Fundamentals of
Air System Design
Second Edition (I-P)
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A Fundamentals of HVAC&R Series
Self Directed Learning Course
1791 Tullie Circle NE • Atlanta, GA 30329 • www.ashrae.org
Air System Design I-P.indd 1
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Fundamentals of Air System Design
Second Edition (I-P)
Prepared by
Robert McDowall, P.Eng.
Consulting Engineer
Winnipeg, Manitoba, Canada
ASHRAE
1791 Tullie Circle NE  Atlanta, GA 30329
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Fundamentals of Air System Design (I-P), Second Edition
A Course Book for Self-Directed or Group Learning
ISBN 978-1-933742-45-8 (paperback)
SDL Number: 98036
© 1996, 2008 ASHRAE
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Karen M. Murray
Manager of Professional Development
Dear Student,
Welcome to this ASHRAE Learning Institute (ALI) self-directed or group learning course. We look forward
to working with you to help you achieve maximum results from this course.
You may take this course on a self-testing basis (no continuing education credits awarded) or on an ALImonitored basis with credits (PDHs or LUs) awarded. ALI staff will provide support and you will have
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Skill Development Exercises at the end of each chapter will gauge your comprehension of the course material. If you take this course for credit, please complete the exercises and either email copies from each
chapter to edu@ashrae.org (preferred method) or fax them to 678-539-2161. Be sure to include your student ID number with each set of exercises. (Your student ID can be the last five digits of your Social Security number or other unique five-digit number you create.) We will return answer sheets to the Skill
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When you finish all exercises, please submit the course evaluation, which is located at the back of your
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We hope your educational experience is satisfying and successful.
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AN INTERNATIONAL ORGANIZATION
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Chapter 1: Fundamentals of Air Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1-1
1.1
1.2
1.3
1.4
Static and Dynamic Compressible Fluid (Air) Laws
Friction Effects
The Friction Chart
Density and Altitude Effects
Chapter 2: Air Distribution System Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-1
2.1
2.2
2.3
2.4
2.5
2.6
Air Distribution System Overview
Air Handling Units
Ducts
Controls
Air Distribution Devices
Sound Absorbers
Chapter 3: Human Comfort and Air Distribution. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-1
3.1 Principles of Human Comfort
3.2 Principles of Space Air Distribution
3.3 Types of Air Distribution Devices
Chapter 4: Relationship of Air Systems to Load and Occupancy Demands . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-1
4.1
4.2
4.3
4.4
4.5
Operating System Selection Criteria
System Types by Heating/Cooling Equipment Type
System Type by Duct Configuration
Economizers
Outdoor Air Intake
Chapter 5: Exhaust and Ventilation Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-1
5.1 Design Considerations
5.2 Ventilation and Exhaust Systems
5.3 Energy Recovery
Chapter 6: Air Movers and Fan Technology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-1
6.1
6.2
6.3
6.4
6.5
6.6
Fan Principles
Fan Drives
Fan Selection
Fan Installation Design
Fan Controls
Effect of Variable Resistance Devices
Chapter 7: Duct System Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-1
7.1
7.2
7.3
7.4
7.5
Duct System Design Overview
Duct Materials
Duct Construction
Duct Design and Sizing
Sample Systems
Chapter 8: Codes and Standards . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-1
8.1
8.2
8.3
8.4
8.5
Building Code Requirements
ASHRAE Standard 90.1-2007
ASHRAE Standard 62.1-2007
Other Codes and Standards
Sources of Information
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Table of Contents
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Chapter 9: Air System Auxiliary Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-1
9.1
9.2
9.3
9.4
9.5
Dampers
Air Filters
Humidifiers
Duct Heaters
Duct Insulation
Chapter 10: Sound and Vibration in Air Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-1
10.1
10.2
10.3
10.4
Fundamentals of Sound
Sound and Vibration Sources
Sound Attenuation
Vibration Control
Chapter 11: Air System Start-Up and Diagnostics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-1
11.1
11.2
11.3
11.4
11.5
Introduction
Design Considerations
Air Volumetric Measurement Methods
Balancing Procedures for Air Distribution Systems
Noise and Vibration Diagnostics
Chapter 12: An Actual Duct Design Problem . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12-1
12.1
12.2
12.3
12.4
Introduction
Duct Design Procedure
The Building and System
Working Through The Problem
Skill Development Exercises Answer and Work Sheets. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Answer Sheet 1
Evaluation Form . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Final Page
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Table of Contents
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Fundamentals of Air
Flow
Contents of Chapter 1
•
•
•
•
•
•
•
1.1 Static and Dynamic Compressible Fluid (Air) Laws
1.2 Friction Effects
1.3 The Friction Chart
1.4 Density and Altitude Effects
Summary
Bibliography
Skill Development Exercises for Chapter 1
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Chapter 1
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Instructions
Read the material of Chapter 1. At the end of the chapter, complete the skill development
exercises without consulting the text.
Study Objectives of Chapter 1
After completing this chapter, you should be able to:
• Explain static pressure, velocity pressure and total pressure, and the relationship between them.
• Calculate change in volume of air with change in temperature at constant pressure.
• Calculate the approximate volume and temperature resulting from mixing airstreams.
• Sketch and explain the Psychrometric Chart parameters of temperature, moisture, relative humidity and specific volume.
• Explain duct frictional losses.
1.1 Static and Dynamic Compressible Fluid (Air) Laws
Because this course is designed to address the needs of people with varying backgrounds
and experience, it is necessary to review the fundamental principles of fluid mechanics.
Your understanding of these principles is essential to the applied system design concepts
that follow in later chapters. The concepts are presented in the context of HVAC applications, defining terms as they are used in that field.
This course makes use of three of the four basic principles of fluid mechanics:
• Fluid Statics
• The Continuity Equation
• The Energy Equation
THE DIFFERENCE BETWEEN MASS AND WEIGHT
A fundamental and often confused point must be addressed here: the difference between
mass and weight. Mass is a property of matter that is invariant with location. For example,
the mass of the astronauts remained essentially the same during their trip and landing on
the moon. However, their weight changed dramatically.
1–2
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The relationship of mass to weight is given by Newton’s Law of Motion:
Force  Mass  Acceleration
In the context of this course, the force is the weight required to hold up the matter (that is,
to keep it from falling), and the acceleration is the acceleration due to gravity. This acceleration due to gravity, g, is virtually constant at sea level on Earth at 32.2 ft/s2 . Note that the
law is stated as a proportionality, so we must insert a proportionality constant to make it an
operationally useful equation.
The symbol for the proportionality constant that has been used for generations is 1/gc. This
is unfortunate because of its similarity with the symbol for acceleration due to gravity, g.
The proportionality constant, gc , is actually used to convert the units of mass  acceleration to units of force. In the system currently used in the American HVAC industry, the
value of gc is 32.2 lbm ft/lbf  sec2. The reason for this choice is that the “weight” of a poundmass, lbm, is numerically equal to a pound-force, lbf , at sea level.
Two important rules result from this:
• Mass and weight are inherently different but related, and pound-mass (lbm) is
completely different from pound-force (lbf ).
•
The symbol gc is a units conversion factor.
As a thermodynamic property, density is the ratio of mass to volume. It has the unit of
lbm/ft3 and is denoted by the symbol, , (lower case Greek letter rho). At times, it is convenient to express the density as “weight” density, , (the lower case Greek letter gamma),
and by this we mean the force of gravity on a unit volume of mass. The conversion is
accomplished by multiplying by g (the acceleration due to gravity), and by taking into
account the need for units conversion with the constant gc.
Thus,  = g/gc . The units of  are:
 lb m  ft 
  ------ g  ----
 ft 3   s 2
lb f
------------------------------- = ------3
ft
 lb m  ft 
g c  -------  ----
 lb f   s 2
Also, numerically g/gc = 1, because both have the value 32.2.
Specific volume is the reciprocal of density, and is defined as the volume of a unit mass of
material. It is expressed in cubic feet per pound-mass: v = ft3/lbm.
1–3
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Fundamentals of Air System Design
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FLUID STATICS
Hydrostatic pressure is something we all experience in a swimming pool; recall how the
pressure on your eardrums increases as you dive deeper in the pool. The pressure is due to
two factors:
• The atmospheric air pressure on the surface of the water; and
• The weight of a column of water equal to the depth below the surface.
Imagine a column of water as shown in Figure 1-1. The weight is equal to the volume times
the weight density, W = hA. The force required to hold the fluid column plus oppose the
pressure of the atmosphere is W + pa. The force, F, is also the pressure at the point of application of force, F, times the area. In equilibrium, we write:
F = pA = pa A + W = pa A + h A
Therefore, p = pa + h.
The pressure difference between A and B is:
 p = yh
(1-1)
Note that the density is the weight density,  = (g/gc ).
Figure 1-1
1–4
Fluid Static System
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THE CONTINUITY EQUATION
The Continuity Equation expresses the idea that all mass is accounted for; none is lost or
created. Mass that enters a space also leaves the space, provided no change in the stored
amount occurs. Filling a tank, or releasing gas from a compressed gas bottle, obviously are
cases where stored mass changes. However, when air flows through a duct, or into and out
of a fan, the amount of air flowing per unit time is the same at the inlet as at the outlet. Figure 1-2 depicts such a situation.
Figure 1-2
Continuity Equation Example
The volume of air that passes through a cross-section of the duct is given as VA, where V is
the velocity and A is the area. Rationalize this by imagining that the flow rate would double
if the area doubles, or if the velocity doubles. The mass associated with a unit of volume is
the density, . Therefore, the mass flow rate is given by AV.
The conservation of mass idea states that no change in mass flow rate occurs under steady
conditions when there is no storage change. Considering two locations, 1 and 2, on the
same duct, we can write:
 AV  1 =  AV  2
Under normal conditions in a short length of duct with no heating or cooling coils, the
pressure and temperature changes are so small that the density is virtually constant. The
Continuity Equation can then be written as:
 AV  1 =  AV  2
(1-2)
1–5
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In the HVAC industry, flow rate almost invariably means volume flow rate, not mass flow
rate. Also, the most common units are cubic feet per minute (cfm), where the velocity has
been specified in feet per minute (fpm), and the areas are in square feet (ft2). The practice
is so common that practitioners use cfm as a word. For example, “How many cfm do you
supply to the room?”
Air behaves as a perfect gas and the change in density is inversely proportional to the absolute temperature. Absolute temperature is the temperature above absolute zero, which is 460°F. To convert from our normal °F to absolute °F, we add 460°F. Thus if outside air is
heated from 35°F to 75°F as it comes in through the air-conditioning system, the density,
, will change from
y 35
to
35 + 460
y 35  --------------------- = y 35  0.925
75 + 460
a 7.5% decrease in density.
Similarly, on a hot day, cooling 100°F outside air down to 55°F will increase the density by
100 + 460
------------------------ = 1.09
55 + 460
or 9%. The difference is even more significant in a cold climate. For example, suppose it is
January and the outside temperature is 30°F. The outside air is brought in over a heating
coil, and supplied at 75°F. The drop in density as the air is heated is from
y –30
to
– 30 + 460
y –30  ------------------------- = y –30  0.804
75 + 460
This means that the air will be approximately 20% less dense. Being 20% less dense, the air
occupies 20% more space at our constant pressure. Thus 1,000 cfm at 30°F when heated
to 75°F becomes 1,200 cfm at 75°F.
The mass of air stays the same. However, with rising temperature, the volume increases as
density drops. Remember, this is in the normal commercial and institutional HVAC system with very small pressure changes.
A very common process in air conditioning is mixing air streams. Suppose we have a situation where 15,000 cfm of air at 75°F from a space is being mixed with 5,000 cfm of out-
1–6
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side air at 40°F. We want to know what the resulting mixed air temperature and volume
will be. The industry practice, which works well for estimating and small temperature differences, is to assume that cfm is equivalent to mass and use Equation 1-3:
 cfm 1  T 1  +  cfm 2  T 2  =   cfm 1 + cfm 2   T 3 
(1-3)
For example, based on Equation 1-3, the resulting volume will be:
15 000 + 5 000 = 20 000cfm
 15 000  75  +  5 000  40  =   15  000 + 5 000   t m 
tm = 66.25°F, which is very close to the correct answer of 65.9°F.
For wide temperature differences, the inaccuracy can be almost eliminated by adjusting the
incoming cfm values to be the cfm at the initially calculated mixed temperature.
THE ENERGY EQUATION (FIRST LAW OF THERMODYNAMICS)
The third principle we will use in this course is the Energy Equation, which is based on the
idea that energy, like mass, is neither created nor destroyed. A major consequence of this
idea is that the forms that energy takes are interchangeable; that one form can be converted
into another. However, there is one caveat to this idea, based on the Second Law of Thermodynamics: heat cannot be completely converted into work in a cyclic process.
The units of the forms of energy are many and varied. Each author seems to have a unique
set of preferences and biases. However, because most studies of energy begin with a definition of mechanical work (force  distance), it is appropriate to say that the “fundamental”
unit of energy is force  distance: ft lbf . Other units of energy are the British thermal unit
(Btu), kilowatt hour (kWh) and horsepower hour (hph).
It is common to write energy conversion in terms of a unit of mass flowing; for example,
ft lbf /lbm . Whatever the units used, it is imperative to have all values expressed in the same
units when comparing or adding. Conversion factors are available in many texts and reference books, and you are expected to obtain and use conversion tables competently. For
example, an experimentally derived relationship is that 1 Btu is equivalent to 778 ft lbf .
Following is a brief listing and discussion of the forms of energy:
•
Work, W, results from a force applied through a distance in the direction of the force.
It is also the result of a torque applied through an angular displacement. Work can be
either internal or external. Machines such as pumps, fans and compressors do mechanical work, Wm , on the fluid. Machines such as turbines produce mechanical work, Wm,
done by the fluid to an external machine such as an electric generator. Fluid friction,
1–7
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Wf , can be considered to be work done by the fluid on the duct or obstruction in
much the same way as aircraft engines do work to overcome air drag in a flying airplane. In the HVAC industry, frictional forces exist in ductwork as the air passes down
a straight section, as it makes a turn, or as it passes through louvers or a heat exchanger.
This work results in a loss of pressure that must always be compensated for by the fan.
1–8
•
Flow work, pv, is energy supplied to the system when fluid crosses the boundary entering the system. It is also done by the system as fluid leaves. Consider the situation
where air discharges from a compressed air tank and makes room for itself in the atmosphere. The air pushes away the atmosphere. The air occupies space, and the atmosphere must make way for it by becoming a little bit higher; therefore increasing its
own potential energy. Flow work is always present even though the amount done on
the system at the inlet may be very nearly equal to the amount done by the system at
the outlet. Flow work is always given by the product of pressure and specific volume,
pv, for one pound of fluid.
•
Heat, Q, is the result of energy transfer due to a temperature difference. That heat can
be transformed into work, and that work can be dissipated into internal energy and
transferred as heat, constitute the main business of thermodynamics. Note that heat is
not stored and it is not “contained” by a fluid. Heat is thermal energy in transit. It is
defined only at the boundary of a system.
•
Internal energy, u, is often confused with heat, but they are totally different concepts.
Internal energy is associated with molecular motion, molecular bonding and other
forms of molecular activity such as spinning or rotation of the molecules. Internal
energy can have units of ft lbf /lbm or Btu/lbm . In an ideal gas and a liquid, u is directly
related to temperature. For example, a 1 Btu increase in internal energy is represented
by a 1°F rise in temperature for a pound of water.
•
Potential energy is energy that represents the work done on a mass in moving it in the
Earth’s gravitational field. For example, if a 1 lbm book is elevated 1 ft above a desk,
work in the amount of 1 ft lbf has been done on it. This work can be recovered by lowering the book and raising a mass someplace else through linkages or pulleys. Or the
force of gravity will accelerate the book if it is allowed to drop, and the potential energy
will have been converted to kinetic energy associated with the velocity. Potential energy
is always measured relative to some datum of zero elevation: PE = (mg/gc )z where z is
measured relative to some assigned datum in the system.
•
Kinetic energy results from motion. For example, an automobile traveling at 55 mph
has kinetic energy, as does a baseball thrown at 90 mph. The kinetic energy is derived
from a steady force applied through a distance required to accelerate the body from rest
to a velocity, V: KE = mV 2/2
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Here is an excellent example where units need to be converted. As it stands, KE has units of
lbm ft2/sec2. The units conversion factor can be used to convert to conventional energy
units by writing:
2
ft
--------2
2
mV
sec
KE = ------------ = lb m --------------------- = ft  lb f
2gc
lb m ft
-------  --------lb f sec 2
The Energy Equation is simply a balance of these forms of energy. It is assumed that the
system is in steady state. If another form was found to be important (such as chemical
energy in combustion), it could be added to the list. If one or more of the forms is not present or important, it can be dropped. If we include those discussed above, the Energy Equation is written:
Wm–W
f
2
g
V+ Q + m u + pv + ----z + ------gc 2gc
in
2
g
V= m u + pv + ----z + ------gc
2gc
(1-4)
out
In air systems, the mass is that of flowing air, heat is added (or removed at a specified rate)
and work is done at a certain rate, such as a 10 hp motor driving a fan. The Energy Equation can be turned into a Rate Equation by considering:
• Heat as a rate, Btu/h, kW or hp; Q·
•
•
·
Work as a rate, Btu/h, kW or hp; W
Mass as a mass flow rate, or lbm /h; m·
A dot over the symbol is commonly used to indicate a rate. Although various units for the
energy terms have been suggested above, the units for all terms in the equation must be the
same:
·
·
W m –W
f
2
g
V
+ Q· + m· u + pv + ----z + -------2gc
gc
in
2
g
V
= m· u + pv + ----z + -------2gc
gc
(1-5a)
out
1–9
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If this form of the energy equation is divided through by the mass flow rate, m· , the following form results where heat and work are on a unit mass flowing basis:
2
g
V
W m – W f + q + u + pv + ----z + -------gc
2gc
in
2
g
V
= u + pv + ----z + -------gc
2gc
(1-5b)
out
The terms of the Energy Equation are depicted in Figure 1-3.
Figure 1-3
Energy Equation Applied to a Flow System
STATIC PRESSURE, VELOCITY PRESSURE AND TOTAL PRESSURE
Now let's discuss a run of ductwork where the following conditions exist:
• No machines, so all work terms are zero;
• No heat transfer because the duct air is the same temperature as the room air;
• No significant changes in elevation, so z is constant; and
• The internal energy, u, is essentially constant.
In this case, we have the simpler form of the Energy Equation:
2
V pv + ------2gc
1–10
2
=
in
pv + V
-------2gc
(1-6a)
out
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The above circumstances exist for a pitot tube as shown in Figure 1-4 where the flow comes
to zero velocity (where the arrow indicates total pressure in direction of flow). The Energy
Equation becomes:
2
V
pv + -------2gc
=  pv  total
(1-6b)
duct or static
Suppose further that the specific volume is constant because of the small pressure changes
involved, and that we change the specific volume to the mass density using v = 1/, multiply through by gc /g, and replace (g/gc) by the “weight” density, . Equation 1-6b then
becomes:
p V2
--- + ------- 2g
duct
p
= --
total
(1-6c)
The location “duct” could be anyplace, and we can say that the “total” duct pressure is constant in the absence of friction and significant heat transfer. The following is known as Bernoulli’s Equation:
p V2
--- + ------ 2g
Figure 1-4
= constant
duct
(1-6d)
Static and Total Pressure
1–11
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Among other things, Bernoulli’s Equation says that as the velocity goes up or down (perhaps due to area changes or takeoff air), the static pressure changes. Note that the units for
Equation 1-6d as written are feet. These are pressure equivalents to the weight of a column
of the fluid on a unit area. Thus the units are feet of air, or feet of water, depending on the
fluid actually flowing and not the instrument that is used for measuring.
Returning to Equation 1-6c, and multiplying through by the weight density, g, we define
the velocity pressure and obtain:
2
V
p +  ------2g
duct
= p total
or
p static + p velocity = p total
(1-6e)
Examining the units of the velocity pressure term, we find that:
 
  2
2
lb f
lb
ft  sec 
V
 ------- =  ------f-   -------------------  = ------ 3
2
2
2g
ft
 ft  ft  sec 
 
2
The unit lbf /ft2 can be converted to pounds per square inch (psi), or inches water gauge
(in. wg.). Here again, the density is for the fluid flowing.
Note that the relationship between velocity and velocity pressure can be used both ways, to
find pressure or velocity. Two equations commonly used in practice are the following:
pv
V = 1096.7 ------- air
V
p v = -----------4005
1–12
2
(1-7)
(1-8)
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where the numbers 1096.7 and 4005 contain the conversion factors appropriate for pv in
inches of water; density, , in lbf /ft3 (standard air density is 0.075 lbm /ft3); and velocity,
V, in fpm.
Standard Air, for the HVAC industry, is dry air at 70°F and 14.969 psia with a mass density of 0.075 lbm/ft3. Sea level pressure is 14.969 psia, so Standard Air can be considered as
typical dry air at sea level. For this reason, most airflow tables and charts are based on Standard Air. Note that defining an air flow in terms of Standard Air also defines the weight
and mass flow. Thus, 1,000 ft3/min of Standard Air is also 1,000  0.075 = 75 lbm/min.
As elevation increases, air density decreases and above 3,000 feet, density corrections
should be considered. Because most projects are located at altitudes from sea level to 3,000
feet, most designs can use Standard Air without correction. Air expands as it is heated and
the density drops. For many air-conditioning systems, this can be ignored, but be careful.
In a cold climate, outside air at –30°F has a density about 20% lower at 75°F.
Standard Air is dry air with no moisture vapor. But the air we experience is never dry.
Atmospheric air always includes water in the form of moisture vapor. Also, the quantity of
moisture vapor varies. It is typically under 2% by weight, and it influences the density and
thermal properties of air. The addition and removal of moisture are common processes in
air systems and can be conveniently shown on a chart called the Psychrometric Chart. The
main axes on the chart are temperature along the bottom x-axis and moisture weight compared to dry air weight, lb/lb, on the y-axis.
There is a maximum proportion of moisture vapor with the air at any given temperature,
so the chart has the characteristic form of Figure 1-5. Shown are:
• Vertical temperature lines, °F
• Horizontal moisture content (humidity ratio) lines, lb of moisture/lb of dry air
•
•
Sloping down left to right specific volume lines, ft3/lb. For example, air at 90°F
and 25% relative humidity has a specific volume of 14.0 ft3/lb. At this specific
volume, 1 lb of air occupies 14 ft3.
Curved relative humidity lines, %. The highest of these lines, labeled 100% rh,
is the maximum moisture that can be in gaseous form at that temperature.
When the air is saturated with moisture, we say the humidity is 100%. When the same volume of air holds only half the weight of water vapor that it has capacity to hold at that temperature, we call it 50% relative humidity, or 50% rh. The chart shows the 25%, 50%,
75% and 100% relative humidity lines. The saturation line is 100% and 0% is the horizontal line along the x-axis. Note that on the chart, the relative humidity lines are not linearly related. Thus at a particular temperature, the 50% relative humidity curve is not at
half the height of the saturation, 100% humidity, line.
1–13
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Figure 1-5
Psychrometric Chart
Psychrometric charts are based on Standard Air, and humidity ratio may be labeled lb, lbm or lbw.
Because lbm and lbw are numerically the same at the same pressure, all the charts are graphically the
same. For most above-ground terrestrial systems, the lbm and lbw issue can be ignored. But be careful with units when dealing with substantial pressure changes as occur in mines, submarines, planes
and space vehicles.
We will return to the Psychrometric Chart in future chapters.
1–14
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AIR HANDLING – A PRACTICAL APPLICATION
How these basic principles apply to air system design is illustrated in Figure 1-6, which
shows a duct with the air coming in the left and going out the right. For this example, we
assume this to be a frictionless process. Notice that the duct reduces in cross-section, with
area A1 greater than A2. There will be one velocity at A1 and another at A2. This process
can be analyzed using the Continuity Equation.
The Continuity Equation says that for a given mass flow, and by the law of conservation of
matter or mass, whatever air we put in on the left side has to come out on the right side
because we can neither destroy nor create air in the duct between the two points. The Continuity Equation says that the cfm in is equal to the cfm out, ignoring any kind of compressibility or temperature change. In other words, the quantity of air in (cfm1 ) is equal to
the quantity of air out (cfm2 ), giving cfm1 = cfm2. Because cfm = AV, then A1V1 = A2V2. If
we measure the duct, we know what A1 and A2 are. If we know V1 , we can solve for V2 ,
and we know that, because A1 is bigger than A2 , V2 must be bigger than V1.
This relationship can be explained by the Continuity Equation:
A 
V 2 = V 1  -----1- 
 A2 
(1-9)
So as the cross-sectional area is reduced, the velocity is increased as predicted by the Continuity Equation.
Figure 1-6
Conversion of Static Pressure to Velocity Pressure
1–15
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Let’s return to the Energy Equation and the relationship that the total pressure is
equal to the velocity pressure plus the static pressure (pt = pv + ps ). If the velocity
increased, the velocity pressure had to increase, because velocity pressure is pv =
V 2/2g. As the air flows from left to right in Figure 1-6, both velocity and kinetic
energy increase.
The simple device shown in Figure 1-6 converts potential energy into kinetic
energy. But how did this happen? In this example, there are two forms of energy:
static pressure (the flow work) and velocity pressure (kinetic energy). If the kinetic
energy increases, then the flow work must decrease in direct proportion.
Consequently, if A2 is one-half as big as A1, then V2 is twice as big as V1. Because
the velocity pressure is proportional to the square of the velocity (V 2/2g), pV2 is
four times pV1 and the static pressure ps is smaller by an equal amount. This is not
too difficult to understand because it is expected that the static pressure will be less
at A2 than A1.
Figure 1-6 shows an accelerator, where air velocity is increased by making the duct
area smaller. Suppose Figure 1-6 is reversed, as in Figure 1-7 which shows a decelerator. The air comes in at a higher velocity than it goes out. Because the air comes
in at a higher velocity through the smaller section, and goes out at a lower velocity
through the larger section, the kinetic energy is reduced. If the velocity is reduced
by a factor of two, the kinetic energy level (and the velocity pressure) is reduced by
four. Consequently, the static pressure increases by an equal amount. Static pressure probe manometers at A1 and at A2 in Figure 1-7 would show that the static
pressure at A2 is greater than at A1.
This phenomenon is called static pressure regain, and it is a very important principle of air system design. One method of designing ducts is called the static pressure
regain method, which is applied to a duct with a series of outlets. After each outlet,
the velocity is reduced and the duct size is reduced so that the static pressure at the
next outlet will be the same.
Figure 1-7
1–16
A Decelerator
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1.2 Friction Effects
Until now, we have considered frictionless systems. But in the real world of air system
design, friction must be taken into account. Viscosity is the property responsible for dissipation of the fluid’s kinetic energy into intrinsic internal energy. In air ducts, the amount
of energy transferred is small, but the effect on pressure drop is major. The frictional pressure drop is commonly characterized by the Darcy-Weisbach Equation:
2
L V
 p f = f ----  ------D 2g
in lbf /ft2
(1-10)
Or in terms of head as feet of fluid flowing, the Darcy-Weisbach Equation can be written
as:
2
pf
L V
---------- = f ---- -------- in feet of fluid flowing

