a closing direction depending on whether the control point is over the upper limit
of the differential or under the lower limit. In Fig. 5.7, when the controlled variable, the duct static
pressure ps, is above the upper limit of the differential, the controller then closes the inlet vanes. If
ps is below the lower limit of the differential, the controller opens the inlet vanes.
A floating control mode is more suitable for control systems with a minimal lag between the
sensor and the control medium. A control medium is the medium in which the controlled variable
exists.
Proportional Control
In a proportional control mode, the controller moves the controlled device to a position such that
the change in its output capacity is proportional to the deviation of the controlled variable from the
set point. The position of the controlled device is linearly proportional to the magnitude of the controlled
variable. Figure 5.8 represents the control of the space temperature through the modulation
of a two-way valve of a cooling coil using proportional control. In Fig. 5.8, the controlled variable
is the space temperature Tr, and the controlled device is the valve.
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.11
FIGURE 5.6 DDC-activated discharge air temperature control for a VAV rooftop packaged system.
In a proportional control mode, the throttling range is the change in the controlled variable when
the controller moves the controlled device from the position of maximum output to the position of
minimum output. The controlled variables range of values that will move the proportional controller
through its operating range is called the proportional band. In a proportional control system,
the throttling range is equal to the proportional band.
5.12 CHAPTER FIVE
FIGURE 5.7 Floating control mode.
FIGURE 5.8 Proportional control mode.
The set point is the desired value of the controlled variable, or the desired control point that the
controller seeks to achieve. The difference between the control point and the set point is called
the offset, or deviation. In proportional control, when the controlled variable is at the bottom line of
the throttling range, the controller will position the actuator at the closed position. At the set
point, the actuator will be at 50 percent of the open position. When the controlled variable is at the
top of the throttling range, the actuator will be at 100 percent of the open position.
In a proportional control mode, since the output signal V of the controller is proportional to the
deviation of the control point from the set point, their relationship can be expressed as:
(5.1)
where Kp  proportional gain, proportional to 1 / (throttling range)
e  error signal, i.e., deviation or offset
M  output value when deviation is zero (usually, output value at middle of output range of
controller)
For space and discharge air temperature control using proportional control, the offset is directly
proportional to the space load and the coil load. The space or discharge temperature T, F (C), can
be calculated as:
(5.2)
where Tt,r  throttling range, expressed in terms of space or discharge temperature, F (C)
Tmin  space and discharge temperature when space load or coil load is zero, F (C)
In Eq. (5.2), Rload represents the load ratio of the space load or coil load, which is dimensionless and
can be evaluated as
(5.3)
The actual load and design load must be in the same units, such as Btu/h (W).
For a proportional control mode, a certain degree of offset, or deviation, is inherent. The position
of the controlled device is a function of the offset. Only when offset exists will the controller
position the actuator and the valve or damper at a position greater or smaller than 50 percent open.
The throttling range is primarily determined by the HVAC&R system characteristics and cannot be
changed after the system is designed. A proportional control mode is suitable for an HVAC&R system
that has a large thermal capacitance, resulting in a slow response and stable system and allowing
comparatively narrower throttling range and thus a smaller offset. Fast-reacting systems need a
large throttling range to avoid instability and short cycling, or hunting.
Proportional plus Integral (PI) Control
In a proportional plus integral control mode, a second component, the integral term, is added to the
proportional action to eliminate the offset. The output of the controller can thus be expressed as
(5.4)
where Ki  integral gain and t  time.
In Eq. (5.4), the second term on the right-hand side (the integral term) indicates that (1) the error,
or offset, is measured at regular time intervals and (2) the product of the sum of these measurements
and Ki is added to the output of the controller to eliminate the offset. The longer the offset
exists, the greater the response of the controller. Such a control action is equivalent to resetting the
set point in order to increase the controller output to eliminate the offset. As such, PI control is
sometimes called proportional plus reset control. Figure 5.9a shows the variations of a controlled
V  Kpe 	 Kie dt 	 M
Rload 
actual load
design load
T  RloadTt,r 	 Tmin
V  Kpe 	 M
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.13
variable for a proportional-integral control mode. In PI control, proportional-integral control signals
are additive.
For PI control, the controlled variable does not have any offset once it has achieved a stable condition,
except due to any inaccuracy of the instrumentation and measurement. Proper selection of
the proportional gain Kp and integral gain Ki as well as proper tuning is important for system stability
and control accuracy. PI control may be applied to fast-acting control systems with a greater
throttling range setting at the controller for better system stability, e.g., discharge air temperature,
discharge chilled water temperature, and duct static pressure control systems.
Proportional-Integral-Derivative (PID) Control
A PID control mode has additional control action added to the PI controller: a derivative function
that opposes to any change and is proportional to the rate of change. The output of such a controller
can be described by the following equation:
(5.5) V  Kpe 	 Kie dt 	 K
de
dt
	 M
5.14 CHAPTER FIVE
FIGURE 5.9 Proportional-integral and proportional-integral-derivative
control modes. (a) Proportional-integral; (b) proportional-integral-derivative.
where K is the derivative gain. The effect of adding the derivative function K de/dt is that the
quicker the control point changes, the greater the corrective action provided by the derivative function.
Figure 5.9b shows the variation of the controlled variable for a PID control mode. As with a PI
control mode, PID control mode also has no offset once the controlled variable has reached a stable
condition, except due to instrument inaccuracy.
Compared to the PI control mode, the PID control mode, which is a combination of proportional,
integral, and derivative actions, exhibits faster corrective action and a smaller overshoot and
undershoot following an offset and a change of the controlled variable. The controlled variable is
brought to the required set point in a shorter time. However, it is more difficult to determine properly
three constants, or gains (Kp, Ki, and K).
Compensation Control, or Reset
Compensation control, or reset, is a type of control mode in which a compensation sensor is generally
used to reset a main sensor to compensate for a variable change sensed by the compensation
sensor. The purpose is to achieve operation that is more effective, energy-efficient, or both.
In the design of a reset mode, the first things to decide are the control point at which the main
sensor will be reset and the variable to be sensed by the compensation sensor. The main sensor,
which senses the mixed air temperature in the air-handling unit, is usually reset by a compensation
sensor that senses the outdoor temperature, as shown in Fig. 5.10.
Another task to decide on is the relationship between the variables sensed by the main and compensation
sensorsthe reset schedule. Many reset schedules have a different relationship between
these two sensors at various stages. For example, in Fig. 5.10, when the outdoor temperature To
is less than 30F (1.1C), it is within the range of stage I. No matter what the magnitude of To is,
the set point of the mixed temperature Tm is 65F (18.3C). When 30F  To  95F (1.1C 
To  35C), it is in stage II. The linear relationship between the temperatures sensed by the main
and compensation sensors in Fig. 5.10 can be expressed by the following equation:
(5.6)
When To  95F (35C), the set point of the mixed temperature is always 55F (12.8C).
Compensation control modes have been widely adopted to reset space temperature, discharged
air temperatures from air-handling units or packaged units, or water discharge temperatures from
central plants.
Applications of Various Control Modes
Selection of a suitable control mode depends on
 Operating characteristics
 Process or system characteristics, such as whether the thermal capacitance should be taken into
consideration
 Characteristics of load changes
If a simpler control mode can meet the requirements (say, two-position control versus PID control),
the simpler control mode is always the first choice. Except for two-position and step control, all the
other control modes are modulation controls. Modulation control is a control mode that is capable
of increasing or decreasing a variable according to the deviation from the required value in small
increments continuously. Also note that there is a significant difference between the two-position
control, step control, floating control, and proportional control modes with an offset and the offset
free PI and PID controls.
Tm  65 
65 55
95 30
(To 30)
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.15
For HVAC&R processes, PI control can satisfy most of the requirements, and it is most widely
used. Microprocessor-based PID control is a very powerful tool. PID control mode is more appropriate
for fast-acting duct static pressure control and airflow control, and it is recommended to set
the controller with a large proportional band for control system stability, a slow reset to eliminate
deviation, and a derivative action to provide a quick response.
5.4 SENSORS AND TRANSDUCERS
A sensor is a device that acts as a component in a control system to detect and measure the
controlled variable and to send a signal to the controller. A sensor consists of a sensing element and
accessories, as shown in Fig. 5.11. The term sensing element often refers to that part of the sensor
that actually senses the controlled variable. In HVAC&R systems, the most widely used sensors are
temperature sensors, humidity sensors, pressure sensors, and flow sensors. Recently, CO2 sensors,
air quality sensors, and occupancy sensors are being used in many new and retrofit projects. Electronic
sensors that send electric signals to electronic controllers, can also be used for the DDC
units, and the current trend is to use solid-state miniaturizing sensing elements. Electric output from
the sensors is usually expressed in 0 to 10 V dc or 4 to 20 mA. In the selection of sensors, accuracy,
sensitivity of response, reliability, long-term stability or drift, required calibration intervals, maintainability,
and especially the possibility of contamination by dust particles due to contact with the
control medium should be considered. Drift is a kind of offset. ASHRAEs Terminology defines
drift as change in an output-input relationship over time with the change unrelated to input, environment,
or load. A calibration period of 1 year and longer is often considered acceptable.
Location of the sensor affects the sensor output directly. A sensor should be located in the critical
area where the controlled variable needs to be maintained within required limits. The sensing
element of an air sensor should be well exposed to the air, so that air can flow through the sensor
without obstruction. Air sensors should not be affected directly by the supply airstream nor should
space air sensors be affected by the outdoor airstream. Sensors should be shielded from radiation
5.16 CHAPTER FIVE
FIGURE 5.10 Reset control and schedule.
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.17
FIGURE 5.11 (a) A temperature sensor and (b) a typical pitot tube flow-measuring station.
and mounted firmly on a structural member or duct wall, free from vibration. If a space air sensor
must be located on a concrete column, thermal insulation should be provided between the sensor
and the column.
Temperature Sensors
Temperature sensors fall into two categories: those that produce mechanical signals and those that
emit electric signals. Bimetal and rod-and-tube sensors that use sensing elements to produce a mechanical
displacement during a sensed temperature change either open or close an electric circuit in
an electric control system, as shown in Fig. 5.1, or adjust the throttling pressure by means of a
bleeding nozzle in a pneumatic control system. For a sealed bellow sensor, a change in temperature
causes a change in the pressure of a liquid in a remote bulb. The expansion and contraction of vapor
then move the mechanism of the controller.
Temperature sensors that produce electric signals, as shown in Fig. 5.11a, are the same as the
sensors for temperature measurement and indication mentioned in Sec. 2.3.
In addition to resistance temperature detectors (RTDs) and thermistors, sensors sometimes use
thermocouples. A thermocouple uses wires of two dissimilar metals, such as copper and constantan,
or iron and nickel, connected at two junctions, to generate an electromotive force between the junctions
that is directly proportional to the temperature difference between them. One of the junctions
is kept at constant temperature and is called the cold junction. Various systems have been developed
to maintain the cold junction at a constant temperature and to provide compensation if the cold
junction is not at 32F (0C). This task makes the use of thermocouples more complicated and
expensive. The electromotive force produced between the two junctions can be used as the signal
input to a controller.
Bimetal and rod-and-tube temperature sensors are simple and low in cost. However, they cannot
provide temperature indication and electric signals for DDC and are often used in electric control
systems. Platinum and nickel RTDs are stable, reliable, and accurate. They are very expensive compared
to thermistors and need calibration to compensate for the effects of having external leading
wires. RTDs are widely adopted in DDC for commercial applications. For a project that needs
precise temperature control, RTD is often the choice. High-quality thermistors can provide stable,
reliable, and interchangeable temperature sensors, and are also widely used in many commercial
applications. Petze (1986) reported that some high-quality thermistors exhibited better than

0.002F (
0.001C) stability for a 2-year period.
A typical space air temperature sensor has the following characteristics:
Sensing element Platinum film element, 3000 
Positive temperature coefficient, 4.8 /F
(8.64 /C)
Operating range 60 to 90F (15.6 to 32.2C)
Humidity Sensors
As mentioned in Sec. 2.8, humidity sensors fall into two categories: mechanical and electronic.
When the same humidity sensor is used for both monitoring and DDC, the ion-exchange resistancetype,
also called the bulk polymer resistance-type, and capacitance-type humidity sensors are often
adopted. Humidity sensors usually have different accuracy at very low, 20 to 80 percent mid-range,
and high relative humidities. The bulk polymer resistance humidity sensor is not accurate at low
relative humidities and generally provides stable and accurate readings within a range of 30 to 90
percent relative humidity. Its performance is affected by the air contamination. Capacitance humidity
sensors are accurate at 10 to 80 percent relative humidity. However, they become unstable at
high relative humidities.
5.18 CHAPTER FIVE
In an EMCS, an electronic humidity sensing element measures the relative humidity and sends
an electric signal to the DDC unit. A typical space air humidity sensor has the following specifications:
Power source 24 V ac
Humidity range (active) 10 to 80 percent
Nominal output range 0 to 5 V dc
Accuracy at 70F (21.1C) 
3 percent at 10 to 60 percent range

4 percent at 60 to 80 percent range
Speed of response 8 min (90 percent of response time)
Mounting Wall mount
Pressure Sensors
A pressure sensor usually senses the difference in pressure between the controlled medium (air and
water) and a reference pressure; or the pressure differential across two points, such as the pressure
differential across a filter. The reference pressure may be an absolute vacuum, atmospheric pressure,
or the pressure at any adjacent point. Output signals from the pressure sensors may be electric
or pneumatic, and analog or binary.
Pressure sensors used for HVAC&R systems can be divided into high-pressure and low-pressure
sensors. High-pressure sensors measure in pound per square inch or feet WC (kilo pascals), and
low-pressure sensors measure in inches WC (pascals). The sensing elements for high-pressure sensors
are usually Bourdon tube, bellows, and sometimes diaphragms. For low-pressure air sensors,
large diaphragms or flexible metal bellows are usually used.
To measure the duct static pressure, if a long section of straight duct of a length greater than
10 ft (3 m) is available, a single-point, pitot-tube type of duct static pressure sensor can be used.
Otherwise, a multipoint pitot-tube array or flow-measuring station with airflow straighteners should
be used. The small hole used to measure static pressure should never be directly opposite to an
airstream with a velocity pressure that can affect the reading. A reference pressure should be picked
up at a point with low air velocity outside the duct, at a point served by the same air system, or in
the ceiling plenum.
In a DDC system, low air pressure is often sensed by measuring the capacitance of two
diaphragms; one is allowed to move toward or away from the fixed one, depending on the pressure
on two sides of the diaphragm. An electric circuit is used to convert the capacitance to a voltage or
milliampere signal. The other method used to sense the low air pressure in DDC systems works because
the air pressure differential compresses or stretches a diaphragm as well as a strain gauge.
The change of the resistance of the gauge is detected and amplified, and this electric signal in the
form of voltage or milliamperes is sent to the controller.
A space air pressure sensor having an operating range of 	0.1 to 0.1 in WC (	25 Pa to 25
Pa) detects the resistance of a silicon diaphragm, and thus the space air pressure is then measured.
Stainless steel, rubber, etc., can also be used as the material of the diaphragms. A space air pressure
sensor should be located in an open area of the conditioned space where the air velocity is less than
40 fpm (0.2 m/s) and where its reading is not affected by the opening of the doors. The reference
pressure pickup is best located at the rooftop, at a level 10 ft (3 m) above the building to avoid the
influence of wind.
Flow Sensors
Flow sensors usually sense the rate of air flow and water flow in cfm (L/s) for air and gpm (L/ s)
for water. For airflow sensors, the average velocity pressure pv, in. WC (Pa), is often sensed and
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.19
measured. The average air velocity va, fpm, is then calculated as
(5.7)
where K is the flow coefficient, which depends on the type of pitot-tube array used and the dimension
of the round or rectangular duct. After that, the volume flow rate can easily be determined as
the product vaA. Here, A represents the cross-sectional area of the air passage, perpendicular to the
airflow, in ft2 (m2). Various forms of pitot-tube array have been developed and tested to determine
the average velocity pressure of a rectangular or circular duct section by measuring the difference
between the total and static pressures of the airstream. A typical pitot-tube flow-measuring station
with flow straighteners to provide more even airflow is shown in Fig. 5.11.
Electronic air velocity sensors such as hot-wire anemometers and thermistors have also been
widely adopted to measure airflow, especially for variable-air-volume (VAV) boxes. Thermistors are
cheaper than hot-wire anemometers. Heated thermistors need periodic calibration.
Water flow sensors of the differential-pressure type, such as orifice plates, flowing nozzles, and
pitot tubes, have only a limited measurement range. Turbine or magnet-type flowmeters can apply
to a wider range, but they are more expensive and also need periodic calibration.
Carbon Dioxide and Air Quality Sensors
A CO2 sensor detects and indicates the amount of carbon dioxide (CO2 ) contained inside the air,
which is a reliable indication of the body odor released by the occupants of the conditioned space.
Because of a certain relationship that exists between the CO2 contained in the outdoor air and
the CO2 released by the occupants, the concentration of CO2 in space air may sometimes be used as
a rough indication of amount of outdoor ventilation air supply to the conditioned space under the
specific conditions.
The sensing process used in CO2 sensors includes potentiometric and amperometric electrochemical
cells, an infrared detector, etc. A CO2 sensor usually has an operating range of 0 to 3000
ppm. For measurements within an accuracy of 
100 ppm, recalibration may be required on the
order of once per year.
An air quality sensor, also called a volatile organic compound (VOC) or a mixed-gas sensor, is
used to monitor and detect the relative concentrations of VOC or mixed gas, or total concentration
including acetone, ammonia, CO, CO2, SO2 , chlorine, formaldehyde, CFC-11, CFC-12, etc. The
concentration of the contaminant is often expressed in tested and commissioned units, such as 0 to
5 or 0 to 10 units.
The sensing element has a tin dioxide surface which is heated to a temperature above 130F
(54C). The change of the conductivity and thus its resistance are then amplified and fed to the controller
in terms of 0 to 10 V dc in proportion to the contamination. Air quality sensors are less
expensive than CO2 sensors and need less maintenance. The drift of an air quality sensor is unpredictable.
Both CO2 sensors and air quality sensors are now used for demand-based ventilation control to
provide the required amount of outdoor air for acceptable indoor air quality.
Occupancy Sensors
An occupancy sensor detects whether a room is occupied by occupant(s). As a result, the HVAC&R
and lighting in this room can be turned off when the room is not occupied, to save energy. There are
two types of occupancy sensors: ultrasonic and infrared. An ultrasonic occupancy sensor sends out
a rather lower level of ultrasonic wave and detects movement of the occupant when there are
changes in receiving patterns. Ultrasonic sensors sweep over the area that surrounds the sensor.
False signals may be received due to the motion of the papers or air movement from a diffuser. The
sensitivity of the ultrasonic sensor should be adjusted to avoid these problems.
va  4005K ?pv
5.20 CHAPTER FIVE
An infrared occupancy sensor senses the movement of the occupant(s) as the sensor receives
heat from the occupant when that person is moving. Infrared sensors need to see the occupant. As
long as there is no obstruction between the sensor and the occupant in the room, the sensing process
of an infrared sensor is effective. Occupancy sensors are often mounted on the ceiling or at a high
level on a wall. They are used in the guest rooms of the hotels and motels and other commercial
buildings. Ultrasonic occupancy sensors show a wider acceptance than infrared occupancy sensors.
Wireless Zone Sensors and Intelligent Network Sensors
In an HVAC&R system, a wireless zone sensor is an indoor spread spectrum radio-frequency transmitter
which sends room temperature and other status information to a local receiver located no
more than 1000 ft (300 m) away. A translator converts these data and sends them to a variable-airvolume
DDC unit controller through a wired communications link. This type of new technology
was developed recently for the sake of providing flexibility for the rearrangements of the office
layout.
Kovacs (1996) noted that intelligent network sensors . . . can linearize the sensor signal, accept
an offset adjustment through the network, may have alarm- or decision-making algorithms on
board, and include self-test diagnostics to continuously validate performance. An intelligent network
sensor may have a sensing element, an analog/ digital transducer, a neutron chipan onboard
microprocessor, and a transceiver. Dew-point temperature, enthalpy, wet-bulb temperature, etc. can
be provided as a combined variable. An intelligent network sensor will help to drive the system
structure towards becoming a more cost-effective distributed control system in the future, a system
that has more locally processed signal and data.
Transducers or Transmitters
A transducer is a device that converts energy from one form to another or amplifies an input or output
signal. In HVAC&R control systems, a transducer may be used to convert an electric signal to a
pneumatic signal (E/P transducer), e.g., a pneumatic proportional relay that varies its branch air
pressure from 3 to 15 psig (20 to 103 kPa g) in direct proportion to changes in the electrical input
from 2 to 10 V. Also E/P transducers are used between microprocessor-based or electronic controllers
and pneumatic actuators. However, a pneumatic signal can be converted to an electric signal
in a P /E transducer. For example, a P/ E relay closes a contact when the air pressure falls and opens
the contact when the air pressure rises above a predetermined value. A transmitter is used to transmit
a signal, pneumatic or electric; through air, water, or other fluids. A sensor is also a transmitter.
However, the difference between them is that a sensor only senses the signal of the controlled variable
and transmits it.
5.5 CONTROLLERS
A controller receives input from the sensor, compares with the set point or implemention based on its
stored computer software, sends an output to or modulates the control device for maintaining a desirable
indoor environment. A thermostat is a combination of a temperature sensor and a temperature
controller, whereas a humidistat is a combination of a humidity sensor and a humidity controller.
Direct-Acting or Reverse-Acting
A direct-acting controller increases its output signal upon an increase in the sensed controlled variable,
and it decreases its output signal upon a decrease in the sensed controlled variable. Conversely,
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.21
a reverse-acting controller decreases its output signal upon an increase in the sensed controlled
variable and increases its output signal upon a decrease in the controlled variable.
Normally Closed or Normally Open
A controlled device, a valve or a damper, that is said to be in the normally closed (NC) position indicates
that when the input signal to the controller falls to zero or below a critical value, the control
device will be in the closed position. The closing of the valve or damper is most likely due to the
action of a spring or a supplementary power supply. A valve or a damper that is said to be normally
open (NO) indicates that when the input signal to the controller falls to zero or below a critical
value, then the valve, damper, or associated process plant will be in the open position.
Pneumatic Controllers
In a pneumatic controller, the basic mechanism used to control the air pressure in the branch line
supplied to the actuators is a nozzle-flapper assembly plus a restrictor. Figure 5.12 shows a pneumatic
controller with such a mechanism. Compressed air supplied from the main line flows through
the restrictor and discharges at the opening between the nozzle and the flapper, which has a spring
pulling it downward. The nozzle and restrictor are sized in such a manner that when the flapper
moves away from the nozzle, all the air escapes from the nozzle and the branch line pressure is
zero. When the flapper covers the nozzle or if there is no airflow in the branch line, the pressure of
the branch line is equal to that in the main line.
If a sensor, such as a bimetal sensor, moves the flapper upward according to the magnitude of
the controlled variable sensed by the sensor, then the input signal from the sensor determines
the opening between the nozzle and flapper and hence the compressed air pressure in the branch
line.
5.22 CHAPTER FIVE
FIGURE 5.12 A pneumatic controller with a nozzle-flapper assembly
and a restrictor.
For a direct-acting pneumatic temperature controller used to control the space temperature
during summer, the branch line pressure may change from 9 to 13 psig (62 to 89 kPag) when the
space temperature increases from 73 to 77F (22.8 to 25C). This nozzle-flapper assembly pneumatic
controller operates in proportional control mode. Many other more complicated pneumatic
controllers have been developed to perform other control modes and additional functions.
Electric and Electronic Controllers
An electric controller uses switches, relays, and a bridge circuit formed by potentiometers to position
the actuators in on-off, floating, and proportional control modes according to the input signal
from the sensor and the predetermined set point. An electronic controller can provide far more
functions than electric controllers can. It may receive input signals from both the main sensor and
the compensation sensor with amplification and combination.
In the control circuit, an electronic controller basically provides proportional or proportionalintegral
control modes. The output signal from the controller can be used to position an actuator or
to provide the sequencing of actuators, or to change to two-position, floating, or even PID control
modes in conjunction with additional circuits.
Direct Digital Controllers
A direct digital controller has a microprocessor to implement computer programs to provide various
control functions. In DDC units, there are analog-to-digital (A/D) and digital-to-analog (D/A)
converters to convert analog input to digital signals for processing, or to convert digital signals to
analog for actuators, if necessary. DDC units are stand-alone and microprocessor-based controllers.
Stand-alone means that the controller has sufficient capacity to execute the assigned control
functions alone. Today, there are mainly two types of DDC units: system controllers and unit controllers.
System Controllers. A system controller, also called a stand-alone panel (SAP), as shown in
Fig. 5.13, has the ability to coordinate communications between system controllers, between the
system controller and the personal computer (PC) in the workstation, and between the system controller
and the supported unit controllers. It also has the ability to provide and execute the control
programs for functional control, and to store user databases and trend log values. A system controller
can support 50 to 200 unit controllers on separate unit controller trunk(s).
Unit Controllers. A unit controller, also called a terminal controller, is shown in Fig. 5.14. It usually
has limited capacity to execute factory-loaded computer programs and to provide functional
control for a terminal or a piece of HVAC&R equipment. Unit controllers are often connected in a
separate network and supported by a system controller. The new-generation unit controllers have
greater memory to handle complicated control programs, and they provide time and calendar scheduling,
data storage, and other functions, such as limited programming.
Hardware. Many system controllers have only a single printed-circuit board. Single-board configuration
often offers lower first cost. Its disadvantage is that any component failure requires the
replacement of the complete board. Another approach is that the controller is made from various
modules. The modular approach isolates the component failures and plugs on the single-module
replacement quickly and inexpensively.
Memory. The types of memory included in DDC units are as follows:
 Read-only memory (ROM), which stores the software provided by the manufacturer and should
not be modified by the user.
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.23
 Random-access memory (RAM), which stores custom control software developed during installation
or prepared by the user. This type of memory is volatile (i.e., it can be read from and written to)
and requires battery backup.
 Electric erasable programmable read-only memory (EEPROM), which stores custom control software
and is volatile. The advantage of EEPROM over RAM is that EEPROM does not need
5.24 CHAPTER FIVE
Universal analog
inputs;
Independent power
supplies for each
electronic module;
Tool-less installation;
Onboard diagnostic
displays and indications.
Context-sensitive
touchpad;
user interface.
Wireway space
Fully integrated
output relays,
overrides and
transducers;
FIGURE 5.13 A typical system controller. (Source: Johnson Controls. Reprinted by permission.)
battery backup. However, EEPROM cannot be used for as many writing, erasing, and rewriting
cycles as RAM.
 Flash erasable programmable read-only memory (flash EPROM), which is a kind of new memory
technology and allows the stored control software to remain untouchable indefinitely without
power.
System controllers often use a 16-bit microprocessor. A typical system controller has the following
memories:
RAM: 256 Kbytes
ROM: 128 Kbytes EEPROM
512 Kbytes flash EPROM
Input/Output (I/O). There are four kinds of I/O: analog input (AI); digital or binary input (BI);
analog output (AO); and pulsed or binary output (BO). Conventionally, a sensor input, a controller
output, or a control value, is referred to as a point or object. Current system controllers allow their
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.25
FIGURE 5.14 A typical unit controller. (Source: Honeywell Inc. Reprinted by
permission.)
input and output connections to be configured with great flexibility. Each input or output point can
be either analog or digital. Typical analog inputs (electric signal) are 0 to 10 V dc or 4 to 20 mA. A
system controller often has a total of 20 to 50 I /O points. Some system controllers can be extended
to 100 points if necessary. A typical system controller has the following I/O points capacity:
Analog/ digital inputs 18
Universal analog / digital input /outputs 6
Digital outputs 12
Totalizer inputs, i.e., pulsed inputs 4
A unit controller or a terminal controller usually has 4 to 20 I /O points. The I/O points in a typical
unit controller may take the following forms:
Analog inputs 0 to 10 V dc
Digital inputs Switch, relay, and transistor
Microbridge sensor 0 to 3 in. WC pressure differential (0 to 750 Pa)
Analog outputs 0 to 10 V dc, 4 to 20 mA
Digital outputs 30 V ac
Pneumatic 3 to 15 psig (20.6 to 103 kPag)
Triac On/off output for electric heater, fan motor, etc.
5.6 WATER CONTROL VALVES AND VALVE ACTUATORS
Water valves are used to regulate or stop water flow in a pipe either manually or by means of automatic
control systems. Water control valves adopted in water systems can modulate water flow rates
by means of automatic control systems.
Valve Actuators
An actuator, sometimes called an operator, is a device which receives an electric or pneumatic analog
control signal from the controller, either directly or through a digital-to-analog converter. It then
closes or opens a valve or damper, modulating the associated process plant, and causes the controlled
variable to change toward its set point. Valve actuators are used to position control valves.
They are mainly of the following types:
Solenoid Actuators. These use a magnetic coil to move a movable plunger connected with the
valve stem. Most solenoid valve actuators operate at two positions (on and off). They are used
mainly for small valves.
Electric Actuators. These move the valve stem by means of a gear train and linkage. Different
electric motor valve actuators can be classified according to the control mode they use:
1. On/off mode. For this type of actuator, the motor moves the valve in one direction, and when the
electric circuit breaks, the spring returns the valve stem to the top position (either open or closed
position depending on whether it is a normally open or closed valve).
2. Modulating mode. The motor can rotate in both directions, with spring return when the electric
circuit breaks.
3. Modulating mode with supplementary power supply. The motor rotates in two directions and
without a spring-return arrangement. When the power is cut off, a bypass signal is usually sent
5.26 CHAPTER FIVE
to the electric motor to drive the valve to its open or closed position, depending on whether it is
a normally open or closed valve.
It may take minutes to fully open a large valve using an electric motor valve actuator. Modern
electronic actuators use solid-state control boards to determine the speed, the action, and other
functions to meet more demanding requirements.
Pneumatic Actuators. A pneumatic valve actuator consists of an actuator chamber whose bottom
is made of a flexible diaphragm or bellows connected with the valve stem. When the air pressure in
the actuator chamber increases, the downward force overcomes the spring compression and pushes
the diaphragm downward, closing the valve. As the air pressure in the actuator chamber decreases,
the spring compresses the diaphragm, moving the valve stem and valve upward. A pneumatic valve
actuator is powerful, simple, and fast to respond.
Because of the increasing popularity of the DDC systems, there is an increasing demand for
electric actuators that can be interfaced with a DDC system.
Types of Control Valves
Water control valves consist mainly of a valve body, one or two valve disks or plugs, one or two
valve seats, a valve stem, and a seal packing. Based on their structure, water control valves can be
classified into the following types:
1. Single-seated. A single-seated valve has only a single valve disk and seat, as shown in Fig. 5.15b
and c. It is usually used for water systems that need a tight shutoff.
2. Double-seated. A double-seated valve has two valve disks connected to the same valve
stem and is designed so that the fluid pressure exerted on the valve disks is always balanced.
Consequently, less force is required for the operation of a double-seated valve, as shown in
Fig. 5.15a.
3. Butterfly. A butterfly valve consists of a cylindrical body, a shaft, and a disk that rotates on an
axis, as shown in Fig. 5.15d. When the valve closes, it seats against a ring inside the body. A
butterfly valve exhibits low flow resistance when it is fully opened. It is compact and is usually
used in large water pipes.
According to the pattern of the water flow, water control valves can again be classified as twoway
valves or three-way valves. A two-way valve has one inlet port and one outlet port. Water flows
straight through the two-way valve along a single passage, as shown in Fig. 5.15a.
In a three-way valve, there are three ports: two inlet ports and one common outlet port for a
three-way mixing valve, as shown in Fig. 5.15b, and one common inlet port and two outlet ports for
a three-way diverting valve, as shown in Fig. 5.15c. In a three-way mixing valve, the main water
stream flows through the coil or boiler, and the bypass stream mixes with the main stream in the
common mixing outlet port. In a three-way diverting valve, the supply water stream divides into
two streams in the common inlet port. The main water stream flows through the coil, and the bypass
stream mixes with the main water stream after the coil. A three-way mixing valve is always located
downstream of the coil. But a diverting valve is always located upstream of the coil.
A diverting valve should never be used as a mixing valve. The unbalance pressure difference between
the two inlet ports and the outlet port at a closed position may cause disk bouncing and valve
wear when the valve disk travels between the two extremes.
Valve Characteristics and Ratings
The different types of control valves and the characteristics that are important during the selection
of these valves are as follows:
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.27
1. Equal-percentage valve. This control valve changes the water flow rate by a certain percentage
for that same percentage of lift in the valve stem when the upstream versus downstream water
pressure difference across the valve (its pressure drop) is constant.
2. Linear valve. This control valve shows a directly proportional relationship between the flow rate
and the lifting of the valve stem for a constant pressure drop.
5.28 CHAPTER FIVE
FIGURE 5.15 Various types of control valves: (a) Double-seated two-way valve; (b) single-seated three-way mixing valve; (c) singleseated
three-way diverting valve; (d) butterfly valve.
3. Quick-opening valve. This control valve gives the maximum possible flow rate when the valve
disk or plug is just lifted from its seat.
Rangeability is defined as the ratio of the maximum flow rate to the minimum flow rate under
control. An equal-percentage valve may have a very good rangeability of 50 : 1. A linear valve may
have a rangeability of 30 :1.
The flow characteristics of equal-percentage, linear, and quick-open valves are shown in
Fig. 5.16.
The following control valve ratings should be considered during the selection and sizing of a
valve:
1. Body rating. The nominal body rating of the valve is the theoretical rating of the valve body
only, in psig. The actual body rating is the permissible safe water pressure for the valve body, in
psig (kPa g), at a specific water temperature.
2. Close-off rating. That is the maximum pressure difference between the inlet and outlet ports that
a valve can withstand without leakage when the valve is fully closed, in psi (kPa).
3. Maximum pressure and temperature. These are the maximum pressure and temperature of water
that the whole valve, including body, disk, seat, packing, etc., can withstand.
Valve Selection
Proper selection of water control valves depends on water system performance, load variations, pipe
size, control modes, etc. Today, the use of scaling factors on analog outputs in a DDC system permits
a nonlinear device to provide an output of linear response. However, select a control valve having a
linear relationship between a change in the controlled variable and the amount of travel of the valve
stem, or a linear system control characteristic over the operating range is still desirable when it is costeffective.
Hence, a linear valve is often used for the water system for which the controlled variable has
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.29
FIGURE 5.16 Typical flow characteristics of various types of control
valve.
a linear relationship with the water flow or valve opening; or for applications that do not have wide
load variations.
When a control valve is used to modulate the water flow rate of a hot or chilled water coil, a
large reduction in the flow rate causes only a small reduction in the heating or cooling output of the
coil. Given such circumstances, the nonlinear behavior of an equal-percentage valve combined with
the nonlinear output performance of a hot or chilled water coil will provide more linear system
behavior.
When three-way valves are used, the water flow rate before or after the common port is approximately
constant, no matter how wide the openings of the various ports in the three-way valves. As
such, three-way valves are used in constant or approximately constant water flow rate systems, even
the coil load changes.
As a two-way valve closes, the flow rate of the water system decreases and its pressure drop
across the two-way valve increases. A two-way valve is thus used for water systems that have variable
volume flow during a variation in system load, as shown in Fig. 5.17.
Valve Sizing
The size of a control valve affects the controllability of a water system. If a control valve is oversized,
then the smallest increment possible may overshoot the controlled variable. An undersized
control valve needs great pumping power. The size of a control valve is also closely related to its
5.30 CHAPTER FIVE
FIGURE 5.17 A typical chilled water system.
design water flow rate , gpm (L/ s), and the pressure drop across the valve pvv, psi (kPa), when
the control valve is fully opened. Their relationship can be expressed as
(5.8)
where Cv is the flow coefficient for a flow rate of 1 gpm at a pressure drop of 1 psi. The flow coeffi-
cient of control valves can be found in manufacturers catalogs.
For a water system that has several cooling coils between supply and return mains, as shown in
Fig. 5.17, modulating the water flow is effective only when the opening and closing of the control
valve V1 in piping section EH has a significant effect on the change in the pressure drop pEH
across piping section EH between the supply and return mains. As control valve V1 opens wider
and pEH decreases, if there is no concurrent change in flow resistance of piping sections JM and
NQ, then a greater water flow will pass through piping section EH. If pEH increases, less water
will flow through EH. If pEH, pJM, and pNQ all increase, based on the characteristic curve of the
centrifugal pump, the water flow rate through pumps P2 and P3 will decrease for an increase in the
system head. If pEH, pJM, and pNQ all drop, the water flow rate through pump P2 or P3 will
increase.
Because of the use of DDC systems with PI or PID control modes and the variable-speed pumping
systems, as well as the results of the previous analysis the following hold:
1. The size of a control valve should be determined according to the flow coefficient calculated
from Eq. (5.8) and listed in the manufacturers catalog.
2. The assumed pressure drop pvv across the control valve should be appropriate. It is affected by
the type of control valve used. It is also a compromise between a higher pvv value to provide
desirable controllability and a lower pvv value to save energy.
3. For a water system using variable-speed pumping and DDC systems with PI or PID control
modes, a pressure drop across an equal-percentage control valve pvv  5 to 10 ft WC (15 to
30 kPa) and a rangeability of 30 : 1 or greater is recommended for energy saving.
Example 5.1. A water system using variable-speed pumping with DDC supplies chilled water to
the cooling coils of three air-handling units (AHUs), as shown in Fig. 5.17. If the chilled water
flowing through two-way valve V1 is at 100 gpm (6.31 L/s), select and size control valve V1.
Solution.
1. To achieve a nearly linear relationship between the coil load and the travel of the valve stem, an
equal-percentage valve is selected for valve V1.
2. For a water system using variable-speed pumping with DDC, it is assumed that the pressure
drop across the control valve pvv  5 ft WC, or 5 0.433  2.165 psi (15 kPa). From
Eq. (5.8), the flow coefficient is
From one of the manufacturers catalog, for Cv  68, the size of equal-percentage control valve V1
is 3 in. (76 mm).
3. From the friction chart of water in steel pipes, for a chilled water flow rate of 100 gpm (6.31
L/ s) and a piping head loss of 2.5 ft /100 ft (2.5 m/100 m) of length, the diameter of the branch
piping section EH is also 3 in. (76 mm). The size of the control valve and the diameter of the
branch pipe are the same.
Cv 
V
?pvv

100
?2.165
 68
Cv 
V?
?pvv
V?
 Cv?pvv
V?
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.31
5.7 DAMPERS AND DAMPER ACTUATORS
A damper is a device that controls the airflow in an air system or ventilating system by changing the
angle of the blades and therefore the area of its flow passage. In HVAC&R systems, dampers can be
divided into volume control dampers and fire dampers. Fire dampers are covered in a later section.
In this section, only volume control dampers are discussed.
Types of Volume Control Dampers
Volume control dampers can be classified as single-blade dampers or multiblade dampers according
to their construction. Various types of volume control dampers are shown in Fig. 5.18.
Butterfly Dampers. A butterfly damper is a single-blade damper. A butterfly damper is made from
either a rectangular sheet mounted inside a rectangular duct or a round disk placed in a round duct,
as shown in Fig. 5.18a. It rotates about an axle and is able to modulate the air volume flow rate of
the duct system by varying the size of the opening of the passage for air flow.
Gate Dampers. A gate damper is a single-blade damper. It also may be rectangular or round. It
slides in and out of a slot in order to shut off or open up a flow passage, as shown in Fig. 5.18b.
Gate dampers are mainly used in industrial exhaust systems with high static pressure.
Split Dampers. A split damper is also a single-blade damper. It is a piece of movable sheet metal
that is usually installed at the Y connection of a rectangular duct system, as shown in Fig. 5.18c.
5.32 CHAPTER FIVE
FIGURE 5.18 Various types of volume control dampers: (a) Butterfly damper; (b) gate damper; (c) split
damper; (d) opposed-blade damper; (e) parallel-blade damper.
The movement of the split damper from one end to the other modulates the volume of air flowing
into the two legs or branches. A split damper is usually modulated only during air balancing after
installation or during periodic air balancing.
Opposed-Blade Dampers. An opposed-blade damper is a type of multiblade damper that is often
rectangular, as shown in Fig. 5.18d. It is usually used for a flow passage of large cross-sectional
area. The damper blades may be made of galvanized steel, aluminum alloy, or stainless-steel sheets,
usually not exceeding 10 in. (25.4 cm) in width. Rubber or spring seals can be provided at the fully
closed position to control the air leakage rating, which often does not exceed 6 cfm/ ft2 (30 L/ s m2)
at a pressure drop across the damper of 4 in. WC (1000 Pa). The bearing used for supporting the
blade axle should be made of a corrosion-resistant material such as copper alloy or nylon. Tefloncoated
bearings may also be used to ensure smooth operation of the damper. Lever linkages are
used to open and close the damper blades. The characteristics of the opposed-blade dampers are
covered later in this section.
The maximum static pressure drop across closed opposed-blade dampers is 6 in. WC (1500 Pa)
for a 36-in.- (914-mm-) long damper (the length of the damper blade) and 4 in. WC for a 48-in.-
(1219-mm-) long damper.
Parallel-Blade Dampers. A parallel-blade damper is also a type of multiblade damper used
mainly for large cross-sectional areas, as shown in Fig. 5.18e. The blade material and the requirement
for the seals and bearings are the same as those for opposed-blade dampers.
Damper Actuators (Motors)
Damper actuators, also called damper motors, are used to position dampers according to a signal
from the controller. As with valve actuators, damper motors can be classified as either electric or
pneumatic.
Electric Damper Motors. These either are driven by electric motors in reversible directions or are
unidirectional and spring-returned. A reversible electric motor is used often for more precise control.
It has two sets of motor windings. When one set is energized, the motors shaft turns in a
clockwise direction; and when the other set is energized, the motors shaft turns in a counterclockwise
direction. If neither motor winding is energized, the shaft remains in its current position. Such
an electric motor can provide the simplest floating control mode, as well as other modes if required.
Pneumatic Damper Motors. Their construction is similar to that of pneumatic valve actuators,
but the stroke of a pneumatic damper motor is longer. They also have lever linkages and crank arms
to open and close the dampers.
Volume Flow Control between Various Airflow Paths
For air conditioning control systems, most of the dampers are often installed in parallel connected
airflow paths to control their flow volume, as shown in Fig. 5.19. The types of airflow volume
control are as follows:
Mixed-Air Control. In Fig. 5.19a, there are two parallel airflow paths: the recirculating path um
in which a recirculating air damper is installed and the exhaust and intake path uom, in which exhaust
and outdoor air dampers are installed. The outdoor air and the recirculating air are mixed together
before entering the coil. Both the outdoor damper and the recirculating damper located just
before the mixing box (mixed plenum) are often called mixing dampers. The openings of the outdoor
and recirculating dampers can be arranged in a certain relationship to each other. When the
outdoor damper is at minimum opening for minimum outdoor air ventilation, the recirculating
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.33
5.34 CHAPTER FIVE
FIGURE 5.19 Airflow paths: (a) mixed- air control, (b) bypass control, and (c) branch flow control.
damper is then fully opened. If the outdoor damper is fully opened for free cooling, the recirculating
damper is closed.
Bypass Control. In the flow circuit for bypass control, as shown in Fig. 5.19b, the entering air at
the common junction m1 is divided into two parallel airflow paths: the bypass path, in which a
bypass damper is installed, and the conditioned path, in which the coil face damper is installed in
series with the coil, or the washer damper with the air washer. The bypass and the conditioned
airstreams are then mixed together at the common junction m2. The face and bypass dampers can
also be arranged in a certain relationship to each other.
Branch Flow Control. In a supply main duct that has many branch take offs, as shown in Fig.
5.19c, there are many parallel airflow path combinations: paths b1s1 and b1b2s2, b2s2 and b2b3s3,
etc. In each branch flow path, there is a damper in the VAV box, and points s1, s2, s3, etc., are the
status points of the supply air.
Parallel airflow paths such as those shown in Fig. 5.19 have the following characteristics:
1. The total pressure losses of the two airflow paths that connect the same endpoints are always
equal; for example, , etc.
2. The relationship between total pressure loss p, in. WC (Pa); flow resistance R, in. WC/ (cfm)2
(Pa s2/m6); and volume flow rate , cfm (m3 / s), can be expressed as
(5.9)
Flow resistance is covered in greater detail in Chap. 10.
3. If the total pressure loss p remains constant and the flow resistance Rn of one parallel path
increases, from Eq. (5.9), the airflow through this path V must be reduced. The airflow in other
parallel paths remains the same.
4. The total pressure loss of an airflow path between two common junctions p determines the volume
flow rate of air passing through that path and can be calculated from Eq. (5.9) as
5. When the flow resistances in most of the branches increase because of the closing of the
dampers to a small opening in their VAV boxes, the flow resistance of the supply duct system
Rsys and the system total pressure loss psys both tend to increase, and thus the total air volume
flow of the supply duct system sys will reduce accordingly.
Flow Characteristics of Opposed- and Parallel-Blade Dampers
A parallel-blade or an opposed-blade damper that is installed in a single airflow path to modulate
airflow is often called a volume control damper (or throttling damper). For volume control dampers,
a linear relationship between the percentage of the damper opening and the percentage of full flow
is desirable for better controllability and cost effectiveness. (Full flow is the air volume flow rate
when the damper is fully opened at design conditions.) The actual relationship is given by the installed
characteristic curves of parallel-blade and opposed-blade dampers shown in Fig. 5.20a and
b. For the sake of energy savings, it is also preferable to have a lower pressure drop when air flows
through the damper at the fully open condition.
In Fig. 5.20,  is called the damper characteristic ratio and may be calculated as
(5.10)


ppath pod
pod

ppath
pod
 1 
pp-od
 pod
V?
?
p
R
V?

p  RV?2
V?
pum  puom, pm1 bym2  pm1 conm2
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.35
5.36 CHAPTER FIVE
FIGURE 5.20 Flow characteristic curves for dampers: (a) parallel-blade
and (b) opposed-blade.
where ppath  total pressure loss of airflow path, in. WC (Pa)
pod  total pressure loss of the damper when it is fully opened, in. WC (Pa)
pp-od  total pressure loss of air flow path excluding damper, in. WC (Pa)
Damper Selection
Butterfly dampers are usually used in ducts of small cross-sectional area or in places like VAV boxes.
For volume control dampers in a single airflow path, in order to have better controllability, an
opposed-blade damper is recommended if many dynamic losses other than the damper itself (such
as coil or air washer, heat exchanger, and louvers) exist in the airflow path. If the damper is the primary
source of pressure drop in the airflow path, a parallel-blade damper is often used.
For mixing dampers, a parallel-blade damper is recommended for the recirculating damper as the
pressure drop across the damper is often the primary source in its airflow path. An opposed-blade
damper is recommended for the outdoor damper and exhaust (relief) damper for better controllability.
The parallel blades of the recirculating damper should be arranged so that the recirculating airstream
will blow toward the outdoor airstream, resulting in a more thorough mixing. Many packaged units
also use parallel-blade outdoor dampers for smaller pressure drop and less energy consumption.
For face and bypass dampers, an opposed-blade coil face damper in an airflow path of greater pressure
drop and a parallel-blade bypass damper will give better linear system control characteristics.
For two-position control dampers, a parallel-blade damper is always used because of its lower
price.
Damper Sizing
Damper sizing should be chosen to provide better controllability (such as a linear relationship between
damper opening and airflow), to avoid airflow noise if the damper is located in the ceiling
plenum, and to achieve an optimum pressure drop at design flow to save energy.
The face area of the damper Adam, ft2 (m2), in most cases is smaller than the duct area Ad, in
ft2 (m2). Based on Alley (1988) paper, the local loss coefficient Cdam of the damper for different setups
can be determined from Fig. 5.21. Then the pressure drop across the damper when the damper
is fully opened pod, in. WC (Pa), can be calculated as
(5.11)
(5.12)
where vdam  face velocity of the damper, fpm.
1. The damper is generally sized when the air flowing through the damper is at a maximum. For
an outdoor damper, the maximum airflow usually exists when the free cooling air economizer cycle
is used. For a recirculating damper, its maximum airflow occurs when the outdoor air damper is at
minimum opening position, to provide outdoor air ventilation.
2. The face velocity of dampers vdam is usually 1000 to 3000 fpm (5 to 15 m/ s), except that the
face velocity of a butterfly damper in a VAV box may drop to only 500 fpm (2.5 m/ s) for energy
savings and to avoid airflow noise. The ratio Adam /Ad is often between 0.5 and 0.9.
3. The outdoor damper may be either made in a one-piece damper or split into two dampers, a
larger and a smaller, to match the needs at free cooling and minimum outdoor ventilation.
4. For a bypass damper, its face area should be far smaller than that of an air washer or than a
water heating or cooling coils face damper. When the air washer or coils face damper is closed, the
area of the bypass damper should provide an airflow that does not exceed the system design airflow.
vdam 
V?
dam
Adam
pod  Cdamvdam
40052
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.37
5.8 SYSTEM ARCHITECTURE
Architecture of a Typical EMCS with DDC
Figure 5.22 shows the system architecture of a typical energy management and control system with
direct digital control (EMCS with DDC) for a medium or large building.
Operating Levels. Such an EMCS has mainly two operating levels:
1. Unit level. This level is controlled by unit controllers. A unit controller is a small and specialized
direct digital controller which is used to control a specific piece of HVAC&R equipment or device
such as a VAV box, a fan-coil unit, a water-source heat pump, an air-handling unit, a packaged
unit, a chiller, or a boiler. For HVAC&R, most of the control operations are performed at the unit
level. Since the software is often factory-loaded, only the time schedules, set points, and tuning
constants can be changed by the user. Some of the most recently developed unit controllers are also
5.38 CHAPTER FIVE
FIGURE 5.21 Local loss coefficient Cdam of air damper. (Source: ASHRAE Transactions 1988, Part I.
Reprinted by permission.)
programmable to various degrees. Sometimes the manufacturer provides a variety of preprogrammed
control sequences, such as monitoring, and diagnostics, and designers can specify the
required control sequence that best fits their designs.
2. System/building level. This level is controlled by system controllers. Since a system controller
has an onboard capacity, programmed by an operator or factory-preprogrammed software, to
execute complicated HVAC&R and other programs, they are the brain of an EMCS. Generic control
software such as scheduling, trending, alarming, diagnostics, and security is also provided in system
controllers. Generic control is covered in detail in a later section.
A system controller is used to coordinate the control operations of an HVAC&R system, such as
the coordination between the duct static pressure and the total air volume of VAV boxes in an air
system, or the sequencing of three centrifugal chillers and cooling towers in a refrigeration plant. A
system controller may interface with sensors/transmitters by means of input/ output (I /O) connections
directly. Unit controllers are also configured on a separate subnetwork and connected to a
system controller.
Operator Personal Computer (PC) Workstation. The operator may interface with the EMCS
primary through an operators PC workstation or purpose-built device, either handheld or fixed to
the system controller. Each workstation shall consist of a PC, autodial telephone modems, and
printers. The central processing unit (CPU) shall be a minimum of an Intel 80486 and operated at a
minimum of 33 MHz. Its memory includes a minimum of 8 Mbytes RAM and 212 Mbytes hard
disk. The communication ports connected to system controllers and other control systems should be
provided. A 14-in. (356-mm) color monitor also shall be provided.
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.39
Operator
interface
System
Unit
Gateway Unit controller System controller
Proprietary Network
PC
GW
BAC net
BAC net
BAC net
BAC net
GW
SC
AHU
Chiller
UC UC UC
BAC net
SC
UC UC UC UC
UC SC
Printer Modem
FIGURE 5.22 System architecture of a typical large EMCS.
The software in the workstation shall do the following:
 Accommodate processes as well as prioritize applications based on their input/output priority level.
 Provide system graphics including the HVAC&R equipment (such as a display of up to 10
graphic screens at once for comparison) to monitor the operating status of the system. The operator
with the proper password is able to add, delete, or change the set points, time scheduling, etc.
 Support the editing of all system applications including the generic control software provided in
the system controllers. The edited or custom programmed software shall be downloaded and executed
at one or more of the system controllers.
 Automatically save the database and restore the database that has been lost in one of the system
controllers.
 Provide scheduling, trends, totalization, alarm processing, security to view and edit data, and
system diagnostics.
Communication Network. A peer-to-peer data communication of a local-area network (LAN)
adopting either Ethernet or ARCNET will be used between system controllers and between system
controllers and PC workstations (or other system). A peer-to-peer communication means that all
system controllers or work stations have equitable access to communication resources. For the communication
subnetwork between system controllers and unit controllers, and between unit
controllers themselves, a master-slave token-passing (MS/TP) technology is often used. A system
controller also acts as a medium to provide data communication between the work station and the
unit controllers. The network technology is covered in detail in a later section.
Power Source. The temperature sensors and humidity sensors may need up to 12 V dc and
24 V ac as a power source. Many DDC units have a power source of 24 V ac or 120 V ac line voltage.
Most of the valve actuators and damper motors need a power source of 24 V ac.
Size of EMCS. The size of an EMCS depends on the number of points (or objects) that belong to
its DDC units. An EMCS of 100 points or less can be considered a small project. An EMCS that
has 1000 points or more can be considered a large project.
System Characteristics
An architecture of EMCS incorporating DDC such as that in Fig. 5.22 has the following characteristics:
1. All DDC units are independent and stand-alone controllers. If any of the controller fails, there
is only a limited effect.
2. The system architecture shows a distributed processing model. Since most of the control operations
are performed at the unit controller level, and partly at the system controller level, such an
architecture has the advantage that it tremendously reduces the data communication between the
unit controllers and the system controller, as well as between the DDC system controller and the PC
workstation.
3. Zimmerman (1996) noted that since the introduction of DDC in the 1980s, . . . microprocessor
and memory have declined rapidly in cost while wiring and installation costs have not
declined at the same rate. To provide more powerful unit controllers, moving the controllers nearer
to the sensors and control devices will reduce a lot of wiring and installation costs as well as the
overall system cost.
4. If each HVAC&R piece of equipment has its own unit controller, it surely will be beneficial to
the HVAC&R equipment manufacturers to fabricate controllers and other control system components
themselves.
5.40 CHAPTER FIVE
Future Development
The development of more powerful, programmable unit controllers using a modular configuration,
will reduce the difference between a system controller and a unit controller. As predicted by
Hartman (1993), the future architecture of an EMCS may have only a single tier of various kinds of
unit controllers. The operators PC workstation, modems, and unit controllers will all be connected
to a peer-to-peer data communication trunk. Such an architecture will simplify the DDC units and
the communication network, move the intelligence nearer to the control devices, and finally create
more effective control at lower system cost.
5.9 INTEROPERABILITY AND OPEN PROTOCOL BACnet
Interoperability
Turpin (1999) defined interoperability as the ability of systems including equipment and components
from different manufacturers to share data and information for the purpose of operation with
plug-and-play connectivity. Interoperability is one of the necessary conditions for system integration.
System integration is a strategy to integrate various HVAC&R systems of various manufacturers
together, and to integrate HVAC&R systems with lighting, fire protection, security, elevator, and
electrical systems in a building together.
Robertson and Moult (1999) note that the advantages of system integration include the following:
 It reduces the installation cost of the building automation system.
 It enhances energy management.
 It can apply building automation system features.
 It provides building operator training.
 There is a single user interface.
The greater the interoperability of a single system and the more systems you try to integrate into a
single system the more complicated and costly the process will be. Interoperability and system integration
is one of the goals of the HVAC&R industry. More and more engineers, facility owners, and
manufacturers recognize this need. It takes time to accomplish such a complicated process.
BACnetOpen Data Communication Protocol
BACnet means building automation and control network. It is an open data communication protocol.
Open means that an independent institution governs its development. All contents are known,
fixed, and accessible. A protocol refers to the rules by which two or more devices communicate
data to each other that must be obeyed. BACnet enables that building automation devices from various
manufacturers can talk to one another, share data, and work together following a standard way.
BACnet defines all the elements of data communication between devices in a building automation
control system. It is specifically tailored for HVAC&R control equipment, but it also provides a basis
for integrating lighting, security, and fire-detecting systems. BACnet was developed from 1987
to 1995 by ASHRAE and was adopted as a national standard in 1995 by ANSI, as ANSI/ASHRAE
Standard 135-1995. For details, refer to ASHRAEs BACnet.
BACnet will assist building owners, designers, contractors, and operators in three areas:
1. It gives more freedom to select the best equipment and components from different manufacturers
in order to have a more efficient system at lower cost.
2. It operates the control system from a single workstation; i.e., it is more effective to operate and
easier to maintain, and there is only one system to learn.
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.41
3. It collects data from different systems and offers greater flexibility for extended systems in
retrofit projects.
Most of the EMCS manufacturers agreed to fabricate BACnet-compliant products from the late 1990s.
There are many other data communication protocols. One was developed by Echelon Corporation,
called LonTalk Protocol, and it became a working system in the mid-1990s and was favored
by members of the LonMark Interoperability Association. LonTalk protocol is neither an opento-
public protocol nor a standard.
An independent consortium called the OPC foundation, formed as a nonprofit organization in
1996 in Boca Raton, Florida, has dedicated itself to provide interoperability with Microsoft technologies
to develop a global specification and multivendor interoperability in industries. In 1990,
OPC is leading 140 member companies including Honeywell , Johnson Controls, and Siemens.
Application Layer
Layered Structure. A data communication system often adopts a hierarchical layered structure so
that a complex problem is broken into smaller and more easily solved problems. BACnet is based
on a four-layer collapsed architecture that corresponds to application, network, data link, and physical
layers in an International Organization for Standardization (ISO) model. This is the result of
careful consideration of the characteristics and requirements of the building automation control
(including HVAC&R) together with a constraint that protocol overhead be as low as possible.
An application layer is the highest layer in BACnet. It serves to define the objects and services
(including control operations, information exchange, and control devices) in a building automation
control (BAC) system. It also provides communication services and data encoding schemes
required by applications to perform monitoring and control functions.
Object Types, Properties, and Devices. The BACnet defines a set of standard object types instead
of conventional points. Analog input, analog output, binary value, command, file, program, schedule,
etc., grouped in 18 types are standard object types, for every device in an EMCS must have a
device object. There are 123 properties that have been identified by BACnet which fully describe
the device, or object type, in the network. Certain properties are required to be specified whereas
others are optional. An object identifier specifies its object name, object type, etc., and optional
properties such as description and device type. In BACnet, a device is defined as any device, real or
virtual, that supports digital communication using the BACnet protocol.
Services. In BACnet, services are the operations by which one device acquires information from
another device, commands another device to do something, or announces that some event
happened. BACnet defines 32 services that can be grouped into five categories:
 Alarm and event services refer to changes in conditions detected by a device, such as acknowledgment
of an alarm and confirmed change of value notification.
 File access services are used to read and manipulate files that are kept in devices, such as only
one read or write operation at a time.
 Object access service provides the means to read, to write, and to modify properties, such as to
add one or more items to a property.
 Remote device management offers disparate operations such as to tell a device to stop accepting
messages.
 Virtual terminal services are used by an operator to establish a bidirectional connection with an
application program implemented in a remote device.
Services are classified as confirmed when a reply is usually expected with data and unconfirmed
when no reply is expected. In BACnet, a given device is not required to implement every service.
However, read property is required to be executed by all the devices.
5.42 CHAPTER FIVE
Conformance Class, Functional Groups, and Protocol Implementation Conformance Statement
(PICS). BACnet defines six levels of conformance for all devices, classes 1 to 6, that are hierarchical
to indicate the difference in requirements that must be met to conform to BACnet. The
requirement of a class includes all the requirements of all the other classes having a lower number.
At the lowest level, conformance class 1 requires only that a device be able to execute (respond to)
a read property service request, such as a sensor. For conformance class 6, a device is required to
implement 21 types of service requests, such as a PC workstation.
A functional group defines a combination of services and object types that are required to perform
certain BAC functions. In BACnet, there are altogether 13 functional groups, such as clock,
workstation, and event initiation.
The protocol implementation conformance statement (PICS) is a document provided by the
manufacturer of a device to identify those options implemented by a particular device. The PICS
covers the conformance class of the device, supported functional groups, standard and proprietary
services executed and initiated, etc.
Data Encoding. In BACnet, application-layer protocol data units (APDUs) are used to convey the
information contained in the service primitives and associated parameters. ISO Standard 8824,
Specification of Abstract Syntax Notation One (ASN.1), has been chosen as the method to represent
the data content of BACnet services. Each data element consists of three components: (1) identifier
octets, (2) length octets, and (3) content octets. The fixed portion of each APDU providing protocol
control information is encoded implicitly, and the variable portion of each APDU providing servicespecific
information is encoded explicitly.
Network Layer
In BACnet, the purpose of the network layer is to provide the means from which messages can be
relayed from one BACnet network to another in the internetwork. Two or more BACnet networks
are connected by routers to form a BACnet internetwork. A router, a BACnet device, is used to interconnect
two disparate BACnet local-area networks (LANs). A network layer directs the messages
to a single remote device or broadcasts the messages on a remote network, or to all devices
on all networks. A device is located by a network number and a medium access control (MAC) address.
Another network layer function is message segmentation and reassembly.
In an EMCS, there are often two networks: one uses Ethernet or ARCNET for high-speed
message transmission between system controllers and the PC workstation, and another adopts lowspeed
message transmission between unit controllers and system controllers. Consequently, a network
layer is required for the BACnet protocol.
Data Link/Physical LayerNetwork Technology
In BACnet, a data link layer has the capability to address messages to a single device or to all
devices. At the data link layer, only incoming BACnet messages received from the physical layer
are passed on to the network layer, and a code is added to the outgoing messages to identify them as
BACnet messages before they are passed to the physical layer. A physical layer provides the physical
medium for message transmission.
For local-area networks, BACnet supports Ethernet, ARCNET, MS/TP, PTP, and LonTalk as
alternatives. Various media are used as physical transmission entities. Typical media are twistedpair
wire, fiber-optic cable, and coaxial cable.
Ethernet. It is one of the commonly used LAN technologies. At any time, multiple devices may
access the network simultaneously, i.e., multiple access with collision detection. Ethernet is using
a peer-to-peer communication with a bus network configuration. Ethernet often runs at a speed of
10 Mbits/ s. Two types of coaxial cable are often used in Ethernet: thick-wire and thin-wire. A
thick-wire Ethernet segment has a maximum length of 1600 ft (488 m), and up to 100 nodes can be
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.43
attached. Thick-wire Ethernet is more expensive. A thin-wire Ethernet segment has a maximum
length of 600 ft (183 m) and 50 attached nodes.
ARCNET (Attached Resources Computer Network). This is also a commonly used LAN technology
and is lower in cost than Ethernet. ARCNET adopts a token-passing network access method.
A token which indicates the permission to use the physical medium is passed from one network
node to the next node in a circular manner. As in Ethernet, ARCNET also uses a peer-to-peer communication
network, and its network nodes reside on a bus.
Both coaxial cable and twisted-pair wire are used in ARCNET and run at a speed of
2.5 Mbits/s. For each ARCNET segment, up to 8 nodes can be connected. All together, up to
255 nodes can be communicated over an ARCNET network.
Master-Slave/Token-Passing (MS/TP). In BACnet, MS/TP divides all the nodes on the network
into two categories: masters and slaves. Only masters can initiate data communication, whereas
slaves cannot initiate. Slaves can only respond to requests from masters. The MS/ TP also adopts a
token-passing network access method. A master node may access the network only when the token
(permission to use the medium) is passed to it from the previous master node. The token never
passes to the slave nodes. Master nodes in an MS/ TP network are at a peer-to-peer communication.
The MS/TP network often uses twisted-pair wires. It can operate at a speed of 9600 bits / s, as well
as 19.2, 38.4, and 76.8 kbits/ s.
Point-to-Point (PTP). This is a data link layer protocol which provides serial communication
between two devices. Such a point-to-point communication typically involves a dial-up phone modem
or hardwired connection between two nodes. PTP has a simpler medium access mechanism
and is often temporary in nature. PTP is much slower than a LAN. Both devices can receive and
transmit simultaneously.
LonTalk LAN. This is an option of the physical medium in BACnet and is at the base of the Lon
Mark protocol. LonTalk supports a number of choices of physical media.
Connection between BACnet and Proprietary Network
For an extension project, it is possible that the original building is still intended to keep the proprietary
network and the extended part is to construct a BACnet network. The proprietary network
needs a gateway to connect to the BACnet network, as shown in Fig. 5.22. A proprietary network
does not open to others. Typically, information can only be exchanged between EMCS components
and the proprietary network of the same manufacturer. For two networks and their computers using
different protocols to communicate, some translation must take place. The device that provides this
translation is called a gateway.
LonTalk Protocol
According to Glinke (1997), the Local Operators Network (LonTalk) protocol is a seven-layer proprietary
protocol developed by Echelon Corp. Data communication is implemented on a neuron
chip. There are actually three microprocessors within the chip; two are used for the network and
one is for specific functions. These devices, which allow different types of existing networks to
communicate, use multiple, low-cost media (Ethernet, ARCNET, fiber optics) and provide flexibility.
Since this technology is essentially peer-to-peer communication without the need for a central
supervisory control node, Lon Mark-certified unit controllers can exist on the same network as
equals.
The two promising and widely used protocols, BACnet and LonTalk, were able to communicate
with each other in the late 1990s. They will improve themselves through actual operation in the
future and produce better choices for engineers and users.
5.44 CHAPTER FIVE
5.10 CONTROL LOGIC AND ARTIFICIAL INTELLIGENCE
In an EMCS, the software in the DDC units determines the control functionality. Since the logic is
separate from the hardware, the control functions and sequences are now limited only by the knowledge
and innovation of the designer and operators. The development of artificial intelligence in the
DDC softwarefuzzy logic, expert systems, artificial neural network, etc.tremendously expands
the control flexibility and improves the operating quality to meet the more complex requirements
for an HVAC&R energy management and control system.
Fuzzy Logic
Basics. Since many systems have become more and more complex over the past decades, an accurate
analytical model based on rigorous and nonlinear mathematics is difficult to develop and to be
accepted in the daily management by operators. In 1965 Professor Lotfi Zadeh of the University of
California, Berkeley, developed fuzzy set theory which provides a new control logic for a system.
Fuzzy logic has the ability to deal with the imprecision that happens in everyday life. Conventional
digital technology is based on bivalence: yes or no, on or off, 1 or 0, and black or white.
Fuzzy logic is multivalent. Things can be partly yes to some degree, partly no to some degree, and
fuzzy logic deals with shades of gray. Conventional set theory observes that a fact must be either
true or false, whereas in fuzzy logic set theory, a fact can be partly true and partly false, can belong
to a set and also belong to another set.
According to Lehr (1996) and Scholten (1995), fuzzy logic offers simplicity in the midst of
complexity and is a real alternative for system operation and control. The obvious benefits of
fuzzy logic are that it makes things more human and more friendly to a person who is less trained,
and it is more easily maintainable. Fuzzy logic really is not a vague theory. It is comprised of a set
of precise rules based on rigorous mathematics which governs the behavior of a system by means
of words and phrases instead of nonlinear models. Fuzzy logic controllers (FLCs) have been
widely used in air conditioners, humidifiers, refrigerators, and many other devices such as elevators.
There are also many applications for which conventional control logic is better than fuzzy
logic controller.
Fuzzy Sets and Membership Function. Fuzzy sets, membership function, and production rules
are three primary elements of fuzzy logic. Figure 5.23 shows the fuzzy sets, membership functions,
and a diagram of fuzzy logic control. In conventional bivalent crisp set theory, sets of thermal
comfort of an occupant in an air conditioned space can be categorized according to the indoor temperature
Tr exactly as cool, a range of 62 to 72F (16.7 to 22.2C); and just right, a range of 72 to
76F (22.2 to 24.6C), etc., as shown in Fig. 5.23. At Tr  74F (23.3C) and Tr  75F (23.9C),
both are just righthave a membership value of 1 with regard to the just-right set and a membership
value of 0 with regard to the cool set.
On the other hand, the fuzzy set just right ranges between a membership value of 0 at 70F
(21.1C), a membership value of 1 at 74F (23.3C), and a membership value of 0 again at 78F
(25.6C). For Tr  75F (23.9C), the membership value of fuzzy sets would be described as
75 percent just right and 25 percent warm. In another way, the assertion of thermal comfort of just
right is 75 percent true and 25 percent false.
If Tr  76F (24.4C), then the conventional crisp set will be difficult to determine. In addition,
for Tr  74F (23.3C) and Tr  75F (23.9C), the assertion of the conventional crisp set for both
is just right, so there is no difference between them. And the assertion of fuzzy sets for Tr  74F
(23.3C) is 100 percent just right, which is different from 75 percent just right and 25 percent warm
when Tr  75F (23.9C).
Production Rules. When the fuzzy set of an occupants thermal comfort is integrated with another
fuzzy set of percentage of fan speed variation (positive large, positive small, zero, negative small,
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.45
and negative large), then the production rules, or fuzzy logic rules, can be described as follows:
 If the indoor temperature is hot, then the fan speed rises a lot.
 If the indoor temperature is warm, then the fan speed rises a little.
 If the indoor temperature is just right, then the fan speed stays unchanged.
 If the indoor temperature is cool, then the fan speed reduces a little.
 If the indoor temperature is cold , then fan speed reduces a lot.
5.46 CHAPTER FIVE
PL
NL
100
80
60
40
32
20
0
0 0.5
Fan speed variation range, %
If hot,
then
decrease
a lot If warm,
then
decrease
a little
If just right,
then do
nothing
If cool,
then increase
a little If cold,
then
increase
a lot
0
62 66 7071
Fuzzy sets
Positive large PL
Positive small PS
Zero ZE
Negative small NS
Negative large NL
Indoor temperature, F
Indoor temperature, F Conventional crisp sets
74 78 82 86 90
62 66 70 74 78 82 86 90
Cold
Cold
Cool
Cool
Warm
Warm
Hot
Hot Just
right
0.5
0.2 0.45
1
0
0.5
1
Centroid
20
40
60
80
100
1
PS
ZE
NS
Just
right
FIGURE 5.23 Fuzzy sets, membership function, and fuzzy logic control.
Fuzzy Logic Controller. An FLC consists of three parts: a fuzzifier converts ordinary inputs to
fuzzy variables, a fuzzy reasoning unit produces fuzzy control signals based on input fuzzy variables,
and a defuzzifier converts fuzzy control signals to conventional control outputs.
If the indoor temperature Tr is 71F (21.7C), then the membership function of the fuzzy sets of
thermal comfort in Fig. 5.23 would be described as 0.2 (20 percent) just right and 0.45 (45 percent)
cool. From the production rules, then, reduce the fan speed a little has a membership value of
0.75 and fan speed remains unchanged has a membership value of 0.2. One way to interpret these
two fuzzy outputs in a conventional crisp output is to determine the centroid based on the area of
the two truncated triangles. From Fig. 5.23, the output value is 32 (reduce the fan speed by 32 percent
of its range).
Huang and Nelson (1994) noted that an FLC is characterized by a set of linguistic fuzzy logic
rules. The initial set of these rules is often based on past experience or analysis of the control
process. The initial set of rules can be modified by analyzing the performance trajectory on the linguistic
plane to obtain an optimal rule set. This is called rule refinement. The second significant in-
fluence on the behavior of an FLC is the choice of membership functions. The overlap of the fuzzy
sets should be moderate to allow for reasoning with uncertainty and the need for completeness of
the control rules. Refer to Huang and Nelsons paper for details.
Knowledge-Based Systems and Expert Systems
Basics. A knowledge-based system (KBS) is a knowledge-rich, logic-oriented computer program
that mimics human knowledge and reasoning in a specific domain to assist in solving more complex
problems. Human knowledge includes expert knowledge, common facts, and knowledge in
any form, whereas an expert system, strictly speaking, mimics only an experts expertise in a given
domain to solve specific problems. An expert system can be considered as the core part of a knowledge-
based system.
In comparing a knowledge-based system with conventional programmed software, Hall and
Deringer (1989) noted that the benefits of a knowledge-based system are due to the following superior
abilities:
 Logical reasoning. Logical problems are solved by means of sorting, comparing, searching, and
reasoning as well as to evaluate alternative choices to a specific problem. More often, several
causes (or solutions) of a problem are identified.
 Resolution of uncertainty. When precise knowledge is not available and for problems that involve
qualitative facts and incomplete data, the certainty of the conclusions of the KBS using heuristic
rules depends on the reliability of the heuristic rules themselves.
 Multiple approaches. Multiple experts are often used to develop a knowledge base. Experts will
rarely agree exactly on the proper approach to solve a problem. Solution of the problem with multiple
approaches is one aspect of KBS.
 Justification of results. A well-developed KBS has the capability to recall the basis of each decision
during the problem-solving process. It can provide a decision trail that will explain the
sequence of decisions used to produce the conclusion.
The limitations of a KBS are as follows. First, a KBS can make no decision that is not explicitly
contained in the knowledge base. Second, two different KBSs developed using the expertise
from different experts may give different advice for the same problem. Third, the KBS is timeconsuming
and expensive.
Knowledge-based systems will not replace experts and engineering professionals. Instead, the
systems will play the role of specialized knowledge-based advisers or consultants.
KBS Structure. An expert system comprises the following four modules, as shown in Fig. 5.24.
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.47
1. Knowledge base. This is the most important part of an expert system. It contains the common
facts and inference rules, often in the form of if-then, and is domain-specific. Every rule in the
knowledge base is extracted from the experts knowledge and experience by the developer of the
KBS, the knowledge engineer, or is based on data and information from published handbooks,
manufacturers engineer manuals, and field survey results. Frames are another common knowledge
representation method. Hall and Deringer (1989) defined frames that . . . facilitate the description
of knowledge about complex objects, which have many subelements, . . . The content of the
knowledge base should be tailored to the users task and requirements. The users task can be determined
through direct interview with potential end users.
2. Inference engine. The inference engine asks for inputs from the user interface, executes
reasoning algorithms by applying the knowledge from the knowledge base, and arrives at the conclusion
based on the rules. There are two kinds of reasoning: (1) data-driven forward chaining or
forward reasoning and (2) goal-driven backward chaining or backward reasoning. They are categorized
according to how new information is inferred. Forward reasoning proceeds forward from the
user inputs about probable outcomes. In most engineering design problems, forward reasoning is
adopted. For instance, knowledge of the building envelope leads to load calculations, which forwards
next to determining the equipment capacity, and so on.
Backward reasoning proceeds backward from the user inputs about probable outcomes. In troubleshooting
or diagnostics, backward reasoning is used. For an example, collected facts lead to
determining the cause of the problem. Forward and backward reasoning were often used in conjunction
to control the flow of questioning as well as provide a dialog between the system and the
end user.
3. User interface. This includes the facilities to support smooth and convenient interaction with
the user. Only simple instruction is needed to start the system, end a session, save the contents, and
5.48 CHAPTER FIVE
FIGURE 5.24 Structure of a knowledge-based system for an EMCS and HVAC&R design.
give a printout. A friendly dialog between the KBS and the user and a simple question-and-answer
format with a provided menu of possible answers are expected. An intelligent operating system
which accesses various application tools is recommended. Modular new application tools can be
plugged into the operating system as required. A user interface with built-in intelligence, graphically
based high-resolution display with standardized menus and format may be the preferable answer.
4. Knowledge acquisition. A knowledge acquisition module provides strategies to capture the
experts knowledge to develop a KBS. Sometimes, it also checks for consistency and completeness.
Knowledge acquisitions are usually accomplished through personal interviews with experts and review
of the application literature. Most HVAC&R-related knowledge systems have long-term plans
to continue the process of knowledge acquisition. A new trend is machine learning; i.e., the computer
learns from the experience how to capture and manipulate new knowledge.
Development of KBSs. Many KBSs are developed by using commercially available development
tools called shells. A shell consists of mainly a rule editor, a knowledge base, an inference engine,
and user interfaces. Usually, the inference engine and user interface are fully developed within a
shell. Thus, the user can focus on the collection and the input of the knowledge. He or she can
change the existing knowledge base and does not need to change the entire system. KBS is expected
to operate on a PC.
The performance of a KBS and the resources needed during development are highly affected by
its knowledge engineer. The knowledge base is built through the cooperation of the knowledge engineer
and experts in a specific domain. The knowledge engineer has the responsibility to choose an
appropriate inference strategy, and a suitable shell and to ensure compliance of the system with the
task.
The development cycle of a KBS is an iterative and incremental process. It begins with the initial
prototype. The next is an improved, and expanded, one. The process is repeated and may take years.
How the KBS Works. For example, if the space air is too humid and the space cooling load is
only one-half of the design load, it is required to find the general cause of these symptoms. The
problem is a diagnosis problem. Brothers and Cooney (1989) stated that if-then rules are composed
of parameters (symptom, set point) and values (too hot, too humid, correct). In the inference
engine, the computer program of the KBS will begin to search for an if-then rule through the
knowledge base that will give a value (too humid) for a general cause. If the values do not match
those required by the rule, the computer program will search for the next rule, until the following
if-then rule has been found:
IF symptom is too humid, and
cold supply air temperature set point is correct, and
cold supply air relative humidity is O.K., and
sensible cooling load is 50 percent of design load
THEN general cause is size of supply airflow rate CF 90
The cause of a too humid space air is that the volume flow of supply air is too great. The term CF is
the abbreviation of certainty factor. That means the confidence in the answer to the problem is 90
percent.
Testing, Verification, and Validation. In Jafar et al. (1991), testing often detects logical errors related
to the knowledge base, syntactic errors, and missing knowledge. Testing only shows errors
and does not explain their cause. Verification and validation should be performed at each stage of
development, to check the knowledge base for internal inconsistencies, mismatches, etc.
Applications. There are three primary areas of HVAC&R-related applications of KBS, as reported
in the paper by Hall and Deringer:
 Monitoringinterpretation of measured data in comparison to expected behavior
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.49
5.50
Outputs
Output
Neuron
1n
Neuron
W12,1
W12,n
Wi1,n
WLo,n
WLo,1
Wi1,1
Wi1,1
Wi1,2
Wi1,n
(a)
0
	1
i
(b)
Neuron
Hidden layer
Input
Input layer Hidden layer
1 i 2...L
Neuron
f(x)
Output layer
o
FIGURE 5.25 An artificial neural network (ANN): (a) Structure of an ANN; (b) sigmoid function.
 Diagnosticsassistance in identifying solutions to complex technical problems
 Designassistance in the selection of the HVAC&R systems and subsystems
Artificial Neural Networks
Basics. An artificial neural network (ANN) is massive interconnected, parallel processing,
dynamic system of interacting processing elements that are in some aspect similar to the human
brain. The fundamental processing element is called the neuron, which is analogous to the neural
cell in human brain. The neurons are set in layers, and thus a network is formed as shown in
Fig. 5.25. Inputs representing the variables that affect the output of the network are feeding forward
to each of the neurons in the following layers with an activation depending on their weighted sum.
Finally, an output can be calculated as a function of the weighted sum of the inputs and an additional
factor, the biases.
The ability to learn is one of the outstanding characteristics of an ANN. The weights of the
inputs are adjusted to produce a predicted output within specified errors. ANNs have been increasingly
used in recent years to predict or to improve nonlinear system performance in HVAC&R. An
ANN system is characterized by its net topology, neuron activations transfer, and learning method.
Net Topology. The structure of the network of an ANN, or net topology, depends on the data flow
mode, the number of layers, and the number of hidden neurons.
 In Miller and Seem (1991), there are two types of data flow modes: state and feed-forward models.
In the state models, all the neurons are connected to all the other neurons. In feed-forward
models, the neurons are connected between layers, as shown in Fig. 5.25a, and the information
flows from one layer to the next. Feed-forward models are the most popular and most often analyzed
models.
 In an ANN, there is always an input layer with the number of inputs equal to the number of parameters
(variables) that affect the output.
 There may be one or more hidden layers of neurons next to the input layer. The selection of number
of hidden layers and the number of neurons in each hidden layer remains an art. Curtiss et al.
(1996) noted that too many hidden layers and hidden neurons tend to memorize data rather than
learning. The hidden layers and hidden neurons must be sufficient to meet the requirement during
the learning process for more complex nonlinear systems. More hidden layers and hidden units
need more calculations and become a burden.
Among the 10 papers published from 1993 to 1996 in ASHRAE Transactions regarding developed
ANNs, most have only one hidden layer, some have two layers, and only one has three hidden
layers. None exceeds three hidden layers. Kawashima (1994) recommended that in an ANN only
one hidden layer is sufficient for load prediction.
 If the relationship between the inputs and output is more complex, i.e., nonlinear, and more inputs
are involved, then more neurons are needed in each hidden layer. Kawashima (1994) also suggested
that the number of neurons in each hidden layer exceed 2m 	 1. Here m indicates the
number of inputs.
 There is always an output layer next to the hidden layer(s). It is preferable to have one neuron
(single output) in the output layer for simplicity. There may be two or more neurons for multiple
outputs.
Neuron Activation Transfer. In Miller and Seem (1991) and Curtiss et al. (1996), for each neuron
in the hidden and output layers:
1. The input activations to a neuron in the first hidden layer h, denoted by i1n, can be
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.51
calculated as
(5.13)
where in  normalized input
Wi1,n  weights of connection between inputs and neurons in first hidden layer
B  biases
(5.14)
where io  original input data
imax, imin  maximum and minimum of original input data
2. The output from the neurons in the first hidden layer o1n, is expressed as a selected sigmoid
activation function:
(5.15)
There is only one output o for a neuron. This output is transmitted through output connections in
which it usually splits into multiple connections with identical activations.
3. The input activation to a neuron in the hidden layer L, denoted iLn, or a neuron in the output
layer o, denoted by ion, is equal to the output of the neuron in the previous layer which split identically
to all the neurons in hidden layer L. The output of a neuron in hidden layer L, denoted by oLn,
or output layer oon can be calculated as
(5.16)
and
(5.17)
Learning Method. The learning method, also called the training process, determines the connection
weights Wxyn through known sets of input/output pairs. The initially assigned connection
weights are adjusted repeatedly during the learning process until the error is within the specified
values. After training, the ANN can predict the outputs from given inputs. Backpropagation is the
most often used systematic method to train multilayer ANNs. Curtiss et al. (1996) and Miller and
Seem (1991) recommended the following training procedure:
1. Assign initial random values for connection weights, often between 0.5 and 	0.5. Select a
training input output pair; calculate the normalized input activations in and the input activation to
the neurons in the first hidden layer according to Eqs. (5.13) and (5.14).
2. Calculate the outputs of the 1, 2, . . ., L hidden layers o1n, o2n, . . ., oLn and then of the output
layer oon from Eqs. (5.15) and (5.16).
V  	iLnWLo,n
T  	i1nW1L,n
oon 
1
1 	 e(VB)
oLn 
1
1 	 e(TB)
o1n 
1
1 	 e(SB)
in 
io imin
imax imin
i1n  	inWi1, n 	 B  S 	 B
5.52 CHAPTER FIVE
3. Evaluate the error  between the calculated oon and the selected training output (target output)
ot by:
(5.18)
4. Adjust the connection weights from Wj to Wj	1 to minimize the error by using the following
rule:
(5.19)
where  is the learning rate whose value lies between 0 and 1.
5. Repeat the previous steps for all the input / output pairs in the training set until the error for the
entire training set is lower than the preset training tolerance. The training set should cover the
operating range of the inputs and outputs.
Applications. An ANN can model multiple parameters simultaneously for nonlinear systems. It
can also be periodically trained to update the weights. ANNs are now widely used for predictive
control, such as energy use prediction, energy optimization, adaptive control, data trending, and
optimum start and stop.
5.11 PROGRAMMING OF DDC SYSTEMS
Evolution of DDC Programming
The EMCSs having more complex functional controls are heavily software-driven. From the 1960s
to the 1970s, the software for HVAC&R control was mostly performed in a workstation computer
and programmed in the control manufacturers factory. Because of the trend to use distributed
microprocessor-based DDC units in the late 1970s, many EMCS manufacturers provide some type
of operator control language (OCL) for field-programmable line programming in the controllers using
BASIC-type language. This meets the requirements of computer program development, periodic
updating, and necessary modifications and improvements.
Since the late 1980s, there is a trend toward the function and object-orientated graphical programming
to provide software for DDC units and the PC in the operators workstation. Today, both
traditional line programming using BASIC-like language (sometimes Pascal and C++ languages are
also used) and graphical programming are used in DDC systems.
Graphical Programming
Davison (1992) described that graphical programming is a schematic drawing of a desirable control
scheme (functional control), such as mechanical cooling or economizer control, using symbols
called templates. The templates are displayed on the computer screen interconnected by lines that
direct the flow of data. The control scheme shows the inputs, through control operations, to outputs.
After the diagram of the graphical programming is completed, it is converted to a program usable
by the DDC units and PC through a computer program.
Compared with traditional line programming using a BASIC-like language, graphical programming
has the following advantages: First, it is intuitive and easier to use and understand by field
HVAC&R engineers and operators. Second, the user is more familiar with it because of meaningful
visual symbols represent functions and devices. It is simpler in documentation and easier in troubleshooting.
Third, it provides a time-saving tool for the specification and documentation of the
HVAC&R energy management and control systems. Finally, graphical programming enables one to
improve programming quality at a reduced cost. The limitations of graphical programming are that
Wj	1  Wj 	
oon