D 2g
(1-11)
For Standard Air, the Darcy-Weisbach Equation can also be written as:
L V
 p f = f ---- -----------D 4005
2
in in. wg
(1-12)
This is a purely empirical formula which states that the frictional pressure drop is proportional to length, L, inversely proportional to diameter, D, and proportional to velocity
pressure or velocity head. One would hope that the proportionality constant, f (a dimensionless constant called the friction factor), would be truly constant, and that turns out to
be partially true. When the flow is fast, f is fairly constant and depends only on the duct
roughness. When the flow is slow, f is inversely proportional to velocity, but the wall
roughness is unimportant. The terms fast and slow must be explained.
Consider all the properties and characteristics involved in fluid friction: velocity, diameter,
viscosity and density. Consider also the variety of motions that we observe: slow, such as
the streamline flow of water out of a hose; or fast, such as the turbulent flow of water flowing out of the same hose when the faucet is fully open. We are fortunate that these phenomena can all be related through a single parameter known as the Reynolds number,
which is defined as:
Re = VD
-----------
(1-13)
where µ = absolute viscosity, lbm/ft sec; V = velocity, fps; D = diameter, ft; and  = density,
lbm /ft3 .
1–17
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The Reynolds number is the ratio of the momentum of the flow (V) to the viscosity, µ. If
the viscosity is high relative to the momentum, the flow is laminar or streamline (like
maple syrup). But if the viscosity is low (as for air), the flow will be turbulent for any realistic duct size. Laminar air flow occurs in laminar flow filters where the pore size, D, is very
small.
So there are two distinct regimes of flow (laminar and turbulent) that depend on the Reynolds number. The effect of these distinctions is manifested in the behavior of the friction
factor as shown on the Moody Chart (Figure 1-8). This chart shows the friction factor as a
function of the Reynolds number. Note that both axes have logarithmic scales.
Figure 1-8
1–18
Moody Chart
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Several interesting features are present on the Moody Chart:
•
The laminar flow region is shown for Reynolds numbers smaller than about
2,000. The dashed extension of the solid line indicates that, under some circumstances, the relationship can be extended up to 4,000. This part of the line
is not to be trusted. In the laminar flow region, the friction factor is inversely
proportional to the Reynolds number:
f = 64  Re
•
If this value is substituted into Equation 1-12:
L
V
 p f = 64v ------ -------------2
2
D 4005
•
•
(1-14)
in in.wg
(1-15)
where , the kinematic viscosity, is / with units of ft2/sec. Note that the frictional pressure drop varies with the first power of velocity.
There is also a dashed line labeled “fully rough.” To the right of this line, the
friction factor is constant for a particular value of roughness, . The relative
roughness values are shown. For example, the roughness of commercial steel
pipe is 0.00015 in. Relative to a 4 in. pipe, /D is about 0.00045. In this “fully
rough” region, a constant value of f can be used regardless of flow rate or velocity, and the frictional pressure drop varies with the second power of velocity.
Between laminar and fully turbulent flow, the friction factor depends on the
Reynolds number and the relative roughness, and an iterative solution to a
problem may be necessary. In this region, pressure drop varies with a power of
velocity between 1 and 2. Unfortunately, many air duct flows occur in this
transition region.
Test work performed by ASHRAE and its predecessor organization ASHVE (American
Society of Heating and Ventilating Engineers) indicated prior to Moody’s work1 that  for
galvanized sheet metal ductwork was about 0.0005 ft. This is based on transverse joints
spaced at 30 in. intervals. When joints are spaced at 46 in. intervals, the value is reduced to
0.0003 ft.
1–19
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THE SYSTEM CONSTANT FORM OF THE DARCY-WEISBACH EQUATION
HVAC air and piping systems usually use a simplified form of the Darcy-Weisbach Equation where it is assumed that the friction factor is constant and that L and D do not change
(although the system may be made up of various L and D and fittings). So we lump all of
the constants together and write two forms that are essentially the same – the second being
an inversion of the first – with two constants, K and Cs :
 p f = K  cfm 
2
(1-16)
or
cfm = C s  p f
(1-17)
Equation 1-16 is the system constant form of the Darcy-Weisbach Equation. It is used
extensively in HVAC systems work. A system curve as shown in Figure 1-9 portrays the frictional pressure drop for a particular system. The curve is a parabola that can be generated
with one known experimental or calculated value for a particular system. One pair of cfm
and pf is required to determine K or Cs .
Figure 1-9
1–20
Typical System Curve
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Consider a complex air handling system where we want to move 10,000 cfm through the
system. The pressure drop in the system is calculated to be 4 in. wg. The system constant
form of the Darcy-Weisbach Equation can be used to find the system constant:
 000
C s = 10
------------------ = 5 000
4
Similarly, we find that K = 4  10-8. Values of pf and cfm can be plotted on a graph.
Other values can be determined by using:
 p f = K  cfm 
2
As long as the system is unchanged, it will operate on this curve.
1.3 The Friction Chart
In 1945, D.K. Wright published “A New Friction Chart For Round Ductwork” in the
ASHVE Transactions.2 A graph from this article has become essentially the standard for
HVAC work. This graph, often known as the Wright Friction Factor Chart, takes the
Darcy-Weisbach relationship and the Moody Chart and converts them into a graphical
presentation that lets us determine frictional pressure drops at various diameters of round
ductwork and at various velocities based on an  value for galvanized sheet metal ductwork
of 0.0005 ft .
Since that time, ASHRAE and the Sheet Metal and Air Conditioning National Contractors' Association (SMACNA) have conducted a series of tests and obtained slightly different numbers than those used by Wright. The new data have been included in the ASHRAE
Handbook–Fundamentals since 1993. The friction factor chart (see Figure 1-10) was
revised based on standard galvanized sheet metal ductwork with an absolute  roughness of
0.0003 ft instead of 0.0005 ft.
Other factors, including the shape of the duct, the roughness of the material of construction, and fittings used must be taken into consideration. These will be discussed later in
Chapter 7.
1–21
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1–22
Friction Factor Chart
Figure 1-10
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1.4 Density and Altitude Effects
Standard psychrometric charts and performance data published by manufacturers generally assume equipment operation at sea level with Standard Air. However, when the project is located at a significantly higher altitude, allowances must be made for the lower pressure. Factors by which the usual data must be multiplied when operating at higher
altitudes are summarized in Table 1-1. For items not listed, consult appropriate sources,
such as Carrier’s Engineering Guide for Altitude Effects.3
Table 1-1 Typical Altitude Correction Factors4
Item
Compressors
Condensers, air-cooled
Condensers, evaporative
Chillers
Induction room terminals (chilled water)
Fan-coil units
Total capacity (*SHF = .40-.95)
Sensible capacity (SHF = .40-.95)
Total capacity (SHF = .95-1.00)
Packaged air-conditioning units, air-cooled condenser
Total capacity (*SHF = .40-.95)
Sensible capacity (SHF = .40-.95)
Total capacity (SHF = .95-1.00)
Altitude (ft above sea level)
2500
1.00
0.95
1.00
1.00
0.93
5000
1.00
0.90
1.01
1.00
0.86
7000
1.00
0.85
1.02
1.00
0.80
10,000
1.00
0.80
1.03
1.00
0.74
0.97
0.92
0.93
0.95
0.85
0.86
0.93
0.78
0.79
0.91
0.71
0.73
0.98
0.92
0.96
0.96
0.85
0.82
0.94
0.78
0.88
0.92
0.71
0.84
*SHF = Sensible Heat Factor
The Next Step
This chapter has introduced the theory needed and included some discussion about air
flowing in ducts. Chapter 2 will introduce the other common components of air systems
that condition air and deliver it to the occupied space. Included will be their function and
main operating characteristics. More detailed issues of choosing components and their
detailed operation will be explained in later chapters.
1–23
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Summary
The chapter began by explaining the difference between mass and weight. Mass is a property of matter that is invariant with location, but weight changes depending on the local
gravitation. Conveniently, for most building designers, gravity is constant, with lbm and
lbw being numerically the same.
Hydrostatic pressure, commonly referred to as static pressure, is the pressure exerted by a
fluid at rest. The pressure is the same in all directions at any point. In a duct with the air
flowing under pressure, the static pressure around the duct will be the same on all sides.
The Continuity Equation states that mass is neither created nor destroyed. Thus, under
steady conditions with no storage, the mass flow into a system must equal the mass flow
out of the system. The volume of air at constant pressure is proportional to the absolute
temperature. Thus, while the mass into a system equals the mass out, the volume in can be
different from the volume out if the temperature is changed.
The useful, but not absolutely correct, formula for calculating the result of mixing airstreams was introduced: (cfm1  T1) + (cfm2  T2) = [(cfm1 + cfm2)  T3]
Energy, like mass, is neither created nor destroyed. It can be converted from one form to
another and measured in different units. However, in consistent units in any process:
energy in = energy out – energy stored in the system
Energy can be in a number of forms: work, done by a force over distance or torque through
an angle; heat, energy transfer due to a temperature difference; internal energy, due to thermal energy relative to some datum; potential energy, work done by movement in the
earth’s gravitational field; and kinetic energy from motion.
Static pressure is the pressure exerted by a fluid at rest. Velocity pressure is the pressure
exerted by a fluid by virtue of its motion. Typically, measuring the pressure at a tapping in
the side of a duct provides the static pressure. The pressure on the open end of a tube facing the flow of air measures both the static and the velocity pressure, called total pressure.
The difference between the static pressure and total pressure is the velocity pressure.
Bernoulli’s Equation states that in a system without energy losses or gains, the sum of static
pressure and velocity pressure are constant:
p V2
---- + ------- 2g
1–24
= constant
duct
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The valuable concept in Bernoulli’s Equation is that if the velocity is reduced due to a
wider duct, the drop in velocity pressure (V 2 reduces) is exactly matched by an increase in
static pressure, friction ignored.
Velocity pressure in inches water gage for Standard Air equals:
2
V
-----------4005
in. wg
Standard Air and the psychrometric chart were introduced to raise the issue of decreasing
density with increasing temperature and the issue of moisture in the air.
Friction effects occur in ducts for several reasons including surface roughness, duct joints,
fittings, equipment and outlets. The theory behind duct friction was discussed including
Reynolds Number, Moody Chart and Darcy-Weisbach Equation. The critical point to
remember is that in a fixed system, the pressure drop through the system will be about pro2
portional to the square of the flow:  p f = K  cfm  Thus, doubling the flow will create
four times the pressure drop.
The pressure drops through ducts can be calculated, but the simplest method is to use a
Friction Chart such as the ASHRAE chart shown in Figure 1-9 for Standard Air.
Standard psychrometric charts and performance data generally assume equipment operation with Standard Air. The lower air density at high altitudes significantly affects some
equipment but not all. Reference tables can be used and manufacturers contacted for assistance in these cases.
Bibliography
1. Moody, L. 1944. "Friction factors for pipe flow." ASME Transactions. New York, NY:
American Society of Mechanical Engineers.
2. Wright, D. 1945. "A new friction chart for round ductwork." ASHVE Transactions.
Atlanta, GA: ASHRAE.
3. Carrier Corp. Engineering Guide for Altitude Effects.
ASHRAE produces four Handbooks: Fundamentals; HVAC Systems and Equipment;
HVAC Applications; and Refrigeration. Each Handbook is updated and reissued on a fouryear cycle. Handbook sections that relate to the material are listed in the bibliography for
each chapter. For this chapter, see HandbookFundamentals for general theory fluid flow
and duct sizing.
1–25
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Skill Development Exercises for Chapter 1
Complete these questions by writing your answers on the sheets at the back of this book.
1–26
1-1.
In the figure below, Area A1 = 2 ft2, Area A2 = 1.25 ft2, and velocity V1 = 1,000
fpm. Calculate V2 (fpm).
a) 1,600 fpm b) 625 fpm c) 1,406 fpm d) 2,569 fpm
1-2.
The total pressure at a certain point in a system is determined to be 5 in. wg, and
the static pressure at that point is determined to be 2 in. wg. What is the velocity
pressure (in. wg) at that point?
a) 21 in. wg b) 7 in. wg c) 3 in. wg d) 2 in. wg
1-3.
Which of the following is the most correct definition of static pressure regain?
a) As the velocity of an airstream decreases, the static pressure increases.
b) As the velocity of an enclosed airstream decreases due to friction, the
static pressure increases.
c) Friction reduces static pressure while velocity pressure increases with
reduction in duct size.
1-4.
An air handling system is determined to have a 6 in. pressure drop through the
system at a flow of 8,000 cfm. What is the system constant?
a) 1.5 b) 1,333 c) 3,265 d) 4,000
1-5.
The product of fluid pressure and specific volume is ______?
a) Internal energy b) Reynolds number c) Kinetic energy
d) Flow work e) Viscosity
1-6.
What does a water manometer measure?
a) Velocity b) Pressure c) Temperature d) All of the above
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1-7.
Fan pressures are typically indicated in what units?
a) in. wg b) in. Hg c) cfm d) None of the above
1-8.
If the cross-sectional area of a duct decreases in size, the velocity of an airstream
passing through the duct will increase.
a) True b) Cannot tell c) False
1-9.
Air is passing through a length of inaccessible duct with a constant cross-sectional area. You suspect that there is a serious leak in the duct. The velocity pressure drops from 0.85 in. wg to 0.60 in. wg along the suspect section of duct.
Approximately what percentage of air is being lost through the leak?
a) 50% b) 31% c) 16% d) 11%
1-10.
In an air-conditioning system, 3000 cfm of outside air at 34°F is drawn in over
a heater and delivered into the building at 74°F. What volume of air is delivered?
a) 6,529 cfm b) 1,378 cfm c) 2,775 cfm d) 3,243 cfm
1-11.
In an air-conditioning system, 30,000 cfm of return air at 78°F is mixing with
4,600 cfm of outside air at 95°F. What is the approximate resulting volume and
temperature?
a) 36,400 cfm, 79.2°F b) 36,400 cfm, 80.3°F
c) 34,600 cfm, 80.3°F d) 34,600 cfm, 79.2°F
1–27
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Chapter 2
Air Distribution
System Components
Contents of Chapter 2
•
•
•
•
•
•
•
•
•
2.1 Air Distribution System Overview
2.2 Air Handling Units
2.3 Ducts
2.4 Controls
2.5 Air Distribution Devices
2.6 Sound Absorbers
Summary
Bibliography
Skill Development Exercises for Chapter 2
2–1
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Instructions
Read the material of Chapter 2. At the end of the chapter, complete the skill development
exercises without consulting the text.
Study Objectives of Chapter 2
The goal of this chapter is to give you an overview of air distribution system components
and their schematic symbols, to serve as a foundation of knowledge as each component is
discussed in detail in later chapters. After completing this chapter, you should be able to:
• List and explain the functions of the components of an air distribution system.
• Identify the schematic symbols of air distribution system components.
2.1 Air Distribution System Overview
An air-distribution system is used to maintain desired environmental conditions within a
space. In almost every application, many options are available to the designer to satisfy that
goal. Air distribution systems are categorized in many ways including: by how they control
the conditioned area; by special equipment arrangement; and by duct configuration.
This chapter will provide an overview of the basic components of an air distribution system: air handling units; fans, fan motors and fan drives; coils; filters; ducts; controls; air
distribution devices; intake and exhaust louvers; and sound absorbers.
2.2 Air Handling Units
An air-handling unit (AHU) combines fans, coils, filters, dampers, connections to supply
and return ducts, and other components into a device that moves air. It may also be used to
clean, heat, cool, humidify, dehumidify and mix the air. Figure 2-1 shows a large typical
central air-handling unit.Types of air-handling units include:
•
•
•
•
•
2–2
A central-station unit is a factory-made, encased assembly consisting of the fan
and other necessary equipment. It does not include a source of heating or cooling, but it may include heating and/or cooling coils.
A cooling unit that includes the means for cooling. It may also perform other
AHU functions.
A heating unit that includes the means for heating. It may also perform other
AHU functions.
A makeup air unit is a factory-assembled fan heater, or cooler, used to supply
tempered fresh air to replace the air that is exhausted.
A ventilating unit has the means to provide ventilation, and may also perform
other AHU functions.
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Figure 2-1
Air Handling Unit
FANS, MOTORS AND FAN DRIVES
A fan is an air pump that creates a pressure difference and causes air flow. The fan impeller
does work on the air, imparting to it both static and kinetic energy, varying in proportion
depending on the fan type.
Fans are generally classified as centrifugal fans or axial flow fans according to the direction
of air flow through the impeller. Figure 2-2 shows the general configuration and schematic
symbol for a centrifugal fan. Figure 2-3 shows the configuration and schematic symbol for
an axial flow fan.
All fans must have some type of power source, usually an electric motor. On packaged fans,
the motor is furnished and mounted by the manufacturer. On larger units, the motor is
mounted separately and coupled directly to the fan or indirectly by a drive mechanism.
The schematic symbol for a motor is also shown in Figure 2-3.
2–3
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2–4
Figure 2-2
Centrifugal Fan Configuration and Symbol
Figure 2-3
Axial Fan Configuration and Symbol
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Two standard fan drive arrangements are available:
• Direct drive, where the fan is mounted directly on the motor shaft or an extension of the motor shaft, offers a more compact assembly and ensures constant
fan speed. Fan speeds used to be limited to available motor speeds, an economical solution when practical. Today, at additional cost, the motor speed can be
adjusted over a wide range by supplying the motor through a variable frequency controller. Capacity is set during construction by variations in fan
impeller geometry and motor speed.
• Belt drive offers flexibility in that the fan speed can be changed by altering the
drive ratio. This allows initial adjustments to match the fan output with the
system actually installed. In some applications, this flexibility allows for
changes in system capacity or pressure requirements due to changes in process,
hood design, equipment location or air cleaning equipment.
COILS
A coil is a cooling or heating element made of pipe or tube. Coils are usually finned, and
are found in a number of shapes (serpentine, helical, etc.). Some coils commonly encountered in air systems include:
• A cooling coil uses refrigerant or secondary coolant to provide cooling, or cooling with dehumidification.
• A heating coil provides heat. Electric heating coils use a resistance element
instead of a fluid to create a heating effect.
• A preheat coil is a heating coil installed upstream of a cooling coil, or at the
inlet end of an air handling system, to preheat air.
• A reheat coil is a heating coil installed downstream of a cooling coil.
Cooling and heating coils are often seen as labeled boxes, as shown in Figure 2-1.
FILTERS
A filter is a device used to remove solids from an airstream. Filter performance is based on
the ability to collect a particular size, or type, of dust and is stated for each filter as a rating.
The rating may denote air cleaning efficiency as a percentage of dust removal or as the ability to remove dust particles of certain size ranges. These efficiencies are defined by standardized ASHRAE test methods that we will discuss in Chapter 9.
A filter used to remove gases is correctly called an adsorber, as the gas is chemically
adsorbed onto the filter material rather than mechanically collected on the filter surface.
Filters encountered in air system design include:
• A disposable filter has elements that are discarded after use. Efficiencies range
from very low to relatively high depending on the construction.
2–5
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•
A pleated filter provides a high ratio of media area to face area, thus allowing
reasonable pressure drop. The filter media may be self-supporting because of
inherent rigidity, or because the air flow inflates it into an extended form, such
as with bag filters.
•
A roll filter (moving curtain filter) has a filter medium on a continuous belt on
movable rolls that brings a clean filter area into the airstream, either automatically or manually. Efficiencies are usually fairly low.
•
A viscous impingement filter has a medium made from materials that have
been impregnated with a viscous oil to increase dust retention.
•
An absolute filter has an efficiency of 99.9% or higher, and can filter particles
down to 0.01 micrometers (microns) in size. A particular type, the High Efficiency Particulate Air (HEPA) filter, is tested and rated to an ASTM standard
to remove at least 99.97% of particles 0.3 microns in diameter.
•
An electrostatic filter (active) has the airstream passing through a high-voltage
ionizing field to impart a positive electrical charge to the particles, which are
then collected on electrically negative plates.
•
An electrostatic filter (static) consists of plastic media that generate an electrostatic attraction by the air flow over the plastic.
•
A carbon filter (adsorber) uses a mass of granulated activated carbon to chemically adsorb certain gases.
Labeled boxes are often used to indicate filters in diagrams (see Figure 2-1). Filters are discussed in more detail in Chapter 9 of this course.
2.3 Ducts
A duct is a tube or conduit for conveying air. Ducts are classified in terms of application
and pressure. HVAC systems in public assembly, business, educational, general factory
and mercantile buildings are usually designed as commercial systems. Air pollution control
systems, industrial exhaust systems and systems outside the pressure range of commercial
system standards are classified as industrial systems.
Ducts may be round, oval or rectangular. They may be made of galvanized steel, aluminum, fibrous glass and other materials. They may be rigid or flexible. Schematic symbols
for ducts are shown in Figure 2-4.
2–6
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Duct Symbols
Figure 2-4
2–7
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2.4 Controls
A control is a device that regulates a variable such as temperature, velocity or pressure.
Controls may be manual or automatic. For example, an air handling unit might have a
manually set minimum flow of outside air and an automatic control to increase the outside
air as the building becomes more densely occupied. If automatic, the implication is that
the control is responsive to a measured change in pressure, temperature or some other variable to be regulated. Two common and important controls are dampers and thermostats.
A damper is a device used to vary the volume of air passing through an outlet, inlet or duct.
A thermostat is an automatic device that is responsive to temperature. Thermostats are
used to maintain a constant temperature in a regulated space, permit the passage of control
air when the temperature of the controlled air is within the limits at which the thermostat
is set, and other temperature control purposes. Schematic symbols for these controls are
shown in Figure 2-5.
Figure 2-5
2–8
Controls Symbols
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2.5 Air Distribution Devices
Air distribution devices are devices or openings through which air is discharged into a conditioned space. Included in this category of devices are registers, grilles and diffusers. Registers and grilles are also used to withdraw air from a conditioned space. Schematic symbols
for these devices are shown in Figure 2-6 and they are fully discussed in the next chapter
.
Figure 2-6
Air Distribution Devices
2–9
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INTAKE AND EXHAUST LOUVERS
A louver is a device consisting of multiple blades that, when mounted in an opening, permits the flow of air, but inhibits the entrance of other elements. An intake louver is used at
the entrance to an air system. An exhaust or relief louver is used at an exit. Schematic symbols for louvers are shown in Figure 2-7.
Figure 2-7
Louvers
2.6 Sound Absorbers
A proper acoustical environment is as important for human comfort as other environmental factors controlled by air-conditioning systems. The objective of sound control is to
achieve an appropriate sound level for all activities and people involved. Sound absorbers
diminish the intensity of sound energy from fans, ducts and other sources. Chapter 10 in
this course provides additional information on acoustical environments.
Sound and vibration isolation are required for most central system fan installations.
Mountings of fiberglass, ribbed rubber, neoprene and springs are available for most fans
and prefabricated units.
Noise transmitted
through ductwork can
be reduced by soundabsorbing units,
acoustical linings and
other means. The
schematic symbol for a
sound absorber in
ductwork is shown in
Figure 2-8.
2–10
Figure 2-8
Sound Absorber
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The Next Step
The primary task of commercial and institutional HVAC systems is to keep the building
occupants comfortable. To achieve this, the system designer requires knowledge of the factors affecting comfort and how air can be distributed in occupied spaces to achieve comfort
conditions. This is the subject of Chapter 3.
Summary
This chapter has briefly introduced the main components of an air-conditioning system.
More details of their construction and operation are included in later chapters.
Air-handling units (AHU) are a combination of fans, coils, filters, controls, louvers and
dampers, which together provide a supply of conditioned air. Depending on the particular
requirements, the air may be filtered, mixed, cooled, dehumidified, heated or humidified,
and the fans provide the necessary static pressure and velocity to the air flow.
A fan is an air pump. The fan creates a pressure difference (static pressure) and causes air
flow (kinetic energy). The first main fan type is the centrifugal fan where the air enters the
center of the drum-shaped impeller and is thrown radially into the fan outlet casing. The
second main fan type is the axial fan where the air flows axially, or parallel, to the fan shaft.
Most fans are driven by an electric motor. The simplest arrangement is mounting the
impeller directly on an extended motor shaft. This arrangement works for smaller sizes but
is limited to the few available motor speeds. Belt drives are a popular mechanical method
of connecting the fan and impeller shaft and they can be adjusted to change speeds. In
addition, electrical speed controllers are available to provide variable speed drive.
A coil is an array of finned pipes containing a flow of cooling or heating fluid. The fins
greatly extend the heat transfer area of the pipes. Coils used for cooling are often cool
enough for condensation to occur, thereby dehumidifying the air.
Filters remove dirt from an airstream. Their performance is rated on the basis of particle
removal based on quantity or particle size. Filters are available in a large range of designs,
each aimed at a specific market segment. This will be discussed in more detail in Chapter
9. Units that remove gases are called adsorbers, although the most common type made of
activated carbon granules is called a carbon filter.
A duct is a tube or conduit for conveying air. Ducts are most commonly made of light galvanized steel with round or rectangular sections. They can be made in many other materials for particular duties. The main criteria for choosing ducts are pressure and contaminants.
2–11
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Controls regulate the performance of a system. Manual controls are preset, such as a
damper preset to restrict flow through a duct. Automatic controls regulate some functions
continuously, such as a thermostat controlling a heater.
Some air distribution devices distribute air into occupied spaces while others allow air out
of the spaces. A louver allows air into, or out of, the building while restricting the entrance
of unwanted rain, snow, animals and birds.
A mechanical plant is inherently noisy. The noise can be distributed either by direct transfer into the building structure or as airborne noise along the ducts. A variety of materials
are used to isolate the vibration and to attenuate the noise distributed through the ductwork. Chapter 10 goes into more detail.
Bibliography
ASHRAE Handbooks:
HandbookFundamentals contains information on air contaminants and odors
HandbookHVAC Systems and Equipment contains detailed information on HVAC components
2–12
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Skill Development Exercises for Chapter 2
Complete these questions by marking your answers on the worksheets at the back of this book.
2-1.
This is the symbol for:
a) Centrifugal fan b) Axial fan c) Diffuser d) None of the above
2-2.
This ductwork is _____________, and the dimension of the side shown is
_________.
a) Dropping, 20 b) Dropping, 12 c) Rising, 12 d) None of the above
2-3.
This is the symbol for a flexible duct:
a) True b) False
c) Cannot be determined from the information given.
2–13
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2-4.
This symbol shows
a) A blanked-off duct, with a top dimension of 12
b) A return air duct, with a side dimension of 18
c) A supply air duct, with a side dimension of 18
d) None of the above
2-5.
The shown dimension of this duct is 24:
a) True b) False
c) Cannot be determined from the information given.
2-6.
A filter that uses a liquid as an adhesive is a:
a) Carbon filter b) Electrostatic filter c) Viscous filter
d) All of the above e) None of the above
2-7.
An air handling unit may be used to:
a) Move air b) Mix air c) Heat air d) All of the above e) None of the above
2–14
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2-8.
This is the symbol for:
a) Manually operated damper b) Electrically controlled damper
c) Manual damper d) All of the above e) None of the above
2-9.
This is the symbol for:
a) Pneumatically operated damper b) Inline psychrometric observation device
c) Fire damper d) All of the above e) None of the above
2–15
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2-10.
This is the symbol for:
a) Temperature relay b) Test station c) Remote bulb thermostat
d) All of the above e) None of the above
2–16
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Chapter 3
Human Comfort and
Air Distribution
Contents of Chapter 3
•
•
•
•
•
•
3.1 Principles of Human Comfort
3.2 Principles of Space Air Distribution
3.3 Types of Air Distribution Devices
Summary
Bibliography
Skill Development Exercises for Chapter 3
3–1
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Instructions
Read the material of Chapter 3. At the end of the chapter, complete the skill development
exercises.
Study Objectives of Chapter 3
After completing this chapter, you should be able to list and explain:
•
•
•
•
The main issues of thermal comfort;
The principles of air distribution as they relate to human comfort;
The principles of space air distribution; and
The functions of the different types of air distribution devices.
3.1 Principles of Human Comfort
Human comfort depends on a variety of factors. There are factors relating to the space, the
individual and the individual’s current activity level. The space temperature, humidity, air
quality and acoustics are controlled, or influenced, by the air conditioning. However,
other space factors such as lighting are not controlled by the air conditioning.
Individuals vary. For example, one person may have a much higher metabolic rate and be
comfortable in a much cooler environment than someone else. In contrast, the elderly are
often more comfortable with a significantly higher temperature than younger people.
Finally, the activity levels of the individuals and their clothing will influence their comfort.
We will start with thermal comfort and then go on to air quality before discussing air delivery and movement in the occupied space.
THERMAL INTERCHANGE BETWEEN PEOPLE AND ENVIRONMENT
One of the first steps in designing an air distribution system for human comfort is to establish comfort criteria for the intended service. These criteria should include space temperature and humidity, ventilation rate, indoor air quality and sound level. The selection of
these criteria is influenced by many conditions including: the ages and activities of the
occupants, the occupant density and the contaminants present in the space.
The human body can be thought of as a total energy plant that operates the same way as
any other power plant. The body takes in raw materials and uses them to generate energy
for daily life and activities. One major function of the body is the heat rejection that occurs
in the thermal processes that the body goes through to produce mechanical energy.
3–2
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As shown in Figure 3-1, the body uses three major heat transfer mechanisms to reject this
heat: radiation, convection and evaporation. Radiation is important occasionally. We feel
radiation when we sit next to a window with the sun shining in, or in the winter when we
are too close to a cold window or wall.
However, in general, the basic modes of heat transfer the body uses are convection and
evaporation. They are similar in magnitude in most cases, although when we begin adjusting dry bulb temperature or humidity, a mechanism in the body reacts to that change and
shifts more of the heat transfer to one mode or the other as needed. The problem is that
both convection and evaporation depend on the same phenomenon, air motion over the
skin surface.
The evaporation from the skin surface is based on two driving forces:
•
The difference between the partial pressure of the water vapor at the skin temperature and the partial pressure of the water vapor at the dewpoint temperature in the room (how humid it is)
•
The velocity of the air past the occupant
If there is no air velocity, the mechanism of moisture diffusion is not very good. Also, the
more humid the room, the lower the mechanism to evaporate water, and consequently, the
lower the evaporative heat transfer.
Similarly, convection is driven by the difference between the skin temperature and the
space temperature. As the space temperature increases, the heat transfer decreases. As the
space temperature decreases, the heat transfer increases. Because the body tries to maintain
the skin temperature at a relatively constant level, room temperature is quite important.
Figure 3-1
Body Heat Mechanisms
3–3
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TEMPERATURE AND HUMIDITY COMFORT ZONE
Comfort is a complex, subjective response to several interacting variables. Not everyone
perceives a given temperature and humidity level with the same degree of satisfaction. The
perception of comfort relates to individual physical conditions, body heat exchange with
the surroundings, and physiological characteristics. The heat exchange between the individual and the surroundings is influenced by several factors, including:
•
•
•
•
•
•
•
Dry-bulb temperature, °F
Relative humidity, rh
Thermal radiation
Air movement, fpm
Insulation value of clothing, clo
Activity level, met
Direct contact with surfaces not at body temperature.
Two units  clo and met  are probably new for you. Clothing has an insulating value and,
in general, the greater the insulating value, the lower the ambient temperature for the same
comfort level. Typical indoor winter clothing is 1 clo; a person with shoes, socks, pants/full
length skirt, underwear, shirt and jacket. Typical light summer clothing, including
shorts/knee-length skirt and short sleeved shirt, is 0.5 clo.
The met is a unit of metabolic activity, resulting in a heat loss of about 18.4 Btu/h/ft2. A
resting adult typically produces 1 met; light office work produces 1 to 1.3 met; and walking at 2 mph produces 2 met.
Figure 3-2, which is adapted from ASHRAE Standard 55-2004, Thermal Environmental
Conditions for Human Occupancy, specifies conditions likely to be thermally acceptable to
at least 80% of the adult occupants in a mechanically conditioned space where:
•
•
•
Activity levels are between 1 and 1.3 met
Clothing is near 0.5 or 1 clo
Air speeds are below 40 fpm
The design space temperature and humidity for both heating and cooling seasons should
be based on Figure 3-2 for most applications. The comfort zone is defined for people in
winter clothing (1 clo) and summer clothing (0.5 clo), primarily engaged in sedentary
activities.
As a practical matter, the higher the conditioned space relative humidity, the cooler the
space needs to be to provide the same thermal comfort for the occupants.
3–4
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Figure 3-2
Acceptable Range of Temperature and Humidity
This has been given as a reason for increasing the humidity indoors in cold dry climates in
winter. The sales pitch is that not having to keep the indoor temperature so high saves on
the heating bill. Unfortunately, the proponents conveniently do not assess the real cost of
humidification, which is higher than the heating saving.
In a hot humid climate, dehumidification is costly in plant and operating costs. So allowing the humidity to rise saves in air-conditioning operating costs. However, allowing the
humidity to rise enough to permit mold growth can make the building uninhabitable until
very expensive remedial work has been completed.
The comfort chart indicates that relative humidity does not have a very significant bearing
on comfort as long as the space dry-bulb temperature is in the comfort range. The upper
moisture level shown as humidity ratio of 0.012 lbmoisture/lbdry air is far higher than acceptable in a building in a moist climate.
Because mold can grow in relative humidities above 60%, it is prudent to maintain buildings in hot humid climates with a humidity ratio significantly lower, at about 0.010
3–5
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lbmoisture/lbdry air. In addition, relative humidity affects odor perceptibility and respiratory
health. Because of these considerations, 40% to 50% rh is the preferred design range.
However, maintaining humidity within this range during winter is complicated by:
•
•
•
Energy costs for humidification
The risk of condensation on windows and window frames during cold weather
The need to provide and maintain humidifying equipment incorporated in the
air-conditioning system.
Where winter humidification is provided for comfort, a minimum relative humidity of
20% is generally acceptable in cold climates.
If a higher humidity level is acceptable under summer conditions, considerable energy savings can be realized, as shown in Figure 3-3. To determine an approximate value of the
energy used for dehumidification at a constant 78°F dry-bulb temperature, enter the
annual wet-bulb degree-hours above 66°F in the occupied space at the bottom left. Next,
intersect this value with the indoor relative humidity chosen, and then draw a vertical line
to the weekly hours of cooling system operation and read the energy used (in million Btu
per 1,000 cfm) on the upper-left scale.
Figure 3-3
3–6
Summer Dehumidification Energy Requirements
Repeating this procedure for a different value of relative
humidity yields the
energy savings
obtainable by raising relative humidity. However, be
cautious about
choosing excessively high humidities. Computer
rooms, particularly
computer printers
and drafting rooms,
are two applications for which relative humidity in
excess of 50% to
55% is undesirable
or unacceptable due
to effects of moisture on the paper
products.
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INDOOR AIR QUALITY
Air contaminants. Indoor air contaminants can be solid or liquid particles, gases or vapors.
Some can be irritants or odiferous, thus affecting occupant comfort. The same contaminants at higher concentrations, as well as others of which occupants may be unaware, can
be health risks. People vary in their sensitivity to contaminants. Even very small concentrations of certain fungi and other impurities can cause serious discomfort and impairment of
sensitive individuals while not affecting most occupants.
Standards for vapors and gases specify a quantity of pollutant per unit volume in parts per
million (ppm) of air. Standards for particles often specify the mass concentration of particles, expressed as micrograms per cubic meter (µg/m3). They include all particle sizes or the
total suspended particulate (TSP) concentration.
Large particles are filtered by the nasal passages and cause no adverse physiological
response unless they are allergenic or pathogenic. Smaller respirable suspended particles
(RSP) are important because they can lodge in the lungs. Respirable particles range in size
up to 5 µm. Particles of specific interest include:
• Respirable particulates as a group
• Tobacco smoke (solid and liquid droplets), which also contains many gases
• Asbestos fibers
• Allergens (pollen, fungi, mold spores and insect feces and parts)
• Pathogens (bacteria and viruses), which are almost always contained in or on
other particulate matter
Vapors and gases of interest include:
• Carbon dioxide (CO2)
• Carbon monoxide (CO)
• Radon (decay products become attached to solids)
• Formaldehyde (HCHO)
• Other volatile organic compounds (VOCs)
Although some contaminants (such as sulfur dioxide) are brought in with outside air by
mechanical ventilation or uncontrolled infiltration, most indoor contaminants come from
inside sources. People are sources of carbon dioxide, biomatter and other contaminants
characterized as body odors. People’s activities (such as smoking, cleaning, cooking, gluing
and refinishing furniture) also cause pollution. In addition, building materials and finishes
can outgas pollutants.
Furnishings, business machines and appliances (particularly unvented or poorly vented
wood- and fossil-fueled heaters and ranges) can be contaminant sources. The soil surrounding a building can be a source of radon and pesticides that enter the building
through cracks or drains or by diffusion. HVAC systems, drains, plumbing systems and
3–7
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poor construction or maintenance practices can have “environmental niches” where pathogenic or allergenic organisms collect and multiply to be reintroduced into the air. Many
microorganisms (such as molds) have accelerated growth rates at relative humidity levels
above 60%.
An additional complicating factor in the buildup of contaminants is the variation in dilution rates and effectiveness of the ventilation delivery systems often found within buildings. Concentrations vary spatially as well as over time. These variations add further nonuniformity to the pollutant concentration. ASHRAE Standard 62.1-2007, Ventilation for
Acceptable Indoor Air Quality tabulates enforceable and guideline maximum concentration
levels of common indoor contaminants. It also includes the US National Primary Ambient-Air Quality Standards for Outdoor Air used for building ventilation. If the outdoor air
source exceeds the contaminant parameters, it may be cleaned or purified prior to introduction into occupied spaces.
Outdoor air requirements. Standard 62.1-2007 provides designers with a means of determining ventilation rates needed to achieve acceptable indoor air quality, which is defined
as:
“air in which there are no known contaminants at harmful concentrations as determined by cognizant authorities and with which a substantial majority (80% or more)
of the people exposed do not express dissatisfaction.”
Two procedures for determining the required ventilation rate are offered to the designer:
the Ventilation Rate Procedure, and the Indoor Air Quality (IAQ) Procedure.
The Ventilation Rate Procedure sets forth prescriptive rates, for a variety of applications.
Unless unusual pollutants are present, these rates are intended to produce acceptable IAQ.
The basis for most of the rates specified is an underlying minimum of 5 cfm per sedentary
occupant plus a minimum of 0.06 cfm/ft2 to deal with pollutants from the space. These
minimums are increased for more active occupants, an example being 20 cfm/person in an
exercise room. Similarly, the space ventilation rate is increased where there are anticipated
contaminants; an example being 0.12 cfm/ft2 in a library.
The IAQ Procedure offers an analytical alternative, allowing the designer to determine the
ventilation rate based on knowledge of the contaminants being generated within the space
and the capability of the ventilation air supply to limit them to acceptable levels.
Exhaust requirements. Exhaust air systems are either general systems that remove air from
large spaces, or local systems that capture aerosols, heat or gases at specific locations within
a space and transport them to where they can be collected, filtered, inactivated or safely discharged to the atmosphere. The air in local exhaust systems can sometimes be dispersed
safely to the atmosphere, but sometimes contaminants must be removed so the emitted air
meets air quality standards. Standard 62.1-2007 specifies the exhaust rate for many spaces
3–8
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in terms of cfm/ft2. Examples are 0.5 cfm/ft2 for barber shops, arenas, locker rooms and
copy/print rooms. Twice the exhaust, 1.0 cfm/ft2, is required for darkrooms, janitor, trash,
recycling and science class rooms due to the higher anticipated pollution to be removed.
Air movement effect. Standard 55-2004 includes no minimum air velocity past the occupant for comfort. In the private residential environment, comfort and negligible air movement are the norm. However, the experience of many commercial building operators has
shown air motion is a significant benefit to comfort in mechanically ventilated spaces. The
standard further prescribes a maximum rate of air movement of 40 fpm to avoid drafts.
Higher air speeds (up to 160 fpm) may be used to enhance cooling, if the air speed is under
the occupant’s control.
Minimum air changes. Low air velocity may affect the ability to maintain uniformity of a
comfortable temperature throughout the occupied zone and the dilution of contaminants
generated within that zone. Occupant comfort has been reported to suffer as a consequence of low total supply air flow in the space, even when the space temperature is within
the comfort envelope. Often, this dissatisfaction is not due to air change but due to a
source of warm or cool radiation, poor temperature/humidity control, or occupant expectations.
However, to ensure adequate air changes, many designers have adopted a minimum total
supply air flow of 0.6 to 0.8 cfm/ft2 for office applications. These values are based on an
all-air system with conventional mixing supply outlets. They can be reduced when outlets
with high induction ratios are employed, because they increase the average room air
motion.
Terminal air velocity. Terminal velocity is the airstream velocity at the end of the throw
(the horizontal or vertical axial distance an airstream travels before the stream velocity is
reduced to a specified terminal velocity). The specified terminal velocity must be high
enough to maintain the desired level of comfort.
Drafts. A draft is a localized effect caused by one or more factors of high air velocity, low
ambient temperature or direction of air flow, where more heat is withdrawn from a person’s skin than is normally dissipated. It can be thought of as any air motion that causes
discomfort. Air movement in excess of 40 fpm may well be considered a draft. The location of the draft has considerable effect. The back of the neck and the ankle are the most
sensitive exposed locations.
Stratification. Stratification in a space (such as an atrium or other high ceilinged room) is
the division of air into a series of temperature layers. If conditioned air is introduced at
about the 10 ft level or below, the space close to the floor is conditioned. The cooling
requirements of the elements above the 10 ft level may be reduced.
3–9
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3.2 Principles of Space Air Distribution
Room air distribution systems include mixing, under-floor, displacement and local systems.
MIXING SYSTEMS
Conditioned air is normally supplied to air outlets at velocities much greater than those
acceptable in the occupied zone. This relatively high velocity jet of air creates mixing and
air movement to create relatively uniform air conditions in the occupied zone. The exception is underfloor systems which supply air from below the floor (see next section).
Mixing air outlets have been classified into five groups:
• Group A outlets are mounted in or near the ceiling and discharge air horizontally (see Figure 3-4). Because these outlets discharge horizontally near the ceiling, the warmest air in the room is mixed immediately with the cool primary
supply air above the occupied zone. Consequently, these outlets can handle relatively large quantities of air at large temperature differentials when cooling.
Figure 3-4
3–10
Air Motion Characteristics of Group A Outlets
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During heating, warm supply air introduced at the ceiling can cause stratification in the space if there is insufficient induction of room air at the outlet.
Selecting diffusers properly, limiting the room supply temperature differential,
and maintaining air supply rates at a level high enough to ensure air mixing by
induction can provide adequate air diffusion and minimize stratification.
•
Group B outlets are mounted in or near the floor and discharge air vertically in
a non-spreading jet (see Figure 3-5). This figure shows that a stagnant zone
forms outside the conditioned air region above its terminal point. Judgment is
needed to determine the acceptable size of the space outside the conditioned
air zone. A distance of 15 ft to 20 ft between the drop region and the exposed
wall is a conservative design value.
Figure 3-5
Air Motion Characteristics of Group B Outlets
A comparison of Figures 3-4 and 3-5 for heating shows that the stagnant region
is smaller for Group B than Group A outlets because the air entrained in the
immediate vicinity of the outlet is taken mainly from the stagnant region,
which is the coolest air in the room. This results in greater temperature equalization and less buoyancy in the total air than would occur with Group A outlets. Cooling effectiveness of Group B is inferior to Group A for the same
reasons.
3–11
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•
Group C outlets are mounted in or near the floor and discharge air in a vertical
spreading jet (see Figure 3-6). Although outlets of this group are related to
Group B, they have wide-spreading jets and diffusing action. Conditioned air
and room air characteristics are similar to those of Group B, but the stagnant
zone formed is larger during cooling and smaller during heating.
Diffusion of the primary air usually causes the conditioned air space to fold
back on the primary air during cooling, instead of following the ceiling. This
diffusing action of the outlets makes it more difficult to project the cool air, but
it also provides a greater area for induction of room air. This action is beneficial
during heating, because the induced air comes from the lower regions of the
room.
Figure 3-6
•
3–12
Air Motion Characteristics of Group C Outlets
Group D outlets are mounted in or near the floor and discharge air horizontally (see Figure 3-7). This group includes baseboard and low sidewall registers
and similar outlets that discharge the primary air in single or multiple jets.
However, because the air is discharged horizontally across the floor, the total
air, during cooling, remains near the floor, and a large stagnant zone forms in
the entire upper region of the room. During heating, the conditioned air rises
toward the ceiling because of the buoyant effect of warm air. The temperature
variations are uniform, except in the conditioned air region.
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•
Group E outlets are mounted in or near the ceiling and project primary air vertically (see Figure 3-8). During cooling, the conditioned air projects to and follows the floor, producing a stagnant region near the ceiling. During heating,
the conditioned air flow reaches the floor and folds back toward the ceiling. If
projected air does not reach the floor, a stagnant zone results.
Figure 3-7
Air Motion Characteristics of Group D Outlets
Figure 3-8
Air Motion Characteristics of Group E Outlets
3–13
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The principles of air diffusion found in these five groups are:
•
The primary air from the outlet down to a velocity of about 150 fpm can be
treated analytically. The heating or cooling load has a strong effect on the characteristics of the primary air.
•
The conditioned air (which is shown by the lightly shaded envelopes in Figures
3-4 through 3-8) is influenced by the primary air and is of relatively high velocity (but less than 150 fpm), with air temperatures generally within 1°F of room
temperature. The conditioned air is also influenced by the environment and
drops during cooling or rises during heating; it is not subject to precise analytical treatment.
•
Natural convection currents form a stagnant zone from the ceiling down during cooling, and from the floor up during heating. This zone forms below the
terminal point of the conditioned air during heating and above the terminal
point during cooling. Because this zone results from natural convection currents, the air velocities within it are usually low (approximately 20 fpm), and
the air stratifies in layers of increasing temperatures. The concept of a stagnant
zone is important in properly applying and selecting outlets, because it considers the natural convection currents from warm and cold surfaces and internal
loads.
•
A return inlet affects the room air motion only within its immediate vicinity.
The intake should be located in the stagnant zone to return the warmest room
air during cooling or the coolest room air during heating. The importance of
the location depends on the relative size of the stagnant zone that results from
various types of outlets.
•
The general room air motion (shown by clear areas in Figures 3-4 through 3-8)
is a gentle drifting of air. Room conditions are maintained by the entrainment
of the room air into the conditioned airstream. The room air motion between
the stagnant zone and the conditioned air is relatively slow and uniform. The
highest air motion occurs in and near the conditioned airstreams.
This review of outlets and their resulting airflows indicates that the air velocity and temperature vary substantially through the occupied space. The airflows are also different in
cooling and heating mode. For cooling mode, a standard method has been developed for
rating diffusers called the Air Diffusion Performance Index (ADPI). The ADPI for an outlet is the percentage of points within the occupied space where the draft temperature, , is
between –3°F and +2F and the air velocity is below 70 fpm.
3–14
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The draft temperature is a measure of perceived difference in temperature at a location as
compared to a location at the average temperature and air velocity of 30 fpm. The draft
temperature is calculated as:
Draft temperature =  = (tactual – taverage) – 0.07(local air velocity – 30)
For example, a location temperature 2F cooler than average and with an air velocity of 20
fpm will have a draft temperature of  = (–2) – 0.07(20–30) = –2 + 0.7 = –1.3 F.
To calculate the ADPI, a test room with air supplied 20 F cooler than room average is
checked at an array of points within the occupied zone, and the percentage within the draft
temperature range is the ADPI. Full details of the ADPI methodology are given in
ASHRAE Standard 113-2005, Method of Testing for Room Air Diffusion. Outlet performance selection data from numerous tests are shown for high sidewall outlets in Table 3-1
below.
Now consider a 20 ft2 office, with a cooling load of 22 Btu/h ft2 being cooled by a sidewall
diffuser. The length of throw by which the air velocity has dropped to 50 fpm, T50, is used
as the criteria. For a room load of 20 Btu/h ft2, the maximum ADPI of 85 is obtained with
a T50/L of 1.5 and, for over 80%, the ADPI range is 1.01.9. Aiming for the maximum,
choose a grille with a 50 fpm terminal velocity throw of 1.5 times the room length; T50/20
= 1.5, so T50 = 30 ft.
Table 3-1 Outlet Performance Selection Data
Terminal
Device
High sidewall
grilles
Room Load T50/L for
(Btu/h ft2) Max. ADPI
80
60
40
20
1.8
1.8
1.6
1.5
Maximum
For ADPI
ADPI
Greater Than
68
72
78
85

70
70
80
Range of
T50/L

1.5  2.2
1.2  2.3
1.0  1.9
3–15
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UNDER-FLOOR SYSTEMS
Under-Floor Air Distribution (UFAD) is supplied from a raised floor through numerous
small floor grilles. The floor typically consists of 24 in.2 metal plates, or tiles, supported by
a 1018 in.-high supporting leg, or column, at each corner. Some of the tiles have outlet
grilles installed in them. The tiles can be lifted and moved around, making grille re-location, addition or removal a simple task, as shown in Figure 3-9. Typically, the floor is covered with carpet tiles, and laying these with their joints not aligned with the tile joints substantially reduces uncontrolled leakage from the floor plenum.
Air, at 58° 64°F, is supplied to the cavity and discharges through the floor grilles. The
floor grilles are designed to create mixing, so that the velocity is below 50 fpm within 4 feet
of the floor. Think of the air as turbulent columns spreading out as they flow toward the
ceiling. Return air is taken from the ceiling or high on the wall. The rising air column takes
contaminants with it up and out of the breathing zone. This sweep-away action is considered more effective than mix-and-dilute. As a result, the ventilation requirements of
ASHRAE Standard 62.1 can be satisfied with 10% less outside air.
Figure 3-9
3–16
Under-Floor Air Distribution
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There are numerous outlets, because the individual outlet volume is typically limited to
100 cfm. The entering air does not sweep past the occupants, as occurs in displacement
ventilation, so there is no restriction on cooling capacity. However, there is a limit on how
well the system will work with rapidly changing loads. For spaces with high solar cooling
loads or high winter perimeter heating loads, thermostatically controlled fan coils or other
methods are required to modulate the capacity to match the changing load.
Because the air is rising toward the ceiling, the convection heat loads above the occupied
zone do not influence the occupied zone temperature. Therefore, the return air temperature can be warmer than the occupied zone and a return air temperature sensor is a poor
indicator of occupied zone temperature.
The cool plenum air flows continuously over the structural floor that somewhat acts as a
passive thermal storage unit. This storage can be used to reduce peak loads, but it means
the system is slow to respond to change. Night setback of temperature is not advisable and
many systems are run continuously, but without outside air, during unoccupied hours.
For perimeter heating, small fan-coil units can be installed under the floor, using finned
hot water pipes or electric coils. The tempering of the plenum air as it flows over the structure often makes it necessary to duct the plenum air some 1015 feet to the perimeter fan
coils to maintain an adequately low supply temperature. In a similar way, conference
rooms that have highly variable loads can use a thermostatically controlled fan to boost the
flow into the room when it is occupied.
A modification of the under-floor system with individual grilles is the use of a porous floor.
The floor tiles are perforated with an array of small holes, and a porous carpet tile allows air
to flow upwards over the entire tile area. This is a modification of the standard grille and
has yet to gain popularity.
The under-floor air delivery system has the following advantages:
• Changing the layouts of partitions, electrical and communications cables is
easy. For buildings with high “churn” (frequent layout changes), this flexibility
may, in itself, make the added cost of the floor economically justified.
• The flow of air across the concrete structural floor provides passive thermal
storage.
• When the main supply duct and branches to the floor plenums are part of a
well-integrated architectural design, the air supply pressure drop can be very
low, resulting in fan-horsepower savings.
• Less ventilation outside air can potentially be used.
Disadvantages include:
• A significant cost per square foot for the floor system supply, installation and
maintenance.
3–17
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•
•
A tendency to require a greater floor-to-floor height, because space for lights
and return air ducts is still required at the ceiling level.
A need for specific, and detailed, knowledge and skills on the part of the
designer and installers. Examples include: coordination between floor layout
and duct layout to avoid floor pedestals through the duct; and sealing the
structure and other service penetrations into the plenum to minimize uncontrolled leakage.
DISPLACEMENT SYSTEMS
In displacement systems, conditioned air with a temperature slightly lower than the
desired room air temperature in the occupied zone is supplied from air outlets at low air
velocities of 100 fpm or less. The outlets are located at or near the floor level for comfort
conditioning, and the supply air is directly introduced to the occupied zone. Returns are
located at or close to the ceiling through which the warm room air is exhausted from the
room.
The supply air is spread over the floor and then rises as it is heated by the heat sources in
the occupied zone. Heat sources (such as people, computers, etc.) in the occupied zone create upward convective flows in the form of thermal plumes. These plumes remove heat and
contaminants because they are less dense than the surrounding air (see Figure 3-10).
In contrast to mixing ventilation, displacement ventilation is designed to minimize mixing
of air within the occupied zone.
Figure 3-10
3–18
Schematic of Displacement Ventilation
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Unidirectional air flow systems. Here, air is either supplied from the ceiling and exhausted
through the floor, typical of many hospital operating theatre systems (or vice versa), or
supplied through the wall and exhausted through returns at the opposite wall, typical of
many industrial cleanroom systems. The outlets are uniformly distributed to provide a low
turbulent air flow across the entire room. This type of system is primarily used for ventilating cleanrooms, or for high air change areas in which the main objective is to remove
contaminant particles within the room. It is also used in areas where a unidirectional air
flow is desired (such as computer rooms, paint booths, etc.).
LOCAL SYSTEMS
Air is supplied locally for occupied
regions, such as desks in offices or
working places in industrial buildings. Conditioned air is supplied
towards the breathing zone of the
occupants to create comfortable conditions and/or to reduce the concentration of pollutants. Several special
air diffusers are available. Figure 3–11
shows one such arrangement, with
diffusers placed on the desks in front
of the occupants and the supply air
coming from a raised floor plenum.
Exhaust and return air pickup. Return
and exhaust air openings should be
located to minimize short-circuiting
of supply air into the return air open- Figure 3-11 Localized Ventilation
ing. If air is supplied by the jets
attached to the ceiling, exhaust openings should be located between the jets or at the other side of the room away from the supply air jets. In a room with temperature stratification along its height, exhaust openings
should be located near the ceiling to collect warm air, odors and fumes.
For industrial rooms with gas release, selection of exhaust opening locations depends on
the specific weight of the released gases and their temperatures. The locations should be
specified for each application.
Exhaust outlets located in walls, depending on their elevation, have the characteristics of
either floor or ceiling returns. In large buildings with many small rooms, return air should
not be brought through door grilles or undercuts into the corridors, then to a common
3–19
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return or exhaust because smoke would accumulate in the main egress pathway in case of
fire. Most building codes restrict this application.
Room system balancing. Room system balancing is adjusting the air flows within the room
so they are in accordance with specified design quantities. In designing a system, ducts and
diffusers should be sized so that the air supply is distributed properly. However, for flexibility and cost considerations, standard sizes are typically used. Consequently the room, as
designed, may not be self-balancing. The results of an unbalanced room system can be
drafts, doors slamming shut or open, and other undesirable effects. Balancing an air system
will be discussed in Chapter 11.
3.3 Types of Air Distribution Devices
Supply air outlets and diffusing equipment introduce air into a conditioned space to
obtain a desired environment. Return and exhaust air are removed from a space through
return and exhaust inlets. This section discusses some common types of diffusing equipment.
SUPPLY AIR OUTLETS
The following basic supply outlet types are commonly available: grille outlets, slot diffuser
outlets, and ceiling diffuser outlets. These differ in their construction features, physical
configurations, and the way they diffuse or disperse supply air, and induce or entrain room
air into a primary airstream.
Grille outlets. A grille outlet may be louvered or perforated, and located in a sidewall, ceiling or floor. Several types of grilles are available:
• Adjustable bar grille. This is the most common type of grille used as a supply
outlet. It is available as either a single-deflection grille (with a single set of
vanes), or double-deflection grille (with two sets of vanes, one in front of the
other, at right angles to each other). Vertical vanes deflect the airstream in the
horizontal plane; horizontal vanes deflect the airstream in the vertical plane.
• Fixed bar grille. This type is similar to the adjustable single-deflection grille,
except that the vanes are not adjustable. The vanes may be straight or set at an
angle. The angle at which the air is discharged from this grille depends on the
type of deflection vanes.
• Stamped grille. This grille is stamped from a single sheet of metal to form openings through which air can pass.
• Variable area grille. This type of grille is similar to the adjustable double-deflection grille, but can vary the discharge area to achieve an air volume change
(variable volume outlet) at constant pressure, so that the variation in throw is
minimized for a given change in supply air volume.
3–20
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Properly selected grilles operate satisfactorily from high side wall and perimeter locations
in the sill, curb or floor. Ceiling-mounted grilles, which discharge the airstream down, are
generally unacceptable in comfort air-conditioning installations in interior zones and may
cause drafts in perimeter applications.
Accessories available for grille outlets include:
• Opposed blade dampers. These can be attached to the backs of grilles (the combination of a grille and a damper is called a register) or installed as separate
units in the duct (see Figure 3–12a). Adjacent blades of this damper rotate in
opposite directions and may receive air from any direction, discharging it in a
series of jets without adversely deflecting the airstream to one side of the duct.
•
Parallel blade dampers. These have a series of gang-operated blades that rotate
in the same direction (see Figure 3–12b). This uniform rotation deflects the
airstream when the damper is partially open.
•
Gang-operated turning vanes (extractors). These are sometimes installed in collar
connections to grilles near the main ducts. The device shown in Figure 3-12c
has vanes that pivot and remain parallel to the duct air flow, regardless of the
setting. This allows for field adjustment which the fixed set of vanes shown in
Figure 3-12d do not allow.
•
Dual blade collector. Figure 3-12e shows a dual blade collector and turning vane
allowing directional control of the air as it enters the outlet.
Figure 3-12
Grille and Register Outlet Accessory Controls
3–21
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Slot diffuser outlets. A slot diffuser is an elongated outlet consisting of single or multiple
slots. It is usually installed in long continuous lengths. Outlets with dimensional aspect
ratios of 25:1 or greater, and a maximum opening of approximately 3 in. generally meet
the performance criteria for slot diffusers.
Slot diffusers are generally equipped with accessory devices for uniform supply air discharge along the entire length of the slot. While accessory devices help correct the air flow
pattern, proper approach conditions for the airstream are also important for satisfactory
performance. When the plenum supplying a slot diffuser is being designed, the transverse
velocity in the plenum should be less than the discharge velocity of the jet, as recommended by the manufacturer and also as shown by experience.
If tapered ducts are used to introduce supply air into the diffuser, they should be sized to
maintain a velocity of about 500 fpm, and tapered to maintain constant static pressure.
Slot diffusers having a single-slot discharge are available for use in conjunction with
recessed fluorescent light troffers. A plenum mates with a light fixture and is concealed
from the room. It discharges air through openings in the fixture, and is available with fixed
or adjustable air discharge patterns, air distribution plenum, inlet dampers for balancing,
and inlet collars suitable for flexible duct connections.
Accessories available for slot diffuser outlets include dampers and flow equalizing vanes.
Ceiling diffuser outlets. A ceiling diffuser is a supply-air diffuser designed for ceiling mounting. A number of designs are available:
• Multi-passage ceiling diffusers. These diffusers consist of a series of flaring rings
or louvers that form a series of concentric air passages. They may be round,
square or rectangular. For easy installation, these diffusers are often made in
two parts: an outer shell with a duct collar, and a removable inner assembly.
3–22
•
Flush and stepped-down diffusers. In the flush unit, all rings or louvers project to
a plane surface. In the stepped-down unit, the rings project beyond the surface
of the outer shell.
•
Perforated diffusers. These meet architectural demands for air outlets that blend
into ceilings. Each has a perforated metal face with an open area of 10% to
50% that determines its capacity. Units are usually equipped with deflection
devices to obtain multipattern horizontal air discharge. Large perforated diffusers are used in laboratories, hospital operating rooms and other spaces having high air change rates to provide laminar flow. Designers are cautioned to
thoroughly investigate the air flow and induction characteristics under both
cooling and heating conditions for this type of diffuser, particularly in applications with varying air flows such as VAV systems.
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•
Variable area diffusers. These feature a means of varying the discharge area to
achieve an air volume change (variable volume outlet) at a constant pressure so
that the variation in throw is minimized for a given change in supply air volume.
•
Antismudge rings. These are round or square metal frames attached to and
extending approximately 4 to 12 in. beyond the outer edge of the diffuser.
Their purpose is to minimize ceiling smudging.
Dampers and accessories of various types are available for ceiling diffusers:
•
Multilouver dampers. Consisting of a series of parallel blades mounted inside a
frame, multilouver dampers are installed in the diffuser collar or the duct system branch. The blades are usually arranged in two groups rotating in opposite
directions, and are key operated from the face of the diffuser (see Figure 3-13a).
•
Opposed-blade dampers. These usually consist of a series of pie-shaped vanes
mounted inside a round frame installed in the diffuser collar or the duct system
branch. The vanes pivot about a horizontal axis and are arranged in two
groups, with adjacent vanes rotating counter to each other (see Figure 3-13b).
The vanes are key-operated from the diffuser face. Another opposed-blade
design is similar in construction to the damper shown in Figure 3-12a, and has
either a round or square frame. Designers should note that volume control
devices near outlets can generate objectionable noise.
Figure 3-13
Ceiling Diffuser Outlet Accessory Controls
3–23
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•
Blankoff baffles. These baffles are used for minor adjustments of the air flow
from a diffuser. They blank off a section of the diffuser and prevent the supply
air from striking an obstruction such as a column, partition or the wall of the
conditioned space by reducing flow in a given direction. Blankoff baffles generally reduce the area and increase supply air velocity, which must be considered when selecting diffuser size. Pattern control in diffusers having removable
directional cores may be accomplished by rearranging the cores, generally without a change in area or increase in velocity.
Due to noise considerations, dampening in the branch duct to the diffuser is preferable to
a damper in the diffuser as long as there is easy access to the damper for balancing.
Procedure for outlet selection. The following procedure is generally used in selecting outlet
locations and types:
•
Determine the amount of air to be supplied to each room based on system
design and heating/cooling load calculations.
•
Select the type and quantity of outlets for each room, considering such factors
as air quantity required, distance available for throw or radius of diffusion,
structural characteristics, and architectural concepts. Table 3-2 is based on
experience and typical ratings of various outlets. It may be used as a guide for
the outlets applicable for use with various room air loadings. Manufacturers’
ratings should be consulted to determine the suitability of the outlets used.
•
Locate outlets in the room to distribute the air as uniformly as possible. Outlets
may be sized and located to distribute air in proportion to the heat gain or loss
in various parts of the room.
•
Select proper outlet size from manufacturers’ ratings according to air quantities, discharge velocities, distribution patterns and sound levels. Obstructions
to the primary air distribution pattern require special consideration.
Table 3-2 Outlet Usage Guide
Outlet Type
Grille
Slot
Perforated panel
Ceiling diffuser
Perforated ceiling
3–24
Floor space air loading
(cfm/ft2)
Approx. max. ACH for
10-ft ceiling
0.6  1.2
0.8  2.0
0.9  3.0
0.8  5.0
1.0  10.0