 oon(1 oon)(ot oon)
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.53
it is bulky to display on screen and that it is new to us and needs to develop required supports such
as a portable interface that acts directly to the DDC units.
Templates
In Davisons (1992) paper, a template is a graphical symbol (icon) in graphical programming that
describes a single or combined control scheme (specific functional control), as shown in Fig. 5.26.
A template consists of input and output connections, a small section of computer code, and private
data storage registers. A template is function- and object-oriented. Each template performs a small
portion of a specific functional control. If several templates are connected by lines, they form a
complete computer program for a specific functional control. The template symbol provides a
visual reminder or memory of the function that the template accomplished.
A template maintains its own set of private data and variables that cannot be altered by the action
of any other template. The state of an instance of a template is contained in its private variables. They
can be only modified by the program code contained in that instance. The private variables of an instance
of a template are hidden from the action from other templates.
In a template, inputs are used to receive data from other templates or constants. Outputs are used
to send data to other templates. Inputs and outputs between two or more templates are connected by
lines that direct the data flow between the templates. An input may receive a value from only one
output in order to prevent overwriting of data at the input. On the other hand, an output data may be
sent and used by any number of inputs.
For a more complex control function that needs several templates, a combination of templates
called a macrotemplate is often used to simplify a graphical programming diagram.
5.54 CHAPTER FIVE
10K
Templates Dead band DB Set point SP
Cooling control
Outdoor temperature
lockout
Average
temperature
2
I
SP
Cool
Cool SP
DB 	
Zone
temperature
west
Zone schedule
10K
Zone temperature
east
10K
Outdoor
temperature

85
74
180
T
T
T
on off Min Min
Economizer
Compressor
0
1
0.5
Time
clock
on off Min Min
I
SP
Cool
DB
3
SO
FIGURE 5.26 Graphical programming for mechanical cooling control in a small packaged unit.
Graphical Programming for Mechanical Cooling Control
Figure 5.26 shows the graphical programming for mechanical cooling control for a small packaged
unit. This packaged unit has a supply fan, a DX coil, a single-scroll compressor, and an economizer.
The graphical programming for the control of this packaged unit includes mechanical cooling,
economizer, heating, and fan control.
During the scheduled occupied time, the zone schedule produces a 1.0 signal, and the cooling
control template SP selects the input labeled 1.0 and passes the value 74F (23.3C) as the cooling
set point. During unoccupied periods, SP passes the value 85F (29.4C). The average temperature
of east and west zones is compared with the current set point of the cooling dead-band control template.
If the average temperature exceeds the set point plus the dead-band value, 0.5F (0.28C), or
74.5F (23.6C), then the output of the cooling control template goes to 1.0, and the compressor is
turned on. If the zone average temperature drops below the set point minus the dead-band value,
73.5F (23.1C), then the output of the cooling control template goes to 0.0 and the compressor is
turned off.
When the outdoor air temperature drops below 50F (10.0C) minus the dead-band value
3F (1.7C), or 47F (8.3C), then the outdoor temperature lockout template produces 1.0, and the
compressor is locked out (turned off). Minimum on and minimum off timer templates prevent the
compressor from operating at short cycles.
Graphical programming of DDC systems may become supplemental to the sequence of operation
and control diagrams, and may become a part of the DDC design documentation.
5.12 TUNING DDC UNITS
Most of the DDC units use PI or PID control modes. Tuning determines the gains and control
parameters which have a direct impact on the steady-state error and transient characteristics of the
DDC unit and the control system. A well-tuned DDC unit minimizes the steady-state error, or offset
from the set point; shows a quick response to disturbance; and provides operating stability at all
operating conditions. HVAC&R processes are nonlinear, and system characteristics change when
seasons are varied. The DDC unit tuned at one condition may not be appropriate at other operating
conditions.
Tuning PI Controllers
For DDC units using PI control modes, proportional gain and integral gain should be properly
selected. A high gain decreases the control stability and the offset or error, whereas a low gain will
produce a slow response, which increases the control stability and offset. In tuning PI controllers,
trial and error is often used by the EMCS subcontractor. The trial-and-error method adjusts the gain
until the desirable response to a set-point change is shown. This response should start with a small
overshoot and rapidly damp to steady-state conditions. The trial-and-error method is time-consuming.
Other tuning methods for PI controllers, such as closed- and open-loop process identification,
are also used. Refer to ASHRAE Handbook 1995, HVAC Applications, for details. Bekker et al.
(1991) recommended a root locus tuning method for first-order processes like most temperature and
humidity control for PI controllers to achieve a critically damped response.
Self-Tuning PI and PID Controllers
During the 1980s, different schemes were developed including Astrom and Hagglunds (1984)
automatic tuning of PI and PID controllers. The tuning procedure suggested by Astrom and
Hagglund is based on the identification of one point on the Nyquist curve of an open-loop system
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.55
with relay feedback. In the earlier 1990s, commercial products of automatic tuning DDC units
started to appear in the EMCS in buildings.
Based on Astrom and Hagglunds principle, Wallenborg (1991) accomplished automatic tuning
of PID controllers in supply air temperature and duct static pressure control experiments. After automatic
tuning, its system performance was improved. The autotuner used in these experiments was
different from a true adaptive controller. An autotuner is operated with fixed parameters during
normal operation. The tuning experiment is initiated between the normal operation periods by the
operator. The required parameters are calculated from the results of the tuning experiments.
Adaptive Control
The controller has the ability to adapt to the control system by determining the optimum PID parameters
and adjusting itself accordingly. An adaptive controller, or self-tuning controller, continuously
updates its parameter during operation based on some on-line process identifier and computer
programs. Self-tuning adaptive controllers are also available as commercial products now.
5.13 FACTORS AFFECTING CONTROL PROCESSES
As defined in Sec. 5.1, the function of an air conditioning (HVAC&R) control system is to modulate
the air conditioning system capacity to match the off-design condition, load variation, and
climate change, to maintain the indoor environment within desirable limits at optimum energy use.
Load
The term load refers to the magnitude of the space load, coil load, refrigeration load, or boiler load
that determines the amount of the supply air, chilled water, or hot water needed to control and
maintain the controlled variable at the desirable value(s). Load variation and disturbances affect the
controlled variable in three circumstances: part-load operation, intermittent operation, and disturbances.
Part-Load Operation. The sizes of an air conditioning system and its components are always selected
according to the magnitude of the load at the design condition, which is often called the design
load or full load. In actual operation, air conditioning or HVAC&R systems operate at part load
most of the time. Many air systems even spend 85 to 90 percent of their annual operating hours at
part-load operation. The load drops below the design load because of
1. Changes in the outdoor climate
2. Changes in the internal loads at the time of operation
During part-load operation, the capacity of an air conditioning system must be appropriately reduced
so that a desirable indoor environment can be maintained; at the same time, the energy use of
the HVAC&R equipment can be saved.
Climate Change
Outdoor climate change affects not only the space load, but also the performance of HVAC&R
systems. Outdoor climate parameters for HVAC&R include the dry-bulb temperature, wet-bulb
temperature, solar radiation, wind speed, and wind direction. They are usually specified at design
conditions for an HVAC&R system or at a rated condition for certain equipment. Outdoor climate
5.56 CHAPTER FIVE
changes during off-design conditions. The general trend in summer when the outdoor dry- and wetbulb
temperatures fall is that the space cooling load drops accordingly. Because each system has its
own characteristics, sometimes it is simpler and more convenient to consider the load ratio and climate
parameter as two different factors during the calculation and analysis of system performance.
For instance, the analysis of the energy use of a water-cooled chiller requires both the load ratio and
the entering condenser water temperature (ECWT) to determine the power input to the compressor.
Intermittent Operation. Many air systems do not operate continuously. Some operate only few
hours within a diurnal cycle. Others operate only in the daytime and shut down at night when the
building is unoccupied. During the warm-up and cool-down periods, space loads vary a great deal
from the design load as well as from that of a continuous part-load operation. The space loads represent
transient loads of a dynamic model. Optimum starting and stopping of intermittently operated
air systems is important for better indoor environmental control and energy savings.
Disturbance. These can be sudden load changes or set-point changes within a short time, say, a
fraction of an hour, that affect the controlled variable. The offset of the controlled variable resulting
from a disturbance can be eliminated by a DDC system with a proportional-integral control mode.
For disturbances resulting from a sudden climate change that affects a 100 percent outdoor airhandling
unit, or from a sudden switch-on of the spotlights in a conditioned space, a DDC system
that incorporates proportional-integral-derivative control mode may be more suitable.
System Capacity
First, the capacity of an air conditioning (HVAC&R) system and its components should be adequate.
An oversized electric heater produces a greater overshoot of space temperature and wastes
more energy than an appropriately sized heater. An undersized chiller always causes a higher space
temperature during the cooling season and the EMCS can become out of control.
Second, from the point of view of capacity control and energy conservation, a heating and
cooling plant installed with multiple units is always better than a single-unit plant. A piece of equipment
will operate more efficiently at a higher turndown ratio than at a lower turndown ratio. The
turndown ratio RTD, often called part-load capability, is usually expressed as the percentage of design
capacity:
(5.20)
Third, modulation of capacity continuously is always better for matching the load to outdoor climate
change than two-position or step controls. Finally, capacity modulation of large fans, pumps,
and compressors using adjustable-frequency variable-speed drives is now often energy-efficient and
cost-effective in many circumstances.
Performance of Control Processes
The performance of an EMCS is often represented by the performance of its control processes, or
more directly its control loops that perform the control functions. HVAC&R control processes are
mostly nonlinear processes whose characteristics cannot be expressed by first-order equations.
Nordeen (1995) listed five criteria to use to evaluate the performance of an HVAC&R control
process:
 Stability. A control system is stable if the controlled variable does not show a continuing trend
away from the set point following a disturbance, as shown in Fig. 5.9a. As described previously, a
wider throttling range and a slower response both are beneficial to system stability. The decay of
RTR 
minimum capacity
design capacity
 100
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.57
the oscillation of the controlled variable is called damping. The more damped a control system is,
the less it oscillates.
 Response time. The length of time needed for the controlled variable to reach the set point after a
step change in set point or load is the response time.
 Overshoot. The difference between the maximum controlled variable and the set point following a
change in set point or load is the overshoot.
 Settling time. The length of time needed for the controlled variable to reach steady-state following
a disturbance.
 Offset. As defined previously, this is the constant error that exists between the controlled variable
and the set point when a steady state is reached.
The problems associated with poor performance include the time lag or dead time, which is large
compared to the reaction rate or time constant of a control system; poor measurements of the controlled
variables; inappropriate capacity control; and hysteresis within control and control components
which create time lag. Time lag or dead time is the delay in time between the change in the
controlled variable and when that change is sensed by the sensor, or when the controlled device is
modulated, or when the capacity of the process plant varies.
During the sizing of control valves and air dampers, a linear relationship between the coil load
and the valve stem travel was the primary design consideration during the past decades. However,
many microprocessor-based DDC systems now permit scaling factors to be applied to the analog
outputs of the DDC units, and thus an inherently nonlinear system will respond in a linear relationship.
Thermal Capacitance
Thermal capacitance, which is sometimes called thermal inertia, is related to the mass of the building
envelope, equipment, and system components. Capacitance is usually calculated by the product
of specific heat of the material and the mass of the material. The thermal capacitance of the
HVAC&R process and the building envelope affects the performance of a control system as follows.
First, the high thermal capacitance of the building envelope, the equipment, chilled and hot water,
etc., reduce the effect of a disturbance to vary the controlled variable. For example, consider a
sudden increase in the lighting load in a conditioned space. As the space air temperature increases
because of heat released from the electric lights, a large portion of heat at the same time will be
transferred to the building envelope because of the additional temperature difference between the
space air and the building structure. The increase in the space temperature resulting from the disturbance
is thus significantly reduced.
Second, during the warm-up and cool-down periods, because of the high thermal capacitance of
the building envelope, heat exchanger, ducts, and pipes, a greater system capacity and longer period
of time are needed to raise or cool down the space air temperature.
Third, the thermal influence of the weather and of external load changes on the conditions of the
indoor space air through the building envelope is rather slow, often needing several hours to reach
full effect.
5.14 FUNCTIONAL CONTROLS
The control functions of a microprocessor-based EMCS in a large high-rise building become more
and more complex and demanding. To provide the required functional controls with satisfactory
system performance is always the primary target of an EMCS. In an EMCS, there are four kinds of
functional controls: generic, specific, safety, and diagnostics.
5.58 CHAPTER FIVE
Generic Controls
Generic controls are usually needed by most of the systems, units (equipment), and components.
The software of generic controls is usually provided in the PC workstation and in the system controllers
except system graphical displays. The editing of the generic software usually takes place in
the PC workstation and can be downloaded to any one of the system controllers. In many EMCSs,
data and information can be monitored and collected from any point in the system and analyzed in
the working station PC and in the system controllers. Graphics displays, trending, totalization,
scheduling, alarming, etc., are examples of the generic controls. The following are features provided
by various manufacturers (or vendors).
Graphical Displays. The graphical displays provided in the PC workstation become the
showpiece of an EMCS. Some EMCS manufacturers demand that the software in the operator PC
workstation be graphically oriented. The graphical displays include the following:
 Building floor plan graphics show the selected floor plan and the space temperatures.
 Equipment graphics are provided for each major piece of equipment such as packaged unit, airhandling
unit, and chiller with status of all points.
 Schematic graphics show the detailed system drawings which can be created, modified, and saved.
 High-resolution digitized photo-quality displays are also accommodated in some EMCSs.
The graphical display system can allow a display of up to 10 graphical screens for comparison and
monitoring of system status. Graphics combined with color coding, such as temperature that is assigned
with different colors at various values, will increase viewers effectiveness. A graphical display
system should provide navigation from a facility map down to a specific floor plan. Associated
temperature and related parameters are provided on the floor plan. An additional click of the
mouse on any questionable area will cause the serving HVAC&R system with its operating information
to appear on the screen.
Speed and convenience of editing are two key issues in graphical displays. If a 486 processor is
used, graphics may take 5 to 7 s to paint. Including the navigating through 5 to 10 screens in search
of a problem, too much time will be needed. Because of the use of the Pentium chip, it is possible
to reduce the graphical response to 2 to 3 s per screen paint. The operator must create, modify, or
edit the graphics easily.
Graphical displays play a small role in increasing the effectiveness of any functional control in
an EMCS. However, they enable the user to interact more efficiently with the data and information
from the EMCS and to make proper decisions to operate the EMCS and the associated HVAC&R
system effectively and efficiently.
Trending. Trending is the ability to provide continuous track of certain parameter(s), or operating
status of a piece of equipment. The commonly used time interval for the trend log is once for every
30 min. Trending is mostly used for troubleshooting. For an example, if the output capacity of a heat
exchanger is gradually reduced, scale may form on the heat-exchange surface. Totalization is another
trend log which records the total accumulated operated time for each of the units in a plant employing
multiple units. Totalization is usually helpful for maintenance and troubleshooting purposes.
The trend graph displays the trend data in graphical form. When the user asks for a trend graph,
select the point(s). The Y axis of the graph is often automatically scaled; the X axis to indicate time
is appropriately labeled. The user can also choose the capacity of trend points, time interval between
two trend data other than the standard 30 min, and selection of both changes of value (COV)
that exceed the defined value and the timed trend. The trend graph is a powerful tool. It is very
helpful for tuning of PID control loops and analyzing operating problems.
Scheduling. Prior to the DDC systems, the use of a time clock to schedule the start and stop of
equipment automatically was most widely adopted. A DDC system usually starts and stops the
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.59
equipment according to the predetermined schedule in the system controllers. They can be in the
form of day, week, month, and season schedules, and the users can define the executing periods.
The graphical schedules are often used for clarification. Modifications of schedules are easily made
at the PC workstation and downloaded to the system controllers. Optimum start and stop are actually
run optimizing of scheduling.
Some manufacturers provide temporary schedules that will be used only once and supplement
the existing schedule. Sometimes an override schedule which supersedes the existing schedule is
also used only once.
Alarming. This is the software for tracking and reporting the alarm conditions, i.e., abnormal
conditions. The operator needs to determine the limits or the difference from the set points that
causes the abnormal condition. The operator also should determine the reactions, logging, printing,
displaying messages and graphics, and producing audible announcements that are taken during an
alarm. An alarm that is not acknowledged by the operator within a specific time will move to a
higher level of priority. The system must automatically lock out alarms when an alarmed system is
turned off and appropriate reactions have been taken.
Discriminator Control. Discriminator control is a kind of optimum control which searches for the
required lowest cooling supply air temperature and the highest heating supply temperature. This
control also minimizes the amount of mixing of cold and warm supply air, terminal reheat, and resets
the cold deck supply temperature to reduce the mechanical cooling and the amount of reheat.
Fault Detection and Diagnostics. Fault detection and diagnostics should monitor the operation of
HVAC&R equipment, components, and control devices; analyze the data; and identify performance
problems to be corrected. Fault detection and diagnostics are discussed in detail in the next section.
Specific Controls
Specific controls including capacity controls are controls for a specific function in an HVAC&R
system, a unit (or equipment), or a zone. The following are specific controls that are discussed in
the corresponding sections later:
 Zone temperature control
Fan-coil unit
VAV reheat
Fan-powered VAV box
 Air system control
Economizer control
Discharge air temperature control
Minimum ventilation control
Duct static pressure control
Space humidity control
Warm-up or cool-down control
 Water system control
Differential pressure control
Chilled water temperature reset
Condenser water temperature control
Low-temperature heating system control
5.60 CHAPTER FIVE
 Central plant control
Multiple-chiller optimizing control
Multiple-boiler optimizing control
Condenser fan cycling
Demand limit control
 Safety controls
Commissioning and Maintenance
A properly designed and installed EMCS needs commissioning to test and tune its controllers and
system components according to the design specification. A sufficient, clear, and well-followed
operations manual and a well-implemented maintenance schedule are key factors for an effective
and efficient EMCS. Commissioning and maintenance are covered in detail in Chap. 32.
5.15 FAULT DETECTION AND DIAGNOSTICS
Basics
Modern air-handling units (AHUs), packaged units (PUs), and chillers become more and more
complicated because of the IAQ, thermal comfort and energy efficiency requirements, and the use
of DDC. An HVAC&R operator is hardly able to monitor and detect the fault operations of the
AHUs, PUs, and chillers; find their causes; and correct them. An automatic fault detection and diagnostic
system is important for the effective operation and control, for the monitoring and maintenance,
and for the optimizing of utilization and continuous improving of the HVAC&R systems.
As of the late 1990s, fault detection and diagnostics are already a standard component in many
large PUs.
In HVAC&R operations, faults occur when the actual measured operating parameters deviate
from the normal operating values. There are two types of detected faults: complete failures and
performance degradations. Complete failures are abrupt faults that often cause discontinuation of
the operation of a system or component. Symptoms of abrupt faults can be easily observed. Peformance
degradation is the result of an evolving fault accumulated during a certain time.
The differences between the actually measured values of an operating parameter, such as temperature
T, F (C); pressure p, in. WG (Pa); volume flow rate , cfm (m3 / s); or mass flow rate ,
lb / s (kg/s) and the expected values (estimated, simulated, or set points) of temperature Texp, pressure
pexp, volume flow rate , or mass flow rate under normal operating conditions are
called residual. A fault can be detected by investigating and analyzing residuals.
A temperature residual Tres, F (C), a pressure residual pres, in. WC (Pa), or a volume flow rate
residual , cfm (m3 / s), can be calculated as
(5.21)
where the subscript exp indicates expected (predicted) values and the units of Texp, pexp, and are
the same as those of T, p, and . Most of the measured operating parameters in a fault detection and
diagnostic system are the same monitored parameters (sensed or measured) in an EMCS.
Residuals are often normalized so that the dominant symptom may have approximately the same
V?
V?
exp
V?
res  V? V?exp
pres  p pexp
Tres  T Texp
V?
res
m?exp V?exp
m? V?
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.61
magnitude for different types of faults. The residual R can be normalized as
(5.22)
where Rnor  normalized residual
Rmax, Rmin  maximum and minimum residuals
In the late 1990s, a large rooftop packaged unit made by one U. S. manufacturer offered a standard
fault detection and diagnostic device to review active and historical lists of diagnostic conditions.
A total of 49 different diagnostics can be read at the human interface panel, and the last 20
diagnostics can be held in an active history buffer log at the panel. A human interface panel provides
2-line 40-character liquid-crystal display and a 16-button keypad for monitoring, setting,
controlling, and diagnostics.
In the late 1980s, the earlier development of diagnostics of HVAC&R system operation was
mainly rule-based expert systems. During the late 1990s, the development of automating fault detection
and diagnosis was emphasized. Inputs and outputs of an HVAC&R operating process can
be mathematically related by using autoregressive models with exogenous inputs (ARX), artificial
neural network (ANN) models, and many other developing models. Both ARX and ANN are
called black-box because they require less physical knowledge of the operating process. These
technologies are expected to be commercially available after laboratory and field tests in the
early 2000s.
Expert System Rule-Based Diagnostics
Expert systems are discussed in Sec. 5.10, and diagnostics is one of the applications used in
HVAC&R system operations. Brambley et al. (1998) reported a toolthe outdoor air /economizer
(OAE) diagnosticianwhich monitors the performance of an air-handling unit to provide outdoor
air as a constant fraction of its supply flow rate and detects problems of outdoor ventilation air control
and economizer operation. An OAE diagnostician uses sensors that are commonly installed for
control purposes and diagnoses the operating problems based on rules derived from engineering
models of proper and improper performance of the AHUs. These rules are implemented in a decision
tree structure in computer software. The diagnostician collects data periodically from the
Building Automation System (BAS) to navigate the decision tree and produce conclusions. At each
point of the tree, a rule is evaluted according to the collected data, and the results determine in
which branch the diagnosis should be. At the end of the branch, a conclusion is reached corresponding
to the current operating condition of the AHU. The operator or installer of the diagnostician enters
data only once during setup.
The OAE diagnostician was installed and operated on three AHUs in a newly constructed and
occupied 200,000-ft2 (18,600-m2) DOE William R. Wiley Environmental Molecular Sciences Laboratory
and on four AHUs in a 72,700-ft2 (6760-m2) Technical Management Center, both in Rishland,
Washington.
For each AHU, data were recorded hourly from sensors in the BAS for outside temperature, return
air temperature, mixed air temperature, discharge air temperature, on/off status of the supply
fan, and open/closed status of the chilled water and hot water valves. No sensors were installed specially
for the diagnostician; all are used by the BAS for the control purpose. The data were automatically
transferred hourly from the BAS to the diagnosticians database. The diagnostician then
processed these data and produced the diagnostic results that can be viewed on the display.
Shortly after initial processing of data, four of the seven AHUs monitored were found to have
problems. The problems included a sensor malfunction, a return damper that was not closed fully, a
mixed-air sensor problem, and a chilled water control problem. All problems were confirmed after
inspection.
Rnor 
R Rmin
Rmax Rmin
5.62 CHAPTER FIVE
The diagnostician used color coding on a display to alert the HVAC&R system operator when
the problem happened and provided assistance in modifying the causes as well as making the
correction. The OAE diagnostician has proved itself effective in identifying installation and operating
problems during the field testing of the outdoor ventilation control and economizer cycle.
Dexter and Benouarets (1996) introduced a fault diagnosis of HVAC&R plant items based on
semiqualitative generic reference models. The scheme uses reference models describing fault-free
and fault operations that are collected from data producd by simulating a number of HVAC&R systems
(or subsystems) of the same type. Results have showed that a fuzzy-logic-based fault diagnosis
requiring no training on the actual plant can successfully identify faults in a simulated HVAC&R
subsystem if the parameter denoting faults is sufficiently large. The proposed fuzzy method of fault
diagnosis is computationally simple enough to be used for more complex subsystems, such as
AHUs.
Stylianou and Nikanpour (1996) present a methodology that uses thermodynamic modeling,
pattern recognition, and expert knowledge to determine whether a reciprocating chiller is fault-free
and to diagnose selected faults. The status of this chiller consists of the off cycle, start-up, and
steady-state operation. The off cycle deals with those faults that are more easily detected when the
chiller is off. Start-up deals with faults which are related to refrigerant flow characteristics and are
generally more apparent during the transient period. During steady-state operation, faults of performance
deterioration are detected.
This methodology requires training data. Data can be collected during commissioning and online
measurement of variables to build up normal steady-state linear regression models and data
collected from manufacturers. This methodology also needs additional treatment to improve the de-
finition of the thresholds to classify emerging patterns and to establish the range of applicability of
these patterns. Further study is required to produce a list of faults and their associated transient
characteristics and steady-state patterns.
ARX and ANN Model-Based Diagnosis
Peitsman and Bakker (1996), Peitsman and Soethout (1997), Lee et al. (1996), and Yoshida et al.
(1996) discussed the application of ARX and ANN model-based fault detection and diagnosis and
the related laboratory testing results. Model-based diagnosis is a technique capable of finding diagnoses
based on the behavior of the system and components. The behavior is best understood as
the interaction of observation and prediction expressed in the difference between the measured
(observed) and predicted parameter values as previously covered, called residual. The behavior of the
system and components is defined using ARX or ANN models. One advantage of a model-based diagnosis
over rule-based diagnosis is that the model-based diagnosis does not rely on symptom-fault
patterns. Such patterns are incomplete, and it is difficult for an expert to anticipate all possible faults
and to predict their symptoms. Another advantage is that a model-based diagnosis can be used for a
new system even when there is no repair experience or only when a system model is available.
ARX Model. Consider a process in an autoregressive model with exogenous inputs (ARX) having
an input signal u(t) and an output signal y(t) and described by a simple input/output relationship as
(5.23)
where t  time
e(t)  enter as a direct error
a  autoregressive parameter (order n)
b  exogenous parameter (order m)
The model relating output and input and extra (exogenous) variables is linear. The actual performance
of a system is nonlinear. However, such an approximation is allowable in fault detection and
y(t) 	 a1y(t 1) 	    	 any(t n)  b1u(t 1) 	    	 bmu(t m) 	 e(t)
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.63
diagnosis. Extra inputs are added, and the use of historical measured inputs (t 1, t 2, . . . )
will make it more nearly a nonlinear relationship.
Procedure to Identify Parameters. After the ARX model has been selected, according to Peitsman
and Bakker (1996), the procedure to identify parameters a and b in Eq. (5.23) is as follows:
 Measure input/output data. Input and output values of the process are collected in a data set that
consists of measured and design (predicted) data under healthy (fault-free) conditions. The
measured data sets should contain sufficient dynamic conditions of the process.
 Model order (m and n) should be selected as the optimum values by trial and error.
 From the collected input/ output data, estimate parameters a and b.
 Validate the selected model. The accuracy of the model is tested by comparing the predicted value
with the measured values.
ANN Models. For an ANN model of three layers with an input layer, a hidden layer, and an output
layer, as discussed in Sec. 5.10, the outputs can be calculated as a function of the weighted sum
of the inputs, and an additional factor, the biases B as expressed in Eqs. (5.16) and (5.17):
where iLn  input activation to a neuron in hidden layer
WLon  weights of connection between hidden layer and outputs
ANN models are nonlinear. A learning or training process in an ANN model is necessary. The initially
assigned connection weights are adjusted repeatedly during the learning and training process
until the error is within specified values, as discussed in Sec. 5.10. After training, an ANN model
can predict the outputs from given inputs.
Reliability of ARX and ANN Models. Peitsman and Bakker (1996) recognized that the reliability
of the ARX and ANN models (the use of mathematic equations to forecast the outputs) is good.
One of the conditions needed to produce a reliable model is a healthy and dynamic data set that is
measured under optimal and good operating conditions over a wide working range. Another necessary
condition is the availability of a fixed time interval (time step) because the regression models
use values of the previous time interval of the inputs and outputs to predict the output.
System and Component Models. In Peitsman and Bakker (1996), two types of models are produced
with the measured and predicted learned data:
 System models. A chiller or an AHU is considered a black box in which multiple independent inputs
estimate outputs. Independent variables are necessary to form a system model. This model
checks the whole system performance. In a system model if the possibility of a malfunction is detected,
the next step is to use the component model to localize this malfunction.
 Component models. Major components are considered as black boxes. The configuration of component
models is accomplished after studying the structure of the modeled HVAC&R systems.
The model becomes a little bit gray (with some physical knowledge) instead of black (without
physical knowledge). After the system model has detected a malfunction, the component model
can pinpoint the cause of the malfunction with a greater accuracy than the system model can.
Fault Detection and Diagnosis. After the parameters (variables) of a system model or component
model are simulated using an ARX or ANN model, the predicted outputs of the model must be
V  	iLnWLon
oon 
1
1 	 e(VB)
5.64 CHAPTER FIVE
compared with the measured outputs of the system or the components (temperature, pressure, volume
flow) as shown in Fig. 5.27. Only when the measured output is not within the threshold of the
predicted output may there be a detected fault or an incorrect model used.
The threshold is usually determined from the statistical properties of the process. A residual
(margin) of is used. Here x indicates the measured value, xexp is the assumed predicted
or expected mean value, and  is the standard deviation of the predicted values. The probability
that x falls within is 0.9973. If , a detected fault may
occur, as shown in Fig. 5.27, in which the difference between the measured and predicted discharge
air temperatures is greater than 3 . The possible causes of the fault include block of chilled water
flow, high chilled water temperature, and a fouled cooling coil.
Comparison of ARX and ANN Models. According to Peitsman and Bakker (1996), the comparison
of ARX and ANN models for fault detection and diagnosis is as follows:
ARX ANN
Prediction of process variables Linear Nonlinear
Results ANN gives slightly
better results than
ARX models
Training period Shorter Longer
 x xexp   3 
[xexp 3 , xexp 	 3 ]
 x xexp   3 
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.65
Measured
Predicted
Detected
fault
0
5
100 200 300 400
Time, s
TresTTexp
500 600 700 800
Discharge air temperature Tdis,F
Discharge air temperature Tdis,C
40
50
60
70
10
15
20
FIGURE 5.27 Comparison of the measured and predicted discharge air temperatures for a VAV system in an ARX
system model. (Source: Peitsman and Bakker (1996) ASHRAE Transactions Part I 6637. Reprinted with permission.)
5.16 CONTROLS IN ASHRAE/IESNA STANDARD 90.1-1999
ASHRAE/IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential
Buildings, specifies the following mandatory provisions for HVAC&R controls. For exceptions and
details refer to Standard 90.1-1999.
General
 Thermostatic controls. The supply of heating and cooling energy to each zone shall be controlled
individually by thermostatic controls using the zone temperature as the response. A dwelling unit
is permitted to be considered as a single zone.
 Dead band. When both heating and cooling are controlled, zone thermostatic controls shall be capable
of providing a temperature range or dead band of at least 5F (2.8C). Within that range the
heating and cooling energy supplied to the zone is shut off.
 Set-point overlap. When heating and cooling supply to a zone are controlled by separate zone
thermostatic controls, means such as limit switches, mechanical stops, and software for DDC systems,
shall be provided to prevent the heating set point exceeding the cooling set point minus any
proportional band.
Off-Hour Controls
HVAC systems with a design heating or cooling capacity greater than 65,000 Btu/h (19,050 W) and
fan system power greater than hp (0.56 kW) shall have all the following off-hour controls: automatic
shutdown, setback controls, optimum start controls, shutoff damper controls, and zone isolation.
 Automatic shutdown. HVAC systems shall be equipped with at least one of the following:
1. Controls that can start and stop the system under different time schedules, are capable of retaining
programming and time setting during loss of power for a period at least 10 hours, and
include an accessible manual override or equivalent that allows temporary operation of the
system for up to 2 hours.
2. An occupancy sensor that can shut the system off when no occupant is sensed for a period of
up to 30 minutes.
3. A manual operated timer that can adjust the system for two hours.
4. An interlock to a security system which shuts the system off when the security system is
activated.
 Setback controls. A heating system located where the heating outdoor design temperature is 40F
(4.4C) or less shall be equipped with controls that can automatically restart and temporarily operate
the system as required to maintain zone temperature above a heating set point adjustable
down to 55F (12.8C) or lower. A cooling system located where the cooling outdoor design temperature
is greater than 100F (37.8C) shall be equipped with controls that can automatically
restart and temporarily operate the system as required to maintain zone temperature below a cooling
set point adjustable up to 90F (32.2C) or higher or to prevent high space humidity.
 Optimum start controls. An air system with a supply volume flow rate exceeding 10,000 cfm
(4720 L/ s), served by one or more supply fans, shall have optimum controls. The optimum start
shall be at least a function of the difference between space temperature and occupied setpoint, and
the amount of time prior to scheduled occupancy.
 Shutoff damper controls. Both outdoor air supply and exhaust system shall be installed with motorized
dampers that will automatically shut off when the systems or the space are not in use. Out-
3	4
5.66 CHAPTER FIVE
door dampers for ventilation can be automatically shut off during morning warm-up and cooldown,
or nighttime setback except when ventilation reduces energy costs, or ventilation must be
supplied. Both outdoor air or exhaust dampers shall have a maximum air leakage of 3 cfm/ ft2
(15 L/ s m2) at a pressure difference of 1.0 in WC (250 Pa). As an exception, gravity dampers
(nonmotorized) are acceptable in buildings less than three stories in height or for buildings of any
height in climates with HDD65 less than 2700. Also excepted are systems with a design outside
air intake of 300 cfm (142 L/s) or less that are equipped with motor operated dampers.
 Zone isolation. HVAC systems serving zones that are intended to operate or be occupied nonsimultaneously
shall be divided into isolation areas. Each isolation area shall be no larger than
25,000 ft2 (2320 m2) of floor area nor include more than one floor. Each isolation area shall be
equipped wih isolation devices that can automatically shut off the supply of conditoned air, outside
air, and exhaust air from the area. Each isolation area can be controlled independently.
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Alpers, R., and Zaragoza, J., Air Quality Sensors for Demand Controlled Ventilation, Heating/Piping/Air
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Anderson, R., Gems to Look for in EMCS, Heating/Piping/Air Conditioning, no. 11, 1991, pp. 4752.
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ASHRAE, ASHRAE Handbook 1995, HVAC Applications, Atlanta, GA, 1995.
ASHRAE/IESNA, Standard 90.11999, Energy Standard for Buildings Except Low-Rise Residential Buildings,
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Bekker, J. E., Meckl, P. H., and Hittle, D. C., A Tuning Method for First Order Processes with PI Controllers,
ASHRAE Transactions, 1991, Part II, pp. 1923.
Brambley, M., Pratt, R., Chassin, D., Katipamula, S., and Hatley, D., Diagnostics for Outdoor Air Ventilation
and Economizers, ASHRAE Journal no. 10, 1998, pp. 4955.
Brothers, P., and Cooney, K., A Knowledge-Based System for Comfort Diagnostics, ASHRAE Journal, no. 9,
1989, pp. 6067.
Bushby, S. T., and Newman, M., BACnet: A Technical Update, ASHRAE Journal, no. 1, 1994, pp. S72S84.
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Dexter, A. L., and Benouarets, M., A Generic Approach to Identifying Faults in HVAC Plants, ASHRAE Transactions,
1996, Part I, pp. 550556.
Elyashiv, T., Beneath the Surface: BACnet Data Link and Physical Layer Options, ASHRAE Journal, no. 11,
ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.67
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ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.69
6.1
CHAPTER 6
LOAD CALCULATIONS
6.1 SPACE LOAD CHARACTERISTICS 6.2
Space, Room, and Zone 6.2
Convective and Radiative Heat 6.2
Space and Equipment Loads 6.3
Night Shutdown Operating Mode 6.3
Influence of Stored Heat 6.6
6.2 COOLING LOAD AND COIL LOAD
CALCULATIONS 6.6
Components of Cooling Load 6.6
Components of Cooling Coil Load 6.7
Difference between Cooling Load and
Cooling Coil Load 6.8
Load Profile 6.9
Peak Load and Block Load 6.9
Characteristics of Night Shutdown
Operating Mode 6.10
Moisture Transfer from the Building
Structures 6.10
6.3 HISTORICAL DEVELOPMENT OF
COOLING LOAD CALCULATIONS 6.11
6.4 METHODOLOGYHEAT
BALANCE 6.12
The Physical Model 6.12
Heat Balance Equations 6.12
Heat Balance of Space Air 6.13
Characteristics of Heat
Balance Method 6.14
6.5 METHODOLOGYTRANSFER
FUNCTION 6.14
Basics 6.14
Transfer Function and Time Function 6.14
Calculation Procedure 6.15
CLTD/SCL/CLF Method 6.15
TETD/TA Method 6.15
6.6 DETAILED CALCULATION
PROCEDURES FOR TFM 6.16
Conduction Heat Gain through Exterior
Walls and Roofs 6.16
Heat Gain through Ceilings, Floors, and
Interior Partition Walls 6.16
Solar Heat Gain and Conductive Heat
Gain through Window Glass 6.17
Internal Heat Gain 6.17
Infiltration 6.24
Cooling Load Conversion Using Room
Transfer Function 6.24
Space Cooling Load Calculation 6.25
Heat Extraction Rate and Space Air
Transfer Function 6.25
Heat Loss to Surroundings 6.25
6.7 DETAILED CALCULATION PROCEDURE
USING CLTD/SCL/CLF METHOD 6.26
Space Cooling Load due to Heat Gain
through Exterior Walls and Roofs and
Conductive Gain through Glass 6.26
Space Cooling Load due to Solar Heat
Gain through Fenestration 6.28
Space Cooling Load due to Heat Gain
through Wall Exposed to
Unconditioned Space 6.28
Calculation of Internal Cooling Loads and
Infiltration 6.29
Space Cooling Load of Night Shutdown
Operating Mode 6.32
6.8 COOLING COIL LOAD 6.32
Basics 6.32
Fan Power 6.33
Duct Heat Gain 6.33
Temperature of Plenum Air and
Ventilation Load 6.34
6.9 COOLING LOAD CALCULATION BY
FINITE DIFFERENCE METHOD 6.34
Finite Difference Method 6.34
Simplifying Assumptions 6.36
Heat and Moisture Transfer at
Interior Nodes 6.36
Heat and Moisture Transfer at
Surface Nodes 6.37
Space Air Temperature and
Cooling Loads 6.38
6.10 HEATING LOAD 6.39
Basic Principles 6.39
Transmission Loss 6.39
Adjacent Unheated Spaces 6.40
Latent Heat Loss and Heat Loss from
Products 6.41
Infiltration 6.41
Setback of Night Shutdown
Operation 6.41
6.11 LOAD CALCULATION
SOFTWARE 6.42
Introduction 6.42
TRACE 600Structure and Basics 6.42
TRACE 600 InputLoad
Methodology 6.43
TRACE 600 InputJob 6.44
TRACE 600 InputExternal Loads 6.45
TRACE 600 InputSchedules 6.45
TRACE 600 InputInternal Loads 6.46
TRACE 600 InputPartition and Shading
Devices 6.47
TRACE 600Minimum Input
Requirements, Run, and Outputs 6.47
REFERENCES 6.49
6.1 SPACE LOAD CHARACTERISTICS
Space, Room, and Zone
Space indicates either a volume or a site without a partition or a partitioned room or group of
rooms. A room is an enclosed or partitioned space that is usually treated as a single load. A conditioned
room often has an individual control system. A zone is a space, or several rooms, or units of
space having some sort of coincident loads or similar operating characteristics. A zone may or may
not be an enclosed space, or it may consist of many partitioned rooms. It could be a conditioned
space or a space that is not air conditioned. A conditioned zone is always equipped with an individual
control system. A control zone is the basic unit of control.
Convective and Radiative Heat
Whether heat enters the conditioned space from an external source or is released to the space from
an internal source, the instantaneous heat gains of the conditioned space can be classified into two
categories: convective heat and radiative heat, as shown in Fig. 6.1. When solar radiation strikes the
outer surface of a concrete slab, most of its radiative heat is absorbed by the slab; only a small portion
is reflected. After absorption, the outer surface temperature of the slab increases. If the slab and
the conditioned space are in thermal equilibrium originally, there is then a convective heat and
radiative heat transfer from the surface of the slab to the space air and other surfaces. Meanwhile,
heat transfer due to conduction takes place from the surface to the inner part of the slab. Heat is
then stored inside the slab. The stored heat is released to the space air when the surface temperature
falls below the temperature of the inner part of the slab.
6.2 CHAPTER SIX
FIGURE 6.1 Convective and radiant heat in a conditioned space and the temperatures of the interior surfaces.
To maintain the preset space air temperature, the heat that has been convected or released to the
conditioned space should be removed from the space instantaneously.
Space and Equipment Loads
The sensible and latent heat transfer between the space air and the surroundings can be classified as
follows:
1. Space heat gain qe, in Btu/h (W), represents the rate at which heat enters a conditioned space
from an external source or is released to the space from an internal source during a given time interval.
2. Space cooling load, often simply called the cooling load Qrc, Btu /h (W), is the rate at which
heat must be removed from a conditioned space so as to maintain a constant temperature and acceptable
relative humidity. The sensible cooling load is equal to the sum of the convective heat transfer
from the surfaces of the building envelope, furnishings, occupants, appliances, and equipment.
3. Space heating load Qrh, Btu /h (W), is the rate at which heat must be added to the conditioned
space to maintain a constant temperature and sometimes a specified relative humidity.
4. Space heat extraction rate Qex, Btu /h (W), is the rate at which heat is actually removed from
the conditioned space by the air system. The sensible heat extraction rate is equal to the sensible
cooling load only when the space air temperature remains constant.
5. Coil load Qc, Btu /h (W), is the rate of heat transfer at the coil. The cooling coil load Qcc,
Btu/h (W), is the rate at which heat is removed by the chilled water flowing through the coil or is
absorbed by the refrigerant inside the coil.
6. The heating coil load Qch, Btu /h (W), is the rate at which heat is added to the conditioned air
from the hot water, steam, or electric heating elements inside the coil.
7. Refrigerating load Qrl, Btu /h (W), is the rate at which heat is absorbed by the refrigerant at
the evaporator. For central hydronic systems, the refrigerating load is the sum of the coil load plus
the chilled water piping heat gain, pump power heat gain, and storage tank heat gain. For most
water systems in commercial buildings, the water piping and pump power heat gain is only about
5 to 10 percent of the coil load. In an air conditioning system using DX coil(s), the refrigerating
load is equal to the DX coil load.
The instantaneous sensible heat gain of a conditioned space is not equal to the instantaneous
sensible cooling load because of storage of part of the radiative heat inside the building structures.
Such phenomenon results in a smaller instantaneous cooling load than that of the heat gain when it
is at its maximum value during a diurnal cycle, as shown in Fig. 6.2.
If the space relative humidity is maintained at an approximately constant value, the storage
effect of the moisture in the building envelope and furnishings can be ignored. Then the instantaneous
space latent heat gain will be the same as the instantaneous space latent cooling load.
Night Shutdown Operating Mode
In many commercial buildings, air systems are often shut down during the nighttime or during unoccupied
periods. The operating characteristics of the conditioned space in a 24-h diel cycle can
then be divided into three periods, as shown in Fig. 6.3.
Night Shutdown Period. This period commences when the air system is switched off and ends
when the air system is switched on again. When the air-handling unit, packaged unit, or terminal
unit is turned off during summer in a hot and humid area, the infiltrated air (through the lift, pipe
shafts, and window cracks) and any heat transfer from the window glass, external wall, or roofs will
cause a sudden increase in indoor temperature Tr of a few degrees Fahrenheit (Celsius). After that,
LOAD CALCULATIONS 6.3
Tr will rise farther or drop slightly, depending on the difference between outdoor and indoor temperatures.
At the same time, the indoor space relative humidity increases gradually because of the infiltration
of hot and humid outdoor air and the moisture transfer from the wetted surfaces in the air
system. Higher space relative humidity causes a moisture transfer from the space air to the building
envelopes and furnishings.
During winter, after the air system is turned off, the indoor temperature drops because of the
heat loss through external windows, walls, and roofs, as well as the infiltration of outdoor cold air.
Meanwhile, the space relative humidity increases mainly owing to the fall of the indoor space temperature.
Cool-down or Warm-up Period. This period commences when the air system begins to operate
and ends when the space temperature or other controlled variables have attained predetermined
limits. During summer, the supply of cold and dehumidified air after the air system is turned on
causes a sudden drop in space air temperature and relative humidity. Both heat and moisture are
transferred from the building envelope to the space air because of the comparatively higher temperature
and moisture content of the building envelope. These heat and moisture transfers form
the cool-down cooling load. Sometimes, the cool-down load can be the maximum summer design
cooling load.
6.4 CHAPTER SIX
FIGURE 6.2 Solar heat gain and heat gain from electric lights and the corresponding space cooling loads
for a night shutdown air system.
6.5
FIGURE 6.3 Operating characteristics of an air conditioned space operated at night shutdown mode: (a) summer
cooling mode (hot and humid area) and (b) winter mode.
If refrigeration is used for cooling during the cool-down period, then it usually lasts less than
1 h, depending mainly on the tightness of the building and the differences of the space temperatures
and humidity ratios between the shutdown and cool-down periods. If the cool-down of the space air
and the building envelope is by means of the free cooling of outdoor air, the cool-down period may
last for several hours.
During winter, the supply of warm air during the warm-up period raises the space temperature
and lowers the space relative humidity. Because of the higher space temperature, heat transfer from
the space air to the colder building envelope and furnishings, and the heat energy required to raise
the temperature of the building envelope, forms the warm-up heating load.
Conditioning Period. This period commences when the space air temperature has fallen or risen
to a value within the predetermined limits. It ends when the air system is shut down. In summer,
cold and dehumidified air is usually supplied to the space to offset the space cooling load and to
maintain a required temperature and relative humidity. During winter, warm air is supplied to
compensate for heat losses from the conditioned space. Space temperature is often controlled and
maintained within predetermined limits by control systems.
In commercial buildings located in areas with a cold winter, the air system may operate in dutycycling
mode during nighttime, unoccupied periods in winter, or intermediate seasons to maintain a
night setback temperature. In duty-cycling mode, the fan and heater turn on and off to maintain a
desired temperature. Space temperature is set back at night, e.g., to 55 or 60F (12.8 or 15.6C), to
prevent freezing of water pipes and to produce a comparatively smaller temperature lift during a
warm-up period. The operation of the air system in a 24-h diel cycle in winter is then divided into
night setback period, warm-up period, and conditioning period. The required load on the heating
coil during a morning warm-up period at winter design conditions is usually the winter design heating
load.
Influence of Stored Heat
The curve of solar heat gain from a west window is shown in Fig. 6.2, as well as the cooling load
curve due to this solar heat gain for a conditioned space operated at night shutdown mode in summer.
The difference between the maximum solar heat gain qhgm and the maximum cooling load Qclm
during the conditioning period indicates the amount of heat stored inside the building structures.
This difference significantly affects the size of air conditioning equipment required. The amount
of heat stored depends mainly on the mass of the building envelope (whether it is heavy, medium,
or light), the duration of the operating period of the air system within a 24-h cycle, and the characteristics
of heat gain, whether radiant heat or convective heat predominates. For solar radiation
transmitted through a west window, Qclm may have a magnitude of only 40 to 60 percent of qhgm.
ASHRAE Handbook divides the mass of the building construction into the following three
groups:
Heavy construction: approximately 130 lb/ ft2 (634 kg/m2) floor area
Medium construction: approximately 70 lb/ ft2 (342 kg/m2) floor area
Light construction: approximately 30 lb/ ft2 (146 kg/m2) floor area
6.2 COOLING LOAD AND COIL LOAD CALCULATIONS
Components of Cooling Load
Cooling load calculations for air conditioning system design are mainly used to determine the volume
flow rate of the air system as well as the coil and refrigeration load of the equipmentto size
6.6 CHAPTER SIX
the HVAC&R equipment and to provide the inputs to the system for energy use calculations in
order to select optimal design alternatives. Cooling load usually can be classified into two categories:
external and internal.
External Cooling Loads. These loads are formed because of heat gains in the conditioned space
from external sources through the building envelope or building shell and the partition walls.
Sources of external loads include the following cooling loads:
1. Heat gain entering from the exterior walls and roofs
2. Solar heat gain transmitted through the fenestrations
3. Conductive heat gain coming through the fenestrations
4. Heat gain entering from the partition walls and interior doors
5. Infiltration of outdoor air into the conditioned space
Internal Cooling Loads. These loads are formed by the release of sensible and latent heat from
the heat sources inside the conditioned space. These sources contribute internal cooling loads:
1. People
2. Electric lights
3. Equipment and appliances
If moisture transfers from the building structures and the furnishings are excluded, only infiltrated
air, occupants, equipment, and appliances have both sensible and latent cooling loads. The remaining
components have only sensible cooling loads.
All sensible heat gains entering the conditioned space represent radiative heat and convective
heat except the infiltrated air. As in Sec. 6.1, radiative heat causes heat storage in the building
structures, converts part of the heat gain into cooling load, and makes the cooling load calculations
more complicated. Latent heat gains are heat gains from moisture transfer from the occupants,
equipment, appliances, or infiltrated air. If the storage effect of the moisture is ignored, all release
heat to the space air instantaneously and, therefore, they are instantaneous cooling loads.
Components of Cooling Coil Load
If the conductive heat gain from the coils framework and the support is ignored, the cooling coil
load consists of the following components, as shown in Fig. 6.4 by the summer air conditioning
cycle O-m-cc-s-r-rf-m of a constant volume of supply air, single supply duct, and serving a
single zone.
1. Space cooling load Qrc, including sensible and latent load
2. Supply system heat gain qss, because of the supply fan heat gain qsf , and supply duct heat
gain qsd
3. Return system heat gain qrs because of heat gains of recessed electric lights and ceiling plenum
qrp, of return duct qrd , and return fan qrf , if any
4. Sensible and latent load because of the outdoor ventilation rates Qo to meet the requirements of
the occupants and others
In Fig. 6.4, the summer air conditioning cycle O-m-cc-s-r-rf-m consists of an adiabatic mixing
process O-m-rf, a cooling and dehumidifying process m-cc, a supply system heat gain process cc-s,
a space conditioning process r-s, and a return system heat gain process r-rf. Here, O indicates
the status of outdoor air, m the mixture of outdoor air and recirculating air, cc the conditioned air
LOAD CALCULATIONS 6.7
leaving the coil, s the supply air, r the conditioned space, and rf the recirculating air. All the cooling
loads and heat gains are in Btu/h (W). Usually, both the supply and return system heat gains are
sensible loads. These components are absorbed by the supply air and return air and appear as cooling
and dehumidifying loads at the cooling coil during summer.
Difference between Cooling Load and Cooling Coil Load
For the same air conditioning cycle shown in Fig. 6.4, note the following:
1. The space cooling load is represented by Qrc, and the cooling coil load is represented by Qc.
Since supply system heat gain qss and return system heat gain qrs are both instantaneous cooling
loads, then
Qc  Qrc  qss  qrs  Qo (6.1)
where Qo  load due to the outdoor ventilation air intake, Btu/h (W).
2. The space cooling load is used to determine the supply volume flow rate , whereas the coil
load is used to determine the size of the cooling coil in an air-handling unit or DX coil in a
packaged unit.
3. A cooling load component influences both and the size of the cooling coil, whereas a cooling
coil load component may not affect .
4. Infiltration heat gain is an instantaneous cooling load. From Fig. 6.4, it is apparent that the load
due to the outdoor ventilaion air Qo, sometimes called the ventilation load, is a coil load. If Qo is
considered a cooling load, the volume flow rate of the air system will be oversized.
V?
s
V?
s
V?
s
6.8 CHAPTER SIX
FIGURE 6.4 Difference between cooling load and cooling coil load on
psychrometric chart.
Load Profile
A load profile shows the variation of space load within a certain time period, such as a 24-h operating
cycle or an annual operating cycle, as shown in Fig. 6.5. In a load profile, the space load is
always plotted against time. For a space cooling load, the magnitude of the curve is positive; for a
space heating load, the magnitude is negative. A load profile may be used to illustrate the load variation
of an air conditioned spacea room, a zone, a floor, a building, or a project. The shape of a
load profile depends on the outdoor climate and, therefore, the latitude, orientation, and structure of
the building. It is also affected by the operating characteristics and the variation of the internal
loads.
The load duration curve is the plot of number of hours versus the load ratio. The load ratio is
defined as the ratio of cooling or heating load to the design full load, both in Btu/ h, over a certain
period. The period may be a day, a week, a month, or a year. The load duration curve is helpful in
many operating and energy consumption analyses.
Peak Load and Block Load
The zone peak load is the maximum space cooling load in a load profile of a control zone of the
same orientation and similar internal load characteristics calculated according to the daily outdoor
dry-bulb and coincident wet-bulb temperature curves containing summer or winter outdoor design
conditions. For a zone cooling load with several components, such as solar load through window
glass, heat transfer through roofs, or internal load from electric lights, the zone peak load is always
LOAD CALCULATIONS 6.9
FIGURE 6.5 Load profile, peak load, and block load.
the maximum sum of these zone cooling load components at a given time. The block load is the
maximum sum of the zone cooling loads of various load profiles of the control zones of a building
floor or a building at the same time. The block load of a space, room, floor, or building is the maximum
sum of cooling load components in that space, room, floor, or building at a given time.
For air systems, the supply volume flow rate required for a control zone is calculated based
on the zone peak load; and the supply volume flow rate for a specific area, room, floor, or building
should be calculated based on the block load (cooling) of this specific area, room, floor, or
building. For conditioned space using variable-air-volume systems and space air conditioning
systems, the required cooling coil load or refrigeration load can be calculated based on block
load of the corresponding specific area that air system serves.
The heating load is usually the peak heating load of a space, room, zone, floor, or building in a
load profile. Block load is not needed in heating load calculations because the solar radiation and
internal loads are not taken into account. Design heating load depends entirely on indoor-outdoor
temperature difference.
Suppose a typical floor of a multistory building has five zones: a north, a south, an east, and a
west perimeter zone and an interior zone. For each zone, there is a corresponding daily cooling load
profile curve and a zone peak load QN,13, Qs,12, QE,8, QW,16, and QI,14. Here, subscripts N, S, E, W,
and I indicate north, south, east, west, and interior, respectively; 8, 12, 13, 14, and 16 represent time
at 8, 12, 13, 14, and 16 h, respectively.
The block load of this typical floor Qmax,t is not calculated as the sum of the zone peak loads
QN,13, Qs,12, QE,8, QW,16, and QI,14. Rather the block load is calculated for a specific time. For example,
block load Qmax.14, Btu /h (W), is the block load at 14 h, i.e.,
Qmax.14  QN,14  QS,14  QE,14  QW,14  QI,14 (6.2)
Characteristics of Night Shutdown Operating Mode
Compared with the continuous operating mode, night shutdown operating mode has the following
characteristics:
1. A greater peak load, therefore a higher air system and initial cost
2. A higher maximum heat extraction rate
3. A smaller accumulated heat extraction rate over the 24-h operating cycle
4. A lower power consumption for the fans, compressors, and pumps
Moisture Transfer from the Building Structures
In hot and humid climates, the frequency of outdoor wet-bulb temperatures higher than 73F
(22.8C) exceeds 1750 h during the six warmest consecutive months annually. At the same time, the
air infiltrated through the elevator and pipe shafts during shutdown periods often causes a temperature
increase of more than 7F (3.9C) and a relative humidity increase exceeding 10 percent. These
result in a significant increase of the latent load because of the moisture transfer from the building
envelope and furnishings to the space air during the cool-down period in summer. Such an increase
in the coils latent load not only necessitates a greater refrigerating capacity, but also lowers
the sensible heat ratio of the space conditioning process. A space conditioning process removes
heat and moisture from the space by the supply air, or sometimes adds moisture and heat to
compensate for space heat losses. In industrial applications where a very low relative humidity is
6.10 CHAPTER SIX
maintained in the conditioned space, the latent load due to the moisture transfer from the building
envelope should be added to the space cooling load calculations during night shutdown operation
mode.
6.3 HISTORICAL DEVELOPMENT OF COOLING LOAD
CALCULATIONS
In the 1930s, Houghton et al. introduced the analysis of heat transmission through the building
envelope and discussed the periodic heat flow characteristics of the building envelope. In 1937,
ASHVE Guide introduced a systematic method of cooling load calculation involving the division of
various load components. In the ASHVE Guide, solar radiation factors were introduced and their
influence on external walls and roofs was taken into consideration. Both the window crack and
number-of-air-changes methods were used to calculate infiltration.
Mackey and Wright first introduced the concept of sol-air temperature in 1944. In the same
paper, they recommended a method of approximating the changes in inside surface temperature
of walls and roofs due to periodic heat flow caused by solar radiation and outside temperature
with a new decrement factor. In 1952, Mackey and Gay analyzed the difference between the
instantaneous cooling load and the heat gain owing to radiant heat incident on the surface of the
building envelope.
In 1964, Palmatier introduced the term thermal storage factor to indicate the ratio between the
rate of instantaneous cooling load in the space and rate of heat gain. One year after, Carrier Corporation
published a design handbook in which the heat storage factor and equivalent temperature
difference (ETD) were used to indicate the ratio of instantaneous cooling load and heat gain because
of the heat storage effect of the building structure. This cooling load calculation method was
widely used by many designers until the current ASHRAE methods were adopted.
In 1967, ASHRAE suggested a time-averaging (TA) method to allocate the radiant heat over
successive periods of 1 to 3 h or 6 to 8 h, depending on the construction of the building structure.
Heat gains through walls and roofs are tabulated in total equivalent temperature differentials
(TETDs). In the same year, Stephenson and Mitalas recommended the thermal response factor,
which includes the heat storage effect for the calculation of cooling load. The thermal response
factor evaluates the system response on one side of the structure according to random temperature
excitations on the other side of the structure. This concept had been developed and forms the basis
of the weighting factor method (WFM) or transfer function method (TFM) in the 1970s. In 1977,
ASHRAE introduced a single-step cooling load calculation procedure that uses the cooling load
factor (CLF) and cooling load temperature difference (CLTD); these are produced from the simpli-
fied TFM.
In 1963, Kusuda and Achenbach used a digital computer to analyze the thermal environment of
occupied underground space. The use of computers to design building mechanical systems was first
accepted in 1965 when a group of mechanical engineering consultants organized Automated Procedures
for Engineers Consultants Inc. (APEC) because of sharing of software and development
costs.
Because of the need for computerized load and energy calculations, ASHRAE established a
Committee on Energy Consumption in 1965 and renamed the Task Group on Energy Requirements
(TGER) for Heating and Cooling in 1967. In the mid-1970s, ASHRAE and the National Bureau of
Standards (NBS) published the computerized calculation of heating and cooling loads in energyestimating
programs. In 1980, the U.S. Department of Energy sponsored a computer program for
energy estimation and load calculation through hour-by-hour detailed system simulation, called
DOE-2, which was published through Los Alamos National Laboratory and Lawrence Berkeley
Laboratory. In this program, a custom weighting factor method for various room configurations is
used for heating and cooling load calculations. Many computerized thermal load and energy calculating
software programs had been developed in the 1980s. Since the 1980s, because of the wide
LOAD CALCULATIONS 6.11
adoption of personal computers, the use of computer-aided HVAC&R design was rapidly increased
and many thermal load and energy analysis programs were developed in this period.
6.4 METHODOLOGYHEAT BALANCE
The Physical Model
The exact method to calculate the space cooling load is to use heat balance equations to determine
the temperature of the interior surfaces of the building structure at time t simultaneously and then to
calculate the space sensible cooling load, which is equal to the sum of the convective heat transfer
from these surfaces, latent cooling loads, and the cooling load due to infiltrated air at time t.
Consider a typical air conditioned room, as shown in Fig. 6.1. In this room, the building envelope
consists of mainly walls, window, ceiling, and floor. There is also heat transfer from occupants,
electric lights, and equipment. The heat transfer between various surfaces takes place under the
following assumptions:
 Only one-dimensional transient heat flow through the building envelope is considered.
 The room air is perfectly mixed with the supply air so that the resulting room air temperature is
uniform.
 The materials of the building envelope are homogeneous. The surface temperature, the surface
heat-transfer coefficient, and the absorptivity for each surface are uniform values. Reflectivity is
very small and can be ignored.
 The cooling load is calculated based on the mean value of a fixed time interval, such as 1 h.
Heat Balance Equations
Any interior surface of this air conditioned room may receive conduction heat transfer at time t,
denoted by qi,t, Btu /h (W), from the adjacent building material. Each interior surface receives shortwave
solar radiation from the window glass and longwave radiative heat transfer from other interior
surfaces and from the surfaces of the lighting fixtures, appliances, equipment, and occupants. Convective
heat transfer is also present between these interior surfaces and room air.
The sensible heat balance at the ith surface (for i  1, 2, . . ., m) at time t can be expressed as
follows:
(6.3)
where hci  convective heat-transfer coefficient of ith surface, Btu /h ft2 F (W/m2 C)
hij  radiative heat-transfer coefficient between interior surface i and j, i  j, Btu/h ft2 F
(W/m2 C)
Tr,t  room air temperature at time t, F (C)
Ti,t  temperature of ith surface at time t, F (C)
Tj,t  temperature of jth surface at time t, F (C)
Ai  area of ith surface, ft2 (m2)
Sir, t  solar radiation transmitted through window glass and absorbed by ith surface at time t,
Btu/h (W)
Lir, t  radiative energy from electric lights and absorbed by ith surface at time t, Btu /h (W)
Eir, t  radiative energy from equipment and absorbed by ith surface at time t, Btu /h (W)
Oir, t  radiative energy from occupants and absorbed by ith surface at time t, Btu /h (W)
qi, t  [hci (Tr, t  Ti, t)  k
j1
hij (Tj, t  Ti, t)] Ai  Sir, t  Lir, t  Eir, t  Oir, t
6.12 CHAPTER SIX
Because the conductive heat and radiative heat in Eq. (6.3) are all continuous functions of time f (t),
the transient conductive heat transfer qi, t either can be solved by a partial differential equation using
numerical solutions (finite difference method) or can be solved by means of the conduction transfer
function [Eq. (6.10)].
Usually, the room air temperature Tr,t is considered as constant, and Eq. (6.3) and the conduction
transfer function equation can be solved simultaneously to determine the interior surface
temperatures Ti, t. Then the sensible cooling load at time t, denoted by Qrs,t, Btu /h (W), can be
calculated as
(6.4)
where  volume flow rate of infiltrated air at time t, cfm [m3/(60 s)]
 air density, lb/ ft3 (kg/m3)
cpa  specific heat of moist air, Btu / lbF (J /kgC)
To,t  temperature of outdoor air at time t, F(C)
Sc,t  solar heat coming through windows and convected into room air at time t, Btu /h (W)
Lc,t  sensible heat from electric lights and convected into room air at time t, Btu /h (W)
Ec,t  sensible heat from equipment and convected into room air at time t, Btu /h (W)
Oc,t  sensible heat from occupants and convected into room air at time t, Btu /h (W)
Since the latent heat gains convert to latent cooling loads instantaneously, therefore the space latent
cooling load at time t, or Qrl,t, Btu /h (W), can be calculated as
(6.5)
where qil, t  latent heat from ith interior surface and convected into room air at time t, Btu /h (W)
wo, t  humidity ratio of infiltrated air at time t, lb/ lb (kg /kg)
wr, t  humidity ratio of room air at time t, lb/ lb (kg /kg)
hfg  latent heat of condensation, Btu/ lb (J /kg)
El, t  latent heat from equipment at time t, Btu /h (W)
Ol, t  latent heat from occupants at time t, Btu /h (W)
Then the space cooling load at time t, denoted by Qrc, t, Btu /h (W), is
Qrc,t  qrs,t  qrl,t (6.6)
Heat Balance of Space Air
The conditioned air for this room or space is supplied from the ceiling diffuser at a mass flow rate
at time t of . Here, indicates the volume flow rate of supply air at time t, cfm
(m3/min), and s is the density of supply air, lb/ ft3 (kg/m3). The supply air is then mixed with the
space air. The resulting mixture absorbs convective sensible heat and latent heat from various surfaces.
At time t, the outdoor air infiltrates into the space at a volume flow rate , cfm [m3 / (60 s)].
The sensible heat balance of the space air can be expressed as
(6.7)
where Ts,t  supply air temperature at time t, F (C). If the space air temperature is allowed to
float, Eq. (6.3), the transient conductive equation, and Eq. (6.7) must be solved simultaneously.
 Sc, t  Lc, t  Ec, t  Oc, t  m
i1
hci Ai(Ti, t  Tr, t)  60V?if, t cpa (To, t  Tr, t)
60 V?s, t scpa (Tr, t  Ts,t)
V?
if, t
V?
s,t m?s,t  V?s, t s
Qrl, t  qil, t  60V?if  (wo, t  wr, t)hfg  El, t  Ol, t
V?
if, t
Qrs, t  m
i1
hci (Ti, t  Tr, t) Ai  60V?if,t cpa(To, t  Tr, t)  Sc, t  Lc, t  Ec, i  Oc, t
LOAD CALCULATIONS 6.13
Characteristics of Heat Balance Method
The heat balance method is more direct and clear in load calculation methodology. Using the heat
balance method, the assumption of linear supposition is not required, and the changing of certain
parameters, such as the surface convective heat-transfer coefficient, can be modeled as required. If
moisture transfer should be included in the cool-down period in nighttime shutdown mode in a
location where the outdoor climate is hot and humid in summer, then the heat balance method will
give comparatively more accurate results.
However, the heat balance method demands laborious work, more computing time, complicated
computer programs, and experienced users. Only expensive mainframe computers could run
computer programs adopting the heat balance methodology in the 1970s and early 1980s. The heat
balance method is generally used for research and analytical purposes.
6.5 METHODOLOGYTRANSFER FUNCTION
Basics
The transfer function method or weighting factor method is a simplification of the laborious heat
balance method. The wide application of the TFM is due to the user-friendliness of the inputs and
outputs of the TFM software and the saving of computing time. In the transfer function method,
interior surface temperatures and the space cooling load were first calculated by the exact heat
balance method for many representative constructions. The transfer function coefficients (weighting
factors) were then calculated which convert the heat gains to cooling loads. Sometimes, transfer
function coefficients were also developed through test and experiments.
The room transfer function coefficients (weighting factors ) were originally generated based on
a typical room configuration of 15 ft 	 15 ft with 10-ft (4.5 m 	 4.5 m with 3-m) ceiling and one
exposure of 30 percent glass in the early 1970s . In the late 1980s, the introduction of 14 influential
parameters of zone characteristics by Sowell (1988) enabled the adopted room transfer function
coefficients to more closely match the room type to be used. Today, TFM is the most widely
adopted computer-aided load calculation method in HVAC&R consulting firms.
Transfer Function and Time Function
The transfer function K of an element or a system is the ratio of the Laplace transform of the output
Y to the Laplace transform of the input or driving force G, or
Y  KG (6.8)
When a continuous function of time f (t) is represented at regular intervals t and its magnitudes are
f (0), f (), f (2), . . ., f (n), the Laplace transform is given as
(6.9a)
where   time interval, h
z  es
The preceding polynomial in z1 is called the z transform of the continuous function f (t).
In Eq. (6.8), Y, K, and G can all be represented in the form of a z transform. Because of the
radiative component and the associated heat storage effect, the space sensible cooling load at time t
can be related to the sensible heat gains and previous cooling loads in the form of a continuous
function of time f (t), which can be expressed as a z transform.
	(z)  f (0)  f ()z1  f (2)z2      f (n)zn
6.14 CHAPTER SIX
Weighting factors are transfer function coefficients presented in the form of z transform functions.
Weighting factors are so called because they are used to weight the importance of current and
historical values of heat gain and cooling load on currently calculated cooling loads.
Calculation Procedure
The calculation of space cooling load using the transfer function method consists of two steps.
First, heat gains or heat loss from exterior walls, roofs, and floors is calculated using response
factors or conduction transfer function coefficients; and the solar and internal heat gains are calculated
directly for the scheduled hour. Second, room transfer function coefficients or room weighting
factors are used to convert the heat gains to cooling loads, or the heat losses to heating loads.
As described in Sec. 6.2, the sensible infiltration heat gain is the instantaneous sensible cooling
load. All latent heat gains are instantaneous latent cooling loads.
The TFM is limited because the cooling loads thus calculated depend on the value of transfer
function coefficients as well as the characteristics of the space and how they are varied from those
used to generate the transfer function coefficients. In addition, TFM assumed that the total cooling
load can be calculated by simply adding the individual componentsthe superposition principle.
However, this assumption can cause some errors.
CLTD/SCL/CLF Method
The cooling load temperature difference (CLTD)/solar cooling load (SCL)/cooling load factor
(CLF) method first calculates the sensible cooling load based on the TFM. The result is divided by
the U value, shading coefficient, or sensible heat gain to generate the CLTD, SCL, or CLF. Thus, it
provides a direct, one-step space cooling load calculation instead of a heat gaincooling load conversion,
a two-step calculation in TFM. Cooling load calculation using the CLTD/SCL/CLF
method can be either computer-aided or performed manually for a check or rough estimate. The
CLTD/SCL/CLF method is one of the members of the TFM family.
In the CLTD/SCL/CLF method, the CLTD is used to calculate the sensible cooling load for the
exterior wall and roofs. Recently, an SCL factor has been added which represents the product of the
solar heat gain at that hour and the fraction of heat storage effect due to various types of room
construction and floor coverings. CLF is used to calculate internal sensible cooling loads.
The limitations of the TFM are also carried through to the CLTD/SCL/CLF results. Furthermore,
the grouping of CLTD/SCL/CLF may cause additional errors.
TETD/TA Method
In the total eqivalent temperature difference (TETD)/time-averaging (TA) method, heat gains of a
number of representative exterior wall and roof assemblies qw, Btu /h (W), are calculated as
qw  AU(TETD) (6.9b)
where A  area of wall or roof, ft2 (m2)
U  overall heat-transfer coefficient of wall or roof, Btu /h ft2 F (W/m2 C)
In Eq. (6.9b), TETD, in F (C), can be evaluated by:
 Using the conduction transfer function as in TFM to determine qw. Then it is divided by the U
value to generate TETD values.
 Using the following relationship:
TETD  Tsol, a  Tr  (Tsol, 
 Tsol, a) (6.9c)
LOAD CALCULATIONS 6.15
where Tsol, a  daily average sol-air temperature, F (C)
Tsol, 
 sol-air temperature at time lag 
 h, F (C)
 effective decrement factor 
The internal heat gains and conductive heat gain are calculated in the same manner as in the TFM.
The radiant fraction of each of the sensible heat gains is then allocated to a period including the current
and successive hours, a total of 1 to 3 h for light construction and 6 to 8 h for heavy construction.
The TETD/ TA method is also a member of the TFM family and is developed primarily for
manual calculation. TETD/TA is simpler in the conversion of heat gains to cooling loads. However,
the time-averaging calculation procedure is subjectiveit is more an art than a rigorous scientific
method. Also the TETD/TA method inherits the limitations that a TFM possesses if the TFM is
used to calculate the TETD.
6.6 DETAILED CALCULATION PROCEDURES FOR TFM
Most of the widely adopted computer software programs for space load calculations are based on
the transfer function method, i.e., the weighting factors method. The following are the detailed
calculation procedures for TFM.
Conduction Heat Gain through Exterior Walls and Roofs
The sensible heat gain through an exterior wall or a roof using TFM can be calculated by the conduction
transfer function. It uses a sol-air temperature at time t, denoted by Tsol, t, to represent the combined
temperature excitation of the outdoor temperature and solar heat on the exterior surface of an exterior
wall or roof, and a constant indoor temperature Tr. Conduction-transfer function coefficients are calculated
based on an outdoor surface heat-transfer coefficient ho  3.0 Btu/h ft2 F (17 W/m2 C), and
an indoor surface heat-transfer coefficient hi  1.46 Btu/h ft2 F (8.3 W/m2 C).
The external heat gain through an exterior wall or roof at time t, denoted by qe,t, Btu /h (W), can
be calculated as
(6.10)
where t  time, h
  time interval, h
n  summation index of number of terms
Tsol,t  n  sol-air temperature at time tn, F (C)
qe,t  n  conduction heat gain at time tn, Btu /h (W)
bn,cn,dn  conduction-transfer function coefficients; refer to ASHRAE Handbook for details
A  interior surface area of wall or roof, ft2 (m2)
Harris and McQuiston (1988) provide the conduction-transfer function coefficients for 41 representive
wall assemblies and 42 roof assemblies with variations in components, insulation, and mass.
Heat Gain through Ceilings, Floors, and Interior Partition Walls
If the temperature of the adjacent space Tad, F (C), is constant, or the variation of Tad is small, the
sensible heat gain due to the heat transfer from the adjacent space through the ceiling, floor, or interior
partition wall at time t, denoted by qp,t, Btu /h (W), can be calculated as
qp,t  UA(Tad  Tr) (6.11)
qe,t   
n0
bnTsol,t  n  
n1
dn
qe,t  n
A
 Tr
n0
cn A
6.16 CHAPTER SIX
where U  overall heat-transfer coefficient of ceiling, floor, or partition wall, Btu/h ft2 F
(W/m2 C)
A  surface area of ceiling, floor, or partition wall, ft2 (m2)
If the temperature of the adjacent space Tad,t varies with time, then the sensible heat gain transferring
through the ceiling, floor, or partition wall at time qp, t, Btu /h (W), can be calculated from
Eq. (6.10), except qe,t should be replaced by qp,t, qe,t  n replaced by qp,t  n, and Tsol,t replaced by
Tad,t. Here qp,t  n indicates the sensible heat gain of the ceiling, floor, or partition wall at time
t  n, Btu /h (W); and Tad,t indicates the adjacent space temperature, F (C).
Solar Heat Gain and Conductive Heat Gain through Window Glass
Although the inward heat flow from the solar radiation absorbed by the window glass and the heat
flow due to the outdoor and indoor temperature difference are actually combined, it is simpler and
acceptably accurate to separate this composite heat gain into solar heat gain and conductive heat
gain; both are sensible heat gains.
Solar heat gain qso,t, Btu /h (W), is given as
qso,t  As, t (SC)(SHGFt)  Ash, t (SC)(SHGFsh, t) (6.12)
where As,t  sunlit area of window glass at time t, ft2 (m2)
Ash,t  shaded area of window glass at time t, ft2 (m2)
SC  shading coefficient
SHGFt  solar heat gain factor at time t considering orientation, latitude, month, and hour,
Btu/h ft2 (W/m2)
SHGFsh,t  solar heat gain factor for shaded area at time t, considering latitude, month, and hour,
Btu/h ft2 (W/m2)
Usually, the SHGFt of the vertical surface facing north orientation is taken as SHGFsh,t. Conductive
heat gain due to the outdoor-indoor temperature difference qwin,t, Btu /h (W), is given as
qwin,t  UwinAwin (To,t  Tr) (6.13)
where Uwin  overall heat-transfer coefficient of window including glass and frame (Table 3.7),
Btu/h ft2 F (W/m2 C)
Awin  gross area of window including glass and frame, ft2 (m2)
To,t  outdoor temperature at time t considering month, hour, and location, F (C)
Internal Heat Gain
The internal sensible heat gain consists of the sensible heat gain from people, from electric lights,
and from equipment and appliances.
People. Human beings release both sensible heat and latent heat to the conditioned space. The
radiative portion of the sensible heat gain is about 70 percent when the indoor environment of
the conditioned space is maintained within the comfort zone. The space sensible heat gain for
occupants staying in a conditioned space at time t, denoted by qsp,t, Btu /h (W), can be calculated as
qsp,t  Np,t (SHGp) (6.14)
where Np,t  number of occupants in conditioned space at time t
SHGp  sensible heat gain of each person, Btu/h (W)
LOAD CALCULATIONS 6.17
Space latent heat gain for occupants staying in a conditioned space at time t, denoted by qlp,t, Btu/h
(W), is given as
qlp,t  Np,t(LHGp) (6.15)
where LHGp  latent heat gain of each person, Btu/h (W). Table 6.1 lists the heat gain from occupants
in conditioned space, as abridged from ASHRAE Handbook 1989, Fundamentals. In Table
6.1, total heat is the sum of sensible and latent heat. The adjusted heat is based on a normally
distributed percentage of men, women, and children among the occupants.
Lighting. The sensible heat gains from the electric lights depend on the types of installation, as
follows:
Inside Conditioned Space. For electric lights installed inside the conditioned space, such as
light fixtures hung below the ceiling, the sensible heat gain released from the electric lights, the
emitting element, and light fixtures qs, l is equal to the sensible heat released to the conditioned
space qes, l, Btu /h (W); both depend mainly on the criteria of illumination and the type and effi-
ciency of electric lights and can be calculated as
qs.l  3.413 Wlamp FuslFal
 3.413 WAAfl (6.16)
where Wlamp  rated input of electric lights,W
WA  wattage per ft2 of floor area,W/ ft2 (W/m2)
Fusl  ratio of wattage in use to installation wattage
In Eq. (6.16), Fal indicates an allowance factor for light fixtures, such as Ballast losses. For rapidstart
40-W fluorescent fixtures, Fal varies from 1.18 to 1.3 with a recommended value of 1.2
(ASHRAE Handbook 1993, Fundamentals).
Recess-Mounted Fixtures Using Return Plenum. For situations in which electric lights are
recess-mounted on the ceiling and the ceiling plenum is used as a return plenum, the fraction of the
6.18 CHAPTER SIX
TABLE 6.1 Rates of Heat Gain from Occupants of Conditioned Spaces*
Total heat Total heat Sensible
of adults, adjusted, heat, Latent heat,
Degree of activity Typical application male, Btu /h Btu/h Btu /h Btu/h
Seated at theater Theatermatinee 390 330 225 105
Seated at theater Theaterevening 390 350 245 105
Seated, very light work Offices, hotels, apartments 450 400 245 155
Moderately active office work Offices, hotels, apartments 475 450 250 200
Standing, light work; walking Department store, retail store 550 450 250 200
Walking; standing Drugstore, bank 550 500 250 250
Light bench work Factory 800 750 275 475
Moderate dancing Dance hall 900 850 305 545
Walking 3 m/h; light machine work Factory 1000 1000 375 625
Heavy work Factory 1500 1450 580 870
Heavy machine work; lifting Factory 1600 1600 635 965
Athletics Gymnasium 2000 1800 710 1090
*Tabulated values are based on 75F room dry-bulb temperature. For 80F room dry-bulb temperature, the total heat remains the same, but the
sensible heat values should be decreased by approximately 20 percent and the latent heat values increased accordingly. All values are rounded to
nearest 5 Btu/ h.
Adjusted heat gain is based on normal percentage of men, women, and children for the application listed, with the postulate that the gain from
an adult female is 85 percent of that for an adult male, and that the gain from a child is 75 percent of that for an adult male.
Sources: Adapted with permission from ASHRAE Handbook 1989, Fundamentals.
sensible heat gain of electric lights that enters the conditioned space qes,l, Btu /h (W), is closely
related to the type of lighting fixture, the ceiling plenum, and the return system.
If the ceiling plenum is used as a return plenum in a multistory building, as shown in Fig. 6.6,
the heat released from recessed fluorescent light fixtures and the heat transfer between the outdoor
air, plenum air, return air, and space air are as follows:
 Radiative and convective heat transfer from the lighting fixture downward directly into the conditioned
space qld, Btu /h (W), which is calculated as
qld  qs,l  qlp  (1  Flp)qs,l (6.17)
where qlp  heat released by electric lights to return air, Btu /h (W)
Flp  fraction of heat released from light fixture to plenum air
The fraction that enters conditioned space qld depends on the volume flow rate of the return air
flowing through the lighting fixture and the type of fixture. Its magnitude should be obtained from
the lighting fixture manufacturer. In Fig. 6.7 is shown the relationship between Flp and the intensity
of volume flow rate of return air if the lighting fixture is ventilated. Usually, Flp varies
between 0.4 and 0.6 for a ventilated fixture in a return air plenum. For unventilated lighting
fixtures recess-mounted in a return air plenum, Flp varies between 0.15 and 0.5.
V?
r / Afl
LOAD CALCULATIONS 6.19
FIGURE 6.6 Heat released from a recess-mounted ventilated lighting fixture.
 Heat carried away by return air from the ceiling plenum qret, Btu /h (W), which is calculated as
(6.18)
where volume flow rate of return air, cfm [m3 / (60 s)]
r  density of return air, lb/ ft3 (kg/m3)
cpa  specific heat of return air, Btu / lbF (J /kgC)
Tp, Tr  temperatures of plenum air and space air, F (C)
 Heat transfer from the plenum air to the conditioned space through the suspended ceiling qcl and
heat transfer from the plenum air to the conditioned space through the composite floor qfl, both in
Btu/h (W), can be calculated as
qcl  UclAcl (Tp  Tr) (6.19a)
qfl  UflAfl (Tp  Tr) (6.19b)
where Ucl, Ufl  overall heat-transfer coefficient of ceiling and composite floor, Btu /h ft2 F
(W/m2 C)
Acl, Afl  area of ceiling and composite floor, ft2 (m2)
 Heat transfer between the outdoor air and the plenum air through the exterior wall of the ceiling
plenum qwp,t, Btu /h (W), can be calculated by conductive transfer function as shown in Eq. (6.10).
Because qwp,t is comparatively small, it is often ignored.
 For a multistory building where all the floors are air conditioned, heat gain from the electric lights
that enters the conditioned space, heat to space, qes,l is given as
qes,l  qld  qcl  qfl (6.20)
Of this, about 50 percent is radiative, and the rest is convective. Heat released from the electric
lights to the return air including radiative transfer upward, heat to plenum, qlp, Btu /h (W), is
V?
r 
qret  60 V?r r cpa (Tp  Tr)
6.20 CHAPTER SIX
FIGURE 6.7 Relationship between Flp and the intensity of return air
/ Afl. V?r
given as
qlp  qret  qcl  qfl (6.21)
In a return air plenum, precise calculation of the plenum air temperature Tp is rather complicated.
A simplified method is to assume a steady-state heat transfer between the plenum air and the conditioned
space air and to use a cooling load temperature difference CLTDwp (CLTDwp can be
found from Table 6.2) to calculate the heat transfer through the exterior wall of the plenum. Then,
based on the heat balance at the plenum air,
(6.22)
where Uwp  overall heat-transfer coefficient of exterior wall, Btu/h ft2 F (W/m2 C)
Awp  area of exterior wall, ft2 (m2)
CLTDwp  cooling load temperature difference of exterior wall of ceiling plenum, F (C)
Plenum air temperature Tp can thus be determined.
Surface-Mounted Fixtures under Ceiling. If the lighting fixtures are surface-mounted under the
suspended ceiling, then the fraction of heat gain downward that enters the conditioned space qes,l,
Btu/h (W), can be calculated as
qes,l  Fes,l qs,l (6.23 )
where Fes,l  fraction of sensible heat gain entering the conditioned space for surface-mounted
lighting fixtures, which varies from approximately 0.8 to 0.95.
Equipment and Appliances. With equipment and appliances, all energy inputs are converted to
heat energy, e.g., in the motor windings, combustion chamber, rubbing surfaces, even at components
where mechanical work is performed. A portion of heat released may be exhausted locally by
a mechanical ventilation system.
In many industrial applications, the space sensible heat gain due to the machine load when a
motor is located inside the conditioned space qs,e, Btu /h (W), can be calculated as
(6.24)
where Php  rated horsepower of machine, hp
Fload  load factor indicating ratio of actual power required to rated power
Fuse  use factor indicating ratio of actually used equipment and appliance to total installed
Fexh  heat removal factor due to mechanical exhaust system
mo  motor efficiency
In Table 6.3 are listed the efficiencies of motors from a fraction of a horsepower to 250 hp
(187 kW). The bigger the motor, the higher the motor efficiency. A high-efficiency motor is often
cost-effective. The Energy Policy Act of 1992 (EPACT), after a phase-in period of 5 years which
ended on October 24 1997, requires that all covered motors meet increased minimum efficiency
levels. If the motor is located outside the conditioned space, then in Eq. (6.24), mo  1.
In many types of equipment and appliances installed with exhaust hoods in the conditioned
space, energy input and Fexh are better determined according to the actual performance of similar
projects. For example, for food preparation appliances equipped with exhaust hoods, only radiation
from the surface of the appliances, which as a fraction of the total heat input is between 0.1
and 0.5, should be counted; or Fexh is between 0.9 and 0.5. If there is no mechanical exhaust
system, Fexh  0.
qs,e  2546Php Fload Fuse
1  Fexh
mo
qlp  Uwp Awp (CLTDwp)  (60 V?r rcpa  Ucl Acl  Ufl Afl) (Tp  Tr)
LOAD CALCULATIONS 6.21
6.22
TABLE 6.2 CLTD for Calculating Sensible Cooling Loads from Sunlit Walls of North Latitude, F
Solar time, h
Facing 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20
Group C walls: typical, outside 1-in. stucco, 2-in. insulation (5.7lb/ft3), 4-in. concrete, 0.75-in. plaster or gypsum, inside U  0.119 Btu/hft2F; mass, 63 lb/ft2
N 15 14 13 12 11 10 9 8 8 7 7 8 8 9 10 12 13 14 15 16
NE 19 17 16 14 13 11 10 10 11 13 15 17 19 20 21 22 22 23 23 23
E 22 21 19 17 15 14 12 12 14 16 19 22 25 27 29 29 30 30 30 29
SE 22 21 19 17 15 14 12 12 12 13 16 19 22 24 26 28 29 29 29 29
S 21 19 18 16 15 13 12 10 9 9 9 10 11 14 17 20 22 24 25 26
SW 29 17 25 22 20 18 16 15 13 12 11 11 11 13 15 18 22 26 29 32
W 31 29 27 25 22 20 18 16 14 13 12 12 12 13 14 16 20 24 29 32
NW 25 23 21 20 18 16 14 13 11 10 10 10 10 11 12 13 15 18 22 25
Group D walls: typical, outside 1-in. stucco, 4-in. concrete, 1- or 2-in. insulation (2 lb/ft3), 0.75-in. plaster or gypsum, inside U  0.1190.20 Btu/hft2F; mass, 63 lb/ft2
N 15 13 12 10 9 7 6 6 6 6 6 7 8 10 12 13 15 17 18 19
NE 17 15 13 11 10 8 7 8 10 14 17 20 22 23 23 24 24 25 25 24
E 19 17 15 13 11 9 8 9 12 17 22 27 30 32 33 33 32 32 31 30
SE 20 17 15 13 11 10 8 8 10 13 17 22 26 29 31 32 32 32 31 30
S 19 17 15 13 11 9 8 7 6 6 7 9 12 16 20 24 27 29 29 29
SW 28 25 22 19 16 14 12 10 9 8 8 8 10 12 16 21 27 32 36 38
W 31 27 24 21 18 15 13 11 10 9 9 9 10 11 14 18 24 30 36 40
NW 25 22 19 17 14 12 10 9 8 7 7 8 9 10 12 14 18 22 27 31
Group G walls: typical, outside 1-in. stucco, airspace; 1-, 2-, or 3-in. insulation (2 lb/ft3); 0.75-in. plaster or gypsum, inside U  0.0810.78 Btu/hft2F; mass, 16 lb/ft2
N 3 2 1 0 1 2 7 8 9 12 15 18 21 23 24 24 25 26 22 15
NE 3 2 1 0 1 9 27 36 39 35 30 26 26 27 27 26 25 22 18 14
E 4 2 1 0 1 11 31 47 54 55 50 40 33 31 30 29 27 24 19 15
SE 4 2 1 0 1 5 18 32 42 49 51 48 42 36 32 30 27 24 19 15
S 4 2 1 0 1 0 1 5 12 22 31 39 45 46 43 37 31 25 20 15
SW 5 4 3 1 0 0 2 5 8 12 16 26 38 50 59 63 61 52 37 24
W 6 5 3 2 1 1 2 5 8 11 15 19 27 41 56 67 72 67 48 29
NW 5 3 2 1 0 0 2 5 8 11 15 18 21 27 37 47 55 55 41 25
Direct applications and adjustments are stated in the text.
Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.
Hours of
maxi- Mini- Maxi- Difference
mum mum mum in
21 22 23 24 CLTD CLTD CLTD CLTD
17 17 17 16 22 7 17 10
23 22 21 20 20 10 23 13
28 27 26 24 18 12 30 18
28 27 26 24 19 12 29 17
25 25 24 22 20 9 26 17
33 33 32 31 22 11 33 22
35 35 35 33 22 12 35 23
27 27 27 26 22 10 27 17
19 19 18 16 21 6 19 13
23 22 20 18 19 7 25 18
28 26 24 22 16 8 33 25
28 26 24 22 17 8 32 24
27 26 24 22 19 6 29 23
38 37 34 31 21 8 38 30
41 40 38 34 21 9 41 32
32 32 30 27 22 7 32 25
11 9 7 5 18 1 26 27
11 9 7 5 9 1 39 40
12 10 8 6 10 1 55 56
12 10 8 6 11 1 51 52
12 10 8 5 14 1 46 47
17 13 10 8 16 0 63 63
20 15 11 8 17 1 72 71
17 13 10 7 18 0 55 55
Because of the widespread installation of microcomputers, display terminals, printers,
copiers, calculators, and facsimile machines, the heat released from machines in office buildings
has increased considerably in recent years. Komor (1997) listed the measured office plug loads
from 17 U.S. office buildings in the 1990s which varied from 0.44 to 1.11 W/ ft2. Precise office
equipment heat gain can be calculated from manufacturers data using a load factor Fload between
0.3 and 0.5.
The latent heat gain from the equipment and appliances ql,e, Btu /h (W), can be calculated from
the mass flow rate of water vapor evaporated , as
(6.25) ql,e  1075 m? w
m? w, lb / h (kg / h)
LOAD CALCULATIONS 6.23
TABLE 6.3 Heat Gain from Typical Electric Motors
Location of motor and driven equipment with
respect to conditioned space or airstream
Motor Full-load Motor in, Motor out, Motor in,
nameplate motor driven driven driven
or rated Nominal effciency, equipment in, equipment in, equipment out,
horsepower rpm percent Btu/h Btu /h Btu/h
Motor type: shaded pole
0.05 1,500 35 360 130 240
0.125 1,500 35 900 320 590
Motor type: split-phase
0.25 1,750 54 1,180 640 540
0.33 1,750 56 1,500 840 660
0.50 1,750 60 2,120 1,270 850
Motor type: three-phase
0.75 1,750 72 2,650 1,900 740
1 1,750 75 3,390 2,550 850
1 1,750 77 4,960 3,820 1,140
2 1,750 79 6,440 5,090 1,350
3 1,750 81 9,430 7,640 1,790
5 1,750 82 15,500 12,700 2,790
7.5 1,750 84 22,700 19,100 3,640
10 1,750 85 29,900 24,500 4,490
15 1,750 86 44,400 38,200 6,210
20 1,750 87 58,500 50,900 7,610
25 1,750 88 72,300 63,600 8,680
30 1,750 89 85,700 76,300 9,440
40 1,750 89 114,000 102,000 12,600
50 1,750 89 143,000 127,000 15,700
60 1,750 89 172,000 153,000 18,900
75 1,750 90 212,000 191,000 21,200
100 1,750 90 283,000 255,000 28,300
150 1,750 91 420,000 382,000 37,800
200 1,750 91 569,000 509,000 50,300
250 1,750 91 699,000 636,000 62,900
For motors operating more than 750 h/year, it is usually cost-effective to use a high-efficiency motor. Typical efficiency ratings
are as follows: 5 hp, 89.5%; 10 hp, 91.0%; 50 hp, 94.1%; 100 hp, 95.1%; 200 hp, 96.2%.
Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.
Infiltration
Infiltration is the uncontrolled inward flow of outdoor air through cracks and openings in the building
envelope due to the pressure difference across the envelope. The pressure difference may be
caused by any of the following:
1. Wind pressure
2. Stack effect due to the outdoor and indoor temperature difference
3. Mechanical ventilation
In summer, for low-rise commercial buildings that have their exterior windows well sealed, and if a
positive pressure is maintained in the conditioned space when the air system is operating, normally
the infiltration can be considered zero.
For high-rise buildings, infiltration should be considered and calculated in both summer and
winter. Infiltration is discussed again in Sec. 20.4.
As soon as the volume flow rate of infiltrated air , cfm (m3 /min), is determined, the space
sensible heat gain from infiltration qinf, Btu /h (W), can be calculated as
(6.26)
where o  density of outdoor air, lb/ ft3 (kg/m3). The space latent heat gain from infiltration ql, inf,
Btu/h (W), can be calculated as
(6.27)
where wo,wr  humidity ratio of outdoor and space air, respectively, lb/ lb (kg /kg)
hfg, 32  latent heat of vaporization at 32F, Btu/ lb (J /kg)
Cooling Load Conversion Using Room Transfer Function
The conversion of space sensible heat gains qrs,t, Btu /h (W), having radiative only or radiative and
convective components to space sensible cooling loads Qrs,t, Btu /h (W), using the transfer function
method and room transfer function coefficients can be expressed as follows:
(6.28)
where i  number of heat gain components in same group
  time interval
tn  time at tn
Here v0, v1, v2, . . ., w1, w2, . . . are the coefficients of the room transfer function; refer to the
ASHRAE Handbook for details. Their relationship can be expressed from Eqs. (6.8) and (6.9) as
(6.29)
The magnitude of the coefficients depends on the duration of the time interval, fraction of
the radiative component, and heat storage capacity because of the 14 influential parameters
of zone characteristics, such as zone geometry, height, exterior wall construction, interior shade,
furniture, zone location, glass percentage, and type of partition, midfloor, slab, ceiling, roof,
and floor.
K(z) 
v0  v1z1  v2z2    
1  w1z1  w2z2    
Qrs,t  
i1
(v0 qs,t  v1qs,t    v2qs,t  2    )  (w1 Qrs,t    w2 Qrs,t  2    )
ql,inf  60V?inf o (wo  wr)hfg, 32
qs, inf  60V?inf ocpa (To  Tr)
V?
inf
6.24 CHAPTER SIX
Space Cooling Load Calculation
The types of conversion of space heat gains to space cooling loads can be grouped into the following
two categories:
1. Space sensible cooling loads Qrs,t that are only a fraction of the space sensible heat gain qrs,t.
These kinds of sensible heat gains have both radiative and convective components, and it is diffi-
cult to separate the convective component from the radiative component, such as sensible heat
gains from exterior walls and roofs, and solar heat gains through windows. Equation (6.28) will
be used to convert these types of sensible heat gains qrs,t to cooling loads Qrs,t.
2. Space heat gains qin,t are instantaneous space cooling loads Qin,t, both in Btu/h (W), or
Qin, t  qin, t (6.30)
This category includes all latent heat gains qrl, convective sensible heat gains, infiltration sensible
heat gain qinf, s, and sensible heat gains whose convective component can be separated from
the radiative component, such as air-to-air heat gain from windows, lights, and people.
The space cooling load Qrc, t, Btu /h (W), is their sum, or
Qrc,t  Qrs,t  Qin, t  Ql,t  Qs,t  Ql,t (6.31)
where Qs,t, Ql,t  space sensible cooling load and latent load, respectively, Btu /h (W).
As mandated in ASHRAE/IESNA Standard 90.1-1999, Optimum Start Controls, pickup loads
either for cooling or heating depend on the difference between space temperature and occupied set
point and the amount of time prior to scheduled occupancy, and is often determined by computer
software.
Heat Extraction Rate and Space Air Transfer Function
In the calculation of the conduction heat gain through the exterior wall and roofs by using the
conduction transfer function [Eq. (6.10)] and then converting to sensible cooling load from the
room transfer function, the space temperature Tr is found to be a constant. Most direct-digital control
(DDC) zone control systems are now adopting proportional-integral control mode. When the
air system and the DDC system are in operation, Tr is a constant once it has achieved a stable
condition.
However, for many air systems operated at nighttime shutdown mode, Tr will drift away from the
set point during the shutdown period. The space air transfer function relates the heat extraction rate at
time t, denoted by Qex, t, Btu/h (W), to the space air temperature at time t, denoted by Tr, t, F (C ), as
(6.32)
where pi, gi  space air transfer function coefficients, refer to ASHRAE Handbook for details
Qex,t  heat extraction rate at time t, Btu /h (W)
Qrc,ti  calculated space cooling load at time ti, Btu /h (W)
Tr, con  assumed constant space air temperature, F (C)
Heat Loss to Surroundings
There is always a radiant heat loss from the outer surface of the building to the sky vault without
clouds because atmospheric temperature is lower at high altitudes, as described in Sec. 3.12. In
many locations in the United States, there are also radiant heat losses to the surroundings at
1
i0
pi (Qex, t    Qrc,t  i)  2
i0
gi (Tr,con  Tr, t  i)
LOAD CALCULATIONS 6.25
nighttime in summer due to the lower ground temperature. ASHRAE Handbook 1993, Fundamentals,
recommends a simplified calculation procedure so that the sum of sensible cooling loads from
heat gain by conduction through exterior walls and roofs, from conduction and solar heat gain
through windows, and from heat gain through interior partitions, ceilings, and floors at time t, denoted
by Qex, t, Btu /h (W), plus the sensible cooling loads from the radiant component of internal
heat gain at time t, denoted by Qin, t, Btu /h (W), will be multiplied by a factor Fsur to take into
account these heat losses. The corrected sensible cooling load at time t, denoted by Qrs, cor, t,
Btu/h (W), is
Qrs, cor, t  Fsur (Qex, t  Qin, t) (6.33)
The factor of heat loss to surroundings Fsur can be calculated as
(6.34)
where Lex  length of space exterior wall, ft
U  overall heat-transfer coefficient (subscript win for window and part for interior
partitions), Btu/h ft2 F (W/m2 C)
A  area of component of building envelope, ft2 (m2)
6.7 DETAILED CALCULATION PROCEDURE USING
CLTD/SCL/CLF METHOD
The following sections describe in greater detail the principles of the CLTD/SCL/CLF method.
They also provide a simple, manual cooling load calculation procedure in case an estimate or a
rough check of computer-aided cooling load calculation is required.
Space Cooling Load due to Heat Gain through Exterior Walls and Roofs
and Conductive Gain through Glass
If the ratio of sensible cooling load to sensible heat gain through the exterior wall or roof
Qrs,w /qrs,w  CLTD/T, for an exterior sunlit wall or roof under the combined effect of solar radiation
and the outdoor-indoor temperature difference, the one-step calculation of space sensible cooling
load Qrs,w, Btu /h (W), can be performed as
Qrs,w  UA (CLTD) (6.35)
where U  overall heat-transfer coefficient of exterior wall or roof, Btu /h ft2 F (W/m2 C)
A  area of exterior wall, roof, or window including frame or sash, ft2 (m2)
CLTD  cooling load temperature difference, F (C)
The CLTD values recommended by ASHRAE for calculating the space sensible cooling load through
flat roofs and sunlit walls of various constructions are listed in Tables 6.2 and 6.4, respectively. The
values in both Table 6.2 and Table 6.4 were calculated under the following conditions. In other words,
these are the conditions under which the listed data can be applied directly without adjustments:
 Indoor temperature of 78F (25.6C)
 Outdoor maximum temperature of 95F (35C) with an outdoor daily mean of 85F (29.4C) and
an outdoor daily range of 21F (11.7C)
 Solar radiation of 40 north latitude on July 21
 Roof with dark, flat surface
Fsur  1  0.02  1
Lex
 (Uroof Aroof  Uwall Awall  Uwin Awin  Upart Apart)
6.26 CHAPTER SIX
6.27
TABLE 6.4 CLTD for Calculating Sensible Cooling Loads from Flat Roofs, F
Solar time, h
Description of Weight, U value,
construction lb/ft2 Btu/hft2F 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Without suspended ceiling
2.5-in. wood with 13 0.093 30 26 23 19 16 13 10 9 8 9 13 17 23 29 36 41 46
2-in. insulation
4-in. wood with 18 0.078 38 36 33 30 28 25 22 20 18 17 16 17 18 21 24 28 32
2-in. insulation
With suspended ceiling
1-in. wood with 10 0.083 25 20 16 13 10 7 5 5 7 12 18 25 33 41 48 53 57
2-in. insulation
2.5-in. wood with 15 0.096 34 31 29 26 23 21 18 16 15 15 16 18 21 25 30 34 38
1-in. insulation
8-in. lightweight 33 0.093 39 36 33 29 26 23 20 18 15 14 14 15 17 20 25 29 34
concrete
4-in. heavyweight 54 0.090 30 29 27 26 24 22 21 20 20 21 22 24 27 29 32 34 36
concrete with
2-in. insulation
2.5-in. wood with 15 0.072 35 33 30 28 26 24 22 20 18 18 18 20 22 25 28 32 35
2-in. insulation
Roof terrace 77 0.082 30 29 28 27 26 25 24 23 22 22 22 23 23 25 26 28 29
system
6-in. heavyweight 77 0.088 29 28 27 26 25 24 23 22 21 21 22 23 25 26 28 30 32
concrete with
2-in. insulation
4-in. wood with 19 0.082 35 34 33 32 31 29 27 26 24 23 22 21 22 22 24 25 27
2-in. insulation 20 0.064
Conditions of direct application and adjustments are stated in the text.
Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.
Hours of
maximum Minimum Maximum Difference
18 19 20 21 22 23 24 CLTD CLTD CLTD in CLTD
49 51 50 47 43 39 35 19 8 51 43
36 39 41 43 43 42 40 22 16 43 27
57 56 52 46 40 34 29 18 5 57 52
41 43 44 44 42 40 37 21 15 44 29
38 42 45 46 45 44 42 21 14 46 32
38 38 38 37 36 34 33 19 20 38 18
38 40 41 41 40 39 37 21 18 41 23
31 32 33 33 33 33 32 22 33 22 11
33 34 34 34 33 32 31 20 21 34 13
30 32 34 35 36 37 36 23 21 37 16
 Outer surface R value Ro  0.333 h ft2 F/Btu (0.06 m2 C/W) and inner surface Ri  0.685
h ft2 F/Btu (0.123 m2 C/W)
 No attic fans or return air ducts in suspended ceiling space
The following formula can be used for adjustments when the conditions are different from those
mentioned:
CLTDcorr  CLTD  78  Tr  Tom  85 (6.36)
where 78  Tr  indoor temperature correction; Tr is indoor temperature, F (C)
Tom  85  outdoor temperature correction; Tom is outdoor mean daily temperature, F (C)
In Table 6.4, the roof terrace system includes the following:
 2-in. (50-mm) lightweight concrete
 Airspace
 2-in. (50-mm) insulation [5.7 lb / ft3 (91.2 kg/m3)]
 0.5-in. (13-mm) slag or stone
 0.375-in. (9.5-mm) felt and membrane
For a pitched roof with a suspended ceiling, the area A in Eq. (6.35) should be the area of the suspended
ceiling. If a pitched roof has no suspended ceiling under it, then the actual CLTD is slightly
higher than the value listed in Table 6.4 because a greater area is exposed to the outdoor air.
Space Cooling Load due to Solar Heat Gain through Fenestration
The space sensible cooling load from solar heat gain transmitted through the window facing a specific
direction Qrs, s, Btu /h (W), can be calculated as follows:
Qrs, s  Qsun  Qsh
 As 	 SCLs 	 SC  Ash 	 SCLsh 	 SC (6.37)
where Qsun  space cooling load from solar heat gain through sunlit area of window glass, Btu/h (W)
Qsh  space cooling load from solar heat gain through shaded area of window glass, Btu/h (W)
As, Ash  sunlit and shaded area, ft2 (m2)
SC  shading coefficient
SCLs  solar cooling load for sunlit glass facing specific direction, Btu/h ft2 (W/m2)
SCLsh  solar cooling load for shaded area as if glass is facing north, Btu/h ft2 (W/m2)
Zone types for use with SCLs and SCLsh tables, single-story building, are listed in Table 6.5.
July solar cooling loads for sunlit glass 40 north latitude are listed in Table 6.6. Refer to ASHRAE
Handbook 1993, Fundamentals, for zone types for multistory buildings and other details.
In Eq. (6.37), at a given time, As  Ash  Aglass. Here Aglass indicates the glass area of the window,
in ft2 (m2). In the northern hemisphere for a conditioned space with southern orientation, the
maximum SCLs may occur in December instead of June.
Space Cooling Load due to Heat Gain through Wall Exposed to
Unconditioned Space
When a conditioned space is adjacent to an area that is unconditioned, and if the temperature fluctuation
in this area is ignored, then the sensible heat gain qrs transferred through the partitioned walls
and interior windows and doors, in Btu/h (W), can be calculated as
6.28 CHAPTER SIX
qrs  AU(Tun  Tr) (6.38)
where Tun,Tr  daily mean air temperature of adjacent area that is unconditioned and space temperature,
respectively, F (C). For an adjacent area that is not air conditioned and has heat sources
inside, such as a kitchen or boiler room, Tun may be 15F (8.3C) higher than the outdoor temperature.
For an adjacent area without any heat source other than electric lights, Tun  Tr may be
between 3 and 8F (1.7 and 4.4C). For floors built directly on the ground or located above a basement
that is neither ventilated nor conditioned, the space sensible cooling load from the heat gain
through the floor can often be ignored.
Calculation of Internal Cooling Loads and Infiltration
The calculation of the internal heat gains of people, lights, equipment, and appliances qint, s and heat
gains of infiltration qinf, all in Btu /h (W), using the CLTD/SCL/CLF method is the same as that
which uses TFM. The sensible internal heat gain that contains the radiative component is then multiplied
by a cooling load factor CLFint to convert to space sensible cooling load Qint, s, Btu /h (W),
and can be calculated as
Qint, s  qint, s (CLFint) (6.39)
In conditioned spaces in which an air system is operated at nighttime shutdown mode, CLFint is
equal to 1 during the occupied period when the air system is operating. Refer to ASHRAE Handbook
1993, Fundamentals, for CLFint when the air system is operated 24 h continuously.
Since internal latent heat gains qint, l are instantaneous internal latent cooling loads Qint, l both in
Btu/h (W),
Qint,l  qint,l (6.40)
LOAD CALCULATIONS 6.29
TABLE 6.5 Zone Types for Use with SCL and CLF Tables, Single-Story Building
Zone parameters Zone type Error band
No. Floor Partition Inside Glass People and
walls covering type shade solar equipment Lights Plus Minus
1 or 2 Carpet Gypsum * A B B 9 2
1 or 2 Carpet Concrete block * B C C 9 0
1 or 2 Vinyl Gypsum Full B C C 9 0
1 or 2 Vinyl Gypsum Half to none C C C 16 0
1 or 2 Vinyl Concrete block Full C D D 8 0
1 or 2 Vinyl Concrete block Half to none D D D 10 6
3 Carpet Gypsum * A B B 9 2
3 Carpet Concrete block Full A B B 9 2
3 Carpet Concrete block Half to none B B B 9 0
3 Vinyl Gypsum Full B C C 9 0
3 Vinyl Gypsum Half to none C C C 16 0
3 Vinyl Concrete block Full B C C 9 0
3 Vinyl Concrete block Half to none C C C 16 0
4 Carpet Gypsum * A B B 6 3
4 Vinyl Gypsum Full B C C 11 6
4 Vinyl Gypsum Half to none C C C 19 1
A total of 14 zone parameters are defined. Those not shown in this were selected to achieve the minimum error band shown in
the right-hand column for solar cooling load.
*The effect of inside shade is negligible in this case.
Source: Adapted from ASHRAE Handbook 1997, Fundamentals. Reprinted with permission.
6.30
TABLE 6.6 July Solar Cooling Load for Sunlit Glass, 40 North Latitude
Glass Solar time, h
face 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15
Zone type A
N 0 0 0 0 1 25 27 28 32 35 38 40 40 39 36
NE 0 0 0 0 2 85 129 134 112 75 55 48 44 40 37
E 0 0 0 0 2 93 157 185 183 154 106 67 53 45 39
SE 0 0 0 0 1 47 95 131 150 150 131 97 63 49 41
S 0 0 0 0 0 9 17 25 41 64 85 97 96 84 63
SW 0 0 0 0 0 9 17 24 30 35 39 64 101 133 151
W 1 0 0 0 0 9 17 24 30 35 38 40 65 114 158
NW 1 0 0 0 0 9 17 24 30 35 38 40 40 50 84
Horiz. 0 0 0 0 0 24 69 120 169 211 241 257 259 245 217
Zone type B
N 2 2 1 1 1 22 23 24 28 32 35 37 38 37 35
NE 2 1 1 1 2 73 109 116 101 73 58 52 48 45 41
E 2 2 1 1 2 80 133 159 162 143 105 74 63 55 48
SE 2 2 1 1 1 40 81 112 131 134 122 96 69 58 49
S 2 2 1 1 1 8 15 21 36 56 74 86 87 79 63
SW 6 5 4 3 2 9 16 22 27 31 36 58 89 117 135
W 8 6 5 4 3 9 16 22 27 31 35 37 59 101 139
NW 6 5 4 3 2 9 16 22 27 31 34 37 37 46 76
Horiz. 8 6 5 4 3 22 60 104 147 185 214 233 239 232 212
Zone type C
N 5 5 4 4 4 24 23 24 27 30 33 34 35 34 32
NE 7 6 6 5 6 75 106 107 88 61 49 47 45 43 40
E 9 8 8 7 8 83 130 148 145 124 89 62 56 52 47
SE 9 8 7 6 6 45 82 107 121 121 107 82 59 51 47
S 7 7 6 5 5 12 18 23 36 54 70 79 79 70 54
SW 14 12 11 10 9 15 21 26 29 33 36 57 86 110 124
W 17 15 13 12 11 17 22 27 31 34 36 37 59 98 132
NW 12 11 10 9 8 14 20 25 29 32 34 36 36 44 73
Horiz. 24 21 19 17 16 34 68 107 144 175 199 212 215 207 189
Zone type D
N 8 7 6 6 6 21 21 21 24 27 29 31 32 31 30
NE 11 10 9 8 9 63 87 90 77 58 49 48 46 44 42
E 15 13 12 11 11 70 107 123 124 110 85 65 60 57 53
SE 14 13 11 10 10 39 68 90 102 104 95 78 60 55 51
S 11 10 9 8 7 12 17 21 32 46 59 67 69 63 52
SW 21 19 17 15 14 18 22 25 28 31 34 51 74 94 106
W 25 23 20 18 17 21 24 28 30 33 34 35 53 84 112
NW 18 16 15 13 12 17 21 24 27 30 32 33 34 41 64
Horiz. 37 33 30 27 24 38 64 95 124 150 171 185 191 188 176
Notes:
1. Values are in Btu /hft2.
2. Apply data directly to standard double-strength glass with no inside shade.
3. Data apply to 21st day of July.
4. For other types of glass and internal shade, use shading coefficients as multiplier. For externally shaded glass, use north orientation.
Source: ASHRAE Handbook 1997, Fundamentals. Reprinted with permission.
16 17 18 19 20 21 22 23 24
31 31 36 12 6 3 1 1 0
32 26 18 7 3 2 1 0 0
33 26 18 7 3 2 1 0 0
34 27 18 7 3 2 1 0 0
42 31 20 8 4 2 1 0 0
152 133 93 35 17 8 4 2 1
187 192 156 57 27 13 6 3 2
121 143 130 46 22 11 5 3 1
176 125 70 29 14 7 3 2 1
32 31 35 16 10 7 5 4 3
36 30 23 13 9 6 5 3 3
41 34 25 15 10 7 5 4 3
42 35 26 15 10 8 6 4 3
46 37 27 16 11 8 6 4 3
138 126 94 46 31 21 15 11 8
166 173 147 66 43 30 21 15 11
108 128 119 51 33 22 16 11 8
180 137 90 53 37 27 19 14 11
29 29 34 14 10 8 7 6 6
36 31 25 16 13 11 10 9 8
43 37 30 20 17 15 13 12 11
42 36 29 19 16 14 13 11 10
40 33 26 16 13 12 10 9 8
125 111 80 37 28 23 20 17 15
153 156 128 50 35 28 24 21 19
102 118 107 39 26 21 17 15 13
160 123 83 53 44 38 34 30 27
28 29 32 17 14 12 11 10 9
39 35 29 22 19 17 15 14 12
48 43 37 29 25 22 20 18 16
47 42 35 27 24 21 19 17 16
41 36 30 22 19 17 15 14 12
109 100 78 45 37 33 29 26 23
130 135 116 57 46 39 35 31 28
87 101 94 42 33 29 25 22 20
156 128 96 72 63 56 50 45 41
Internal load density (ILD), W/ ft2 (W/m2), indicates the total internal heat gains of people, lights,
and equipment, and it can be calculated as
(6.41)
where SHGp, LHGp  sensible and latent heat gains for occupants, respectively, Btu /h (W)
Afl  floor area, ft2 (m2)
WA,l, WA,e  lighting and equipment power density, respectively,W/ ft2 (W/m2)
Both infiltration sensible heat gain qinf, s and infiltration latent heat gain qinf, l, in Btu/h (W), are instantaneous
space cooling loads; they also can be expressed as
Qinf, s  qinf, s Qinf, l  qinf, l (6.42)
where Qinf, s, Qinf, l  infiltration sensible and latent cooling loads, respectively, Btu /h (W).
Example 6.1. A return air plenum in a typical floor of a multistory building has the following
construction and operating characteristics:
ILD 
SHGp  LHGp
3.413 Afl
 WA,l  WA,e
LOAD CALCULATIONS 6.31
1.5 W/ ft2 (16.1 W/m2)
11,800 cfm (334 m3 /min)
0.073 lb/ ft3 (1.168 kg/m3)
0.5
0.2 Btu/h ft2 F (1.136 W/m2 C)
0.32 Btu/h ft2 F (1.817 W/m2 C)
0.21 Btu/h ft2 F (1.192 W/m2 C)
11,900 ft2 (1106 m2)
1920 ft2 (178 m2)
1.0
24F (13.3C)
75F (23.9C)
Wattage of electric lights
Return air volume flow rate
Density of return air r
Fraction of heat to plenum Flp
U values: exterior wall of plenum
suspended ceiling
floor
Area of ceiling and floor
Area of exterior wall of plenum
CLF of electric lights
CLTD of exterior wall of plenum
Space temperature
Determine the return air plenum temperature and the space cooling load from the electric lights
when they are recess-mounted on the ceiling and the ceiling plenum is used as a return plenum.
Solution. From Eq. (6.17),
qlp  Flpqs, l  0.5 	 1.5 	 11,900 	 3.413  30,461 Btu/h (W)
Since , and given Eq. (6.22), the
temperature of return air in the return plenum is
 75 
39,933
18,866
 77.12F (25.07C)
 75 
30,717  0.2 	 1920 	 24
12,559  0.32 	 11,900  0.21 	 11,900
Tp  Tr 
qlp  Uwp AwpCLTDwp
60V?r r cpa  Ucl Acl  UflAfl
60V?r r cpa  60 	 11,800 	 0.073 	 0.243  12,559 Btu / hF
From Eqs. (6.19a) and (6.19b), heat transfer from the plenum air to conditioned space through the
ceiling and floor is calculated as
Then, from Eq. (6.20), heat to space is calculated as
qes, l  qld  qcl  qfl
 30,461  7997  5248  43,706 Btu/h (W)
From Eq. (6.39), the space sensible cooling load from electric lights is
Qrs, l  CLFint qes, l  1.0 	 43,692  43,692 Btu/h (12,802 W)
Space Cooling Load of Night Shutdown Operating Mode
In commercial buildings, air systems are often operated in night shutdown mode during unoccupied
hours in summer. The accumulated stored heat because of the external heat gains increases the
space cooling load during cool-down and conditioned periods. On the other hand, heat losses to the
surroundings due to the radiant heat exchange between the outer surface of the building and the sky
vault and surroundings decrease the accumulated stored heat as well as the space cooling load, although
the radiative heat losses to the sky vault and surroundings partly compensate the stored heat
released to the space. However, overlooking the remaining stored heat released to the space during
cool-down and conditioned periods in summer is the limitation of the CLTD/SCL/CLF method
compared to TFM, especially when peak load occurs during the cool-down period. As in TFM, an
increase of up to 10 percent is recommended by ASHRAE/IES Standard 90.1-1989 for pickup load
during the cool-down period for air systems operated at nighttime shutdown mode.
6.8 COOLING COIL LOAD
Basics
Based on the principle of heat balance, the cooling coil load is given as
Total enthalpy of entering air  total enthalpy of leaving air
 cooling coil load (or cooling capacity)  heat energy of condensate
Since the heat energy of the condensate is small and can be ignored, the cooling coil load Qcc,
Btu/h (W), can be calculated as
(6.43)
where volume flow rate of supply air, cfm [m3 / (60 s)]
s  density of supply air, lb/ ft3 (kg/m3)
hae, hcc  enthalpy of entering air and conditioned air leaving coil, respectively, Btu / lb (J /kg)
Of this, the sensible cooling coil load Qcs, Btu /h (W), is
(6.44) Qcs  60V?s s cpa(Tae  Tcc)
V?
s 
Qcc  60V?s s (haehcc)
 0.21 	 11,900(77.1  75)  5248 Btu / h (W)
qfl  Ufl Afl (Tp  Tr)
 0.32 	 11,900(77.1  75)  7997 Btu / h (W)
qcl  Ucl Acl (Tp  Tr)
6.32 CHAPTER SIX
where Tae, Tcc  temperature of entering air and conditioned air leaving coil, respectively, F (C).
And the latent coil load Qcl, Btu /h (W), is
(6.45)
where wae, wcc  humidity ratio of entering air and conditioned air leaving coil, respectively, lb/ lb
(kg/ kg). Also,
Qcc  Qcs  Qcl (6.46)
From Fig. 6.4, alternatively, the sensible cooling coil load can be calculated as
Qcs  Qrs  qs, s  qr, s  Qo, s (6.47)
where qs, s, qr, s  supply and return system heat gain (as mentioned in preceding section, both are
instantaneous cooling loads), Btu/h (W)
Qo, s  sensible load from outdoor air intake, Btu/h (W)
And the latent coil load can be calculated as
Qcl  Qrl  Qo, l (6.48)
where Qo, l  latent load from outdoor air intake, Btu/h (W). The supply system heat gain consists
of mainly the supply fan power heat gain qsf and supply duct heat gain qsd; and the return system
heat gain comprises the return fan power heat gain qrf, return duct heat gain qrd, and ceiling plenum
heat gain qrp, all in Btu /h (W).
Fan Power
In the air duct system, the temperature increase from the heat released to the airstream because of
frictional and dynamic losses is nearly compensated by the expansion of air from the pressure drop
of the airstream. Therefore, it is usually assumed that there is no significant temperature increase
from frictional and dynamic losses when air flows through an air duct system.
Fan power input is almost entirely converted to heat energy within the fan. If the fan motor is
located in the supply or return airstream, the temperature increase across the supply (or return fan)
Tf, F (C), can be calculated as
(6.49)
where pt  fan total pressure, in. WC
f, m  total efficiency of fan and efficiency of motor
If the motor is located outside the airstream, then, in Eq. (6.49), m  1.
The pt of the return fan for a central hydronic air conditioning system in commercial buildings
is usually 0.25 to 0.5 of the pt of the supply fan. Therefore, the temperature increase of the return
fan is far smaller than that of the supply fan. The temperature increase of the relief fan or exhaust
fan affects only the relief or exhaust airstream. It is not a part of the supply and return system heat
gain. A relief fan is used to relieve excess space pressure when 100 percent outdoor air is flowing
through the supply fan for free cooling.
Duct Heat Gain
Duct heat gain is the heat transfer caused by the temperature difference between the ambient air and
the air flowing inside the air duct. Duct heat gain is affected by this temperature difference, the
Tf 
0.37pt
f m
Qcl  60V?s s (wae  wcc)hfg, 32
LOAD CALCULATIONS 6.33
thickness of the duct insulation layer, air volume flow rate, and the size and shape of the duct.
Detailed calculations are presented in Secs. 17.4 and 20.16. A rough estimate of the temperature increase
of the supply air for an insulated duct is as follows:
Supply air velocity Air temperature rise
 2000 fpm (10 m/ s) 1F/ 100 ft (0.6C/30 m) duct length
 2000 fpm (10 m/ s) 0.75F/ 100 ft (0.45C/30 m) duct length
Temperature of Plenum Air and Ventilation Load
For a ceiling plenum using a return plenum, the plenum air temperature can be calculated from
Eq. (6.22). The temperature increase of the plenum air, caused by the heat released from the electric
lights (Tp  Tr), is affected by their power input, type of lighting fixture, return air volume flow
rate, and construction of the ceiling plenum. The temperature increase of plenum air Tp  Tr is usually
between 1 and 3F (0.6 and 1.7C).
From Eqs. (6.26) and (6.27), the sensible and latent loads Qo, s and Qo, l, Btu /h (W), which are
attributable to the outdoor air intake, can be similarly calculated, except in Eqs. (6.26) and
(6.27) should be replaced by the volume flow rate of outdoor air , cfm (m3 /min).
System heat gains are mainly due to convective heat transfer from the surfaces. For simplification,
they are considered instantaneous cooling coil loads.
6.9 COOLING LOAD CALCULATION BY FINITE
DIFFERENCE METHOD
When both heat and moisture transfer from the surface of the walls, ceiling and carpet or floors
should be considered in the space cooling load calculation during the cool-down period in summer
in a location where the outdoor climate is hot and humid, the finite difference method might be the
best choice.
Finite Difference Method
Because of the rapid increase in the use of microcomputers in the HVAC&R calculations, it is now
possible to use a finite difference method, a numerical approach, to solve transient simultaneous
heat- and moisture-transfer problems in heating and cooling load calculations and energy
estimations. The finite difference method divides the building structures into a number of sections.
A fictitious node i is located at the center of each section or on the surface, as shown in Fig. 6.8. An
energy balance or a mass balance at each node at selected time intervals results in a set of algebraic
equations that can be employed to determine the temperature and moisture for each node in terms
of neighboring nodal temperatures or moisture contents, nodal geometry, and the thermal and moisture
properties of the building structure. The stored heat energy and moisture are expressed as an
increase of internal energy and moisture content at the nodes.
Heat conduction can be approximated by using the finite difference form of the Fourier law, as
(6.50)
where k  thermal conductivity, Btu / h ft F (W/mC)
Ai  area of building structure perpendicular to direction of heat flow, ft2 (m2)
qi1:i 
kAi(T ti
1  T t
1)
x
V?
o
V?
inf
6.34 CHAPTER SIX
LOAD CALCULATIONS 6.35
FIGURE 6.8 Building structures and nodes for a typical room.
T  temperature, F (C)
x  spacing between the nodal points, ft (m)
In Eq. (6.50), superscript t denotes at time t. Each nodal equation is solved explicitly in terms of the
future temperature of that node. The explicit method is simpler and clearer than the more complex
implicit method. The time derivative is then approximated by a forward finite difference in time, or
(6.51)
Compared with the transfer function method to calculate the cooling load, the finite difference
method has the following benefits:
1. It solves heat- and moisture-transfer load calculations simultaneously.
2. The concept and approach are easily understood.
3. It permits custom-made solutions for special problems.
4. It allows direct calculation of cooling loads and energy estimates.
Its drawbacks are mainly due to the great number of computerized calculations and comparatively
fewer computer programs and less information and experience are available.
Simplifying Assumptions
When the finite difference method is used to calculate space cooling loads, simplifications are often
required to reduce the number of computer calculations and to solve the problem more easily. The
errors due to simplification should be within acceptable limits.
For a typical room in the building, as shown in Fig. 6.8, the following are the simplifying
assumptions:
 Heat and moisture flow are one-dimensional.
 Thermal properties of the building materials are homogeneous.
 The properties of the airstream flowing over the surface of the building structures are homogeneous.
 The surface temperature differences between the partition walls, ceiling, and floors are small;
therefore, the radiative exchange between these surfaces can be ignored.
 The radiative energy received on the inner surface of the building structures can be estimated as
the product of the shape factor and the radiative portion of the heat gains, and the shape factor is
approximated by the ratio of the receiving area to the total zone area.
 During the operating period, the heat capacity of the space air is small compared with other heat
gains; therefore, it can be ignored.
 When the air system is not operating during the night shutdown period, the heat capacity of the
space air has a significant influence on the space air temperature; therefore, it should be taken into
account.
 Different heat- and mass-transfer coefficients and analyses are used for the operating period and
shutdown periods.
Heat and Moisture Transfer at Interior Nodes
Consider an interior node i as shown in the upper part of Fig. 6.8. For a one-dimensional heat flow,
if there is no internal energy generation, then according to the principle of heat balance,
Ti
t