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Other supply air outlet considerations. Other supply-air outlet considerations include: the
surface effect, smudging and sound level.
The induction or entrainment characteristics of a moving airstream cause a surface effect.
An airstream moving adjacent to, or in contact with, a wall or ceiling surface creates a lowpressure area immediately adjacent to that surface, causing the air to remain in contact
with the surface substantially throughout the length of throw. The surface effect counteracts the drop of horizontally projected cool airstreams.
Smudging will occur when using ceiling and slot diffusers. Dirt particles held in suspension in the room air are subjected to turbulence at the outlet face. This turbulence is primarily responsible for smudging. The cleanliness of the room will affect when the smudging becomes visible.
An outlet’s sound level is a function of the damper arrangement, discharge velocity and
transmission of systemic noise, which is a function of the size of the outlet and the duct
velocity. Higher frequency sounds can be the result of excessive outlet velocity but may
also be generated in the duct by the moving airstream. Lower pitched sounds are generally
the result of mechanical equipment noise transmitted through the duct system and outlet.
The cause of higher frequency sounds can be pinpointed as outlet or systemic sounds by
removing the outlet during operation. A reduction in sound level with the outlet removed
indicates the portion of the noise caused by the outlet. If the sound level remains essentially
unchanged, the system is at fault. Sound will be covered in greater detail in Chapter 10.
Suggested duct velocities where takeoffs to grilles or diffusers are close to the outlet are:
• Acceptable high noise levels: 1,500 fpm maximum
• General office of classroom: 1,000 fpm maximum
• Noise sensitive areas:
800 fpm maximum
Return inlets may either be connected to a duct, or be simple vents that transfer air from
one area to another. Exhaust inlets remove air directly from a building and are always connected to a duct or directly to outside. Whatever the arrangement, inlet size and configuration determine velocity and pressure requirements for the required air flow.
In general, the same types of equipment (for example, grilles, slot diffusers and ceiling diffusers) used for supplying air may also be used for air return and exhaust. Return and
exhaust inlets do not require the accessory devices used in supply outlets. However, dampers are necessary when it is desirable to balance the air flow in the return duct system.
Return and exhaust inlets may be mounted in almost any location including ceilings, high
or low side walls, and floors when using mixing systems for supply. When using displacement and underfloor supply, distributed ceiling exhaust is required. The opposed blade
dampers shown in Figure 3-12a are used in conjunction with grille return and exhaust
inlets. The type of damper does not affect the inlet’s performance. Usually no other accessory devices are required.
3–25
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The Next Step
This chapter has discussed human comfort and air distribution within the occupied space.
This air is supplied by a distribution system, and Chapter 4 will introduce the various air
systems that provide the conditioned air.
Summary
The space, the individual and the individual’s current activity affect human comfort. Thermal comfort depends on individual characteristics and the ability of the body to reject heat,
primarily by convection and evaporation. Radiation is usually less important.
The main thermal factors affecting comfort are: dry-bulb temperature; relative humidity;
thermal radiation; air movement; insulation value of clothing; activity level; and direct
contact with warmer, or cooler, surfaces. ASHRAE Standard 55–2004 details requirements
for thermal comfort.
ASHRAE Standard 62.1-2007 prescribes supply ventilation rates, requirements for contaminant removal and exhaust rates for human satisfaction with air quality.
Room air distribution systems are classified as mixing, under-floor, displacement and local
systems. In mixing systems, the air enters the occupied zone at a fairly high velocity and
mixes with zone air to be at an acceptable velocity and average temperature in the occupied
space. Inlets are divided into five groups for air movement classification.
With mixing systems, the profile of the primary air jet can be forecast with some certainty,
but as it mixes with room air, behavior is modified by the temperature difference between
primary air and room air and the shape of the space. When the primary air is cooler, there
is a tendency for the air to drop and for stagnant areas to occur near the floor. In most
comfort situations, a general drift of air occurs through the space. This general drift is not
significantly influenced by the location of the return air outlet.
The performance of various types of mixing outlets has been analyzed for their relative
ability to maintain comfort conditions throughout a space. This data is presented as ADPI
and can be used to make choices about outlets.
Under-Floor Air Distribution (UFAD) uses a plenum created above the structural floor
using 2 ft2 metal panels on support columns. The air, at 58° 64°F, is supplied up through
diffusers distributed among the floor panels. The system uses the vertical supply and convection to lift the air towards ceiling outlets. As the air flows across the structure, the structure acts as a thermal buffer and the system is slow to change.
3–26
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In UFAD systems, perimeter heating and cooling can be challenging. The use of perimeter
fan coils and some ducting may be required to provide adequate capacity.
The UFAD system has advantages in layout flexibility, structural thermal storage, lower
fan power in some cases, and a 10% lower requirement for outside ventilation air. However, these advantages must be balanced with the cost of the floor, possibly greater floor-tofloor height, and a need for very competent design and construction.
Displacement systems, for comfort, supply a large volume of low velocity air near room
temperature. Outlets are close to, or at floor level, so the air sweeps across the space, with
convection lifting the contaminated air to high level return outlets. The system minimizes
mixing.
A wide range of grille outlets with fixed or/and adjustable vanes provide a supply of air
shaped from a narrow jet perpendicular to the room surface for a long throw to a wide
spreading, short-throw jet. The flow and throw may be adjustable by using the grille blades
or adjustable damper and turning vane accessories.
A slot diffuser is a grille with one long dimension and is often designed to be installed endto-end for long continuous air supply. Due to their length, the air supply must be carefully
designed to obtain consistent performance. Their length makes them aesthetically suitable
to install alongside fluorescent lighting fixtures.
Supply-air ceiling diffusers have flaring vanes with an open, or perforated, face. They
spread the air across the ceiling, entraining room air to produce a large volume of wellmixed circulating air.
Having established the quantity of supply air, a preliminary choice of outlet style and layout can be made. Using manufacturers’ data on air flow and sound generation, the final
airflows and layout are determined. The diffuser choice will often be significantly influenced by the available duct space for bringing air to the space and the room aesthetics.
Return air outlets can use the same grilles or diffusers, but no direction control is needed,
although a damper for balancing may be required. The location is not critical for mixing
systems, but must be high in the room for floor supply and displacement systems.
3–27
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Bibliography
ANSI/ASHRAE Standard 55-2004, Thermal Environmental Conditions for Human Occupancy.
Atlanta, GA: ASHRAE.
ANSI/ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality. Atlanta, GA:
ASHRAE.
ASHRAE Standard 113-2005, Method of Testing for Room Air Diffusion. Atlanta, GA: ASHRAE.
ASHRAE. 1993. Air-Conditioning Systems Design Manual. Atlanta, GA: ASHRAE.
ASHRAE Handbook-Fundamentals: thermal comfort, indoor environmental health, odor, space air
diffusion, heating and cooling load calculations
ASHRAE Handbook-HVAC Applications: control of gaseous indoor air contaminants
ASHRAE Handbook-Systems and Equipment: air-diffusing equipment
Skill Development Exercises for Chapter 3
Complete these questions by writing your answers on the worksheet at the back of this
book.
3–28
3-1.
The human body uses which of the following heat transfer mechanisms:
a) Radiation b) Convection c) Evaporation d) All of the above
3-2.
The perception of comfort relates to:
a) Individual physical condition
b) Body heat exchange with the surroundings
c) Physiological characteristics d) All of the above e) None of the above
3-3.
Which of the following would be within the acceptable range of temperature
and humidity for human comfort when wearing light summer clothing?
a) 72°F, 20% rh b) 70°F, 65% rh c) 80°F, 30% rh d) All of the above
3-4.
In a system with 8,000 annual wet-bulb degree hours above 66°F, with a 60%
indoor relative humidity desired, and 56 hours of cooling system operation per
week, the energy used will be _______ Btu  106 per year per 1,000 cfm.
a) 51 b) 25 c) 36 d) None of the above
3-5.
The ____________________ Procedure for determining the required ventilation rate is based on knowledge of the contaminants being generated within the
space and the capability of the ventilation air supply to limit them to acceptable
levels.
a) Indoor Air Quality b) Ventilation Rate c) Contaminant Mitigation
d) All of the above e) None of the above
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3-6.
Many designers have adopted a minimum total supply air flow of _____ for
office applications.
a) 0.1 to 0.3 cfm/ft2 b) 0.6 to 0.8 cfm/ft2 c) 0.2 to 2.0 cfm/ft2
d) All of the above e) None of the above
3-7.
The airstream velocity at the end of the throw is called:
a) Terminal velocity b) Primary velocity c) Airstream velocity
d) All of the above e) None of the above
3-8.
_________________ air distribution systems create relatively uniform air conditions in the occupied zone.
a) Unidirectional b) Local c) Mixing d) All of the above
e) None of the above
3-9.
The stagnant region of a Group B mixing outlet in a heating only system is
___________ the stagnant region of a Group A mixing outlet.
a) Larger than b) The same as c) Smaller than d) All of the above
3-10.
In displacement systems, the outlets are frequently located:
a) At or near the floor level b) In the walls c) In the ceiling
d) A and B e) None of the above
3-11.
Smudging is most likely to occur from dirt particles held in suspension in:
a) The room air b) The supply air c) The return air
d) All of the above e) None of the above
3-12.
The fan horsepower for under-floor supply systems can often be less than
required for a ceiling supply mixing system due to which of the following?
a) Much cooler supply air
b) The low resistance to air flow in the plenum
c) The insulating value of the floor and carpet
3-13.
The under-floor supply systems work well for large open areas and the most
effective control is a thermostat in the return duct. True or false?
a) True b) False
3–29
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Chapter 4
Relationship of Air
Systems to Load and
Occupancy Demands
Contents of Chapter 4
•
•
•
•
•
•
•
•
4.1 Operating System Selection Criteria
4.2 System Types by Heating/Cooling Equipment Type
4.3 System Type by Duct Configuration
4.4 Economizers
4.5 Outdoor Air Intake
Summary
Bibliography
Skill Development Exercises for Chapter 4
4–1
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Study Objectives of Chapter 4
After completing this chapter, you should be able to describe:
•
•
•
•
Operating system criteria;
Air systems by heating/cooling equipment type;
Air systems by duct configurations; and
Considerations for outdoor air intake.
4.1 Operating System Selection Criteria
To select an operating system, detailed building design and use information plus weather
data at selected design conditions are required. Although a detailed discussion of load calculations is outside the scope of this course, the air system designer should be aware that
generally all of the following are considered when performing load calculations:
4–2
•
Building characteristics. Determine building materials, areas, external surface
colors and shapes from building plans and specifications.
•
Building configuration. Determine building location, orientation and external
shading from building plans and specifications. Shading from adjacent buildings should be carefully evaluated as to its probable permanence before including it in the calculation. The possibility of abnormally high ground-reflected
solar radiation (for example, from adjacent water, sand or parking lots), or solar
load from adjacent reflective buildings should be considered.
•
Thermal zones. The thermal zones within the building should be identified. For
example, external offices with windows will have different thermal characteristics than windowless rooms in the building’s interior. Additionally, some areas
of the building may have to be kept at different temperatures than others.
•
Room pressures. Room pressure relationships should be considered. For example, in a building with a natatorium (swimming pool), the air pressure gradients within the building should draw air into the natatorium from the rest of
the building rather than vice versa. This will prevent the rest of the building
from smelling like a swimming pool. The same concept applies in buildings
with laboratories or areas where noxious smells may be generated.
•
Building uses. The uses to which the building will be put will affect the levels of
noise permissible in the building. For example, an office environment is typically less tolerant of noise from the HVAC system than a warehouse.
•
Outdoor design conditions. Obtain appropriate weather data (wet- and dry-bulb
temperatures, daily range, heating and cooling degree days, elevation, etc.) and
select outdoor design conditions from local weather stations. The ASHRAE
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Handbook–Fundamentals also lists outdoor design conditions for many
weather stations across the world. The National Climatic Center, in Asheville,
North Carolina, has additional data.
•
Space psychrometric requirements. Select indoor design conditions, such as
indoor dry-bulb temperature range, and indoor wet-bulb temperature (or relative humidity) range. Note that ASHRAE Standard 55-2004 has wide limits on
moisture content in the space as it is dealing only with comfort. The maximum
and minimum levels specified for comfort are often excessively wide for ensuring no mold growth in the building fabric or occupant complaints about low
humidity in cold climates. Include permissible variations and control limits.
Different areas within a building may have different psychrometric requirements (for example, a facility having a cleanroom, temperature-controlled laboratory and general office space).
•
Outdoor air ventilation requirements. For each space, ASHRAE Standard 62.1
specifies the methods of calculating the required supply ventilation rates and
exhaust rates; for example, polluted areas such as toilets.
•
System design and sizing. The proper design and sizing of central heating and
air-conditioning systems require more than calculation of the cooling load in
the space to be conditioned. The type of heating and air-conditioning system,
fan energy, fan location, duct heat loss and gain, duct leakage, heat extraction
lighting systems, and type of return air system all affect system load and component sizing. Adequate system design and component sizing require that system performance be analyzed as a series of psychrometric processes. The
ASHRAE Handbook–HVAC Systems and Equipment and the ASHRAE Handbook–Fundamentals describe elements of this technique in detail.
•
Operating schedule. Obtain a proposed schedule of lighting, occupants, internal
equipment, appliances and processes that contribute to the internal thermal
load. Determine the probability that the cooling equipment will be operated
continuously or shut off during unoccupied periods (such as nights and weekends). Performance of the system at part-load conditions must be considered.
•
Date and time. Frequently, several different times of day and several different
months must be analyzed to determine the peak load time. For example, in
buildings having a large amount of glass located at 32°N latitude, the peak load
times are shown in Table 4-1.
4–3
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Table 4-1 Peak Load Times
Perimeter Zone
East
West
South
North
Northeast & Southeast
Northeast & Southwest
Peak Load Time
Month
8:00 AM
4:00 PM
12:00 NOON
12:00 NOON
10:00 AM
2:00 PM
August
August
December
June
March & October
March & October
Interior zones peak at times of peak occupancy.
•
Owning and operating costs. The total cost of a facility includes the cost of the
HVAC system. The cost of an HVAC system is customarily broken down into
owning costs and operating costs. Owning costs include the initial cost of the
system and annual fixed charges that will be present whether or not the system
is used at all (taxes, insurance, etc.). Operating costs are what it costs to run the
system including energy and maintenance. The ASHRAE Handbook–HVAC
Applications provides a detailed discussion of this subject.
4.2 System Types by Heating/Cooling Equipment Type
UNITARY EQUIPMENT SYSTEMS
Unitary equipment systems are systems that are factory-assembled into an integrated package including fans, filters, heating coil, cooling coil, refrigerant compressors, refrigerantside controls, airside controls and condenser.
This equipment is manufactured in various configurations to meet a wide range of applications. Window air conditioners, through-the-wall room air conditioners, rooftop packaged units, air source heat pumps and water source heat pumps are examples. This equipment can be applied in single units and as multiple units to form a complete airconditioning system for a building.
Single-space applications. Window-mounted and through-the-wall mounted air conditioners and heat pumps are designed to cool or heat individual room spaces. They include a
complete system in an individual package. Each room is an individually controlled zone.
They are installed in buildings requiring many temperature control zones (such as motels,
apartments and dormitories). These systems are applicable for renovation of existing buildings because existing systems can still be used. However, the user should be cautioned that
these systems do not dehumidify and tend to be noisy and cause drafts.
4–4
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Entire building applications. Unitary equipment is used in both outdoor and indoor locations to cool and heat entire buildings. The complete system consists of a unit with a condenser, air distribution system and temperature controls. The equipment may be single or
multizone, installed outdoors on the roof or at grade-level, or indoors in service areas adjacent to the conditioned space. Totally indoor condenser installations require that the unit
be water-cooled.
Multiple-unit systems generally use single-zone units with a unit for each zone (see Figure
4-1). Zoning is determined by cooling and heating loads, occupancy considerations, flexibility requirements and thermal zones. Appearance considerations, costs and equipment
and duct space availability may dictate compromises in selecting the ideal zoning. Designers are also cautioned to carefully evaluate the use of unitary equipment in cases of more
than 25% outside air. Many unitary systems will not remove sufficient moisture at high
outside air quantities. For adequate part-load performance at high outside air quantities,
direct expansion systems may require a hot gas bypass to prevent coil freezing.
In both all-air systems, and air-and-water systems, air is used to perform the heating and
cooling function within the occupied space. Unitary systems are discussed in detail in the
ASHRAE Handbook–HVAC Systems and Equipment.
Figure 4-1
Multiple Packaged Units
4–5
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ALL-AIR SYSTEMS
An all-air system provides complete sensible and latent cooling, preheating and humidification capacity in the air supplied by the system. No additional cooling or humidification
is required at the zone, except in special cases. Heating may be accomplished by the same
airstream, either in the central system or at a particular zone.
All-air systems may be adapted to many applications for comfort or process work. They are
used in buildings that require individual control of multiple zones (such as office buildings, schools, universities, laboratories, hospitals, stores, hotels and ships).
All-air systems are also commonly used in special applications for close control of temperature and humidity (including clean rooms, computer rooms, hospital operating rooms,
research and development facilities), as well as many industrial/manufacturing facilities.
All-air systems have the following advantages:
4–6
•
The central mechanical equipment room location for major equipment allows
operation and maintenance to be performed in unoccupied areas and permits
the maximum range of choices of filtration equipment, and vibration and
noise control.
•
The complete absence within the conditioned area of piping, electrical equipment, wiring, filters, and vibration- and noise-producing equipment reduces
potential harm to occupants, furnishings and processes, thereby minimizing
service needs.
•
These systems have the greatest potential for the use of outside air and “free”
cooling systems to augment the use of mechanical refrigeration for cooling.
•
Seasonal changeover is simple and readily adaptable to automatic control.
•
A wide choice of zoning, flexibility and humidity control under all operating
conditions is available, including simultaneous heating and cooling, even during off-season periods.
•
Air-to-air and other heat recovery systems may be readily incorporated.
•
Good design flexibility is permitted for optimum air distribution, draft control
and adaptability to varying load requirements.
•
These systems are well suited to applications requiring unusual exhaust or
makeup air quantities (negative or positive pressurization, etc.).
•
All-air systems adapt well to winter humidification.
•
The primary system may be used to introduce outside air required for ventilation without the need for supplemental systems.
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•
By increasing the air change rate, these systems are able to maintain operating
conditions of ±1.0°F dry-bulb and ±5% relative humidity fairly simply. Some
systems can essentially maintain constant space conditions.
All-air systems have the following disadvantages:
•
They require additional duct clearance, which reduces usable floor space and
increases the height of the building.
•
Depending on layout, vertical shaft space may be needed for distribution,
thereby requiring larger floor planes.
•
The accessibility of terminal devices requires close cooperation between architectural, mechanical and structural designers.
•
Air balancing, particularly on large systems, can be more difficult.
•
Heating systems are not always available for use in providing temporary heat
during construction.
Heating and cooling calculations. Basic calculations for air flow, temperatures, relative
humidity, loads and psychrometrics are covered in the ASHRAE Handbook–Fundamentals.
It is important that the designer understands the operation of the various system components, their relationship to the psychrometric chart, and their interaction under various
operating conditions and system configurations.
Categories of all-air systems. All-air systems are classified in two basic categories: single-duct
and dual-duct. These classifications may be further divided as follows:
•
•
•
•
•
Constant volume: single zone; multiple zoned reheat; bypass
Variable air volume (VAV): reheat; induction; fan powered; dual conduit; variable diffusers
Dual-duct: constant volume; variable volume
Multizone: constant volume; variable volume; three-deck; Texas multizone
Combinations of the above systems
Constant volume single-duct. Single-duct systems contain the main heating and cooling
coils in a series flow air path. A common duct distribution system at a common air temperature feeds all terminal apparatus. These systems change the supply air temperature in
response to the space load. Variations of the constant volume single-duct system include:
single-zone systems, zoned reheat systems and bypass systems.
The single-zone system is the simplest all-air system, using a supply unit to serve a single
temperature control zone (see Figure 4-2). The unit may be installed within or remote
from the space it serves, and it may operate with or without distributing ductwork.
4–7
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In Figure 4-2, heat flows and air flows are indicated by arrows, and temperatures are indicated by t. The subscripts in the sequence of the air flow are:
R = room
rp = return plenum
o = outside air
m = mixed air
r = return
cc = cooling coil
hc = heating coil
sf = supply fan
s = supply
In the psychrometric chart in Figure 4-2, all pertinent points are identified by the same
subscripts. The room sensible and latent loads are denoted by qSR and qLR, respectively,
and the outside air sensible and latent loads are denoted by qSo and qLo, respectively. The
cooling load qcc is the difference in enthalpies between states m and cc.
Note that the cooling coil discharge air draws heat from the supply air fan and the supply
air ducts, accounting for the difference in dry-bulb temperatures between points cc and s in
Figure 4-2 before entering the room. Room sensible and latent loads (due to occupants,
lights, machinery, solar radiation, transmission, etc.) are picked up and carried to the
return air plenum. Additional heat may be picked up from recessed ceiling lights, floors
above, the roof and the return air fan, accounting for the increase in temperature between
points R and r. Some of the air is exhausted, while outside (ventilation) air o is taken in,
resulting in a mixed airstream m, which is cooled and dehumidified by the cooling coils,
producing the state of air at cc. A heating coil is provided immediately downstream of the
cooling coil to raise the air temperature in winter when required.
Properly designed systems can maintain temperature and humidity closely and efficiently
and can be shut down when desired without affecting the operation of adjacent areas. They
are energy efficient, easy to control, and easily adaptable to economizers.
Their disadvantage is that they respond to only one set of space conditions. Therefore,
their use is limited to situations where variations occur approximately uniformly throughout the zone served or where the load is stable..
Single-zone systems are applicable to small department stores, small individual stores in a
shopping center, individual classrooms in a school, computer rooms, hospital operating
rooms, and large open areas such as gymnasiums. For example, a rooftop unit, complete
with refrigeration system, serving an individual space is considered a single-zone system.
However, the refrigeration system may be remote and may serve several single-zone units
in a larger installation. A return fan may be necessary to maintain proper space pressure in
4–8
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Single Zone Schematic and Psychrometric Chart
Figure 4-2
4–9
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relation to the outside air inlet pressure and ambient. The designer should consider relief
air fans in place of return air fans if the relief path has high pressure losses. A system using
multiple fan-coil units is a collection of single-zone systems put together to control different zones.
A single-zone system can be controlled by varying the quantity and/or the temperature of
the supply air, by providing reheat, by face-and-bypass dampers, or by a combination of
these.
The multiple-zoned reheat system is a modification of the single-zone system. It provides
zone or space control for areas of unequal load, simultaneous heating or cooling of perimeter areas with different exposures, and close tolerance of control for process or comfort
applications and better performance for dehumidification. As the word reheat implies, heat
is added as a secondary simultaneous process to preconditioned primary air. Single-duct
systems without reheat offer cooling flexibility but cannot control summer humidity independent of temperature requirements.
Single-duct systems with reheat provide flexibility for both temperature and humidity control; the cooling coil cools the air to the desired humidity level, and the reheat coil raises
the dry-bulb temperature to the desired value. However, ASHRAE Standard 90.1 severely
restricts the application of reheat, limiting this option to special cases because of the high
energy consumption of the system.1 If high humidity and low dry-bulb temperatures are
desired, a humidifier may have to be included in the system.
The bypass system is a variation of the constant volume reheat system, using face-andbypass dampers in place of reheat. This system is essentially a constant volume primary system and may have a VAV secondary system.
Variable air volume single-duct. A VAV system (as shown in Figure 4-3) controls temperature within a space by varying the quantity of supply air rather than varying the supply air
temperature. A VAV terminal device is used at the zone to vary the quantity of supply air
to the space. The supply air temperature is held relatively constant, depending on the season.
VAV systems are easy to control, are highly energy efficient, allow fairly good room control, and are easily adaptable to economizers. A potential drawback includes the possibility
of poor ventilation, particularly under low zone loads. They are suitable for offices, classrooms and many other applications, and are currently widely used for commercial and
institutional buildings despite the fact that humidity control under widely varying latent
loads is difficult.
4–10
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VAV System Schematic and Psychrometric Chart
Figure 4-3
4–11
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With the current concern for indoor air quality, care should be exercised to provide minimum ventilation in any occupied space and required outside air quantities under all operating conditions. The pressure relationships of the system change when the supply fan is
throttled. Means such as outside air injection fans with capacity control may be required.
The typical return air fan generally should not be used because it is difficult to control supply and return fans in tandem. If relief is necessary, a relief fan with capacity control may
be used.
VAV systems are available in a number of configurations including:
•
Simple VAV. This system applies to cooling-only service with no requirement
for simultaneous heating and cooling in different zones; a typical application is
the interior of an office building. To permit system volume variations without
fan volume control, on chilled water systems, the air supply can ride the fan
curve down to the lowest acceptable air flow, usually at least 50% of the full air
flow. Care must be exercised in the selection of air outlets to maintain the
desired mixing and throw conditions. Avoid varying zone air volume while
keeping fan and system volume substantially constant by dumping excess air
into a return air ceiling plenum or directly into the return air duct system.
Dumping cold air into the return air plenum wastes energy and can cause overcooling under low load conditions due to radiation from the cool ceiling surface to the zone below. Dumping can also cause a shortage of system volume if
it is used for system balancing as well as temperature control. Dumping and
bypassing are generally not desirable. Fan speed control is preferred.
Three VAV box arrangements are shown in Figure 4-4. The first is the simplest.
It is a pressure-independent box, which means that it adjusts to allow for variations in supply duct pressure. The unit has a velocity sensor that is used to
control for constant velocity, and hence volume. The room thermostat requests
more, or less, flow to maintain the room temperature. The box is lined with
acoustically absorbent material, typically protected fiberglass, to reduce any
noise from the higher pressure air going over the control damper.
The second diagram shows the VAV box with a reheat coil. Typically, the supply volume is throttled to minimum flow before the coil is operated to provide
heating.
The final diagram shows a series fan box. This type of box can be used to maintain the air distribution within the space by keeping a constant volume flowing
into the space. The fan capacity meets, or exceeds, the maximum primary supply air flow. When the primary airflow is reduced, the fan draws more air from
the ceiling plenum, maintaining the constant flow. A heating coil may also be
included so that the fan and coil can run as a fan-coil heater unit, with the primary air system off during unoccupied hours.
4–12
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VAV Box Arrangements
Figure 4-4
4–13
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4–14
•
VAV reheat or VAV dual-duct. Full heating/cooling flexibility can be achieved
more energy efficiently after throttling the cold air supply to the zone.
•
VAV perimeter system. All-air cooling and heating can be accomplished by a
constant volume system serving interior spaces in connection with a VAV
perimeter system. The constant volume system provides cooling year-round,
taking care of all variations in all zone internal heat gains. The perimeter system can use an outdoor/indoor temperature schedule VAV air supply, which
simply offsets the skin transmission gains or losses. The perimeter system
requires individual zone control based on solar exposure. If a hydronic perimeter heating system is provided, the air system accomplishes all cooling in all
zones year round, while the perimeter heating system offsets the winter transmission heat losses.
•
VAV with constant zone volume. Individual zone fans may be used to maintain
minimum or constant supply air to the zone while the system primary air fed
to the zone is throttled. Terminals in these systems are commonly referred to as
fan-powered terminals. The load is satisfied by recirculating return air, thus
keeping the sum of the throttled system air and the recirculated return air substantially constant. This technique is particularly useful for zones with large
variations of internal loads (such as conference rooms), and it may be combined with terminal reheat. Fan-powered terminals can be used to ensure good
air circulation in occupied spaces during periods of reduced cooling load. Care
should be taken to ensure that proper outside air will still be delivered to the
occupied zone when the primary air is throttled. A distributed outside air duct
system may be required.
•
VAV with economizer. When the enthalpy of the outside air is lower than that of
the return air, chiller power can be reduced by taking in more outside air than
required for ventilation and relieving the excess return air. Under favorable
conditions, all of the return air can be relieved and replaced by outside air. This
mode of operation is called an economizer cycle. While this cycle requires large
outside air intakes and exhausts, it improves the economy of operation except
in areas such as the southeastern United States, where these favorable conditions occur so rarely that the additional first-cost of providing for economizer
operation is not justified.1 Even so, some Southern states have adopted energy
codes that require the use of an economizer.
•
VAV with induction terminal. The VAV induction system uses a terminal unit
to reduce cooling capacity by simultaneously reducing primary air and inducing room air or air from the ceiling return plenum to maintain a relatively constant room supply volume.
•
Dual-conduit VAV. The dual-conduit system is designed to provide two air supply paths: one to offset exterior transmission cooling or heating loads, and the
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other where cooling is required throughout the year. The typical terminal
device (box) will have two inlets, one for cold air and one for hot air or bypass
air. Each inlet will have a throttling damper and actuator. Typically, the cold
damper will be throttled to a preset minimum condition before the hot damper
is opened.
•
VAV with variable diffusers. These devices reduce the discharge aperture of the
diffuser. This keeps the discharge velocity relatively constant while reducing
the conditioned supply air flow. Under these conditions, the induction effect
of the diffuser is kept high, and cold air mixes in the space.
One important, and difficult, issue with VAV systems is providing enough ventilation air
to each space all the time. Consider a simple example where 20% outside air is required at
full system flow. If the system, as a whole, throttles back to 80% capacity, the proportion
of outside air will rise to 25% (20 out of 80). However, if one of the zones is throttled back
to 60% flow, it will only receive 0.6  0.25 = 0.15, or 15% outside air. ASHRAE Standard
62.1 provides rules for dealing with this issue.2
A second issue is ensuring adequate air distribution in the space when the volume is throttled back. Diffusers that maintain their performance at reduced flows must be chosen to
ensure that ventilation effectiveness is maintained even at times of low air flow.
Constant volume dual-duct systems. Dual-duct systems contain the main heating and cooling coils in parallel flow or series-parallel flow air paths with either: a separate cold and
warm air duct distribution system that blends the air at the terminal apparatus (dual duct
systems); or a separate supply air duct to each zone, with the supply air blended to the
required temperature at the main unit mixing dampers (multizone). The two types of constant volume dual-duct systems are:
•
Single fan – No reheat. This is similar to a single-duct system except that it
contains a face-and-bypass damper at the cooling coil arranged to bypass a
mixture of outdoor and recirculated air as the latent heat load fluctuates in
response to a zone thermostat.
•
Single fan – Reheat. This is similar to a conventional reheat system. The difference is that reheat is applied at a central point instead of at individual zones
(see Figure 4-5).
4–15
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Figure 4-5
Dual Duct System
Variable air volume dual-duct systems. Dual-duct VAV systems blend cold and warm air in
various volume combinations. These systems include:
•
Single fan. A single supply fan is sized for the coincident peak of the hot and
cold decks. Control of the fan is from two static pressure controllers: one
located in the hot deck, and the other in the cold deck. The duct requiring the
highest pressure governs the fan air flow
•
Dual fan. The volume of each supply fan is controlled independently by the
static pressure in its respective duct. The return fan is controlled based on the
sum of the hot and cold fan volumes using flow-measuring stations (see Figure
4-6).
Multizone dual-duct systems. Multizone systems supply several zones from a single centrally
located air-handling unit. Different zone requirements are met by mixing cold and warm
air through zone dampers at the central air handler in response to zone thermostats. The
mixed, conditioned air is distributed throughout the building by a system of single-zone
ducts. The return air is handled in a conventional manner. A Texas multizone system has a
heating coil in each mixed air zone, which is energized only when the cooling damper is
closed.
4–16
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Figure 4-6
Dual Duct, Dual Fan System
AIR-AND-WATER SYSTEMS
Air-and-water systems condition spaces by distributing air and water sources to terminal
units installed in habitable space throughout the building. The air and water are cooled or
heated in central mechanical rooms. Sometimes a separate electric heating coil is included
instead of a hot water coil.
The room terminal may be an induction unit, a fan-coil unit or a conventional supply air
outlet combined with a radiant panel. Generally, the air supply has a constant volume, and
is called primary air to distinguish it from room air or secondary air that has been induced.
Induction systems. Figure 4-7 shows a basic arrangement for an air-water induction terminal. Centrally conditioned primary air is supplied to the unit plenum at medium to high
pressure. The acoustically treated plenum attenuates part of the noise generated in the unit
and duct system. A balancing damper adjusts the primary air quantity within design limits.
These systems are not used very often anymore.
4–17
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Figure 4-7
Air-Water Induction Terminal
Fan-coil systems. Figure 4-8 shows a typical fan-coil unit. The basic elements of fan-coil
units are a finned-tube coil, filter and fan section. The fan recirculates air continuously
from the space through the coil or coils. The unit may contain an additional electric resistance, steam or hot-water heating coil.
Panel heating and cooling systems. The sensible heating and cooling loads in a zone can be
met by using ceiling panels. An example of one type is shown in Figure 4-9. If the panels
are used for cooling, the panel temperature must not go below the air dewpoint to avoid
any possibility of condensation. The proportion of load is thus limited in cooling applications, less so in heating applications. One very effective system is to use ceiling panels with
a dedicated outdoor air system (DOAS). The DOAS provides a constant volume of conditioned outdoor air for ventilation, humidity control and some cooling. The balance of
the cooling load is absorbed by the ceiling panels.
For heating, the floor may also be used as the heating panel. Pipes cast into a concrete floor
with warm water pumped through provide a large area for low temperature heating of the
space. For wooden floors, the pipes can be run on the underside of the floor with insulation
below to maximize the upward heat flow.
4–18
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.
Figure 4-8
Fan-Coil Unit
Figure 4-9
Ceiling Panel Example
4–19
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EVAPORATIVE COOLING SYSTEMS
Evaporative coolers exchange sensible heat for latent heat. Evaporative air cooling evaporates water into an airstream. Figure 4-10 illustrates the thermodynamic changes that occur
between the air and water in direct contact in a moving airstream. The continuously recirculated water reaches an equilibrium temperature equal to the wet-bulb temperature of the
entering air. The heat and mass transfer between the air and water lowers the air dry-bulb
temperature and increases the humidity ratio at a constant wet-bulb temperature.
The extent to which the leaving air temperature approaches the thermodynamic wet-bulb
temperature of the entering air or the extent to which complete saturation is approached is
expressed as a percentage evaporative cooling or saturation effectiveness and is defined:
t 1 – t 2 
e c = ------------------- t 1 – t' 
where:
ec = evaporative cooling or saturation effectiveness, percent
t1 = dry-bulb temperature of the entering air
t2 = dry-bulb temperature of the leaving air
t' = thermodynamic wet-bulb temperature of the entering air.
Evaporative air-cooling
equipment can be classified
as either direct or indirect.
Direct evaporative equipment cools air by direct contact with the water, either by
an extended wetted-surface
material (as in packaged air
coolers) or with a series of
sprays (as in an air washer).
Indirect systems cool air in a
heat exchanger, which transfers heat to either a secondary
airstream that has been evaporatively cooled (air-to-air) or
to water that has been evaporatively cooled (by a cooling
tower).
4–20
Figure 4-10
Thermodynamic Interaction of
Water and Air
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4.3 System Type by Duct Configuration
Duct construction is classified in terms of application and pressure. HVAC systems in
public assembly, business, educational, general factory and mercantile buildings are usually
designed as commercial systems. Air pollution control systems, industrial exhaust systems
and systems outside the pressure range of commercial system standards are classified as
industrial systems.
The designer must select a numerical static pressure class or classes that satisfy the requirements of the particular system. Duct pressure classification and duct construction will be
discussed in Chapter 7.
4.4 Economizers
Air-handling systems that have access to 100% outside air can provide full cooling without
the assistance of mechanical refrigeration whenever the outside temperature is lower than
the required supply air temperature. This so-called airside economizer (see Figure 4-11) is
progressively more effective in northern latitudes, saving up to 70% of mechanical refrigeration energy. In southern areas (such as Florida), the airside economizer is seldom used.
This is because the number of hours during which the outside enthalpy falls below the controlled space temperature is insufficient to justify the investment in the return air fan, airmixing chambers and louvers necessary to dissipate the air pressure caused by supplying
100% outside air.
Figure 4-11
Airside Economizer
4–21
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More energy savings are achieved with an economizer when:
•
The outdoor air enthalpy is lower than the supply air enthalpy required to
meet the space-cooling load; compressors and chilled water pumps are turned
off; and outdoor air, return air and exhaust air dampers are positioned to attain
the required space temperature.
•
The outdoor air enthalpy is higher than the supply air enthalpy but is lower
than the return air enthalpy; compressor and chilled water pumps are energized; and the dampers are positioned for 100% outside air.
•
The outdoor air enthalpy exceeds the return air enthalpy; the dampers are positioned to bring in the minimum outdoor air required for ventilation.
As a simple rule-of-thumb, airside economizers can be based on dry-bulb temperature (Figure 4-12). But, to be truly effective, economizer operation should be based on enthalpy, as
shown in Figure 4-13.
Figure 4-12
Airside Temperature Economizer Cycle
Compartmented air handling systems that lack the potential for 100% outside air may
adopt a winter “free cooling” concept by adding a heat exchanger in the supply airstream to
circulate the cooling tower water for cooling rather than the chilled water. This adds capital cost for the heat exchanger. Waterside free cooling is less energy conserving than airside
free cooling, depending on climate.
Another form of free cooling involves purging the conditioned areas with cool night air
prior to occupancy the following morning. This can avoid the cooling energy necessary to
4–22
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overcome the heat buildup from lights, office equipment and heat flowing back to the conditioned space from concrete floor slabs absorbing heat during the day. This purging cycle
is highly effective in dry climates with low nighttime temperatures, such as in the southwestern United States. Do not use it in hot humid climates because of the potential moisture buildup.
Figure 4-13
Enthalpy Economizer Cycle
4.5 Outdoor Air Intake
Outdoor air is air outside a building, or taken from outdoors and not previously circulated
through the system. Outdoor air that flows through a building either intentionally as ventilation air, or unintentionally as infiltration, is important for two reasons:
•
•
Outdoor air is used to dilute indoor air contaminants; and
The energy associated with heating or cooling this outdoor air is a significant
space-conditioning load.
In large buildings, the effects of infiltration and ventilation on distribution and interzone
air flow patterns, which include smoke circulation patterns in the event of fire, should be
determined, see “Fire and Smoke Management” in the ASHRAE Handbook–HVAC Applications. Outdoor air can be used to pressurize the building and minimize infiltration.
Outdoor air intakes should be located so that cross-contamination from exhaust fans to the
intake louver does not occur. Outdoor air is typically drawn in through louvers designed to
minimize the entry of snow, water, birds, trash and other foreign matter into the equip-
4–23
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ment. Figure 4-14 depicts a typical outdoor air louver design. The screen and louver are
located sufficiently above the roof to minimize the pickup of roof dust and the probability
of snow accumulating. This height is determined by the annual snowfall. However, a minimum of 2.5 ft is recommended for most areas. In some locations, doors are added outside
the louver for closure during very bad weather (such as hurricanes and blizzards).
When outdoor air must be drawn in through the roof, a gooseneck outdoor air intake like
the one in Figure 4-15 may be used. Codes also restrict the location of inlets to minimize
drawing in contaminated air. ASHRAE Standard 62.1 requires: “Use rain hoods sized for
no more than 500 fpm (2.5 m/s) face velocity with a downward-facing intake such that all
intake air passes upward through a horizontal plane that intersects the solid surfaces of the
hood before entering the system” to minimize rain entrainment.2
Figure 4-14
4–24
Outdoor Air Louver
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Figure 4-15
Gooseneck Outdoor Air Intake3
The Next Step
We have been considering supply air systems in this chapter. In Chapter 5, we will consider exhaust systems to remove excess air and contaminants from the building.
Summary
System selection depends on many factors including:
•
•
•
•
•
•
Building construction
Building layout
Schedule of operation and use of spaces
Summer and winter external design conditions
Internal design requirements and limits for ventilation, filtration, temperature,
humidity and pressure
Owning and operating costs
4–25
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Once the requirements are known, the most appropriate system can be selected.
Complete, factory-assembled units range from the small window air-conditioner serving a
single room to large packages serving a whole building. In the larger sizes, the unit may be
supplied as a set of bolt-together parts. These units range from the economical mass-produced window unit up to the best one-off designed unit.
An all-air system provides complete sensible and latent cooling, heating, humidification
ventilation and filtration through the air supplied to each space. Their main advantages are
that the equipment is outside the occupied space, which is particularly important in many
clean spaces in manufacturing and medical facilities. These systems allow for free-cooling
with outside air and heat recovery from the exhaust. They also provide air for processes
with high exhaust needs as well as flexibility in zoning and control performance.
All-air systems have disadvantages including requiring space for ducting to each zone from
the mechanical room, careful integration with the architectural layout and other services.
Systems provide temperature control by either varying the air volume and/or temperature
to each zone. For a system serving a single zone, this can be achieved at the main unit. For
multiple zones, the varying loads in each zone can be served by one of the following main
system types, or a modification of them:
•
•
•
•
Multizone: Mixing of warm and cool air at the main unit to provide a separately ducted supply to each zone.
VAV: Single supply duct supplying cool air to a variable-air-volume damper on
the branch to each zone (with a reheat coil if required).
Reheat: Single supply duct supplying a constant volume of cool air with a
reheat at each zone branch.
Dual duct: Two ducts, one with warm air, one with cold air run through the
building. At each zone, air from each duct is connected to a dual-duct box that
chooses the proportion of warm and cool air to deliver to the zone to maintain
temperature control.
Air-and-water systems provide ventilation and humidity control by supplying air to each
zone while most of the cooling and heating loads are handled by water coils in the zone.
The ventilation air may be used as the power source for inducing room air over the coil as
in induction systems, or fan-coil units may be used.
Evaporative coolers evaporate water into the air. The water absorbs latent heat to evaporate. This heat comes from the air, which lowers the air temperature. In direct evaporative
coolers, cooler wetter air is produced. In indirect evaporative coolers, water is cooled by
evaporation and used in coils to cool the air with no increase in air moisture content.
Duct construction is classified in terms of application and pressure and will be discussed in
Chapter 7.
4–26
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Mechanical cooling can be minimized by using outside air whenever the outside air
enthalpy is lower than the return air enthalpy. Depending on the climate, this may occur
most of the year or almost never. The saving in mechanical cooling operating cost is somewhat offset by the additional first-cost of larger intake, exhaust and control dampers.
Outdoor air is normally drawn in through louvers designed to minimize the entry of rain,
snow, water, birds, trash and other foreign matter into the equipment. The intake also
should be located to minimize drawing in pollutants.
Bibliography
1. ASHRAE. 2004. ASHRAE/IESNA Standard 90.1-2004, Energy Efficient Design of New Buildings
Except Low-Rise Residential Buildings. Atlanta, GA: ASHRAE.
2. ASHRAE. 2007. ANSI/ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air
Quality. Atlanta, GA: ASHRAE.
3. Carrier Corp. 1965. Handbook of Air Conditioning System Design. New York, NY: McGrawHill.
ASHRAE HandbookFundamentals: load calculations, psychrometrics; HandbookHVAC Systems
and Equipment: HVAC system analysis and selection, system types and equipment, heat
recovery; HandbookHVAC Applications: energy use, owning and operating costs,
building intake and exhaust design, evaporative cooling
4–27
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Skill Development Exercises for Chapter 4
Complete these questions by writing your answers on the worksheet at the back of this book.
4–28
4-1.
External offices with windows will have different thermal characteristics than
windowless rooms in the interior of the building:
a) True b) False
4-2.
In a building with a natatorium, the air pressure gradients within the building
should ____________________:
a) Draw air from the natatorium into the rest of the building
b) Draw air into the natatorium from the rest of the building
c) Relieve the natatorium air intake
d) All of the above e) None of the above
4-3.
Which of the following is an advantage of an all-air system?
a) Additional duct clearance is not required
b) Air balancing in large systems is less difficult
c) Vertical shaft space is not required
d) All of the above e) None of the above
4-4.
Single-duct, single-zone systems can respond simultaneously to more than one
set of space conditions, in more than one area at a time:
a) True b) False
4-5.
In air-and-water systems, the air supply generally has a constant volume:
a) True b) False
4-6.
Evaporative coolers____________________:
a) Evaporate water into an airstream b) Exchange sensible heat for latent heat
c) Can be either direct or indirect d) All of the above e) None of the above
4-7.
An economizer can achieve energy savings when _______:
a) The outdoor air enthalpy is lower than the supply air enthalpy
b) The outdoor air enthalpy is higher than the supply air enthalpy, but lower
than the return air enthalpy
c) Both of the above d) None of the above
4-8.
A minimum height of _________ above the roof surface is recommended for
locating outside air louvers where light snowfall is expected:
a)1.0 ft b) 2.5 ft c) 4.0 ft d) All of the above e) None of the above
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Chapter 5
Exhaust and
Ventilation Systems
Contents of Chapter 5
•
•
•
•
•
•
5.1 Design Considerations
5.2 Ventilation and Exhaust Systems
5.3 Energy Recovery
Summary
Bibliography
Skill Development Exercises for Chapter 5
5–1
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Instructions
Read the material of Chapter 5. At the end of the chapter, complete the skill development
exercises without consulting the text.
Study Objectives of Chapter 5
After completing this chapter, you should be able to describe design considerations for
exhaust and ventilation systems and some energy recovery systems.
5.1 Design Considerations
Ventilation and exhaust systems control heat, odors and contaminants. The two types of
exhaust systems are:
•
General exhaust, in which an entire workspace is exhausted without considering
specific operations; and
•
Local exhaust, which is applied to specific areas. Local exhaust offers better control with minimum air volumes, thereby lowering the cost of air cleaning and
replacement air equipment. Local exhaust is required for hazardous contaminant exhaust.
Ventilation may be provided by natural draft, by a combination of general supply and
exhaust air fan and duct systems, by exhaust fans only (with makeup air through inlet louvers and doors), or by supply fans only (exhaust through relief louvers and doors).
VENTILATION SYSTEM SELECTION AND DESIGN
Some factors to consider in ventilation system selection and design include:
5–2
•
Local exhaust systems provide general ventilation for the work area.
•
A balance of the supply and exhaust systems is required for either system to
function as designed.
•
Natural ventilation systems are most applicable when internal heat loads are
high and the building is tall enough to produce a significant stack effect (such
as steelmaking plants and glass-melting furnaces).
•
To provide effective general ventilation for heat relief by either natural or
mechanical supply, the air must be delivered low in the work zones. A sufficient
exhaust volume is necessary to remove the heat liberated in the space. Local
relief systems may require supplemental supply air for heat removal.
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•
Supply and exhaust air cannot be used interchangeably. Supply air can be delivered where it is wanted at controlled velocities, temperature and humidity.
Exhaust systems should be used to capture heat and fumes at the source.
•
General building exhaust may be required in addition to local exhaust systems.
•
The exhaust discharge should not be located where it will be recirculated into
the outdoor air intake.
•
The inlet air quantity of the exhaust is established by the volume and velocity
required to contain and remove heat and contaminants. For human occupancy,
ASHRAE Standard 62.1, Ventilation for Acceptable Indoor Air Quality has
requirements for ventilation air and exhaust, as was described in Chapter 2.1
For industrial applications, minimum values are prescribed for local exhaust
systems in Industrial Ventilation, A Manual of Recommended Practice2 and
sometimes by code.
•
Properly sized ductwork keeps contaminants flowing. This requires high velocities for heavy materials. The selection of materials and the construction of
exhaust ductwork and fans depend on the nature of the contaminant, the
ambient temperature, the lengths and arrangement of duct runs, the method of
fan operation, and the flame and smoke spread rating.
•
Care must be taken to minimize the following:
• Corrosion, destruction by chemical or electrochemical action.
• Dissolution, a dissolving action. Coatings and plastics are subject to dissolution, particularly by solvent fumes.
• Melting, which can occur in certain plastics and coatings at such elevated
temperatures as may be found in an exhaust system.
• Abrasion due to conveyed particles impacting the duct, particularly at fittings.
•
Low temperatures that cause condensation in ferrous metal ducts may increase
corrosive attack. Ductwork is less subject to attack when the runs are short, and
direct to the terminal discharge point. The longer the runs, the longer the
period of exposure to fumes and the greater the degree of condensation. Horizontal runs allow moisture to remain longer than it can on vertical surfaces.
Intermittent fan operation can contribute to longer periods of wetness
(because of condensation) than continuous operation. Exhaust ducts from
high-moisture areas (such as shower rooms) must have drains and watertight
bottoms. Corrosion-resistant material should be considered.
•
The national and local Clean Air Acts have requirements for controlling the
discharge of contaminants to the atmosphere.
5–3
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MAKEUP AIR
For safe, effective operation, most industrial plants require makeup air to replace the large
volumes of air exhausted. If makeup air is provided consistently with good air distribution,
more effective cooling can be provided in the summer, and more efficient and effective
heating will result in the winter. Using windows or other inlets that cannot be used in
stormy weather should be discouraged. The needs for makeup air include:
•
To replace air exhausted through combustion processes and local and general
exhaust systems.
•
To eliminate uncomfortable cross-drafts by proper arrangement of supply air,
and prevent infiltration through doors, windows and similar openings that
may make exhaust hoods unsafe or ineffective, defeat environmental control,
bring in or stir up dust, or adversely affect processes.
•
To obtain clean air. Supply air can be filtered, infiltration air cannot. Also, supply air can be preheated to prevent spot freeze-up, infiltration air cannot.
•
To control building pressure and air flow from space to space. Such control is
necessary:
• To avoid positive or negative pressures that will make it difficult or unsafe to
open doors and to avoid the conditions that are detailed in Table 5-1.
• To confine contaminants, reduce their concentration, and control temperature, humidity and air movement positively.
• To recover heat and conserve energy.
Table 5-1 Negative Pressures That May Cause Unsatisfactory Building Conditions
Negative
Pressure (in.wg)
0.01  0.02
0.01  0.05
0.02  0.05
0.03  0.10
0.05  0.10
0.10  0.25
5–4
Adverse Conditions
Worker Draft Complaints: High velocity drafts through doors and windows
Natural Draft Stacks Ineffective: Ventilation through roof exhaust ventilators, flow
through stacks with natural draft greatly reduced
Carbon Monoxide Hazard: Back-drafting will occur in hot water heaters, unit heaters,
furnaces and other combustion equipment not provided with induced draft
General Mechanical Ventilation Reduced: Air flows reduced in propeller fans and low
pressure supply and exhaust systems
Doors Difficult to Open: Serious injury may result from nonchecked, slamming doors
Local Exhaust Ventilation Impaired: Centrifugal fan exhaust reduced
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STACK EFFECT
Temperature differences between indoors and outdoors cause density differences and,
therefore, pressure differences that drive infiltration. During the heating season, the
warmer air rises and flows out of the building near its top. It is replaced by colder outdoor
air that enters the building near its base. During the cooling season, the stack effect is
reduced and pressures reversed, because the indoor-outdoor temperature differences are
smaller and reversed. Qualitatively, the pressure distribution over the building in the heating season due to the stack effect takes the form shown in Figure 5-1.
The height at which the interior and exterior pressures are equal is called the neutral pressure level (NPL). Above this point (during the heating season), the interior pressure is
greater than the exterior; below this point, the greater exterior pressure causes air flow into
the building.
Figure 5-1
Pressure Differences Due to Stack Effect (Heating Season)
5–5
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The pressure difference due to the stack effect at height h is:
 p s = C 2   o –  i  g  h – h NPL  = C 2  i g  h – h NPL   T i – T o T o
(5-1)
where:
ps = pressure difference due to stack effect, in. wg
 = air density, lbm /ft3 (about 0.075 for indoor conditions)
g = gravitational constant, 32.2 ft/s2
h = height of observation, ft
hNPL = height of neutral pressure level, ft
T = average absolute temperature, °R
C2 = unit conversion factor, 0.00598
Regarding the unit conversion factor, water weighs 62.4 lbm per ft3. So 1 inch of water is
62.4/12 lbm per inch depth. For ps to be in in. wg, ps must be multiplied by (62.4/12) 
32.2 = 167.17. Moving 167.17 to the right hand side of the equation changes it to
1/167.17 = 0.00598.
Subscripts: i = inside; o = outside
5.2 Ventilation and Exhaust Systems
This section describes some of the more common ventilation and exhaust systems.
VENTILATION FOR HEAT RELIEF
Many situations involve processes that release heat and moisture to the environment. Ventilation is one of many controls that may be used to mitigate heat stress conditions.
The HVAC designer must distinguish between the control needs for hot-dry and warmmoist conditions. In the first case, the process gives off only sensible and radiant heat without adding moisture to the air. The heat load on exposed workers is increased, and the rate
of cooling by evaporation of sweat is increased. Heat balance may be maintained, although
possibly at the expense of excessive sweating. In the warm-moist situation, the wet process
gives off mainly latent heat. The rise in the heat load on workers may be small, but the
increase in moisture content of the air reduces heat loss by evaporation of sweat by the
workers.
Hot-dry work situations occur around hot furnaces, forges, metal-extruding and rolling
mills, glass-forming machines, and so forth. Typical warm-moist operations are found in
5–6
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many textile mills, laundries, dye houses and deep mines where water is used extensively
for dust control. However, these industrial applications are outside the scope of this course.
Where appropriate, local exhaust ventilation can remove the natural convection column of
heated air rising from a hot process with a minimum of air from the surrounding space.
TOILET EXHAUST
The ventilation of locker rooms, toilets and shower spaces is important to remove odor and
reduce humidity. Supply air may be introduced through door or wall grilles. In some cases,
plant air may be so contaminated that filtration, or mechanical ventilation, may be
required. When mechanical ventilation is used, the supply system should have supply fixtures such as wall grilles, ceiling diffusers or supply plenums to distribute the air adequately
throughout the area. Pressure relationships must be carefully considered to prevent air flow
from locker rooms, toilets and shower spaces to other occupied spaces.
ASHRAE Standard 62.1 includes general exhaust requirements including those shown in
Table 5-2. Note that where the lockers are being used for laboring employees with wet,
sweaty clothes, the rate should be increased to the higher of 1 cfm/ft2 or 7 cfm exhausted
from each locker. Where heavy labor is involved and the clothes may be wet and have
picked up odors, the rate should be increased to the higher of 3 cfm/ft 2 or 10 cfm
exhausted from each locker.
Table 5-2 Ventilation for Locker Rooms, Ancillary and Toilet Spaces
Space
Locker rooms
Locker/dressing rooms
Janitor, trash, recycling
Toilets - public
Toilets - private
CFM
per Unit
CFM/FT2
0.50
0.25
1.00
50/70*
25/50*
*The toilet rate is per water closet and/or urinal. Provide the
higher rate where periods of heavy use are expected to occur;
for example, toilets in theaters, schools and sports facilities.
The lower rate may be used where use is intermittent.
5–7
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KITCHEN EXHAUST
Kitchens typically have a great concentration of noise, sensible and latent heat load, smoke
and odors. Ventilation is the chief means of removing and preventing these elements from
entering other occupied spaces. Kitchen air pressure should be kept negative relative to
other areas to ensure odor control. Maintenance of reasonably comfortable working conditions is important.
Kitchens present common load problems encountered in other occupied space, with additional factors including:
•
•
•
•
Extremely variable loads with high peaks, in many cases occurring twice daily
High sensible and latent heat gains because of gas, steam and electric appliances, people and food
Heavy infiltration of outdoor air through doors during rush hours in commercial establishments
Grease in the ductwork
Codes require exhaust hoods with grease filters for cooking equipment where grease is generated, and hoods over all gas-fired appliances. Other equipment that generates a lot of
heat, or moisture, should be located under hoods.
SMOKE CONTROL
When a fire occurs in a building, smoke often flows to locations remote from the fire,
threatening life and damaging property. Stairwells and elevators frequently become smokefilled, blocking or inhibiting evacuation. Smoke causes the most deaths in fires. Smoke
control describes systems that use pressurization produced by mechanical fans to limit
smoke movement in fire situations.
A smoke control system must be designed so that it is not overpowered by the driving
forces that cause smoke movement, including:
5–8
•
Stack effect. As discussed earlier, when the air outside a building is colder than
the building air, the building air moves upward within building shafts (such as
stairwells, mechanical shafts and elevator shafts). This is the normal stack
effect. When the outside air is warmer than the building air, a downward, or
reverse stack effect, occurs. Smoke movement from a building fire can be dominated by stack effect. In a building with normal stack effect, the existing air
currents can move smoke considerable distances from the fire origin.
•
Buoyancy. High temperature smoke from a fire has a buoyancy force due to its
reduced density. As smoke travels away from the fire, its temperature drops due
to heat transfer and dilution. Therefore, the effect of buoyancy generally
decreases with distance from the fire.
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•
Expansion. In addition to buoyancy, the energy released by a fire can move
smoke by expansion. The ratio of volumetric flows can be expressed as a ratio
of absolute temperatures:
Q out
T out
---------- = ---------Q in
T in
(5-2)
where:
Qout = volumetric flow rate of smoke out of the fire compartment, cfm
Qin = volumetric flow rate of smoke into the fire compartment, cfm
Tout = absolute temperature of smoke leaving the fire compartment, °R
Tin = absolute temperature of smoke entering the fire compartment, °R
•
Wind. Frequently in fire situations, a window breaks in the fire compartment.
If the window is on the leeward side of the building, the negative pressure
caused by the wind vents the smoke from the fire compartment. This reduces
smoke movement throughout the building. However, if the broken window is
on the windward side, the wind forces the smoke throughout the fire floor and
to other floors, which endangers the lives of building occupants and hampers
firefighting. Pressures induced by the wind in this situation can be large and
can dominate air movement throughout the building.
•
HVAC system. The HVAC system frequently transports smoke during fires.
Before the concept of using the HVAC system for smoke control, systems were
shut down when fires were discovered. Although shutting the system down
prevents it from supplying air to the fire, it does not prevent smoke movement
through the supply and return air ducts, air shafts and other building openings
due to stack effect, buoyancy or wind.
Additional information on smoke control can be found in the ASHRAE Handbook–HVAC
Applications.
STAIR PRESSURIZATION SYSTEMS
Many pressurized stairwells have been designed and built to provide a tenable escape route
in the event of a building fire. They also provide a staging area for firefighters. On the fire
floor, a pressurized stairwell must maintain a positive pressure difference across a closed
stairwell door so that smoke does not enter the stairwell.
During building fire situations, some stairwell doors are opened intermittently during
evacuation and firefighting, and some doors may even be blocked open. Ideally, when the
stairwell door is opened on the fire floor, air flow through the door should be sufficient to
5–9
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prevent smoke backflow. Designing such a system is difficult because of the many combinations of open stairwell doors and weather conditions affecting air flow.
The stairwell pressurization fan must be sized to allow for doors to be open to floors and
often to the outside during the fire. If no doors are open, the static pressure could easily rise
high enough to make opening doors very difficult. To avoid this over-pressurization, some
form of pressure control is often provided. A simple barometric relief damper with wind
shield can be used to relieve any excess pressure to atmosphere. Alternatively, pressure sensors measuring the pressure between a floor and the stairwell can control a damper on a
short-circuit duct around the fan. When the pressure rises above the setpoint pressure, the
damper opens to let air short-circuit around the fan, thereby lowering its capacity.
The maximum allowed design pressure difference across a door is typically 0.20.3 in. wg
so that it can be opened. The minimum pressure to hold back smoke is about 0.08 in. wg,
so the pressure control should be designed to hold the pressure from floor to stairwell in
that range. Controls to limit differential pressures at the doors are very complicated and
difficult to maintain.
Stairwell pressurization systems may be single and multiple injection systems. A single
injection system has pressurized air supplied to the stairwell at one location, usually at the
top. Associated with this system is the potential of smoke entering the stairwell through the
pressurization fan intake. Therefore, automatic shutdown during such an event should be
considered. For tall stairwells, single injection systems can fail when a few doors are open
near the air supply injection point. Such a failure is especially likely when a ground-level
stairwell door is open in bottom injection systems.
Multiple injection points are recommended no more than 45 ft apart (see Figures 5-2a and
5-2b). Compartmentation of a stairwell is illustrated in Figure 5-3.
Additional information on stair pressurization can be found in the ASHRAE Handbook–
HVAC Applications.
5–10
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Figure 5-2a
Stairwell Pressurization,
Ground-Level Fan
Figure 5-3
Figure 5-2b
Stairwell Pressurization,
Roof-Mounted Fan
Compartmentation of Pressurized Stairwell
5–11
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HEALTHCARE FACILITIES
The application of air conditioning to healthcare facilities presents many problems not
encountered in the usual comfort conditioning system. The basic differences between air
conditioning for medical facilities and other types of facilities stem from:
•
•
•
•
The need to restrict air movement in and between the various departments
Specific requirements for ventilation and filtration to dilute and remove contamination in the form of odor, airborne microorganisms and viruses, and hazardous chemical and radioactive substances
The need for different temperature and humidity requirements for various
areas
The need for sophistication in design to permit accurate control of environmental conditions
The specific environmental conditions required by a particular medical facility can be
complex, and vary depending on the planned use of the facility and the agency responsible
for the facility environmental standard. Healthcare facilities are discussed in greater detail
in the ASHRAE Handbook–HVAC Applications.
5.3 Energy Recovery
Much of this chapter has focused on exhaust. Air that has been cooled or heated before
exhausting is taking energy from the building. In many situations, it is both possible and
practical to recover some of that energy. Recovery may be of sensible heat or sensible and
latent heat.
ENERGY RECOVERY COILS
Run-Around Coils: One way to achieve energy recovery is with run-around energy recovery
coils. A typical run-around coil arrangement is shown in Figure 5-4.
In summer, the conditioned exhaust air cools the fluid in the exhaust air coil. This fluid is
then pumped over to the supply air coil to pre-cool the incoming outside air.
In winter, the heat transfer works the other way; the warm exhaust air heats the fluid in the
exhaust air coil, which is then pumped over to the supply air coil to heat the cold incoming
air.
At intermediate temperatures, the system is shut off, because it is not useful.
When outside temperatures are below freezing, the three-way valve is used with a glycol
antifreeze mixture in the coils. In cold weather, some of the fluid bypasses the supply air
coil, to avoid overcooling. The mixture of very cold fluid from the supply air coil and
5–12
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diverted fluid mix to a temperature that is high enough to avoid causing frost on the
exhaust air coil. The maximum amount of cooling that can be achieved with the exhaust
air coil is limited by the temperature at which frost starts to form in the coil. This frosting
of the exhaust coil effectively sets a limit to the transfer possible at low temperatures.
The run-around coil system has three particular advantages:
•
There is no possibility of cross-contamination between the two airstreams.
This factor makes it suitable for hospital or fume hood exhaust heat recovery.
The exhaust coil must be resistant to corrosion from any chemicals in the
exhaust.
•
The two coils do not have to be adjacent to one another. A laboratory building
could have the outside air intake low in the building and the fume hood
exhaust on the roof, with the run-around pipes connecting the two coils.
•
The run-around coils transfer sensible heat, and under favorable conditions,
condense the water in the exhaust to recover latent heat. This makes them particularly suitable for natatoriums in some climates.
Figure 5-4
Run-Around Energy Recovery Coils
5–13
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HEAT PIPES
A heat pipe is a length of pipe with an interior wick that contains a refrigerant charge, as
shown in Figure 5-5.
The type and quantity of refrigerant that is installed is chosen for the particular temperature requirements. In operation, the pipe is approximately horizontal and one end is
warmed, which evaporates refrigerant. The refrigerant vapor fills the tube. If the other half
of the tube is cooled, the refrigerant will condense and flow along the wick to the heated
end, to be evaporated once more. This heat-driven refrigeration cycle is surprisingly efficient.
The normal heat pipe unit consists of a bundle of pipes with external fins and a central
divider plate. Figure 5-6 shows a view down onto a unit that is mounted in the relief and
intake airstreams to an air-handling unit. Flexible connections are shown that facilitate the
tipping. To adjust the heat transfer, one end or the other end of the tubes would be lifted
The outside air is cold as it comes in over the warm coil. This warms the air, and the tube
is cooled. The cooled refrigerant inside condenses, giving up its latent heat, which heats the
air. The re-condensed refrigerant wicks across to the exhaust side and then absorbs heat
from the exhaust air. This heat evaporates the refrigerant back into a vapor that fills the
pipe, and is again available to warm the cold outside air.
The usual heat-pipe unit must be approximately horizontal to work well. A standard way
to reduce the heat transfer is to tilt the evaporator (cold) end up a few degrees. This tilt
control first reduces, and then halts, the flow of refrigerant to the evaporator end, and the
process stops.
Figure 5-6 was based on winter operation. In summer, the unit only has to be tilted to
work the other way and cool the incoming outside air as it heats the outgoing exhaust air.
The unit is designed as a sensible heat transfer device; although allowing condensation to
occur on the cold end can transfer worthwhile latent heat. Effectiveness ratings range up to
80% with 14 rows of tubes. However, each additional row contributes proportionally less
to the overall performance. As a result, the economic choice is ten or fewer rows.
A major advantage of the units is very low cross-contamination.
5–14
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Figure 5-5
Cutaway Section of a Heat Pipe
Figure 5-6
Heat Pipe Assembly in Exhaust and Outside Air Entry Pipe
5–15
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DESICCANT WHEELS
Desiccants are chemicals that are quick to pick up heat and moisture, and quick to give
them up again if exposed to a cooler, drier atmosphere. A matrix, as shown on the left of
Figure 5-7, may be coated with such a chemical and made up into a wheel several centimeters thick. In use, the supply air is ducted through one half of the wheel and the exhaust air
through the other half.
Suppose it is a hot summer day, so the exhaust is cooler and drier than the supply of outside air. The chemical coating in the section of the coil in the exhaust stream becomes relatively cool and dry. Now the wheel is slowly rotated and the cool, dry section moves into
the incoming hot, humid air, drying and cooling the air. Similarly, a section is moving
from hot and humid into cool and dry, where it gives up moisture and becomes cooler.
The wheel speed  a few revolutions per minute  is adjusted to maximize the transfer of
heat and moisture. Control of wheel speed to truly maximize savings is a complex issue,
because the transfers of sensible and latent heat do not vary in direct relation to each other.
The depth of the wheel is filled with exhaust air as it passes into the supply airstream, so
there is some cross-contamination. There are ways of minimizing this cross-contamination, but it cannot be eliminated. In most comfort situations, the cross-contamination in a
well-made unit is quite acceptable.
The use of heat recovery is required in many energy codes, particularly for larger systems
and systems with a high proportion of outside air. ASHRAE Standard 90.1-2007Energy
Standard for Buildings Except Low-Rise Residential Buildings has several mandatory requirements for the use of heat recovery equipment.3
Figure 5-7
5–16
Desiccant Wheel Matrix and Operation Pipe
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The Next Step
This chapter has covered ventilation and exhaust. The next chapter will cover fans and the
movement of air through systems.
Summary
Ventilation and exhaust systems control heat, odors and contaminants. Exhausts can be:
•
Local: removing the contaminant before it mixes with the air in the space
•
General: changing the air in the space on a regular basis to keep the concentration of contaminants down to an acceptable level
All the air exhausted must enter the building, so there is a balance. Failure to provide adequate supply air makeup can create problems of pressure difference. Therefore, exhausts
must be designed with the supply system. For human comfort, the supply of outside air
was the criteria. But for many commercial and industrial processes, the exhaust is the criteria. The process often determines the volume and the construction of the exhaust system
to deal with corrosion and erosion.
In natural exhausts, stack effect can be used as the motive power where there are sufficient
and reliable temperature differences.
Ventilation for heat relief under hot working conditions is used in many industries. It is
less effective in moist conditions because sweating is less effective.
Locker rooms and toilets should be kept at a slightly negative pressure relative to surrounding areas to contain smells. Building codes usually dictate the minimum exhaust per fixture.
Kitchen exhaust fumes are typically warm, aromatic and grease laden. Most codes require
the use of grease filters to reduce the quantity of grease (which deposits in the ducts) and to
reduce the likelihood of fire entering the ducts. The large quantity of exhaust makes kitchens a challenge for supplying adequate makeup air at a reasonable operating cost.
Smoke control systems are designed to provide a small pressure difference between the fire
zone and other zones. Maintaining this difference, less than 0.1 in. wg, can be very difficult
due to:
•
Stack effect where the difference between inside and outside temperatures
causes pressure differences
•
Buoyancy of the hot gases from a fire
5–17
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•
Expansion of the air due to temperature around the fire
•
Wind blowing past the building, creating a higher pressure on the windward
side and a lower pressure on the leeward side
•
The HVAC system, if it is left running
Stairwell pressurization is provided to keep smoke out of the means-of-escape and firefighter access routes. Design is a challenge as the pressure must be maintained even with
doors open but limited to prevent doors being held shut by the pressure. Barometric
dampers and short-circuit ducts on fans are used to regulate the effective supply fan capacity.
Energy recovery from large exhausts is often economically very attractive and is mandated
in energy codes.
Energy recovery coils: One coil in the exhaust piped to another coil in the makeup air system
allows the energy to literally be pumped from exhaust to intake. In freezing climates, an
antifreeze mixture is used. The system has the advantages of enabling the intake and
exhaust to be separated, and there is zero cross-contamination.
Heat pipes: Transfer heat using the boiling and condensation of refrigerant in sealed
lengths of pipe to transfer heat from one end of the tube to the other. Capacity control is
by tilting the pipes. Some cross-contamination may occur.
Desiccant wheels: Desiccant wheels are deep porous wheels coated in a chemical to collect
heat and moisture. The wheel slowly rotates in the two airstreams, collecting moisture and
energy from one airstream and giving up energy and moisture to the other airstream. Their
value is in very high recovery rates. However, there is some cross-contamination, which is
an issue in processes with toxic exhaust contaminants.
Bibliography
1. ANSI/ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality. Atlanta, GA:
ASHRAE.
2. ACGIH. 2004. Industrial Ventilation, A Manual of Recommended Practice. Cincinnati, OH:
American Conference of Governmental Industrial Hygienists.
3. ASHRAE Standard 90.1-2007, Energy Standard for Buildings Except Low-Rise Residential
Buildings. Atlanta, GA: ASHRAE.
ASHRAE HandbookFundamentals: ventilation, stack effect; HandbookHVAC Applications:
ventilation and exhaust for specific applications, energy recovery, fire and smoke
management; HandbookHVAC Systems and Equipment: air-to-air energy recovery
5–18
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Skill Development Exercises for Chapter 5
Complete these questions by writing your answers on the worksheets at the back of this
book.
5-1.
Natural ventilation systems are most applicable when the building will produce
a significant stack effect:
a) True b) False
5-2.
Care must be taken in exhaust systems to minimize:
a) Corrosion b) Dissolution c) Melting d) All of the above
e) None of the above
5-3.
All other things being equal, ductwork is least subject to condensation corrosion
when the runs are:
a) Long and horizontal b) Short and vertical
c) Direct to the terminal discharge
d) All of the above e) None of the above
5-4.
Kitchen air pressure should be kept ______________ relative to other areas.
a) Positive b) Neutral c) Negative d) All of the above e) None of the above
5-5.
Smoke movement is driven by:
a) Stack effect b) Buoyancy c) Expansion d) All of the above
e) None of the above
5-6.
To prevent smoke infiltration on a fire floor, a pressurized stairwell must maintain a _________________ pressure difference across a closed stairwell door.
a) Positive b) Neutral c) Negative d) All of the above e) None of the above
5-7.
Health facility ventilation requires:
a) Little need for accurate control of temperature and humidity
b) Free movement of air between departments
c) Removal of airborne microorganisms
d) All of the above e) None of the above
5–19
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Chapter 6
Air Movers and Fan
Technology
Contents of Chapter 6
•
•
•
•
•
•
•
•
•
6.1 Fan Principles
6.2 Fan Drives
6.3 Fan Selection
6.4 Fan Installation Design
6.5 Fan Controls
6.6 Effect of Variable Resistance Devices
Summary
Bibliography
Skill Development Exercises for Chapter 6
6–1
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Study Objectives of Chapter 6
After completing this chapter, you should be able to:
•
•
•
•
List and explain fan principles;
List and describe the main types of HVAC fans, fan drives and fan controls;
Explain factors to be considered when selecting an appropriate fan for a given
set of conditions; and
Explain factors to be considered when installing a fan, once it has been
selected.
6.1 Fan Principles
A fan is an air pump that creates a pressure difference and causes air flow. The impeller
does work on the air, imparting both static and kinetic energy, varying in proportion
depending on the fan type.
Symbols and definitions commonly encountered when working with fans include:
V
Wo
= fan outlet area, ft2
= fan size or impeller diameter
= rotational speed, rpm (sometimes revolutions per second)
= volume flow rate moved by fan at fan inlet conditions, cfm
= fan total pressure rise; fan total pressure at outlet minus fan total pressure at
inlet, in. wg
= fan velocity pressure; pressure corresponding to average velocity determined
from the volume flow rate and fan outlet area, in. wg
= fan static pressure rise; fan total pressure rise diminished by fan velocity pressure, in. wg. The fan inlet velocity head is assumed equal to zero, because the
inlet is not connected to ductwork and unobstructed for fan rating purposes.
= fan inlet or outlet velocity, fpm
= power output of fan; based on fan volume flow rate and fan total pressure, hp
Wi
= power input to fan; measured by power delivered to fan shaft, hp
ht
= mechanical efficiency of fan (or fan total efficiency); the ratio of power output to power input (ht = Wo /Wi )
hs
= static efficiency of fan; mechanical efficiency multiplied by the ratio of static
pressure to fan total pressure, hs = (ps /pt )ht