T i
t t  T i
t
t
6.36 CHAPTER SIX
Conduction heat from node i  1  conduction heat from node i  1
 rate of change of internal energy of node i
(6.52)
where Ui  internal energy of node i, Btu/lb (J/kg)
b , cb  density and specific heat of building material, lb / f t 3 (kg/m3) and Btu/ lbF
(J/kgC)
t  selected time interval, s or min
Substituting Eq. (6.50) into (6.52) and solving for Ti
tt, we have
(6.53)
In Eq. (6.53), Fo is the Fourier number and is defined as
(6.54)
Subscript b indicates the building material.
The choice of spacing x and the time interval t must meet some criterion to ensure convergence
in the calculations. The criterion is the stability limit, or
(6.55)
Similarly, for moisture transfer at the interior nodes,
(6.56)
and
(6.57)
where X  moisture content, dimensionless
Dlv  mass diffusivity of liquid and vapor, ft2/s (m2/s)
Heat and Moisture Transfer at Surface Nodes
For a one-dimensional heat flow, the energy balance at surface node i of the partition wall as shown
in Fig. 6.8 is
Conductive heat from node i  1  convective heat transfer from space air
 latent heat of moisture transfer from space air  radiative heat from internal loads
 rate of change of internal energy of node i
 F1:iLr  Fp:iOr  Fm:i Mr 
bcb Ai
x
2
T i
t t  T i
t
t
kAi (T i1
t  T ti
)
x
 hci Ai(T t
r  T ti
) 
ahmi Ai Xti (wt
r  wt
is) hfg
Fomass 
Dlv t
(x)2  1
2
Xi tt  Fomass (Xti
1  Xti
1)  (1  2 Fomass) Xti
Fo  1
2
Fo 
kb
bcb
t
(x)2
T i
tt  Fo(T ti
1  T ti
1)  (1  2Fo)T ti

bcb Aix (T i
tt  T i
t)
t
qi1:i  qi1:i 
Ui
t
LOAD CALCULATIONS 6.37
where hci  convective heat-transfer coefficient of surface i, Btu/h ft2 F (W/m2 C)
a  density of space air, lb/ f t 3 (kg/m3)
hmi  convective mass-transfer coefficient of surface i, ft/s (m/s)
wt
is  humidity ratio corresponding to surface i at time t, lb/ lb (kg /kg)
hf g  latent heat of vaporization at surface temperature, Btu/lb (J /kg)
 shape factor between surface of lights, occupants, appliances and surface i
Lr,Or, Mr  radiative portion of heat energy from lights, occupants, and appliances and machines,
Btu/ h (W)
Solving for gives
(6.58)
In Eq. (6.58) Bi is the Biot number and can be expressed as
(6.59)
The stability limit of the surface nodes requires that
(6.60)
Similarly, according to the principle of conservation of mass, the moisture content at surface node i
is given as
(6.61)
The temperature and moisture content at other surface nodes such as the floor, ceiling, exterior
walls, glass, and Plexiglas of the lighting fixture can be found in the same manner.
According to ASHRAE Handbook 1993, Fundamentals, the radiative and convective portions of
the heat gains are as follows:
Radiative, percent Convective, percent
Fluorescent lights 50 50
People 33 67
External walls and roofs 60 40
Appliance and machines 2080 8020
Space Air Temperature and Cooling Loads
If the infiltrated air is ignored, the heat balance on the space air or the plenum air can be described
by the following relationship:
Internal energy of supply air  convective heat transfer from building structures
 convective heat transfer from internal loads  internal energy of space air
60V?s scpaT s
t t  n
k1
hck Ak(T i
t t  T r
t t)  Lc  Oc  Mc  60V?s scpaT r
t t
Xi tt  (1  2Fomass)Xi t  2Fomass Xti
1 
2 ahmi t Xi t (wt
r  wt
is)
b x
Fo(1  Bi)  1/2
Bi 
hc x
k

F1:i Lr
hci Ai

Fp:i Or
hci Ai

Fm:i Mr
hci Ai 	  [1  2Fo(1  Bi)]T i
t
Ti
tt  2FoT ti
1  Bi	T r
t 
a hmi hfgXi t (wr t  wis
t )
hci
T i
tt
F1:i, Fp:i, Fm:i
6.38 CHAPTER SIX
Solving for Tr
tt, we have
(6.62)
where ,Ts  volume flow and temperature of supply air, cfm (m3 / (60 s)) and F (C)
Lc, Oc, Mc  convective heat from lights, occupants, and appliances, Btu/h (W)
The temperature of the return plenum air can be similarly calculated. The space sensible cooling
loads, therefore, can be calculated as
(6.63)
The latent heat gains are instantaneous latent cooling loads.
6.10 HEATING LOAD
Basic Principles
The design heating load, or simply the heating load, is always the maximum heat energy that might
possibly be required to supply to the conditioned space at winter design conditions to maintain the
winter indoor design temperature. The maximum heating load usually occurs before sunrise on the
coldest days. The following are the basic principles of heating load calculation that are different
from those for the cooling load calculation:
 All heating losses are instantaneous heating loads. The heat storage effect of the building structure
is ignored.
 Solar heat gains and the internal loads are usually not taken into account except for those internal
loads Qin, Btu /h (W), that continuously release heat to the conditioned space during the operating
period of the whole heating season.
 Only that latent heat Ql , Btu /h (W), required to evaporate liquid water for maintaining necessary
space humidity is considered as heating load.
For a continuously operated heating system, the heating load Qrh, Btu /h (W), can be calculated as
(6.64)
where Qtran  transmission loss, Btu/h (W)
Qinf, s  sensible heat loss from infiltrated air, Btu /h (W)
Qmat  heat added to entering colder product or material, Btu/ h (W)
Transmission Loss
Transmission loss Qtran, Btu /h (W), is the sum of heat losses from the conditioned space through
the external walls, roof, ceiling, floor, and glass. If the calculation is simplified to a steady-state
heat flow, then
(6.65) Qtran   AU(Tr  To)
Qrh  Qtran  Qif,.s  Ql  Qmat  Qin
Qrs
tt  n
k1
hck Ak(T i
t t  T r
tt )  Lc  Oc  Mc
V?
s
T r
tt 
60V?s scpaT s
t t  n
k1
hck AkT i
t t  Lc  Oc  Mc
60V?s scpa  n
k1
hck Ak
LOAD CALCULATIONS 6.39
where A  area of walls, roof, ceiling, floor, or glass, ft2 (m2)
U  overall heat-transfer coefficient of walls, roof, ceiling, floor, or glass, Btu/h ft2 F
(W/m2 C)
When the winter outdoor design temperature is used for To, the heat loss calculated by transient
heat transfer will be less than that from Eq. (6.65) because of the cyclic fluctuations of the outdoor
temperature and the heat storage in the building structures.
For concrete slab floors on a grade, heat loss Qfl, Btu /h (W), is mostly through the perimeter instead
of through the floor and the ground. It can be estimated as
(6.66)
where P  length of perimeter, ft (m)
Cfl  heat loss coefficient per foot (meter) of perimeter length, Btu/h ft F (W/mC)
For areas having an annual total of heating degree-days HDD65  5350 and for a concrete wall
with interior insulation in the perimeter having an R value of 5.4 h ft2 F/Btu (0.97 m2 C/W),
Cfl  0.72 Btu/h ft F (1.25 W/mC). Refer to ASHRAE Handbook 1989, Fundamentals, chapter
25, for more details.
For basement walls, the paths of the heat flow below the grade line are approximately concentric
circular patterns centered at the intersection of the grade line and the basement wall. The thermal
resistance of the soil and the wall depends on the path length through the soil and the construction
of the basement wall. A simplified calculation of the heat loss through the basement walls and floor
Qb,g, Btu /h (W), is as follows:
(6.67)
where Ab,g  area of basement wall or floor below grade, ft2 (m2)
Ub,g  overall heat-transfer coefficient of wall or floor and soil path, Btu/h ft2 F
(W/m2 C)
The values of Ub,g are roughly given as follows:
0 to 2 ft below grade Lower than 2 ft
Uninsulated wall 0.35 0.15
Insulated wall 0.14 0.09
Basement floor 0.03 0.03
Refer to ASHRAE Handbook 1989, Fundamentals, for details.
The space heating load in the perimeter zone, including mainly transmission and infiltration
losses, is sometimes expressed in a linear density qh, ft, in Btu/ h per linear foot of external wall, or
Btu/hft (W/m).
Adjacent Unheated Spaces
Heat loss from the heated space to the adjacent unheated space Qun, Btu /h (W), is usually assumed
to be balanced by the heat transfer from the unheated space to the outdoor air, and this can be calculated
approximately by the following formula:
(6.68) Qun  n
i1
AiUi(Tr  Tun)  (m
j1
AjUj  V?inf o cpa)(Tun  To)
Qb, g  Ab, gUb, g(Tbase  To)
Qfl  PCfl(Tr  To)
6.40 CHAPTER SIX
where Ai,Ui  area and overall heat-transfer coefficient of partitions between heated space and
unheated space, ft2 (m2) and Btu/h ft2 F (W/m2 C)
Aj,Uj  area and overall heat-transfer coefficient of building structures exposed to outdoor
air in unheated space, ft2 (m2) and Btu/h ft2 F (W/m2 C)
The temperature of the unheated space Tun, F (C), can be calculated as
(6.69)
Latent Heat Loss and Heat Loss from Products
In Eq. (6.64), Ql, Btu /h (W), represents the heat required to evaporate the liquid water to raise the
relative humidity of the space air or to maintain a specific space relative humidity, i.e.,
(6.70)
where  mass flow of water evaporated, lb/h (kg/ s)
 volume flow rate of outdoor ventilation air and infiltrated air, cfm [m3 / (60 s)]
 mass flow of water evaporated from minimum number of occupants that always stay
in conditioned space when heating system is operated, lb /h (kg / s)
In Eq. (6.70), hfg,57 indicates the latent heat of vaporization at a wet-bulb temperature of 57F
(13.9C), that is, 72F (22.2C) dry-bulb temperature and a relative humidity of 40 percent. Its
value can be taken as 1061 Btu/ lb (2.47 	 106 J / kg).
For factories, heat added to the products or materials that enter the heated space within the occupied
period Qmat, Btu /h (W), should be considered part of the heating load and can be calculated as
(6.71)
where  mass flow rate of cold products and cold material entering heated space, lb/h (kg/ s)
cpm  specific heat of product or material, Btu/ lb F (J /kgC)
Infiltration
Infiltration can be considered to be 0.15 to 0.4 air changes per hour (ach) at winter design conditions
only when (1) the exterior window is not well sealed and (2) there is a high wind velocity. The
more sides that have windows in a room, the greater will be the infiltration. For hotels, motels, and
high-rise domicile buildings, an infiltration rate of 0.038 cfm / ft2 (0.193 L/ s m2) of gross area of
exterior windows is often used for computations for the perimeter zone.
As soon as the volume flow rate of infiltrated air , cfm (m3 /min), is determined, the sensible
heat loss from infiltration Qinf, s, Btu /h (W), can be calculated as
(6.72)
where o  density of outdoor air, lb/ ft3 (kg/m3).
Setback of Night Shutdown Operation
During a nighttime or unoccupied period, when the space temperature is set back lower than the
indoor temperature during the operating period, it is necessary to warm up the conditioned space
Qinf, s  V?inf
ocpa(Tr  To)
V?
inf
m? mat
Qmat  m? matcpm(Tr  To)
m? p
V?
oinf
m? w
Ql  m? w hfg,57  [60V?oinf
o(wr  wo)  m? p]hfg,57
Tun  n
i1
AiUiTr 
(60V?inf ocpa  m
j1
AjUj)To
n
i1
AiUi  60V?inf ocpa  m
j1
AjUj
LOAD CALCULATIONS 6.41
the next morning before the arrival of the occupants in offices or other buildings. The warm-up or
pickup load depends on the pickup temperature difference, the outdoor-indoor temperature difference,
the construction of the building envelope or building shell, and the time required for warm-up.
There are insufficient data to determine the oversizing factor of the capacity of the heating plant for
morning warm-up. According to tests by Trehan et al. (1989), one can make a rough estimate of the
energy required to warm up or raise a space temperature by 12F (6.7C) for a single- or two-story
building with a roof-ceiling U value of 0.03 Btu/h ft2 F (0.17 W/m2 C) and a U value for external
walls of 0.08 Btu /h ft2 F (0.45 W/m2 C) as follows:
Outdoor-indoor temperature difference, F Warm-up period, h Oversizing factor, percent
35 1 40
55 2 40
55 1 100
As discussed in Sec. 5.16, heating pickup load during warm-up period depends on the difference
between space temperature and occupied set point, and the amount of time prior to scheduled occupancy,
and is often determined by computer software.
6.11 LOAD CALCULATION SOFTWARE
Introduction
Today, most of the load calculations are performed by personal computers. Among the widely
adopted load calculation and energy analysis software, only Building Load Analysis and System
Thermodynamics (BLAST) developed by the University of Illinois adopts the heat balance method.
All the others are based on the transfer function or weighting factors method.
Load calculation software can also be divided into two categories. The first includes those
developed by government or public institutions, such as the Department of Energy (DOE-2.0),
National Bureau of Standards Load Program (NBSLP), and BLAST, which are white box, or
transparent to the user, and called public domain software. The second category consists of
those programs developed by the private sector, such as TRACE 600, developed by The Trane
Company; HAP E20-II, developed by Carrier Corporation; and HCC (loads) and ESP (energy),
developed by Automatic Procedures for Engineering Consultants Inc. (APEC). The privatesector-
developed software programs were based on the published literature of government,
research, and public institutions such as ASHRAE. The most widely used, reliable, user-friendly,
and continuously supported load and energy calculation programs in design are TRACE 600,
HAP E20-II, and DOE-2.1E. BLAST is the most elaborate load calculation program developed in
the United States and is usually considered a research tool. Most HVAC&R designers do care
about the accuracy of the computational results of the software; however, the priority is the userfriendly
inputs and outputs.
Trace 600Structure and Basics
In this section we include the basics and inputs of the most widely used software program in load
calculations, TRACE 600 written for DOS, in the United States. Recently, Windows 95-based
TRACE 700 Load Design has become available. TRACE 600 Load and TRACE 700 Load Design
are similar in structure, basics, and engineering capabilities. Most of the load and energy calculation
software consists of four principal programs: loads, systems, plant, and economics (LSPE).
TRACE 600 divides into five phases:
6.42 CHAPTER SIX
 Load. In the load phase, all external and internal heat gains are calculated. The heat gain profile is
then converted to a cooling load profile.
 Design. In the design phase, based on the maximum block space sensible load, the volume flow
and the size of the air system are determined from the psychrometric analysis. In design phase it
also calculates the coils load and selects the size of the coils.
 System. In the system simulation, the program predicts the load that is imposed on the equipment
according to the space load profile and the type of air system selected.
 Equipment. In the equipment (plant) phase, the energy consumption of the fans, furnaces or boilers,
and refrigeration systems is determined based on the hourly coil loads.
 Economic. During the economic analysis, utility cost is calculated from the energy use. When it
is combined with installation cost and maintenance cost, a life cycle cost is available for
comparison.
Only the load calculation program is covered in this section. The available documentation for
TRACE 600 includes Users Manual, Engineering Manual, Quick Reference, Cook Book, Software
BulletinGetting Started, etc.
The hardware requirements for TRACE 600 are as follows:
LOAD CALCULATIONS 6.43
IBM AT 286 processor or higher
Program: 7.5 Mbytes hard disk space (3 M
bytes for load calculation)
Job: 3 to 10 Mbytes depending on the size of
job
640 kbytes
Personal computer
Hard disk
Ram requirement
DOS Version 3.1 or higher
The TRACE 600 data sheet can be created by using the input editor. The input file is organized
in card format. In each input card, every input is shown by a given field. Sometimes, the value and
unit of input are separated into two fields. Entering the inputs can be done either by using field
mode or by using full-screen mode. In field mode, the user can access help and selection to determine
and select a specified input. Field mode input is preferable. In full-screen mode, the user can
view a full screen of raw input. Each line shows the inputs of a card and is prefixed by a twocharacter
code. Each field in a card is separated by a forward slash.
Input data are subdivided into five groups: job, load, system, equipment, and economic. Many
input cards and fields inside the input cards are optional, i.e., may be left blank.
Trace 600 InputLoad Methodology
In TRACE 600, card 10 lists various methods of cooling and heating load calculations for the users
selection. There are five cooling load calculating methods:
 CEC-DOE2. This method adopts the transfer function method for both heat gain and space cooling
load calculations. The space load calculations adopt the precalculated weighting factors listed
in DOE 2.1 Engineering Manual. This is a comparatively exact cooling load and energy calculation
method, especially for air systems operated at nighttime shutdown mode. CEC-DOE2 needs
more computational calculations. However, it is often not a primary problem when the computations
are performed by a powerful PC.
 CLTD-CLF. This method also adopts the transfer function method for both heat gain and space
cooling load calculations. Cooling load temperature difference (CLTD) and cooling load factor
(CLF) tables are prepared based on the exact transfer function coefficients or weighting factors
from the TFM method. For many commercial buildings, when the air system is shut down during
an unoccupied period, the external heat gains entering the space cannot be reasonably allocated
over the cool-down period and the successive operating hours.
 TETD-TA1 and TETD-TA2. The TETD-TA1 method adopts the transfer function method to calculate
heat gain, and TETD-TA2 uses an approximate TETD to calculate heat gain which adopts a
decrement factor and a time lag 
to describe the amplitude and time-delay characteristics of the
heat wave inside an exterior wall or a concrete roof slab. Both TETD-TA1 and TETD-TA2 use the
time-averaging (TA) method to convert the heat gain to space cooling load. The TA technique
lacks scientific support.
 TETD-PO. This method also adopts an approximate TETD to calculate the heat gain. It uses Post
Office RMRG weighting factors to convert a heat gain of 100 percent radiative to space cooling
load, and Post Office RMRX weighting factors to convert a heat gain which is not 100 percent
radiative, such as heat gains from people, lights, and equipment.
There are six heating load calculation methods. Five of themCEC-DOE2, CLTD-CLF,
TETD-TA1, TETD-TA2, and TETD-POhave already been described in the cooling load calculation
methodology. The sixth heating load calculation method is called the UATD method. In the
UATD method, heat losses are calculated based on the U value 	 area 	 temperature difference,
which is the temperature difference of the outdoor and indoor design temperatures. Heat losses are
also considered as instantaneous heating load.
For peak load calculation at winter design conditions, internal heat gains are not taken into
account in the UATD method. However, for energy use calculation, internal heat gains need to be
taken into account according to the load schedules. Otherwise the heating energy use becomes too
conservative.
Trace 600 InputJob
There are 11 job cards: Cards 01 to 05 are used to describe the name of the project, its location, the
client of the project, the program user, and comments. Card 08, climatic information, lists the name
of weather file, summer and winter clearness numbers, outdoor dry- and wet-bulb temperatures,
ground reflectance, as well as the building orientation. Card 09, load simulation periods, covers the
first and last month of cooling design, summer period, and daylight savings time. It also covers the
peak cooling load hour.
In card 10, load simulation parameters (optional), the selection of cooling and heating load
methodology has been described in preceding paragraphs. In addition, the input data cover airflow
input and output units, percentage of wall load included in the return air, and room circulation rate
when the transfer function method is used in the conversion of heat gain to space cooling load. Outdoor
air dry-bulb temperature and humidity ratio are also required to determine the state of the supply
air during psychrometric analysis.
The sensible and latent loads due to the ventilation outdoor air intake are a component of cooling
coil load. They do not affect the supply airflow and the size of the supply fan if the mixture of
outdoor and recirculating air is extracted by the supply fan.
Card 11, energy simulation parameters (optional), describes the first and last months of energy
simulation, holiday and calendar type, and the conditioned floor area. It also determines the input
data of calculation level whether it is at room, zone, or air system level.
Card 12, resource utilization factors (optional), covers the input data of energy utilization
factors which indicate the inefficiency of producing and transfering energy, of electricity, gas,
oil, steam, hot water, chilled water, and coal. Card 13, daylighting parameters (optional),
describes the atmospheric moisture, atmospheric turbidity, or a measure of aerosols that
affects daylighting, inside visible reflectivity, and geometry method, such as glass percentage
(GLAS-PCT).
6.44 CHAPTER SIX
Trace 600 InputExternal Loads
There are seven cards for external load input.
Card 19, load alternative description, will save separate files for each of the Load, Design, System,
and Equipment alternatives.
Card 20, general room parameters, determines the room number, zone reference number, room
description, floor length and width, construction type (a 2- to 12-in. (50- to 300-mm), light- or
heavyweight concrete construction), plenum height, R value of the acoustic ceiling, floor-to-
floor height, perimeter length, and duplicate floor and duplicate room/zone multipliers to save
input data for a similar floor or room/zone.
Card 21, thermostat parameters, lists input data of room design cooling and heating dry-bulb
temperatures and relative humidities; the highest and lowest room temperatures allowable to
drift up during low occupancy or nonoccupancy; cooling and heating thermostat schedules;
location of the thermostat; light, medium, or heavy room construction and the corresponding
2 to 8 h of time averaging; and carpet covering.
Card 22, roof parameters, describes the room number, alternate roof numbers, whether roof area
is equal to floor area, roof length and width, roof U value, roof construction type, roof direction,
roof tilting angle, and roof absorptivity .
Card 23, skylight parameters (optional), covers the room number, roof number, length and width of
skylight, number of skylights or percentage of glass of roof area, skylight U value, skylight shading
coefficient, external and internal shading types, percentage of solar load to be picked up by return
plenum air because a portion of skylight is exposed to the plenum air, visible light transmissivity,
and the fraction of inside visible light being reflected on the inside surface of the glazing.
Card 24, wall parameters, lists the input data of room number; wall number; the length, height,
U value, and construction type of the wall; wall direction; wall tilting angle; wall absorptivity;
and ground reflectance.
Card 25, wall glass parameters, describes the room number, wall number, glass length and height,
number of windows or percentage of glass (of gross wall area), glass U value, shading coefficient,
external and internal shading type, percentage of solar load to return plenum air, visible light
transmissivity, and the inside visible light being reflected on the inside surface of the glazing.
Trace 600 InputSchedules
In TRACE 600, card 26 specifies the load and operating schedules of the internal loads, fans, reheating,
and daylighting (optional). The operating schedule of internal loads and equipment affects
the design load (maximum block load), especially the annual energy use.
1. For the design load calculation of internal loads include people, lighting, and miscellaneous
equipment:
 TRACE 600 assumes that the internal loads are scheduled at 100 percent, 24 h/day during
cooling design months operation and at 0 percent, 24 h/day during heating design months, a
schedule of cooling only (CLGONLY).
 If the pickup, cool-down, or warm-up load is greater than the design load, the user should
consider oversizing the cooling capacity and heating capacity to handle the pickup loads.
 For the energy use calulation of internal loads, choose or create a schedule at which the internal
load varies 24 h /day.
2. For the outdoor ventilation air and infiltration loads:
 TRACE 600 assumes that the design outdoor ventilation air and infiltration are scheduled at
100 percent, 24 h/day for both cooling and heating design months.
LOAD CALCULATIONS 6.45
 For the calculation of the energy use, only when the outdoor ventilation air system is separated
from the main supply system is it then possible to choose and to create a schedule in
which the airflow of outdoor ventilation air varies according to peoples occupancy.
 During the calculation of energy use for a variable-air-volume (VAV) system, the airflow of
the outdoor ventilation air may be varied based on its control systems into the following
schedule: 100 percent, 24 h/day; varying according to supply volume flow of the air system;
varying according to peoples occupancy.
 For the calculation of energy use due to infiltration, choose and create a schedule in which
the infiltration varies inversely with supply fan schedule, such as for hours fan schedule
is 100 percent, infiltration is zero; then when fan schedule is 0 percent, the infiltration is
100 percent.
3. Operating schedules of fans that affect the design load and energy use calculations are as
follows:
 At design load, TRACE 600 assumes that the main supply fan and auxiliary fan, if any, are
scheduled at 100 percent, 24 h/day as the default value during cooling mode operation, and
fan heat will be included in the cooling coil load. Similarly, at design load, main supply and
auxiliary fans are scheduled at 0 percent, 24 h/day during heating mode operation, and no
fan heat will be taken into account.
 In the energy use calculation, main supply and auxiliary fans in a constant-volume air system
are scheduled at 100 percent, 24 h/day. Or choose or create a schedule at which fans will be
cycling (on and off) depending on the space cooling and heating load.
 In the energy use calculation, main supply and auxiliary fans in a VAV system are scheduled
depending on the space cooling and heating load.
 For a room exhaust fan operated at cooling design load, it is scheduled at 100 percent,
24 h/day as the default value. For the energy use calculation, choose or create an appropriate
operating schedule.
4. The reheat minimum percentage determines the ratio of reheat minimum airflow for that room
to the design supply airflow.
5. Daylighting schedule describes the availability of daylighting each hour. Choose an available
schedule or create an appropriate daylighting schedule.
Trace 600 InputInternal Loads
Four cards cover the input data of internal loads and airflows.
Card 27, people and lights, describes the room number, peoples density and corresponding unit,
peoples sensible and latent loads, lighting heat gain and corresponding unit, type of lighting
fixture, ballast factor, and daylighting reference points 1 and 2, to show whether there is
daylighting control or one or two control areas. According to the types of lighting fixtures and
the return airflow per unit floor area, the percentage of lighting load to return air should be
determined more precisely. Refer to Sec. 6.6 for details.
Card 28, miscellaneous equipment (optional), covers the room number, miscellaneous equipment
reference number, equipment description, energy consumption value and corresponding
units, schedule and energy metering code, percentage of sensible load, percentage of miscellaneous
equipment load to space and to air path, radiant fraction, and whether the plenum air,
exhaust air, or return air path picks up the miscellaneous load.
Card 29, room airflows, covers the flow rates of outdoor ventilation air during cooling and heating
operation and their corresponding units, the flow rates of infiltration air for cooling and
heating and their corresponding units, and the amount of minimum reheat and its units.
Card 30, fan flow rates, describes the room number, airflow of the main and auxiliary cooling and
heating supply fans and their corresponding units, and airflow of room exhaust fan and its unit.
6.46 CHAPTER SIX
Trace 600 InputPartition and Shading Devices
Five cards cover the input data of partition and shading devices.
Card 31, partition parameters (optional), describes the room number, adjacent room number,
partition number, and the length, height, U value, and construction type of the partition wall.
The adjacent space temperature during cooling and heating periods can follow these profiles:
constant, sine fit, prorate the adjacent space temperature against the outdoor air temperature,
vary as the outdoor air temperature, or there is no heat transfer from the adjacent space through
the partition.
Card 32, exposed floor parameters (optional), lists the room number, adjacent room number, exposed
floor number, the perimeter length and loss coefficient of the slab on grade; the area, U
value, and construction type of the exposed floor; and adjacent space temperature during cooling
and heating periods whether it is constant, sine curve fit, or prorated curve fit.
Card 33, external shading devices (optional), describes its constructional type, glass height and
glass width, the height above glass and the horizontal projection (projection out) of the overhang,
the projection left and right, the left and right projection out, and the shading due to adjacent
buildings. TRACE 600 does not allow for shading because of both adjacent building and
overhang.
Card 34, internal shading devices, covers the types, overall U value, overall shading coefficient,
location, operating schedule, and overall visible light transmittance of internal shading
device; requires that below the minimum outdoor air dry-bulb temperature and above the
maximum solar heat gain, the internal shading device is deployed; and sun control and glare
control probability.
Card 35, daylight sensor (optional), describes the daylight sensor reference number, percentage
of space affected, lighting set point in footcandles (luxes), type of control (continuous or
stepped), minimum power or minimum light percentage, light control steps, manually operated
control probability, height of the reference point and its distance from the glass, window sill
height, ratio of the glass length seen by a sensor to the distance between the reference point and
the glass, and skylight length and distance from the reference point.
Trace 600Minimum Input Requirements, Run, and Outputs
For load calculation and energy analysis, the following are the minimum input requirements:
Card 08: climatic information, field 2: weather filename.
Card 20: general room parameters, field 1: room number.
Card 40: system type, field 1: system number, and field 2: system type.
Card 41: zone assignment, field 1: system number, field 2: system serving these zones begins at,
and field 3 : end at.
After the user has completed all the input required for the job, he or she must edit for errors,
check for possible data loss, and run and save. Choose output from the TRACE 600 menu. Select
the desired section of the output. Print the compressed report in a desired order.
Example 6.2. The plan of a typical floor in a multistory office building in New York City is
shown in Fig. 6.9. The tenants and the partition walls of this floor are unknown during the design
stage. The core part contains the restrooms, stairwells, and mechanical and electrical service
rooms, and it has mechanical exhaust systems only. This building has the following construction
characteristics:
LOAD CALCULATIONS 6.47
This typical floor has the following internal loads: The occupant density is 7 persons per 1000
ft2 (93 m2). For each person, there is 250 Btu/h (73 W) sensible heat gain, and 250 Btu/h (73 W)
latent heat gain. The lighting load, including ballast allowance, is 1 W/ ft2 (10.7 W/m2). The miscellaneous
equipment load due to personal computers and appliances is 1 W/ ft2 (10.7 W/m2).
6.48 CHAPTER SIX
FIGURE 6.9 Typical plan of a high-rise office building.
6300 ft2 (586 m2)
5600 ft2 (520 m2)
13 ft (4 m)
9 ft (2.74 m)
0.345
0.67 Btu/h ft2 F (3.8 W/m2 C)
0.36
120 ft (36.6 m)
13 ft (2.74 m)
0.12 Btu/ hft2 F (0.68 W/ (m2 C)
200 ft (61 m)
9 ft (2.74 m)
0.12 Btu/h ft2 F (0.68 W/m2 C)
Area of perimeter zone
Area of interior zone
Floor-to-floor height
Floor-to-ceiling height
Window:
Percentage of glass/exterior wall
U value
Interior shading, SC
Exterior wall:
Length
Height
U value
Partition wall:
Length
Height
U value
The summer indoor design temperature is 75F (23.9C). If the infiltration load and the shading
due to the adjacent building are ignored, calculate the maximum space cooling load during summer
outdoor design conditions of this typical floor by using TRACE 600 Load program with the CLTDCLD
method.
Solution. The maximum space cooling loads for this typical floor, calculated by using TRACE
600 Load, with CLTD-CLF method on July 7, 5 P.M., are as follows:
Manually calculated in
TRACE 600, Btu/h Example 7.2 of first edition, Btu/h
Envelope loads
Glass solar 36,846
Glass conduction 11,095
Wall conduction 9,927
Partition wall 1,296
Ceiling load 4,226
63,390
Internal loads
Lights 36,147
People 31,446
Miscellaneous 37,994
105,587
Grand total 168,977 198,770
Compare the TRACE 600 calculated space cooling load of 168,977 Btu/h (49,520 W) with the
manually calculated space cooling load of the same typical floor with the same building and operating
characteristics (in Example 7.2 of Handbook of Air Conditioning and Refrigeration, 1st ed.) of
198,770 Btu/h (58,240 W). The TRACE 600 calculated space cooling load is only 85 percent of the
manually calculated value. This is due to a lower solar CLF and therefore lower solar cooling loads
in the TRACE 600 calculation.
REFERENCES
Amistadi, H., Energy Analysis Software Review, Engineered Systems, no. 10, 1993, pp. 3445.
ASHRAE, ASHRAE/IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential
Buildings, ASHRAE Inc., Atlanta, GA, 1989.
ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, GA, 1997.
Ayres, J. M., and Stamper, E., Historical Development of Building Energy Calculations, ASHRAE Journal,
no. 2, 1995, pp. 4753.
Carrier Air Conditioning Co., Handbook of Air Conditioning System Design, 1st ed., McGraw-Hill, New York, 1965.
Deringer, J. J., An Overview of Standard 90.1: Building Envelope, ASHRAE Journal, February 1990, pp. 3034.
Harris, S. M., and McQuiston, F. C., A Study to Categorize Walls and Roofs on the Basis of Thermal Response,
ASHRAE Transactions, 1988, Part II, pp. 688715.
Johnson, C. A., Besent, R.W., and Schoenau, G. J., An Economic Parametric Analysis of the Thermal Design of
a Large Office Building under Different Climatic Zones and Different Billing Schedules, ASHRAE Transactions,
1989, Part I, pp. 355369.
Kerrisk, J. F., Schnurr, N. M., Moore, J. E., and Hunn, B. D., The Custom Weighting-Factor Method for Thermal
Load Calculations in the DOE-2 Computer Program, ASHRAE Transactions, 1981, Part II, pp. 569584.
Kimura, K. I., and Stephenson, D. G., Theoretical Study of Cooling Loads Caused by Lights, ASHRAE Transactions,
1968, Part II, pp. 189197.
LOAD CALCULATIONS 6.49
Komor, P., Space Cooling Demands from Office Plug Loads, ASHRAE Journal, no. 12, 1997, pp. 4144.
Kreith, F., and Black,W. Z., Basic Heat Transfer, Harper & Row, New York, 1980.
Mackey, C. O., and Gay, N. R., Cooling Load from Sunlit Glass, ASHVE Transactions, 1952, pp. 321330.
Mackey, C. O., and Wright, L. T., Periodic Heat FlowHomogeneous Walls or Roofs, ASHVE Transactions,
1944, pp. 293312.
McQuiston, F. C., and Spitler, J. D., Cooling and Heating Load Calculation Manual, 2d ed., ASHRAE Inc.,
Atlanta, GA, 1992.
Mitalas, G. P., Transfer Function Method of Calculating Cooling Loads, Heat Extraction Rate and Space Temperature,
ASHRAE Journal, no. 12, 1972, p. 52.
Palmatier, E. P., Thermal Characteristics of Structures, ASHRAE Transactions, 1964, pp. 4453.
Persily, A. K., and Norford, L. K., Simultaneous Measurements of Infiltration and Intake in an Office Building,
ASHRAE Transactions, 1987, Part II, pp. 4256.
Romine, T. B., Cooling Load Calculation: Art or Science? ASHRAE Journal, no. 1, 1992, pp. 1424.
Rudoy,W., and Duran, F., Development of an Improved Cooling Load Calculation Method, ASHRAE Transactions,
1975, Part II, pp. 1969.
Rudoy,W., and Robins, L. M., Pulldown Load Calculations and Thermal Storage during Temperature Drift,
ASHRAE Transactions, 1977, Part I, pp. 5163.
Snelling, H. J., Duration Study for Heating and Air Conditioning Design Temperature, ASHRAE Transactions,
1985, Part II B, p. 242.
Sowell, E. F., Classification of 200,640 Parametric Zones for Cooling Load Calculations, ASHRAE
Transactions, 1988, Part II, pp. 754777.
Sowell, E. F., and Chiles, D. C., Zone Descriptions and Response Characterization for CLF/CLTD Calculations,
ASHRAE Transactions, 1985, Part II A, pp. 179200.
Sowell, E. F., and Hittle, D. C., Evolution of Building Energy Simulation Methodology, ASHRAE Transactions,
1995, Part I, pp. 851855.
Stephenson, D. G., and Mitalas, G. P., Cooling Load Calculations by Thermal Response Factor Method,
ASHRAE Transactions, 1967, Part III, pp. 1.11.7.
Sun, T. Y., Air Conditioning Load Calculation, Heating/Piping/Air Conditioning, January 1986, pp. 103113.
The Trane Company, TRACE 600: Engineering Manual, The Trane Company, LaCrosse, WI, 1992.
The Trane Company, TRACE 600: Users Manual, LaCrosse, WI, 1992.
Trehan, A. K., Fortmann, R. C., Koontz, M. D., and Nagda, N. L., Effect of Furnace Size on Morning Pickup
Time, ASHRAE Transactions, 1989, Part I, pp. 11251129.
Wang, S. K., Air Conditioning, vol. 1, Hong Kong Polytechnic, Hong Kong, 1987.
Williams, G. J., Fan Heat: Its Source and Significance, Heating/Piping/Air Conditioning, January 1989,
pp. 101112.
6.50 CHAPTER SIX
CHAPTER 7
WATER SYSTEMS
7.1
7.1 FUNDAMENTALS 7.2
Types of Water System 7.2
Volume Flow and Temperature
Difference 7.4
Water Velocity and Pressure Drop 7.5
7.2 WATER PIPING 7.7
Piping Material 7.7
Piping Dimensions 7.7
Pipe Joints 7.13
Working Pressure and Temperature 7.13
Expansion and Contraction 7.14
Piping Supports 7.15
Piping Insulation 7.15
7.3 VALVES, PIPE FITTINGS, AND
ACCESSORIES 7.16
Types of Valves 7.16
Valve Connections and Ratings 7.17
Valve Materials 7.18
Piping Fittings and Water System
Accessories 7.18
7.4 WATER SYSTEM PRESSURIZATION
AND THE PRESENCE OF AIR 7.19
Water System Pressurization
Control 7.19
Open Expansion Tank 7.20
Closed Expansion Tank 7.21
Size of Diaphragm Expansion Tank 7.21
Pump Location 7.23
Air in Water Systems 7.23
Penalties due to Presence
of Air and Gas 7.24
Oxidation and Waterlogging 7.24
7.5 CORROSION AND DEPOSITS IN
WATER SYSTEM 7.25
Corrosion 7.25
Water Impurities 7.25
Water Treatments 7.26
7.6 CLOSED WATER SYSTEM
CHARACTERISTICS 7.27
System Characteristics 7.27
Changeover 7.28
7.7 CENTRIFUGAL PUMPS 7.30
Basic Terminology 7.30
Performance Curves 7.32
Net Positive Suction Head 7.33
Pump Selection 7.33
7.8 PUMP-PIPING SYSTEMS 7.34
System Curve 7.34
System Operating Point 7.34
Combination of Pump-Piping
Systems 7.35
Modulation of Pump-Piping
Systems 7.36
Pump Laws 7.37
Wire-to-Water Efficiency 7.37
7.9 OPERATING CHARACTERISTICS OF
CHILLED WATER SYSTEM 7.38
Coil Load and Chilled Water
Volume Flow 7.38
Chiller Plant 7.39
Variable Flow for Saving Energy 7.40
Water Systems in Commercial
Buildings 7.40
7.10 PLANT-THROUGH-BUILDING
LOOP 7.40
Bypass Throttling Flow 7.40
Distributed Pumping 7.41
Variable Flow 7.41
7.11 PLANT-BUILDING LOOP 7.43
System Description 7.43
Control Systems 7.43
System Characteristics 7.45
Sequence of Operations 7.46
Low T between
Chilled Water Supply and Return Temperatures
7.49
Variable-Speed Pumps Connected in
Parallel 7.49
Use of Balancing Valves 7.49
Common Pipe and Thermal
Contamination 7.51
7.12 PLANT-DISTRIBUTED PUMPING 7.52
7.13 CAMPUS-TYPE WATER
SYSTEMS 7.53
Plant-Distribution-Building Loop 7.54
Plant-Distributed Building Loop 7.56
Multiple Sources-Distributed
Building Loop 7.57
Chilled and Hot Water Distribution
Pipes 7.58
7.14 COMPUTER-AIDED PIPING DESIGN
AND DRAFTING 7.58
General Information 7.58
Computer-Aided Drafting
Capabilities 7.58
Computer-Aided Design
Capabilities 7.59
REFERENCES 7.60
7.1 FUNDAMENTALS
Types of Water System
Water systems that are part of an air conditioning system and that link the central plant,
chiller / boiler, air-handling units (AHUs), and terminals may be classified into the following categories
according to their use:
Chilled Water System. In a chilled water system, water is first cooled in the water chillerthe evaporator
of a reciprocating, screw, or centrifugal refrigeration system located in a centralized plantto
a temperature of 40 to 50F (4.4 to 10.0C). It is then pumped to the water cooling coils in AHUs and
terminals in which air is cooled and dehumidified. After flowing through the coils, the chilled water
increases in temperature up to 60 to 65F (15.6 to 18.3C) and then returns to the chiller.
Chilled water is widely used as a cooling medium in central hydronic air conditioning systems.
When the operating temperature is below 38F (3.3C), inhibited glycols, such as ethylene glycol or
propylene glycol, may be added to water to create an aqueous solution with a lower freezing point.
Evaporative-Cooled Water System. In arid southwestern parts of the United States, evaporativecooled
water is often produced by an evaporative cooler to cool the air.
Hot Water Systems. These systems use hot water at temperatures between 450 and 150F (232
and 66C) for space and process heating purposes. Hot water systems are covered in greater detail
in Chap. 8.
Dual-Temperature Water System. In a dual-temperature water system, chilled water or hot water
is supplied to the coils in AHUs and terminals and is returned to the water chiller or boiler mainly
through the following two distribution systems:
 Use supply and return main and branch pipes separately.
 Use the common supply and return mains, branch pipe, and coil for hot and chilled water supply
and return.
The changeover from chilled water to hot water and vice versa in a building or a system depends
mainly on the space requirements and the temperature of outdoor air. Hot water is often produced
by a boiler; sometimes it comes from a heat recovery system, which is discussed in later chapters.
Condenser Water System. In a condenser water or cooling water system, the latent heat of condensation
is removed from the refrigerant in the condenser by the condenser water. This condenser
water either is from the cooling tower or is surface water taken from a lake, river, sea, or well. For
an absorption refrigeration system, heat is also removed from the solution by cooling water in the
absorber. The temperature of the condenser water depends mainly on the local climate.
Water systems also can be classified according to their operating characteristics into the following
categories:
Closed System. In a closed system, chilled or hot water flowing through the coils, heaters,
chillers, boilers, or other heat exchangers forms a closed recirculating loop, as shown in Fig. 7.1a.
In a closed system, water is not exposed to the atmosphere during its flowing process. The purpose
of recirculation is to save water and energy.
Open System. In an open system, the water is exposed to the atmosphere, as shown in Fig. 7.1b.
For example, chilled water comes directly into contact with the cooled and dehumidified air in the
air washer, and condenser water is exposed to atmosphere air in the cooling tower. Recirculation of
water is used to save water and energy.
7.2 CHAPTER SEVEN
7.3
FIGURE 7.1 Types of water systems. (a) Closed system; (b) open system; (c) once-through system.
Open systems need more water treatments than closed systems because dust and impurities in
the air may be transmitted to the water in open systems. A greater quantity of makeup water is required
in open systems to compensate for evaporation, drift carryover, or blow-down operation.
Once-Through System. In a once-through system, water flows through the heat exchanger only
once and does not recirculate, as shown in Fig. 7.1c. Lake, river, well, or seawater used as condenser
cooling water represents a once-through system. Although the water cannot recirculate to the
condenser because of its rise in temperature after absorbing the heat of condensation, it can still be
used for other purposes, such as flushing water in a plumbing system after the necessary water
treatments, to conserve water. In many locations, the law requires that well water be pumped back
into the ground.
Volume Flow and Temperature Difference
The heating and cooling capacity of water when it flows through a heat exchanger Qw, Btu /h (W),
can be calculated as
(7.1)
where volume flow of water, ft3 /h (m3/s)
volume flow rate of water, gpm (L/ s)
density of water, lb/ ft3 (kg/m3)
cpw  specific heat of water, Btu / lb F (J /kgC)
Twe ,Twl  temperature of water entering and leaving heat exchanger,F (C)
temperature drop or rise of water after flowing through heat exchanger,F (C)
Here, the equivalent of pounds per hour is gpm  60 min/h  0.1337 ft3/gal  62.32 lb/ ft3  500.
Equation 7.1 also gives the relationship between Tw and Vgal during the heat-transfer process.
The temperature of water leaving the water chiller should be no lower than 37F (2.8C) to
prevent freezing. If the chilled water temperature is lower than 37F (2.8C), brine, ethylene glycol,
or propylene glycol should be used. Brine is discussed in a later chapter. For a dual-temperature
water system, the hot water temperature leaving the boiler often ranges from 100 to 150F (37.8 to
65.6C), and returns at a between 20 and 40F (11.1 and 22.2C).
For most dual-temperature water systems, the value of and the pipe size are determined
based on the cooling capacity requirement for the coils and water coolers. This is because chilled
water has a smaller than hot water does. Furthermore, the system cooling load is often higher
than the system heating load. For a chilled water system to transport each refrigeration ton of cooling
capacity, a of 8F (4.4C) requires a of 3 gpm (0.19 L/ s), whereas for a of 24F
(13.3C), is only 1.0 gpm (0.063 L/ s).
The temperature of water entering the coil Twe, the temperature of water leaving the coil Twl, and
the difference between them are closely related to the performance of a chilled
water system, air system, and refrigeration system:
 Temperature Twe directly affects the power consumption in the compressor.
 The temperature differential is closely related to the volume flow of chilled water and
thus the size of the water pipes and pumping power.
 Both Twe and influence the temperature and humidity ratio of air leaving the coil.
If the chilled water temperature leaving the water chiller and entering the coil is between 44 and
45F (6.7 and 7.2C), the off-coil temperature in the air system is usually around 55F (12.8C) for
conventional comfort air conditioning systems. In low-temperature cold air distribution systems,
Tw
V?
gal Tw
Tw  Twl  Twe
V?
gal
Tw V?gal Tw
Tw
V?
gal
T