= gas density, lb/ft3
A
D
N
Q
ptf
pvf
psf
6–2
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PRINCIPLES OF FAN OPERATION
Fans produce pressure by altering the velocity vector of the flow. Fans produce pressure
and/or flow because the rotating blades of the impeller impart kinetic energy to the air by
changing its velocity. This velocity change is the result of tangential and radial velocity
components in the case of centrifugal fans, and of axial and tangential velocity components
in the case of axial flow fans.
Centrifugal fan impellers produce pressure from:
•
The centrifugal force created by rotating the air column enclosed between the
blades
•
The kinetic energy imparted to the air by virtue of its velocity leaving the
impeller
Axial flow fans produce pressure from the change in velocity passing through the impeller,
with none being produced by centrifugal force.
The basic fan types can be further subdivided as follows:
•
Centrifugal fans: airfoil, backward inclined/backward curved, forward curved
and radial
•
Axial fans: propeller, tubeaxial and vaneaxial
•
Special designs: tubular centrifugal, centrifugal power roof ventilator and axial
power roof ventilator
•
Plug fans
Figure 6-1 illustrates most of these fans, and provides details of the impeller design, housing design, performance characteristics and typical applications.
The plug, or plenum, fan is not shown. A single inlet impeller, similar to one for a centrifugal fan is mounted on the end of the drive shaft. The impeller is mounted between two
walls. It draws the air into the centre of the impeller and blows it out evenly in all directions. This fan design can be particularly useful in compact air-handling units and in
industrial situations (such as ovens) when the only components subjected to the high temperature are the impeller and drive shaft.
6–3
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6–4
Types of Fans
Figure 6-1
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Figure 6-1 Types of Fans (cont.)
6–5
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FAN LAWS
The fan laws (see Table 6-1) relate the performance variables for any dynamically similar
series of fans. Fan Law 1 shows the effect of changing size, speed or density on volume
flow, pressure and power. Fan Law 2 shows the effect of changing size, pressure or density
on volume flow rate, speed and power. Fan Law 3 shows the effect of changing size, volume flow or density on speed, pressure and power.
Table 6-1 Fan Laws
Fan Law 1
1a
Q1 = Q2