Tw 
	w 
V?
gal 
V?
w 
 500V?gal T
 500V?gal (Twe  Twl)
Qw  V?w 	wcpw (Twe  Twl)
7.4 CHAPTER SEVEN
WATER SYSTEMS 7.5
chilled water leaving the chiller may be as low as 34F (1.1C), and the off-coil temperature is
often between 42 and 47F, typically 44F (5.6 and 8.3C, typically 6.7C).
The greater the value of for chilled water, the lower the amount of water flowing through
the coil. Current practice is usually to use a value of between 10 and 18F (5.6 and 10.0C) for
chilled water systems in buildings. Kelly and Chan (1999) noted a greater results at lower total
power consumption in water pumps, cooling tower fans, and chillers. However, a greater
means a larger coil and air-side pressure drop. For chilled water systems in a campus-type central
plant, a value of between 16 and 24F (8.9 and 13.3C) is often used.
Water Velocity and Pressure Drop
The maximum water velocity in pipes is governed mainly by pipe erosion, noise, and water hammer.
Erosion of water pipes is the result of the impingement of rapidly moving water containing air
bubbles and impurities on the inner surface of the pipes and fittings. Solden and Siegel (1964)
increased their feedwater velocity gradually from 8 ft/ s (2.4 m/ s) to an average of 35.6 ft / s (10.8
m/ s). After 3 years, they found no evidence of erosion in the pipe or a connected check valve. Erosion
occurs only if solid matter is contained in water flowing at high velocity. Velocity-dependent
noise in pipes results from flow turbulence, cavitation, release of entrained air, and water hammer
that results from the transient pressure impact on a sudden closed valve. Ball and Webster (1976)
performed a series of tests on -in. copper tubes with elbows. At a water velocity of 16.4 ft / s (5.0
m/ s), the noise level was less than 53 dBA. Tests also showed that cold water at a speed up to 21
ft / s (6.4 m/ s) did not cause cavitation. In copper and steel pipes, water hammer at a water velocity
of 15 ft / s (4.6 m/ s) exerted a pressure on 2-in.- (50-mm-) diameter pipes that was less than 50 percent
of their design pressure.
Given the above results, excluding the energy cost for the pump power, the maximum water velocity
in certain short sections of a water system may be raised to an upper limit of 11 ft/s (3.35 m/s)
for a special purpose, such as enhancing the heat-transfer coefficients.
Normally, water flow in coils and heat exchangers becomes laminar and seriously impairs the
heat-transfer characteristics only when its velocity drops to a value less than 2 ft / s (0.61 m/s) and
its corresponding Reynolds number is reduced to about 10,000 (within the transition region). In
evaporators, Redden (1996) found that at low tube water velocity at 1.15 ft / s (0.35 m/s) and low
heat flux, instability of heat transfer occurred and caused the chilled water leaving temperature to
fluctuate by 4F (2.2C). In condensers, condensing operation is not affected even if the condenser
water velocity in the tube was about 1 ft / s (0.31 m/ s). Water velocity should also be maintained at
not less than 2 ft / s (0.61 m/ s) in order to transport the entrained air to air vents.
When pipes are being sized, the optimum pressure drop , commonly expressed in feet
(meters) of head loss of water per 100 ft of pipe length (p in pascals of pressure drop per meter
length), is a compromise between energy costs and investments. At the same time, the agecorrosion
of pipes should be considered.
Generally, the pressure drop for water pipes inside buildings is in a range of 1 ft / 100 ft to
4 ft /100 ft (100 to 400 Pa/m), with a mean of 2.5 ft /100 ft (250 Pa/m) used most often. Because of
a lower increase in installation cost for smaller-diameter pipes, it may be best to use a pressure
drop lower than 2.5 ft /100 ft (250 Pa/m) when the pipe diameter is 2 in. or less.
Age corrosion results in an increase in the friction factor and a decrease in the effective diameter.
The factors that contribute to age corrosion are sliming, caking of calcareous salts, and corrosion.
Many scientists recommended an increase in friction loss of 15 to 20 percent, resulting in a design
pressure drop of 2 ft/100 ft (200 Pa/m), for closed water systems; and a 75 to 90 percent increase in
friction loss, or a design pressure drop of 1.35 ft/100 ft (135 Pa/m), for open water systems.
Figures 7.2, 7.3, and 7.4 show the pressure drop charts for steel, copper, and plastic pipes, respectively,
for closed water systems. Each chart shows the volume flow (gpm), pressure drop
(ft / 100 ft), water velocity vw (ft /s), and water pipe diameter D (in.). Given any two of these
parameters, the other two can be determined. For instance, for a steel water pipe that has a water
volume flow of 1000 gpm, if the pressure drop is 2 ft / 100 ft, the diameter is 8 in. and the corresponding
velocity is about 8 ft / s.
Hf
V?
gal
Hf
Hf
3
8
Tw
Tw
Tw
Tw
Tw
It is a common practice to limit the water velocity to no more than 4 ft/ s (1.2 m/s) for water
pipes 2 in. (50 mm) or less in diameter in order to prevent an excessive Hf. The pressure drop
should not exceed 4 ft /100 ft (400 Pa/m) for water pipes of greater than 2-in. (50-mm) diameter.
An open water system or a closed water system that is connected with an open expansion tank,
all the pressure differences between two points or levels, and pressure drops across a piece of
equipment or a device are expressed in feet of water column or psi (head in meters of water column
or pressure loss in kPa). The total or static pressure of water at a certain point in a water system is
actually measured and expressed by that part of pressure which is greater or smaller than the atmospheric
pressure, often called gauge pressure, in feet of water column gauge or psig (meters gauge or
kPa  g). The relationships between the steady flow energy equation and the fluid pressure, and between
the pressure loss and fluid head, are discussed in Sec.17.1.
7.6 CHAPTER SEVEN
FIGURE 7.2 Friction chart for water in steel pipes (Schedule 40). (Source: ASHRAE Handbook 1989 Fundamentals. Reprinted with
permission.)
FIGURE 7.3 Friction chart for water in copper tubing (types K, L, and M). (Source: ASHRAE Handbook 1989 Fundamentals.
Reprinted with permission.)
7.2 WATER PIPING
Piping Material
For water systems, the piping materials most widely used are steel, both black (plain) and galvanized
(zinc-coated), in the form of either welded-seam steel pipe or seamless steel pipe; ductile iron
and cast iron; hard copper; and polyvinyl chloride (PVC). The piping materials for various services
are shown below:
WATER SYSTEMS 7.7
FIGURE 7.4 Friction chart for water in plastic pipes (Schedule 80). (Source: ASHRAE Handbook 1989 Fundamentals. Reprinted with
permission.)
Black and galvanized steel
Black steel, hard copper
Black steel, galvanized ductile iron, PVC
Chilled water
Hot water
Cooling water and drains
Copper, galvanized steel, galvanized ductile iron, and PVC pipes have better corrosion resistance
than black steel pipes. Technical requirements, as well as local customs, determine the selection of
piping materials.
Piping Dimensions
The steel pipe wall thicknesses currently used were standardized in 1930. The thickness ranges
from Schedule 10, light wall, to Schedule 160, very heavy wall. Schedule 40 is the standard for a
pipe with a diameter up to 10 in. (250 mm). For instance, a 2-in. (50-mm) standard pipe has an outside
diameter of 2.375 in. (60.33 mm) and an inside diameter of 2.067 in. (52.50 mm). The nominal
pipe size is only an approximate indication of pipe size, especially for pipes of small diameter.
Table 7.1 lists the dimensions of commonly used steel pipes.
The outside diameter of extruded copper is standardized so that the outside diameter of the copper
tubing is 1/8 in. (3.2 mm) larger than the nominal size used for soldered or brazed socket joints.
As in the case with steel pipes, the result is that the inside diameters of copper tubes seldom equal
the nominal sizes. Types K, L, M, and DWV designate the wall thickness of copper tubes: type K is
the heaviest, and DWV is the lightest. Type L is generally used as the standard for pressure copper
tubing. Type DWV is used for drainage at atmospheric pressure.
7.8
TABLE 7.1 Dimensions of Commonly Used Steel Pipes
Working pressure
Nominal Surface area Cross-sectional Weight of ASTM A538B to 400F
size and Schedule Wall Inside Metal Flow
pipe OD number thickness diameter Outside, Inside, area, area, Pipe, Water, Mfr. Joint
D, in. or weight* t, in. d, in. ft2/ft ft2/ft in.2 in.2 lb/ft lb/ft process type psig
40 ST 0.088 0.364 0.141 0.095 0.125 0.104 0.424 0.045 CW Thrd 188
D  0.540 80 XS 0.119 0.302 0.141 0.079 0.157 0.072 0.535 0.031 CW Thrd 871
40 ST 0.091 0.493 0.177 0.129 0.167 0.191 0.567 0.083 CW Thrd 203
D  0.675 80 XS 0.126 0.423 0.177 0.111 0.217 0.141 0.738 0.061 CW Thrd 820
40 ST 0.109 0.622 0.220 0.163 0.250 0.304 0.850 0.131 CW Thrd 214
D  0.840 80 XS 0.147 0.546 0.220 0.143 0.320 0.234 1.087 0.101 CW Thrd 753
40 ST 0.113 0.824 0.275 0.216 0.333 0.533 1.13 0.231 CW Thrd 217
D  1.050 80 XS 0.154 0.742 0.275 0.194 0.433 0.432 1.47 0.187 CW Thrd 681
1 40 ST 0.133 1.049 0.344 0.275 0.494 0.864 1.68 0.374 CW Thrd 226
D  1.315 80 XS 0.179 0.957 0.344 0.251 0.639 0.719 2.17 0.311 CW Thrd 642
1 40 ST 0.140 1.380 0.435 0.361 0.669 1.50 2.27 0.647 CW Thrd 229
D  1.660 80 XS 0.191 1.278 0.435 0.335 0.881 1.28 2.99 0.555 CW Thrd 594
1 40 ST 0.145 1.610 0.497 0.421 0.799 2.04 2.72 0.881 CW Thrd 231
D  1.900 80 XS 0.200 1.500 0.497 0.393 1.068 1.77 3.63 0.765 CW Thrd 576
2 40 ST 0.154 2.067 0.622 0.541 1.07 3.36 3.65 1.45 CW Thrd 230
D  2.375 80 XS 0.218 1.939 0.622 0.508 1.48 2.95 5.02 1.28 CW Thrd 551
2 40 ST 0.203 2.469 0.753 0.646 1.70 4.79 5.79 2.07 CW Weld 533
D  2.875 80 XS 0.276 2.323 0.753 0.608 2.25 4.24 7.66 1.83 CW Weld 835
3 40 ST 0.216 3.068 0.916 0.803 2.23 7.39 7.57 3.20 CW Weld 482
D  3.500 80 XS 0.300 2.900 0.916 0.759 3.02 6.60 10.25 2.86 CW Weld 767
4 40 ST 0.237 4.026 1.178 1.054 3.17 12.73 10.78 5.51 CW Weld 430
D  4.500 80 XS 0.337 3.826 1.178 1.002 4.41 11.50 14.97 4.98 CW Weld 695
6 40 ST 0.280 6.065 1.734 1.588 5.58 28.89 18.96 12.50 ERW Weld 696
D  6.625 80 XS 0.432 5.761 1.734 1.508 8.40 26.07 28.55 11.28 ERW Weld 1209
1
2
1
2
1
4
3
4
1
2
3
8
1
4
7.9
8 30 0.277 8.071 2.258 2.113 7.26 51.16 24.68 22.14 ERW Weld 526
D  8.625 40 ST 0.322 7.981 2.258 2.089 8.40 50.03 28.53 21.65 ERW Weld 643
80 XS 0.500 7.625 2.258 1.996 12.76 45.66 43.35 19.76 ERW Weld 1106
30 0.307 10.136 2.814 2.654 10.07 80.69 34.21 34.92 ERW Weld 485
10 40 ST 0.365 10.020 2.814 2.623 11.91 78.85 40.45 34.12 ERW Weld 606
D  10.75 XS 0.500 9.750 2.814 2.552 16.10 74.66 54.69 32.31 ERW Weld 887
80 0.593 9.564 2.814 2.504 18.92 71.84 64.28 31.09 ERW Weld 1081
30 0.330 12.090 3.338 3.165 12.88 114.8 43.74 49.68 ERW Weld 449
12 ST 0.375 12.000 3.338 3.141 14.58 113.1 49.52 48.94 ERW Weld 528
D  12.75 40 0.406 11.938 3.338 3.125 15.74 111.9 53.48 48.44 ERW Weld 583
XS 0.500 11.750 3.338 3.076 19.24 108.4 65.37 46.92 ERW Weld 748
80 0.687 11.376 3.338 2.978 26.03 101.6 88.44 43.98 ERW Weld 1076
30 ST 0.375 13.250 3.665 3.469 16.05 137.9 54.53 59.67 ERW Weld 481
14 40 0.437 13.126 3.665 3.436 18.62 135.3 63.25 58.56 ERW Weld 580
D  14.00 XS 0.500 13.000 3.665 3.403 21.21 132.7 72.04 57.44 ERW Weld 681
80 0.750 12.500 3.665 3.272 31.22 122.7 106.05 53.11 ERW Weld 1081
16 30 ST 0.375 15.250 4.189 3.992 18.41 182.6 62.53 79.04 ERW Weld 421
D  16.00 40 XS 0.500 15.000 4.189 3.927 24.35 176.7 82.71 76.47 ERW Weld 596
ST 0.375 17.250 4.712 4.516 20.76 233.7 70.54 101.13 ERW Weld 374
18 30 0.437 17.126 4.712 4.483 24.11 230.3 81.91 99.68 ERW Weld 451
D  18.00 XS 0.500 17.000 4.712 4.450 27.49 227.0 93.38 98.22 ERW Weld 530
40 0.562 16.876 4.712 4.418 30.79 223.7 104.59 96.80 ERW Weld 607
20 ST 0.375 19.250 5.236 5.039 23.12 291.0 78.54 125.94 ERW Weld 337
D  20.00 30 XS 0.500 19.000 5.236 4.974 30.63 283.5 104.05 122.69 ERW Weld 477
40 0.593 18.814 5.236 4.925 36.15 278.0 122.82 120.30 ERW Weld 581
*Numbers are schedule number per ASTM B36.10; ST  standard weight; XS  extra strong.
Working pressures have been calculated per ASME/ANSI B31.9 using furnace butt weld (continuous weld, CW) pipe through 4 in. and electric resistance weld (ERW)
thereafter. The allowance A has been taken as (a) 12.5 percent of t for mill tolerance on pipe wall thickness, plus (b) an arbitrary corrosion allowance of 0.025 in. for pipe sizes
through NPS 2 and 0.065 in. from NPS 2 through 20 plus (c) a thread cutting allowance for sizes through NPS 2.
Because the pipe wall thickness of threaded standard weight pipe is so small after deducting the allowance A, the mechanical strength of the pipe is impaired. It is good practice
to limit standard-weight threaded pipe pressures to 90 psig for steam and 125 psig for water.
Source: ASHRAE Handbook 1988, Equipment. Reprinted with permission.
1
2
TABLE 7.1 Dimensions of Commonly Used Steel Pipes (Continued)
Working pressure
Nominal Surface area Cross-sectional Weight of ASTM A538B to 400F
size and Schedule Wall Inside Metal Flow
pipe OD number thickness diameter Outside, Inside, area, area, Pipe, Water, Mfr. Joint
D, in. or weight* t, in. d, in. ft2/ft ft2/ft in.2 in.2 lb/ft lb/ft process type psig
7.10
TABLE 7.2 Dimensions of Copper Tubes
Working pressure*
Surface area Cross-sectional Weight of ASTMB 888 to 250F
Nominal Wall Outside Inside
diameter, thickness diameter diameter Outside, Inside, Metal Flow Tube, Water, Annealed, Drawn,
in. Type t, in. D, in. d, in. ft2/ft ft2/ft area, in.2 area, in.2 lb/ft lb/ft psig psig
K 0.035 0.375 0.305 0.098 0.080 0.037 0.073 0.145 0.032 851 1596
L 0.030 0.375 0.315 0.098 0.082 0.033 0.078 0.126 0.034 730 1368
K 0.049 0.500 0.402 0.131 0.105 0.069 0.127 0.269 0.055 894 1676
L 0.035 0.500 0.430 0.131 0.113 0.051 0.145 0.198 0.063 638 1197
M 0.025 0.500 0.450 0.131 0.008 0.037 0.159 0.145 0.069 456 855
K 0.049 0.625 0.527 0.164 0.138 0.089 0.218 0.344 0.094 715 1341
L 0.040 0.625 0.545 0.164 0.143 0.074 0.233 0.285 0.101 584 1094
M 0.028 0.625 0.569 0.164 0.149 0.053 0.254 0.203 0.110 409 766
K 0.049 0.750 0.652 0.196 0.171 0.108 0.334 0.418 0.144 596 1117
L 0.042 0.750 0.666 0.196 0.174 0.093 0.348 0.362 0.151 511 958
K 0.065 0.875 0.745 0.229 0.195 0.165 0.436 0.641 0.189 677 1270
L 0.045 0.875 0.785 0.229 0.206 0.117 0.484 0.455 0.209 469 879
M 0.032 0.875 0.811 0.229 0.212 0.085 0.517 0.328 0.224 334 625
1 K 0.065 1.125 0.995 0.295 0.260 0.216 0.778 0.839 0.336 527 988
L 0.050 1.125 1.025 0.295 0.268 0.169 0.825 0.654 0.357 405 760
M 0.035 1.125 1.055 0.295 0.276 0.120 0.874 0.464 0.378 284 532
1 K 0.065 1.375 1.245 0.360 0.326 0.268 1.217 1.037 0.527 431 808
L 0.055 1.375 1.265 0.360 0.331 0.228 1.257 0.884 0.544 365 684
M 0.042 1.375 1.291 0.360 0.338 0.176 1.309 0.682 0.566 279 522
DWV 0.040 1.375 1.295 0.360 0.339 0.168 1.317 0.650 0.570 265 497
1 K 0.072 1.625 1.481 0.425 0.388 0.351 1.723 1.361 0.745 404 758
L 0.060 1.625 1.505 0.425 0.394 0.295 1.779 1.143 0.770 337 631
M 0.049 1.625 1.527 0.425 0.400 0.243 1.831 0.940 0.792 275 516
DWV 0.042 1.625 1.541 0.425 0.403 0.209 1.865 0.809 0.807 236 442
2 K 0.083 2.125 1.959 0.556 0.513 0.532 3.014 2.063 1.304 356 668
L 0.070 2.125 1.985 0.556 0.520 0.452 3.095 1.751 1.339 300 573
M 0.058 2.125 2.009 0.556 0.526 0.377 3.170 1.459 1.372 249 467
DWV 0.042 2.125 2.041 0.556 0.534 0.275 3.272 1.065 1.416 180 338
1
2
1
4
3
4
5
8
1
2
3
8
1
4
7.11
2 K 0.095 2.625 2.435 0.687 0.637 0.755 4.657 2.926 2.015 330 619
L 0.080 2.625 2.465 0.687 0.645 0.640 4.772 2.479 2.065 278 521
M 0.065 2.625 2.495 0.687 0.653 0.523 4.889 2.026 2.116 226 423
3 K 0.109 3.125 2.907 0.818 0.761 1.033 6.637 4.002 2.872 318 596
L 0.090 3.125 2.945 0.818 0.771 0.858 6.812 3.325 2.947 263 492
M 0.072 3.125 2.981 0.818 0.780 0.691 6.979 2.676 3.020 210 394
DWV 0.045 3.125 3.035 0.818 0.795 0.435 7.234 1.687 3.130 131 246
3 K 0.120 3.625 3.385 0.949 0.886 1.321 8.999 5.120 3.894 302 566
L 0.100 3.625 3.425 0.949 0.897 1.107 9.213 4.291 3.987 252 472
M 0.083 3.625 3.459 0.949 0.906 0.924 9.397 3.579 4.066 209 392
4 K 0.134 4.125 3.857 1.080 1.010 1.680 11.684 6.510 5.056 296 555
L 0.110 4.125 3.905 1.080 1.022 1.387 11.977 5.377 5.182 243 456
M 0.095 4.125 3.935 1.080 1.030 1.203 12.161 4.661 5.262 210 394
DWV 0.058 4.125 4.009 1.080 1.050 0.741 12.623 2.872 5.462 128 240
5 K 0.160 5.125 4.805 1.342 1.258 2.496 18.133 9.671 7.846 285 534
L 0.125 5.125 4.875 1.342 1.276 1.963 18.665 7.609 8.077 222 417
M 0.109 5.125 4.907 1.342 1.285 1.718 18.911 6.656 8.183 194 364
DWV 0.072 5.125 4.981 1.342 1.304 1.143 19.486 4.429 8.432 128 240
6 K 0.192 6.125 5.741 1.603 1.503 3.579 25.886 13.867 11.201 286 536
L 0.140 6.125 5.845 1.603 1.530 2.632 26.832 10.200 11.610 208 391
M 0.122 6.125 5.881 1.603 1.540 2.301 27.164 8.916 11.754 182 341
DWV 0.083 6.125 5.959 1.603 1.560 1.575 27.889 6.105 12.068 124 232
8 K 0.271 8.125 7.583 2.127 1.985 6.687 45.162 25.911 19.542 304 570
L 0.200 8.125 7.725 2.127 2.022 4.979 46.869 19.295 20.280 224 421
M 0.170 8.125 7.785 2.127 2.038 4.249 47.600 16.463 20.597 191 358
DWV 0.109 8.125 7.907 2.127 2.070 2.745 49.104 10.637 21.247 122 229
10 K 0.338 10.125 9.449 2.651 2.474 10.392 70.123 40.271 30.342 304 571
L 0.250 10.125 9.625 2.651 2.520 7.756 72.760 30.054 31.483 225 422
M 0.212 10.125 9.701 2.651 2.540 6.602 73.913 25.584 31.982 191 358
12 K 0.405 12.125 11.315 3.174 2.962 14.912 100.554 57.784 43.510 305 571
L 0.280 12.125 11.565 3.174 3.028 10.419 105.046 40.375 45.454 211 395
M 0.254 12.125 11.617 3.174 3.041 9.473 105.993 36.706 45.863 191 358
*When using soldered or brazed fittings, the joint determines the limiting pressure.
Working pressures calculated using ASME B31.9 allowable stresses. A 5 percent mill tolerance has been used on the wall thickness. Higher tube ratings can be calculated using allowable
stress for lower temperatures.
If soldered or brazed fittings are used on hard-drawn tubing, use the annealed ratings. Full-tube allowable pressures can be used with suitably rated flare or compression-type fittings.
Source: ASHRAE Handbook 1988, Equipment. Reprinted with permission.
1
2
1
2
TABLE 7.2 Dimensions of Copper Tubes (Continued)
Working pressure*
Surface area Cross-sectional Weight of ASTMB 888 to 250F
Nominal Wall Outside Inside
diameter, thickness diameter diameter Outside, Inside, Metal Flow Tube, Water, Annealed, Drawn,
in. Type t, in. D, in. d, in. ft2/ft ft2/ft area, in.2 area, in.2 lb/ft lb/ft psig psig
Copper tubes are also categorized as hard and soft copper. Soft pipes should be used in applications
for which the pipe will be bent in the field. Table 7.2 lists the dimensions of copper tubes.
Thermoplastic plastic pipes are the most widely used plastic pipes in air conditioning. They are
manufactured with dimensions that match steel pipe dimensions. The advantages of plastic pipes include
resistance to corrosion, scaling, and the growth of algae and fungi. Plastic pipes have smooth
surfaces and negligible age allowance. Age allowance is the allowance for corrosion and scaling for
plastic pipes during their service life. Most plastic pipes are low in cost, especially compared with
corrosion-resistant metal tubes.
The disadvantages of plastic pipes include the fact that their pressure ratings decrease rapidly
when the water temperature rises above 100F (37.8C). PVC pipes are weaker than metal pipes
and must usually be thicker than steel pipes if the same working pressure is to be maintained. Plastic
pipes may experience expansion and contraction during temperature changes that is 4 times
greater than that of steel. Precautions must be taken to protect plastic pipes from external damage
and to account for its behavior during fire. Some local codes do not permit the use of some or allplastic
pipes. It is necessary to check with local authorities.
Pipe Joints
Steel pipes of small diameter (2 in. or 50 mm less) threaded through cast-iron fittings are the most
widely used type of pipe joint. For steel pipes of diameter 2 in. (50 mm) and more, welded joints,
7.12 CHAPTER SEVEN
TABLE 7.3 Maximum Allowable Pressures at Corresponding Temperatures
System
Maximum
allowable
Fitting
pressure at
Temperature, temperature,
Application Pipe material Weight Joint type Class Material F psig
Recirculating water
2 in. and smaller Steel (CW) Standard Thread 125 Cast iron 250 125
Copper, hard Type L 95-5 solder  Wrought copper 250 150
PVC Sch. 80 Solvent Sch. 80 PVC 75 350
CPVC Sch. 80 Solvent Sch. 80 CPVC 150 150
PB SDR-11 Heat fusion  PB 160 115
Insert crimp  Metal 160 115
2.512 in. A53 B ERW Standard Weld Standard Wrought steel 250 400
steel Flange 150 Wrought steel 250 250
Flange 125 Cast iron 250 175
Flange 250 Cast iron 250 400
Groove  MI or ductile iron 230 300
PB SDR-11 Heat fusion PB 160 115
Refrigerant
Copper, hard Type L or K Braze  Wrought copper  
A53 B SML
steel Standard Weld Wrought steel  
Note: Maximum allowable working pressures have been derated in this table. Higher system pressures can be used for lower temperatures and
smaller pipe sizes. Pipe, fittings, joints, and valves must all be considered.
Note: A53 ASTM Standard A53
PVC Polyvinyl chloride
CPVC Chlorinated polyvinyl chloride
PB Polybutylene
Source: Abridged with permission from ASHRAE Handbook 1988, Equipment.
bolted flanges, and grooved ductile iron joined fittings are often used. Galvanized steel pipes are
threaded together by galvanized cast iron or ductile iron fittings.
Copper pipes are usually joined by soldering and brazing socket end fittings. Plastic pipes are
often joined by solvent welding, fusion welding, screw joints, or bolted flanges.
Vibrations from pumps, chillers, or cooling towers can be isolated or dampened by means of
flexible pipe couplings. Arch connectors are usually constructed of nylon, dacron, or polyester and
neoprene. They can accommodate deflections or dampen vibrations in all directions. Restraining
rods and plates are required to prevent excessive stretching. A flexible metal hose connector
includes a corrugated inner core with a braided cover. It is available with flanged or grooved
end joints.
Working Pressure and Temperature
In a water system, the maximum allowable working pressure and temperature are not limited to the
pipes only; joints or the pipe fittings, especially valves, may often be the weak links. Table 7.3 lists
types of pipes, joint, and fittings and their maximum allowable working pressures for specified
temperatures.
Expansion and Contraction
During temperature changes, all pipes expand and contract. The design of water pipes must take
into consideration this expansion and contraction. Both the temperature change during the operating
period and the possible temperature change between the operating and shutdown periods should
WATER SYSTEMS 7.13
FIGURE 7.5 Expansion loops. (a) U bends; (b) L bends; (c) Z bends.
also be considered. For chilled and condenser water, which has a possible temperature change of 40
to 100F (4.4 to 37.8C), expansion and contraction cause considerable movement in a long run of
piping. Unexpected expansion and contraction cause excess stress and possible failure of the pipe,
pipe support, pipe joints, and fittings.
Expansion and contraction of hot and chilled water pipes can be better accommodated by using
loops and bends. The commonly used bends are U bends, Z bends, and L bends, as shown in
Fig. 7.5. Anchors are the points where the pipe is fixed so that it will expand or contract between
them. Reaction forces at these anchors should be considered when the support is being designed.
ASHRAE Handbook 1992, HVAC Systems and Equipment, gives the required calculations and data
for determining these stresses. Guides are used so that the pipes expand laterally.
Empirical formulas are often used instead of detailed stress analyses to determine the dimension
of the offset leg Lo [ft (m)]. Waller (1990) recommended the following formulas:
U bends: Lo  0.041D0.48Lac
0.46 T
Z bends: Lo  (0.13DLacT)0.5 (7.2)
L bends: Lo  (0.314DLacT)0.5
where D  diameter of pipe, in. (mm)
Lac  distance between anchors, hundreds of ft (m)
T  temperature difference, F (C)
If there is no room to accommodate U, Z, or L bends (such as in high-rise buildings or tunnels),
mechanical expansion joints are used to compensate for movement during expansion. Packed
expansion joints allow the pipe to slide to accommodate movement during expansion. Various types
of packing are used to seal the sliding surfaces in order to prevent leakage. Another type of mechanical
joint uses bellows or flexible metal hose to accommodate movement. These types of joints
should be carefully installed to avoid distortion.
7.14 CHAPTER SEVEN
TABLE 7.4 Recommended Pipe Hanger Spacing, ft
Standard-weight
Nominal pipe steel pipe Copper Rod size,
diameter, in. (water) tube (water) in.
7 5
7 5
1 7 6
1 9 8
2 10 8
2 11 9
3 12 10
4 14 12
6 17 14
8 19 16
10 20 18
12 23 19
14 25 1
16 27 1
18 28 1
20 30 1
Note: Spacing does not apply where concentrated loads are placed between
supports such as flanges, valves, and specialties.
Source: ASHRAE Handbook 1988, Equipment. Reprinted with permission.
1
4
1
4
7
8
3
4
5
8
1
2
1
2
3
8
3
8
1
2
3
8
3
8
1
2
1
4
1
4
3
4
1
4
1
2
Piping Supports
Types of piping support include hangers, which hang the pipe from above; supports, which usually
use brackets to support the pipe from below; anchors to control the movement of the piping; and
guides to guide the axial movement of the piping. Table 7.4 lists the recommended spacing of pipe
hangers. Piping support members should be constructed based on the stress at their point of connection
to the pipe as well as on the characteristics of the structural system. Pipe supports must have
sufficient strength to support the pipe, including the water inside. Except for the anchors, they
should also allow for expansion movement.
Pipes should be supported around the connections to the equipment so that the pipes weight and
expansion or contraction do not affect the equipment. For insulated pipes, heavy-gauge sheet-metal
half-sleeves are used between the hangers and the insulation. Corrosion protection should also be
carefully considered.
Piping Insulation
External pipe insulation should be provided when the inside water temperature 105F  Tw  60F
(41C  Tw  15.6C) for the sake of energy saving, surface condensation, and high-temperature
safety protection. The optimum thickness of the insulation of pipes depends mainly on the operating
temperature of the inside water, the pipe diameter, and the types of service. There is a compromise between
initial cost and energy cost. ASHRAE/IESNA Standard 90.1-1999 specifies the minimum pipe
insulation thickness for water systems, as listed in Table 7.5. Insulation shall be protected from damage
including that because of sunlight, moisture, equipment maintenance, and wind.
WATER SYSTEMS 7.15
TABLE 7.5 Minimum Pipe Insulation Thickness*, in.
Fluid design
Insulation conductivity
Nominal pipe or tube size, in.
operating temperature Conductivity, Mean rating
range, F Btuin./h  ft2F temp. F <1 1 to <1 1 to <4 4 to <8 ?8
Heating systems (steam, steam condensate, and hot water)
>350 0.320.34 250 2.5 3.0 3.0 4.0 4.0
251350 0.290.32 200 1.5 2.5 3.0 3.0 3.0
201250 0.270.30 150 1.5 1.5 2.0 2.0 2.0
141200 0.250.29 125 1.0 1.0 1.0 1.5 1.5
105140 0.220.28 100 0.5 0.5 1.0 1.0 1.0
Domestic and service hot water systems
105+ 0.220.28 100 0.5 0.5 1.0 1.0 1.0
Cooling systems (chilled water, brine, and refrigerant)
4060 0.220.28 100 0.5 0.5 1.0 1.0 1.0
<40 0.220.28 100 0.5 1.0 1.0 1.0 1.5
*For insulation outside the stated conductivity range, the minimum thickness T shall be determined as follows:
where T  minimum insulation thickness (in.), r  actual outside radius of pipe (in.), t  insulation thickness listed in this table
for applicable fluid temperature and pipe size, K  conductivity of alternate material at mean rating temperature indicated for
the applicable fluid temperature (Btu in.[h ft2F]); and k  the upper value of the conductivity range listed in this table for the
applicable fluid temperature.
These thicknesses are based on energy efficiency considerations only. Additional insulation is sometimes required relative
to safety issues/surface temperature.
Piping insulation is not required between the control valve and coil on run-outs when the control valve is located within 4
ft of the coil and the pipe size is 1 in. or less.
These thicknesses are based on energy efficiency considerations only. Issues such as water vapor permeability or surface
condensation sometimes require vapor retarders or additional insulation.
Source: ASHRAE/IESNA Standard 90.1-1999. Reprinted with permission.
T  r[(1 	 t/r)K/k  1]
1
2
1
2
Insulation exposed to weather shall be suitable for outdoor service, such as, protected by aluminum,
sheet metal, painted canvas, or plastic cover. Cellular foam insulation shall be protected as
above or with a painted coating which itself is a water retardant and also provides shielding from
solar radiation.
Insulation of chilled water piping or refrigerant suction piping shall include an exterior vapor retardant
covering the insulation (unless the insulation is inherently vapor retardant). All penetrations
and joints of the vapor retardant shall be sealed.
7.3 VALVES, PIPE FITTINGS, AND ACCESSORIES
Types of Valve
Valves are used to regulate or stop the water flow in pipes either manually or by means of automatic
control systems. Valves used in automatic control systems are called control valves, discussed in
Chap. 5. In this section, only manually operated valves, or simply valves, are discussed.
Hand-operated valves are used to stop or isolate flow, to regulate flow, to prevent reverse flow,
and to regulate water pressure. The basic construction of a valve consists of the following (see
Fig. 7.6): a disk to open or close the water flow; a valve body to seat the disk and provide the
flow passage; a stem to lift or rotate the disk, with a handwheel or a handle and corresponding
mechanism to make the task easier; and a bonnet to enclose the valve from the top.
Based on the shape of the valve disk, the valve body, or its function, commonly used valves can
be classified into the following types:
Gate Valves. The disk of a gate valve is in the shape of a gate or wedge, as shown in Fig. 7.6a.
When the wedge is raised at the open position, a gate valve does not add much flow resistance. The
wedge can be either a solid wedge, which is most commonly used, or a split wedge, in which two
disk halves being forced outward fit tightly against the body seat. Gate valves are used either fully
opened or closed, an on/off arrangement. They are often used as isolating valves for pieces of
equipment or key components, such as control valves, for service during maintenance and repair.
Globe Valves. They are so named because of the globular shape of the valve body, as shown in
Fig. 7.6b. Globe valves have a round disk or plug-type disk seated against a round port. Water flow
enters under the disk. Globe valves have high flow resistances. They can be opened or closed
7.16 CHAPTER SEVEN
FIGURE 7.6 Types of valves. (a) Gate valve; (b) Globe valve; (c) Check valve, swing check.
substantially faster than gate valves. Angle valves are similar to globe valves in their seats and
operation. The basic difference is that the valve body of an angle valve can also be used as a 90
elbow at that location.
Globe valves are used to throttle and to regulate the flow. They are sometimes called balancing
valves. They are deliberately designed to restrict fluid flow, so they should not be used in applications
for which full and unobstructed flow is often required.
Check Valves. Check valves, as their name suggests, are valves used to prevent, or check, reverse
flow. There are basically two types of check valves: swing check and lift check. A swing check
valve has a hinged disk, as shown in Fig. 7.6c. When the water flow reverses, water pressure pushes
the disk and closes the valve. In a lift check valve, upward regular flow raises the disk and opens the
valve, and reverse flow pushes the disk down to its seat and stops the backflow. A swing check
valve has a lower flow resistance than a lift check valve.
Plug Valves. These valves use a tapered, cylindrical plug disk to fit the seat. They vary from fully
open to fully closed positions within a quarter-turn. Plug valves may be used for throttling control
during the balancing of a water system.
Ball Valves. These valves use a ball as the valve disk to open or close the valve. As with plug
valves, they vary from fully open to fully closed positions within a quarter-turn. As with gate valves,
ball valves are usually used for open/ close service. They are less expensive than gate valves.
Butterfly Valves. A butterfly valve has a thin rotating disk. Like a ball or plug valve, it varies
within a quarter-turn from fully open to fully closed. As described in Sec. 5.6, a butterfly valve
exhibits low flow resistance when it is fully opened. The difference between a butterfly valve
used for control purposes and a hand-operated butterfly valve is that the former has an actuator
and can be operated automatically. Butterfly valves are lightweight, easy to operate and install,
and lower in cost than gate valves. They are primarily used as fully open or fully closed, but they
may be used for throttling purposes. Butterfly valves are gaining in popularity, especially in large
pipes.
Balance Valves. These valves are used to balance the water flow in a water system. There are two
kinds of balancing valves: manual balance valves and automatic balance valves. A globe valve can
be used as a manual balance valve. A manual balance valve can also be a valve with integral pressure
taps for flow measurement and a calibrated port to adjust the flow. An automatic balancing
valve is also called an automatic flow-limiting valve. There is a moving element that adjusts the
flow passage area according to the water pressure differential across the valve.
Pressure Relief Valves. These valves are safety valves used to prevent a system that is overpressurized
from exceeding a predetermined limit. A pressure relief valve is held closed by a spring or
rupture member and is automatically opened to relieve the water pressure when it rises above the
system design working pressure.
Valve Connections and Ratings
The type of connection used between a valve and the pipes is usually consistent with the type of
joint used in the pipe system. A water piping system with flanged joint requires a valve with flanged
ends. The commonly used types of valve connection are as follows:
 Threaded ends. These connections are mainly used for small pipes with diameters from to 2 in.
(6 to 50 mm). Threaded-end valves are usually inexpensive and simpler to install.
1
4
WATER SYSTEMS 7.17
 Flanged ends. These connections are commonly used for larger pipes ( in. or 63 mm and
above). Flanged ends are more easily separated when necessary.
 Welded ends. Steel valves, when used at higher pressure and temperature, are often connected
with welded ends. Welded ends exhibit the fewest instances of leakage.
 Grooved ends. These connections use circumferential grooves in which a rubber gasket fits and
are enclosed by iron couplings. Butterfly valves are often connected with grooved ends.
 Soldered ends. Bronze valves in copper piping systems use soldered ends. Tin-alloy soldering is
the type of soldering commonly used. Lead soldering cannot be used in a potable water system
because it will contaminate the water.
Valves are usually rated according to their ability to withstand pressure at a specified temperature.
Metal valves have two different ratings, one for steam [working steam pressure (WSP)], which
should correspond to its operating temperature, and the other for cold water, oil, or gas (WOG). The
following are the commonly used ratings:
 125 psig (862 kPa g) WSP, 250 psig (1724 kPag) WOG
 150 psig (1034 kPa g) WSP, 300 psig (2068 kPag) WOG
 300 psig (2069 kPa g) WSP, 600 psig (4138 kPag) WOG
Here psig represents gauge pressure in pounds per square inch (kPag gauge pressure in kPa). As
listed in Table 7.3, a wrought steel valve flange joint with a 150 psig (1034 kPag) rating can be used
for a hot or chilled water system with a maximum allowable pressure of 250 psig (1724 kPa g) at a
temperature below 250F (121C), for pipes of diameter between 2.5 and 12 in. (63 and 300 mm).
Valve Materials
Valve materials are selected according to their ability to withstand working pressure and temperature,
their resistance to corrosion, and their relative cost. The most commonly used materials for
valves are as follows:
 Bronze. It has a good corrosive resistance and is easily machined, cast, or forged. Bronze is
widely used for water valves up to a size of 3 in. (75 mm) because of its high cost. For valves
above 3 in. (75 mm), bronze is still often used for sealing elements and stems because it is
machinable and corrosion-resistant.
 Cast iron and ductile iron. These materials are used for pressure-containing parts, flanges, and
glands in valves 2 in. (50 mm) and larger. Ductile iron has a higher tensile strength than cast iron.
 Steel. Forged or cast steel provides a higher tensile strength as well as toughness in the form of resistance
to shock and vibration than do bronze, cast iron, and ductile iron. Steel is used in applications
that require higher strength and toughness than bronze and ductile iron can provide.
 Trim materials. These include the elements and components that are easily worn as well as those
parts that need to be resistant to corrosion, such as the disk, seating elements, and stem. Stainless
steel, stellite (a kind of cobalt-chromium-tungsten alloy), and chromium-molybdenum steel are
often used for trim material in valves.
Pipe Fittings and Water System Accessories
Water pipe fittings include elbows, tees, and valves. Water pipe elbows and tees are often made of
cast iron, ductile iron, or steel. Pressure losses due to the water pipe fittings are usually expressed in
terms of an equivalent length of straight pipe, for the sake of convenience. Table 7.6 allows calculation
of the pressure losses for various types of piping fittings, in terms of an equivalent length of
21
2
7.18 CHAPTER SEVEN
straight pipe. The equivalent length for a fitting can be estimated by multiplying the elbow equivalent
to that fitting by the equivalent length for a 90 elbow.
Water system accessories include drains, strainers, and air vents. Drains should be equipped at
all low points of the system. Arrangements should be made so that a part of the system or individual
components can be drained rather than draining the entire system. A condensate drain pipe is
always required for cooling and dehumidifying coils. Galvanized steel is often used for this
purpose. It is usually piped to a plumbing drain or other suitable location. A condensate drain pipe
should be insulated so as to avoid surface condensation.
Water strainers are often installed before the pumps, control valves, or other components to protect
them from dirt and impurities. Air vents are discussed in the next section.
7.4 WATER SYSTEM PRESSURIZATION AND
THE PRESENCE OF AIR
Water System Pressurization Control
For an open water system, the maximum operating gauge pressure is the pressure at a specific point
in the system where the positive pressure exerted by the water pumps, to overcome the pressure
WATER SYSTEMS 7.19
TABLE 7.6 Pressure Losses for Pipe Fittings and Valves, Expressed in Terms of an Equivalent Length (in ft) of Straight Pipe
Equivalent length, ft of pipe, for 90 elbows
Pipe size, in.
Velocity, ft/s 1 1 1 2 2 3 3 4 5 6 8 10 12
1 1.2 1.7 2.2 3.0 3.5 4.5 5.4 6.7 7.7 8.6 10.5 12.2 15.4 18.7 22.2
2 1.4 1.9 2.5 3.3 3.9 5.1 6.0 7.5 8.6 9.5 11.7 13.7 17.3 20.8 24.8
3 1.5 2.0 2.7 3.6 4.2 5.4 6.4 8.0 9.2 10.2 12.5 14.6 18.4 22.3 26.5
4 1.5 2.1 2.8 3.7 4.4 5.6 6.7 8.3 9.6 10.6 13.1 15.2 19.2 23.2 27.6
5 1.6 2.2 2.9 3.9 4.5 5.9 7.0 8.7 10.0 11.1 13.6 15.8 19.8 24.2 28.8
6 1.7 2.3 3.0 4.0 4.7 6.0 7.2 8.9 10.3 11.4 14.0 16.3 20.5 24.9 29.6
7 1.7 2.3 3.0 4.1 4.8 6.2 7.4 9.1 10.5 11.7 14.3 16.7 21.0 25.5 30.3
8 1.7 2.4 3.1 4.2 4.9 6.3 7.5 9.3 10.8 11.9 14.6 17.1 21.5 26.1 31.0
9 1.8 2.4 3.2 4.3 5.0 6.4 7.7 9.5 11.0 12.2 14.9 17.4 21.9 26.6 31.6
10 1.8 2.5 3.2 4.3 5.1 6.5 7.8 9.7 11.2 12.4 15.2 17.7 22.2 27.0 32.0
Iron and copper elbow equivalents
Source: ASHRAE Handbook 1997, Fundamentals. Reprinted with permission.
Fitting Iron pipe Copper tubing
Elbow, 90 1.0 1.0
Elbow, 45 0.7 0.7
Elbow, 90 long turn 0.5 0.5
Elbow, 90 welded 0.5 0.5
Reduced coupling 0.4 0.4
Open return bend 1.0 1.0
Angle radiator valve 2.0 3.0
Radiator or convector 3.0 4.0
Boiler or heater 3.0 4.0
Open gate valve 0.5 0.7
Open globe valve 12.0 17.0
1
2
1
2
1
2
1
4
3
4
1
2
drops across the equipment, components, fittings, and pipes plus the static head due to the vertical
distance between the highest water level and that point, is at a maximum.
In a closed chilled or hot water system, a variation in the water temperature will cause an expansion
of water that may raise the water pressure above the maximum allowable pressure. The purposes
of system pressurization control for a closed water system are as follows:
 To limit the pressure of the water system to below its allowable working pressure
 To maintain a pressure higher than the minimum water pressure required to vent air
 To assist in providing a pressure higher than the net positive suction head (NPSH) at the pump
suction to prevent cavitation
 To provide a point of known pressure in the system
Expansion tanks, pressure relief valves, pressure-reducing valves for makeup water, and corresponding
controls are used to achieve water system pressurization control. There are two types of
expansion tanks for closed water systems: open and closed.
Open Expansion Tank
An expansion tank is a device that allows for the expansion and contraction of water contained in a
closed water system when the water temperature changes between two predetermined limits. Another
function of an expansion tank is to provide a point of known pressure in a water system.
An open expansion tank is vented to the atmosphere and is located at least 3 ft (0.91 m) above
the highest point of the water system, as shown in Fig. 7.7. Makeup water is supplied through a
7.20 CHAPTER SEVEN
FIGURE 7.7 Open expansion tank.
float valve, and an internal overflow drain is always installed. A float valve is a globe or ball valve
connected with a float ball to regulate the makeup water flow according to the liquid level in the
tank. An open expansion tank is often connected to the suction side of the water pump to prevent
the water pressure in the system from dropping below the atmospheric pressure. The pressure of the
liquid level in the open tank is equal to the atmospheric pressure, which thus provides a reference
point of known pressure to determine the water pressure at any point in the water system. The minimum
tank volume should be at least 6 percent of the volume of water in the system Vs, ft3 (m3). An
open expansion tank is simple, more stable in terms of system pressure characteristics, and low in
cost. If it is installed indoors, it often needs a high ceiling. If it is installed outdoors, water must be
prevented from freezing in the tank, air vent, or pipes connected to the tank when the outdoor temperature
is below 32F (0C). Because the water surface in the tank is exposed to the atmosphere,
oxygen is more easily absorbed into the water, which makes the tank less resistant to corrosion than
a diaphragm tank (to be described later). Because of these disadvantages, an open expansion tank
has only limited applications.
Closed Expansion Tank
A closed expansion tank is an airtight tank filled with air or other gases, as shown in Fig. 7.8. When
the temperature of the water increases, the water volume expands. Excess water then enters the
tank. The air in the tank is compressed, which raises the system pressure. When the water temperature
drops, the water volume contracts, resulting in a reduction of the system pressure.
To reduce the amount of air dissolved in the water so as to prevent corrosion and prevent air
noise, a diaphragm, or a bladder, is often installed in the closed expansion tank to separate the filled
air and the water permanently. Such an expansion tank is called a diaphragm, or bladder, expansion
tank. Thus, a closed expansion tank is either a plain closed expansion tank, which does not have a
diaphragm to separate air and water, or a diaphragm tank.
For a water system with only one air-filled space, the junction between the closed expansion
tank and the water system is a point of fixed pressure. At this point, water pressure remains
constant whether or not the pump is operating because the filled air pressure depends on only the
volume of water in the system. The pressure at this point can be determined according to the ideal
gas law, as given by Eq. (2.1): pv  RTR. The pressure in a closed expansion tank during the initial
filling process or at the minimum operating pressure is called the fill pressure pfil, psia. The fill
pressure is often used as the reference pressure to determine the pressure characteristics of a water
system.
Size of Diaphragm Expansion Tank
If a closed expansion tank with its filled volume of air is too small, the system pressure will easily
exceed the maximum allowable pressure and cause water to discharge from the pressure relief
valve, thus wasting water. If the closed tank is too large, when the water temperature drops, the system
pressure may decrease to a level below the minimum allowable value and cause trouble in the
air vent. Therefore, accurate sizing of a closed expansion tank is essential.
For diaphragm expansion tanks, the minimum volume of the water tank, Vt, gal (m3), can be calculated
by the following formula, recommended by ASHRAE Handbook 1996, HVAC Systems and
Equipment:
(7.3)
where T1  lower temperature, F (C)
T2  higher temperature, F (C)
Vs  volume of water in system, gal (m3 )
p1  absolute pressure at lower temperature, psia (kPa abs.)
Vt  Vs  v2 / v1  1  3(T2  T1)
1  p1 / p2 
WATER SYSTEMS 7.21
7.22 CHAPTER SEVEN
FIGURE 7.8 Closed expansion tank for a water system. (a) Diaphragm expansion tank in a chilled water system.
(b) Diaphragm expansion tank in a hot water system. (c) Plain closed expansion tank.
p2  absolute pressure at higher temperature, psia (kPa abs.)
v1, v2  specific volume of water at lower and higher temperature, respectively, ft3 /lb (m3 /kg)
 linear coefficient of thermal expansion; for steel,  6.5  106 in. / in F (1.2 
105 per C); for copper,  9.5  106 in. / in. F (1.7  105 per C)
In a chilled water system, the higher temperature T2 is the highest anticipated ambient temperature
when the chilled water system shuts down during summer. The lower temperature in a heating
system is often the ambient temperature at fill conditions (for example, 50F or 10C).
Pump Location
The location of the pump in a water system that uses a diaphragm expansion tank should be
arranged so that the pressure at any point in the water system is greater than the atmospheric pressure.
In such an arrangement, air does not leak into the system, and the required net positive suction
head (NPSH) can be maintained at the suction inlet of the water pump. NPSH is discussed in
detail in Sec. 7.7.
A water pump location commonly used for hot water systems with diaphragm expansion tanks
is just after the expansion tank and the boiler, as shown in Fig. 7.8b. In this arrangement, the pressure
at the pump suction is the sum of the water pressure and the fill pressure. In another often-used
arrangement, the diaphragm expansion tank is moved to the highest point of the water system, and
the pump is still located after the boiler. In a chilled water system, the location of the chilled water
pump is usually before the water chiller, and the diaphragm expansion tank is usually connected to
the suction side of the water pump.
Air in Water Systems
In a closed recirculated water system, air and nitrogen are present in the following forms: dissolved
in water, free air or gas bubbles, or pockets of air or gas. The behavior of air or gas dissolved in liquids
is governed and described by Henrys equation. Henrys equation states that the amount of gas
dissolved in a liquid at constant temperature is directly proportional to the absolute pressure of that
gas acting on the liquid, or
(7.4)
where x  amount of dissolved gas in solution, percent by volume
p  partial pressure of that gas, psia
H  Henrys constant; changes with temperature
The lower the water temperature and the higher the total pressure of the water and dissolved gas,
the greater the maximum amount of dissolved gas at that pressure and temperature.
When the dissolved air or gas in water reaches its maximum amount at that pressure and temperature,
the water becomes saturated. Any excess air or gas, as well as the coexisting water vapor,
can exist only in the form of free bubbles or air pockets. A water velocity greater than 1.5 ft / s
(0.45 m/ s) can carry air bubbles along with water. When water is in contact with air at an air-water
interface, such as the filled airspace in a plain closed expansion tank, the concentration gradient
causes air to diffuse into the water until the water is saturated at that pressure and temperature. An
equilibrium forms between air and water within a certain time. At specific conditions, 24 h may be
required to reach equilibrium.
The oxygen in air that is dissolved in water is unstable. It reacts with steel pipes to form oxides
and corrosion. Therefore, after air has been dissolved in water for a long enough time, only nitrogen
remains as a dissolved gas circulating with the water.
x 
p
H
WATER SYSTEMS 7.23
Penalties due to Presence of Air and Gas
The presence of air and gas in a water system causes the following penalties for a closed water system
with a plain closed expansion tank:
 Presence of air in the terminal and heat exchanger, which reduces the heat-transfer surface
 Corrosion due to the oxygen reacting with the pipes
 Waterlogging in plain closed expansion tanks
 Unstable system pressure
 Poor pump performance due to gas bubbles
 Noise problems
There are two sources of air and gas in a water system. One is the air-water interface in a plain
closed expansion tank or in an open expansion tank, and the other is the dissolved air in a city water
supply.
Oxidation and Waterlogging
Consider a chilled water system that uses a plain closed expansion tank without a diaphragm, as
shown in Fig. 7.8. This expansion tank is located in a basement, with a water pressure of 90 psig
(620 kPa  g) and a temperature of 60F (15.6C) at point A. At such a temperature and pressure, the
solubility of air in water is about 14.2 percent. The chilled water flows through the water pump, the
chiller, and the riser and is supplied to the upper-level terminals. During this transport process, part
of the oxygen dissolved in the water reacts with the steel pipes to form oxides and corrosion. At
upper-level point B, the water pressure is only 10 psig (69 kPa g) at a chilled water temperature of
about 60F (15.6C). At this point, the solubility of air in water is only about 3.3 percent. The
difference in solubility between point A and B is 14.23.310.9 percent. This portion of air,
containing a higher percentage of nitrogen because of the formation of oxides, is no longer
dissolved in the chilled water, but is released from the water and forms free air, gas bubbles, or
pockets. Some of the air pockets are vented through air vents at the terminals, or high points of the
water system. The chilled water returns to point A again and absorbs air from the air-water interface
in the plain closed expansion tank, creating an air solubility in water of about 14.2 percent. Of
course, the actual process is more complicated because of the formation of oxides and the presence
of water vapor.
Such a chilled water recirculating process causes the following problems:
 Oxidation occurs because of the reaction between dissolved oxygen and steel pipes, causing corrosion
during the chilled water transport and recirculating process.
 The air pockets vented at high levels originally come from the filled air in the plain closed expansion
tank; after a period of recirculation of the chilled water, part of the air charge is removed to
the upper levels and vented. The tank finally waterlogs and must be charged with compressed air
again. Waterlogging also results in an unstable system pressure because the amount of filled air in
the plain closed expansion tank does not remain constant. Oxidation and water logging also exist
in hot water systems, but the problems are not as pronounced as in a chilled water system.
Oxidation and waterlogging can be prevented or reduced by installing a diaphragm expansion
tank instead of a plain closed expansion tank. Air vents, either manual or automatic, should be
installed at the highest point of the water system and on coils and terminals at higher levels if a
water velocity of not less than 2 ft / s (0.61 m/ s) is maintained in the pipes, in order to transport the
entrained air bubbles to these air vents.
In a closed chilled water system using a diaphragm expansion tank, there is no air-water interface
in the tank. The 3.3 percent of dissolved air, or about 2.6 percent of dissolved nitrogen, in
7.24 CHAPTER SEVEN
water returning from point B to A cannot absorb more air again from the diaphragm tank. If there is
no fresh city water supply to the water system, then after a period of water recirculation the only
dissolved air in water will be the 2.6 percent nitrogen. No further oxidation occurs after the initial
dissolved oxygen has reacted with the steel pipe. Waterlogging does not occur either.
Because of the above concerns, a closed water system should have a diaphragm or bladder
expansion tank. An open expansion tank at high levels causes fewer problems than a plain closed
expansion tank. A diaphragm tank may be smaller than an equivalent plain tank. An air eliminator
or air separator is usually required for large water systems using a diaphragm tank to separate
dissolved air from water when the water system is charged with a considerable amount of city
water.
7.5 CORROSION AND DEPOSITS IN WATER SYSTEM
Corrosion
Corrosion is a destructive process that acts on a metal or alloy. It is caused by a chemical or electrochemical
reaction of a metal. Galvanic corrosion is the result of contact between two dissimilar
metals in an electrolyte. The corrosion process involves a flow of electricity between two areas of a
metal surface in a solution that conducts the electric current. One area acts as the anode and
releases electrons, whereas the other area acts as the cathode, which accepts electrons and forms
negative ions. Corrosion, or the formation of metal ions by means of oxidation and disintegration
of metal, occurs only at the anodes. In iron and steel, ferrous ions react with oxygen to form ferric
hydroxide (the rust).
Moisture encourages the formation of an electrolyte, which is one of the basic elements that give
rise to corrosion. Oxygen accelerates the corrosion of ferrous metals by means of a reaction with
hydrogen produced at the cathode. This creates the reaction at the anode. Some alloys, such as
those of stainless steel and aluminum, develop protective oxide films to prevent further corrosion.
For iron and steel, solutions such as those containing mineral acids accelerate the corrosion, and
solutions such as those containing alkalies retard it. Because the corrosion reaction at the cathode
depends on the concentration of hydrogen ions, the more acidic the solution, the higher the concentration
of hydrogen ions and the greater the corrosion reaction. Alkaline solutions have a much
higher concentration of hydroxyl ions than hydrogen, and as such the ions decrease the corrosion
rate.
Water Impurities
In hot and chilled water systems, the problems associated with water mainly concern waters dissolved
impurities, which cause corrosion and scale, and the control of algae, bacteria, and fungi.
Typical samples of dissolved impurities in public water supplies are listed in Table 7.7.
Calcium hardness, sulfates, and silica all contribute to the formation of scale. Scale is the
deposit formed by the precipitation of water-insoluble constituents on a metal surface. Chlorides
cause corrosion. Iron may form deposits on a surface through precipitation. All these increase the
fouling factor of water.
In addition to dissolved solids, unpurified water may contain suspended solids, which may be either
organic or inorganic. Organic constituents may be in the form of colloidal solutions. At high
water velocities, hard suspended solids may abrade pipes and equipment. Particles that settle at the
bottom may accelerate corrosion.
In open water systems, bacteria, algae, and fungi cause many operating problems. The possibility
of bacteria existing in the cooling tower and causing Legionnaires disease necessitates microbiological
control.
WATER SYSTEMS 7.25
Water Treatments
Scale and Corrosion Control. One effective corrosion control method is to reduce the oxygen
composition in water systems. In past years, acids and chromates were the chemical compounds
commonly used to eliminate or to reduce scale and corrosion. On January 3, 1990, however, Proposed
Prohibition of Hexavalent Chromium Chemicals in Comfort Cooling Towers was posted by
the Environmental Protection Agency because chromates are suspected carcinogens and disposal
problems are associated with these chemicals.
There has been a significant improvement in water treatment chemistry in recent years. Currently
used chemical compounds include crystal modifiers and sequestering chemicals. Crystal modifiers
cause a change in the crystal formation of scale. As a result, scale ions cannot interlace with ions of
other scale-forming elements. Another important characteristic of these crystal modifiers and sequestering
chemicals is that they can be applied to water systems that have a wide range of pH values.
(The pH value indicates the acidity or alkalinity of a solution. It is the negative logarithm of the hydrogen
ion concentration of a solution.) Even if the chemical is over- or underfed, it will not cause operating
problems. Crystal modifiers and sequestering chemicals create fewer environmental problems.
Microbiological Control. The growth of bacteria, algae, and fungi is usually treated by biocides
to prevent the formation of an insulating layer on the heat-transfer surface, which would promote
corrosion and restrict water flow. Chlorine and chlorine compounds have been effectively and
widely used. Bromine has the same oxidizing power as chlorine and is effective over a wide pH
range. Biocide chemicals are detrimental to the environment if they are used in excess, however.
Biostat is a new type of chemical used in algae growth control. It prevents the algae spores from
maturing, which is an approach different from that of a biocide.
Blow-down or bleed-off operation is an effective water treatment. It should be considered as
important as treatments that use chemicals.
7.26 CHAPTER SEVEN
TABLE 7.7 Analyses of Typical Public Water Supplies
Location or area*,
Substance Unit (1) (2) (3) (4) (5) (6) (7) (8) (9)
Silica SiO2 2 6 12 37 10 9 22 14 
Iron Fe2 0 0 0 1 0 0 0 2 
Calcium Ca 6 5 36 62 92 96 3 155 400
Magnesium Mg 1 2 8 18 34 27 2 46 1,300
Sodium Na 2 6 7 44 8 183 215 78 11,000
Potassium K 1 1 1  1 18 10 3 400
Bicarbonate HCO3 14 13 119 202 339 334 549 210 150
Sulfate SO4 10 2 22 135 84 121 11 389 2,700
Chloride Cl 2 10 13 13 10 280 22 117 19,000
Nitrate NO3 1  0 2 13 0 1 3 
Dissolved solids 31 66 165 426 434 983 564 948 35,000
Carbonate hardness CaCO3 12 11 98 165 287 274 8 172 125
Noncarbonate hardness CaSO4 5 7 18 40 58 54 0 295 5,900
*All values are ppm of the unit cited to nearest whole number.
Numbers indicate location or area as follows:
(1) Catskill supply, New York City
(2) Swamp water (colored), Black Creek, Middleburg, FL
(3) Niagara River (filtered), Niagara Falls, NY
(4) Missouri River (untreated), average
(5) Well waters, public supply, Dayton, OH, 30 to 60 ft (9 to 18 m)
(6) Well water, Maywood, IL, 2090 ft
(7) Well water, Smithfield, VA, 330 ft
(8) Well water, Roswell, NM
(9) Ocean water, average
Source: ASHRAE Handbook 1987, HVAC Systems and Applications. Reprinted with permission.
7.27 WATER SYSTEMS
Chemical Feeding. Improper chemical feeding causes operating problems. A water treatment program
with underfed chemicals results in an ineffective treatment, whereas an overfed program not
only increases the operating cost but also may cause environmental problems. Generally, a continuous
feeding of very small amounts of chemicals often provides effective and economical water
treatment.
System Characteristics
Closed water systems, including hot, chilled, and dual-temperature systems, can be characterized as
follows:
Constant or Variable Flow. A constant-flow water system is a system for which the volume flow
at any cross-sectional plane in the supply or return mains remains constant during the operating
period. Three-way mixing valves are used to modulate the water flow rates to the coils. In a variable-
flow system, all or part of the volume flow varies when the system load changes during the operating
period. Two-way valves are used to modulate the water flow rates to the coils or terminals.
Direct Return or Reverse Return. In a direct-return water system, the various branch piping circuits,
such as ABGBA and ABCFGBA, are not equal in length (see Fig. 7.9a). Careful balance is often
required to establish the design flow rates for a building loop when a direct-return distribution
loop is used as described in later sections. In a reverse-return system, the piping lengths for each
branch circuit, including the main and branch pipes, are almost equal (see Fig. 7.9b).
FIGURE 7.9 Direct-return and reverse-return water systems. (a) Direct-return; (b) reverse-return
Two Pipe or Four Pipe. In a dual-temperature water system, the water piping from the boiler or
chiller to the coils and the terminals, or to various zones in a building, can be either a two-pipe system,
with a supply main and a return main, as shown in Fig. 7.10a; or a four-pipe system, with a hot
water supply main, a hot water return main, a chilled water supply main, and a chilled water return
main, as shown in Fig. 7.10b. For a two-pipe system, it is impossible to heat and cool two different
coils or terminals in the same zone simultaneously. Changeover from summer cooling mode operation
to winter heating mode operation is required. A four-pipe system does not need changeover operation.
Chilled and hot water can be supplied to the coils or terminals simultaneously. However, a
four-pipe system requires a greater installation cost.
Several decades earlier, there was also a three-pipe system with a hot water supply main, a chilled
water supply main, and a common return main. ASHRAE/IESNA Standard 90.1-1999 clearly speci-
fies that hydronic systems that use a common return system for both hot water and chilled water shall
not be used. This is because of the energy loss during the mixing of the hot and chilled water.
Changeover
In a dual-temperature two-pipe system, changeover refers to when the operation of one zone or the
entire water system in a building changes from heating mode to cooling mode, or vice versa. During
changeover, the water supplied to the terminals is changed from hot water to chilled water, or
vice versa. The changeover temperature Tco, F (C), is the outdoor temperature at which the space
sensible cooling load can be absorbed and removed by the combined effect of the conditioned
outdoor air, the primary air, and the space transmission and infiltration loss. Such a relationship can
be expressed as:
(7.5)
where Tr  space temperature, F (C)
sum of internal sensible loads from electric lights, occupants, and appliances,
Btu/h (W)
sum of external sensible loads through building shell, Btu/h (W)
volume flow rate and density of conditioned outdoor air, cfm (m3/min) and lb/ ft3
(kg/m3)
cpa  specific heat of air, Btu / lb F (J/kg C)
Tso  supply temperature of outdoor air or primary air, F (C)
qtl  transmission and infiltration losses per 1F of outdoor-indoor temperature difference,
Btu/h F (W/C)
Changeover usually takes from 3 to 8 h to complete. The greater the size of the water system, the
longer the changeover period. To prevent more than one changeover per day, the changeover temperature
Tco may have a tolerance of 
2F (1.1C).
Changeover may cause a sudden flow of a large amount of hot water into the chiller or of chilled
water into the boiler. Such a rapid change in temperature imposes a thermal shock on the chiller or
boiler and may damage the equipment. For chillers, the temperature of water entering the chiller
should be no higher than 80F (26.7C) to prevent excessive refrigerant pressure in the evaporator.
For boilers, a temperature control system bypasses most of the low-temperature water until the water
temperature can be gradually increased.
Changeover may be performed either manually or automatically. Manual changeover is simple
but may be inconvenient during periods when daily changeover is required. With sufficient
safety controls, automatic changeover reduces the operating duties significantly. A compromise is
V?
so, 	so 
Qres 
Qris 
K  60 	socpa
Tco  Tr 
Qris 	 Qres  KV?so(Tr  Tso)
qtl
7.28 CHAPTER SEVEN
7.29
FIGURE 7.10 Multiple-zone, dual-temperature water systems. (a) Two-pipe system; (b) four-pipe system.
a semiautomatic changeover system in which the changeover temperature is set by a manual
switch.
Outdoor reset control is often used to vary the supply water temperature Tws [F(C)], in response
to the outdoor temperature To for a hot water system. Typically, Tws is 130F (54.4C) at the
winter design temperature and drops linearly to 80F (26.7C) at the changeover temperature.
ASHRAE/IESNA Standard 90.1-1999 specifies that two-pipe changeover systems are acceptable
when the following requirements are met:
 The designed deadband width between changeover from one mode to the other is of at least 15F
(8.3C) outdoor temperature.
 System controls will allow operation in one mode for at least 4 h before changing over to
another mode.
 At the changeover point, reset controls allow heating and cooling supply temperatures to be no
more than 30F (16.7C) apart.
7.7 CENTRIFUGAL PUMPS
Centrifugal pumps are the most widely used pumps for transporting chilled water, hot water, and
condenser water in HVAC&R systems because of their high efficiency and reliable operation.
Centrifugal pumps accelerate liquid and convert the velocity of the liquid to static head. A typical
centrifugal pump consists of an impeller rotating inside a spiral casing, a shaft, mechanical seals
and bearings on both ends of the shaft, suction inlets, and a discharge outlet, as shown in Fig. 7.11.
The impeller can be single-stage or multistage. The vanes of the impeller are usually backwardcurved.
The pump is usually described as standard-fitted or bronze-fitted. In a standard-fitted construction,
the impeller is made of gray iron, and in a bronze-fitted construction, the impeller is made of
bronze. For both constructions, the shaft is made of stainless steel or alloy steel, and the casing is
made of cast iron.
Three types of centrifugal pumps are often used in water systems in HVAC&R systems: doublesuction
horizontal split-case, frame-mounted end suction, and vertical in-line pumps, as shown in
the upper part of Fig. 7.12. Double-suction horizontal split-case centrifugal pumps are the most
widely used pumps in large central hydronic air conditioning systems.
Basic Terminology
Volume flow rate [gpm (m3/s)] is the capacity handled by a centrifugal pump. On a cross-sectional
plane perpendicular to fluid flow in a water system, the static head Hs [ft (m)] is the pressure
expressed in feet (meters) of water column that is exerted on the surrounding fluid and enclosure.
On a cross-sectional plane, velocity head Hv [ft (m)] can be calculated as
(7.6)
where vo  velocity of water flow at pump outlet, ft / s (m/ s)
g  gravitational acceleration, 32.2 ft / s2 (9.81 m/ s2 )
Total head Ht [ft (m)] is the sum of static head and velocity head, i.e.,
Ht  Hs 	 Hv (7.7)
Hv 
Vo 2
2g
V?
p
7.30 CHAPTER SEVEN
WATER SYSTEMS 7.31
(a)
(b)
Suction
inlet
Enclosed double-suction
impeller
Ball bearings
Shaft
Case wear rings
Bronze shaft sleeve
FIGURE 7.11 A double-suction, horizontal split-case, single-stage centrifugal pump. (a) Sectional view;
(b) centrifugal pump and motor assembly.
Net static head [ft (m)] is the head difference between the discharge static head Hdis and suction
static head Hsuc, both in feet (meter), as shown in Fig. 7.13.
Pump power Pp (hp) is the power input on the pump shaft; and pump efficiency p is the ratio of
the energy output from water to the power input on the pump shaft, and both can be calculated as
(7.8)
where gs  specific gravity, i.e., the ratio of the mass of liquid handled by the pump to the mass of
water at 39F (4C).
Performance Curves
Pump performance is often illustrated by a head-capacity curve and a power-capacity
curve, as shown in Fig. 7.12. The head-capacity curve illustrates the performance of a centrifugal
pump from maximum volume flow to the shutoff point. If the total head at shutoff point Hso is 1.1
to 1.2 times the total head at the point of maximum efficiency Hef , the pump is said to have a flat
head-capacity curve. If Hso  1.1Hef to 1.2Hef, it is a steep-curve pump.
Pp-V?p Ht-V?p
p 
V?
p t gs
3960Pp
Pp 
V?
p Ht gs
3960 p
Hs
7.32 CHAPTER SEVEN
FIGURE 7.12 Performance curves for centrifugal pumps.
Net Positive Suction Head
The lowest absolute water pressure at the suction inlet of the centrifugal pump (shown in Fig. 7.11)
must exceed the saturated vapor pressure at the corresponding water temperature. If the absolute
pressure is lower than the saturated vapor pressure, the water evaporates and a vapor pocket forms
between the vanes in the impeller. As the pressure increases along the water flow, the vapor pocket
collapses and may damage the pump. This phenomenon is called cavitation.
The sum of the velocity head at the suction inlet and the head loss (due to friction and
turbulence) between the suction inlet and the point of lowest pressure inside the impeller is
called the net positive suction head required (NPSHR), in feet (meters). This factor is determined by
the pump manufacturer for a given centrifugal pump. For a specific water system using a centrifugal
pump, the net positive suction head available (NPSHA), in feet (meters), can be calculated as
NPSHA  Hat 	 Hsuc  Hf  2.31pvap (7.9)
where Hat  atmopheric pressure, usually expressed as 34 ft (10.3 m or 101 kPa) of water
column
Hsuc  static suction head, ft (m)
Hf  head loss due to friction and dynamic losses of suction pipework and fittings, ft (m)
pvap  saturated water vapor pressure corresponding to water temperature at suction inlet,
psia (kPa abs.)
NPSHA must be greater than NPSHR to prevent cavitation.
Pump Selection
First, the selected pump must satisfy the volume flow and total head requirements and should operate
near maximum efficiency most of the time. Second, for comfort air conditioning systems, quiet operation
is an important consideration. A noise generated in a water system is very difficult to isolate and
WATER SYSTEMS 7.33
FIGURE 7.13 Net static head.
remove. In most cases, the lowest-speed pump that can provide the required and pump head Ht is
the quietest and often the most economical choice. Third, today a variable-speed pump is often costeffective
for a variable-flow system. However, when a constant-speed pump is used to serve a variable-
flow system that operates with minor changes of head, a flat-curve pump should be selected.
7.8 PUMP-PIPING SYSTEMS
System Curve
When a pump is connected with a pipe system, it forms a pump-piping system. A water system may
consist of one pump-piping system or a combination of several pump-piping systems.
The speed of a variable-speed pump in a variable-flow water system is often controlled by a
pressure-differential transmitter installed at the end of the supply main, with a set point normally
between 15 and 20 ft WC (4.5 and 6 m WC). This represents the head loss resulting from the control
valve, pipe fittings, and pipe friction between the supply and return mains at the farthest branch
circuit from the variable-speed pump. Therefore, the head losses of a pump-piping system can be
divided into two parts:
 Constant, or fixed, head loss Hfix, which remains constant as the water flow varies. Its magnitude
is equal to the set point of the pressure-differential transmitter Hset, or the head difference between
the suction and the discharge levels of the pump in open systems Hsd [ft WC (m WC)].
 Variable head loss Hvar, which varies as the water flow changes. Its magnitude is the sum of the head
losses caused by pipe friction Hpipe, pipe fittings Hfit, equipment Heq (such as the pressure drop
through the evaporator, condenser, and coils), and components Hcp, all in ft WC (m WC), that is,
Hvar  Hpipe 	 Hfit 	 Heq 	 Hcp (7.10)
Head losses Hfix and Hvar are shown in Fig. 7.14. The relationship between the pressure loss
p [ft WC (kPa)]; flow head Hvar [ft WC (m WC)]; flow resistance of the water system Rvar [ft
WC/(gpm)2 (m WC s 2/m6)]; and water volume flow rate [gpm (m3/s)], can be expressed as
(7.11)
where 	w  density of water, lb/ft3 (kg/m3)
g  gravitational acceleration, ft / s2 (m/s2)
gc  dimensional constant, 32.2 lbmft / lbf  s2
The curve that indicates the relationship between the flow head, flow resistance, and water volume
flow rate is called the system curve of a pump-piping system, or a water system.
System Operating Point
The intersection of the pump performance curve and the water system curve is the system operating
point of this variable-flow water system, as shown by point P in Fig. 7.14. Its volume flow rate is
represented by [gpm (m3/s)], and its total head is HP  Hfix 	 Hvar [ft WC (m WC)].
Usually, the calculated system head loss is overestimated, and the selected pump is oversized
with a higher pump head, so that the actual system operation point is at point P. Therefore, for a
V?
P
Hvar  R
var V?
w 2
p  Rvar V?w
2
Hvar  p
gc
	wg
V?
w
V?
p
7.34 CHAPTER SEVEN
variable-flow water system installed with a constant-speed pump, the design system operating
point is preferably located to the left of the region of pump maximum efficiency, because the system
operating point of an oversized pump moves into or nearer to the region of pump maximum
efficiency.
Combination of Pump-Piping Systems
When two pump-piping systems 1 and 2 are connected in series as shown in Fig. 7.15a, the volume
flow rate of the combined pump-piping system, [gpm (m3/s)] is
(7.12)
where and are the volume flow rate of pump-piping systems 1 and 2, gpm (m3/s). The total
head lift of the combined system Hcom [ft WC (m WC)] is
Hcom  H1 	 H2 (7.13)
where H1 and H2 are the head of pump-piping systems 1 and 2, ft WC (m WC).
It is simpler to use one system curve to represent the whole system, i.e., to use a combined system
curve. The system operating point of the combined pump-piping system is illustrated by point
P with a volume flow of and head of HP. The purpose of connecting pump-piping systems in
series is to increase the system head.
When a pump-piping system has parallel-connected water pumps, its volume flow rate [gpm
(m3/s)] is the sum of the volume flow rates of the constituent pumps , , etc. The head of each
constituent pump and the head of the combined pump-piping system are equal. It is more convenient
to draw a combined pump curve and one system curve to determine their intersection, the
system operating point P, as shown in Fig. 7.15b. The purpose of equipping a water system with
parallel-connected water pumps is to increase its volume flow rate.
V?
2 V?1
V?
V?
P
V?
2 V?1
V?
com  V?1  V?2
V?
com
WATER SYSTEMS 7.35
FIGURE 7.14 Water system curve and system operating point.
Modulation of Pump-Piping Systems
Modulation of the volume flow rate of a pump-piping system can be done by means of the following:
 Throttle the volume flow by using a valve. As the valve closes its opening, the flow resistance of
the pump-piping system increases. A new system curve is formed, which results in having a new
system operating point that moves along the pump curve to the left-hand side of the original
curve, with a lower volume flow rate and higher total head, as shown in Fig. 7.16a. Such behavior
is known as riding on the curve. Using the valve to modulate the volume flow rate of a
pump-piping system always wastes energy because of the head loss across the valves Hval in
Fig. 7.16a.
 Turn water pumps on or off in sequence for pump-piping systems that have multiple pumps in a
parallel connection. Modulation of the volume flow rate by means of turning water pumps on and
off often results in a sudden drop or increase in volume flow rate and head, as shown by system
operating points P, Q, and T in Fig. 7.16b.
 Vary the pump speed to modulate the volume flow and the head of a pump-piping system. When
the speed of the pump is varied from n1 to n2 and then to n3, new pump curves P2 and P3 are
formed, as shown in Fig. 7.16b. The system operating point will move from point P to Q and then
to T along the system curve, with a lower volume flow rate, head, and input pump power. The system
curve becomes the modulating curve and approaches Hfix  Hset when the volume flow
rate is zero. Here Hset is the set point of the pressure-differential transmitter, ft WC (m WC).
Varying the pump speed requires the lowest pump power input in comparison with other modulation
methods.
7.36 CHAPTER SEVEN
FIGURE 7.15 Combination of pump-piping systems. (a) Two pump-piping systems connected in series; (b) three parallel-connected
pumps.
Pump Laws
The performance of geometrically and dynamically similar pump-piping systems 1 and 2 can be expressed
as follows:
(7.14a)
(7.14b)
(7.14c)
where  volume flow rate of pump-piping system, gpm (m3/s)
Ht  total head lift, ft WC (m WC)
P  pump power input at shaft, hp (kW)
D  outside diameter of pump impeller, ft (m)
n  speed of pump impeller, rpm
Equations (7.14a) through (7.14c) are known as the pump laws. They are similar to the fan laws and
are discussed in detail in Chap. 17.
Wire-to-Water Efficiency
A pump may be directly driven by a motor, or it may be driven by a motor and belts. When the energy
cost of a water system is evaluated, the pump total efficiency p, the motor efficiency mot, and
the efficiency of the variable-speed drives dr should all be considered.
V?
P2
P1

n2 3
n1 3
Ht2
Ht1

n2 2
n1 2
V?2
V?
1

D2 3n2
D1 3n1
WATER SYSTEMS 7.37
FIGURE 7.16 Modulation of pump-piping systems. (a) Using a valve; (b) varying the pump speed.
The wire-to-water efficiency of a water system ww, expressed either in dimensionless form or as
a percentage, is defined as the ratio of energy output from water to the energy input to the electric
wire connected to the motor. It can be calculated as
ww  p dr mot (7.15)
The total efficiency of the centrifugal pump p can be obtained from the pump manufacturer or calculated
from Eq. (7.8). The pump efficiency p depends on the type and size of pump as well as the
percentage of design volume flow rate during operation. Pump efficiency usually varies from 0.7 to
0.85 at the design volume flow rate. Drive efficiency dr indicates the efficiency of a direct drive,
belt drive, and various types of variable-speed drives. For direct drive, dr  1. Among variablespeed
drives, an adjustable-frequency alternating-current (ac) drive has the highest drive efficiency.
For a 25-hp (18.7-kW) motor, dr often varies from 0.96 at design flow to 0.94 at 30 percent design
flow to 0.80 at 20 percent design flow. Motor efficiency mot depends on the type and size of motor.
It normally varies from 0.91 for a 10-hp (7.5-kW) high-efficiency motor to 0.96 for a 250-hp (187-
kW) motor as listed in Table 6.2.
7.9 OPERATING CHARACTERISTICS OF CHILLED
WATER SYSTEM
Many chilled and hot water systems used in commercial central hydronic air conditioning systems
often have their central plant located in the basement, rooftop, or equipment floors of the building.
The hot/chilled water from the boiler/chiller in the central plant is then supplied to the coils and terminals
of various zones in one building or in adjacent buildings by means of supply main pipes.
Water returns from the coils and terminals to the central plant via the return mains.
Coil Load and Chilled Water Volume Flow
In AHUs or fan coils, two-way control valves are currently widely used to modulate the water
volume flow rate so as to maintain a predetermined air discharge temperature or space temperature
at reduced system loads. Coils, especially oversized coils, operate at design load usually less than 5
percent of their total operating time. For a typical coil, nearly 60 percent of the operating time may
correspond to a coil load of 35 to 65 percent of the design value.
During part-load operation, the required fraction of design volume flow rate of chilled water
flowing through a coil is not equal to the fraction of design sensible coil load Qcs Btu/h (W)
which is the sensible heat transfer from the coil to the conditioned air, as shown in Fig. 7.17a. In
Fig. 7.17a, indicates the design chilled water volume flow rate, gpm (m3/min), and Qcs,d the
design sensible coil load, Btu/h (W). This is because of the characteristics of sensible heat transfer
described by
Qcs  AoUoTm (7.16)
where Ao  outer surface area of coil, ft2 (m2)
Uo  overall heat-transfer coefficient based on outer area, Btu/h  f t 2 F (W/m2C)
Tm  logarithmic temperature between conditioned air and chilled water,F (C)
When the volume flow rate of chilled water is reduced, the decrease in the product of AoUoTm
is not the same as the reduction in the chilled water volume flow rate . When drops, the outer
surface area Ao remains the same and Uo is slightly reduced. Only a considerable rise in chilled
water temperature across the coil Tw,c  Twl  Twe, as shown in Fig. 7.17b, can reduce Tm suffi-
V?
w V?w
V?
w
V?
w,d
V?
w
7.38 CHAPTER SEVEN
ciently to match the reduction of Qcs. Figure 7.17 is obtained for entering water and entering air
temperatures that remain constant at various fractions of the design flow.
Theoretically, when the sensible coil load Qcs is reduced to 0.6 of the design value, the chilled
water volume flow rate should be decreased to about 0.25 of the design volume flow rate to match
the reduction of Qcs. Meanwhile, the power input at the shaft of the variable-speed pump is only
about 8 percent of its design brake horsepower. There is a tremendous savings in pump power for a
variable-flow system compared to a constant-flow system.
A two-way control valve for the coil must be carefully selected. First, an equal-percentage contour
valve should be used. As described in Sec. 5.6 and shown in Figs. 5.16 and 7.17a, when a coil
is equipped with an equal-percentage valve, the sensible coil load is directly related to the valve
stem travel or control output signal through the percentage water flow rate and thus provides better
control quality. Second, the control valve closes its opening to provide both the pressure drop for
the modulated required flow and an additional pressure drop for the coils nearer to the pump for
water flow balance if it is a direct-return piping system.
Chiller Plant
For chilled water systems, a central plant is often installed with multiple chillers, typically two to
four chillers. Multiple chillers are usually connected in parallel. Each chiller is often installed with
a chilled water pump that has the same volume flow rate as the water chiller. In such an arrangement,
it is more convenient to turn the chillers on or off in sequence. ASHRAE/IESNA Standard
90.1-1999 specifies that when a chilled water plant is equipped with more than one chiller, provisions
shall be made so that the chilled water flow in the chiller plant can be automatically reduced
when a chiller is shut down. Chillers must provide adequate chilled water flow and cooling capacity
that the AHUs and fan coils require.
Usually, a fairly constant-volume flow in the evaporator of the water chiller is preferable to
avoid an extremely high temperature drop in the chiller and to prevent water from freezing at a reduced
flow during part-load operation. A constant flow of chilled water in chillers is also beneficial
to the capacity control of multiple chillers.
WATER SYSTEMS 7.39
FIGURE 7.17 Relationship between fraction of design volume flow rate , coil load Qcs, and water temperature
rise Tw,c. (a) Qcs versus ; (b) Tw,c versus . V?w,d V?w
V?
w,d
Variable Flow for Saving Energy
ASHRAE/IESNA Standard 90.1-1999 specifies that water systems having a total pump system
power exceeding 10 hp (7.5 kW) shall include control valves to modulate or step open and close as
a function of load and shall be designed for variable flow for energy savings. Water systems should
be able to reduce system flow to 50 percent of design flow or less. Individual pumps serving variable
flow systems having a pump head exceeding 100 ft (30 m) and motors exceeding 50 hp (37
kW) shall have controls and devices (such as variable speed control) that will reduce pump motor
demand of no more than 30 percent of design wattage at 50 percent of design water flow.
Water Systems in Commercial Buildings
The following types of water systems are currently used in commercial buildings in the United States:
 Plant-through-building loop: bypass throttling flow
 Plant-through-building loop: distributed pumping
 Plant-building loop
 Plant-distributed pumping loop
 Plant-distribution-building loop
 Plant-distributed building loop
 Multiple sources-distributed building loop
Since the 1960s, one of the old, energy-inefficient water systems, a constant-flow system using
three-way valves which has constant flow in its chiller/boiler and its supply and return mains, is
gradually being replaced by a water system that is equipped with cheaper and more effective twoway
valves and uses energy-efficient variable flow for distribution. Constant-flow systems using
three-way valves are not discussed here.
7.10 PLANT-THROUGH-BUILDING LOOP
In a plant-through-building loop system, water is transported only by plant (chiller/boiler) pump(s)
or by distributed pump(s). Plant-through-building loops can be classified into three categories: bypass
throttling flow, distributed pumping, and variable flow.
Bypass Throttling Flow
A plant-through-building loop water system using bypass throttling flow is one of the older
chilled /hot water systems that has been adopted in commercial buildings since the use of two-way
control valves. For each chiller/ boiler, a corresponding plant constant-speed water pump is
equipped as shown in Fig. 7.9a. The chilled or hot water is supplied to the coils and terminals
through the supply and return mains and branches, and is then returned to the chiller/boiler. There is
a crossover bridge that connects the supply and return mains at junctions P and Q. A bypass twoway
control valve is often installed on the crossover bridge. A pressure-differential transmitter and
a pressure relief valve are used to maintain a set pressure differential across the supply and return
mains by modulating the opening of the bypass two-way control valve when the system pressure
7.40 CHAPTER SEVEN
tends to increase during part-load operation. A portion of the water is throttled in the control valve
and flows through the bypass crossover. It is then combined with water from the return main and returns
to the chiller/boiler. A constant flow (or approximately constant flow) is maintained in the
chiller / boiler.
A plant-through-building loop using bypass throttling control cannot save pumping energy if the
set point of the pressure-differential transmitter is fixed at the value of pressure drop across the
crossover bridge at design load during part-load operation. In comparison with a plant-building
loop water system, it is simpler, and lower in first cost and needs a smaller pump room space.
However, a plant-building loop (primary-secondary loop, which is discussed in a later section)
water system saves far more pump energy than a plant-through-building loop using bypass throttling
control.
Plant-through-building loop using bypass throttling control still has applications in small projects
and especially in retrofit where space may not be available for a plant-building loop system.
Distributed Pumping
A plant-through-building loop water system using distributed pumping is often used for hot water
systems. Hot water supplied from a boiler is divided into several distributed-piping loops, as
shown in Fig. 8.7. For each of the distributed-piping loop in a control zone, there is either a variable-
speed on-line hot water circulating pump or a constant-speed on-line pump with a two-way
control valve, finned-tube baseboard heaters, supply and return pipes, accessories, and controls.
Hot water flows through the distributed-piping loops and then returns to the boiler for heating
again. The amount of hot water flowing through the boiler is reduced within a limit during partload
operation.
A space temperature sensor sends a signal to a direct digital control (DDC) unit controller and
modulates the amount of hot water extracted from the supply header by means of a variable-speed
on-line pump, or a two-way control valve and an on-line pump to maintain a preset zone temperature.
A temperature sensor is also mounted at the water outlet of the boiler to maintain a preset hot
water leaving temperature through the modulation of the burners capacity by a DDC unit
controller.
A plant-through-building loop with distributed-pumping water system is a simple, energy-effi-
cient system.
Variable Flow
A plant-through-building loop using variable flow, as shown in Fig. 7.18, is an ideal chilled water
system with variable flow in chiller, supply main, coils, return main, and variable-speed
pump(s) during part-load operation. For each chiller, there is often a corresponding variablespeed
pump. Chilled water is supplied to various control valves and coils in branches through
supply and return mains. It is then extracted by the pump(s) and returned to the chillers for cooling
again.
Three related direct digital controls are equipped for this system:
 A control of chilled water temperature leaving the chiller
 An air discharge temperature control by modulation of the control valve and thus the water flow
rate entering the coil
 A pressure-differential control using pressure-differential transmitter with a DDC unit controller
that modulates the variable-speed pump to maintain a preset pressure differential between supply
and return mains
WATER SYSTEMS 7.41
Today, although many chiller manufacturers allow a reduction of chilled water flow of 30 to 40 percent
of the design volume flow rate, multiple variable-speed pumps must run at approximately the
same speed at the same pump head. If Qcs /Qcs,d drops to 0.5, the system flow ratio needs to
reduce to only about 0.28 (refer to Table 7.8). This may cause operating troubles, control problems,
and system instability.
Until the middle of 1997, there hasnt been a single project actually using plant-throughbuilding
loop with variable-flow chilled water system reported in HVAC&R publications in the
United States and operated successfully.
V?
w/V?w,d
7.42 CHAPTER SEVEN
TABLE 7.8 Magnitudes of bg/ bg,d and Calculated Results for Example 7.1
Qcs/Qcs,d
0.3 0.4 0.5 0.6 0.65 0.7 0.8 0.9 1.0
1 chiller 2 chillers 2 chillers 3 chillers
bg/ bg,d 0.17 0.17 0.22 0.275 0.34 0.38 0.38 0.45 0.60 0.8 1.0
Tw,c, F 36 36 36 36 35.2 34.2 34.2 31.1 26.7 22.5 20
pt, gpm 350 700 700 700 700 700 1050 1050 1050 1050 1050
cn, gpm 180 530 480 425 360 320 670 600 450 250 0
Tee, F 57.5 48.8 51.3 54.2 57.1 58.6 52.4 53.3 55.3 57.1 60
V? V?
V?
V?
V?
V?
2 1 1 2
Coil
Chiller
Coil
Coil
Coil
PD T
PD Pressure-differential transmitter Variable-speed drive
VSD
VSD
Chiller
VSD
Chiller
VSD
FIGURE 7.18 A chilled water system of plant-through-building loop using variable flow.
7.11 PLANT-BUILDING LOOP
System Description
Plant-building loop water systems, also called primary-secondary loop (or circuit) water systems,
are the widely adopted water systems for large new and retrofit commercial HVAC&R installations
in the United States today. A plant-building loop chilled water, hot water, or dual-temperature water
system consists of two piping loops:
 Plant loop (primary loop). In a plant loop, there are chiller(s)/boiler(s), circulating water pumps,
diaphragm expansion tank, corresponding pipes and fittings, and control systems, as shown by
loop ABFG in Fig. 7.19. A constant volume flow rate is maintained in the evaporator of each
chiller. For a refrigeration plant equipped with multiple chillers, the chilled water volume flow
rate in the plant loop will vary when a chiller and its associated chiller pump are turned on or off.
 Building loop (secondary loop). In a building loop, there are coils, terminals, probably variablespeed
water pumps, two-way control valves and control systems, and corresponding pipes, fittings,
and accessories, as shown by loop BCDECF in Fig. 7.19. The water flow in the building
loop is varied as the coil load is changed from the design load to part-load.
A short common pipe, sometimes also called a bypass, connects these two loops and combines
them into a plant-building loop.
Control Systems
For a plant-building loop water system, there are four related specific control systems:
Coil Discharge Air Temperature Control. A sensor and a DDC system or unit controller are used
for each coil to modulate the two-way control valve and the water flow into the coil. The discharge
temperature after the coil can be maintained within predetermined limits.
Water Leaving Chiller Temperature Control. In a chiller, chilled water temperature leaving the
chiller is always maintained at a preset value within a specified period by varying the refrigerant
flow in the chiller. In a boiler, the leaving temperature of hot water is maintained at a predetermined
value by varying the fuel flow to the burner.
During part-load, the chilled water temperature leaving the chiller is reset to a higher value,
such as between 3 and 10F (1.7 and 5.6C) according to system loads or outdoor temperature
both to reduce the pressure lift between evaporating and condensing pressure and to save compressing
power. ASHRAE/IESNA Standard 90.1-1999 specifies that a chilled water system with a
design capacity exceeding 300,000 Btu/h (87,900 W) supplying chilled water to comfort air conditioning
systems shall be equipped with controls that automatically reset supply chilled water
temperature according to building loads (including return chilled water temperature) or outdoor
temperature.
Staging Control. Chillers are turned on and off in sequence depending on the required system
cooling capacity Qsc Btu/h (W), or the sum of the coils loads. The required system cooling capacity
can be found by measuring the product of the temperature difference across the supply and
return mains, as shown by temperature sensors T8 and T7 and the water volume flow rate by the
flowmeter F2 in Fig. 7.19. If the produced refrigeration capacity Qrf , Btu /h (W), measured by the
product of chilled water supply and return temperature differential (T6 and T5) and the flowmeter
(F1) is less than Qsc, a DDC system controller turns on a chiller. If Qrf Qsc is greater than the refrigeration
capacity of a chiller Qrfc, the system controller turns off a chiller. Chillers should not be
staged on or off based on the chilled water volume flow rate flowing through the common pipe.
WATER SYSTEMS 7.43
7.44
6 5 3 3
7 3 3
3 3 1 4 2
5
4 3 1
6
6
Boiler
Evaporator
Condenser
Plant
hot water
pump
Building
hot water
pumps
Chiller
pump
Condenser
pump
Chilled water
return main
Hot water
return main
Temperature sensor
Hot water
supply main
Chiller water
supply main
Coils
Building variablespeed
pumps
Common pipe
T2
T1
T
Pressure-differential
transmitter
PD
Flow meter F
Flow switch FS
Adjustable frequency
drives
VSD
T5
T7
T1 PD1
T9
T6 T8 F2
FS1
A
F
B C
C
D
E
G FS2
VSD
VSD
FIGURE 7.19 A dual-temperature water system with plant-building loop.
Pressure-Differential Control. These controls are used to maintain the minimum required pressure
differential between the supply and return mains at a specific location, as shown by PD1 in
Fig. 7.19. If only one differential-pressure transmitter is installed for chilled or hot water supply
and return mains, it is usually located at the end of the supply main farthest from the building pump
discharge. If multiple differential-pressure transmitters are installed, they are often located at places
remote from the building pump discharge, with a low signal selector to ensure that any coil in the
building loop has an adequate pressure differential between the supply and return mains.
The set point of the differential-pressure transmitter should be equal to or slightly greater than
the sum of the pressure drops of the control valve, coil, pipe fittings, and piping friction of the
branch circuit between the supply and return mains. A low set point cannot ensure adequate water
flow through the coils. A high set point consumes more pump power at a reduced flow. A set point
of 15 to 25 ft (4.5 to 7.5 m) of head loss may be suitable.
System Characteristics
For a plant-building loop chilled water system, when the volume flow rate of the chilled water in the
building loop is at its design value , the volume flow rate in the plant loop is equal to that in
the building loop theoretically, all in gpm (m3/min). In actual practice, is slightly (less than 3
percent) higher than to guarantee a sufficient chilled water supply to the building loop.
At design load, chilled water leaving the chiller(s) at point A flows through the junction of the
common pipe, plant loop, and building loop (point B), is extracted by the variable-speed building
pump; and is supplied to the coils. From the coils, chilled water returns through another junction of
the building loop, common pipe, and plant loop (point F). There is only a very small amount of
bypass chilled water in the common pipe flows in the direction from point B to F. The chilled water
return from the coils is then combined with the bypass water from the common pipe and is
extracted by the chiller pump(s) and enters the chiller(s) for cooling again.
When the coils load drops during part-load operation, and the water volume flow rate bg
reduces in the building loop because the control valves have been partially closed, is now
considerably greater than bg. Chilled water then divides into two flows at junction B: water at the
reduced volume flow rate is extracted by the variable-speed building pump in the building loop and
is supplied to the coils; the remaining water bypasses the building loop by flowing through the common
pipe, is extracted by the chiller pump(s), and returns to the chiller(s).
For a water system that includes a plant-building loop with a common pipe between the two
loops, Carlson (1968) states the following rule: One pumped circuit affects the operation of the
other to a degree dependent on the flow and pressure drop in piping common to both circuits. The
lower the pressure drop in the common pipe, the greater the degree of isolation between the plant
and building loops. The head-volume flow characteristics of these loops act as two separate
systems.
A plant-building loop has the following characteristics:
 It provides variable flow at the building loop with separate building pump(s) and constant flow
through the evaporator of the chiller and thus saves pumping power during periods of reduced
flow in the building loop. According to Rishel (1983), the annual pump energy consumption of a
plant-building loop with variable flow in a building loop that uses a variable-speed building pump
is about 35 percent that of a plant-through-building loop constant-flow system using three-way
control valves.
 It separates the building loop from the plant loop and makes the design, operation, and control of
both loops simpler and more stable.
 Based on the principles of continuity of mass and energy balance, if differences in the density
of chilled water are ignored at junctions B and F, the sum of the volume flow rates of chilled
water entering the junction must be equal to the sum of volume flow rates of water leaving that
junction. Also, for an adiabatic mixing process, the total enthalpy of chilled water entering the
V?
V?
pt
V?
V?
bg,d
V?
pt V?bg,d
V?
pt V?bg,d
WATER SYSTEMS 7.45
junction must be equal to the total enthalpy of water leaving the junction. At junction B or F,
chilled water has the same water pressure and temperature.
Sequence of Operations
Consider a chilled water system in a dual-temperature water system that is in a plant-building loop,
as shown in Fig. 7.19. There are three chillers in the plant loop, each of which is equipped with a
constant-speed chiller pump. In the building loop, there are two variable-speed building pumps connected
in parallel. One is a standby pump. Chilled water is forced through the water cooling coils in
AHUs that serve various zones in the building. For simplicity, assume that the latent coil load
remains constant when the coil load varies. Based on the data and information from Ellis and
McKew (1996), for such a chilled water system, the sequence of operations of the DDC system is
as follows:
1. When the system controller of the water system is in the off position, the chiller pump is off,
condenser pump is off, building pump is off, and the cooling tower fan is off.
2. If the system controller is turned on, then the chillers on/off switch in the unit controller is
placed in the on position; and interlock signals are sent to three chiller pumps, one variable-speed
building pump, and three condenser pumps and start all these pumps. The variable-speed building
pump is always started from zero speed and increases gradually for safety and energy saving. As
the chilled water flow swiches confirm that all the pumps are delivering sufficient water flow, the
compressor of the leading chiller (first chiller) is turned on.
3. Temperature sensor T2 tends to maintain the set point of the chilled water leaving chiller temperature
often at 45F (7.2C) by means of refrigerant flow control through multiple on/off compressors,
modulation of inlet vanes, or variable-speed compressor motor (details are discussed in
later chapters). Temperature sensors T7 and T8 and flowmeter F2 measure the required system
cooling capacity Qsc; and sensors T5 and T6 and flowmeter F1 measure the produced refrigeration
capacity Qrf. If Qsc  Qrf, chiller is staged on in sequence, until Qrf  Qsc.
4. Condenser water temperature sensor T measures the water temperature entering the condenser
so that it will not be lower than a limit recommended by the manufacturer for normal operation.
5. When the coils control valves in AHUs close, the chilled water flow drops below the design
flow. As the pressure-differential transmittter DP1 senses that the pressure differential between
chilled water supply and return mains increases to a value which exceeds the set point, such as 15 ft
(4.5 m), the system controller modulates the variable-speed drive (VSD) and reduces the speed of
the variable-speed pump to maintain a 15-ft (4.5-m) pressure differential.
6. At the design system load, three chillers shall provide nearly their maximum cooling capacity,
and the veriable-speed building pump shall provide maximum flow through pump speed control.
All two-way valves shall be nearly opened fully. A constant chilled water flow is maintained in the
evaporator of each chiller. Cooling tower fan shall be continuously operated at full speed.
7. During part-load operation as the sum of the coils load (system load) decreases, the two-way
valves close their openings to reduce the chilled water flowing through the coils. At a specific fraction
of design sensible coil load Qcs /Qcs,d, there is a corresponding water volume flow rate in the
building loop, expressed as a fraction of design flow , that offsets this coil load. The building
variable-speed pump should operate at this building volume flow rate [gpm (m3/min)] with a
head sufficient to overcome the head loss in the building loop through the modulation of the variable-
speed pump.
The supply and return temperature differential of the building loop, or the mean chilled water
temperature rise across the coils Twc [FC] depends on the fraction of the design sensible coil load
Qcs /Qcs,d and the fraction of the design volume flow rate through the coils . The smaller the
value of , the greater the temperature rise Tw,c. At part load (Qcs /Qcs,d  1), Tw,c is always
greater than that at the design load, as shown in Fig. 7.20d.
V?
bg/V?bg,d
V?
bg/V?bg,d
V?
bg
V?
bg / V?bg,d
7.46 CHAPTER SEVEN
WATER SYSTEMS 7.47
FIGURE 7.20 System performance curves for plant-building loop. (a) Schematic diagram; (b) head of building
variable-speed pump at various volume flow rates; (c) versus Qcs /Qcs, d and versus Qcs /Qcs,d;
(d) Tw,c versus Qcs /Qcs,d.
V?
bg / V?bg, d V?bg
Coil
Chiller
Chiller
Chiller
Coil
Coil
Coil
pFG
pCK
A
E D C B F
H J K L G
(a)
M
8. During part-load operation, temperature sensors T5, T6, T7, and T8 and flowmeters F1 and F2
measure the readings which give the produced cooling capacity Qrf and required system cooling
capacity Qsc. If Qrf  Qsc  Qrf,c (one chillers cooling capacity, Btu/ h), none of the chillers is staging
off. When Qrf  Qs  Qrf,c, one of the chillers is then turned off until Qrf  Qsc  Qsc,c. To turn
off a chiller, the compressor is turned off first, then the chiller pump, condenser pump, and cooling
tower fans corresponding to that chiller.
9. During part-load operation, a constant flow of chilled water is still maintained in the evaporator
of each turned-on chiller. However, the volume flow rate in the plant loop depends on the
number of operating chillers and their associated chiller pumps. The staging on or off of the chillers
and their associated pump causes a variation of chilled water volume flow rate in the plant loop.
The difference between the volume flow rate of chilled water in the plant loop and the volume
flow rate in the building loop gives the volume flow rate of chilled water in the com- V?pt  V?bg
V?
pt
V?
pt
7.48 CHAPTER SEVEN
FIGURE 7.20 (Continued)
mon pipe , that is, . At part-load operation, there is always a bypass flow of
chilled water from the plant loop returning to the chiller via the common pipe.
10. During part-load operation, the set point of the chilled water leaving temperature is often reset
to a higher value according to either the outdoor temperature or the reduction of system load.
11. During part-load operation, as the two-way control valves close, the chilled water pressure in
the supply main of the building loop tends to increase. The pressure-differential transmitter PD1
senses this increase and reduces the speed of the variable-speed building pump by means of a
variable-speed drive to maintain a constant 15-ft (4.5-m) pressure differential between supply and
return mains. When the system load increases, the two-way control valves open wider and PD1
senses the drop of the pressure differential, increases the speed of the pump, and still maintains a
required 15-ft (4.5-m) pressure differential.
12. When the water system is shut down, the system controller should be in the off position.
First, the compressor(s) are turned off, then the variable-speed building pump is off, condenser
pump(s) are off, chiller pump(s) are off, and cooling tower fan(s) are off. The speed of the variablespeed
pump is gradually reduced to zero first, and then the pump is turned off.
Low T between Chilled Water Supply and Return Temperatures
Many chilled water systems suffer a lower actual T between chilled water supply and return temperatures
compared to the design value. There is also argument that a primary-secondary control
scheme that depends on system flow to gauge system load is virtually blind to load variation.
First, from Eq. (7.1), T  Qsc /(500 ), a lower T is due to an overestimated coil load (system
load) Qsc, or an underestimated water volume flow rate , or both. Second, a plant-building
(primary-secondary) loop should measure system load Qsc, which is the product T and is not
the only water flow to stage on and off the chillers. Third, the space load and coil load of many
projects are often overestimated, the equipment is oversized, and pump head is often calculated on
the safe side. All these result in a far greater actual flow and cause a low T. Fourth, the cleanliness
of the coils including the air-side cleanliness has an influential effect on low T. Finally, for a
chiller plant equipped with three chillers, if the design T  20F (11.1C), when one of the
chillers is operated at 50 percent of the design load, T will be lower, to about 16F (8.9C) only. A
chilled water system using a plant-building loop is not an essential factor causing a low T between
the supply and return mains.
Variable-Speed Pumps Connected in Parallel
In a chilled water system, if two or more variable-speed pumps are connected in parallel, all the
variable-speed pumps must generate the same head. The purpose of variable-speed pumps connected
in parallel is to increase the volume flow rate. Their flow is additive.
For identical variable-speed pumps connected in parallel, the best overall efficiency is often
obtained if the pumps are operated at identical speeds. Parallel-connected variable-speed pumps
should be reduced or increased to approximately the same speed. If two parallel-connected identical
variable-speed pumps are operated at different speeds, or a large pump with higher head is connected
to a small pump with lower head, then it is possible that the lower-speed pump or pump with
lower head sometimes may contribute negative effects since they must operate at the same head.
Different speed pumps or different sized pumps are hardly operated at higher efficiency at the same
time. The performance of two or more variable-speed pump-piping systems connected in parallel is
further complicated in that only variable-speed pumps are connected in parallel.
V?
gal
V?
gal
V?
gal
V?
cn  V?pt  V?bg V?cn
WATER SYSTEMS 7.49
Use of Balancing Valves
Equal-percentage two-way control valves are widely used to modulate the flow rate of chilled water
flowing through the coils during part-load operation. Because of the lower installation cost and
since often there is only limited space available inside the ceiling plenum, the direct-return piping
system is often the best choice for the chilled water system in a multistory building. For a variable-
flow building loop using a direct-return piping arrangement, the argument concerns whether a balance
valve is necessary for each branch pipe to balance the water flow according to its requirement,
such as for branches CK, DJ, EH, and FG, as shown in Fig. 7.20a. If there are no balancing valves
installed in the branch pipes, after the control valve balances the water flow at design load, can it
still effectively adjust the amount of chilled water entering the coil as required in part-load operation.
This depends mainly on the type of control valve, the control mode adopted, the variation in
pressure drop between various branches, and the difference in main pipe pressure drop between the
farthest and the nearest branch regardless of whether equal-percentage two-way control valve with
modulation control (such as proportional plus integral control ) or two-way control valve with twoposition
on/off control (including pulse-width-adjusted two-position control) is used.
For an equal-percentage two-way control valve with modulation control for many AHUs, consider
a plant-building loop in a chilled water system, as shown in Fig. 7.20a. At the design load, the
chilled water volume flow rate through branch FG is 80 gpm (0.30 m3/min), the corresponding
pressure drop of its fully opened equal-percentage two-way valve is 7.5 ft WC (3.3 psi or 2.3 m
WC), and the pressure drop across the farthest branch FG is 20 ft WC (8.7 psi or 6.1 m WC). From
Eq. (5.8), the flow coefficient
Usually, the difference between the pressure drop across the farthest branch from the building pump
FG and the pressure drop across the nearest branch CK is often within 60 ft WC (26 psi or 18 m
WC). At design load, for a fully opened two-way control valve in branch CK, even if all this 26 psi
(18 m head loss) has been added, the chilled water volume flow rate is then
From Fig. 5.16, for a typical equal-percentage two-way valve, when the percentage of the valve
stem travel lies between 80 and 100 percent, the relationship between the percentage of full-range
travel of valve stem z and the percentage of water flow rate when the valve is fully opened is
 Kekz  0.004e5.5 (7.17)
where k  proportional constant
K  flow parameter affected by size of valve
Since 80/238  0.336, substituting into Eq. (7.17) gives
0.336  0.004e5.5z
And the percentage of full-range travel of valve stem z  0.805. That is, an equal-percentage twoway
control valve in branch CK will close its opening from 100 percent fully open to 80.5 percent
for water flow balance at design load. Avery et al. (1990) emphasized that If properly selected
valves (those with equal percentage ports and with the correct actuators) are used, 20 percent or less
of the stroke will be used to balance the flow. The rest of the stroke will still be available to modulate
the flow within the design limits. Rishel (1997) also stated that manual balance valves and automatic
balance valves should not be used on variable-volume, direct-return, modulating type, coil
control valve, chilled water systems.
Therefore, an equal-percentage, two-way control valve, direct-return VAV system can balance
the water flow in a direct-return chilled water system, and at least 80 percent of its stroke is still
V?
v
V?
v
V?
 Cv?pvv  44?3.3 	 26  238 gpm (15 L / s or 0.90 m3 / min)
Cv 
V?
?pvv