(D1/D2)3 (N1/N2)
1b
p1 = p2

(D1/D2)2 (N1/N2)2 1/2
1c
W1 = W2

(D1/D2)5 (N1/N2)3 1/2
2a
Q1 = Q2

(D1/D2)2 (p1/p2)1/2 ( 2 /1)1/2
2b
N 1 = N2

(D2/D1) (p1/p2)1/2 (2 /1)1/2
2c
W1 = W2

(D1/D2)2 (p1/p2)3/2 (2 /1)1/2
3a
N 1 = N2

(D2/D1)3 (Q1/Q2)
3b
p1 = p2

(D2/D1)4 (Q1/Q2)2 1/2
3c
W1 = W2

(D2/D1)4 (Q1/Q2)3 1/2
Fan Law 2
Fan Law 3
Notes:
1. Subscript 1denotes the variable for the fan under consideration. Subscript 2 is the variable for the
tested fan.
2. For all fan laws,
3. p equals either ptf or psf
The fan laws simplify analyzing a given fan because there is no change in fan size (D) or
density (). Another way to remember these relationships is that the air quantity is directly
proportional to fan speed. Static pressure varies as the square of the speed change. Power
input to the fan varies as the cube of the speed.
6–6
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Figure 6-2 illustrates the application of the fan laws for a change in fan speed N for a specific size fan. The computed ptf curve is derived from the base ptf curve. For example,
Point E (N1 = 650) is computed from Point D (N2 = 600) as follows:
At Point D,
Q2 = 6,000 cfm and ptf2 = 1.13 in. wg
Using Fan Law 1a at Point E,
Q1 = 6,000  650/600 = 6,500 cfm
Using Fan Law 1b,
ptf1 = 1.13 (650/600)2 = 1.33 in. wg
The completed total pressure curve, the ptf1 at N=650 curve, may be generated by computing additional points from data on the base curve, such as Point G from Point F.
Figure 6-2
Sample Application of the Fan Laws
6–7
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If equivalent points of rating are joined (as shown by the dotted lines in Figure 6-2), these
points will form parabolas that are defined by the relationship expressed in the following
equation:
Q
p