80
?3.3
 44
7.50 CHAPTER SEVEN
available to perform the control actions at part-load if the variations in pressure drop of various
branches are small and the difference in main pipe pressure drop between the farthest and nearest
branches is within 60 ft WC (18 m WC).
Common Pipe and Thermal Contamination
For a plant-building loop in a chilled water system, if there is a backflow of a portion of the return
chilled water from the building loop that enters the common pipe at junction L and is mixed with
the supply chilled water from the plant loop at junction B, as shown in Fig. 7.20a, then the thermal
contamination of building return chilled water occurs. Lizardos (1995) suggested that the length
of the common pipe (expressed as the number of diameters of the common pipe) should be as
follows:
The diameter of the common pipe should be equal to or greater than the diameter of the return main
of the building loop.
Example 7.1. Consider a chilled water system using a plant-building loop. The design chilled water
flow rate is 1000 gpm (63.1 L/s) with a chilled water temperature rise across the coils of
Tw,c  20F (11.1C). There are three chillers in the central plant, each equipped with a constantspeed
chiller pump that provides 350 gpm (22 L/ s) at 40-ft (12-m) total head. In the building loop,
there are two variable-speed building pumps, each with a volume flow rate of 1000 gpm (63.1 L/ s)
at 60-ft (18-m) head. One of these building pumps is a standby pump. At design conditions, chilled
water leaves the chiller at a temperature Tel  40F (4.4C) and returns to the chiller at 60F
(15.6C). Chilled water leaving the chiller is controlled at 40F (4.4C) for both design and partload
operation. Once the fouling and inefficiency of the coils have been taken into account, the fractions
of design volume flow rate required to absorb the coil load at various fractions of the
sensible load Qcs /Qcs,d are listed in Table 7.8.
When the system load drops, plant chiller 1 will turn off when Qcs /Qcs.d equals 0.65 and chiller 2
will turn off when Qcs /Qcs,d equals 0.30. When the system load increases, chiller 2 turns on at
Qcs /Qcs,d  0.35, and chiller 1 turns on at Qcs /Qcs,d  0.7. Plant chiller 3 operates continuously.
Calculate the following based on the chillers on/off schedule at various fractions of the sensible
coil load:
1. Mean chilled water temperature rise across the coil
2. Water flow in the common pipe
3. Temperature of water returning to the water chiller
Solution
1. From the given information in Table 7.8, and Eq. (7.1), the mean water temperature rise across
the coil for Qs,c /Qsc,d  0.9 is
Values of Tw,c at other values Qsc /Qsc,d can be similarly calculated and are listed in Table 7.8.
Tw,c  20QcsV?bg,d / Qcs,dV?bg 
20  0.9
0.8
 22.5F (12.5C)
V?
bg / V?bg,d
Chilled water velocity in return main, ft / s (m/ s) Minimum length of common pipe
 5 (1.5) 3 diameters, or  2 ft (0.6 m)
 5 (1.5) 10 diameters
WATER SYSTEMS 7.51
2. For Qcs /Qcs,d  0.9, because all three chillers are operating, the water volume flow rate in the
plant loop  3(350)  1050 gpm. The water volume flow rate in the building loop is
Then, the water volume flow rate in the common pipe is
When Qcs /Qcs,d  0.65, chiller 1 is turned off. Just before the chiller is turned off, the volume
flow rate in the common pipe is
 1050  0.38(1000)  670 gpm (42.3 L/ s)
Immediately after chiller 1 is turned off
 2(350)  0.38(1000)  320 gpm (20.2 L/ s)
Values of for other values of Qcs /Qcs,d can be similarly calculated and are listed in Table 7.8.
3. For Qcs /Qcs,d  0.9, after the adiabatic mixing of water from the building loop and common
pipe, the temperature of chilled water returning to the water chiller can be calculated as
For Qcs /Qcs,d  0.65, just before chiller 1 is turned off,
Immediately after chiller 1 is turned off,
Other chilled water temperatures upon entering the chiller can be similarly calculated and are
listed in Table 7.8.
7.12 PLANT-DISTRIBUTED PUMPING
A water system using plant-distributed pumping consists of two loops: a plant loop and a distributed
pumping loop connected by a bypass (common pipe), as shown in Fig. 7.21. As in the plantbuilding
loop, the plant loop comprises chiller(s) / boiler(s), constant-speed plant pumps, piping, and
controls. A constant flow is maintained in the evaporator of each chiller / boiler.
In each of the distributed pumping loops connected to an associated air-handling unit, there is a
corresponding variable-speed distributed pump, a coil, two isolating valves, a drain valve, and other
accessories. The discharge air temperature of the AHU is controlled by the DDC unit through the
modulation of the water flow rate by a variable-speed drive (VSD) and the asssociated variablespeed
pump. There is no two-way control valve, and no pressure-differential transmitter is equipped
to maintain a fixed pressure differential between the supply and return mains at the farthest branch.
In summer, as the AHU is operated at part-load operation for cooling, if a temperature sensor
senses the discharge air temperature drops below the preset limit, a DDC unit controller modulates
Tee 
350
700 (40) 	
0.35(1000)(40 	 37.1)
700
 58.6F (14.8C)
Tee 
700
1050 (40) 	
0.35(1000)(40 	 34.2)
1050
 51.4F (10.8C)
Tee 
250
1050 (40) 	
0.8(1000)(40 	 22.5)
1050
 57.1F (13.9C)
V?
cn
V?
cn
V?
cn
V?
cn  V?pt  V?bg  1050  800  250 gpm (15.8 L / s)
V?
bg 
1000V?bg
V?
bg,d
 1000(0.8)  800 gpm (50.5 L / s)
V?
pt
7.52 CHAPTER SEVEN
the VSD and the associated variable-speed pump to reduce the amount of chilled water flowing
through the coil of the AHU, to maintain an approximately constant discharge air temperature. In
the plant-distributed pumping loop, a portion of chilled water supply from the plant loop will return
to the chiller by means of the bypass. During part-load operation for winter heating, the discharge
air temperature increases, and the controller modulates the VSD and the variable-speed pump to reduce
the amount of hot water flowing through the coil of the AHU.
Compared to a plant-building loop, a plant-distributed pumping loop has the following advantages:
 A variable-speed pump replaces the two-way control valve and balancing valve.
 No pressure-differential transmitter and control is required.
 When a distributed pumping loop is nearer to the plant (chiller/boiler), less distributing energy to
transport water in the mains is needed.
The disadvantages include higher first cost and that more maintenance is required.
Distributed pumping is suitable to apply for the water system that serves large coils in AHUs
and where the AHU must be installed inside a fan room to avoid inconvenient maintenance and
noise.
7.13 CAMPUS-TYPE WATER SYSTEMS
Chilled water or chilled and hot water is often supplied to many buildings separated from a central
plant in universities, medical centers, and airports. The benefits of using a campus-type central
plant chilled water system instead of individual building installations are cost savings, minimal
environmental impact (e.g., from cooling towers), effective operation and maintenance, and
reliability.
WATER SYSTEMS 7.53
Constantspeed
pumps
Chiller
Boiler
C
pCK
pbranch
F
K
K
G
1
1
Chiller
Boiler
Variable-speed
pumps Supply
main
Return
main
VSD VSD VSD
AHUn
VSD
AHU3
AHU2
AHU1
Bypass
FIGURE 7.21 Schematic diagram of a plant-distributed pumping loop .
The following are three types of currently used campus-type water systems: plant-distributionbuilding
loop, plant-distributed building loop, and multiple sources-distributed building loop.
Plant-Distribution-Building Loop
System Description. Many recently developed campus-type central plant chilled water systems
use a plant-distribution-building loop, as shown in Fig. 7.22a. As in a plant-building loop, constant
flow is maintained in the evaporator of each chiller in the plant loop. Each chiller also has its own
constant-speed chiller pump. Chilled water leaves the chiller at a temperature of 40 to 42F (4.4 to
5.6C). It is then extracted by the distribution pumps and forced to the supply main of the distribution
loop.
Chilled water in the supply and return mains of the distribution loop operates under variable
flow. Multiple variable-speed pumps are often used to transport chilled water at a volume flow rate
slightly higher than the sum of the volume flow rates required in the building loops. At each building
entrance, the variable-speed building pump extracts the chilled water and supplies it to the coils
in AHUs and terminals in various zones by means of building supply mains. Chilled water is then
returned to the water chillers through building return mains, a distribution-loop return main, and
chiller pumps.
The system performances of the plant loop and building loop are similar to those in a plantbuilding
loop.
Pressure Gradient of Distribution Loop. A campus-type chilled water central plant may transport
several thousand gallons of water per minute to the farthest building at a distance that may be several
thousand feet away from the plant. The pressure gradient of the distribution supply and return
mains due to the pipe friction and fitting losses causes uneven pressure differentials among the supply
and return mains [Hs,r (ft WC or m WC)] of buildings along the distribution loop, as shown in
Fig. 7.22b. Buildings nearer to the central plant have a greater Hs,r than buildings farther from the
plant. Along the distribution loop, a smaller pressure gradient results in a lower pumping power but
a larger diameter of chilled water pipe. A more even Hs,r does not impose excessive pressure drop
across the control valves of coils.
Using a lower pressure drop Hf is an effective means of reducing the pressure gradient and
pressure differential Hs,r and saves energy. For a distribution loop, a value of Hf between 0.5 and
1 ft /100 ft (0.5 and 1 m/100 m) pipe length, sometimes even lower, may be used. Low values of
Hs,r can be offset at the coil control valves without affecting the coils proper operation when there
is no building pump in the building loop. A life-cycle cost analysis should be conducted to determine
the optimum Hf .
Using two-way distribution from the central plant, with two supply and return distribution loops,
may reduce the pipe distance and the diameter of the supply and return mains. Such a distribution
loop layout depends on the location of the central plant and the air conditioned buildings as well as
the cost analyses of various alternatives.
Using a reverse-return piping arrangement instead of a direct-return one does even the pressure
differentials Hs,r along the supply and return mains of the distribution loop. However, having an
additional pipe length equal to that of the return main significantly increases the piping investment.
A simpler and cheaper way is to install a pressure throttling valve at each building entrance to offset
the excess Hs,r. Usually, direct return is used for a distribution loop.
Variable-Speed Building Pumps. The function of a variable-speed building pump is (1) to provide
variable-flow and corresponding head to overcome the pressure drop of the building loop at
design and reduced coil loads and (2) to provide different magnitudes of head for the building loop
according to the needs of various types of buildings. When only variable-speed distribution
pump(s) are used instead of both variable-speed distribution pump(s) and variable-speed building
pump(s), it must provide sufficient pump head to overcome the pressure drop of the building
7.54 CHAPTER SEVEN
WATER SYSTEMS 7.55
FIGURE 7.22 Chilled water system using plant-distribution-building loop. (a) Schematic diagram; (b) pressure gradient for distribution
loop; (c) crossover bridge with temperature control valve.
loops. However, the pressure characteristics of the supply and return mains of the distribution loop at
reduced flows make it difficult to satisfy various load profiles in different buildings during part-load
operation. Having a variable-speed building pump for each building also saves more pump energy.
Therefore, the use of variable-speed pumps for both distribution and building loops is preferable.
Control of Variable-Speed Distribution Pump. Two types of controls can be used to modulate a
variable-speed distribution pump to transport the required chilled water volume flow to various
buildings:
 A differential-pressure transmitter may be located near the farthest end of the distribution supply
main, as shown in Fig. 7.22a. Theoretically, the head of the building pump should extract the exact
required amount of chilled water corresponding to the sum of the coil loads in the building
loop, force it through the coils, and discharge it to the distribution return main. Therefore, a set
point for the pressure differential of about 5 ft (1.5 m) may be appropriate. This type of control is
widely used.
 A DDC system measures the total water flow that returns from each building by means of
flowmeters and modulates the variable-speed distribution pump to supply exactly the required
amount to various building loops. This type of control is more precise but more expensive.
Building Entrance. Chilled water is usually supplied directly from the distribution supply main
to the building supply main. A pressure throttling valve may be used to offset the excess pressure
differential Hs,r along the distribution loop.
Although using a heat exchanger at the building entrance entirely isolates the chilled water in
the distribution loop from the chilled water in the building loop, a temperature increase of about 3
to 7F (1.7 to 3.9C) is required for a chilled water heat exchanger. Because chilled water has a supply
and return temperature differential of only about 15 to 20F (8.3 to 11.1C), a heat exchanger is
seldom used at a building or zone entrance for a chilled water system. Because a hot water system
has a greater supply and return temperature differential, a heat exchanger is sometimes used at the
building entrance for a hot water system.
A chilled water building loop may be divided into various zones based on different height levels
within the building. In such an arrangement, the coils in the lower floors of the building loop will
not suffer a high static pressure because the low-level water loop is often isolated from the highlevel
water loop by means of a heat exchanger at the zone entrance.
If a building requires a chilled water supply temperature higher than that given by the distribution
supply main, a crossover bridge with a temperature control valve can be arranged for this purpose,
as shown in Fig. 7.22c.
The return temperature from the coils in a building loop is affected by the cleanliness of the coil,
including air-side cleanliness, and the control system in the building loop. The buildings variablespeed
pump is often controlled by the pressure-differential transmitter located at the end of the
building supply main, as shown in Fig. 7.22a and c.
Plant-Distributed Building Loop
A plant-distributed building loop water system has nearly the same configuration as a plant-distribution-
building loop system except that there is no distribution pump in the distribution loop. Constant-
speed chiller /boiler pumps in the plant loop supply water to the beginning of the supply main
of the distribution loop, point S, and extract water from the end of the return main of the distribution
loop, point R. Various variable-speed building pumps in the building loops extract water from
the distribution main point S. They also overcome the pressure loss of the distribution supply main
piping up to the building entrance, the pressure loss of the building loop pbg including building
supply and return mains, coils, two-way control valve, and fittings; and the pressure loss of the distribution
return main piping from the building outlet to point R.
7.56 CHAPTER SEVEN
Compared to a plant-distribution-building loop, a plant-distributed building loop has the following
advantages:
 It saves the installation cost of variable-speed distribution pumps, related controls, and pump
room space.
 If the pressure differential between the distribution supply and return mains at the building entrance
Hs,r is greater than the pressure loss of the building loop pbg either at design load or at
part-load, then a plant-distributed building loop can save more pumping energy.
On the other hand, a plant-distributed building loop requires variable-speed building pumps of
higher head which means more attention to pump noise control in the buildings mechanical room.
Multiple Sources-Distributed Building Loop
Many conversions and retrofits of campus-type chilled water systems require existing chilled water
plants in addition to the developed central plant. In these cases, there are buildings with chilled water
sources (chillers); buildings with chilled water coils and loads; and buildings with chilled water
sources and loads connected to the same plant-distributed building loop, as shown in Fig. 7.23a. A
WATER SYSTEMS 7.57
FIGURE 7.23 A multiple sources-distributed building loop. (a) Schematic diagram; (b) building with sources and loads.
DDC system controller is used for each type of building. There is a central microprocessor for the
whole chilled water system. The optimization program may proceed as follows:
1. Measure the chilled water temperature across each load and source, as well as its rate of water
flow.
2. Add all the loads.
3. Turn on the most efficient source first, including the demand and downtime, according to available
sources. A chillers running capacity should match the required loads. There should be a
time delay to start or stop the chiller.
4. Use trending and expert system control strategies to predict load changes according to past experience
and outdoor conditions.
Chilled and Hot Water Distribution Pipes
Chilled and hot water distribution pipes are large pipes mounted in underground accessible tunnels
or trenches. They are well insulated, although underground return chilled water mains for which return
temperature Tret  60F (15.6C) may not be insulated, depending on a detailed cost analysis.
Factory-made conduits consist of inner steel pipe, insulation, airspace, and outer conduit; or steel
pipe, with and without insulation, and outer casing. Expansion loops or couplings should be included,
and a good drainage system is important to protect the insulating quality. Please refer to
ASHRAE Handbook 1996, HVAC Systems and Equipment, chapter 11, for details on design and installation.
7.14 COMPUTER-AIDED PIPING DESIGN AND DRAFTING
General Information
According to Selecting Piping System Software by Amistadi (1994), three currently widely used
piping design and drafting computer programs were reviewed: University of Kentuckys
KYCAD/KYPIPE, Tranes Water Piping Design, and Softdesks Piping. The Trane and Softdesk
computer programs require AutoCAD as a base product whereas the University of Kentucky software
has its own integrated CAD system.
The University of Kentucky programs have extensive hydraulic modeling capacities including
transient analysis. They are intended for mechanical engineers and are used to design and analyze
large, complex water systems. Tranes and Softdesks software programs are limited to steady-state
imcompressible fluids. They are intended for contractors, mechanical engineers, and drafters. Kentuckys
and Softdesks programs support both inch-pound and metric units, while Tranes package
supports only inch-pound units. The University of Kentucky software requires 2.0 Mbytes of disk
space and 2.0 Mbytes RAM. Softdesks software requires 10 Mbytes disk space and 8 Mbytes
RAM. Tranes software requires the most disk space10 Mbytes and 12 Mbytes RAM. The
University of Kentucky program is MS-DOS applications, whereas Tranes and Softdesks are
available for Windows platforms.
Computer-Aided Drafting Capabilities
Tranes piping software allows designers to create a schematic two-dimensional (2D) piping system
in AutoCAD and to link to the computer programs of piping size calculations for use in piping system
design. Softdesks software is intended for drafting in 2D or 3D graphics. Design information
databases and engineering analysis software programs are linked to the drawing for support. The
7.58 CHAPTER SEVEN
University of Kentucky program is aimed for the design and diagnosis of large, complicated water
systems.
Softdesks software provides all the graphical components, such as pipes, valves, fittings, equipment,
pumps, tanks, controls, and structural components to form a piping system. Software by
Trane and the University of Kentucky only provides components that are needed for piping sizing
or hydraulic analysis. The University of Kentucky software provides annotation options for all hydraulic
parameters. Softdesk does not offer all, and Trane provides only selections of annotation.
Trane and the University of Kentucky only offer schematic layout and symbols to represent the
piping system and components and provide plan view engineering drawings as the standard option.
Softdesk offers schematic, double-line, and 3D model with wire-frame 3D version of its components.
Softdesk is the only computer program to translate automatically from single-line to 3D
model. The translation is bidirectional. It also allows the translation to go from 3D wire-frame to
double- or single-line. Softdesk offers a full range of engineering drawings including plan, section,
elevation, isometric, and perspective views.
Softdesks and University of Kentuckys software recognize most of the types of layouts, such as series,
parallel, branching, and network. The check of continuity is the basis of the hydraulic calculation,
which begins with the system continuity. The Softdesk software provides the most complete checking
of graphical elements of the piping system including gaps, overlaps, pump location, and size compatibility
of adjacent sections, followed by the University of Kentucky and Trane programs.
Computer-Aided Design Capabilities
System Size. The University of Kentucky software is for large piping systems and supports a system
up to 1000 legs. The Trane and Softdesk software support 400 legs.
Pipe Sizing. Tranes and Softdesks computer programs allow the designer many options based on
maximum head loss, such as 2.5 ft/100 ft (2.5 m/100 m) or velocity for pipe sizing. The University
of Kentucky computer program has extensive constraint capabilities that are linked to meet the
pressure at given node(s). A node is the junction of pipes and is the place where the flow rate
changes.
Pump and System Operations. Tranes and Softdesks programs allow the designer to first set the
system water flow rate and then calculate the pressure drop and the flow of the system components.
The University of Kentucky software is able to determine system operating points for series, parallel,
plant-building loop (primary and secondary), and variable-speed pumping applications by
means of pump and system curves. The University of Kentucky software also has the capabilities to
set control and pressure-regulating valves, or to locate check valves which affect the system hydraulic
calculations.
Pressure Losses and Network Technique. All three programs use the Darcy-Weisbach equation
with empirical fits to the Moody diagram to calculate pipe frictional losses. Unique pipe roughness
is used in Tranes and Softdesks calculations, but it can be varied for each pipe in the University of
Kentucky software. Tranes and Softdesks software uses equivalent length and Cv to calculate dynamic
losses for pipe fittings and control valves, whereas the University of Kentucky software
adopts the local loss coefficient k method. Tranes and Softdesks software read the drawing and automatically
places the node at points where there is a change in flow. The University of Kentucky
software expects the designer to input the nodes as fittings. All automatically number the nodes and
edit them if necessary.
Regarding piping network technique, Tranes and Softdesks software programs adopt sequential
method, a once-through stepwise approach, and use arithmetic sum of flow in parallel circuits to
determine the combined flow and the component head loss. The University of Kentucky program
uses the simultaneous method by solving simultaneous algebraic-equations through successive
approximation.
WATER SYSTEMS 7.59
Input Data and Reports. Tranes software relies on the AutoCAD attribute functions. Its data are
stored with the drawing data. Softdesks software uses extended entity capabilities, and data are
stored in external AutoCADs drawing exchange format (DXF) files. The University of Kentucky
software has a dedicated CAD system. Tranes and Softdesks software programs let the designer
select a kind of fluid and temperature, and the computer program calculates the fluid properties. The
University of Kentucky software requires the designer to input the data each time. Tranes and
Softdesks databases include size, cost, and hydraulic data, whereas the University of Kentucky
only includes hydraulic data.
All three computer programs provide tabular reports of pipe diameter, length, flow, velocity, and
head. Tranes software identifies the critical path, and Tranes and Softdesks software programs
offer quantity and cost bill of material part, whereas the University of Kentucky software offers
multiple design condition reports, such as cavitation and metering reports.
REFERENCES
Ahmed, O., Life-Cycle Cost Analysis of Variable-Speed Pumping for Coils Application, ASHRAE Transactions,
1988, Part I, pp. 194211.
Amistadi, H., Selecting Piping System Software, Engineered Systems, no. 6, 1994, pp. 5762.
ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996.
ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, GA, 1997.
ASHRAE, ASHRAE Handbook 1999, HVAC Applications, Atlanta, GA, 1999.
Avery, G., Microprocessor Control for Large Chilled Water Distribution Systems, Heating/Piping/Air Conditioning,
October 1987, pp. 5961.
Avery, G., Stethem,W. C., Coad,W. J., Hegberg, R. A., Brown, F. L., and Petitjean, R., The Pros and Cons of
Balancing a Variable Flow Water System, ASHRAE Journal, no.10, 1990, pp. 3055.
Ball, E. F., and Webster, C. J. D., Some Measurements of Water Flow Noise in Copper and ABS Pipes with
Various Flow Velocities, The Building Services Engineer, May 1976, pp. 3340.
Binkowski, R. O.,Water Treatment for HVAC Systems,/Heating/Piping/Air Conditioning, October 1989, pp.
131133.
Braun, J. E., Klein, S. A., Mitcell, J. W., and Beckman,W. A., Applications of Optimal Control to Chilled Water
Systems without Storage, ASHRAE Transactions, 1989, Part I, pp. 663675.
Burr, G. C., and Pate, M. E., Conversion of Campus Central Plant from Constant Flow to Variable Flow at
University of West Florida, ASHRAE Transactions, 1984, Part I B, pp. 891901.
Carlson, G. F., Hydronic Systems: Analysis and EvaluationPart I, ASHRAE Journal, October 1968, pp.
211.
Carlson, G. F., Central Plant Chilled Water SystemsPumping and Flow Balance Part I, ASHRAE Journal,
February 1972, pp. 2734.
Coad,W. J., Centrifugal Pumps: Construction and Application, Heating/Piping/Air Conditioning, September
1981, pp. 124129.
Ellis, R., and McKew, Howard, Back to Basics: Test 9Chilled Water System Using Centrifugal Chiller
Advanced Energy Efficient Design, Engineered Systems, no. 11, 1996, p. 11.
Eppelheimer, D. M., Variable FlowThe Quest for System Energy Efficiency, ASHRAE Transactions, 1996,
Part II, pp. 673678.
Griffith, D., Distribution Problems in Central Plant Systems, Heating/Piping/Air Conditioning, November 1987,
pp. 5976.
Haines, R. W., Bahnfleth, D. R., Luther, K. R., Landman,W. J., and Kirsner,W., Open for Discussion: Primary-
Secondary Pumping, HPAC, no. 3, 1997, pp. 6773.
Hansen, E. G., Parallel Operation of Variable-Speed Pumps in Chilled Water Systems, ASHRAE Journal, no. 10,
1995, pp. 3438.
Hull, R. F., Effect of Air on Hydraulic Performance of the HVAC System, ASHRAE Transactions, 1981, Part I,
pp. 13011325.
7.60 CHAPTER SEVEN
Kelly, D. W., and Chan, T., Optimizing Chilled Water Plants, HPAC, no. 1, 1999, pp. 145147.
Lizardos, E. J., Engineering Primary-Secondary Chilled-Water Systems, Engineered Systems, no. 8, 1995, pp.
3034.
MacDonald, K. T., Valves: An Introduction, Heating/Piping/Air Conditioning, October 1988, pp. 109117.
Mannion, G. F., High Temperature Rise Piping Design for Variable Volume Systems: Key to Chiller Energy
Management, ASHRAE Transactions, 1988, Part II, pp. 14271443.
Miller, R. H., Valves: Selection, Specification, and Application, Heating/Piping/Air Conditioning, October
1983, pp. 99118.
Ocejo, J., Program Estimates Expansion Tank Requirements, Heating/Piping/Air Conditioning, November
1986, pp. 8993.
Peterson, P. A., Medical Center Expands Utilities Distribution, Heating/Piping/Air Conditioning, May 1985,
pp. 8496.
Pompei, F., Air in Hydronic Systems: How Henrys Law Tells Us What Happens, ASHRAE Transactions, 1981,
Part I, pp. 13261342.
Prescher, R., Hydronic System Design Guidelines, Heating/Piping/Air Conditioning, May 1986, pp. 132134.
Redden, G. H., Effect of Variable Flow on Centrifugal Chiller Performance, ASHRAE Transactions , 1996, Part
II, pp. 684687.
Rishel, J. B., Energy Conservation in Hot and Chilled Water Systems, ASHRAE Transactions, 1983, Part II B,
pp. 352367.
Rishel, J. B., Twenty Years Experience with Variable Speed Pumps on Hot and Chilled Water Systems,
ASHRAE Transactions, 1988, Part I, pp. 14441457.
Rishel, J. B., Distributed Pumping for Chilled- and Hot-Water Systems, ASHRAE Transactions, 1994, Part I,
pp. 15211527.
Rishel, J. B., Use of Balance Valves in Chilled Water Systems, ASHRAE Journal, no. 6, 1997, pp. 4551.
Scientific Computing, Software Review: Up for Review (Again), Engineered Systems, no. 1, 1998, pp. 7684.
Solden, H. M., and Siegel, E. J., The Trend toward Increased Velocities in Central Station Steam and Water
Piping, Proceedings of the American Power Conference, vol. 26, 1964.
Stewart,W. E., Jr., and Dona, C. L.,Water Flow Rate Limitations, ASHRAE Transactions, 1987, Part II, pp.
811825.
Uglietto, S. R., District Heating and Cooling Conversion of Buildings, ASHRAE Transactions, 1987, Part II,
pp. 20962106.
Utesch, A. L., Variable Speed CW Booster Pumping, Heating/Piping/Air Conditioning, May 1989, pp. 4958.
Waller, B., PipingFrom the Beginning, Heating/Piping/Air Conditioning, October 1990, pp. 5171.
Wilkins, C., NPSH and Pump Selection:Two Practical Examples, Heating/Piping/Air Conditioning, October
1988, pp. 5558.
Zell, B. P., Design and Evaluation of Variable Speed Pumping Systems, ASHRAE Transactions, 1985, Part I B,
pp. 214223.
WATER SYSTEMS 7.61
CHAPTER 8
HEATING SYSTEMS,
FURNACES, AND BOILERS
8.1
8.1 HEATING SYSTEMS 8.1
Selection of a Heating System 8.2
8.2 WARM AIR FURNACES 8.3
Types of Warm Air Furnace 8.3
Upflow Gas-Fired Furnace 8.3
Horizontal Gas-Fired Furnace 8.6
Furnace Performance Factors 8.6
Saving Energy 8.7
Control and Operation 8.8
8.3 HOT WATER BOILERS 8.9
Selection of Fuel 8.9
Types of Hot Water Boiler 8.10
Fire-Tube Boiler 8.10
Scotch Marine Packaged Boiler 8.10
Cast-Iron Sectional Boiler 8.12
Gas and Oil Burners 8.13
Condensing and Noncondensing
Boilers 8.13
Boiler Efficiency 8.13
Chimney, or Stack 8.14
Operation and Safety Controls 8.14
8.4 ELECTRIC FURNACES, HEATERS, AND
BOILERS 8.15
Electric Heating Fundamentals 8.15
Electric Furnaces, Electric Heaters, and
Duct Heaters 8.16
Electric Hot Water Boilers 8.17
8.5 LOW-PRESSURE DUCTED WARM AIR
HEATING SYSTEMS 8.17
System Characteristics 8.17
Types of Low-Pressure Ducted Warm Air
Heating System 8.18
Heat Supplied to Conditioned
Space 8.18
Duct Efficiency and System Efficiency for
Heating 8.20
Location of Furnace and Duct
Insulation 8.20
Duct Leakage 8.20
Thermal Stratification 8.21
Part-Load Operation and Control 8.21
8.6 DUCTED WARM AIR HEATING
SYSTEMS 8.22
8.7 HOT WATER HEATING SYSTEMS
USING FINNED-TUBE HEATERS 8.23
Types of Hot Water Heating System 8.23
Two-Pipe Individual Loop System 8.23
Finned-Tube Heaters 8.24
Design Considerations 8.25
Part-Load Operation and Control 8.26
8.8 HYDRONIC RADIANT FLOOR-PANEL
HEATING SYSTEMS 8.27
System Description 8.27
Radiant Floor Panel 8.27
Thermal Characteristics of
Floor Panel 8.28
Design Nomograph 8.30
Design Considerations 8.30
Control and Operations of Multizone
Hydronic Radiant Floor-Panel
Heating System 8.30
System Characteristics and
Applications 8.31
8.9 INFRARED HEATING 8.31
Basics 8.31
Gas Infrared Heaters 8.32
Electric Infrared Heaters 8.32
Design and Layout 8.33
REFERENCES 8.35
8.1 HEATING SYSTEMS
According to the data in DOE/EIA Commercial Buildings Consumption and Expenditures 1995 and
Household Energy Consumption and Expenditures 1993 for the 54.3 billion ft2 (5.05 billion m2)
of heated commercial buildings and the 96.6 million households in the United States, the percentage
of use of various types of heating systems may be estimated as follows:
In commercial buildings, the use of a boiler as a primary heat source had a share of about 28 percent
to supply hot water or steam to the heating coils, radiators, baseboard heaters, and heating panels.
Heating systems in new constructed commercial projects that use boilers are mainly hot water
systems because these are energy-efficient, simpler to operate, and easier to maintain than steam
heating systems. In new constructed residential buildings, the heat sources are mainly gas-fired warm
air furnaces and heat pumps.
Heating systems that use finned coils to heat the air, including water heating coils, electric heating
coils, steam coils, or condensing heating coils of heat pumps in air-handling units and packaged
units, are convective or warm air heating systems. Heating systems that use high-temperature
infrared heaters, radiant panels, or radiators in which emited radiant heat exceeds the released convective
heat are radiant heating systems.
Selection of a Heating System
Several factors must be considered prior to selection of a suitable heating system. They include the
following:
 Whether it is a separate heating system or part of the air conditioning system
 An open or enclosed space, or a space with high infiltration
 Size of heating system (small, medium, or large)
 Available existing heat source, such as hot water or steam
 Cost of gas, oil, or electricity
 Design criteria and local customs
In both commercial and residential buildings, a warm air furnace is generally linked with
unitary packaged air conditioning systems and, therefore, packaged units. In buildings with
rooftop packaged units, a warm air heating system using a warm air furnace is often the
most direct, economical, and suitable choice. In locations where the outdoor climate is mild
in winter, the heat pump is often a convenient and energy-efficient heatsource. For buildings
with central hydronic air conditioning systems, a hot water heating system using a boiler is
often a suitable choice because the central primary plant and boilers are usually far away from
the air-handling units and conditioned space. According to DOE/EIA 0318 (1998), heat
pumps consumed the least annual energy use, district heating and boilers the most, the
others in between. ASHRAE/IESNA Standard 90.1-1999 specifies that radiant heating shall
be used when heating is required for unenclosed spaces except loading docks equipped with air
curtains.
Commercial, percent Residential, percent
Heating systems using boilers 28 33
Warm air furnaces and packaged heating units 24 37
Heat pumps 10 30
Individual space heaters including electric, gas,
and radiant heaters 28
District heating 10
8.2 CHAPTER EIGHT
8.2 WARM AIR FURNACES
Types of Warm Air Furnace
A warm air furnace is a combustion and heating device in which gas or oil is directly fired to heat
the air through a heat exchanger, or air is directly heated by electric resistance elements in order to
supply warm air to the conditioned space. Warm air furnaces can be classified into various categories
according to
 Fuel types: natural gas, liquefied petroleum gas (LPG), oil, electric energy, or wood
 Airflow directions: upflow, horizontal, or downflow
 Applications: residential, commercial, or industrial
Natural gas is the primary fuel used in warm air furnaces. Warm air furnaces used for residences
usually have a heating capacity up to 175,000 Btu/h (51.3 kW). Upflow models are the most popular
models in residences. For commercial applications, heating capacities are usually greater than
150,000 Btu/h (44 kW). Horizontal models mounted inside a rooftop packaged unit or packaged
heat pump for supplementary heating are widely used.
Upflow Gas-Fired Furnace
An upflow natural gas-fired warm air furnace consists of: a one or more gas burners, a heat exchanger,
a forced-draft circulating fan or blower, a venting system, a filter, and an outer casing, as
shown in Figs. 8.1a and b.
Gas Burners. Gas burners in small furnaces in residences are often atmospheric burners or fanassisted
burners. An atmospheric burner consists of an air shutter, a gas orifice, and outlet ports and
is usually die-formed and made of aluminum painted, heavy-gauge steel or aluminized steel (or
sometimes stainless steel). Atmospheric burners are either in-shot or upshot and are installed in
single or multiple ports. In-shot burners are installed horizontally and are also used for Scotch marine
boilers. Upshot burners are placed vertically and are suitable for vertical fire-tube boilers. Both
Scotch marine and fire-tube boilers are discussed in later sections. Atmospheric burners are simple,
require only a minimal draft of air, and maintain sufficient gas pressure for normal functioning. A
fan-assisted gas burner uses a small fan to induce the combustion air through the heat exchanger
and forces it to the outdoors via an air vent.
A power burner that uses a fan to supply and control combustion air is often employed for large
furnaces. Conversion burners are integrated with furnaces for safety and efficiency. Older gas furnaces
often use conversion burners.
Ignition. The ignition device is often a standing pilot ignition. These pilots are small. The pilots
flame is monitored by a sensor that shuts off the gas supply if the flame is extinguished. Another
type of ignition is called spark ignition. It ignites intermittently only when required. Spark ignition
saves more gas fuel than a standing pilot if the furnace is not operating.
Heat Exchangers. They are in the shape of clamshell, bent tube, or crossflow plate type. The
combustion air is burned and flows inside clamshells, or tubes. Air to be heated flows over the outer
surface of the heat exchanger. A heat exchanger is usually made of aluminum painted heavy-gauge
steel, aluminized steel, or stainless steel.
Circulating Fan or Blower. A circulating fan or blower is always installed in a warm air furnace
to force the air to flow over the heat exchanger, to distribute it to the conditioned space (except
HEATING SYSTEMS, FURNACES, AND BOILERS 8.3
when the warm air furnace is a part of an air-handling unit or packaged unit in which a supply
fan is always provided), and to extract warm space air for heating again. A centrifugal fan with
forward-curved blades and a double inlet intake is usually used.
Filter. A disposable filter, which is discussed in Chap. 17, is used to remove dust from the recirculating
air. It is often located upstream of the fan.
Venting Arrangements. In a natural-vent warm air furnace, a draft hood is employed to connect
the flue gas exit at the top of the heat exchanger section to a vent pipe or chimney. A relief air opening
is also used to guarantee that the pressure at the flue gas exit is always atmospheric. A draft
hood diverts the backdraft from the chimney, bypassing the burner without affecting the combustion
operation. A direct-vent warm air furnace does not have a draft hood. If the vent pipe or chimney is
blocked, a control system shuts down the warm air furnace. A power vent including a fan-assisted
combustion furnace uses a fan to force or induce the combustion products or flue gas flowing
through the heat exchanger and air vent.
Casing. The outer casing is usually made of heavy-gauge steel with removable access panels.
Furnace Operation. Gas is generally brought from the main to the pressure regulator. The regulator
reduces the gas pressure to about 3.5 in. WG (870 Pa). Gas then flows through a gas valve controlled
by a room thermostat. When a solenoid gas valve opens, the gas flows to the burners and
mixes with the necessary outside primary air for combustion. The primary airgas mixture then
8.4 CHAPTER EIGHT
FIGURE 8.1 Upflow gas-fired furnaces. (a) Natural vent; (b) fan-assisted combustion.
discharges from the port slots, mixes with ambient secondary air, and is burned. For a natural-vent
furnace, the combustion products flow through the heat exchanger and the draft diverter and
discharge outdoors through the vent pipe or chimney. For a fan-assisted combustion furnace, the
induced-draft fan extracts the combustion products and the flue gas flowing through the heat
exchanger, and forces them to discharge outdoors through the vent pipe or chimney.
A mixture of space recirculating air and outdoor air is pulled by the circulating fan from the
return ducts and enters the bottom inlet of the warm air furnace. This air is then forced through
the heat exchanger for heating again. For a heating-only furnace, the air temperature is raised by
50 to 80F (28 to 45C). Warm air discharges from the top outlet and is distributed to various
conditioned spaces through the supply duct and outlets. An upflow gas-fired furnace is usually
installed indoors and is often installed within a vented closet or vented basement in residential
buildings.
Most natural gas furnaces can use liquefied petroleum gas. The main difference is their gas
pressure. Natural gas usually needs a pressure of 3 to 4 in. WG (746 to 994 Pa) at the manifold,
whereas LPG needs a higher pressure of about 10 in. WG (2486 Pa) and more primary air for gas
burners.
HEATING SYSTEMS, FURNACES, AND BOILERS 8.5
Vent
(b)
Induceddraft
assisting
fan
Heat
exchanger
Gas
burner
Circulating fan
Warm air
supply plenum
Vent
FIGURE 8.1 (Continued).
Horizontal Gas-Fired Furnace
A typical horizontal gas-fired furnace in a rooftop packaged unit is shown in Fig. 8.2. It mainly
consists of multiple gas burners, a heat exchanger, a combustion blower, and an ignition device. The
supply fan used to force air to flow through the gas-fired furnace is the same supply fan used to
force air through the filters and DX coil in the rooftop unit.
A power burner is often used in a horizontal gas-fired furnace. This type of burner provides
better combustion and higher efficiency than atmospheric burners. A power draft centrifugal blower
may be added to extract the combustion products and discharge them to a vent or chimney. The gas
supply to the burner is controlled by a pressure regulator and a gas valve for the purpose of controlling
the firing rate. In a premix power burner, gas and primary air are mixed first; then the mixture
is forced to mix with secondary air in the combustion zone. A power burner usually has a higher gas
pressure than the atmospheric burner used in residences.
The heat exchanger usually has a tubular two-pass arrangement, typically with 16-gauge
(1.5-mm-thickness) stainless steel for primary surfaces and 18-gauge (1.2-mm-thickness) stainless
steel for secondary surfaces, as shown in Fig. 8.2. The primary surface is the heat-transfer surface
of the combustion chamber. The secondary surface is the surface of the tubes through which the
flue gas flows after the combustion chamber. A cone-shaped flame is injected into a tubular or
drum-shaped combustion chamber. A centrifugal blower is used to provide secondary air for forced
combustion. Another small centrifugal blower may be used to induce the flue gas at the exit of the
heat exchanger, the power vent, to maintain a negative pressure at the heat exchanger section so as
to prevent the mixing of any leaked flue gas with the heated air. The mixture of outdoor and recirculating
air is forced by the circulating blower and is heated when it flows over the primary and
secondary surfaces.
Furnace Performance Factors
The performance of a gas-fired furnace is usually measured by the following parameters:
 Thermal efficiency Et, in percent, is the ratio of the energy output of the fluid (air or water) to the
fuel input energy. Input and output energy should be expressed in the same units. The value of Et
can be calculated as
(8.1) Et 
100 (fluid energy output)
fuel energy input
8.6 CHAPTER EIGHT
FIGURE 8.2 Horizontal gas-fired furnace in a rooftop packaged unit.
 Annual fuel utilization efficiency (AFUE) is the ratio of annual output energy from air or water to
the annual input energy, expressed in the same units:
(8.2)
AFUE also includes nonheating-season standing-pilot input energy loss. AFUE is similar to thermal
efficiency Et, except that AFUE is the ratio of annual energy output to energy input, whereas Et is
the ratio of energy output to energy input at specific test periods and conditions.
 The steady-state efficiency (SSE) is the efficiency of a given furnace according to an ANSI test
procedure and is calculated as
(8.3)
The steady-state efficiency of gas-fired furnaces varies from 65 to 95 percent.
Test data in Jakob et al. (1986) for ASHRAE Special Project SP43, based on a nighttime setback
period of 8 h and a setback temperature of 10F (5.6C), gave the following performance factors for
gas-fired furnaces of two test houses with different construction characteristics:
Saving Energy
Based on the above results, factors that affect the energy saving of the fuel and improve the furnace
performance are as follows:
 Condensing or noncondensing. When the water vapor in the flue gas is condensed by indirect
contact with recirculating air, part of the latent heat released during condensation is absorbed by
the recirculating air, which increases furnace efficiency and saves fuel. The difference in the
values of AFUE between condensing and noncondensing power vent furnaces may be around
10 percent. Corrosion-resistant materials such as stainless steel must be used for condensing heat
exchangers due to the presence of chloride compounds in the condensate.
 Power vent or natural vent. Power vent guarantees the air supply for combustion and also forces
the combustion product flows through heat exchangers including condensing heat exchangers
with higher flow resistance.
 Preheated outdoor combustion air. If combustion air is taken from outdoors and preheated
through an annular pipe that surrounds the flue pipe, part of the heat in the flue gas will be saved.
 Automatic vent damper. An automatic thermal or electric vent damper located in the vent pipe
closes the vent when the furnace is not in operation. This decreases exfiltration from the house
and restricts the amount of heat escaping from the heat exchanger.
Construction characteristics AFUE, percent SSE, percent
Natural vent
Pilot ignition 64.5 77
Intermittent ignition 69 77
Intermittent ignition  venting damp 78 77
Power vent
Noncondensing 81.5 82.5
Condensing 92.5 93
SSE 
100(fuel input  fuel loss)
fuel input
AFUE 
100(annual output energy)
annual input energy
HEATING SYSTEMS, FURNACES, AND BOILERS 8.7
 Intermittent ignition. Intermittent ignition saves more energy than standing-pilot ignition, especially
during times when the furnace is not operating.
ASHRAE/IESNA Standard 90.1-1999 mandates the minimum efficiency of warm air furnaces,
warm air duct furnaces, and warm air unit heaters for sizes as follows:
In 1987, the National Appliance Energy Conservation Act (NAECA) set minimum efficiency
standards for gas furnaces at an AFUE of 78 percent, effective January 1, 1992. According to the
venting systems and the steady-state efficiency, ANSI/AGA Standard Z21.47 classifies the central
gas furnaces into the following four categories:
Here Tflue represents the vent gas (flue gas) temperature, and Tf, dew the dew-point temperature of the
flue gas, both in F.
Paul et al. (1993) reported that a category I fan-assisted gas furnace typically operated at an SSE
of 80 to 83 percent, resulting in AFUEs greater than 78 percent (a 78 percent AFUE corresponds to
an SSE of about 80.5 percent). Prior to 1987, the majority of gas furnaces installed in the United
States for residences were draft-hood-equipped models. Because of the effect of the NAECA, for
new installations and replacements, gas furnaces using natural vent draft hood systems will be gradually
phased out in favor of the fan-assisted power vent furnaces.
Control and Operation
Capacity. The heating capacity of a gas-fired furnace is controlled by a gas valve and ignition device.
For small furnaces the gas valve is usually controlled by a room thermostat that has on/off
control. Either standing-pilot or intermittent ignition may be used. For large furnaces a two-stage
gas valve, controlled by a two-stage thermostat, may be operated with a full gas supply, or at a reduced
rate when the outdoor weather is mild. Energy or fuel savings may not be significant unless
both the gas and the combustion air supply are controlled.
Venting system Tflue  Tf, dew,F Approximately SSE, percent
Category I Nonpositive (natural vent)  140 (noncondensing )  83
Category II Nonpositive (natural vent)  140 (condensing) 	 83
Category III Positive (power vent)  140 (noncondensing)  83
Category IV Positive (power vent)  140 (condensing) 	 83
Efficiency as of
Equipment type Size, Btu/h Rating conditions Minimum efficiency 10/29/2001
Gas-fired furnaces  225,000 78% AFUE or 80% Et 78% AFUE or 80% Et
 225,000 Maximum capacity 80% Et 80% Ec
Oil-fired furnaces  225,000 78% AFUE or 80% Et 78% AFUE or 80% Et
 225,000 Maximum capacity 81% Et 81% Et
Gas-fired duct furnaces All sizes Maximum capacity 78% Et 80% Ec
Minimum capacity 75% Et
Gas-fired unit heaters All sizes Maximum capacity 78% Et 80% Ec
Minimum capacity 74% Et
Oil-fired unit heaters All sizes Maximum capacity 81% Et 80% Ec
Minimum capacity 81% Et
8.8 CHAPTER EIGHT
Nighttime Setback. ASHRAE research project SP43 found that a nighttime thermostat setback
of 10F (5.6C) lower for 8 h improved Et slightly (only a 0.4 percent increase). However, there is a
10 to 16 percent annual savings in energy input compared to those furnaces without nighttime
setbacks.
Oversizing the furnace shortens the morning pickup time caused by nighttime setback. Use of an
oversized furnace has a significant effect on the swing of space air temperature when an on/off control
is used for the gas valve. Project SP43s results showed that when the furnace size corresponded
to 1.4 times the design heating load (DHL), the furnace had a space temperature swing of
4.9F (2.7C) and a morning pickup time of about 1 h. If the furnace size was based on 1.7 times
the DHL, the space temperature swing increased to 5.9F (3.3C), and the pickup time reduced to
about 0.5 h. A furnace size based on 1.4 times the DHL is more suitable for a nighttime setback
period of 8 h and a setback temperature of 10F (5.6C).
Fan Operation. In the past, continuous fan operation in small upflow gas-fired burners was said to
offer the benefits of better air circulation, reduced noise (because the fan did not start and stop), and
an even temperature distribution. Project SP43 showed that continuous fan operation resulted in a
higher furnace efficiency. However, continuous operation consumes more electricity than intermittent
operation in which the fan shuts off and the supply temperature drops below 90F (32.2C). In
many locations with a high electricity-to-fossil-fuel cost ratio, an energy cost analysis based on
AFUE may determine whether continuous or intermittent operation is more efficient.
The fan often starts about 1 min after the burner starts. Such a delay allows the heat exchanger
to warm up and prevents a flow of cold air. The fan will shut down 2 to 3 min after the burner is
shut off. The supply of residual heat from the heat exchanger also improves the performance of the
furnace.
Safety. For safety, the vicinity of the furnace should be free of combustible gas, vapor, and material.
Any passage to provide combustion air must be carefully planned. Gas and vent pipes should be
installed according to local and federal codes.
8.3 HOT WATER BOILERS
A hot water boiler for space heating is an enclosed pressure vessel in which water is heated to a
required temperature and pressure without evaporation. Hot water boilers are manufactured according
to the American Society of Mechanical Engineers (ASME) boiler and pressure vessel codes.
Boilers are usually rated according to their gross output heat capacity, i.e., the rate of heat delivered
at the hot water outlet of the boiler, in MBtu/ h, or thousands of Btu/h (kW). Hot water boilers are
available in standard sizes up to 50,000 Mbtu/h (14,650 kW).
Selection of Fuel
Natural gas, oil, coal, and electricity are energy sources that can be used in hot water boilers. It is
necessary to provide for an adequate supply during normal and emergency conditions and to take
into account the limitations imposed by any building and boiler codes for certain types of equipment
due to safety and environmental concerns. In addition, storage facilities and cost should be
considered before a fuel is selected.
Whereas natural gas and electricity are supplied by a utility, LPG, oil, and coal all need space
for storage within and outside the boiler plant.
Cost includes energy cost, initial cost, and maintenance cost. A gas-fired boiler plant requires
the lowest initial costs and maintenance costs, oil-fired boiler plants are moderately higher, and
coal-fired boiler plants are significantly higher (although their energy cost is lowest). An electric
boiler is simple to operate and maintain. In addition, it does not require a combustion process,
HEATING SYSTEMS, FURNACES, AND BOILERS 8.9
chimney, or fuel storage. In locations where electricity costs are low, electric boilers become
increasingly more attractive.
According to the data in the Commercial Building Characteristics 1992 by the EIA, the percentages
of floor area in all commercial buildings served by different kinds of primary energy source in
hot water and steam boilers in 1992 in the United States are as follows:
Gas-fired boilers 71 percent
Oil-fired boilers 15 percent
Electric boilers 11 percent
Others 2 percent
Types of Hot Water Boiler
According to their working temperature and pressure, hot water boilers can be classified as follows:
1. Low-pressure boilers. These hot water boilers are limited to a working pressure of 160 psig
(1103 kPag) and a working temperature of 250F (121C).
2. Medium- and high-pressure boilers. These boilers are designed to operate at a working pressure
above 160 psig (1103 kPag) and a temperature above 250F (121C).
A low-pressure hot water boiler is generally used for a low-temperature water (LTW) heating
system in a single building, regardless of the buildings size. Medium- and high-pressure boilers are
often used in medium-temperature water (MTW) and high-temperature water (HTW) heating systems
for campuses or building complexes in which hot water temperature may range from 300 to
400F (150 to 205C).
Based on their construction and materials, hot water boilers can also be classified as fire-tube
boilers, water-tube boilers, cast-iron sectional boilers, and electric boilers. Water-tube boilers,
mainly used for steam at higher pressure and temperature, are not discussed here. Electric boilers
are discussed in the next section.
Fire-Tube Boiler
A fire-tube boilers combustion chamber and flue gas passages are in tubes, which are all enclosed
in a shell filled with water. Heat released from the combustion process and the flue gases is absorbed
by the surrounding water, the temperature of which is increased to a required value.
Many kinds of fire-tube boilers have been developed. One of the more recently developed models
is known as the modified Scotch marine boiler, which is a compact and efficient design originally
used on ships. The Scotch marine boiler is probably the most popular hot water boiler manufactured
today.
Scotch Marine Packaged Boiler
A Scotch marine packaged boiler consists mainly of a gas, oil, or gas/ oil burner; a mechanical draft
system; a combustion chamber; fire tubes; and a flue vent. A packaged boiler is a one-piece, integrated,
factory-assembled boiler that includes a burner, outer steel shell, fire tubes, draft system, external
insulation, controls, interconnecting piping, and wiring. A schematic diagram of a typical
Scotch marine packaged boiler is shown in Fig. 8.3a and a cutaway illustration in Fig. 8.3b.
Flow Processes. Gas or oil and air are taken into the burner in measured quantities. The burner
then injects a combustible air-fuel mixture containing the necessary combustion air into the
8.10 CHAPTER EIGHT
HEATING SYSTEMS, FURNACES, AND BOILERS 8.11
FIGURE 8.3 Scotch marine packaged boiler. (a) Schematic diagram; (b) a cutaway photograph of a typical product.
(Source: Cleaver-Brooks. Reprinted with permission.)
combustion chamber, where it is initially ignited by an ignition device. The mixture burns, and the
combustion process sustains itself once the heat it generates is greater than the heat it transfers to
the water, i.e., once a high enough temperature in the combustion chamber is attained. The injected
air-fuel mixture is then burned spontaneously.
The direct radiation from the flame and the high temperature in the combustion chamber both
conduct heat through the wall of the chamberthe primary surface adjacent to the water that fills
the shell. The combustion product from the combustion chamberflue gasis directed into fire
tubes by headers. As the flue gas at a higher temperature flows through the fire tubes, it transfers
heat to the surrounding water through the pipewall of the tubesthe secondary surface. The flue
gas temperature drops, and its volume contracts. Because the number of fire tubes continuously decreases
in the second, third, and fourth passes of fire tubes, which matches the volume contraction
of flue gas, the gas velocity is maintained in a more uniform manner. Flue gas leaves after the
fourth pass and is vented into the stack.
Return water enters the side of the boiler, sinks to the bottom, is heated, and rises again. Hot
water is finally supplied at the top outlet. Such an arrangement prevents cold return water from
surrounding the combustion chamber and producing thermal shock. It also promotes good water
circulation.
Construction Characteristics. The constructional characteristics for Scotch marine packaged
boilers are as follows:
 Four flue gas passes. Each pass means a horizontal run of the fluid flow passage. A four-pass flow
arrangement for the flue gas, with a gradual reduction in gas flow area as the flue gas becomes
colder, helps maintain a higher gas velocity and, therefore, a clean surface and a higher rate of
heat transfer.
 Sufficient heat-transfer surface area. In a packaged Scotch marine boiler, the combustion chamber,
fire tubes, and tube sheets at both ends are all heating surfaces. The amount of heating surface
area directly affects a boilers output. And 3 to 5 ft2 (0.9 to 1.5 m2) of heating surface area for
each boiler horsepower, or 33,475 Btu/h (9.8 kW), of output, is the key value for providing suffi-
cient capacity and a long-lasting boiler. A smaller heating surface area for each boiler horsepower
of output often results in a higher heat flux, probably a higher surface temperature, and therefore a
shorter life for the boiler.
 Forced-draft arrangement. In a forced-draft arrangement, the fan is located adjacent to the burner.
The fan forces air into the combustion chamber. It also forces the flue gas to flow through the fire
tubes and to discharge from the vent or chimney. A forced-draft fan supplies a controllable quantity
of combustion air to the combustion chamber. The forced-draft fan is located upstream of the
combustion chamber, so it handles dense, clean boiler room air at a comparatively lower temperature
and lower volume flow rather than the hot, dirty, high-temperature, expanded flue gas downstream.
The temperature fluctuation for the room air is smaller than that for the exhausted flue
gas, so more accurate control of combustion air is possible than with an induced-draft design.
Cast-Iron Sectional Boiler
A cast-iron sectional boiler consists of many vertical cast-iron hollow sections in the shape of an
inverted U filled with water. When the sections are linked together by bolts, the center part of the
inverted U forms the furnace or the combustion chamber.
Because of its thick, heavy sections, a cast-iron sectional boiler has a large heat storage capacity
and thus is slow to heat up. This reduces temperature swings when heat demand varies. Thick,
heavy sections are also helpful in extending the boilers service life by preventing corrosion.
The heating capacity of a cast-iron sectional boiler depends on the number of sections connected.
Because of its relatively large combustion chamber and lower flow resistance, such a boiler
is able to use atmospheric gas burners with a lower chimney height. Cast-iron sectional boilers are
low-pressure boilers used in residences and in small and medium-size commercial buildings. They
can be field-assembled and fitted in existing buildings.
8.12 CHAPTER EIGHT
Gas and Oil Burners
When natural gas is used as the fuel in boilers, as in gas-fired furnaces, atmospheric and power gas
burners are usually used. Oil burners used in boilers differ from gas burners in that oil requires atomization
and vaporization prior to combustion. Combustion air is supplied by either natural drafts
or forced drafts. Ignition is usually provided by an electric spark or a standing pilot using gas or oil.
Oil burners used in residences are usually pressure atomizing-gun burners. This type of burner
either directly atomizes oil at 100 to 300 psig (690 to 2069 kPag) high pressure or uses combustion
air to atomize the oil. Combustion air is supplied by a blower, and electric spark ignition is
generally adopted. Oil burners for boilers in commercial buildings inject oil and atomize it into fine
sprays. Burners also force combustion air to mix with atomized oil and ignite the mixture, often
with an electric spark. Complete combustion can be sustained with as little as 20 percent excess air.
Sometimes a boiler can be designed to use either gas or oil. A combination of a ring-type gas
burner with an oil burner at the center is often used. Combination gas-oil burners are used more
often than single-fuel burners in large boilers.
Condensing and Noncondensing Boilers
In condensing boilers, the water vapor in the flue gas is condensed and drained by means of a heat
exchanger, as in gas furnaces. Thus, the latent heat of condensation can be recovered. In hot water
boilers, the cooling medium is usually the return water from the conditioned space. The lower the
temperature of the return water, the higher the amount of heat recovered. If the dew point of the flue
gas is 130F (54C), return water or service water at a temperature below 125F (51.7C) may be
used as the condensing cooling medium. Corrosion in the heat exchanger and flue gas passage
caused by the condensate should be avoided. Noncondensing boilers have no way to condense the
water vapor contained in the flue gas.
Boiler Efficiency
Boilers can be assessed by two efficiency values. The combustion efficiency Ec, in percent, is the
ratio of heat output from the hot water or steam Qout to the heat content rate of the fuel consumed
Qfuel, that is,
(8.4)
Both Qout and Qfuel are expressed in Btu/h (kW). Annual fuel utilization efficiency (AFUE) is also
an efficiency value used for hot water boilers. For noncondensing boilers, Ec varies from 80 to 85
percent. For condensing boilers, Ec varies from 85 to 90 percent.
ASHRAE/IESNA Standard 90.1-1999 mandates that the minimum efficiency requirements of
gas- and oil-fired hot water boilers are as follows:
Efficiency as of
Equipment type Size, Btu/h Rating condition Minimum efficiency 10/29/2001
Gas-fired boilers  300,000 Hot water 80% AFUE 80% AFUE
 300,000 and Maximum capacity 80% Ec 75% Et
	 2,500,000
	 2,500,000 Hot water 80% Ec 80% Ec
Oil-fired boilers  300,000 80% AFUE 80% AFUE
 300,000 and Maximum capacity 83% Ec 78% Et
	 2,500,000
	 2,500,000 Hot water 83% Ec 83% Ec
(Continued)
Ec 
100Qout
Qfuel
HEATING SYSTEMS, FURNACES, AND BOILERS 8.13
Chimney, or Stack
The chimney, or stack, is the vertical pipe or structure used to discharge flue gas. The breeching, or
lateral, is the part of the horizontal duct that connects the flue gas outlet of the boiler and the vertical
chimney. Breeching is commonly fabricated from 10-gauge steel (3.4-mm thickness) covered
with a high-temperature insulation layer. Flue gas discharge from hot water boilers that burn gas or
oil usually has a temperature rise of 300 to 400F (167 to 222C) over the ambient temperature.
Chimneys for burning gas or oil should be extended to a certain height above adjacent buildings according
to the local codes and topographical conditions. In high-rise buildings, locating the boiler
room at the basement level may be very expensive because of the space occupied by the vertical
stack and insulation inside the building.
Operation and Safety Controls
Heating Capacity Control. For hot water boilers, heating capacity control during periods of reducing
heat demand is achieved by sensing the return water temperature and controlling the firing
rate of the gas and oil burners. The firing rate of burners can be controlled in on /off, off / low/high,
and modulating modes. An on/off control mode is usually used for small boilers in which hot water
temperature control is not critical. A solenoid valve is used to open or close the fuel supply line to
the burner. An off / low/high control mode offers better output control than an on /off mode. If the
demand for heat is low, the boiler starts to fire at approximately one-half of its full capacity (the low
fire setting). Upon further increase in demand, the burner fires at its maximum capacity (the high
fire setting). A two-stage control valve and damper are used to control the supply of fuel and combustion
air to the burners.
With modulating control, the boiler starts at a low fire of about one-third of the full-load capacity.
As the demand for heat increases, an increasing amount of fuel and combustion air is supplied
to the burner through the modulation of either a butterfly valve in a gas burner or an orifice of variable
size in an oil burner. Both types of burner are linked to a damper to vary the supply of combustion
air. For gas burners, two pressure sensors are often provided to maintain the gas pressure
within a narrow range for proper operation. If one of the sensors detects an improper pressure, the
gas supply is cut off and the boiler shuts down.
To maintain an optimum air-fuel ratio, the amount of O2 in the stack flue gas can be maintained
at 1 to 5 percent by modulating the air damper of combustion air to provide higher combustion effi-
ciency. Usually, CO2 is measured and monitored in small boilers, and O2 is measured and monitored
in large boilers.
Safety Control. One or more pressure or temperature relief valves should be equipped for each
boiler. These are mechanical devices that open when the boiler pressure or temperature exceeds the
rated value. An additional limit control is installed to open the switch and shut down the boiler as
soon as the sensed water pressure or temperature exceeds the predetermined limit. When the pressure
or temperature falls below the limit, the switch closes and the firing starts again.
A flame detector is used to monitor the flame. When the flame is extinguished, the controller
closes the fuel valve and shuts down the boiler. An airflow sensor is also used to verify a continuous
supply of combustion air. Once the combustion air is not present, the fuel valve closes before a dangerous
fuel-air ratio can form.
Efficiency as of
Equipment type Size, Btu/h Rating condition Minimum efficiency 10/29/2001
Oil-fired (residual)  300,000 and Maximum capacity 83% Ec 78% Ec
	 2,500,000
	 2,500,000 Hot water 83% Ec 83% Ec
8.14 CHAPTER EIGHT
Modern packaged boilers often include a totally enclosed and factory-assembled controlled
cabinet that offers the latest microprocessor-based programmable direct digital controls, including
flame safeguard and control systems.
8.4 ELECTRIC FURNACES, HEATERS, AND BOILERS
Electric Heating Fundamentals
Except for hot water electric boilers using electrodes, most electric furnaces, electric heaters, and
electric heating coils used in HVAC&R installations are of the resistance type. When an electric
current flows through a resistor under electric potential, heat is released to the ambient air or water.
Electric resistance heaters of small wattage (e.g., an electric furnace in a residence with a heating
capacity of less than 8 kW) usually use single-phase 120/240-V supply. Large-capacity electric
resistance heaters for commercial and industrial applications usually use a three-phase 240/480-V
supply.
For a single-phase electric heater, the electric power input P, in W, which is equal to the heat
released from the resistor, can be calculated as
P  EI  I 2R (8.5)
where E  electric voltage, V
I  electric current, A
R  electric resistance, 
For a three-phase electric heater, the electric power input is given as
P  1.73El Il 1.73Il
2R (8.6)
where El  line voltage, V, and Il  line current, A. When a three-phase electric heater is in a delta
connection, as shown in Fig. 8.4b, then
and (8.7)
where Ep  phase voltage, V, and Ip  phase current, A. When a three-phase electric heater is in a
wye connection, then
and (8.8)
Several kinds of overload protection devices are used for electric heaters to open the circuit if
the electric current reaches a value that will cause a dangerous temperature:
 Fuses are alloys in the form of links to be inserted in the electric circuit. A fuse melts at a preset
temperature, opening the circuit, when an overload current passes through.
 Circuit breakers are mechanical devices that open their contacts, by means of a warp action of a
bimetallic strip or disk, when an excessive load passes through and heats them.
There are also thermal overload current-sensing devices and magnetic overload relays, both of
which open a circuit when the electric circuit is overloaded.
National Electrical Codes (NEC) and local codes must be followed during the design, installation,
and operation of electric heaters.
Ip  Il Ep 
El
1.73
Ip 
Il
1.73
Ep  El
HEATING SYSTEMS, FURNACES, AND BOILERS 8.15
Electric Furnaces, Electric Heaters, and Duct Heaters
Electric furnaces installed in rooftop packaged units, electric heaters equipped in air-handling units
and packaged units, electric duct heaters mounted in air systems, and electric unit heaters installed
directly above the conditioned space are all electric resistance warm air heaters. Air is forced
through the electric heating element by the supply fan in a rooftop packaged unit or in an airhandling
unit, or by the separately installed propeller fan in an electric unit heater.
Because the mixture of outdoor air and recirculating air or the outdoor air alone is clean and
nonhazardous after passing through the air filter, most electric resistance heaters use heating
8.16 CHAPTER EIGHT
FIGURE 8.4 Typical electric heater wiring diagram. (a) Single-phase line voltage control with thermal cutout and fan interlock;
(b) three-phase line voltage control with fan interlock and backup contactor.
elements made of bare wire of various alloys, such as 80 percent nickel and 20 percent chromium
wire. Bare wire is supported by bushes and brackets of insulating materials such as ceramics.
Sheathed elements are also used for better protection.
The face velocity of air flowing through warm air electric heaters in rooftop packaged units is
approximately the same as the face velocity of air at the cooling coil in AHUs. For a face velocity
of 500 to 600 fpm (2.5 to 3 m/ s), the temperature rise of warm air is often between 30 and 60F
(16.7 and 33.4C).
The total pressure loss of the warm air flowing through an electric resistance heater in AHU or
packaged units is usually less than 0.3 in. WC (75 Pa). For electric duct heaters, the air velocity and
pressure loss may be higher. Figure 8.4a shows a typical wiring diagram for an electric heater using
a single-phase electric power supply. Figure 8.4b shows a wiring diagram for a three-phase electric
power supply. In the design and installation of an electric duct heater, the following requirements
should be fulfilled:
 An electric duct heater should be installed at least 4 ft (1.2 m) downstream of the fan outlet,
elbow, or other obstructions. Otherwise, devices should be installed to provide for an even distribution
of airflow over the face of the heater.
 Do not install two electric duct heaters in series. Install electric duct heaters at least 4 ft (1.2 m)
from the heat pumps and air conditioners. Following these suggestions prevents excessively high
temperatures and the accumulation of moisture in the heater.
 A fan interlock circuit must be provided so that the electric heater does not operate unless the fan
is on.
 Automatic-reset overload cutout safety control should be placed in series with the thermostat. A
manual-reset overload cutout safety control is used in the backup safety control circuit. Limit
controls may be used instead of a manual-reset cutout control. Disconnecting devices should be
installed within sight of the electric heater.
 On each electric circuit, the maximum load of the electric heater is 48 A. The maximum fuse size
is 60 A. Divide a larger electric heater into smaller heaters with smaller loads. The loading of an
electric heater is often divided into stages for step control. Small heaters may be divided into two
or three stages; large electric heaters may have as many as 16 stages.
Electric Hot Water Boilers
There are two types of electric hot water boiler: resistance and electrode. Resistance electric boilers
use metal-sheathed heating elements, such as thin nickel chromium alloy wire, submerged in hot
water. Such safety valves as a relief valve, thermal overload cutout, and high-limit switch should be
provided. As with electric heaters, the electric load is divided into stages. The heating capacity is
controlled in steps by sensing the return water temperature. Resistance boilers are more commonly
used than electrode boilers. Resistance boilers are available in sizes with outputs up to about
3360 kW. Electrode boilers use water as a resistor. They are generally found in very large installations
and require high-quality water and may operate at high voltages.
8.5 LOW-PRESSURE DUCTED WARM AIR HEATING SYSTEMS
System Characteristics
A low-pressure ducted warm air (forced-air) heating system has the following characteristics:
 It is often equipped with an upflow gas-fired furnace, oil-fired furnace, or electric heater as well
as a centrifugal fan to force air through the furnace or electric heater.
HEATING SYSTEMS, FURNACES, AND BOILERS 8.17
 The external static pressure loss for the supply and return duct system is usually no greater than
0.5 in. WC (125 Pa).
 It uses a furnace heating capacity Qf to airflow ratio ranging from 50 to 70 Btu/hcfm (31 to
43 W s /L) and a temperature rise immediately after the furnace between 50 and 70F (28 and
40C).
 Because of the heat storage capacity of the supply duct system, the supply temperature differential
Ts  Tr is often 20 to 35F (11.1 to 19.4C). Here Ts represents the supply air temperature at
the supply register, and Tr indicates the space temperature, both in F (C).
 The heating system is often integrated with the cooling system to form a heating/cooling air
conditioning system.
 Most low-pressure warm air heating systems have a capacity no greater than 100,000 Btu/h
(29.3 kW).
Low-pressure ducted warm air heating systems are usually used in residences and sometimes in
small commercial buildings.
Types of Low-Pressure Ducted Warm Air Heating System
Low-pressure ducted warm air heating systems can be divided into two categories according to
their duct systems:
 Supply and return duct system
 Supply duct and return plenum
Figure 8.5a shows a typical low-pressure ducted warm air heating system with a supply and return
duct system for a single-family house. Recirculating air from the living room, dinning room, and
bedrooms flows through return ducts and enters the lower part of the upflow gas furnace. It mixes
with outdoor air. The mixture is forced through a gas furnace and heated to a required supply temperature
at the supply plenum. Warm air is then supplied to various rooms through supply ducts and
registers.
Figure 8.5b shows a typical low-pressure ducted warm air heating system with supply ducts and
a return plenum. There are no return ducts. Warm air is supplied to various rooms through supply
ducts. Recirculating air flows back to the return plenum by means of undercut or louvered doorways
and corridors. Once an interior door between rooms is closed, air squeezes through the door
undercuts and louvers to the return plenum as well as other cracks and gaps to the attic or outdoors.
The room is therefore maintained at a positive pressure. This causes a reduction in the supply flow
rate and an increase in the infiltration rate of the whole house.
Field tests performed by Cummings and Tooley (1989) on five Florida houses show that, for a
low-pressure ducted warm air heating system with supply ducts and a return plenum, the average
whole-house infiltration was 0.31 ach (air change per hour) when all the interior doors between
rooms were open. If the interior doors were closed, the average whole house infiltration increased to
0.91 ach. The typical pressure difference between rooms separated by a closed door was 0.032 in.
WC (8 Pa), for a door undercut of 0.5 in. (13 mm).
Heat Supplied to Conditioned Space
Heat supplied to the conditioned space by a low-pressure ducted warm air heating system, denoted
by Qsh, Btu /h (W), can be calculated as
(8.9) Qsh  60V?s 
scpa(Ts  Tr)
V?
a
8.18 CHAPTER EIGHT
where volume flow rate of supply air, cfm [m3 / (60 s)]
Ts, Tr  temperature of supply and space air, F (C)
In Eq. (8.9), 
s represents the density of supply air. For a supply air temperature of 100F (38C)
and a relative humidity around 20 percent, 
s can be taken as 0.07 lb / ft3 (1.12 kg/m3). The term cpa
indicates the specific heat of moist air. For simplicity, its value can still be taken as 0.243 Btu/ lb F
(1017 J/kgC).
V?
s 
HEATING SYSTEMS, FURNACES, AND BOILERS 8.19
FIGURE 8.5 Typical low-pressure ducted warm air heating system. (a) Supply and return duct; (b) supply
duct and return plenum.
Duct Efficiency and System Efficiency for Heating
Duct efficiency for heating du, h, in percent, can be calculated by dividing the heat energy output
from the supply and return duct system by the heat energy input to it, i.e.,
(8.10)
where qho, s, qho, r  output heat energy from supply ducts and return ducts, Btu/h (W)
qhi, s, qhi, r  input heat energy to supply ducts and return ducts, Btu/h (W)
qhl, s, qhl, r  heat loss from supply ducts and return ducts, Btu/h (W)
and the supply duct efficiency for heating sd, h , in percent, is
(8.11)
The system efficiency for heating sy, h, in percent, can be calculated as
(8.12)
where Qf, h  jacket loss and equipment losses from furnace and released to conditioned space,
Btu/h (W)
Qf, in  total energy input to furnace, including auxiliary energy input, Btu/h (W)
Location of Furnace and Duct Insulation
The location of the furnace has a significant effect on the system efficiency of heating. In Jakob et al.
(1986), if the gas furnace is installed in a closet, the supply duct is mounted inside the conditioned
space, and the equipment losses of the furnace become the direct heat gains of the conditioned space,
then sy, h might be 20 percent higher than for those installations where the furnace and supply ducts
are in the attic or basement.
Jakob et al. (1986) also showed that if ducts had an exterior or interior insulation of R5, that
is, an insulation layer with an R value of 5 h  ft2  F/Btu (0.9 m2  C/W), the duct efficiency increased
from 61 percent (without duct insulation) to about 78 percent. The system efficiency saw
a smaller increase because a portion of the heat loss from the duct without insulation had been
used to heat the attic or basement, which in turn reduced the heat loss from the conditioned
space.
Duct Leakage
ASHRAE Standard 90.1-1999 mandates the minimum duct seal level as discussed in Sec. 17.8.
Field tests in many houses have showed that actual duct leakage in many low-pressure ducted warm
air heating systems is considerably higher.
Lambert and Robinson (1989) analyzed the duct leakage, whole-house leakage, and heat energy
use of 800 electric-heated houses built since 1980 in the Pacific northwest. Tested houses were
 100 Et du, h 
Qf, h
Qf, in 
hsy, h  100
qho, s  Qf, h
Qf, in
sd, h  100
qho, s
qhi, s
 100
1  qhl, s
qhi, s
 1001 
qhl, s  qhl, r
qhi, s  qhi, r 
du, h  100
qho, s  qho, r
qhi, s  qhi, r
8.20 CHAPTER EIGHT
divided into two groups: highly energy-efficient Model Conservation Standard (MCS) houses and a
control group built to current (regional) practice (CP) or standards. Ducted heating systems in CP
houses had 26 percent more air leakage than unducted systems and used 40 percent more heating
energy. Ducted heating systems in MCS houses had 22 percent more air leakage and used 13 percent
more heating energy than unducted MCS houses.
Gammage et al. (1986) studied the ducted warm air systems of 31 Tennessee houses. They
found that the air infiltration rate was 0.44 ach when the forced-air systems were off and 0.78 ach
when the forced-air systems were on.
Field tests have also shown that leakages are greater for the return duct than for the supply
duct because of the greater importance of the supply air. Also, the return plenum is often not
carefully sealed. In five Florida homes with low-pressure ducted warm air systems with supply
and return plenums, Cummings and Tooley (1989) found that when repairs were made to seal
the return plenums, the infiltration in these five homes dropped from an average of 1.42 ach to
0.31 ach.
When a ducted warm air heating system is operating, supply duct leakage in such nonconditioned
space as the attic or basement raises the space pressure to a positive value and promotes
exfiltration. Return duct leakage extracts the space air, lowers the space pressure to a negative
value, and promotes infiltration. Both types of leakage increase the whole-house infiltration.
Suggested remedies to reduce duct leakage and energy use are as follows:
 Externally seal the duct with tape; seal the ducts internally if possible.
 Seal the return plenum and equipment if there is leakage. Seal the duct, pipes, and cable penetrations
through the structures.
 Avoid locating ducts in unconditioned spaces. Provide insulation for ducts running in unconditioned
spaces.
 For a low-pressure ducted warm air system without return ducts, an adequate door undercut or
door louver should be provided.
Thermal Stratification
A low-pressure ducted warm air heating system with a gas-fired furnace always has a high supply
temperature differential Ts. If Ts exceeds 30F (16.7C), or if there is a high ceiling, thermal
stratification may form in the conditioned space. The vertical temperature difference may be greater
than 5F (2.8C) and cause discomfort, as mentioned in Sec. 4.8. In addition, a higher temperature
near the ceiling may increase heat transfer through the ceiling, attic, and roof.
A greater supply air volume flow rate, lower Ts, higher downward air jet velocity, and suitable
location for the supply grille are remedial measures that reduce thermal stratification and vertical
temperature differences.
Part-Load Operation and Control
For low-pressure ducted warm air heating systems, a thermostat or a temperature sensor is usually
installed in a representative space to control the gas valve of a furnace operating under an on/off or
two-stage step control mode to maintain the required space temperature. The proportion of on and
off time in each operating on/off cycle varies to meet the varying space heating load. Figure 8.6
shows a typical on/off control for a gas furnace, resulting in a temperature variation for a point
20 ft (6.1 m) downstream of the supply plenum and at the supply register.
When the space temperature increases above the upper limit as the space heating load falls, the
controller shuts off the gas valve so that the heat supply to the space is cut off. When the space temperature
falls below the lower limit, the gas valve will open again and raise the space temperature.
HEATING SYSTEMS, FURNACES, AND BOILERS 8.21
The time period for an on/off operating cycle is generally between 5 and 15 min. Too short a cycle
may result in unstable operationa condition known as hunting.
8.6 DUCTED WARM AIR HEATING SYSTEMS
Ducted warm air heating systems are heating systems that are part of the air-handling unit, packaged
units, packaged heat pumps, or fan-coil units that use water heating coils, gas furnaces, electric
heaters, or indoor coils in heat pumps to heat the air. Warm air is then supplied to the conditioned
space through the duct system.
A ducted warm air system has the following characteristics:
 It is often integrated with a cooling system, forming a heating/cooling system.
 Its external pressure loss for a duct system is usually between 0.5 and 2 in. WC (125 and 500 Pa),
except for fan-coil units.
 The heating capacity is controlled by modulating the water flow rate to the water heating coil, by
modulating the gas valve, or by controlling the refrigeration flow in an on/off or step control
mode in the heat pumps.
 Heating systems often have a capacity greater than 100,000 Btu/h (29 kW).
Because the perimeter zones of office buildings need strict control of the vertical temperature
difference in their conditioned space, a smaller supply air temperature differential, such as Ts 	
8.22 CHAPTER EIGHT
FIGURE 8.6 Typical on/off control of a gas furnace and the variation of temperature of supply plenum and supply register.
15F (8.3C) is usually used. The performance and design of these types of heating and cooling
systems will be discussed in later chapters.
8.7 HOT WATER HEATING SYSTEMS USING
FINNED-TUBE HEATERS
Types of Hot Water Heating System
Like hot water boilers, hot water heating systems can be classified according to their operating temperature
into two groups:
 Low-temperature water systems. These operate at a temperature not exceeding 250F (120C),
typically 190F (88C) supply and 150F (65C) return, and a maximum working pressure not
exceeding 150 psig (1034 kPa g), usually less than 30 psig (207 kPa g).
 Medium- and high-temperature water systems. In medium-temperature water systems, the operating
temperature is 350F (177C) or less, and the operating pressure is 150 psig (1034 kPa g) or
less. In high-temperature water systems, the maximum operating temperature can be from 400 to
450F (205 to 232C), and the maximum operating pressure is 300 psig (2070 kPag). In both
medium- and high-temperature systems, the hot water supplied to and returned from the heating
coils and space heaters is typically 190 and 150F (88C and 65C), respectively.
Low-temperature hot water heating systems are widely used for space heating in residential and
commercial buildings. Medium- and high-temperature hot water heating systems are often used in
central heating plants for university campuses or groups of buildings, or in industrial applications
for process heating. Only low-temperature hot water systems are discussed in this section.
In low-temperature hot water heating systems, the two-pipe individual loop and radiant floorpanel
heating systems are currently used in residential, commercial, and industrial buildings.
Two-Pipe Individual Loop System
Most current low-temperature hot water heating systems that use finned-tube heaters are equipped
with zone control facilities. Without such a control system, rooms that face south may be overheated
and rooms that face north may be underheated in northern latitudes because of the effects
of solar radiation. The typical piping arrangement for a low-temperature hot water heating system
is the two-pipe individual-loop system, as shown in Fig. 8.7. On a cold winter day, water in a typical
individual-loop system returns from various finned-tube baseboard heaters through different
individual return loops at a temperature between 150 and 155F (65 and 68C). It is heated at the
hot water boiler to a temperature of 190F (88C), extracted by the on-line circulating pump on
each individual loop, and distributed to the finned-tube baseboard heaters by means of supply and
branch pipes. Because of the higher average temperature of hot water in the baseboard heaters,
heat is released to the space air to offset the heating load by means of radiation and convection in
various control zones, such as north, south, east, and west zones. The temperature of hot water
then drops to between 150 and 155F (65 and 68C) and returns to the hot water boiler again via a
return main.
In Fig. 8.7, there are several individual loops, or control zones. In each individual loop, several
finned-tube baseboard heaters in a large room can be connected in series. Finned-tube baseboard
heaters in small rooms belonging to the same individual loop can be connected in a reverse-return
arrangement. In such an arrangement, the length of pipeline that hot water travels is nearly the same
for each baseboard heater for a better system balance.
If a hot water heating system using finned-tube heaters is part of an air conditoning system, then
the outdoor ventilation air during winter heating will be provided for the occupants in the heating
HEATING SYSTEMS, FURNACES, AND BOILERS 8.23
space by the air conditioning system that cools the conditioned space in summer. In locations
of cold climate, if only winter heating is required, then a separate outdoor ventilation system
sometimes is required instead of opening the windows manually. Outdoor ventilating systems are
discussed in Chaps. 23 and 24.
The design procedure for a hot water heating system involves the following:
 Calculation of the space heating load, as discussed in Chap. 6
 Selection of suitable finned-tube heaters
 Division of the heating space into various control zones, or individual loops
 Planning of a piping layout, containing branches and necessary piping fittings
 Location of the boiler, circulating pumps, air separator, and expansion tank in the mechanical
equipment room, and determination of their capacities
 Specification of the control functions of the capacity and safety control systems, including the
sequence of operation
Finned-Tube Heaters
A finned-tube heater is a terminal unit installed directly inside the conditioned space to deliver heat
energy to the space by means of radiation and convection. A finned-tube heater consists of a finnedtube
element and an outer enclosure, as shown in Fig. 8.8. The tubes are often made of copper and
steel. Copper tubes are usually of 0.75-, 1-, and 1.25-in. (19-, 25-, and 32-mm) diameters, and steel
tubes of 1.25- and 2-in. (32- and 50-mm) diameters. For copper tubes, fins are often made of aluminum.
For steel tubes, the fins are made of steel. Fin spacing varies from 24 to 60 fins per foot (79
to 197 fins per meter). A finned-tube heater may have a length up to 12 ft (3.6 m).
8.24 CHAPTER EIGHT
FIGURE 8.7 A two-pipe individual-loop low-temperature hot water heating system for a factory.
Although various configurations of the outer enclosure have been designed and manufactured,
each enclosure must have a bottom inlet and a top outlet for better convection. The enclosure is usually
made of 18-gauge (1.2-mm-thickness) steel with corrosion-resistant coating to provide protection
and improve appearance. To allow a higher heating capacity, two finned-tube elements can be
set in a two-tier (two-row) arrangement. The most widely used finned-tube heater is the baseboard
heater, which is often installed at a level of 7 to 10 in. (175 to 250 mm) above the floor. It is usually
3 in. (75 mm) deep and has only one pier of finned-tube elements. Baseboard heaters are usually
mounted on cold walls and release heat nearly at the floor level. They also interfere less with indoor
decor than other heaters.
A wall finned-tube heater has a greater height and is available in various shapes to meet the
architectural interior design. A convector has a cabinet-type enclosure with one or two tiers of
finned-tube heating elements installed under the window sill. The rated heating capacity of a
finned-tube heater depends mainly on its length, the fin spacing, the average water temperature
flowing through the heating element, and the temperature of air entering the heater. Table 8.1 lists
the rated heating capacity of typical finned-tube heaters for entering air at 65F (18.3C) and different
average hot water temperatures.
Design Considerations
In the older low-temperature hot water heating system design, a water temperature drop of Tw 
20F (11.1C) at the finned-tube element was usually used. The current trend is to use a greater
temperature drop, such as 20 to 50F (11.1 to 27.8C). A detailed cost analysis should be performed
to determine the optimum temperature drop in a finned-tube heater.
The sizing of low-temperature hot water pipes is usually based on a pressure drop of 0.5 to
1.5 psi (1 to 3 ft head loss per 100 ft or 1 to 3 m head loss per 100 m of pipe length). Friction charts
for hot water steel pipes and the equivalent length of pipe fittings are discussed in Chap. 7.
HEATING SYSTEMS, FURNACES, AND BOILERS 8.25
FIGURE 8.8 Baseboard finned-tube heater.
TABLE 8.1 Heating Capacity of Finned-Tube Elements for Entering Air Temperature of 65F, Btu/h ft
Average water temperature, F
Type of finned-tube heater No. of rows 220 210 200 190 180 170
Steel tube: 1.25-in. dia., steel fin 1 1260 1140 1030 940 830 730
2 2050 1850 1680 1520 1350 1190
Copper tube: 1-in. dia., aluminum fin 1 1000 900 820 740 660 580
2 1480 1340 1210 1100 970 860
Source: Adapted with permission from Handbook of HVAC Design 1990.
For small low-temperature hot water heating systems, an open expansion tank is usually used.
For medium-size and large systems, a diaphragm tank may be more suitable. Circulating water
pumps are often on-line pumps (circulators) with low pump head.
Part-Load Operation and Control
In a low-temperature hot water heating system, one of the basic part-load controls is the variation of
the hot water supply temperature from the boiler in response to a variation in outdoor temperature.
For instance, in the midwest, a low-temperature hot water heating system has a winter design
outdoor temperature To  0F (17.8C). At the winter design outdoor temperature, the hot water
supply temperature Tws is 190F (88C). When the outdoor temperature drops, Tws is reset as follows:
A hot water temperature sensor located at the hot water exit of the boiler, whose set point is
reset by an outdoor temperature sensor, is used to control the firing rate of the boiler by means of a
DDC unit. Zone control can be performed better by sensing the hot water temperature that returns
from each individual loop and then modulating the control valve to vary the mass flow rate of hot
water supplied to that zone through an on-line circulating pump (see Fig. 8.7).
ASHRAE/IESNA Standard 90.1-1999 specifies that a hot water system with a design heating
capacity exceeding 300,000 Btu/ h (88 kW) shall include controls to automatically reset to a lower
hot water supply temperature according to building loads or outdoor temperature during part-load
for energy saving except for systems that use variable flow to reduce pumping energy.
For a low-temperature hot water heating system installed with multiple boilers, the control strategy
is to decide when to turn a boiler on or off. This optimum operation control can be accomplished
by using a microprocessor-based controller that fires a standby boiler according to a preprogrammed
software instruction. This strategy depends on not only the increase or reduction in heating demand
but also how the operating cost of such a hot water heating system can be minimized.
ASHRAE/IESNA Standard 90.1-1999 also specifies that for a boiler plant equipped with more than
one boiler, provisions shall be made so that the hot water flow in the boiler plant can be automatically
reduced when a boiler is shut down.
Example 8.1. A two-pipe individual-loop low-temperature hot water heating system is used to
heat a factory that has a layout shown in Fig. 8.7. At winter design conditions, hot water is supplied
to the heated space at a temperature of 190F (88C) and returns from the baseboard finned-tube
heaters at a temperature of 150F (65C).
1. If the space heating load for the largest room (facing north) is 90,000 Btu/h (26.4 kW) and for
the northwest corner room is 11,000 Btu/h (3.2 kW), and if steel tubes and fins are used, determine
the number of feet of finned tubing required for each of these two rooms.
2. If a pressure drop of 1 ft/100 ft (1 m/100 m) of pipe is used and the hot water system is
equipped with an open expansion tank, determine the diameter of the hot water supply main for
these two rooms.
3. Divide this hot water system into appropriate control zones, or individual loops.
Solution
1. For the largest room, if a two-row finned-tube heater is used, then from Table 8.1, for an average
hot water temperature of (190  150)/2  170F (77C), the heat output of each foot of two-row
Outdoor temperature To, F (C) Supply temperature Tws, F (C)
0 (17.8) 190 (88)
32 (0) 135 (57)
60 (16) 85 (29)
8.26 CHAPTER EIGHT
finned tube is 1190 Btu/h (350 W). The number of feet required is therefore
For the northwest corner room, a single-row finned-tube heater is used. The number of feet required
is
Along the north sides external wall, an 8-ft (2.5-m) finned tube is mounted. On the west sides
external wall, a 7-ft (2-m) finned tube is mounted.
2. Because the density of water at 170F is 60.8 lb / ft3, from Eq. (7.1), the flow rate of hot water
supplied to the largest room and the northwest corner room can be calculated as
Then from the water friction chart for water in steel pipes (Fig. 7.2 in Chap. 7), for a pressure
drop of 1 ft/ 100 ft, the pipe diameter for the hot water supply main to the largest room and
northwest corner room is 1.25 in. (32 mm).
3. This two-pipe individual-loop low-temperature hot water heating system should be divided into
four zones: north, south, east, and west. Each zone, or individual loop, includes the rooms
whose outer walls face that direction (except the south zone, which also includes interior rooms
without external walls). Such a setup can offset the variation in solar radiation and the operating
conditions for each zone to prevent overheating of rooms facing south and underheating of
rooms facing north. For each individual loop, or control zone, a thermostat senses the return hot
water temperature and modulates the control valve and the flow rate of hot water supplied to that
loop by a DDC unit in order to meet the variations in zone heating load.
8.8 HYDRONIC RADIANT FLOOR-PANEL HEATING SYSTEMS
System Description
A hydronic (hot water) radiant floor-panel heating system provides heating and thermal comfort to
the occupants in an enclosed indoor environment primarily by means of the radiant heat transfer
from the floor panels. As with the hot water systems using finned-tube heaters, a hydronic radiant
floor-panel heating system consists of a hot water boiler, several individual-loop pipings with radiant
floor panels, an air separator, an expansion tank, circulating hot water pumps, control valves,
and accessories, as shown in Fig. 8.9a and b. Instead of finned-tube heaters, radiant floor panels are
used. For each individual loop that forms a control zone, there may be only one large radiant floor
panel, or several small panels connected in series.
Radiant Floor Panel
Tubes made of plastic, rubber, steel, or copper of diameters from 0.3875 to 1.0 in. (10 to 25 mm)
are often embedded in a floor slab with a tube spacing from 6 to 15 in. (150 to 375 mm) to form a
heating floor panel, as shown in Fig. 8.9c. Tubes are laid in the shape of serpentine to ensure an
even panel surface temperature. Floor slabs are often made of concrete with a thickness between