---------2  =  ------2- 
 p 
Q 
1
2
(6-1)
1
Each point on the base curve ptf determines only one point on the computed curve. For
example, Point H cannot be calculated from either Point D or Point F. However, Point H
is related to some point between these two points on the base curve, and only that point
can be used to locate Point H. Furthermore, Point D cannot be used to calculate Point F
on the base curve. The entire base curve must be defined by test.
Finally, the horsepower required by a fan is related to both the volume flow rate, and the
pressure. The relationship can be expressed in several ways:
hp~p  Q
or hp~Q 3
or hp~N 3
This is an important observation because when dealing with an existing system where all of
the components are fixed in place, if the amount of air moving can be reduced by changing
the speed, the power requirement is reduced by the cube of the reduction in cfm. For
example, if the air flow rate is reduced by 20% to 80% of the previous, Q2 /Q1 = 0.80.
Therefore,
hp
Q 3
3
--------2- =  ------2-  = 0.8 = 0.512
Q 
h p1
1
(6-2)
The fan power is reduced to 51.2% of the original amount. Another way to express this is
that a 20% reduction in air flow results in a 48.8% reduction in power.
FAN AND SYSTEM PRESSURE RELATIONSHIPS
As previously stated, a fan impeller imparts static and kinetic energy to the air. This energy
is represented in the increase in total pressure and can be converted to static or velocity
pressure. These two quantities are interdependent; fan performance cannot be evaluated
by considering one or the other alone. The conversion of energy, indicated by changes in
velocity pressure to static pressure and vice versa, depends on the efficiency of conversion.
Energy conversion occurs in the discharge duct connected to a fan being tested in accordance with the joint standards AMCA Standard 210 and ASHRAE Standard 51, and the
efficiency is reflected in the rating.1,2
6–8
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Fan total pressure ptf is a true indication of the energy imparted to the airstream by the fan.
System pressure loss (p) is the sum of all individual total pressure losses plus system effects
imposed by the arrangement of duct elements on both the inlet and outlet sides of the fan.
An energy loss in a duct system can be defined only as a total pressure loss. The measured
static pressure loss in a duct element equals the total pressure loss only in the special case
where air velocities are the same at both the entrance and exit of the duct element. By using
total pressure for both fan selection and air distribution system design, the design engineer
is assured of proper design. These fundamental principles apply to both high- and lowvelocity systems. (ASHRAE Handbook-Fundamentals has further information.)3
A very important relationship is:
V 2
p v =  ------------  in. wg.
 4005 
(6-3)
To specify the pressure performance of a fan, the relationship of ptf , psf and pvf must be
understood, especially when negative pressures are involved. Most importantly, psf is a
defined term in AMCA Standard 210 and ASHRAE Standard 51 as psf = ptf – pvf . Except
in special cases, psf is not necessarily the measured difference between static pressure on the
inlet side and static pressure on the outlet side of the fan.
Figures 6-3 through 6-6
illustrate the relationships among these various pressures. Note that,
as defined, p tf = p t2 –
p t1 . Figure 6-3 illustrates a fan with an outlet system but no connected inlet system. In
this particular case, the
fan static pressure p sf
equals the static pressure rise across the fan.
Figure 6-4 shows a fan
with an inlet system but
Figure 6-3 Pressure Relationships of Fan With Outlet
no outlet system. Figure
System Only
6-5 shows a fan with
both an inlet system and
an outlet system. In both cases, the measured difference in static pressure across the fan
( ps2  ps1) is not equal to the fan static pressure.
6–9
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Figure 6-4
Figure 6-5
6–10
Pressure Relationships of Fan With Inlet System Only
Pressure Relationships of Fan With Equal-Sized Inlet and
Outlet Systems
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All of the systems illustrated in Figures 6-3 to 6-5 have inlet or outlet ducts that match the
fan connections in size. Usually the duct size desired is not identical to the fan outlet or the
fan inlet, so a further complication is introduced. To illustrate the pressure relationships in
this case, Figure 6-6 shows a diverging outlet cone, which is a commonly used type of fan
connection. In this case, the pressure relationships at the fan outlet do not match the pressure relationships in the flow section. Furthermore, the static pressure in the cone increases
in the direction of flow because the velocity pressure is decreased. The static pressure
changes throughout the system, depending on velocity.
The total pressure (which, as noted in the figure, decreases in the direction of flow) more
truly represents the loss introduced by the cone or by flow in the duct. Only the fan
changes this trend (that is, the decrease of total pressure in the direction of flow). Therefore, total pressure is a better indication of fan and duct system performance. In this rather
normal fan situation, the static pressure across the fan ( ps2  ps1) does not equal the fan
static pressure ( psf ). This phenomenon is known as system effect, which is discussed later
in this chapter.
Figure 6-6
Pressure Relationships of Fan With Diverging Cone Outlet
6–11
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FAN TESTING AND RATING
Fan efficiency ratings are based on ideal conditions. Some fans are rated at more than 90%
total efficiency. However, necessary inlet and outlet arrangements often make it impossible
to achieve ideal efficiencies in the field.
Fans are tested in accordance with the strict requirements of ASHRAE Standard 51 and
AMCA Standard 210. These joint standards specify the procedures and test setups to be
used in testing the various types of fans and other air-moving devices.
Figure 6-7 depicts one of the most common procedures for developing the characteristics
of a fan. The fan is tested from shutoff conditions to nearly free delivery conditions. At
shutoff, the duct is completely blanked off; at free delivery, the outlet resistance is reduced
to zero. Between these two conditions, various flow restrictions are placed on the end of
the duct to simulate various conditions on the fan. Sufficient points are obtained to define
the curve between shutoff point and free delivery conditions. A nozzle chamber is often
used to determine the air flow rate. The point of rating may be any point on the fan performance curve. For each case, the specific point on the curve must be defined by referring
to the flow rate and the corresponding total pressure.
Other test setups, also described in AMCA Standard 210 and ASHRAE Standard 51,
should produce the same performance curve.
Figure 6-7
6–12
Method of Obtaining Fan Performance Curves
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Fans designed for use with duct systems are tested with a length of straight duct between
the fan discharge and the measuring station on a flow-through test setup. This length of
duct smooths the flow of the fan and provides stable, uniform flow conditions at the plane
of measurement. This allows the centrifugal fan with a cutoff, and the vaneaxial or propeller fan discharge velocities to equalize along the duct, and the difference in velocity pressure is converted to available static pressure. In the case of free discharge or duct fittings
near the fan outlet or inlet, some of the pressure conversion is not realized. The measured
pressures are corrected back to fan outlet. Fans designed for use without ducts (including
almost all propeller fans and power roof ventilators) are tested without ductwork.
Not all sizes are tested for rating. Test information may be used to calculate the performance of larger fans that are geometrically similar, but such information should not be
extrapolated to smaller fans. For the performance of one fan to be determined from the
known performance of another, the two fans must be dynamically similar. Strict dynamic
similarity requires that the important nondimensional parameters vary in only insignificant ways. These nondimensional parameters include those that affect aerodynamic characteristics such as Mach number, Reynolds number, surface roughness and gap size. (For
more specific information, consult the manufacturer.)
6.2 Fan Drives
A proper motor and drive selection aids in long life and minimum service requirements.
Direct drive fans are normally used on applications where exact air quantities are not
required (such as with small fan-coil units), because ample heat transfer surface is available
at more than enough temperature difference to compensate for any lack of air quantity that
may exist. For example, this could apply to a unit heater application. Direct drive fans are
also used on applications where system resistance can be accurately determined. However,
most air-conditioning applications use belt drives.
V-belts must be applied in matched sets and used on balanced sheaves to minimize vibration problems and to ensure long life. They are particularly useful on applications where
adjustments may be required to obtain more exact air quantities. These adjustments can be
accomplished by varying the pitch diameter on adjustable sheaves, or by changing one or
both sheaves on a fixed sheave drive system.
Belt guards are required for safety on all V-belt drives, and coupling guards are required for
direct drive coupling equipment.
The fan motor must be selected for the maximum anticipated brake horsepower requirements of the fan plus drive losses. The motor must be large enough to operate within its
rated horsepower capacity including drive losses and reductions in line voltages and shortterm conditions. Normal torque motors are generally used for fan duty.
6–13
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6.3 Fan Selection
Figure 6-8 shows two fan characteristic curves for the same fan. They are constant speed
curves. Curve 1 is run at one speed, curve 2 at a lower speed. In terms of fan selection, the
objective is always to keep the operating point somewhere in the optimum selection zone
illustrated in Figure 6-8. If the fan is to operate in zone A, a larger fan will be more efficient. Conversely, if the fan is to operate in zone B, a smaller fan will be more efficient.
Keep in mind that a fan is a constant volume device.
There is no magic number to defining the optimum zone, although it should include maximum efficiency. The application will also dictate the appropriate width of the optimum
zone. Some HVAC applications allow a fairly wide optimum zone. In areas where big fans
requiring a lot of energy are needed (such as mills or power plants), the optimum zone is
much narrower because it is more important to operate near peak efficiency.
Figure 6-8
6–14
Optimum Fan Selection Zone
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Figure 6-9 shows a series of maximum efficiency curves for various fan sizes. It is plotted on
log-log paper to show the exponential curve as a straight line. The fan sizes shown are a
standard range, where 365 is a fan with a 36.5 in. diameter impeller or wheel.
The value of a chart like Figure 6-9 is that, once it is prepared for a given type of fan, you
can enter the X-axis with the cfm and the Y-axis with total pressure, defining a point in the
graph. The fan represented by the curve closest to that point is the most efficient fan on the
chart for that cfm/pressure combination.
In theory, the chart indicates the best fan. However, both the next smaller and the next
larger fans should be evaluated for the particular application, even though they are both
less efficient and possibly noisier.
In practice, the AMCA sizes are so close together that it is quite likely that the next larger
or smaller size will probably be acceptable. For example, suppose the chart suggests a 36.5
in. fan. It is quite likely that you can go down to the 33 in. fan. While it will be less efficient, it will be down only a few points, it will not be that much noisier, and the first-costs
will be lower. For variable volume applications, the next smaller size fan should always be
evaluated.
Note that the curves in Figure 6-9 are for one type of fan. If you have another type, a series
of curves must be obtained from the manufacturer for that type.
Another important point is that you cannot satisfy all applications simply by speeding up
the fan. Recall from the fan laws earlier in this chapter that the horsepower goes up as the
cube of the speed ratio (hp = cfm3). Suppose you have 100% of design air in a system, and
it is determined that an additional 10% is required. If the fan is sped up by 10%, the pressure goes up by the square of the speed increase, to about 1.12 = 1.21 or 121%. However,
the power requirements go up from 100% to 133% (1.13 = 1.331), and few systems can
easily tolerate that big an increase.
Happily, this works in reverse. That is, if the cfm can be reduced, the whole process is
reversed. Instead of being worried about overloading the system and requiring new equipment, you are cutting the power bill appreciably because of the cube ratio.
DENSITY, TEMPERATURE AND ALTITUDE
Unless otherwise identified, fan performance data are based on dry air at standard conditions: 14.696 psi and 70°F, with a density of 0.075 lb/ft3. In actual applications, the fan
may be required to handle air or gas at some other density. The change in density may be
because of temperature, composition of the gas, or altitude. As indicated by the fan laws,
6–15
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6–16
Maximum Efficiency Lines for Various Fan Sizes
Figure 6-9
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fan performance is affected by gas density. With constant size and speed, the power and
pressure vary in accordance with the ratio of gas density to the standard air density.
Most of the time, air handling systems are operated at or near sea level, so altitude is not a
consideration. However, at higher altitudes, atmospheric density becomes a factor. At
higher altitudes, or when handling gases lighter than standard air, the pressure is lowered.
When working with a gas of lower density than air at sea level, the air cannot build up the
pressure that the original standard air could.
However, flow rate does not change. If a fan produces 10,000 cfm at sea level, it will produce 10,000 cfm at 5,000 ft above sea level, but not at the same pressure. The flow rate will
remain the same no matter what the density. The change is strictly in pressure. Happily, a
reduction in horsepower also occurs, which comes down as the density reduces. The point
to remember is that catalog information is developed at standard density, and it has to be
converted to lower density, and lower density air will not transfer as much heat at higher
altitudes as it will at sea level. Consequently, the required air flow for a given energy delivered may need to be increased, which results in higher pressures and possibly higher horsepower requirements.
STATIC PRESSURE VERSUS TOTAL PRESSURE
Fan data in catalogs for unitary equipment are usually specified in static pressure, not total
pressure. This can cause errors in selection. The objective of this section is to show the
problem and alert you to the prospective difficulties you may encounter.
For example, assume the duct system for two systems has a static pressure loss of 1 in. wg,
and assume a fan that delivers 4,000 cfm across an outlet area of 1 ft2, giving a velocity of
4,000 fpm. From the equation for velocity pressure:
4000 2
V 2
p v =  ------------  =  ------------  = 1in.
 4005 
 4005 
(6-4)
If we arbitrarily establish a static pressure of 1 in. at the fan outlet, the total pressure
becomes 1 in. + 1 in. = 2 in.
Now consider another fan. Here is the same 4,000 cfm, but this fan has 2 ft2 of output
area. Therefore, the velocity is 2,000 fpm. Again using Equation 6-1, the velocity pressure
equals (2000/4005)2, or 0.25 in. To make things equal, we want 2 in. total pressure just as
before. Subtracting the 0.25 in. velocity pressure from the total pressure leaves 1.75 in. of
static pressure.
The trouble starts when we consider efficiency. The equation for efficiency, , is:
6–17
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 cfm   p 
 = ---------------------------- 6362   hp 
where,
(6-5)
 = efficiency
p = pressure, in. wg
If we use static pressure in this equation, we get static efficiency. If we use total pressure, we
get total efficiency.
Some computations will demonstrate the effect of output area on static efficiency. Assume
that from tests it has been determined that hp = 1.57.
 cfm   p t 
 100%
 t = ---------------------------- 6362   hp 
(6-6)
Total efficiency in both cases:
 4000   2 
 t = ----------------------------------  100% = 80%
 6362   1.57 
Static efficiency in case 1:
 4000   1.00 
 s = ----------------------------------  100% = 40%
 6362   1.57 
(6-7)
Static efficiency in case 2:
 4000   1.75 
 s = --------------------------------6362   1.57   100% = 70%
(6-8)
So here are two fans with the same horsepower (1.57), the same total pressure (2 in.), the
same cfm (4,000), and the same total efficiency (80%). However, by doubling the outlet
area, the static efficiency has been increased from 40% to 70%. All because the outlet areas,
and consequently the outlet velocities, of the two fans are different. By using total efficiency, you can avoid costly mistakes that can easily occur by looking at static efficiencies.
The reason that fans are sometimes specified in terms of static pressure is that, particularly
in older systems, when velocities are low, the difference between static pressure and total
pressure is relatively small.
6–18
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However, particularly in systems with higher velocities (>1,500 fpm), it is important to deal
with total pressure, not static pressure.
A fan introducing unheated outside air will discharge a larger cfm of air after the air is
heated. The fan motor should be selected for this added horsepower.
The density of air varies as the absolute temperature difference where standard air in
degrees Rankine (°R) is (70° + 460°) = 530°R. Therefore, a 4,000 cfm rated fan whose discharge air is heated to 170°F (630°R) introduces (4000  630°/530°) = 4,750 cfm into the
duct system.
FAN PERFORMANCE UNDER INSTALLED CONDITIONS
It is not unusual for a fan
and system combination to
operate at a volume flow
rate and pressure different
from those for which the
system was designed. There
are two basic reasons why
this may occur. First, if a
system is not the same system as specified in the
design, the point of operation will not be at the
design point on the fan
curve. Referring to Figure
6-10, Point B is the specified point of operation, but
the system actually operates
at Point A.
The different point of operation produces a different
combination of capacity
Figure 6-10 Operating Points
and pressure; in the case
shown, a higher pressure
and a lower flow rate. If the
original design volume flow rate must be retained, the situation can be corrected by changing the fan speed until the fan curve and the system curve pass through the required capacity point. Another way of correcting this situation is to reduce the pressure loss in the system so that the point of rating moves out on the curve to point B, as shown in Figure 6-10.
This change in the system characteristics may be accomplished by a change in damper set-
6–19
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ting, a change in outlet grille setting, or an actual change in the duct design to achieve the
lower pressure characteristic.
The important note in this case is that the difference between the specified point of rating
and the actual point of rating is due to a change in the system characteristic curve and not
a difference in the fan. The fan curve is in its original position; the challenge is simply to
get the system characteristic curve to cross the fan curve at the desired point.
Second, an entirely different change in the operation between the fan and the fan system
can occur by an actual change in the fan performance curve. Remember, all fans impart
energy to the air by some form of rotational motion. Fans are designed so they depend on
uniform, straight flow into the fan inlet. If this flow is upset in any way, the fan will not
perform on the original performance curve, but rather will work on a new one. Why this
happens is a system effect.
SYSTEM EFFECT FACTORS
Figure 6-11 illustrates deficient fan/system performance resulting from one
or more undesirable flow
conditions (improper outlet connections, non-uniform inlet flow and/or swirl
at the fan inlet). It is
assumed that the system
pressure losses have been
accurately determined
(Point 1, Curve A) and a
suitable fan selected for
operation at that point.
However, no allowance has
been made for the effect of
the system connections on the
fan’s performance. To compensate for this system
effect, a system effect factor Figure 6-11 Deficient Duct System Performance3
must be added to the calculated system pressure losses
to determine the actual system curve. The system effect is treated as a pressure loss even
though it cannot be accurately measured as such in the field. The system effect factor for
any given configuration is velocity dependent and will, therefore, vary across the range of
flow volumes of the fan (see Figure 6-12).
6–20
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In Figure 6-11, the point of intersection between the fan performance curve and the actual
system curve is Point 4. The actual flow volume will, therefore, be deficient by the difference from 1–4. To achieve design flow volume, a system effect factor equal to the pressure
difference between Points 1 and 2 should have been added to calculate system pressure
losses and the fan selected to operate at Point 2. Note that because the system effect is
velocity related, the difference represented between Points 1 and 2 is greater than the difference between Points 3 and 4.
Figure 6-12 shows a series of 24 system effect curves (labeled A through X); determination
of which curve to use is discussed later in this section. By entering the chart at the appropriate air velocity (on the abscissa), it is possible to read across from any curve (to the ordinate) to find the system effect factor for a particular configuration. The system effect factor
is given in in. wg, and must be added to the total system pressure losses, as shown in Figure
6-11.
The velocity figure used in entering the chart will be either the inlet or the outlet velocity
of the fan. This will be dependent on whether the configuration in question is related to
the fan inlet or the outlet. Most catalog ratings include outlet velocity figures but, for centrifugal fans, it may be necessary to calculate the inlet velocity. A more detailed discussion
of system effects and tables detailing system effects for a wide range of equipment and configurations can be found in the AMCA publication Fans and Systems.3
6–21
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6–22
System Effect Curves3
Figure 6-12
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6.4 Fan Installation Design
COMPUTING THE EFFECT OF FAN OUTLET CONDITIONS
Imagine an ideal uniform flow downstream from the fan. However, the reality is quite different. Figure 6-13 shows the flow patterns of a centrifugal fan and an axial fan. In either
case, the flow is non-uniform at the fan discharge.
Ideally, the outlet duct should be the same size as the fan outlet. To best use the energy
developed by the fan, the length of duct known as the 100% effective duct length should
be provided at the fan outlet. Acceptable flow can be obtained if the duct is not greater in
area than 110%, nor less in area than 85% of the fan outlet, and system effects can usually
be tolerated at fan outlet velocities below 2,000 fpm. The slope of transition elements
should not be greater than 15° for the converging elements, nor greater than 7° for the
diverging elements.
Figure 6-13
Blast Areas for Centrifugal and Axial Fans3
6–23
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There will be a system effect for most fans at effective duct lengths of less than 100% of
straight duct. Closer than that, and there will be an effect such as the one illustrated in
Table 6-2, and the losses at other duct components (elbows, tees, etc.) will be higher than
listed in the ASHRAE or SMACNA handbooks.
Table 6-2 Blast Area Ratios for Various Fan Types
Fan Type
Blast Area Ratio
Centrifugal
Airfoil
Backward-curved
Backward-inclined
Modified radial
Radial
Forward-curved
0.70
0.70
0.70
0.60
0.80
0.50
Propeller
0.90
Axial
Hub ratio:
0.3
0.4
0.5
0.6
0.7
0.90
0.85
0.75
0.65
0.50
Note: Use actual manufacturer’s data when available.
One way to calculate effective duct length for round duct is as follows:
•
If the duct velocity is greater than 2,500 fpm:
V A 
 o o
L e = ---------------------10 600
•
If the duct velocity is less than 2,500 fpm:
A
L e = ----------o
4.3
6–24
(6-9)
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where:
Vo = duct velocity, fpm
Le = effective duct length, ft
Ao = duct area, in.2
If the duct is rectangular, the equivalent duct diameter is calculated by:
4HW
D h = -------------------------2H + W 
(6-10)
where:
Dh = equivalent duct diameter, in.
H = rectangular duct height, in.
W = rectangular duct width, in.
In those cases where you use a shorter discharge length than one effective duct length, an
additional pressure loss will result. This additional pressure must be added to the fan total
pressure requirements. The additional pressure loss may be calculated by Equation 6-11,
which is also used to calculate additional pressure losses for other inlet and outlet configurations:
 V 
p = K 1   ------------ 
 1096 
2
(6-11)
where:
p = pressure loss, in. wg
V = velocity at outlet plane, fpm
K1 = factor from appropriate tables
 = density, lb/ft3
The blast area ratio is calculated by:
Blast Area Ratio = Blast Area/Outlet Area
Typical blast areas for centrifugal and axial fans are identified in Figure 6-13. The blast area
for centrifugal fans is the outlet area minus the area of the cutoff plate. The blast area for
axial fans is the area of the annular space between the hub and the fan housing. The blast
area should be obtained from the fan manufacturer for the particular fan being considered.
For estimating purposes, the values of blast area ratio shown in Table 6-2 may be used if
actual areas cannot be determined.
6–25
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Elbows can contribute to additional pressure loss. To obtain the rated performance from a
fan, the first elbow fitting should be at least one effective duct length from the fan outlet
(see Figure 6-14). If this length cannot be provided, an additional pressure loss will result,
and this additional pressure must be added to the fan total pressure requirements using the
curve letter designation shown in Figure 6-12 and Table 6-3. The additional pressure loss
may also be determined by using Equation 6-11.
Figure 6-14
6–26
Outlet Duct Elbows3
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Table 6-3
System Effect Curves for Outlet Elbows3
Blast Area Outlet Elbow No Outlet 12% Effective 25% Effective 50% Effective 100% Effective
Outlet Area
Position
Duct
Duct
Duct
Duct
Duct
0.4
A
B
C
D
N
M
LM
LM
O
MN
M
M
PQ
O
N
N
S
R
Q
Q
0.5
A
B
C
D
P
NO
MN
MN
Q
OP
NO
NO
R
PQ
OP
OP
T
S
RS
RS
0.6
A
B
C
D
Q
B
NO
O
QR
Q
OP
P
RS
R
PQ
QR
U
T
S
ST
0.7
A
B
C
D
ST
RS
QR
R
T
S
R
RS
U
T
S
ST
W
V
UV
UV
0.8
A
B
C
D
S
R
Q
QR
ST
RS
QR
R
TU
ST
RS
S
VW
UV
U
UV
0.9
A
B
C
D
S-T
R-S
R
RS
T
S
RS
S
U
T
ST
T
W
V
UV
V
1.0
A
B
C
D
RS
ST
RS
RS
S
T
S
S
T
U
T
T
V
W
V
V
No System Effect
Factor
These factors are for single-width single-inlet (SWSI) fans. For double-width double-inlet (DWDI) fans, apply the following multipliers:
Elbow position B = P  1.25
Elbow position D = P  0.85
Positions A & C = P  1.00
6–27
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COMPUTING THE EFFECT OF FAN INLET CONDITIONS
If an elbow must be installed on the fan inlet, a straight run of duct should be put between
the elbow and the fan and a long radius elbow should be used. Inlet elbows without the
straight duct run create an additional loss that must be added to the fan total pressure
requirements. The additional loss may also be calculated by using Equation 6-11.
COMPUTING THE EFFECT OF INLET OBSTRUCTIONS
For obvious reasons, every effort should be made to keep the fan inlet free of obstructions.
The fan inlet should be located so it is not obstructed (by other equipment, walls, pipes,
beams, columns, etc.), because such obstructions will degrade the fan’s performance.
Where such obstructions are unavoidable, the resulting pressure losses can be estimated by
using Equation 6-11. The K Factors for inlet area obstructions are shown in Table 6-4.
Table 6-4 K Factor for Inlet Area Obstructions
% Inlet Area
Obstructed
K Factor
5
10
15
25
50
75
0.22
0.40
0.53
0.80
1.20
1.60
When you estimate the percentage of inlet area remaining obstructed, use that part of the
projected area of the obstruction perpendicular to the air flow and subtract this area from
the area of the inlet plane to obtain the net area. Divide the flow rate by this net area to
determine the flow for V in the above equation.
INLET SPIN
Figure 6-15a shows top and front views of two inlet duct combinations. Fans are normally
tested with open inlets and uniform flow to the wheel. When angled ductwork is too close
to the fan inlet (as shown in the figure), a spin component is imparted to the air. The flow
is no longer uniform and nonstandard fan performance results. This means that the fan is
no longer operating along the expected curve, and the fan performance is different than
specified. It does not matter if it is spun in the direction of wheel rotation, or against the
direction of wheel rotation.
If there is uncontrolled spin in the direction of the wheel, pressure is lost and the flow rate
is reduced. A good indication of this is that the horsepower goes down. If you are getting
lower horsepower than the manufacturer’s data indicates, this may be the cause.
If there is uncontrolled spin against the direction of the wheel, the results are slightly
higher pressure, lower flow and higher than expected power draw. If there is enough spin,
6–28
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enough power will be drawn to blow the circuit breakers or heaters on your system. If you
know your fan is overloaded, but you cannot figure out why, because the system appears in
good order otherwise, look for uncontrolled spin. The best remedy is to enlarge the duct
approaching the fan to reduce the velocity and, therefore, the loss. Another remedy for this
condition may be turning vanes, as shown in Figure 6-15b.
Figure 6-15a
Figure 6-15b
Inlet Duct Connections Causing Inlet Spin3
Corrections for Inlet Spin3
6–29
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COMPUTING THE EFFECT OF ENCLOSURE RESTRICTIONS
In cases where a fan (or several fans) is built into a fan cabinet construction or installed in a
plenum, the walls should be at least one inlet diameter from the fan housing and a space of
at least two inlet diameters should be provided between fan inlets. If these recommendations cannot be met, additional pressure losses will result. These additional losses must be
added to the fan total pressure requirements as shown in Table 6-5. The additional pressure losses may also be calculated using Equation 6-11.
Table 6-5 K Factor for Enclosure Restrictions3
Length (L)
0.75  inlet diameter
0.50  inlet diameter
0.40  inlet diameter
0.30  inlet diameter
System Effect
Curves*
VW
U
T
S
Where D1 = diameter of the fan inlet
* See Table 6-3
COMPUTING THE EFFECT OF INLET AND OUTLET RESTRICTIONS
Normally, fan performance data do not include the effects of any accessories supplied with
the fan. The loss caused by fan accessories (such as bearings, bearing pedestals, inlet vanes,
inlet dampers, belt guards and motors) should be determined from tests by the fan manufacturer. The losses should be subtracted from the original fan performance and the resulting fan curve presented as the installed performance curve. If such data are not available,
the losses due to accessories may be estimated as explained below for inlet obstructions.
PARALLEL FAN OPERATION
The combined performance curve for two fans operating in parallel may be plotted by
using the appropriate pressure for the ordinates and the sum of the volumes for the abscissas. When two fans having a pressure reduction to the left of the peak pressure point are
operated in parallel, a fluctuating load condition may result if one of the fans operates to
the left of the peak static point on its performance curve. This problem may be reduced
using two fans on a single shaft.
The pressure curves (pt ) of a single fan and two identical fans operating in parallel are
shown in Figure 6-16. Curve A–A shows the pressure characteristics of a single fan. Curve
C–C is the combined performance of the two fans. The unique figure-8 shape is a plot of
all possible combinations of volume flow at each pressure value for the individual fans. All
6–30
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points to the right of CD are the result of each fan operating at the right of its peak point of
rating. Stable performance results for all systems with less obstruction to air flow than is
shown on the p curve D–D.
At points of operation to the left of CD, system requirements may be satisfied with one fan
operating at one rating point while the other fan is at a different rating point. For example,
consider p E–E, which requires a pressure of 1.0 in. wg and a volume of 5,000 cfm. The
requirements of this system can be satisfied with each fan delivering 2,500 cfm at 1.0 in.
wg, at Point CE. The system can also be satisfied at Point CE by one fan operating at 1,400
cfm at 0.9 in. wg, while the second fan delivers 3,400 cfm at the same 0.90 in. wg.
Note that system curve E–E passes through the combined performance curve at two
points. Under such conditions, unstable operation can result. Under conditions of CE,
one fan is underloaded and operating at poor efficiency. The other fan delivers most of the
system requirements and uses substantially more power than the underloaded fan. This
imbalance may reverse and shift the load from one fan to the other.
Figure 6-16
Two Forward Curve Centrifugal Fans in Parallel Operation
6–31
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6.5 Fan Controls
In many heating and ventilating systems, the volume of air handled by the fan varies. The
choice of the proper method for varying flow for any particular case is influenced by two
basic considerations: the frequency with which changes must be made; and the balancing
of reduced power consumption against increases in first-cost.
To control flow, the characteristic of either the system or the fan must be
changed. The system characteristic curve may be
altered by installing dampers or orifice plates. This
technique reduces flow by
increasing the system pressure required and, therefore, increases power consumption. Figure 6-17
shows three different system curves (A, B and C)
such as would be obtained
by changing the damper
setting or orifice diameter.
Dampers are usually the
lowest first-cost method of
achieving flow control; they
Figure 6-17 System Total Pressure Loss Curves
can be used even in cases
where essentially continuous control is needed.
However, a system effect loss is created even at the full-open position.
Changing the fan characteristic (pt curve) for control can reduce power consumption.
From the standpoint of power consumption, the most desirable control method is to vary
the fan speed to produce the desired performance. If the change is infrequent, belt-driven
units may be adjusted by changing the pulley on the fan’s drive motor. Variable speed
motors or variable speed drives (whether electrical or hydraulic) may be used when frequent or essentially continuous variations are desired. When speed control is used, the
revised pt curve can be calculated by the fan laws.
Inlet vane control is frequently used. Figure 6-18 illustrates the change in fan performance
with inlet vane control. Curves A, B, C, D and E are the pressure and power curves for various vane settings between wide open (A) and nearly closed (E).
6–32
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Figure 6-18
Effect of Inlet Vane Control on Backward Curve
Centrifugal Fab Performance
Tubeaxial and vaneaxial fans offer adjustable pitch blades to permit balancing of the fan
against the system or to make infrequent adjustments. Vaneaxial fans are also produced
with controllable pitch blades (pitch that can be varied while the fan is in operation) for
frequent or continuous adjustment. Varying pitch angle retains high efficiencies over a
wide range of conditions. Figure 6-19 shows the performance of a typical fan with variable
pitch blades. From the standpoint of noise, variable speed is somewhat better than variable
blade pitch. However, both control methods give high operating efficiency control and
generate much less noise than inlet vane or damper control.
Figure 6-19
Effect of Blade Pitch on Controllable Pitch
Vaneaxial Fan Performance
6–33
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6.6 Effect of Variable Resistance Devices
Variable resistance devices (such as dampers and louvers) can have significant effects on a
system. As discussed earlier, the system curve is a composite of several components in series
with each other. If one component varies, the system curve will also change. Some system
components are truly fixed, such as the ductwork. Others are variable, either by design or
operationally over time. Components that vary by design are referred to as varying with a
purpose and the others as varying without a purpose.
Examples of components
that vary without a purpose
are filters and coils. As
shown in Figure 6-20, dirty
filters will push the system
curve to the left, while dry
coils will push it to the
right. If a coil is not dehumidifying and becomes dry,
there will be less of a pressure drop, the system curve
will slide to the right, and
more air will be delivered.
As the coil begins to dehumidify, or take moisture
out, the pressure drop will
be greater and the system
curve will slide back up to
its original range.
Figure 6-21 shows two system curves for a variable
Figure 6-20 Fan Curve and System Curve
volume system. In this case,
the volume will be varied
with dampers at the terminal devices. The original operating point was Point X on system
Curve A. The thermostat in this system is activated, and causes the damper to close down.
The operating point now shifts to Point Y on system Curve B, which gives 75% of the previous volume flow and a higher operating pressure. Therefore, the damper has to function
to reduce the pressure to Point Z on the original system curve.
6–34
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If the flow rate is halved, as shown in
Figure 6-22, these dampers continue
to close down. The damper pressure
differential is now quite large and
can contribute to both noise and
operating flow instabilities. Consequently, it is usually necessary to
provide some type of capacity control at the fan. This will reduce the
effective pressure available at the
fan, and will keep the available system pressure at or near the original
system curve. On systems with a
minor variation between maximum
and minimum flow, designs may be
based on riding the fan curve. Note
that duct leakage is based on the
pressure of the system operating at
Point Y.
Figure 6-21
Figure 6-22
Variable Volume System at
3/4 Flow
Variable Volume System at
1/2 Flow
6–35
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The Next Step
Having learned how fans work and produce airstreams with static and dynamic pressure,
the next chapter covers ducts that distribute the air around the facility.
Summary
A fan is an air pump with rotating blades that creates an increase in static and velocity pressure. The two main types of fan are the centrifugal where the air enters the eye of a barrel
and is thrown radially out into the spiral scroll, and the axial fan.
The fan laws enable one to calculate fan performance with changes in rpm, fan size, and air
density. For a specific fan connected to a system, the volume rises with the rpm1, static
pressure with rpm2, and power input with rpm3.
A fan creates a velocity pressure and rise in static pressure in a system. Because the system
can, and usually does, influence fan performance, both velocity pressure and static pressure
must be addressed. The “easy” estimation of system static pressure loss and choosing a fan
with that static pressure rise may produce acceptable results on a low velocity system, but
probably will not on a higher velocity system.
The reason is that the inlet and outlet conditions can significantly influence fan performance, a phenomenon known as “system effect.”
Fan efficiency ratings are based on ideal conditions, a new fan, unobstructed inlet and
same-size duct outlet. However, in real installations, the designed inlet and outlet conditions are often not ideal. The reduction in fan performance due to inlet and outlet conditions can greatly reduce effective fan performance. To minimize the risk of error, designs
should be based on total fan pressure and not just on fan static pressure.
Direct drives are used on smaller systems where oversizing the fan is easier than matching
the fan to the load. On larger systems, belt drives are commonly used to adjust from the
motor speed to the required fan speed. Motors and drives must be sized for the maximum
anticipated load.
Selecting a fan involves finding one that provides the required flow and total pressure at a
good efficiency and noise level. The type of fan may be influenced by system operation, so
a very flat characteristic might be sought in a system where variations in flow are required
without any fan adjustment. The horsepower rises as the cube of the flow (fan law), so the
fan must not be significantly undersized. Equally, if a fan is significantly oversized, changing pulleys to reduce capacity to what is actually required will save substantial horsepower
and operating costs.
6–36
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Fan performance data are normally given for Standard Air. At constant speed, the power
and pressure vary with gas density, proportional to absolute temperature, R.
At altitudes significantly above sea level, the air density drops. A fan will provide the same
volume flow, but the same volume will transfer less thermal energy due to the lower air
density. The design volume will typically be higher at higher altitudes, so a design for sea
level operation should be reevaluated before being built at higher altitude.
Fan data are usually specified in terms of static pressure, not total pressure. Due to velocity
changes at inlet and outlet connections the use of static pressure alone can cause errors.
Particularly with system velocities over 1,500 fpm, it is important to work with total pressure not just static pressure. Remember, velocity pressure equals (V/4005)2 in. wg.
When installed, a fan may not provide the expected flow. This may be due to the system
pressure losses being different from design calculations or the fan is being affected by the
way the inlet and outlet are configured. A difference in system pressure loss will result in
the fan riding its curve until the system curve and fan curve meet. Correction may be possible by fan speed adjustment. Remember, if the reason is inlet or outlet configuration, this
is a system effect.
System effects are due to one, or more, of the following: outlet connection geometry;
uneven flow across the inlet; and swirl at inlet. The outlet effects are due to the uneven
velocity profile coming out of the fan and the difference between the fan air outlet size
(blast area) and the fan connection size. At the inlet, an uneven flow across the fan effectively overloads some blade positions and underloads other blade conditions, causing loss
of efficiency. Swirling of the entering air effectively changes the velocity of the air as it
meets the blade, again jeopardizing efficiency.
The flow at a fan discharge is very uneven and takes a length of duct to even out. The
required length, and static pressure loss, can be calculated based on manufacturers’ data of
blast area versus outlet area. If this duct length is not available, the pressure drop will be
increased. Care must be taken to transition from fan outlet to duct system with minimal
losses.
For centrifugal fans, a bend close to the fan outlet will cause an additional static pressure
loss, which must also be factored into total static pressure losses.
The way the air enters a fan can significantly influence fan performance and power consumption. If the entering airstream is biased to one side of the inlet, or is swirling, the fan
performance will be reduced and power may be increased. Careful analysis of these inlet
effects can be very important in ensuring that the system performs as required and that
energy is not wasted due to poor design.
6–37
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Bibliography
1. AMCA. 1985. Standard 210, Laboratory Methods of Testing Fans for Rating. Arlington Heights,
IL: Air Movement and Control Association Inc.
2. ASHRAE. 1999. ASHRAE Standard 51, Laboratory Methods of Testing Fans for Rating. Atlanta,
GA: ASHRAE.
3. AMCA. 1990. Fans and Systems. Arlington Heights, IL: Air Movement and Control Association
Inc. Publication 201-90.
ASHRAE HandbookFundamentals and HandbookSystems and Equipment
Skill Development Exercises for Chapter 6
Complete these questions by writing your answers on the worksheet at the back of this
book.
6-1.
A fan is delivering 6,000 cfm at a pressure of 1.5 in. wg at a rotational speed of
750 rpm. If the fan speed is reduced to 600 rpm, how much air will the fan
deliver, and at what pressure?
a) 4,800 cfm, 1.2 in. wg b) 4,800 cfm, 0.96 in. wg
c) 3,840 cfm, 0.96 in. wg d) 3,840 cfm, 1.2 in. wg.
e) None of the above
6-2.
Given a fan operating at 4,000 cfm, 3 in. wg total pressure, and 2.5 hp, what is
the fan total efficiency?
a) 85% b) 80% c) 75% d) All of the above e) None of the above
6-3.
Given a fan operating at 4,000 cfm, using 1.5 hp, what is the fan total efficiency?
a) 85% b) 75% c) 65% d) None of the above
6-4.
What is one effective duct length for a duct with a duct velocity of 4,000 fpm
and an area of 200 in.2?
a) 80 ft b) 3.3 ft c) 5.66 ft d) None of the above
e) Cannot be determined from the information given
6–38
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6-5.
What is one effective duct length for a duct with a duct velocity of 2,000 fpm
and an area of 225 in.2?
a) 3.5 ft b) 3.0 ft c) 52.3 ft d) None of the above
e) Cannot be determined from the information given
6-6.
A rectangular duct is 10 in. high and 20 in. wide. What is the equivalent duct
diameter of this duct?
a) 200 in.2 b) 254 in. c) 16 in. d) None of the above
e) Cannot be determined from the information given
6-7.
For any given system, the system effect factor is constant across the range of flow
volumes of the fan.
a) True b) False c) Cannot be determined from the information given
6-8.
A fixed fan system is drawing 3 hp to deliver 10,000 cfm. If the air flow requirement can be reduced to 7,000 cfm by decreasing the fan speed, the horsepower
requirement will be reduced to:
a) 2.1 hp b) 1.0 hp c) 0.44 hp d) All of the above e) None of the above
f) Cannot be determined from the information given
6-9.
The ___________________ is the highest efficiency centrifugal fan design.
a) Radial b) Forward-curved c) Backward-inclined, backward-curved
d) All of the above e) None of the above
6-10.
Power roof ventilators ___________ :
a) Usually operate without discharge ductwork b) Operate at low pressure
c) Operate at high volume d) All of the above e) None of the above
6–39
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Chapter 7
Duct System Design
Contents of Chapter 7
•
•
•
•
•
•
•
•
7.1 Duct System Design Overview
7.2 Duct Materials
7.3 Duct Construction
7.4 Duct Design and Sizing
7.5 Sample Systems
Summary
Bibliography
Skill Development Exercises for Chapter 7
7–1
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Study Objectives of Chapter 7
After completing this chapter, you should be able to:
•
Layout and size a simple duct system that will transport the required quantity
of air from the fan to the conditioned space using appropriate methods and
materials; and
•
Calculate the pressure losses in a duct system.
7.1 Duct System Design Overview
Air duct system design must consider: space availability; space air diffusion; noise; duct
leakage; duct heat gains and losses; balancing; fire and smoke control; and initial investment and system operating cost.
This chapter presents duct construction, system design considerations and calculating a
system’s frictional and dynamic resistance to air flow.
7.2 Duct Materials
A variety of materials are used in the construction of ducts. Selection of duct materials
should receive the same careful consideration as the other system components. The material used in a duct system can substantially affect the overall system performance. The
advantages and disadvantages of the available materials should be considered.
Materials used for ducts include: galvanized steel, carbon (black) steel, aluminum, stainless
steel, copper, fiberglass reinforced plastic (FRP), polyvinyl chloride (PVC), polyvinyl steel
(PVS), concrete, fibrous glass (duct board), and gypsum board. These materials are compared in Table 7-1.
Duct sizing and construction specifications are generally stated in terms of use of galvanized steel, and correction factors for other materials must be used. Unless otherwise
noted, this chapter will consider galvanized steel exclusively.
Consideration must also be given to selection of duct construction components other than
those materials used for the duct walls including: flexible ducts, duct liner, pressure sensitive tapes, sealants, reinforcements and hangers. Lined duct must be sized to include the
lining. The duct drawing must clearly state that the duct dimension is the metal size, or the
airway size.
7–2
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Table 7-1 Duct Materials1
Material
Applications
Galvanized Steel Widely used for most air handling applications. Not recommended for corrosive product handling or temperatures
above 400°F.
Carbon Steel
Breechings, flues, stacks, hoods, other
(Black Iron)
high temperature duct systems, kitchen
exhausystems, ducts requiring paint or
special coatings
Aluminum
Duct systems for moisture-laden air, louvers, special exhaust systems, ornamental
duct systems. Often substituted for galvanized steel in HVAC duct systems.
Stainless Steel
Duct systems for kitchen exhaust, moisture-laden air fume exhaust
Copper
Advantages
Limitations
High strength, rigidity, durability, rust resistance in ordinary
conditions, availability, nonporous, workability
High strength, rigidity, durability, availability, paintability,
weldability, non-porous
Weldability, paintability, weight, corrosion
resistance
Corrosion resistance,
weight
Weight, resistance to some forms Low strength, material
of corrosion, availability
cost, weldability, thermal expansion
High resistance to many common forms of corrosion (but care
is definitely required in alloy
selection)
Duct systems for exposure to outside ele- Accepts solder readily, durable,
ments and moisture-laden air
resists corrosion, non-magnetic
Fiberglass
Reinforced
Plastic (FRP)
Chemical exhaust, scrubbers, underground duct systems
Corrosion resistant, ease of modification
Polyvinyl
Chloride (PVC)
Polyvinyl
Steel (PVS)
Exhaust systems for chemical fumes and
hospitals, underground duct systems
Underground duct systems, moistureladen air, corrosive air systems
Corrosion resistance, weight,
weldability, ease of modification
Corrosion resistance, weight,
workability fabrication, rigidity
Concrete
Underground ducts, air shafts
Rigid Fibrous
Glass
Interior HVAC low-pressure duct systems
Compressive strength, corrosion
resistance (steel reinforcement in
concrete must be properly
treated)
Weight, thermal insulation and
vapor barrier, acoustical qualities,
ease of modification, inexpensive tooling for fabrication
Gypsum Board
Ceiling plenums, corridor ducts, airshafts
Cost, availability
Material cost, workability, availability
Cost, electrolytic
action if in contact
with galvanized steel,
thermal expansion,
stains
Cost, weight, range of
chemical and physical
properties, brittleness,
fabrication, code
acceptance
Cost, fabrication, code
acceptance
Susceptible to coating
damage, temperature
limitations (250°F
max.), weldability,
code acceptance
Cost, weight, porous,
fabrication (requires
forming processes)
Cost, susceptible to
damage, system pressure, code acceptance,
questionable cleanability
Weight, code acceptance, leakage, deterioration when damp
7–3
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7.3 Duct Construction
DUCT TYPES
Rectangular metal ducts. Table 7-2 lists construction requirements for rectangular steel
ducts. It shows that construction requirements are determined by duct thickness, duct
dimension and duct pressure. The table indicates what kind of reinforcing (if any) is
required for any given combination of these factors. Combinations of factors that are not
allowed are also indicated. For 4 in. wg and higher systems, this is for positive pressure
only. For negative pressure systems with internal duct wall supports, consult the
SMACNA Industrial Duct Construction standards.1
Reinforcing is indicated by a letter and number (for example, D-10). The letter indicates
the rigidity class of the required reinforcing, and the number indicates the spacing of the
reinforcements, in feet. Specifications for rigidity class and transverse joint reinforcement
are shown in Table 7-3.
Tables 7-2 and 7-3 are samples of the many tables detailing duct construction requirements. For additional detail, consult the ASHRAE Handbook–HVAC Systems and Equipment.2 The SMACNA publication HVAC Duct Construction Standards–Metal and Flexible
gives the functional criteria on which Tables 7-2 and 7-3 are based.3
Transverse joints and, when necessary, intermediate structural members are designed to
reinforce the duct system. Ducts larger than 96 in. require internal tie rods to maintain
their structural integrity. Tie rods allow the use of smaller reinforcements than would otherwise be required.
Fittings must be reinforced similarly to sections of straight duct. On size change fittings,
the greater fitting dimension determines material thickness. Where fitting curvature or
internal member attachments provide equivalent rigidity, such features may be credited as
reinforcement.
7–4
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Table 7-2 Rectangular Ferrous Metal Ducts for Commercial Systems
Table Notes
a  Table 7-2 is reproduced
from the 1992 ASHRAE
Handbook-Systems and Equipment, and is based on Tables
1-3 through 1-9 in HVAC
Duct Contruction StandardsMetal and Flexible (SMACNA
1985). For tie rod details, refer
to this standard.
b  For a given duct thickness,
numbers indicate maximum
spacing (feet) between duct
reinforcement; letters indicate
type (rigidity class) of duct
reinforcement.
Transverse joint spacing is
unrestricted on unreinforced
ducts. To qualify joints on
reinforced ducts, select transverse joints from Table 7-3.
Tables are based on steel construction. Designers should
specify galvanized, uncoated or
painted steel joint and intermediate reinforcement.
Use the same metal duct thickness on all duct sides. Evaluate
duct reinforcement on each
duct side separately. When
required on four sides, for +4,
+6 and +10 in. wg systems,
corners must be tied. When
required on two sides, corners
must be tied with rods or
angles at the ends for +4, +6
and +10 in. wg systems.
Duct sides over 18 in. width
with less than 0.0356 in.
thickness, which have more
than 10 ft2 of unbraced panel
area, must be cross-broken or
beaded, unless they are lined
or insulated externally. Lined
or externally insulated ducts
are not required to have crossbreaking or beading.
continued on next page
7–5
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Table 7-2 (cont.)
Table Notes, cont.
c  The reinforcement tables are
based on galvanized steel of the indicated thickness. They apply to galvanized, painted, uncoated and stainless
steel whenever the base metal thickness is not less than 0.0015 in. below
that indicated for galvanized steel.
d  Blank spaces indicate that no
reinforcement is required.
e  See SMACNA’s publication for
alternative reinforcements using tie
rods or tie straps for positive pressure.
f  Sheet metal 0.0466 in. thick is
acceptable.
g  Tie rods with a minimum diameter of 0.375 in. (or 0.25 in. if the
maximum length is 36 in.) must be
used on these constructions. The rods
for positive pressure ducts are spaced
a maximum of 60 in. apart along
joints and reinforcements.
7–6
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Table 7-3 Transverse Joint Reinforcement
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Pressure classification in relation to the fan curve must be considered, especially with VAV
systems where the dampers may throttle the air flow, raising the duct pressure. Manual balancing dampers may be inadvertently closed, with a resulting rise in system pressure. Supply ducts sometimes blow apart and return ducts sometimes collapse as a result of these
effects. Table 7-4 shows the SMACNA Duct Pressure Classification scheme.
Table 7-4 SMACNA Duct Pressure Classifications3
Static Pressure
Pressure Class Operating Pressure
0.5 in. wg
1 in. wg
2 in. wg
3 in. wg
4 in. wg
6 in. wg
10 in. wg
Up to 0.5 in. wg
Over 0.5 in. wg to 1 in. wg
Over 1 in. wg to 2 in. wg
Over 2 in. wg to 3 in. wg
Over 3 in. wg to 4 in. wg
Over 4 in. wg to 6 in. wg
Over 6 in. wg to 10 in. wg
Round metal ducts. Round ducts are inherently strong and rigid, and are generally the most
efficient and economical ducts for air systems. The dominant factor in round duct construction is the ability of the material to withstand the physical abuse of installation and
negative pressure requirements. Construction requirements are a function of static pressure, type of seam (spiral or longitudinal), and diameter.
Flat-oval ducts. Hanger designs and installation details for rectangular ducts generally apply
to flat-oval ducts.
Fibrous glass ducts. Fibrous glass ducts are a composite of rigid fiberglass and a factoryapplied facing (typically aluminum or reinforced aluminum), which serves as a finish and
vapor barrier. This material is available in molded round sections, or in board form for fabrication. Duct systems of round and rectangular fibrous glass are generally limited to 2,400
fpm and ±2 in. wg. Molded round ducts are available in higher pressure ratings.
Flexible ducts connect mixing boxes, light troffers, diffusers and other terminals to the air
distribution system. Because unnecessary length, offsetting and compression of these ducts
significantly increases air flow resistance, they should be kept as short as possible, and fully
extended.
For further information on fibrous glass ducts, consult the SMACNA publication Fibrous
Glass Duct Construction Standards4, or manufacturers’ construction standards.
7–8
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DUCT SEALING
Ducts must be sufficiently airtight to ensure economical and quiet performance of the system. Both leakage and the noise of leaks increase with increasing duct pressure.
A variety of materials and techniques have been developed for duct sealing including: liquids, mastics, gaskets, pressure sensitive tapes, heat-applied materials and embedded fabric.
Surfaces to receive sealant should be free from oil, dust, dirt, rust, moisture, ice crystals and
any other substances that would inhibit or prevent bonding.
It should be realized that no sealant system is recognized as a substitute for mechanical
joining. Also, the designer should carefully evaluate proposed duct sealants. Some use solvents that are toxic to workers applying the sealant. Some deteriorate or crystallize as they
dry, and do not provide adequate sealing only a few months after being installed.
7.4 Duct Design and Sizing
HVAC system duct design follows after the room loads and desired air quantities have
been determined. Consider the type of duct system needed, based on an economic analysis
of the building design and use, unless the owner or architect specifies a preference for a particular type. In any event, the specific type of system will affect the type of air handling
apparatus selected.
AIR DISTRIBUTION
First, locate the supply air outlets, and then select the size and type required for proper air
distribution in each conditioned space (refer to Chapter 3 of this course). Air distribution
in the conditioned space is highly important in influencing the comfort of the occupants.
Good air distribution is ensured by proper consideration of the basic factors in the selection of the outlet terminal devices. Drafts caused by too much air or physical flow disturbances within the room should be avoided.
The outlet terminal devices should provide the proper air velocities within the room’s
occupied zone (floor to 6 ft above the floor) and the proper temperature equalization.
Entrainment of the room air by the primary (or supply) airstream at the outlet terminal to
attain the required temperature equalization and to counteract the effects of natural room
air convection is very important.
Select air distribution terminal devices from industry standard types or configurations so
that they can be obtained from many sources. Most terminal device manufacturers’ catalogs furnish data on air flow throw, drop, air pattern, terminal velocities, acoustics, ceiling
heights, etc.
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Supply outlets on the same branch should be chosen with approximately the same pressure
loss (no more than 0.05 in. wg variation) through the outlet. Mixing ceiling supply diffusers with sidewall supply grilles on the same branch should be avoided unless there is no significant difference in pressure drops between the different types.
For a comprehensive review of considerations in the selection of air distribution equipment, refer to the ASHRAE Handbook–HVAC Systems and Equipment and to air distribution equipment manufacturers’ application engineering data. However, some of the basic
procedures used in the selection of air distribution equipment are:
•
Consider the ambient conditions that could affect comfort.
•
Decide on the location of air supply outlets (such as in the floor, sill, sidewall,
exposed duct or ceiling), taking into account the type of system serving them.
Locate return and exhaust air devices.
•
Consider the special requirements affecting outlets when used with systems
such as a variable air volume (VAV) system.
•
Place balancing dampers to be used with outlet devices at a convenient location, preferably well upstream from the outlet as long as access is available.
•
Refer to manufacturer’s data regarding throw, spread, drop, noise level, etc.
ZONING
With the outlet devices selected and before duct layout and duct sizing can begin, the
designer must determine how many zones of temperature control will be required for both
perimeter zones and interior zones. In general, the exterior zone will be divided into zones
that will be determined by building exposure (north, east, south or west exposure).
These perimeter zones may be further subdivided into smaller control zones, depending on
variations in internal load or a requirement for individual occupant control. Typical situations would include private executive offices, where the owner may want individual control, or areas of high heat gain or loss such as computer rooms, conference rooms or corner
rooms with two exposed walls.
Similarly, the interior zones may also be divided into control zones to satisfy individual
room requirements or variations created by internal loads, such as lights, people or equipment.
PRELIMINARY LAYOUT
The next step is to draw a preliminary schematic diagram for the ductwork that will convey
the design air quantity to the selected zones and outlets by the most efficient and economical path. This layout should be made on a reproducible tracing of the architectural floor-
7–10
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plans. By doing this, the designer will have a better feel for the final relationship of air terminals, branch ducts, main ducts, risers and apparatus. This procedure will help the
designer coordinate the ductwork with the structural limitations of the building and other
building systems and services.
On this preliminary layout, the designer should indicate the design air flows throughout
the system. If a constant volume system is chosen, it will be the arithmetic sum of the cfm
of each terminal (including branches) working back from the end of the longest run to the
fan. However, if a VAV system is chosen, the designer must apply the proper diversity factors to allow a summarization of the peak design air flows to determine their impact on
branch and main duct sizes coming from the supply fan.
The same procedure must also be followed for return air and exhaust air systems. This is to
size the ductwork properly, and to enable the designer to evaluate the effect of the total
HVAC system design, balancing the proper proportions of supply air to return air, exhaust
air and outside makeup air. Pressure losses due to fittings and transitions must also be
included in the calculation.
DUCT SIZING
Having completed the preliminary HVAC system duct layout, the designer will then proceed to use one of the methods for sizing the duct system discussed later in this chapter.
Generally, these methods will give the equivalent round duct sizes and the pressure losses
for the various elements of the duct system. The designer will then incorporate this information into the preliminary duct layout.
If round ductwork is to be used throughout, the duct sizing efforts are completed, providing the ductwork will physically fit into the building. If rectangular or flat oval ductwork is
chosen, the proper conversions must be made from the equivalent round duct sizes to rectangular or flat oval sizes. Applying the appropriate duct friction loss correction factors and
using the duct fitting loss coefficients, the duct system total pressure loss can be calculated.
With HVAC system duct sizes now selected and the total pressure or static pressure losses
calculated, the designer must determine if the ductwork will fit into the building. At this
point, the designer must consider the additional space required beyond the bare sheet
metal sizes for reinforcing and circumferential joints. In addition, consideration must be
given to external insulation or duct liner that may be required, clearance for piping, conduit, light fixtures, etc., where applicable, and clearance for the removal of ceiling tiles. A
further consideration in the sizing and routing of a ductwork system is the space and access
requirements for air terminals, mixing boxes, VAV boxes, fire and smoke dampers, balancing dampers, reheat coils and other accessories.
7–11
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DESIGN METHODS
No single design method will automatically provide the most economical duct system for
all conditions. A careful evaluation of all cost variables entering into a duct system should
be made with each design method or combination of methods. The cost variables to consider include the cost of the duct material (the aspect ratios are a large factor), duct insulation or lining (duct heat gain or loss), type of fittings, space requirements, fan power, balancing requirements, sound attenuation, air distribution terminal devices and heat
recovery equipment.
Slightly different duct system pressure losses can be obtained using the different design
methods. Some require a broad background of design knowledge and experience. The
careful use of these methods will allow the designer to efficiently size HVAC duct systems
for larger residences, institutional and commercial buildings, including some light industrial process ducts. Traditionally used duct design methods include the following:
•
•
•
•
•
•
Equal Friction
Static Regain
T-Method
Extended Plenums
Velocity Reduction
Constant Velocity
Equal friction (equal friction rate). The equal friction method of duct sizing (where the
pressure loss per foot of duct is the same for the entire system) is probably the most universally used means of sizing lower pressure supply air, return air and exhaust air duct systems.
It normally is not used for higher pressure systems. With supply air duct systems, this
design method “automatically” reduces air velocities in the direction of the air flow, thus
reducing the possibility of generating noise (against the air flow in return or exhaust duct
systems). The major disadvantage of the equal friction method is that there is no provision
for equalizing pressure drops in duct branches (except in symmetrical layouts). A manual
balance of short runs, to achieve the same pressure drop as a long branch run, is required.
The Friction Chart (Figure 7-1) has the pressure drop in in. wg per 100 ft, with the shaded
area indicating the suggested design limits. Many designers use 0.1 in. wg per 100 ft for
ductwork with no acoustic treatment. For systems with VAV boxes, which provide a measure of sound attenuation, 0.2 in. wg per 100 ft might be used from the supply from fan to
VAV boxes and dropping to 01 in. wg per 100 ft from VAV box to outlet.
Whatever equal friction choices are made, the data can be extracted from the Friction
Chart and tabulated to provide a quick reference to the data needed. The beginning of
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Friction Chart
Figure 7-1
7–13
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such a table is shown in Table 7-5. The reason for including velocity and velocity pressure
will become obvious when calculating the pressure drop through fittings in the ductwork.
Table 7-5 Sample Data for Duct Sizing*
Flow
(cfm)
Diameter
(in.)
Velocity
(fpm)
Velocity
Pressure
(in. wg)
50
100
200
5
6
7
480
580
630
0.35
0.38
0.40
* At 0.1 in. wg per 100 ft
The Friction Chart is for round duct, but often rectangular duct must be used. For equal
flow and pressure drop, the equivalent rectangular duct can be read from a table such as
Table 7-6. Note that the velocity will be lower in the equivalent rectangular duct.
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Table 7-6 Equivalent Round and Rectangular Duct Sizes
Circular Duct
Diameter (in.)
5
5.5
6
6.5
7
7.5
8
8.5
9
9.5
10
10.5
11
11.5
12
12.5
13
13.5
14
14.5
15
16
17
18
19
20
21
22
23
24
25
26
27
28
29
30
31
32
33
34
35
36
37
38
39
Length One Side Of Rectangular Duct (a), in.
4
5
6
8
9
11
13
15
17
20
22
25
29
32
5
5
6
7
8
10
11
13
15
17
19
21
23
26
29
32
35
38
6
6
7
8
9
10
12
13
15
16
18
20
22
24
27
29
32
35
38
45
7
8
9
10 12 14 16 18 20
Length Adjacent Side of Rectangular Duct (b), in.
7
8
9
10
11
12
14
15
17
18
20
22
24
26
28
30
36
41
47
54
8
9
10
12
13
14
15
17
18
20
22
24
25
30
34
39
44
50
57
64
9
10
11
12
13
15
16
17
19
20
22
25
29
33
38
43
48
54
60
66
10
11
12
13
14
15
17
18
19
22
25
29
33
37
41
46
51
57
63
69
76
12
13
14
15
16
18
20
23
26
29
33
36
40
44
49
54
59
64
70
76
82
89
96
14
15
17
19
22
24
27
30
33
36
40
44
48
52
56
61
66
71
76
82
88
95
101
108
16
17
19
21
23
26
28
31
34
37
40
43
47
51
55
59
64
68
73
78
83
89
95
18
19
20
23
25
27
29
32
35
38
41
44
47
51
54
58
62
67
71
76
80
20
22
24
26
28
31
33
36
39
41
44
48
51
54
58
62
66
70
22
24
22
24
26
28
30
32
35
37
40
42
45
48
51
55
58
62
24
25
27
29
31
34
36
38
41
44
46
49
52
55
7–15
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The whole business of reading off data from the Friction Chart and then another table to
obtain equivalent sizes can be done with a simple cardboard device called a Ductulator. By
rotating one sheet of the Ductulator, you can input two variables from volume, velocity,
round duct diameter or rectangular duct sides, pressure drop per 100 ft and read off the
other corresponding variables. For example, setting the pressure drop at 0.2 in.wg and volume at 4,000 fpm, you can read off duct diameter, the combinations of equivalent rectangular duct sides and duct velocity.
Static regain. The static regain method of duct sizing may be used to design supply air systems of any velocity or pressure. It normally is not used for return air systems where the air
flow is toward the HVAC unit fan. This method is more complex than the equal friction
method, but it is a theoretically sound method that meets the requirements of maintaining
uniform static pressure at all branches and outlets.
Duct velocities are systematically reduced, allowing a large portion of the velocity pressure
to convert to static pressure that offsets the friction loss in the succeeding section of duct.
The duct system will stay in balance because the losses and gains are proportional to a function of the velocities. This static regain, which is often assumed at 75% for average duct
systems, could be as low as 50% or as high as 100+% under ideal conditions. The assumed
regain factors can create installed systems that are quite different than the design requirements. The classical static regain method should not be used without a computer program
to make actual mass flow calculations at branches, due to the unpredictable regain factor.
A disadvantage of the static regain method is the oversized ducts that can occur at the ends
of long branches, especially if one duct run is unusually long. Often, the resultant very low
velocities require the installation of additional thermal insulation on that portion of the
duct system to prevent unreasonable duct heat gains or losses.
Note: The loss coefficients for duct fittings found in the ASHRAE Handbook–Fundamentals include static pressure regain or loss for the velocity condition changes that occur at
divided flow or change-of-size duct fittings.7 Additional duct static pressure regain (or loss)
must not be calculated and added to (or subtracted from) the total duct system pressure
losses when those fitting losses are used.
The Total Pressure Method is a further refinement of the static regain method that allows
the designer to determine the actual friction and dynamic losses at each section of the duct
system. The advantage is having the actual pressure losses of the duct sections and the fan
total pressure requirements provided.
T-method. The T-method of duct sizing is a comprehensive duct design optimization procedure that includes system initial costs and operating costs, energy costs, hours of operation, annual escalation, interest rates, etc. A description of the method and main procedures and equations may be found in the ASHRAE Handbook–Fundamentals chapter on
7–16
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duct design. The method requires computer software, and an extensive evaluation of
acoustic results.
Extended plenums. An extended plenum is a trunk duct (usually at the discharge of a fan,
fan-coil unit, mixing box, variable air volume box, etc.), extended as a plenum to serve
multiple outlets and/or branch ducts with essentially equal pressure.
A semi-extended plenum is a trunk duct system utilizing the concept of the extended plenum incorporating a minimum number of size reductions. This modification can be used
with equal friction and static regain design methods. Some of the advantages may be: lower
first-costs, lower operating costs, ease of balancing, and adaptability to branch duct or outlet changes. A disadvantage is that low air flow velocities could result in additional heat
gain or loss to the airstream through the duct walls.
Velocity reduction. In this method, a system velocity is selected at the section next to the fan
and arbitrary reductions in velocity are made after each branch or outlet. The resultant
pressure loss differences in the various sections of the duct system are not taken into
account and balancing is attempted mainly by the use of good dampers at strategic locations. An experienced designer who can use sound judgment in selecting arbitrary velocities may design a relatively simple duct system using the velocity reduction method. Other
practitioners should not attempt to use this method except for estimating purposes unless
the system has only a few outlets and can be easily balanced.
Constant velocity. With adequate experience, many designers are able to select an optimum
velocity that is used throughout the design of a duct system. This method is best adapted
to the higher pressure systems that use attenuated terminal boxes to reduce the velocity and
noise before distribution of the air to the occupied spaces. Industrial exhaust systems often
use the constant velocity method to ensure particulate movement along with the exhaust
airstream.
OTHER DESIGN CONSIDERATIONS
The amount of duct leakage in an HVAC system may be determined by the system
designer using data from the SMACNA HVAC Duct Construction Standards-Metal and
Flexible3 and the SMACNA HVAC Air Duct Leakage Test Manual.8 Leakage in ducts varies
with the fabricating machinery used, the methods of assembly, and the quality of the
installation workmanship, plus the effectiveness of any sealants, if used, and the workmanship in their application.
7–17
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A variety of sealed and unsealed duct leakage tests have confirmed that longitudinal seam,
transverse joint and assembled duct leakage can be represented by Equation 7-1, and that
for the same construction, leakage is not significantly different in the negative and positive
modes:
N
Q = C p S
(7-1)
where:
Q = leakage rate, cfm
C = constant reflecting area characteristics of leakage path
ps = static pressure differential from duct interior to exterior, in. wg
N = exponent relating turbulence or laminar flow in leakage path
Analysis of the AISI/ASHRAE/SMACNA/TIMA data resulted in the categorization of
duct systems into a leakage class, CL, the accepted value of N = 0.65, and Q now defined in
terms of surface area of the duct:
0.65
Q = C L  pS
(7-2)
where:
Q
= Leakage rate per unit surface area, cfm/100 ft2
CL = Leakage class, cfm per 100 ft2 duct surface at 1 in. wg static pressure
Figure 7-2 shows how duct pressure affects the leakage rate for each leakage class.
ASHRAE Standard 90.1 prescribes minimum sealing requirements for supply, return, and
exhaust ducts run outside, in conditioned spaces and unconditioned spaces.5 Specifying
allowable leakage rates of less than CL3 should be avoided due to both cost and difficulty.
Leakage class is defined as the cfm leaked per 100 ft2 of duct surface area at 1 in. wg.
A selected series of leakage classes based on Equation 7-2 is shown in Figure 7-2. Table 7-7
is a summary of the leakage class attainable for good duct construction and sealing practices. Note that connections of ducts to grilles, diffusers and registers are not represented in
the test data. The HVAC system designer is responsible for assigning acceptable leakage
rates.
7–18
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Figure 7-2
Duct Leakage Classifications1
Table 7-7 Applicable Leakage Classes1
Duct Class
Applicable
Sealing
Rectangular
Metal
Round and
Oval Metal
Rectangular
Fibrous Glass
Round
Fibrous Glass
0.5, 1, 2 in.wg
3 in. wg
4, 6, 10 in. wg
N/A
Transverse
Joints Only
Transverse
Joints and
Seams
All Joints, Seams
and Wall
Penetrations
48
24
12
6
30
12
6
3
N/A
6
N/A
N/A
N/A
3
N/A
N/A
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EXAMPLE 7.1
Question: Given the system shown below, with average pressure of 2.5 in. wg, find the leakage, in cfm, of the supply ductwork between points A and B.
Answer: From Table 7-6, it is determined that the leakage class for a 3 in. wg duct class
round metal duct is 6. Using Figure 7-2, the leakage factor is determined to be 10.6
cfm/100 ft2. The 30 in. diameter duct from A to B has 785 ft2 of duct surface, and the
leakage is determined using Equation 7-2:
2
10.6 cfm
Leakage = ---------------------  785 ft = 83 cfm
2
100 ft
(7-3)
DUCT HEAT GAIN OR LOSS
At the beginning of this chapter, it was stated that duct design follows building load calculations. An often overlooked factor in load calculations is duct heat gain or loss. The
method of calculating this load is well described in other texts, such as the ASHRAE Handbook–Fundamentals. In this section, some of the practical considerations in duct design
that affect duct heat gain or loss are noted.
Consider first a conditioned air supply system with the air handling apparatus and
ductwork in the conditioned space, and with no additional load imposed on the system.
However, if the ductwork is long and the velocities are low, the designer should check that
air flows are proportioned properly. The air in the ductwork still gets warmer or cooler as it
passes through the conditioned space, thus decreasing the temperature difference. As a
result, less air is required to supply the outlets at the start of the supply run and more is
required at the end.
Naturally, when a duct or plenum carrying conditioned air is located outside the conditioned space, the heat gain or loss must be accounted for in both the design air quantity
and total sensible load. This system load must be calculated by the designer when running
7–20
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conditioned air ductwork through boiler rooms, attics, outdoors or other unconditioned
spaces. Alternate routing might be more desirable than increasing the system load.
With certain exceptions, ASHRAE Standard 90.1 requires thermal insulation of all duct
systems and their components (such as ducts, plenums and enclosures) installed in or on
buildings.5 To estimate duct heat transfer and entering or leaving air temperatures, use
Equations 7-4, 7-5 and 7-6:
 t e + t 1
Q 1 = UPL
- – t
-----------  -------------12  2  a
(7-4)
t 1  y + 1  – 2t a
t e = ---------------------------------- y – 1
(7-5)
t e  y – 1  + 2t a
t 1 = --------------------------------- y + 1
(7-6)
where:
y = 2.4 AV/UPL for rectangular ducts = 0.6 DV/UL for round ducts
A = cross-sectional area of duct, in.2
V = average velocity, fpm
D = diameter of duct, in.
L = duct length, ft
Ql = heat loss/gain through duct walls
U = overall heat transfer coefficient of duct wall, Btu/h ft2 F
P = perimeter of bare or insulated duct, in.
 = density, lbm /ft3
te = temperature of air entering duct, °F
tl = temperature of air leaving duct, °F
ta = temperature of air surrounding duct, °F
Use Figure 7-3 to determine the U-values for insulated and uninsulated ducts. For a 2 in.
thick, 0.75 lb/ft3 fibrous glass blanket compressed 50% during installation, the heat transfer rate increases approximately 20%, as shown in Figure 7-3. Pervious flexible duct liners
also influence heat transfer significantly, as shown in Figure 7-3. At 2,500 fpm, the pervious liner U-value is 0.33 Btu/h ft2  F. For an impervious liner, the U-value is 0.19
Btu/h ft2 F.
7–21
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7–22
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Duct System Design
Heat Transfer Coefficients
Figure 7-3
SDL_AirSystemsDesign_IP.book Page 23 Wednesday, August 16, 2017 2:45 PM
EXAMPLE 7.2
Example 7.2 is adapted from SMACNA HVAC Systems Duct Design, p. 5.27–28.
Question: A 65 ft length of 24 in.  36 in. uninsulated sheet metal duct, freely suspended,
conveys heated air through a space maintained above freezing at 40°F. Based on heat loss
calculations for the heated zone, 17,200 cfm of standard air at a supply air temperature of
122°F is required. The duct is connected directly to the heated zone. Determine the
required air temperature entering the duct, and the duct heat loss.
Answer:
a. Calculate the duct velocity:
cfm
17 200 cfm
V = -------- = ----------------------------------------------------------------------------- = 2900 fpm
2
2
A
 24 in.  36 in.    144 in.  ft 
Select U = 0.73 Btu/h ft2 F (from Figure 7-3)
Calculate P = 2(24 in. + 36 in.) = 120 in.
3
 2.4   24 in.   36 in.   2900 fpm   0.075 lb/ft 
y =  2.4 AV    UPL  = ---------------------------------------------------------------------------------------------------------------- = 79.2
 0.73   120 in.   65 ft 
b. Calculate the entering air temperature:
122F  79.2 + 1  –  2  40F 
t e = ------------------------------------------------------------------------ = 124.1F
 79.2 – 1 
c. Calculate the duct heat loss:
 0.73   120 in.   65 ft 
124.1F + 122F
Q 1 = ------------------------------------------------------   ----------------------------------------- – 40F = 39 200 Btu/h