90,000  11,000
500(190  150)
 5.05 gpm
Vgal 
Qw
500(Twe  Twl)
11,000
730
 15.06 ft (15 ft, or 4.5 m)
90,000
1190
 75.63 ft (76 ft, or 23 m)
HEATING SYSTEMS, FURNACES, AND BOILERS 8.27
2 and 6 in. (50 and 150 mm). The thickness of the floor slab has a significant effect on its thermal
storage and slow response to heating load changes. Floor coverings like carpet and wood affect the
overall thermal resistance of the floor panel. The water temperature, therefore, should be increased
accordingly in order to provide the required amount of heat that is released from the floor panel. An
insulating layer and a moisture barrier under the concrete slab are often needed to reduce the heat
losses, as shown in Fig. 8.9c.
Thermal Characteristics of Floor Panel
According to ASHRAE Handbook 1996, HVAC Systems and Equipment, the heat flux of the radiant
heat transfer irradiated from a heating floor panel qr, Btu/h ft2 (W/m2), can be calculated as
(8.13)
where TRp  absolute temperature of effective panel surface,R (K)
TR,AUST  absolute temperature of area-weighted average temperature of unheated surfaces exposed
to panels (AUST),R (K)
The heat flux due to natural convection flows upward from a heated floor panel qc, Btu/h ft2 (W/m2), is
(8.14) qc  0.31(Tp  Ta)0.31(Tp  Ta)
qr  0.15  108(T 4
Rp  T 4
R,AUST )
8.28 CHAPTER EIGHT
(a)
(c)
(b)
DDC
unit controller
Outdoor
Supply
sensor
Supply main
Return main Mixing
valve
Boiler
Return
sensor
Plastic
tube
Concrete
Boiler
Floor
panel
(zone 1)
Pump
Thermostat
Floor
panel
(zone 2)
TO
TS
T1
TR
Control
value
T2
T1 T2
Insulation
layer
FIGURE 8.9 Hydronic radiant floor heating systems. (a) Zone control valves; (b) outdoor reset plus zone control valves; (c) floor panel;
(d) design nomograph. (Source: ASHRAE Transactions 1995 Part I, p. 1204, Kilkis and Coley. Reprinted by permission.)
where Tp  effective panel surface temperature, F (C)
Ta  space air temperature that surrounds floor panel, F (C)
Then the heat flux from the radiant heating floor panel to the space air due to the combined effect of
radiation and convection qu, Btu / h ft2 (W/m2), can be calculated as
qu  qr  qc (8.15)
If the tubes are embedded in the concrete slab with a tube spacing of 1 ft (12 in. or 0.3 m), the characteristic
panel R value ru, h ft2 F/Btu (m2 C/W), can be calculated as
(8.16)
where rt  thermal resistance of tube wall per unit tube spacing, h ft2 F/Btu (m2C/W)
M  tube spacing, ft (m)
rp  R value of floor panel, h ft2 F/Btu (m2  C/W)
rc  R value of floor-panel covering, h  ft2  F/Btu (m2  C/W)
ru  rt M  rp  rc
HEATING SYSTEMS, FURNACES, AND BOILERS 8.29
(AUST-Ta) F
8
0.1 0.2 0.3 0.4 0.5
0.6
0.7
0.8
0.9
1.0
1.2
1.4
1.6
1.8
2.0
2.5
3.0
4.0
6 4 2 0 2 Ceiling heating
Floor heating
P S
R
Use when
rc /rp ? 4
Floor
heating
0F
AUST 10
20
30 40 50
50
60
70
80
90
100 110
60
70 80 90 100 110
Ceiling
M
ru  rprcrtrs
Panel resistance line
ru  0.05
xp
xp
qu
Do DI
Ta
Tp
qd
Floor
Temperature difference (TpTa), F
Heating flux qu, Btu/hft2
40
35
30
25
20
15
10
5
0 0 0
10
20
30
40
50
60
70
6
9
M12
15
18
50
45
40
35
30
20
25
15
10
5
(d)
FIGURE 8.9 (Continued)
Design Nomograph
Kilkis and Coley (1995) noted that the mean hot water temperature needed to maintain a required
heat flux from the heating panel depends primarily upon the heating panels effective surface temperature,
its tube spacing, and its thermal resistance. Kilkis and Coley also recommended a design
nomograph for the heating floor panel to determine the required design parameters, as shown in
Fig. 8.9d, which comprises the following parameters:
 Mean hot water temperature Tw, m,  (AUST  Tw, add), F (C); here, Tw, add indicates the additional
water temperature,F (C).
 Heat flux released from the floor panel qu  Qrh /Ap, Btu/h ft2 /(W/m2). Here Qrh represents the
space heating load, Btu /h (W), and Ap the area of the floor panel(s), ft2 (m2).
 Tube spacing M, in. (mm).
 Temperature differential AUST  Ta , F (C).
 Panel temperature differential Tp  Ta , F (C).
During the design of a hydronic radiant floor heating system, for a given indoor design space air
temperature Ta, first AUST  Ta must be predicted. For a heating space having small outdoor exposures
like windows, skylights, external walls, and roof, AUST  Ta. For a room equipped with a normal
size of external window and outdoor exposures, AUST  Ta can be predicted to be 2F (1.1C).
For a given room, its winter indoor design temperature Ta is 68F (20C), the heat flux released
from the floor panel qu is 25 Btu/h ft2 (W/m2), tube spacing M  12 in. (300 mm), ru is 0.8
h ft2 F/Btu (0.14 m2 C/W), and rc /rp  4. Then at the intersection of horizontal line qu  25
Btu/h ft2 (79 W/m2) and incline line ru  0.8 h ft2 F/Btu (0.14 m2 C/W), point P (Fig. 8.9d),
draw a vertical line downward that intersects the tube spacing line M  12 in. (300 mm) at point R.
The reading of Tw, add is then 45F (7.2C). The required mean hot water temperature is
Tw, m  AUST  Tw, add  68  2  45  111F (43.9C)
To determine the effective panel surface temperature Tp (floor temperature), extend the horizontal
line of qu  25 Btu/h ft2 (79 W/m2) until it intersects the vertical line AUST  Ta  2F at point
S. From Fig. 8.9d, Tp  Ta  13F (7.2C). Therefore, Tp  Ta  16  68  13  81F (27.2C).
Design Considerations
 Because of a higher space mean radiant temperature when a hydronic radiant floor heating system
is used, a lower winter indoor design temperature Ta, such as 68F (20.0C), is often adopted.
 For occupant comfort, the surface temperature of the floor heating panels Tp in occupied space
should be less than 84F (28.9C).
 In an interior heating space if only a hydronic floor heating system is installed instead of a
cooling/heating (air conditioning) system, an outdoor ventilating system is required to provide
outdoor air for the occupants.
 Setback of the space temperature during nighttime unoccupied periods produces less satisfactory
results for heavy floor heating panels. The surface temperature of the floor panels cannot drop
quickly when the temperature is set back and cannot pick up quickly either during the warm-up
period the next morning.
Control and Operations of Multizone Hydronic Radiant
Floor-Panel Heating System
Gibbs (1994) simulated and compared two different control strategies for hydronic radiant floor
heating systems:
8.30 CHAPTER EIGHT
 Pulse-width-modulated zone control. In this control, hot water is supplied from the boiler at a
constant temperature, as shown in Fig. 8.9a. A room thermostat with a proportional control mode
has a 3F (1.7C) proportional band and pulse-width-modulates the zone control valves 4 times
per hour. If the zone air temperature is at a set point Tset of 68F (20C), the on time of the zone
control valve is 7.5 min and the off time is also 7.5 min.
During a cold day in winter, hot water at a constant temperature of 120F (48.9C) is supplied to
various zones and returned at a temperature of 100F (37.8C). The surface of the floor panel is
maintained at a mean temperature of 82F (27.8C), and the space air temperature of zone 1 is
Tr1  68F (20C). As soon as the space heating load reduces, Tr1 increases to 68.3F (20.2C),
the thermostat in zone 1 pulse-width-modulates the control valve so that it is now on 6 min and off
9 min.
Pulse-width-modulated zone control is simple, effective, and lower in first cost. However, the
boiler is not protected from cold return water.
 Outdoor reset plus pulse-width-modulated zone control. In this control, a four-way mixing valve
is added to pulse-width-modulaed zone control, and a proportional-integral (PI) control is used
instead of proportional control, as shown in Fig. 8.9b. The hot water supply temperature Tws from
the boiler is no longer constant; Tws is reset according to outdoor temperature To. The zone
requires that the hottest supply temperature provide the indoor temperature feedback to the outdoor
reset control. When the zone air temperature is at the set point Trl  Tset, the control valve is
on (opened) all the time and cycles off only when the Trl 	 Tset.
To prevent excessive thermal stress in the boiler, when the return temperature is too cold to enter
the boiler and is sensed by the return temperature sensor TR as shown in Fig. 8.9b, part of the
return water is mixed with the recirculated hot water supply, until the return water temperature is
raised above a safe limit.
Outdoor reset plus pulse-width-modulated zone control is safer and more flexible and has more
continuously distributed hot water. On the other hand, it is more expensive.
System Characteristics and Applications
Compared with a warm air heating system, a hydronic radiant floor-panel heating system has the
following characteristics:
 It provides a better thermal comfort for occupants, because of less cold draft, and a higher mean
radiant temperature and an even temperature distribution.
 There is a lower natural infiltration rate and a lower winter indoor design temperature, and therefore,
it provides greater potential for saving energy.
 The peak load may be shaved (leveled) due to the thermal energy stored in the floor panels.
 Response to the load changes and heating capacity modulation are slower.
 It is higher in first cost if a separate outdoor ventilating system is installed.
Because of the use of plastic and rubber tubes for floor panels in recent years, radiant floor-panel
heating systems were used by many residences in Europe since 1980s, as well as for buildings in
locations where winter heating is dominant.
8.9 INFRARED HEATING
Basics
Infrared heating uses radiant heat transfer from a gas-fired or electrically heated high-temperature tube
or panel to provide heating to a localized area for the comfort of the occupants or the maintenance of a
HEATING SYSTEMS, FURNACES, AND BOILERS 8.31
suitable environment for a manufacturing process. Heat radiates from an infrared heater in the form of
electromagnetic waves in all directions. Most infrared heaters have reflectors to focus the radiation on
a localized target; hence, they are often known as beam radiant heaters. Infrared heating is widely
used in factories, warehouses, garages, gymnasiums, skating rinks, outdoor loading docks, and racetrack
stands.
In environments with a low indoor ambient air temperature during the winter, the warmth that an
occupant feels depends on the radiant energy absorbed by the occupant from all sources that have
temperatures higher than that of the indoor ambient air. This warmth is closely related to the effective
radiant flux Irad, or ERF Btu / h ft2/(W/m2). And ERF indicates the effective radiant flux that
must be added to an unheated space, Btu /h ft2/(W/m2). ERF and ERF can be calculated as
ERF  hr(Trad  Ta)
 h (To  Ta) (8.17a)
ERF  h(To  Tuo) (8.17b)
where hr  radiative heat-transfer coefficient, Btu/h ft2 F (W/m2 C)
h  convective and radiative combined heat-transfer coefficient, Btu/h ft2 F (W/m2 C)
Trad  mean radiant temperature, F (C)
Ta  indoor ambient air temperature, F (C)
To  operative temperature as defined in Eq. (4.6), F (C)
Tuo  operative temperature of unheated space, F (C)
In Eq. (8.17), Trad  Ta indicates the amount Trad is above the indoor ambient air Ta; it is important
because it influences the infrared radiant heat transfer to the occupant. There are two types of infrared
heaters: gas infrared heaters and electric infrared heaters.
Gas Infrared Heaters
Gas infrared heaters can be divided into indirect infrared heaters and porous matrix infrared heaters.
An indirect infrared heater consists of a burner, a radiating tube, and a reflector. Combustion takes
place within the radiating tube at a temperature up to 1200F (650C). In a porous matrix infrared
radiation heater, a gas-air mixture is supplied to an enclosure and distributed evenly through a
porous ceramic, stainless-steel, or metallic screen that is exposed at the other end. Combustion
takes place at the exposed surface and has a maximum temperature of about 1600F (870C).
Gas infrared heaters are usually vented and have a small conversion efficiency. Only 10 to 20
percent of the energy output of an open combustion gas infrared heater is in the form of infrared
radiant energy. Gas infrared heaters should not be operated under conditions in which the ambient
air contains ignitable gas or materials that may decompose to hazardous or toxic gases or
vapors.
Adequate combustion air must be provided. Venting is preferred in order to prevent a buildup of
combustion products. Usually, 4 cfm (1.9 L/s) of makeup air is required for 1000 Btu/h (293 W)
of gas input. If unvented infrared heaters are used, humidity and condensation control should be
provided to account for the accumulation of water vapor that forms during combustion.
A thermostat usually controls the supply of gas by means of a gas valve in on/off mode. For
standing-pilot ignition, a sensing element and controller are also used to cut off the gas supply
when the pilot flame extinguishes. If the combustion air is blocked, the gas supply is also cut off for
safety.
Electric Infrared Heaters
Electric infrared heaters are usually made of nickel-chromium wire or tungsten filament inside an
electrically insulated metal tube or quartz tube with or without an inert gas. The heaters also contain
8.32 CHAPTER EIGHT
a reflector, which directs the radiant beam to the area that needs heating. Nickel-chromium wires
often operate at a temperature between 1200 and 1800F (650 and 980C). The tungsten filament
can stand a temperature as high as 4000F (2200C) while the outer envelope is at a temperature of
about 1000F (540C).
Electric infrared heaters also use a thermostat to switch on or cut off the electric current to control
the input to the nickel-chromium electric heater. Some input controllers can preset the on/off
time period. For a quartz tube, the output can be controlled by varying its voltage supply. As with
gas infrared heaters, electric infrared heaters should not be used where there is a danger of igniting
flammable dust or vapors or decomposing contaminated matter into toxic gases.
Electric infrared heaters have a far higher infrared radiant energy conversion efficiency than gas
infrared heaters. They are clean and more easily managed. Although the cost of electric energy is
several times higher than that of natural gas, a comprehensive analysis should be conducted before
selecting an electric or gas infrared heater.
Design and Layout
Determination of Watt Density. According to tests, an acceptable temperature increase Trad  Ta
for normal clothed occupants in an indoor environment using infrared heating is often between 20
and 25F (11.1 and 13.9C). The effective radiant flux, or required watt density, provided by the
infrared heaters has been determined from field experiments, with the data listed in Table 8.2 for
different operating conditions.
Single or Multiple Heaters. Based on the required coverage area and the width and length of the
floor area that can be covered with the specified watt density listed in Table 8.3, it is possible to
determine whether a single heater or multiple heaters should be installed. It is also necessary to
select the coverage pattern: three-fourths overlap or full overlap. If there are many occupants within
the covering area or if the occupant will stay for a long period, a full-overlap pattern is preferable.
Mounting Height and Clearances. The mounting height of infrared heaters should not be below
10 ft (3 m); otherwise, the occupant may feel discomfort from the radiant beam overhead. Because
of the spread of the radiant beam, a higher mounting height results in a smaller watt density and a
greater coverage, and allows a greater spacing between heaters. An optimum mounting height
should be selected based on the ceiling height of the building or the outdoor structures.
Adequate clearance (recommended by the manufacturer) between the infrared heater and any
combustible material, especially between the heater and the roof, must be maintained to prevent fire.
Other Considerations. When infrared heaters are used for total indoor space heating, the following
must be taken into account:
HEATING SYSTEMS, FURNACES, AND BOILERS 8.33
TABLE 8.2 Required Watt Density,W/ ft2
Temperature Tight Drafty indoors Outdoor shielded, Outdoor unshielded,
increase uninsulated or large glass Loading area, less than 5 mph less than 10 mph
TradTa, qc building area one end open wind speed wind speed
20 30 40 50 55 60
25 37 50 62 70 75
30 45 60 75 85 90
40 60 80 100 115 120
50 75 100 125 145 150
Source: Abridged with permission from Handbook of HVAC Design 1990, Chapter 7 (Lehr Associates).
8.34 CHAPTER EIGHT
TABLE 8.3 Watt Density and Coverage
Two asymmetric heaters
Watt Wattt
Mounting density, density,
height, ft W/ft2 W  L, ft Spacing s, ft W/ft2 W  L, ft Spacing s, ft
10 33 13  12 8.5 38 11  12 6
11 27 14  13 9 33 12  13 6.5
12 24 15  14 10 29 12  14 7
13 21 16  15 10.5 25 13  15 7.5
14 18 17  16 11 22 14  16 8
15 16 18  17 12 20 15  17 8.5
16 14 20  18 12.5 18 16  18 9
18 11 22  20 14 14 18  20 10
20 9.6 24  22 15.5 11 20  22 11
22 8.0 26  24 17 9.6 22  24 12
25 6.4 30  27 19 8 24  27 13.5
Single heater
Watt Wat Wattt
Mounting density, density, density,
height, ft W/ft2 W  L, ft W/ft2 W  L, ft W/ft2 W  L, ft
10 18 12  12 25 8.5  12 19 11  12
11 15 13  13 22 9  13 16 12  13
12 13 14  14 18 10  14 14 12  14
13 11 15  15 16 10  15 13 13  15
14 10 16  16 14 11  16 11 14  16
15 9 17  17 13 12  17 9.6 15  17
16 8 18  18 11 12  18 8.8 16  18
18 6.4 20  20 8.8 14  20 7.2 18  20
20 5.2 22  22 7.6 15  22 5.8 20  22
22 4.4 24  24 6.2 17  24 4.8 24  24
25 3.5 27  27 5.0 19  27 3.9 24  27
W  width, L  length, both in ft.
Source: Abridged with permission from Handbook of HVAC Design 1990, Chapter 7 (Lehr Associates).
 The space heating load can be lower because of a lower space temperature.
 The area near the external walls needs a greater intensity of infrared heating.
 Heaters are often mounted around the perimeter at a suitable height and are usually arranged 15
to 20 ft (4.5 to 6 m) from the corners of the building.
 The key to successful design is to supply a proper amount of heat evenly to the occupying zone.
REFERENCES
Adams, C. W., Performance Results of a Pulse-Combustion Furnace Field Trial, ASHRAE Transactions, 1983,
Part I B, pp. 693699.
The American Boiler Manufacturer Association, Why Packaged Firetube Boilers? Heating/Piping/Air Conditioning,
November 1990, pp. 7983.
Andrews, J. W., Impact of Reduced Firing Rate on Furnace and Boiler Efficiency, ASHRAE Transactions, 1986,
Part I A, pp. 246262.
ASHRAE, ASHRAE Handbook 1995, HVAC Applications, ASHRAE Inc., Atlanta, GA, 1995.
ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, Atlanta, GA, 1996.
Axtman,W. H., Boiler Types and General Selection Requirements, Heating/Piping/Air Conditioning, November
1987, pp. ABMA 59.
Census Bureau, Heating and Cooling Unit in Home 1991, AC Heating & Refrigeration News, August 28, 1995
pp. 2324.
Cummings, J. B., and Tooley, J. J., Infiltration and Pressure Differences Induced by Forced Air Systems in
Florida Residences, ASHRAE Transactions, 1989, Part II, pp. 551560.
DOE/EIA, 1998 Nonresidential Buildings Energy Consumption Survey: Commercial Buildings Consumption
and Expenditures 1995, DOE/EIA-0318(95).
Fischer, R. D., Jacob, F. E., Flanigan, L. J., and Locklin, D. W., Dynamic Performance of Residential Warm-Air
Heating SystemsStatus of ASHRAE Project SP43, ASHRAE Transactions, 1984, Part II B, pp. 573590.
Gammage, R. B., Hawthorne, A. R., and White, D. A., Parameters Affecting Air Infiltration and Air Tightness in
Thirty-one East Tennessee Homes, Measured Air Leakage in Buildings, ASTM STP 904, American Society of
Testing Materials, Philadelphia, 1986.
Gibbs, D. R., Control of Multizone Hydronic Radiant Floor Heating Systems, ASHRAE Transactions, 1994,
Part I, pp. 10031010.
Grimm, N. R., and Rosaler, R. C., Handbook of HVAC Design, McGraw-Hill, New York, 1990.
Int-Hout, D., Analysis of Three Perimeter Heating Systems by Air-Diffusion Methods, ASHRAE Transactions,
1983, Part I B, pp. 101112.
Jakob, F. E., Fischer, R. D., and Flanigan, L. J., ASHRAE Transactions, 1987, Part I, pp. 14991514.
Jakob, F. E., Fischer, R. D., Flanigan, L. J., and Locklin, D. W., Validation of the ASHRAE SP43 Dynamic Simulation
Model for Residential Forced-Warm-Air Systems, ASHRAE Transactions, 1986, Part II B, pp.
623643.
Kesselring, J. P., Blatt, M. H., and Hough, R. E., New Option in Commercial Heating, Heating/Piping/Air Conditioning,
July 1990, pp. 4143.
Kilkis, B., and Coley, M., Development of a Complete Design Software for Hydronic Floor Heating of Buildings,
ASHRAE Transactions, 1995, Part I, pp. 12011213.
Lambert, L. A., and Robinson, D. H., Effects of Ducted Forced-Air Heating Systems on Residential Air Leakage
and Heating Energy Use, ASHRAE Transactions, 1989, Part II, pp. 534541.
Modera, M. P., Residential Duct System Leakage: Magnitude, Impacts, and Potential for Reduction, ASHRAE
Transactions, 1989, Part II, pp. 561569.
Palm, Jr., R. B., Pulse Combustion: A New Approach, Heating/Piping/Air Conditioning, January 1989,
pp. 148150.
Parker, D. S., Evidence of Increased Levels of Space Heat Consumption and Air Leakage Associated with
Forced Air Heating Systems in Houses in the Pacific Northwest, ASHRAE Transactions, 1989, Part II,
pp. 527533.
HEATING SYSTEMS, FURNACES, AND BOILERS 8.35
Patani, A., and Bonne, U., Operating Cost of Gas-Fired Furnace Heating Systems with Add-on Heat Pumps,
ASHRAE Transactions, 1983, Part I B, pp. 319329.
Paul, D. D., Whitacre, G. R., Fischer, R. D., Rutz, A. L., Dewerth, D. W, Borgeson, R. A., and Leslie, N. P.,
Development of Vent Capacity Tables for Category I Gas Appliances with Fan Assisted Combustion Systems,
ASHRAE Transactions, 1993, Part I, pp. 11631179.
Robison, D. H., and Lambert, L. A., Field Investigation of Residential Infiltration and Heating Duct Leakage,
ASHRAE Transactions, 1989, Part II, pp. 542550.
Slattery, L. T., A Look at Packaged Boilers, Heating/Piping/Air Conditioning, no. 12, 1995, pp. 6572.
Spolek, G. A., Herriott, D. W., and Low, D. M., Air Flow in Rooms with Baseboard Heat: Flow Visualization
Studies, ASHRAE Transactions, 1986, Part II A, pp. 528536.
Tao,W., Modern Boiler Plant Design, Heating/Piping/Air Conditioning, November 1984, pp. 6982.
The Trane Company, Boilers for Steam and Hot Water, La Crosse, WI, 1971.
The Trane Company, Basics of Heating with Electricity, La Crosse, WI, 1973.
Trehan, A. K., Fortmann, R. C., Koontz, M. D., and Nagda, N. L., Effect of Furnace Size on Morning Picking
Time, ASHRAE Transactions, 1989, Part I, pp. 11251129.
Trewin, R. R., Langdon, F. M., Nelson, R. M., and Pate, M. B., An Experimental Study of A Multipurpose Commercial
Building with Three Different Heating Systems, ASHRAE Transactions, 1987, Part I, pp. 467481.
8.36 CHAPTER EIGHT
CHAPTER 9
REFRIGERANTS,
REFRIGERATION CYCLES,
AND REFRIGERATION SYSTEMS
9.1
9.1 REFRIGERATION AND REFRIGERATION
SYSTEMS 9.2
9.2 REFRIGERANTS 9.3
Refrigerants, Cooling Media, and Liquid
Absorbents 9.3
Azeotropic, Near Azeotropic, and
Zeotropic 9.3
Numbering of Refrigerants 9.4
9.3 PROPERTIES AND CHARACTERISTICS
OF REFRIGERANTS 9.5
Safety Requirements 9.5
Effectiveness of Refrigeration Cycle 9.5
Evaporating and Condensing
Pressures 9.6
Oil Miscibility 9.6
Inertness 9.6
Thermal Conductivity 9.6
Refrigeration Capacity 9.6
Physical Properties 9.6
Operating Characteristics 9.6
9.4 PHASEOUT OF OZONE DEPLETION
REFRIGERANTS 9.7
Refrigerant Use 9.7
Ozone Depletion and Global Warming
Potentials 9.7
Phaseout of CFCs, Halons, and
HCFCs 9.10
Montreal Protocol and Clean Air Act 9.10
Action and Measures 9.11
Status of CFC Replacements 9.13
9.5 CLASSIFICATION OF
REFRIGERANTS 9.13
Hydrofluorocarbons 9.13
Azeotropic HFC 9.14
Near-Azeotropic HFC 9.14
Zeotropic HFC 9.15
HCFCs and Their Zeotropes 9.15
Inorganic Compounds 9.16
CFCs, Halons, and Their Zeotropes 9.16
9.6 REFRIGERATION PROCESSES AND
REFRIGERATION CYCLES 9.16
Refrigeration Processes 9.16
Refrigeration Cycles 9.17
Unit of Refrigeration 9.17
9.7 GRAPHICAL AND ANALYTICAL
EVALUATION OF REFRIGERATION 9.17
Pressure-Enthalpy Diagram 9.17
Temperature-Entropy Diagram 9.18
Analytical Evaluation of Cycle
Performance 9.19
9.8 CARNOT REFRIGERATION CYCLE 9.19
Performance of Carnot Refrigeration
Cycle 9.19
9.9 COEFFICIENT OF PERFORMANCE OF
REFRIGERATION CYCLE 9.21
9.10 SINGLE-STAGE IDEAL VAPOR
COMPRESSION CYCLE 9.22
Flow Processes 9.22
Cycle Performance 9.22
Determination of Enthalpy by
Polynomials 9.24
Refrigeration Effect, Refrigerating Load,
and Refrigerating Capacity 9.25
9.11 SUBCOOLING AND
SUPERHEATING 9.26
Subcooling 9.26
Superheating 9.26
9.12 MULTISTAGE VAPOR COMPRESSION
SYSTEMS 9.29
Compound Systems 9.29
Interstage Pressure 9.30
Flash Cooler and Intercooler 9.31
9.13 TWO-STAGE COMPOUND SYSTEM
WITH A FLASH COOLER 9.31
Flow Processes 9.31
Fraction of Evaporated Refrigerant in
Flash Cooler 9.31
Enthalpy of Vapor Mixture Entering
Second-Stage Impeller 9.32
Coefficient of Performance 9.33
Characteristics of Two-Stage Compound
System with Flash Cooler 9.33
9.14 THREE-STAGE COMPOUND SYSTEM
WITH A TWO-STAGE FLASH
COOLER 9.35
Flow Processes 9.35
Fraction of Refrigerant Vaporized in
Flash Cooler 9.35
Coefficient of Performance of
Three-Stage System 9.38
9.15 TWO-STAGE COMPOUND
SYSTEM WITH A VERTICAL
INTERCOOLER 9.38
Comparison between Flash Cooler and
Vertical Coil Intercooler 9.40
9.1 REFRIGERATION AND REFRIGERATION SYSTEMS
Refrigeration is defined as the process of extracting heat from a lower-temperature heat source,
substance, or cooling medium and transferring it to a higher-temperature heat sink. Refrigeration
maintains the temperature of the heat source below that of its surroundings while transferring
the extracted heat, and any required energy input, to a heat sink, atmospheric air, or surface
water.
A refrigeration system is a combination of components and equipment connected in a sequential
order to produce the refrigeration effect. The refrigeration systems commonly used for air
conditioning can be classified by the type of input energy and the refrigeration process as
follows:
1. Vapor compression systems. In vapor compression systems, compressors activate the refrigerant
by compressing it to a higher pressure and higher temperature level after it has produced
its refrigeration effect. The compressed refrigerant transfers its heat to the sink and is
condensed to liquid form. This liquid refrigerant is then throttled to a low-pressure, lowtemperature
vapor to produce refrigerating effect during evaporation. Vapor compression
systems are the most widely adopted refrigeration systems in both comfort and process air
conditioning.
2. Absorption systems. In an absorption system, the refrigeration effect is produced by thermal
energy input. After absorbing heat from the cooling medium during evaporation, the vapor
refrigerant is absorbed by an absorbent medium. This solution is then heated by direct-fired
furnace, waste heat, hot water, or steam. The refrigerant is again vaporized and then condensed
to liquid to begin the refrigeration cycle again.
3. Air or gas expansion systems. In an air or gas expansion system, air or gas is compressed
to a high pressure by mechanical energy. It is then cooled and expanded to a low pressure.
Because the temperature of air or gas drops during expansion, a refrigeration effect is
produced.
9.2 CHAPTER NINE
9.16 CASCADE SYSTEMS 9.40
Advantages and Disadvantages 9.40
Performance of Cascade System 9.42
9.17 AIR EXPANSION REFRIGERATION
CYCLES 9.45
Thermodynamic Principle 9.45
Flow Processes of a Basic Air Expansion
System for Aircraft 9.47
Air Expansion Cycle 9.48
9.18 REFRIGERATION SYSTEMS
CLASSIFICATIONS AND
DEVELOPMENTS 9.49
Classifications 9.49
Recent Developments 9.51
9.19 REFRIGERATION
COMPRESSORS 9.51
Positive Displacement and NonPositive
Displacement Compressors 9.51
Hermetic, Semihermetic, and Open
Compressors 9.53
Direct Drive, Belt Drive, and
Gear Drive 9.53
9.20 PERFORMANCE OF
COMPRESSORS 9.53
Volumetric Efficiency 9.53
Motor, Mechanical, and Compression
Efficiency 9.54
Isentropic and Polytropic Analysis 9.54
Energy Use Index 9.55
9.21 SAFETY REQUIREMENTS AND
MACHINERY ROOM 9.56
Refrigerant Safety 9.56
Application Rules for High-Probability
Systems 9.56
Application Rules for Low-Probability
Systems 9.57
Refrigerating Systems of 100 hp
(74.6 kW) or less 9.58
Refrigerating Machinery Room 9.58
Storage of Refrigerants 9.59
REFERENCES 9.59
9.2 REFRIGERANTS
Refrigerants, Cooling Media, and Liquid Absorbents
Refrigerants. A refrigerant is the primary working fluid used for absorbing and transmitting heat
in a refrigeration system. Refrigerants absorb heat at a low temperature and low pressure and
release heat at a higher temperature and pressure. Most refrigerants undergo phase changes during
heat absorptionevaporationand heat releasingcondensation.
Cooling Media. A cooling medium is the working fluid cooled by the refrigerant to transport the
cooling effect between a central plant and remote cooling units and terminals. In a large, centralized
system, it is often more economical to use a coolant medium that can be pumped to remote locations
where cooling is required. Chilled water, brine, and glycol are used as cooling media in many
refrigeration systems. The cooling medium is often called a secondary refrigerant, because it obviates
extensive circulation of the primary refrigerant.
Liquid Absorbents. A solution known as liquid absorbent is often used to absorb the vaporized
refrigerant (water vapor) after its evaporation in an absorption refrigeration system. This solution,
containing the absorbed vapor, is then heated at high pressure. The refrigerant vaporizes, and the
solution is restored to its original concentration for reuse.
Lithium bromide and ammonia, both in a water solution, are the liquid absorbents used most
often in absorption refrigerating systems.
Azeotropic, Near Azeotropic, and Zeotropic
A refrigerant can either be a single chemical compound or a mixture (blend) of multiple compounds.
Azeotropic. These are blends of multiple components of volatilities (refrigerants) that evaporate
and condense as a single substance and do not change their volumetric composition or saturation
temperature when they evaporate or condense at a constant pressure. Components in a mixture of
azeotropes cannot be separated from their constituents by distillation. Properties of azeotropic
refrigerants are entirely different from those of their components and may be conveniently treated
as a single chemical compound.
Near Azeotropic. Near-azeotropic refrigerants are blends whose characteristics are near to
azeotropic. Although properties of near-azeotropic refrigerants are nearer to azeotropic than to nonazeotropic
(zeotropic), near-azeotropic refrigerants are defined as zeotropic or nonazeotropic.
Zeotropic. These are blends of multiple components of volatilities (refrigerants) that evaporate
and condense as a single substance and do change volumetric composition or saturation temperature
when they evaporate or condense at a constant pressure.
Blends. Mixtures of refrigerants of two or more chemical compounds are blends. The advantage
of a blend of multiple chemical compounds compared to a single compound is that the required
properties of the blend can possibly be achieved by varying the fractional composition of the
components.
Glide. Zeotropic mixtures, including near-azeotropic blends, show changes in composition because
of the leaks, the difference between liquid and vapor phases, or the difference between the
charge and circulation, or their combined effect. The shift in composition causes the change in
evaporating and condensing temperature and pressure. The difference in dew point and bubble point
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.3
in the temperature-concentration diagram of a zeotropic refrigerant during evaporation and condensation
is called glide, expressed in F (C). A near-azeotropic refrigerant has a smaller glide than
a zeotropic one. The midpoint between the dew point and bubble point is usually taken as the
evaporating or condensing temperature for a nonazeotropic and near-azeotropic refrigerant. Hwang
et al. (1997) showed that temperature drops during condensation and temperature increases during
evaporation. Ideal or perfect azeotropic refrigerants are uncommon, whereas near-azeotropic ones
are fairly common.
Numbering of Refrigerants
Before the invention of chlorofluorocarbons (CFCs), refrigerants were called by their chemical names.
Because of the complexity of these names, especially the CFCs, the fully halogenated CFCs, and
hydrochlorofluorocarbons (HCFCs), the not fully halogenated HCFCs (see Fig. 9.1), a numbering system
was developed for hydrocarbons and halocarbons, and is used widely in the refrigeration industry.
According to ANSI/ASHRAE Standard 34-1997, the first digit from the right is the number of fluorine
atoms in the compound. The second digit from the right is one more than the number of hydrogen
atoms in the compound. The third digit from the right is one less than the number of the carbon atoms
in the compound. If the digit is zero, it is omitted from the number. The fourth digit from the right is
the number of unsaturated carbon-carbon bonds in the compound. If the digit is zero it is also omitted
from the number. For example, the chemical formula of HCFC-123 is CHCl2CF3:
There are 3 fluorine atoms, first digit from the right is 3
There is 1 hydrogen atom, second digit from the right is 1  1  2
There are 2 carbon atoms, third digit from the right is 2  1  1
No unsaturated C9C bonds, the fourth digit from the right is 0
9.4 CHAPTER NINE
FIGURE 9.1 Fully halogenated CFCs and not fully halogenated HCFCs.
9.3 PROPERTIES AND CHARACTERISTICS OF REFRIGERANTS
Today, the preservation of the ozone layer is the first priority of refrigerant selection. In addition,
the global warming effect and the following factors should be considered.
Safety Requirements
Refrigerant may leak from pipe joints, seals, or component parts during installation, operation, or
accident. Therefore, refrigerants must be acceptably safe for humans and manufacturing processes,
with little or no toxicity or flammability.
In ANSI/ASHRAE Standard 34-1997, the toxicity of refrigerants is classified as class A or B.
Class A refrigerants are of lower toxicity. A class A refrigerant is one whose toxicity has not been
identified when its concentration is less than or equal to 400 ppm, based on threshold limit
valuetime-weighted average (TLV-TWA) or equivalent indices. The TLV-TWA concentration is a
concentration to which workers can be exposed over an 8-h workday and a 40-h workweek without
suffering adverse effect. Concentration ppm means parts per million by mass.
Class B refrigerants are of high toxicity. A class B refrigerant produces evidence of toxicity
when workers are exposed to a concentration below 400 ppm based on a TLV-TWA concentration.
Flammable refrigerants explode when ignited. If a flammable refrigerant is leaked in the area of a
fire, the result is an immediate explosion. Soldering and welding for installation or repair cannot be
performed near such gases.
ANSI/ASHRAE Standard 34-1997 classifies the flammability of refrigerants into classes 1, 2,
and 3. Class 1 refrigerants show no flame propagation when tested in air at a pressure of 14.7 psia
(101 kPa) at 65F (18.3C). Class 2 refrigerants have a lower flammability limit (LFL) of more than
0.00625 lb/ ft3 (0.1 kg/m3) at 70F (21.1C) and 14.7 psia (101 kPa abs.), and a heat of combustion
less than 8174 Btu/lb (19,000 kJ/kg). Class 3 refrigerants are highly flammable, with an LFL less
than or equal to 0.00625 lb/ ft3 (0.1 kg/m3) at 70F (21.1C) and 14.7 psia (101 kPa abs.) or a heat
of combustion greater than or equal to 8174 Btu /lb (19,000 kJ/ kg).
A refrigerants safety classification is its combination of toxicity and flammability. According to
ANSI/ASHRAE Standard 34-1997, safety groups are classified as follows:
 A1 lower toxicity and no flame propagation
 A2 lower toxicity and lower flammability
 A3 lower toxicity and higher flammability
 B1 higher toxicity and no flame propagation
 B2 higher toxicity and lower flammability
 B3 higher toxicity and higher flammability
For zeotropic blends whose flammability and toxicity may change as their composition changes, a
dual safety classification should be determined. The first classification denotes the classification of
the formulated composition of the blend. The second classification lists the classification of the
blend composition at the worst case of fractionation.
Effectiveness of Refrigeration Cycle
The effectiveness of refrigeration cycles, or coefficient of performance (COP), is one parameter that
affects the efficiency and energy consumption of the refrigeration system. It will be clearly defined
in a later section. The COP of a refrigeration cycle using a specific refrigerant depends mainly upon
the isentropic work input to the compressor at a given condensing and evaporating pressure differential,
as well as the refrigeration effect produced.
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.5
Evaporating and Condensing Pressures
It is best to use a refrigerant whose evaporating pressure is higher than that of the atmosphere so
that air and other noncondensable gases will not leak into the system and increase the condensing
pressure. The condensing pressure should be low because high condensing pressure necessitates
heavier construction of the compressor, piping, condenser, and other components. In addition, a
high-speed centrifugal compressor may be required to produce a high condensing pressure.
Oil Miscibility
When a small amount of oil is mixed with refrigerant, the mixture helps to lubricate the moving
parts of a compressor. Oil should be returned to the compressor from the condenser, evaporator, accessories,
and piping, in order to provide continuous lubrication. On the other hand, refrigerant can
dilute oil, weakening its lubricating effect; and when the oil adheres to the tubes in the evaporator
or condenser, it forms film that reduces the rate of heat transfer.
Inertness
An inert refrigerant does not react chemically with other materials, thus avoiding corrosion, erosion,
or damage to the components in the refrigerant circuit.
Thermal Conductivity
The thermal conductivity of a refrigerant is closely related to the efficiency of heat transfer in the
evaporator and condenser of a refrigeration system. Refrigerant always has a lower thermal conductivity
in its vapor state than in its liquid state. High thermal conductivity results in higher heat transfer
in heat exchangers.
Refrigeration Capacity
The cubic feet per minute (cfm) suction vapor of refrigerant required to produce 1 ton of refrigeration
(liters per second to produce 1 kW of refrigeration) depends mainly on the latent heat of vaporization
of the refrigerant and the specific volume at the suction pressure. It directly affects the size
and compactness of the compressor and is one of the criteria for refrigerant selection.
Physical Properties
Discharge Temperature. A discharge temperature lower than 212F (100C) is preferable because
temperatures higher than 300F (150C) may carbonize lubricating oil or damage some of the
components.
Dielectric Properties. Dielectric properties are important for those refrigerants that will be in direct
contact with the windings of the motor (such as refrigerants used to cool the motor windings in
a hermetically sealed compressor and motor assembly).
Operating Characteristics
Leakage Detection. Refrigerant leakage should be easily detected. If it is not, gradual capacity reduction
and eventual failure to provide the required cooling will result. Most of the currently used
refrigerants are colorless and odorless. Leakage of refrigerant from the refrigeration system is often
detected by the following methods:
9.6 CHAPTER NINE
 Halide torch. This method is simple and fast. When air flows over a copper element heated by a
methyl alcohol flame, the vapor of halogenated refrigerant decomposes and changes the color of
the flame (green for a small leak, bluish with a reddish top for a large leak).
 Electronic detector. This type of detector reveals a variation of electric current due to ionization
of decomposed refrigerant between two oppositely charged electrodes. It is sensitive, but cannot
be used where the ambient air contains explosive or flammable vapors.
 Bubble method. A solution of soup or detergent is brushed over the seals and joints where leakage
is suspected, producing bubbles that can be easily detected.
9.4 PHASEOUT OF OZONE DEPLETION REFRIGERANTS
Refrigerant Use
The use of CFCs and HCFCs is a global concern. Approximately two-thirds of all fully halogenated
CFCs were used outside the United States in the mid-1980s. In 1985, the total use of halocarbons
in the United States was 611 million lb (0.28 million ton). These halocarbons were used in foam insulation,
automotive air conditioners, new systems of Air Conditioning and Refrigeration Institute
(ARI) members, and other products. Foam insulation blown by CFCs was the largest user. Automotive
air conditioners made up 19 percent of the total and CFCs purchased by ARI members for new
systems made up 5 percent of the total use. Of the CFCs and HCFCs purchased by ARI members,
HCFC-22 made up 77 percent, while CFC-11 and CFC-12 each made up about 10 percent.
Ozone Depletion and Global Warming Potentials
To compare the relative ozone depletion caused by various refrigerants, an index called the ozone
depletion potential (ODP) has been proposed. ODP is the ratio of the rate of ozone depletion of 1 lb
of any halocarbon to that of 1 pound of CFC-11. The ODP of CFC-11 is assigned a value of 1. The
following are the ODP values for various refrigerants:
Similar to the ODP, the halocarbon global warming potential (HGWP) is the ratio of calculated
warming for each unit mass of gas emitted to the calculated warming for a unit mass of reference
gas CFC-11. The HGWPs of various refrigerants are listed in Table 9.1. In addition to the HGWP,
another global warming index uses CO2 as a reference gas. For example, 1 lb of HCFC-22 has the
Refrigerant Chemical formula ODP value
CFC-11 CCl3F 1.0
CFC-12 CCl2F2 1.0
CFC-13B1 CBrF3 10
CFC-113 CCl2FCClF2 0.8
CFC-114 CClF2CClF2 1.0
CFC-115 CClF2CF3 0.6
CFC/HFC-500 CFC-12 (73.8%)/HFC-152a (26.2%) 0.74
CFC/HCFC-502 HCFC-22 (48.8%)/CFC-115 (51.2%) 0.33
HCFC-22 CHClF2 0.05
HCFC-123 CHCl2CF3 0.02
HCFC-124 CHClFCF3 0.02
HCFC-142b CH3CClF2 0.06
HFC-125 CHF2CF3 0
HFC-134a CF3CH2F 0
HFC-152a CH3CHF2 0
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.7
9.8
TABLE 9.1 Properties of Commonly Used Refrigerants at 40F Evaporating and 100F Condensing
Halocarbon
Ozone global
depletion warming Evaporating Condensing
Chemical Molecular potential potential pressure, pressure, Compression Refrigeration
formula mass (ODP) (HGWP) psia psia ratio effect, Btu/ lb
Hydrofluorocarbons (HFCs)
HFC-32 Difluoromethane CH2F2 52.02 0 0.11 135.6 340.2 2.51
HFC-125 Pentafluoroethane CHF2CF3 120.02 0 0.84 112.4 276.8 2.46 36.4
HFC-134a Tetrafluoroethane CF3CH2F 102.3 0 0.28 49.8 138.9 2.79
HFC-143a Trifluoroethane CH3CF3 84.0 0 1.1
HFC-152a Difluoroethane CHF2CH3 66.05 0 0.03 44.8 124.3 2.77
HFC-245ca Pentafluoropropane CF3CF2CH3 134.1 0 0.09
Azeotropic HFC
HFC-507 HFC-125/HFC-143a(45/55) 0 0.98
HFC-507A HFC-125/HFC-143a(50/50) 0 104.6 257.6 2.46
Near-azeotropic HFC
HFC-404A HFC-125/HFC-143a/
HFC-134a (44/52/4) 0 0.95 109.16 251.18 2.30
HFC-410A HFC-32/HFC-125(50/50) 0 0.48 132.90 332.33 2.50
HFC Zeotropic
HFC-407A HFC-32/HFC-125/
HFC-134a (20/40/40) 0 0.47
HFC 407C HFC-32/HFC-125/
HFC-134a (23/25/52) 0 0.38 86.13 225.22 2.61
Hydrochlorofluorocarbons (HCFCs) and their zeotropes
HCFC-22 Chlorodifluoromethane CHClF2 86.48 0.05 0.34 82.09 201.5 2.46 69.0
HCFC-123 Dichlorotrifluoroethane CHCl2CF3 153.0 0.016 0.02 5.8 20.8 3.59 62.9
HCFC-124 Chlorotetrafluoroethane CHFClCF3 136.47 0.02 0.10 27.9 80.92 2.90 52.1
Near-azeotropic HCFC
HCFC-402A HCFC-22/HFC-125/
PRO-290 (38/60/2) 0.02
Zeotropic HCFC
HCFC-401A HCFC-22/HCFC-124/
HFC-152a (53/34/13) 0.037 0.22
HCFC-401B HCFC-22/HCFC-124/
HFC-152a (61/28/11) 0.04 0.24
Inorganic compounds
R-717 Ammonia NH3 17.03 0 0 71.95 206.81 2.87 467.4
R-718 Water H2O 18.02 0
R-729 Air 28.97 0
Chlorofluorocarbons (CFCs), halons, and their azeotropes
CFC-11 Trichlorofluoromethane CCl3F 137.38 1.00 1.00 6.92 23.06 3.33 68.5
CFC-12 Dichlorodifluoromethane CCl2F2 120.93 1.00 3.1 50.98 129.19 2.53 50.5
BFC-13B1 Bromotrifluoromethane CBrF3 148.93 10
CFC-113 Trichlorotrifluoroethane CCl2FCClF2 187.39 0.80 1.4 2.64 10.21 3.87 54.1
CFC-114 Dichlorotetrafluoroethane CCl2FCF3 170.94 1.00 3.9 14.88 45.11 3.03 42.5
CFC-500 CFC-12/HFC-152a (73.8/26.2) 99.31 0.74 59.87 152.77 2.55 60.5
CFC-502 HCFC-22/CFC-115 (48.8/51.2) 111.63 0.22 3.7
Source: Adapted with permission from ASHRAE Handbook 1997, Fundamentals, and ANSI/ASHRAE Standard 34-1992 and Addenda 1997.
9.9
Specific
volume of Power Critical Discharge
Trade suction Compressor, consumption, temperature, temperature,
Alternatives name vapor, ft3/ lb cfm/ton hp/ton F F Flammability Safety
Hydrofluorocarbons (HFCs)
HFC-32 0.63 173.1 Lower flammability A2
HFC-125 0.33 150.9 103 Nonflammable A1
HFC-134a CFC-12, HCFC-22 0.95 213.9 Nonflammable A1
HFC-143a Lower flammability A2
HFC-152a 1.64 236.3 Lower flammability A2
HFC-245ca CFC-11, HCFC-123 353.1
Azeotropic HFC
HFC-507 CFC-502, CFC-12 Genetron AZ-50 Nonflammable A1
HFC-507A 0.43 Nonflammable A1
Near-azeotropic HFC
HFC-404A CFC-502, CFC-12 SUVA HP-62 0.44 A1/A1
HFC-410A HCFC-22, CFC-502 AZ-20 0.45 A1/A1
Zeotropic HFC A1/A1
HFC-407A CFC-502, CFC-12 KLEA60, AC-9000 A1/A1
HFC-407C HCFC-22, CFC-502 KLEA66 0.63
Hydrochlorofluorocarbons (HFCs) and their zeotropes
HCFC-22 0.66 1.91 0.696 204.8 127 Nonflammable A1
HCFC-123 CFC-11 5.88 18.87 0.663 362.6 Nonflammable B1
HCFC-124 1.30 5.06 0.698 252.5 Nonflammable A1
Near-azeotropic HCFC
HCFC-402A CFC-502 SUVAHP-80 A1/A1
Zeotropic HCFC
HCFC-401A CFC-12 MP-39 A1/A1
HCFC-401B CFC-12 MP-66 A1/A1
Inorganic compounds
R-717 3.98 1.70 0.653 271.4 207 Lower flammability B2
R-718 Nonflammable A1
R-729 Nonflammable
Chlorofluorocarbons (CFCs), halons, and their azeotropes
CFC-11 5.43 15.86 0.636 388.4 104 Nonflammable A1
CFC-12 0.79 3.08 0.689 233.6 100 Nonflammable A1
BFC-13B1 0.21 152.6 103 Nonflammable A1
CFC-113 10.71 39.55 0.71 417.4 86 Nonflammable A1
CFC-114 2.03 9.57 0.738 294.3 86 Nonflammable A1
CFC-500 CFC-12/HFC152a (73.8/26.2) 0.79 3.62 0.692 221.9 105 Nonflammable A1
CFC-502 HCFC-22/CFC115 (48.8/51.2) 98 Nonflammable A1
same effect on global warming as 4100 lb (1860 kg) of CO2 in the first 20 years after it is released
into the atmosphere. Its impact drops to 1500 lb (680 kg) at 100 years.
Phaseout of CFCs, Halons, and HCFCs
The theory of depletion of the ozone layer was proposed in 1974 by Rowland and Molina. (The
1995 Nobel Prize was awarded to F. Sherwood Rowland, Mario Molina, and Paul Crutzen for their
work in atmospheric chemistry and theory of ozone depletion.) Network station in Halley Bay,
Antarctica, established a baseline trend of ozone levels that helped scientists to discover the ozone
hole in 1985. National Aeronautics and Space Administration (NASA) flights into the stratosphere
over the arctic and antarctic circles found CFC residue where the ozone layer was damaged. Approximately
the same ozone depletion over the antarctic circle was found in 1987, 1989, 1990, and
1991. By 1988, antarctic ozone levels were 30 percent below those of the mid-1970s. The most severe
ozone loss over the antarctic was observed in 1992. Ground monitoring at various locations
worldwide in the 1980s has showed a 5 to 10 percent increase in ultraviolet radiation. Although
there is controversy about the theory of ozone layer depletion among scientists, as discussed in
Rowland (1992), action must be taken immediately before it is too late.
Montreal Protocol and Clean Air Act
In 1978, the Environmental Protection Agency (EPA) and the Food and Drug Administration (FDA)
of the United States issued regulations to phase out the use of fully halogenated CFCs in nonessential
aerosol propellants, one of the major uses at that time. On September 16, 1987, the European
Economic Community and 24 nations, including the United States, signed the Montreal Protocol.
This document is an agreement to phase out the production of CFCs and halons by the year 2000.
The Montreal Protocol had been ratified by 157 parties.
The Clean Air Act Amendments, signed into law in the United States on November 15, 1990,
governed two important issues: the phaseout of CFCs and a ban (effective July 1, 1992) on the deliberate
venting of CFCs and HCFCs. Deliberate venting of CFCs and HCFCs must follow the regulations
and guidelines of the EPA. In February 1992, then-President Bush called for an accelerated
phaseout of the CFCs in the United States. Production of CFCs must cease from January 1, 1996,
with limited exceptions for service to certain existing equipment.
In late November 1992, representatives of 93 nations meeting in Copenhagen also agreed to the
complete cessation of CFC production beginning January 1, 1996, and of halons by January 1,
1994, except continued use from existing (reclaimed or recycled) stock in developed nations. In
addition, the 1992 Copenhagen amendments and later a 1995 Vienna meeting revision agreed to
restrict the production of HCFCs relative to a 1989 level beginning from 2004 in developed nations
according to the following schedule:
Consumption indicates the production plus imports minus exports and feedstocks. The value of
2.8 percent cap is the revised value of the Vienna meeting in 1995 to replace the original value of
Date Production limit
January 1, 1996 100 percent cap
Cap  2.8 percent of ODP of 1989 CFC consumption
plus total ODP of 1989 HCFC consumption
January 1, 2004 65 percent cap
January 1, 2010 35 percent cap
January 1, 2015 10 percent cap
January 1, 2020 0.5 percent cap
January 1, 2030 Complete cessation of production
9.10 CHAPTER NINE
3.1 percent in the Copenhagen amendments. The Copenhagen amendments had been ratified by 58
parties.
Action and Measures
The impact of CFCs on the ozone layer poses a serious threat to human survival. The following
measures are essential:
Conversions and Replacements. Use alternative refrigerants (substitutes) to replace the CFCs in
existing chillers and direct-expansion (DX) systems. During the conversion of the CFC to
nonozone depletion alternative refrigerants, careful analysis should be conducted of capacity, ef-
ficiency, oil miscibility, and compatibility with existing materials after conversion. For many refrigeration
systems that already have a service life of more than 15 years, it may be cost-effective to
buy a new one using non-CFC refrigerant to replace the existing refrigeration package.
 HFC-134a and HCFC-22 are alternative refrigerants to replace CFC-12.
 HCFC-123, and HFC-245ca are alternative refrigerants to replace CFC-11 in large chillers.
It is important to realize that HCFC-123 and HCFC-22 themselves are interim refrigerants and will
be restricted in consumption beginning in 2004. HCFC-123 has a very low global warming potential and
is widely used in centrifugal chillers. HCFC-22 is widely used in small and medium-size DX systems.
 HFC-134a, HFC-407C, and HFC-410A are alternative refrigerants to replace HCFC-22. HFC-
407C is a near-azeotropic refrigerant of HFC-32/HFC-125/HFC-134a (23/25/52) [means
(23%/25%/525)], and HFC-410A also a near-azeotropic refrigerant of HFC-32/HFC-125 (50/50).
 HFC-245ca or another new HFC possibly developed before 2004 will be the hopeful alternative
to replace HCFC-123.
In supermarkets, CFC-502 is a blend of HCFC-22/CFC-115 (48.8/ 51.2).
 HFC-404A, HFC-507, and HFC-410A are alternative refrigerants to replace CFC-502. HFC-
404A is a near-azeotropic refrigerant of HFC-125/HFC-143a/HFC-134a(44/52/4); and HFC-
507 is an azeotropic refrigerant of HFC-125/HFC-143a (45/ 55).
Reducing Leakage and Preventing Deliberate Venting. To reduce the leakage of refrigerant from
joints and rupture of the refrigeration system, one must detect the possible leakage, tighten the
chillers, improve the quality of sealing material, and implement preventive maintenance.
Prevent the deliberate venting of CFCs and HCFCs and other refrigerants during manufacturing,
installation, operation, service, and disposal of the products using refrigerants.
Avoid CFC and HCFC emissions through recovery, recycle, and reclaiming. According to
ASHRAE Guideline 3-1990, recovery is the removal of refrigerant from a system and storage in an
external container. Recycle involves cleaning the refrigerant for reuse by means of an oil separator
and filter dryer. In reclamation, refrigerant is reprocessed for new product specifications.
To avoid the venting of CFCs and HCFCs and other refrigerants, an important step is to use an
ARI-certified, portable refrigerant recovering/ recycling unit to recover all the liquid and remaining
vapor from a chiller or other refrigeration system. An outside recovery/ reclaiming service firm may
also be employed. A typical refrigerant recovery unit is shown in Fig. 9.2. It includes a recovery
cylinder, a vacuum pump or compressor, a water-cooled condenser, a sight glass, a shutoff float
switch, necessary accessories, pipes, and hoses.
To recover refrigerant from a chiller that has been shut down involves two phases: liquid recovery
and vapor recovery. Liquid recovery is shown in Fig. 9.2a. The vacuum pump or compressor
in the recovery unit creates a low pressure in the recovery cylinder. Liquid refrigerant is then
extracted from the bottom of the chiller into the recovery cylinder. If the recovery cylinder is not
large enough, the shutoff float switch ceases to operate the vacuum pump or compressor when the
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.11
recovery cylinder is 80 percent full. Another empty recovery cylinder is used to replace the filled
cylinder. If the vapor enters the sight glass, which means that the liquid refrigerant is all extracted,
then the vacuum pump or compressor is stopped and the vapor recovery phase begins.
Vapor recovery is shown in Fig. 9.2b. The vacuum pump or compressor extracts the refrigerant
vapor from the top of the chiller. Extracted refrigerant vapor is then condensed to liquid form that
flows through the water-cooled condenser and is stored in the recovery cylinder. Noncondensable
gases are purged into the atmosphere from the recovery unit. Water at a temperature between 40
and 85F (4.4 and 29.4C) is often used as the condensing cooling medium. The recovered refrigerant
can be recycled or reclaimed as required.
In addition to the recovery of refrigerants from the chiller or other refrigeration system, refrigerant
vapor detectors should be installed at locations where refrigerant from a leak is likely to
concentrate. These detectors can set off an alarm to notify the operator to seal the leak.
9.12 CHAPTER NINE
FIGURE 9.2 A typical refrigerant recovery unit: (a) liquid recovery; (b) vapor recovery.
Because of the worldwide effort to phase out CFCs, the latest result of a study conducted by the
National Oceanic and Atmospheric Administration (NOAA) and published in the journal Science
was reported by UPI science writer Susan Milius . . . scientists are cheering a 1 percent reduction
during 1995 in the chemicals that slowly carry chlorine and bromine aloft to the stratosphere.
(Air Conditioning, Heating and Refrigeration News, June 10, 1996, p. 2).
Status of CFC Replacements
Dooley (1997) showed that according to the ARI survey, in the United States there were about
80,000 CFC large-tonnage chillers in 1992, and most used CFC-11 as refrigerant. At the beginning
of 1997, some 18,981 chillers, or 24 percent of the total 80,000, had been phased-out CFC refrigerants.
Of these, 4813 chillers were converted to non-CFC refrigerant, and 14,168 chillers were
replaced by new chillers which used non-CFC refrigerant. The ratio of new replacements to conversions
was about 3 to 1. It is often cost-effective to replace an old chiller instead of to convert the
CFC refrigerant to alternatives in an existing chiller.
The ARI estimates that 53 percent of the 80,000 chillers in 1992 will remain in service on
January 1, 2000. This shows that the actual phaseout process was slower than expected.
For CFC-12, automotive cooling and supermarkets acount for more than 90 percent of their servicing
requirements. Due to the vibrations and unsteady operating conditions, automotive cooling
required greater servicing losses and a faster phaseout schedule than large-tonnage chillers. The annual
amount of CFCs required to compensate for servicing losses came from stockpiles of virgin
CFCs and reclaimed CFCs. Because of the slower phaseout of CFC chillers, servicing demands
were greater than could be supplied from the reclaimed CFCs. The using up of the stockpiles of
virgin CFCs and thus a shortage of CFC supply may occur at the begining of the twenty-first century
in the United States. The slower phaseout of CFCs indicated that there is a possibility of a
considerably longer period of servicing of HCFC equipment than called for in the production
phaseout schedule in the twenty-first century.
9.5 CLASSIFICATION OF REFRIGERANTS
Before the introduction of chlorofluorocarbons in the 1930s, the most commmonly used refrigerants
were air, ammonia, sulfur dioxide, carbon dioxide, and methyl chloride. Until 1986, nontoxic
and nonflammable halogenated hydrocarbons with various ozone depletion potentials were used
almost exclusively in vapor compression refrigeration systems for air conditioning. The impact of
ozone depletion of CFCs, halons, and HCFCs since the 1980s caused a worldwide decision to
phase out these refrigerants. A new classification of refrigerants into six groups based mainly on
ozone depletion will be helpful for the selection of nonozone depletion refrigerants as well as
replacement of CFCs by alternative refrigerants (Table 9.1).
Hydrofluorocarbons
HFCs contain only hydrogen, fluorine, and carbon atoms. They contain no chlorine atoms, therefore
are environmentally safe, and cause no ozone depletion. They are designated by the prefix HFC.
HFC-134a is an attractive, long-term alternative to replace CFC-12 in reciprocating, scroll,
screw, and centrifugal compressors; and a long-term alternative for HCFC-22. It has a low 0.28
HGWP. HFC-134a is nonflammable, has an extremely low toxicity, and is classified as AI in
ANSI/ASHRAE Standard 34-1997 safety rating.
HFC-134a has a molecular mass of 102.3 instead of CFC-12s molecular mass of 120.93. At a
condensing temperature of 100F (37.8C), HFC-134as condensing pressure is 138.83 psia (957
kPa abs.), whereas CFC-12s is 131.65 psia (908 kPa abs.). A larger impeller of higher speed is
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.13
needed for a centrifugal chiller to provide the same cooling capacity. Parsnow (1993) reported a
capacity loss of direct conversion from CFC-12 to HFC-134a of 8 to 10 percent, and an efficiency
loss of 1 to 2 percent. In Lowe and Ares (1995), in the conversion from CFC-12 to HFC-134a in the
Sears Tower centrifugal chillers, the compressors speed increased about 8.5 percent, there was a
cooling capacity loss of 12 to 24 percent, and efficiency was 12 to 16 percent worse.
HFC-134a has a poor mutual solubility with mineral oil because of a higher interfacial tension
between them. Polyolester-based synthetic lubricants should be used. Polyolester-based synthetic
oils are hygroscopic, so monitoring of the moisture content of the refrigerant is important. Halocarbons,
including HFC-134a, are compatible with containment materials. Concerning nonmetallic or
elastomer (such as gaskets) compatibility, Corr et al. (1993) reported that HFC-134a, an ester-based
synthetic oil mixture, has a smaller volume change of elastomer than CFC-12 and mineral oil.
HFC-134a may become one of the most widely used single-chemical-compound refrigerants during
the first half of the twenty-first century.
HFC-245ca also does not contain chlorine and bromine atoms, and its ODP is 0. Compared to
CFC-11, its efficiency will be 3 to 4 percent lower. Synthetic polyolester lubricant oil will be used.
Except for neoprene at high moisture content, common materials used in the refrigeration system
were shown to be compatible with HFC-245ca in tests. Smith et al. (1993) reported that mixtures of
HFC-245ca in air with a relative humidity of 43 percent and a HFC-245ca concentration range of
7 to 14.4 percent were observed to be flammable in tests.
Because of the higher isentropic work required by HFC-245ca compared to CFC-11 and HCFC-
123, for direct-drive centrifugal chillers, a large impeller is required during the conversion from
CFC-11 to HFC-245ca or from HCFC-123 to HFC-245ca.
HFC-245ca is a possible long-term alternative to CFC-11 and HCFC-123 in large centrifugal
chillers in the future. In the HFC group, HFC-32, HFC-125, HFC-143a, and HFC-152a all are seldom
used as a refrigerant of single compound only.
Azeotropic HFC
HFC-507 is an azeotrope of refrigerant blends of HFC-125/HFC-143a (45/55) of zero ozone
depletion and an HGWP of 0.96. It is a long-term alternative refrigerant to replace CFC-502 and
CFC-12 in low-temperature refrigeration systems whose evaporating temperatures are below10F
(12.2C). HFC-507 needs synthetic lubricant oil. According to ANSI/ASHRAE Standard 34-
1997, HFC-507 is allowed alternative designations for HFC-507A, a refrigeration blend of HFC-
125/HFC-143a (50/ 50).
The Linton et al. (1995) test results showed that compared to CFC-502, the cooling capacity of
HFC-507 was between 0.95 and 1.05. HFC-507 had an energy efficiency of 0.87 to 0.97 compared
to CFC-502.
Near-Azeotropic HFC
Near-azeotropic HFC is a refrigerant blend of zero ozone depletion and having rather small changes
in volumetric composition or saturation temperature, a small glide, when it evaporates or condenses
at a constant pressure. Near-azeotropic HFC-404A and HFC-410A require synthetic lubricant oil
instead of mineral oil and are nontoxic and nonflammable with a safety classification of A1/A1.
HFC-404A is a blend of HFC-125/HFC-143a/HFC-134a (44/52/4) of zero ozone depletion
and an HGWP of 0.94. It is a long-term alternative refrigerant for CFC-502 and CFC-12 both in
low-temperature refrigeration systems. HFC-404A has a temperature glide of 0.9F (0.5C) during
evaporation and a temperature glide of 0.6F (0.33C) during condensation. Snelson et al. (1995)
compared HFC-404A with CFC-502 from their test results. HFC-404A had the same, slightly
higher, or lower evaporating capacity at various condensing and evaporating temperatures. The energy
efficiency ratio of 0.89 to 0.99 was found at different evaporating temperatures Tev. The lower
the Tev, the lower the energy efficiency ratio, because of the higher compressor pressure ratio.
9.14 CHAPTER NINE
HFC-410A is a blend of HFC-32/HFC-125 (50/50) of zero ozone depletion and an HGWP of
0.43. It is a long-term alternative refrigerant to replace HCFC-22 and CFC-502. HFC-410A has a
temperature glide of 0.2F (0.11C) during evaporation and condensation. Hickman (1994) showed
that the compressor displacement, cfm/ton (L/ s kW), for HFC-410A is about 50 percent smaller
than that for HCFC-22; and the discharge pressure for 130F (54.4C) condensing is about 490 psia
(3379 kPa abs.), which is much higher than that for HCFC-22. It is often necessary to change the
original reciprocating compressor to a scroll compressor. A higher energy efficiency was reported
by a refrigerant manufacturer.
Zeotropic HFC
Zeotropic (nonazeotropic) HFCs are refrigerant blends of zero ozone depletion that have greater
temperature glide during evaporation and condensation. Zeotropic HFC-407A and HFC-407C also
require synthetic lubrication oil, instead of mineral oil; and both are nontoxic and nonflammable
with a safety classification of A1/A1.
HFC-407A is a blend of HFC-32/HFC-125/HFC-134a (20/40/40) of zero ozone depletion
with an HGWP of 0.49. It is a long-term alternative refrigerant for CFC-502 and CFC-12 in lowtemperature
refrigeration sysems. HFC-407A showed a reduction in heat transfer in the evaporator
of a low-temperature system during tests. The system performance of HFC-407A was the lowest
compared to HFC-404A and HFC-507.
HFC-407C is a blend of HFC-32/HFC-125/134a (23/25/52) of zero ozone depletion with an
HGWP of 0.38. It is a long-term alternative refrigerant to replace HCFC-22 and CFC-502. Bivens
et al. (1994) compared HFC-407C to HCFC-22 during tests. For cooling and heating, the capacity
ratio ranged from 0.93 to 1.06, and the energy ratio ranged from 0.94 to 0.97. In-tube heat-transfer
coefficients during evaporation and condensation were 85 to 95 percent of HCFC-22 values.
HCFCs and Their Zeotropes
HCFCs contain hydrogen, chlorine, fluorine, and carbon atoms and are not fully halogenated.
HCFCs have a much shorter atmospheric life than CFCs and cause far less ozone depletion (0.02 to
0.1 ODP). They are designated by the prefix HCFC. Their consumption is scheduled to be reduced
gradually starting from 2004 and will be completely phased out in 2030 in developed nations, except
for a limited amount for service, as discussed previously.
HCFC-22 has an ODP of 0.05 and an HGWP of 0.40. It is nonflammable with a safety classifi-
cation of A1. HCFC-22 is partially miscible with mineral oil. At 40F (4.4C), its evaporating pressure
is 82.09 psia (566 kPa abs.), and at 100F (37.8C) its condensing pressure is 201.5 psia (1389
kPa abs.), the highest of currently used HCFC and CFC refrigerants. HCFC-22 has a smaller compressor
displacement among the HCFCs and CFCs. All these factors make it an interim alternative
to replace CFC-12. HCFC-22 was the most widely used refrigerant in reciprocating and scroll compressors
in small and medium-size packaged units in the 1990s in the United States.
HCFC-123 is an interim alternative to replace CFC-11 in low-pressure centrifugal chillers. It has
an ODP of 0.02 and a very low HGWP of 0.02. HCFC-123 is nonflammable and of lower toxicity
with a safety classification of B1. In 1997, DuPont raised the allowable exposure limit (AEL) of
HCFC-123 to 50 ppm. Smithhart and Crawford (1993) reported that for chillers with direct conversion
from CFC-11 to HCFC-123, there was about a decrease of 0 to 5 percent in capacity and a 2 to
4 percent decrease in efficiency. A conversion of refrigerant from CFC-11 to HCFC-123 in an existing
chiller may require changing its lubricants, seals, and motor windings of hermetic compressors.
Because HCFC-123 has a low ODP, a very low HGWP, and a 15 percent higher energy effi-
ciency in centrifugal chillers than HFC-134a does, if no acceptable alternative can be found, the use
of HCFC-123 in centrifugal chillers may be considered longer than the cap specified in the Vienna
meeting in 1995, as listed in Sec. 9.4, in the twenty-first century.
HCFC-124 has an ODP of 0.02. It is nonflammable and has a safety classification of A1. HCFC-
124 is an interim alternative refrigerant to replace CFC-114.
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.15
Near-azeotropic HCFC-402A is a blend of HCFC-22/HFC-125/PRO-290 (38/60/2) with an
ODP of 0.02 and an HGWP of 0.63. Here PRO-290 represents propane, which is a more highly
flammable refrigerant with a safety classification of A3. HCFC-402A is nonflammable and has a
safety classification of A1/A1. It needs polyolester or alkyl-benzene-based lubricant oil. HCFC-
402A is an interim alternative refrigerant to replace CFC-502.
Zeotropic HCFC-401A is a blend of HCFC-22/HCFC-124/HFC-152a (53/34/13) with an ODP
of 0.037 and an HGWP of 0.22; and HCFC-401B is a blend of HCFC-22/HCFC-124/HFC-152a
(61/28/11) with an ODP of 0.04 and an HGWP of 0.24. Both HCFC-401A and HCFC-401B are
nonflammable and have a safety classification of A1/A1. They both need alkyl-benzene-based lubricant
oil. HCFC-401A is an interim alternative refrigerant to replace CFC-12, and HCFC-401B is an
interim alternative refrigerant to replace CFC-12 in low-temperature refrigeration systems.
Inorganic Compounds
These compounds include ammonia (NH3), water (H2O), and gases used in the gas expansion systems.
As refrigerants, they were used far earlier than the halocarbons. Air is a mixture of nitrogen,
oxygen, argon, rare gases, and water vapor. Air has zero ozone depletion and is a zeotropic blend that
has a temperaure glide of 5.5F (3C) at atmospheric pressure. Ammonia also has zero ozone
depletion. It has a high operating pressure at 40F (4.4C) evaporating and 100F (37.8C) condensing.
Ammonia compressors show a smaller cfm/ton displacement and higher energy efficiency than
HCFC-22 compressors. Leakage of ammonia is easily detected due to its objectionable odor.
Ammonia attacks copper even in the presence of a small amount of moisture. It is of higher toxicity.
An ammonia-air mixture is flammable if the concentration of NH3 by volume is within 16 to
25 percent. The mixture may explode if the ignition source is above 1200F (650C). Because the
safety classification of ammonia is B2lower flammability and higher toxicityit is not allowed
to be used in comfort air conditioning in the United States.
Water has a zero ODP and is readily available. At 40F (4.4C) evporating and 100F (37.8C)
condensing, waters evaporating and condensing pressures are both below atmospheric. Air and
other noncondensable gases must be purged out of the refrigeration system periodically.
CFCs, Halons, and Their Zeotropes
CFCs including CFC-11, CFC-12, CFC-113, and CFC-114, have an ODP from 0.8 to 1.0. Halons
including BFC-13B1 have an ODP of 10. Their azeotropics CFC-500 and CFC-502 have ODPs of
0.74 and 0.22, respectively. Production of all these CFCs, halons, and their azeotropes ceased in developed
nations since January 1, 1996. However, a limited amount of these refrigerants, used to service
the refrigeration systems that have not been converted or replaced by nonozone depletion refrigerants,
may be extended to the beginning of the twenty-first century.
9.6 REFRIGERATION PROCESSES AND REFRIGERATION CYCLES
Refrigeration Processes
A refrigeration process indicates the change of thermodynamic properties of the refrigerant and the
energy transfer between the refrigerant and the surroundings. The following refrigeration processes
occur during the operation of a vapor compression refrigerating system:
 Evaporation. In this process, the refrigerant evaporates at a lower temperature than that of its
surroundings, absorbing its latent heat of vaporization.
 Superheating. Saturated refrigerant vapor is usually superheated to ensure that liquid refrigerant
does not flow into the compressor.
9.16 CHAPTER NINE
 Compression. Refrigerant is compressed to a higher pressure and temperature for condensation.
 Condensation. Gaseous refrigerant is condensed to liquid form by being desuperheated, then condensed,
and finally subcooled, transferring its latent heat of condensation to a coolant.
 Throttling and expansion. The higher-pressure liquid refrigerant is throttled to the lower evaporating
pressure and is ready for evaporation.
The following refrigeration processes occur during the operation of an air or gas expansion
refrigeration system:
 Compression. Air or gas is compressed to a higher pressure and temperature.
 Heat release. Heat is released to the surroundings at constant pressure in order to reduce the temperature
of the air or gas.
 Throttling and expansion. Air or gas is throttled and expanded so that its temperature is lowered.
 Heat absorption. Heat is absorbed from the surroundings because of the lower air or gas temperature.
Refrigeration Cycles
Most refrigerants undergo a series of evaporation, compression, condensation, throttling, and expansion
processes, absorbing heat from a lower-temperature reservoir and releasing it to a highertemperature
reservoir in such a way that the final state is equal in all respects to the initial state. It is
said to have undergone a closed refrigeration cycle. When air or gas undergoes a series of compression,
heat release, throttling, expansion, and heat absorption processes, and its final state is not
equal to its initial state, it is said to have undergone an open refrigeration cycle.
Both vapor compression and air or gas expansion refrigeration cycles are discussed in this chapter.
Absorption refrigeration cycles are covered in Chap. 14.
Unit of Refrigeration
In inch-pound (I-P) units, refrigeration is expressed in British thermal units per hour, or simply Btu/h.
A British thermal unit is defined as the amount of heat energy required to raise the temperature of one
pound of water one degree Fahrenheit from 59F to 60F; and 1 Btu/h  0.293 watt (W).
Another unit of refrigeration widely used in the HVAC&R industry is ton of refrigeration, or
simply ton. As mentioned before, 1 ton  12,000 Btu/h of heat removed. This equals the heat
absorbed by 1 ton (2000 lb) of ice melting at a temperature of 32F over 24 h.
Because the heat of fusion of ice at 32F is 144 Btu / lb,
also 1 ton  3.516 kW
9.7 GRAPHICAL AND ANALYTICAL EVALUATION
OF REFRIGERATION
Pressure-Enthalpy Diagram
The pressure-enthalpy p-h diagram is the most common graphical tool for analysis and calculation
of the heat and work transfer and performance of a refrigeration cycle. A single-stage refrigeration
1 ton 
1  2000  144
24
 12,000 Btu / h
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.17
cycle consists of two regions: the high-pressure region, or high side, and the low-pressure region, or
low side. The change in pressure can be clearly illustrated on the p-h diagram. Also, both heat and
work transfer of various processes can be calculated as the change of enthalpy and are easily shown
on the p-h diagram.
Figure 9.3 is a skeleton p-h diagram for refrigerant HCFC-22. Enthalpy h (in Btu/ lb) is the
abscissa, and absolute pressure (psia) or gauge pressure (psig), both expressed in logarithmic scale,
is the ordinate. The saturated liquid line separates the subcooled liquid from the two-phase region in
which vapor and liquid refrigerants coexist. The saturated vapor line separates this two-phase
region from the superheated vapor. In the two-phase region, the mixture of vapor and liquid is
subdivided by the constant-dryness-fraction quality line.
The constant-temperature lines are nearly vertical in the subcooled liquid region. At higher temperatures,
they are curves near the saturated liquid line. In the two-phase region, the constant-temperature
lines are horizontal. In the superheated region, the constant-temperature lines curve down
sharply. Because the constant-temperature lines and constant-pressure lines in the two-phase region
are horizontal, they are closely related. The specific pressure of a refrigerant in the two-phase
region determines its temperature, and vice versa.
Also in the superheated region, the constant-entropy lines incline sharply upward, and constantvolume
lines are flatter. Both are slightly curved.
Temperature-Entropy Diagram
The temperature-entropy T-s diagram is often used to analyze the irreversibilities in a refrigeration
cycle, as well as in the system, in order to select optimum operating parameters and improve performance
of the system. In a temperature-entropy T-s diagram, entropy s, Btu / lbR, is the abscissa of
9.18 CHAPTER NINE
FIGURE 9.3 Skeleton of pressure-enthalpy p-h diagram for HCFC-22.
the diagram and temperature T, R, is the ordinate. A T-s diagram is more suitable for evaluating
the effectiveness of an air expansion refrigeration cycle.
Analytical Evaluation of Cycle Performance
Swers et al. (1972) proposed a thermodynamic analysis of degradation of available energy and irreversibility
in a refrigerating system, and Tan and Yin (1986) recommended a method of exergy
analysis. The exergy of a working substance e, Btu / lb (kJ / kg), is defined as
e  h  ha  TRa (s  sa) (9.1)
where h, ha  enthalpy of working substance and ambient state, Btu / lb (kJ /kg)
TRa  absolute temperature of ambient state, R (K)
s, sa  entropy of working substance and ambient state, Btu / lbR (kJ/kgK)
Both analyses are effective tools in the selection of optimum design and operating parameters by
means of complicated analysis. They require extensive supporting data and information.
For most analyses of refrigeration cycle performance and design and operation of refrigeration
systems in actual applications, satisfactory results can be obtained by using the steady flow energy
equation, heat and work transfer, and energy balance principle. If a more precise and elaborate
analysis is needed in research or for detailed improvements of refrigeration systems, the references
at the end of this chapter can be consulted.
9.8 CARNOT REFRIGERATION CYCLE
The Carnot refrigeration cycle is a reverse engine cycle. All processes in a Carnot refrigeration
cycle are reversible, so it is the most efficient refrigeration cycle.
Figure 9.4a is a schematic diagram of a Carnot cycle refrigerating system, and Fig. 9.4b shows
the Carnot refrigeration cycle using gas as the working substance. This Carnot cycle is composed of
four reversible processes:
1. An isothermal process 4-1 in which heat q#1 is extracted at constant temperature TR1 per lb (kg)
of working substance
2. An isentropic compression process 1-2
3. An isothermal process 2-3 in which q#2 is rejected at constant temperature TR2 per lb (kg) of
working substance
4. An isentropic expansion process 3-4
Figure 9.4c shows the Carnot refrigeration cycle using vapor as the working substance. Wet vapor
is the only working substance where heat supply and heat rejection processes can occur easily
at constant temperature. This is because the temperatures of wet vapor remain constant when latent
heat is supplied or rejected.
As in the gas cycle, there are two isothermal processes 4-1 and 2-3 absorbing heat at temperature
TR1 and rejecting heat at TR2, respectively, and two isentropic processes, one for compression
1-2 and another for expansion 3-4.
Performance of Carnot Refrigeration Cycle
According to the first law of thermodynamics, often called the law of conservation of energy, when
a system undergoes a thermodynamic cycle, the net heat supplied to the system is equal to the net
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.19
work done, or
Heat supply  heat rejected  net work done (9.2)
Referring to Fig. 9.4a, in a Carnot refrigeration cycle,
q#1  q#2   W
or q#1  q#2  W (9.3)
q#2  q#1  W
where q#1  heat supplied from surroundings per lb (kg) of working substance at temperature T1;
sign of q#1 is positive
q#2  heat rejected to sink per lb (kg) of working substance at temperature T2; sign of q#2
is negative
W  net work done by system; sign is positive, or if it is a work input to system, sign is
negative
9.20 CHAPTER NINE
FIGURE 9.4 Carnot refrigeration cycle: (a) schematic diagram; (b) gas cycle; (c) vapor cycle.
The heat extracted from the source at temperature TR1 by the working substance, i.e., the refrigerating
effect per lb (kg) of working substance, is
q#1  TR1(s1  s4) (9.4)
where s1, s4  entropy at state points 1 and 4, respectively, Btu / lbR (kJ/kgK). Heat rejected to
the heat sink at temperature TR2 can be calculated as
q#2   TR2(s3  s2)  TR2 (s2  s3) (9.5)
where s2, s3  entropy at state points 2 and 3, respectively, Btu / lbR (kJ/kgK). Because in the
isentropic process 1-2, s1  s2, and in isentropic process 3-4, s3  s4,
q#2  TR2(s1  s4) (9.6)
9.9 COEFFICIENT OF PERFORMANCE
OF REFRIGERATION CYCLE
The coefficient of performance is an index of performance of a thermodynamic cycle or a thermal
system. Because the COP can be greater than 1, COP is used instead of thermal efficiency. The
coefficient of performance can be used for the analysis of the following:
 A refrigerator that is used to produce a refrigeration effect only, that is, COPref
 A heat pump in which the heating effect is produced by rejected heat COPhp
 A heat recovery system in which both the refrigeration effect and the heating effect are used at the
same time, COPhr
For a refrigerator, COP is defined as the ratio of the refrigeration effect q#1 to the work input Win,
both in Btu/ lb (kJ / kg), that is,
COPref  Refrigeration effect /Work input
 q#1 /Win (9.7)
For the Carnot refrigeration cycle, from Eq. (9.3),
(9.8)
With a heat pump, the useful effect is the heating effect because of the rejected heat q#2, so COPhp is
the ratio of heat rejection to the work input, or
(9.9)
For a heat recovery system, the useful effect is q#1 and q#2. In such a condition, COPhr is defined as
the ratio of the sum of the absolute values of refrigerating effect and heat rejection to the absolute
value of the work input, i.e.,
(9.10) COPhr 
q#1   q#2 
Win
COPhp 
q#2
Win

TR1(s1  s4)
(TR2  TR1)(s1  s4)

TR1
TR2  TR1
COPref 
q#1
q#2  q#1
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.21
9.10 SINGLE-STAGE IDEAL VAPOR COMPRESSION CYCLE
The Carnot cycle cannot be achieved for the vapor cycle in actual practice because liquid slugging
would occur during compression of the two-phase refrigerant. In addition, the mixture, mostly liquid,
does very little work when it expands after condensation in the heat engine. Therefore, a single-
stage ideal vapor compression cycle is used instead of the Carnot cycle.
Figure 9.5 shows an ideal single-stage vapor compression cycle in which compression occurs in
the superheated region. A throttling device, such as an expansion valve, is used instead of the heat
engine. Single-stage means that there is only one stage of compression. An ideal cycle is one in
which the compression process is isentropic and the pressure losses in the pipeline, valves, and
other components are negligible. All the refrigeration cycles covered in this chapter are ideal cycles
except the air expansion refrigeration cycle.
Vapor compression means that the vapor refrigerant is compressed to a higher pressure, and then
the condensed liquid is throttled to a lower pressure to produce the refrigerating effect by evaporation.
It is different from the absorption or air expansion refrigeration cycle.
Flow Processes
Figure 9.5b and c shows the refrigeration cycle on p-h and T-s diagrams. The refrigerant evaporates
entirely in the evaporator and produces the refrigerating effect. It is then extracted by the compressor
at state point 1, compressor suction, and is compressed isentropically from state point 1 to 2. It
is next condensed to liquid in the condenser, and the latent heat of condensation is rejected to the
heat sink. The liquid refrigerant, at state point 3, flows through an expansion valve, which reduces it
to the evaporating pressure. In the ideal vapor compression cycle, the throttling process at the expansion
valve is the only irreversible process, usually indicated by a dotted line. Some of the liquid
flashes into vapor and enters the evaporator at state point 4. The remaining liquid portion evaporates
at the evaporating temperature, thus completing the cycle.
Cycle Performance
For the evaporating process between points 4 and 1, according to the steady flow energy equation,
(9.11)
where h1, h4  enthalpy of refrigerant at points 1 and 4, respectively, Btu / lb (J /kg)
v1, v4  velocity of refrigerant at points 1 and 4, respectively, ft / s (m/ s)
q#  heat supplied per lb (kg) of working substance during evaporation process, Btu/ lb
(J /kg)
gc  dimensional conversion factor, 32 lbmft / lbf  s2
Because no work is done during evaporation, the change of kinetic energy between points 4 and 1 is
small compared with other terms in Eq. (9.11), and it is usually ignored. Then
h4  q#  h1  0
The refrigerating effect qrf, Btu / lb (J / kg), is
qrf  q#  h1  h4 (9.12)
h4  v4 2
2gc  778
 q# 
h1  v1 2
2gc  778
 W
9.22 CHAPTER NINE
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.23
FIGURE 9.5 Single-stage ideal vapor compression cycle: (a)
schematic diagram; (b) p-h diagram; (c) T-s diagram.
For isentropic compression between points 1 and 2, applying the steady flow energy equation and
ignoring the change of kinetic energy, we have
h1  0  h2  W
 W  h2  h1
Work input to the compressor Win, Btu / lb (kJ / kg), is given as
Win  h2  h1 (9.13)
Similarly, for condensation between points 2 and 3,
h2  q#  h3  0
The heat released by the refrigerant in the condenserq#, Btu / lb (kJ / kg), is
q#  h2  h3 (9.14)
For the throttling process between points 3 and 4, assuming that the heat loss is negligible,
h3  0  h4  0
or h3  h4 (9.15)
The COPref of the single-stage ideal vapor compression cycle is
(9.16)
The mass flow rate of refrigerant , lb/h (kg/ s), flowing through the evaporator is
(9.17)
where Qrc  refrigerating capacity, Btu /h (W). From Eq. (9.16), the smaller the difference between
the condensing and evaporating pressures, or between condensing and evaporating temperatures,
the lower the Win input to the compressor at a specific Qrc and, therefore, the higher the COP.
A higher evaporating pressure pev and evaporating temperature Tev or a lower condensing pressure
pcon and condensing temperature Tcon will always conserve energy.
Determination of Enthalpy by Polynomials
During the performance analysis of a refrigeration cycle, the enthalpies h of the refrigerant at
various points must be determined in order to calculate the refrigeration effect, work input, and
COP. The enthalpy of refrigerant at saturated liquid and saturated vapor state is a function of
saturated temperature or pressure. In other words, saturated temperature Ts and saturated pressure
pss of the refrigerant are dependent upon each other. Therefore, it is more convenient to evaluate
the enthalpy of refrigerant in terms of saturated temperature Ts within a certain temperature
range
h  f (Ts) (9.18)
m?r 
Qrc
qrf
m?r

qrf
Win

h1  h4
h2  h1
COPref 
refrigerating effect
work input
9.24 CHAPTER NINE
The enthalpy differential along the constant-entropy line within a narrower temperature range can
be calculated as
h2  h1  F(Ts2  Ts1) (9.19)
where h1, h2  enthalpy of refrigerant on constant-entropy line at points 1 and 2, Btu/ lb (kJ /kg)
Ts1, Ts2  temperature of saturated refrigerant at points 1and 2,F (C)
From the refrigerant tables published by ASHRAE, the following polynomial can be used to calculate
the enthalpy of saturated liquid refrigerant hlr, Btu / lb (kJ / kg), from its temperature Ts1 at a saturated
temperature from 20 to 120F ( 7 to 50C) with acceptable accuracy:
hlr  a1  a2Ts1  a3Ts1
2  a4Ts1
3 (9.20)
where a1, a2, a3, a4  coefficients. For HCFC-22,
a1  10.409 a2  0.26851 a3  0.00014794 a4  5.3429  107
Similarly, the polynomial that determines the enthalpy of saturated vapor refrigerant hvr, Btu / lb
(kJ/ kg), from its temperature Tsv in the same temperature range is
hvr  b1  b2Tsv  b3Tsv
2  b4Tsv
3 (9.21)
where b1, b2, b3, b4  coefficients. For HCFC-22,
b1  104.465 b2  0.098445 b3  0.0001226 b4  9.861  107
The polynomial that determines the enthalpy changes of refrigerant along the constant-entropy line
for an isentropic compression process between initial state 1 and final state 2 is
(9.22)
where c1, c2, c3, c4  coefficients
Ts1, Ts2  saturated temperature of vapor refrigerant corresponding to its pressure at initial
state 1 and final state 2,F (C)
For HCFC-22 within a saturated temperature range of 20 to 100F:
c1  0.18165 c2  0.21502 c3  0.0012405 c4  8.198  106
Computer programs are available that calculate the coefficients based on ASHRAEs refrigerant tables
and charts.
Refrigeration Effect, Refrigerating Load, and Refrigerating Capacity
The refrigeration effect qrf , Btu / lb (J /kg or kJ/ kg), is the heat extracted by a unit mass of refrigerant
during the evaporating process in the evaporator. It can be calculated as
qrf  hlv  hen (9.23)
where hen, hlv  enthalpy of refrigerant entering and leaving evaporator, Btu / lb (J / kg). Refrigerating
load Qrl, Btu /h (W), is the required rate of heat extraction by the refrigerant in the evaporator. It
h2  h1  c1  c2(Ts2  Ts1)  c3(Ts2  Ts1)2  c4(Ts2  Ts1)3
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.25
can be calculated as
(9.24)
where mass flow rate of refrigerant flowing through evaporator, lb/h (kg/ s).
Refrigerating capacity, or cooling capacity, Qrc, Btu /h (W), is the actual rate of heat extracted by
the refrigerant in the evaporator. In practice, the refrigeration capacity of the equipment selected is
often slightly higher than the refrigerating load. This is because the manufacturers specifications
are a series of fixed capacities. Occasionally, equipment can be selected so that its capacity is just
equal to the refrigeration load required. Refrigeration capacity Qrc can be calculated as
(9.25)
where hren, hrlv  enthalpy of refrigerant actually entering and leaving evaporator, Btu / lb (J /kg)
9.11 SUBCOOLING AND SUPERHEATING
Subcooling
Condensed liquid refrigerant is usually subcooled to a temperature lower than the saturated temperature
corresponding to the condensing pressure of the refrigerant, shown in Fig. 9.6a as point 3	.
This is done to increase the refrigerating effect. The degree of subcooling depends mainly on the
temperature of the coolant (e.g., atmospheric air, surface water, or well water) during condensation,
and the construction and capacity of the condenser.
The enthalpy of subcooled liquid refrigerant hsc, Btu / lb (J / kg), can be calculated as
hsc  hs,con  cpr(Ts,con  Tsc) (9.26)
where hs,con  enthalpy of saturated liquid refrigerant at condensing temperature, Btu/ lb (J /kg)
cpr  specific heat of liquid refrigerant at constant pressure, Btu/ lb F (J /kgC)
Ts,con  saturated temperature of liquid refrigerant at condensing pressure,F (C)
Tsc  temperature of subcooled liquid refrigerant,F (C)
Enthalpy hsc is also approximately equal to the enthalpy of the saturated liquid refrigerant at subcooled
temperature.
Superheating
As mentioned before, the purpose of superheating is to avoid compressor slugging damage. Superheating
is shown in Fig. 9.6b. The degree of superheat depends mainly on the type of refrigerant
feed and compressor as well as the construction of the evaporator. These are covered in detail in
Chap. 11.
Example 9.1. A 500-ton (1760-kW) single-stage centrifugal vapor compression system uses
HCFC-22 as refrigerant. The vapor refrigerant enters the compressor at dry saturated state. The
compression process is assumed to be isentropic. Hot gas is discharged to the condenser and
condensed at a temperature of 95F (35C). The saturated liquid refrigerant then flows through a
throttling device and evaporates at a temperature of 35F (1.7C). Calculate:
1. The refrigeration effect
2. The work input to the compressor
Qrc  m?r (hrlv  hren)
m?r 
Qrl  m?r (hlv  hen)
9.26 CHAPTER NINE
3. The coefficient of performance of this refrigeration cycle
4. The mass flow rate of the refrigerant
Recalculate the COP and the energy saved in work input if the refrigerant is subcooled to a temperature
of 90F (32.2C).
Solution
1. From Eq. (9.20), the enthalpy of the saturated liquid refrigerant at a temperature of 95F (35C),
point 3 as shown in Fig. 9.5a, is
 10.409  25.508  1.335  0.458  37.71 Btu / lb
 10.409  0.26851(95)  0.0001479(95)2  5.3429  107(95)3
h3  h4  10.409  0.26851Ts  0.0001479T s
2  5.3429  107T s
3
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.27
FIGURE 9.6 (a) Subcooling and (b) superheating.
From Eq. (9.21), the enthalpy of saturated vapor refrigerant at a temperature of 35F (1.7C),
point 1, is
Then the refrigeration effect is calculated as
qrf  h1  h4  107.72  37.71  70.01 Btu/ lb (162.8 kJ /kg)
2. From Eq. (9.22), the enthalpy differential h2  h1 on the constant-entropy line corresponding to
a saturated temperature differential Ts2  Ts1  95 35  60F in the two-phase region is
That is, the work input Win  h2  h1  10.024 Btu/ lb (23.73 kJ / kg).
3. According to Eq. (9.16), COPref of the refrigerating system is calculated as
4. From Eq. (9.17), the mass flow rate of the refrigerant can be calculated as
If the liquid refrigerant is subcooled to a temperature of 90F, (32.2C), then
h3  h4  10.409  0.26851(90)  0.0001479(90)2  5.3429  107(90)3
 10.409  24.166  1.198  0.389  36.16 Btu/ lb
Refrigeration effect is then increased to
qrf  107.72  36.16  71.56 Btu/ lb (166 kJ/kg)
Also COPref is increased to
Since 1 ton  200 Btu/min and 1 hp  42.41 Btu/min, electric power input to the compressor Pin
without subcooling is
Pin 
500  200
42.41  6.98
 337.8 hp (252 kW)
COPref 
71.56
10.024
 7.14
m? r 
Qrc
qrf

500  12,000
70.01
 85,702 lb / h (38,874 kg / h)
COPref 
qrf
Win

70.01
10.024
 6.98
 0.182  12.901  4.466  1.771  10.02 Btu / lb
  0.18165  0.21502(60)  0.0012405(60)2  8.1982  106(60)3
8.1982  106(Ts2  Ts1)3
h2  h1  0.18165  0.21502(Ts2  Ts1)  0.0012405(Ts2  Ts1)2
 104.47  3.445  0.150  0.042  107.72 Btu / lb
 104.47  0.098445(35)  0.0001226(35)2  9.861  107(35)3
h1  104.465  0.098445Ts  0.0001226T s
2  9.861  107T s
3
9.28 CHAPTER NINE
With subcooling,
Savings in electric energy are calculated as
9.12 MULTISTAGE VAPOR COMPRESSION SYSTEMS
When a refrigeration system uses more than single-stage compression process, it is called a multistage
system (as shown in Fig. 9.7), and may include the following:
1. A high-stage compressor and a low-stage compressor
2. Several compressors connected in series
3. Two or more impellers connected internally in series and driven by the same motor or prime
mover, as shown in Fig. 9.7
4. A combination of two separate refrigeration systems
The discharge pressure of the low-stage compressor, which is equal to the suction pressure of the
high-stage compressor, is called the interstage pressure.
The reasons for using a multistage vapor compression system instead of a single-stage system
are as follows:
1. The compression ratio Rcom of each stage in a multistage system is smaller than that in a singlestage
unit, so compressor efficiency is increased. Compression ratio Rcom is defined as the ratio
of the compressors discharge pressure pdis, psia (kPa abs.), to the suction pressure at the compressors
inlet psuc, psia (kPa abs.), or
(9.27)
2. Liquid refrigerant enters the evaporator at a lower enthalpy and increases the refrigeration effect.
3. Discharge gas from the low-stage compressor can be desuperheated at the interstage pressure.
This results in a lower discharge temperature from the high-stage compressor than would be
produced by a single-stage system at the same pressure differential between condensing and
evaporating pressures.
4. Two or three compressors in a multistage system provide much greater flexibility to accommodate
the variation of refrigeration loads at various evaporating temperatures during part-load operation.
The drawbacks of the multistage system are higher initial cost and a more complicated system
than that for a single-stage system.
Compound Systems
Multistage vapor compression systems are classified as compound systems or cascade systems.
Cascade systems are discussed in a later section.
A compound system consists of two or more compression stages connected in series. For
reciprocating, scroll, or screw compressors, each compression stage usually requires a separate
Rcom 
pdis
psuc
337.8  330.2
337.8
 0.022, or 2.2%
Pin, s 
500  200
42.41  7.14
 330.2 hp (246 kW)
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.29
compressor. In multistage centrifugal compressors, two or more stages may be internally compounded
by means of several impellers connected in series.
Interstage Pressure
Interstage pressure is usually set so that the compression ratio at each stage is nearly the same for
higher COPs. For a two-stage compound system, interstage pressure pi, psia (kPa abs.), can be
9.30 CHAPTER NINE
FIGURE 9.7 Two-stage compound system with a flash cooler: (a)
schematic diagram; (b) refrigeration cycle.
calculated as
(9.28)
where pcon  condensing pressure, psia (kPa abs.)
pev  evaporating pressure, psia (kPa abs.)
For a multistage vapor compression system with z stages, the compression ratio Rcom for each stage
can be calculated as
(9.29)
Flash Cooler and Intercooler
In compound systems, flash coolers are used to subcool liquid refrigerant to the saturated temperature
corresponding to the interstage pressure by vaporizing part of the liquid refrigerant. Intercoolers
are used to desuperheat the discharge gas from the low-stage compressor and, more often, to
subcool also the liquid refrigerant before it enters the evaporator.
9.13 TWO-STAGE COMPOUND SYSTEM WITH
A FLASH COOLER
Flow Processes
Figure 9.7a is a schematic diagram of a two-stage compound system with a flash cooler, and Fig. 9.7b
shows the refrigeration cycle of this system. Vapor refrigerant at point 1 enters the first-stage impeller
of the centrifugal compressor at the dry saturated state. It is compressed to the interstage pressure pi at
point 2 and mixes with evaporated vapor refrigerant from the flash cooler, often called an economizer.
The mixture then enters the second-stage impeller at point 3. Hot gas, compressed to condensing pressure
pcon, leaves the second-stage impeller at point 4. It is then discharged to the condenser, in which
the hot gas is desuperheated, condensed, and often subcooled to liquid state at point 5	. After the condensing
process, the subcooled liquid refrigerant flows through a throttling device, such as a float
valve, at the high-pressure side. A small portion of the liquid refrigerant flashes into vapor in the flash
cooler at point 7, and this latent heat of vaporization cools the remaining liquid refrigerant to the
saturation temperature corresponding to the interstage pressure at point 8. Inside the flash cooler, the
mixture of vapor and liquid refrigerant is at point 6. Liquid refrigerant then flows through another
throttling device, a small portion is flashed at point 9, and the liquid-vapor mixture enters the evaporator.
The remaining liquid refrigerant is vaporized at point 1 in the evaporator. The vapor then flows to
the inlet of the first-stage impeller of the centrifugal compressor and completes the cycle.
Fraction of Evaporated Refrigerant in Flash Cooler
In the flash cooler, out of 1 lb of refrigerant flowing through the condenser, x lb of it cools down the
remaining portion of liquid refrigerant (1  x) lb to saturated temperature T8 at interstage pressure
pi. Because h5	 is the enthalpy of the subcooled liquid refrigerant entering the flash cooler, h6 is the
enthalpy of the mixture of vapor and liquid refrigerant after the throttling device, for a throttling
process, h5	  h6. Enthalpies h7 and h8 are the enthalpies of the saturated vapor and saturated
liquid, respectively, at the interstage pressure, and h9 is the enthalpy of the mixture of vapor and
liquid refrigerant leaving the flash cooler after the low-pressure side throttling device. Again, for a
throttling process, h8  h9.
Rcom   pcon
pev 1/z
pi  ?pcon pev
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.31
If the heat loss from the insulated flash cooler to the ambient air is small, it can be ignored. Heat
balance of the refrigerants entering and leaving the flash cooler, as shown in Fig. 9.8a, gives
Sum of heat energy of refrigerant entering flash cooler
 Sum of heat energy of refrigerant leaving flash cooler
that is,
h5	  xh7  (1  x)h8
The fraction of liquid refrigerant evaporated in the flash cooler x is given as
(9.30)
The fraction x also indicates the quality, or dryness fraction, of the vapor and liquid mixture in the
flash cooler at the interstage pressure.
Enthalpy of Vapor Mixture Entering Second-Stage Impeller
Ignoring the heat loss from mixing point 3 to the surroundings, we see that the mixing of the
gaseous refrigerant discharged from the first-stage impeller at point 2 and the vaporized refrigerant
from the flash cooler at point 7 is an adiabatic process. The heat balance at the mixing point before
the second-stage impeller, as shown in Fig. 9.7b, is given as
h3  (1  x)h2  xh7 (9.31)
where h2  enthalpy of gaseous refrigerant discharged from first-stage impeller, Btu/ lb (kJ /kg)
h3  enthalpy of mixture at point 3, Btu/ lb (kJ /kg)
h7  enthalpy of saturated vapor refrigerant from flash cooler at point 7, Btu/ lb (kJ /kg)
x 
h5	  h8
h7  h8
9.32 CHAPTER NINE
FIGURE 9.8 Heat balance of entering and leaving refrigerants in a
flash cooler and at the mixing point: (a) in the flash cooler; (b) at the
mixing point 3 before entering the second-stage impeller.
Coefficient of Performance
For 1 lb (kg) of refrigerant flowing through the condenser, the amount of refrigerant flowing
through the evaporator is (1  x) lb (kg). The refrigeration effect qrf per lb (kg) of refrigerant flowing
through the condenser, Btu / lb, (kJ/ kg), can be expressed as
qrf  (1  x)(h1  h9) (9.32)
where h1  enthalpy of saturated vapor leaving evaporator, Btu / lb (kJ /kg)
h9  enthalpy of refrigerant entering evaporator, Btu / lb (kJ /kg)
Total work input to the compressor (including the first- and second-stage impeller) Win per lb (kg)
of refrigerant flowing through the condenser, Btu / lb (kJ / kg), is
Win  (1  x)(h2  h1)  h4  h3 (9.33)
where h4  enthalpy of the hot gas discharged from the second-stage impeller, Btu/ lb (kJ / kg). The
coefficient of performance of the two-stage compound system with a flash cooler COPref is
(9.34)
The mass flow rate of refrigerant at the condenser , lb/h (kg/ s), is
(9.35)
where Qrc  refrigeration capacity, Btu /h (W).
Characteristics of Two-Stage Compound System with Flash Cooler
In a two-stage compound system with a flash cooler, a portion of the liquid refrigerant is flashed
into vapor going directly to the second-stage suction inlet, so less refrigerant is compressed in
the first-stage impeller. Furthermore, the remaining liquid refrigerant is cooled to the saturated
temperature corresponding to the interstage pressure, which is far lower than the subcooled liquid
temperature in a single-stage system. The increase in refrigeration effect and the drop in compression
work input lead to a higher COPref than in a single-stage system.
Although the initial cost of a two-stage compound system is higher than that for a single-stage
system, the two-stage compound system with a flash cooler is widely used in large central hydronic
air conditioning systems because of the high COPref.
Example 9.2. For the same 500-ton (1758-kW) centrifugal vapor compression system as in
Example 9.1, a two-stage compound system with a flash cooler is used instead of a single-stage
centrifugal compressor. Vapor refrigerant enters the first-stage impeller at a dry saturated state, and
the subcooled liquid refrigerant leaves the condenser at a temperature of 90F (32.2C). Both
compression processes at the first-stage impeller and the second-stage impeller are assumed to be
isentropic. Evaporating pressure is 76.17 psia (525kPa abs.), and the condensing pressure is 196.5
psia (1355 kPa abs.). Other conditions remain the same as in Example 9.1. Calculate
1. The fraction of liquid refrigerant vaporized in the flash cooler
2. The refrigeration effect per lb (kg) of refrigerant flowing through the condenser
3. The total work input to the compressor
m? r 
Qrc
qrf

Qrc
(1  x)(h1  h9)
m? r

(1  x)(h1  h9)
(1  x)(h2  h1)  (h4  h3)
COPref 
qrf
Win
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.33
4. The coefficient of performance of this refrigerating system
5. The mass flow rate of refrigerant flowing through the condenser
6. The percentage of saving in energy consumption compared with the single-stage vapor compression
system
Solution
1. Based on the data calculated in Example 9.1, enthalpy of vapor refrigerant entering the firststage
impeller h1  107.72 Btu/lb and the enthalpy of the subcooled liquid refrigerant leaving
the condenser h5	  36.162 Btu/ lb, as shown in Fig. 9.6b. From Eq. (9.28) and the given data,
the interstage pressure can be calculated as
From the Table of Properties of Saturated Liquid and Vapor for HCFC-22 in ASHRAE Handbook
1989, Fundamentals, for pi  122.34 psia, the corresponding interstage saturated temperature Ti in
the two-phase region is 63.17F.
From Eq. (9.20), the enthalpy of liquid refrigerant at saturated temperature 63.17F is
h8  h9  10.409  0.26851(63.17)  0.00014794(63.17)2  5.3429(63.17)3
 10.409  16.961  0.59  0.135  28.095 Btu/ lb
Also, from Eq. (9.21), the enthalpy of the saturated vapor refrigerant at a temperature of 63.17F is
h7  104.465  0.098445(63.17)  0.0001226(63.17)2  9.861  107(63.17)3
 104.465  6.219  0.489  0.249  109.946 Btu/ lb
Then, from Eq. (9.30), the fraction of vaporized refrigerant in the flash cooler is
2. From Eq. (9.32), the refrigeration effect is
qrf  (1  x)(h1  h9)  (1  0.09856)(107.72  28.095)
 71.78 Btu/ lb (167 kJ/kg)
3. From Eq. (9.22), the enthalpy differential h2  h1 corresponding to a saturated temperature differential
Ts2  Ts1  63.17  35  28.17F in the two-phase region is
h2  h1  0.18165  0.21502(28.17)  0.0012405(28.17)2  8.1982  106(28.17)3
 0.182  6.057  0.984  0.183  5.074 Btu/ lb
Similarly, from Eq. (9.22), the enthalpy differential h4  h3 corresponding to a saturated temperature
differential Ts  Ti  95  63.17  31.83F is
h4  h3  0.020  0.16352(31.83)  0.00035106(31.83)2  1.9177  106(31.83)3
 0.020  5.205  0.356  0.062  4.891 Btu/ lb

36.162  28.095
109.946  28.095
 0.09856
x 
h5	  h8
h7  h8
pi  ?pcon pev  ?196.5  76.17  122.34 psia
9.34 CHAPTER NINE
Then, from Eq. (9.33), the total work input to the compressor is calculated as
Win  (1  x)(h2  h1)  h4  h3
 (1  0.09856)(5.074)  4.891  9.465 Btu/ lb (22.0 kJ /kg)
4. The coefficient of performance of this two-stage compound system is
5. The mass flow rate of refrigerant flowing through the condenser can be evaluated as
6. From the results in Example 9.1, the power input to the single-stage system is 330.2 hp. The
power input to the two-stage compound system is
Energy saving compared with the single-stage system is calculated as
9.14 THREE-STAGE COMPOUND SYSTEM WITH
A TWO-STAGE FLASH COOLER
To reduce the energy consumption of refrigeration systems in air conditioning, the three-stage compound
system with a two-stage flash cooler became a standard product in the 1980s. Fig. 9.9 shows
the schematic diagram and refrigeration cycle of this system.
Flow Processes
In Fig. 9.9, vapor refrigerant enters the first-stage impeller of the centrifugal compressor at a dry
saturated state, point 1. After the first-stage compression process, at point 2, it mixes with the vaporized
refrigerant coming from the low-pressure flash cooler at point 12. At point 3, the mixture
enters the second-stage impeller. After the second-stage compression process at point 4, it mixes
again with vaporized refrigerant from the high-pressure flash cooler at point 9. The mixture, at
point 5, is then compressed to the condensing pressure in the third-stage impeller. At point 6, the
hot gas enters the condenser, condenses to liquid, and subcools to a temperature below the condensing
temperature. Subcooled liquid refrigerant leaves the condenser at point 7	 and flows through a
two-stage flash cooler and the associated throttling devices, in which a small portion of liquid refrigerant
is successively flashed into vapor at interstage pressure pi1, point 9, and interstage pressure
pi2, point 12. The liquid-vapor mixture enters the evaporator at point 14, and the remaining liquid
refrigerant evaporates into vapor completely in the evaporator.
Fraction of Refrigerant Vaporized in Flash Cooler
Based on the heat balance of the refrigerants entering and leaving the high-pressure flash cooler, as
shown in Fig. 9.10a, the fraction of liquid refrigerant vaporized in the high-pressure flash cooler
330.2  311.1
330.2
 0.058, or 5.8%
Pin 
500  200
42.41  7.58
 311.1 hp (232 kW)
m? r 
Qrc
qrf

500  12,000
71.78
 83,588 lb / h (37,916 kg / h)
COPref 
qrf
Win

71.78
9.465
 7.58
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.35
9.36
FIGURE 9.9 Three-stage compound system with a two-stage flash
cooler: (a) schematic diagram; (b) refrigeration cycle.
x1 at interstage pressure pi1 can be calculated as
h7	  (1  x1)h10  x1h9
Therefore,
(9.36)
where h7	, h9, h10  enthalpies of refrigerants at points 7	, 9, and 10, respectively, Btu / lb (kJ / kg).
In the same manner, the heat balance of the refrigerant entering and leaving the low-pressure flash
cooler, as shown in Fig. 9.10b, may be expressed as
(1  x1)h10  x2h12  (1  x1  x2)h13
where h12, h13  enthalpies of the refrigerant at points 12, and 13, respectively, Btu / lb (kJ / kg). The
fraction of liquid refrigerant vaporized in the low-pressure flash cooler x2 at an interstage pressure
pi2 can be evaluated as
(9.37) x2 
(1  x1)(h10  h13)
h12  h13
x1 
h7	  h10
h9  h10
REFRIGERANTS, REFRIGERATION CYCLES, AND REFRIGERATION SYSTEMS 9.37
FIGURE 9.10 Heat balance of refrigerants entering and leaving the high-pressure and low-pressure flash
cooler and mixing points. (a) High-pressure flash cooler; (b) low-pressure flash cooler; (c) at the mixing