12
2
FITTING LOSSES
Pressure loss at fittings is a critical element of duct system design. The simplest way to
incorporate fitting losses into the design is to use loss coefficients taken from the ASHRAE
Duct Fitting Database tables such as the ones found in the ASHRAE Handbook–Fundamentals. These loss coefficients represent the ratio of total pressure loss to the dynamic
7–23
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pressure (in terms of velocity pressure). They do not include duct friction loss (which is
picked up by measuring the length of duct sections to fitting center lines). However, the
loss coefficients do include static regain (or loss) where there is a change in velocity.
The total pressure (pt) loss of a fitting is determined using the loss coefficient in the following equation:
 pt = C o  pv
(7-7)
where:
pt = total pressure loss (in. wg)
Co = Dimensionless local loss coefficient
Pv = Velocity pressure (in. wg)
By using duct fitting loss coefficients that include static pressure regain or loss, accurate
duct system fitting pressure losses are obtained. When combined with the friction losses of
the straight duct sections sized by the modified equal friction method, the result will be the
closest possible approximation of the actual system total pressure requirements for the fan
EXAMPLE 7.3
Question: To demonstrate the use of the loss coefficient tables, assume a velocity of 2,500
fpm in a 9 in. 7 Gore (segment), 90° elbow, as shown in Figure 7-4. According to the Figure 7-4 table, Co for this fitting is 0.10.
Figure 7-4
CD3-10 Elbow
CO Values for CD3-10 Elbow
D (in.)
Co
3
0.16
6
0.12
* 7 Gore, 90 degree, r/D = 2.5
7–24
9
0.10
12
0.08
15
0.07
18
0.06
27
0.05
60
0.03
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Answer: Using Equation 1-8, we determine that the velocity pressure is 0.39 in. wg:
V 2
2500 2
p v =  ------------  =  ------------  = 0.39 in. wg
 4005 
 4005 
Using Equation 7-7, we determine that the total pressure loss is 0.039 in. wg:
 p t = C o  p v = 0.10  0.39 = 0.039 in. wg
7.5 Sample Systems
The following two simplified sample systems illustrate how to calculate pressure drop in
“real” systems.
SYSTEM 1
As depicted in Figure 7-5, this system consists of a fan, a straight length of galvanized steel
duct and an outlet. The duct is the same size as the fan outlet, so no system effect factor
needs to be added. The outlet is a VAV box with a 1 in. wg pressure drop at 4,000 cfm.
The fan speed will be adjusted to deliver 4,000 cfm. What is the pressure drop in the system?
Figure 7-5
System 1
7–25
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Solution:
1. Calculate the circular equivalent of the rectangular duct using the formula or use Table
7-6.
  ab  0.625 
D e = 1.30  ---------------------------
  a + b  0.250
= 12.9
where:
De = circular equivalent of rectangular duct for equal length, fluid resistance,
and air flow, in.
a = length of one side of duct, in.
b = length of adjacent side of duct, in.
2. Calculate the velocity of the airstream in the duct:
4000 - = 4114 fpm
---------------------- 14
------  10
------
 12 12
2
2
 De 
12.9 

A e =   ------------------ =  ------------------ = 0.9076
 4  144
 4  144
Q
4000 cfm
V = ----- = ------------------------ = 4407 fpm
2
Ae
0.9076 ft
3. Calculate the fan outlet velocity pressure from Equation 1-8:
 V 
p v =  ------------
 4005
2
 4407
=  ------------
 4005
2
= 1.211 in. wg
4. Find the pressure drop in the duct per 100 ft. using the Friction Chart shown in Figure
7-1 (also in the ASHRAE Handbook–Fundamentals):
1.9 in. wg per 100 ft
5. Calculate the pressure drop in the 30 ft run of duct:
30 ft-  1.9 in. wg = 0.57 in. wg
------------100 ft
7–26
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6. Add the pressure drops:
Pressure drop in 30 ft duct run
0.57 in.wg
Fan outlet velocity pressure
1.21 in.wg
Required VAV outlet pressure drop 1.00 in. wg
Total pressure drop required at fan
2.78 in. wg
SYSTEM 2
Figure 7-6 is the same system as in Figure 7-5, except that the outlet at the end of the duct
run was removed, and a 20° 14  10 in. to 24  10 in. rectangular transition was added.
Attached to the transition outlet are two 12  10 in. elbows with r = 1.5 (Co = 0.2).
Attached to each elbow is a branch duct. One branch is 17 ft long, the other branch is 12 ft
long. There is a balancing damper in the 12 ft branch. At each end of the duct extension is
a VAV box with a 1 in. wg pressure drop at 2,000 cfm. What is the pressure drop in the
system, and how should the balancing damper be adjusted to equalize the pressure in both
branch runs?
Figure 7-6
System 2
7–27
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Solution:
1. Calculate the pressure drop, velocity and pressure at the end of the 30 ft run. Because
this is the same configuration as System 1, on the main 30 ft run, the pressure drop is 0.57,
the velocity is 4,407 fpm, and the velocity pressure is 1.211 in., the same as in System 1.
2. Calculate the pressure drop in the transition.
a. Calculate the ratio of the inlet area to the outlet area:
Inlet area = 14 in.  10 in. = 140 in.2
Outlet area = 24 in.  10 in. = 240 in.2
Outlet area
240
-------------------------- = --------- = 1.7
Inlet area
140
b. Refer to the table below (from the ASHRAE Handbook–Fundamentals), which
gives Co values for rectangular transitions. Because the table does not give an
exact value for an outlet/inlet ratio of 1.71, by interpolation, the Co value is
estimated to be 0.43. Multiply the pressure at the inlet by the Co value to
calculate the pressure drop across the transition:
1.211 in. wg  0.43 = 0.52 in. wg
7–28
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SR4-1 Rectangular Transition*
Co Values
Ao/A1
10
15
20

45
60
90
120
150
180
0.07
0.05
0.06
0.06
0.00
1.40
9.60
69.00
181.76
0.08
0.07
0.07
0.07
0.00
1.48
10.88
82.00
220.16
0.19
0.18
0.17
0.14
0.00
1.52
11.20
93.00
256.00
0.29
0.28
0.27
0.20
0.00
1.48
11.04
93.00
253.44
0.37
0.36
0.35
0.26
0.00
1.44
10.72
92.00
250.88
0.43
0.42
0.41
0.27
1.00
1.40
10.56
91.00
250.88
30
0.10 0.05 0.05 0.05 0.05
0.17 0.05 0.04 0.04 0.04
0.25 0.05 0.04 0.04 0.04
0.50 0.06 0.05 0.05 0.05
1.00 0.00 0.00 0.00 0.00
2.00 0.56 0.52 0.60 0.96
4.00 2.72 3.04 3.52 6.72
10.00 24.00 26.00 36.00 53.00
16.00 66.56 69.12 102.40 143.36
*Two sides parallel, symmetrical. supply air systems
3. Calculate the pressure drop in the elbows. The calculation for each elbow is the same.
The velocity of the airstream entering each elbow is:
Q
4000
V = ---- = -------------------------------------------- = 2400 fpm
A
 2   12  10  144 
a. Convert the 12 in. x 10 in. rectangular duct to circular measurement (or use
the table of equivalents found in the ASHRAE Handbook–Fundamentals):
0.625
0.625
 ab 
 12  10 
D e = 1.30 ------------------------ = 1.30 --------------------------------- = 12.0
0.25
0.25
a + b
 12 + 10 
b. Calculate the velocity pressure in the elbows:
 2400
------------
 4005
2
= 0.359 in. wg
c. Multiply the pressure at the inlet by the Co value (given in the problem
statement as 0.2) to calculate the pressure drop across each elbow:
0.359 in. wg  0.2 = 0.072 in. wg
4. Calculate the pressure drop in the 17 ft branch, the longest branch.
7–29
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a. Find the pressure drop in the duct per 100 ft. using the Friction Chart in
Figure 7-1: 0.6 in. wg per 100 ft
b. Calculate the pressure drop for the 17 ft branch run:
17 ft
--------------  0.6 in. wg = 0.102 in. wg
100 ft
5. The loss of the VAV box is given as 1.0 in. wg (which is assumed to include the pressure
losses downstream of the box). Also note that the duct size and the box inlet are the same
size. If this is not the case, then there would be losses or gains depending on whether the
inlet is smaller or larger than the branch duct. If the inlet is smaller, there would be an
additional loss due to increasing the velocity, which is equal to the difference in velocity
pressures, which must be included.
6. The balancing damper in the 12 ft branch should be throttled so that the pressure drop
in the 12 ft branch is equal to the pressure drop in the 17 ft branch (0.102 in. wg). To calculate the pressure drop required across the damper:
a. Calculate the pressure drop in the 12 ft branch run without a damper:
12 ft
--------------  0.7 in. wg = 0.084 in. wg
100 ft
b. Subtract the pressure drop in the 12 ft branch run from the pressure drop in
the 17 ft branch run:
0.119 in. wg – 0.084 in. wg = 0.035 in. wg
c. Therefore the balancing damper must be adjusted to obtain a 0.035 in. wg
pressure drop to obtain equal flow in each branch.
7. Add the pressure requirements:
7–30
Pressure drop in 30 ft duct run
Pressure drop at transition
12 in.  10 in. elbow
17 ft branch duct
Required VAV outlet pressure
Fan outlet velocity pressure
0.57 in. wg
0.52 in. wg
0.07 in. wg
0.10 in. wg
1.00 in. wg
1.21 in. wg
Total pressure drop
3.47 in. wg
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The Next Step
The next chapter deals with codes and standards that are relevant for air system design and
energy usage.
Summary
Air duct system design must consider: space availability; space air diffusion; noise; duct
leakage; duct heat gains and losses; balancing; fire and smoke control; and initial plus operating costs.
Many materials are used for ductwork, but the vast majority is galvanized steel. For this
reason, duct design information is for galvanized steel, with corrections for other materials.
Other materials offer better chemical, moisture, acoustic and high-temperature performance typically at a premium cost.
Lined duct must be sized to include the lining. The duct drawing must clearly state that
the duct dimension is the metal size, or the airway size.
Rectangular metal ducts are manufactured to SMACNA standard specifications for size,
static pressure (positive or negative), material thickness, jointing, reinforcing and supports.
When choosing the pressure rating, take care to allow for probable maximum and minimum pressures on all but the smallest systems.
Round and oval metal ducts are inherently strong and rigid, and are generally the most
efficient and economical ducts for air systems. However, their shape may not fit the available access.
Fibrous glass ducts are a composite of faced rigid fiberglass available in molded round sections, or in board form for fabrication. Duty is generally limited to 2,400 fpm and ±2 in.
wg.
Flexible ducts are typically manufactured from a coiled wire and fabric, and are used for
connection of other ducts to diffusers.
Ducts must be sufficiently airtight to ensure economical and quiet performance of the system. Leakage classifications are given in cfm/100ft2 at 1 in. wg. Actual leakage is Q =
CLp0.65. Many materials, gaskets and tapes are available, but many have an unfortunately
short life. ASHRAE Standard 90.1 prescribes minimum sealing requirements for many
duct situations.
Once the room loads have been calculated and temperature difference chosen, the air volumes to each room can be calculated. Depending on the duct insulation and the tempera-
7–31
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ture of the space the duct runs through, there will be some heat gain or loss which should
be approximately included at this stage. For designs to meet Standard 90.1, the minimum
insulation values for energy conservation must be met.
Once the air volume to the room, room layout and architectural features and requirements
are known, a preliminary layout for outlets is made (as discussed in Chapter 3). Generally,
all outlets on the same branch duct should have the same pressure drop, particularly if they
are of different types.
Due to variations in loads, the HVAC system will be zoned. Typically, interior and exterior spaces will be on separate zones, and the duct layout must accommodate the zoning
and associated air control devices.
Next the preliminary layout is drawn, ideally over the architectural layout for supply,
return and exhaust ducts. Some very preliminary sizing will be done at this stage to ensure
adequate space for the main duct runs. At this stage, extra space for duct joints and insulation must be included, as well as allowing for the other services. The need to accommodate
all three may significantly influence the final choice of layout when crossovers and available space are evaluated.
With the preliminary HVAC system duct layout done, accurate duct sizing must be undertaken either with a computer program or manually. Sizing is more straightforward if all
ducts are round, as any rectangular ducts must be converted to equivalent round size for
calculating the resistances.
Once the ducts are sized, the final calculation of system pressure drop and location of all
outlets, control items, fire and smoke dampers can be fixed.
Duct design is somewhat of an art. There is a choice of design methods; technical design
must be balanced with cost and ease of installation and balancing. Slightly different pressure losses are obtained using different design methods and source data and these will often
be changed somewhat as the installation contractor deals with coordination with other
trades.
With the equal friction method, a fixed pressure drop per 100 feet is chosen and used to do
the duct sizing. This method is simple and decreases the velocity towards outlets, which
provides quiet systems. Care must be taken to avoid very unequal branch resistances which
can cause significant energy waste and noise due to damper pressure drops.
With the static regain method, the velocity pressure is systematically reduced to offset the
prior duct run pressure drop. The method is not easy to do manually and may need to be
modified in cases where branch ducts are very different in length.
7–32
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The T-method is an optimization procedure, ideally run in a computer program, that
designs on the basis of finding the most economic design based on initial costs and operating costs.
An extended plenum is a trunk duct maintained at full size to provide a relatively equal
supply pressure to each branch. A variation  semi-extended plenum  keeps the duct size
up for a greater length than necessary often reducing the cost of numerous size reductions.
In the velocity reduction method, velocity reductions are chosen by experienced designer.
It can be considered as the constant pressure drop method improved by experience.
For the experienced designer, a constant velocity can be chosen for sizing, especially where
noise is not an issue or where all outlets include sound attenuation.
Industrial exhaust systems often use constant velocity sizing to ensure particulate movement along with the exhaust airstream.
Bibliography
1. SMACNA. 2004. Rectangular Industrial Duct Construction Standards. Chantilly, VA: Sheet
Metal and Air Conditioning Contractors' National Association Inc.
2. SMACNA. 1990. HVAC Systems–Duct Design. Chantilly, VA: Sheet Metal and Air
Conditioning Contractors' National Association Inc.
3. SMACNA. 2005. HVAC Duct Construction Standards–Metal and Flexible. Chantilly, VA: Sheet
Metal and Air Conditioning Contractors National Association Inc.
4. SMACNA. 2003. Fibrous Glass Duct Construction Standards. Chantilly, VA: Sheet Metal and Air
Conditioning Contractors' National Association Inc.
5. ASHRAE. 2004. ASHRAE/IESNA Standard 90.1-2004, Energy Efficient Design of New Buildings
Except Low-Rise Residential Buildings. Atlanta, GA: ASHRAE.
6. SMACNA. 1985. HVAC Air Duct Leakage Test Manual. Chantilly, VA: Sheet Metal and Air
Conditioning Contractors' National Association Inc.
ASHRAE HandbookHVAC Systems and Equipment, duct construction; HandbookFundamentals,
duct design and sizing.
7–33
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Skill Development Exercises for Chapter 7
Complete these questions by writing your answers on the worksheets at the back of this
book.
7–34
7-1.
As depicted in the figure below, this system consists of a fan, ductwork and outlets. The duct is the same size as the fan outlet, so no system effect factor needs
to be added. The outlets are VAV boxes with a 1 in. wg pressure drop at 2,500
cfm. The fan speed will be adjusted to deliver 5,000 cfm. The Co value of the
elbow is 0.2. What is the total pressure drop in the system?
a) 3.2 in. wg b) 1.8 in. wg c) 1.6 in. wg d) None of the above
7-2.
Air duct system design must consider:
a) Noise b) Duct leakage, heat gains and heat losses
c) Fire and smoke control d) All of the above e) None of the above
7-3.
Duct sizing and construction specifications are generally stated in terms of the
use of:
a) Galvanized steel b) Aluminum c) Fiberglass reinforced plastic
d) All of the above e) None of the above
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7-4.
Generally the most efficient and economical ducts for air systems are:
a) Rectangular b) Oval c) Round
d) All of the above e) None of the above
7-5.
Duct systems of rectangular fibrous glass are generally limited to:
a) 2,400 fpm and ±2 in. wg b) 4,000 fpm and ±3 in. wg
c) 1,000 fpm and ±3 in. wg d) None of the above
7-6.
Compression of flexible ducts significantly decreases air flow resistance.
a) True b) False
c) Cannot be determined from the information given
7-7.
Sealant systems have been developed that can substitute for mechanical joining
of ductwork.
a) True b) False
7–35
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Chapter 8
Codes and Standards
Contents of Chapter 8
•
•
•
•
•
•
•
•
8.1 Building Code Requirements
8.2 ASHRAE Standard 90.1-2007
8.3 ASHRAE Standard 62.1-2007
8.4 Other Codes and Standards
8.5 Sources of Information
Summary
Bibliography
Skill Development Exercises for Chapter 8
8–1
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Study Objectives of Chapter 8
After completing this chapter, you should be able to list the principle codes and standards
affecting air system design, and briefly state what they cover and why they are important:
•
•
•
•
•
•
ASHRAE Standard 90.1-2007
ASHRAE Standard 62–20047
NFPA 90A
NFPA 90B
NFPA 96
SMACNA HVAC Duct Construction Standards
8.1 Building Code Requirements
In the private sector, each new construction or renovation project is normally governed by
state laws or local ordinances that require compliance with specific health, safety, property
protection and energy conservation regulations. Figure 8-1 depicts relationships between
laws, ordinances, codes and standards that can affect the design and construction of
HVAC duct systems.
However, Figure 8-1 may not list all applicable regulations and standards for a specific
locality. Specifications for federal government construction are promulgated by such agencies as the Federal Construction Council, General Services Administration, Department of
Defense, Department of Energy, and by Executive Orders.
Model code changes require long cycles for approval by the consensus process. Because the
development of safety codes, energy codes and standards proceed independently, the most
recent edition of a code or standard may not have been adopted by a local jurisdiction.
HVAC designers must know which code compliance obligations affect their designs. If a
provision conflicts with the design intent, the designer should resolve the issue with local
building officials. New or different construction methods can be accommodated by the
provisions for equivalency that are incorporated into codes. Staff engineers from the model
code agencies are available to help resolve conflicts, ambiguities and equivalencies.
8–2
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Hierarchy of Building Codes and Standards
Figure 8-1
8–3
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8.2 ASHRAE Standard 90.1-2007
Codes and standards have become much more important. With the substantial increase in
energy demands, many codes now also incorporate a minimum energy performance
requirement. Specifically, many model codes in the United States reference ASHRAE Standard 90.1-2007, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings.1
Originally drafted in 1975, ASHRAE Standard 90 was revised and reissued in 1980, 1989,
1999, 2004 and 2007. The original standard dealt with all buildings, but it was split into
90.1 for all but low-rise residential buildings and 90.2 for low-rise residential buildings.
Standard 90.1 is referenced in the Energy Act of 1992 and has been revised into code language to make it code enforceable. The standard is on ANSI continuous maintenance;
addenda are issued for review when ready and approved when they have passed the public
review process. To assist with code enforcement, the standard is reprinted with all addenda
every three years, with the latest printed revision in 2007.
The original Standard 90 was very important because it was one of the first documents that
truly addressed what can be done in the design of buildings to conserve energy. It went
through an extensive review, and was commented on by thousands of engineers across the
country. Much of the information in Standard 90.1 has been adopted by model building
codes.
The standard’s format is intended to be general and flexible, so it may be applied to many
different climates, building types and HVAC system types. The standard deals with all
aspects of building energy use including the Building Envelope in Section 5, Service Water
Heating in Section 7, Power in Section 8, and Lighting in Section 9. The most relevant
section for this course is Section 6, Heating, Ventilating and Air-Conditioning, although
choices made in the other sections will affect the air system choice, sizing and zoning.
HVAC systems are one of the most significant energy users in the buildings covered by
Standard 90.1. However, the designer has significant latitude in the energy costs and consumption of HVAC systems; a poorly designed system can easily have twice the annual
energy costs of an energy-conserving design.
Analyzing the energy use and cost of an HVAC system is complicated by system interactions. An efficient system is not merely characterized as one that uses efficient equipment.
System level efficiency must account for installation, control, maintenance, system losses
and component interactions (such as reheat or heat recovery).
As a conceptual model, overall HVAC system efficiency may be defined as the ratio of
loads the system must handle (space heating and cooling as well as water heating) to the
energy the system consumes.
8–4
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An efficient system will minimize energy use by minimizing system losses, maximizing
equipment efficiencies, using free heating/cooling, and recovering heat where possible. A
very efficient system could have an overall efficiency greater than one.
Section 6 addresses the following fundamental factors of system efficiency:
•
Specifying minimum equipment efficiencies
•
Reducing system losses from ductwork through sealing and insulation
•
Reducing system losses from piping through insulation
•
Reducing system operation through the use of automatic time controls and
zone isolation
•
Reducing system inefficiencies by minimizing simultaneous heating and cooling
•
Reducing system inefficiencies by shutting off outdoor ventilation during setback and warm-up
•
Reducing system operation through requirements for zone controls
•
Reducing system inefficiencies by limiting equipment oversizing
•
Reducing distribution losses, limiting HVAC fan energy demand and requiring efficient balancing practices
•
Requiring systems to take advantage of cool weather to provide free cooling
•
Requiring energy recovery on systems over 5,000 cfm and 70% outside air
Although compliance with Section 6 assures a minimum level of HVAC system performance, designers are encouraged to view the requirements as a starting point and investigate designs that exceed these requirements. Careful design and application of heat recovery, solar energy or high efficiency equipment can create systems that are more efficient
than the standard requires, and offer excellent returns on investment. The process of lifecycle costing is used to determine that proposed alternates have an economic payback.
COMPLIANCE METHODS
There are three primary subsections in Section 6. First, there is a simplified approach for
smaller buildings with simple HVAC systems. Then there are mandatory requirements in
Section 6.4 that must be met for either compliance path. Lastly, the prescriptive requirements in Section 6.5 include measures that must be met to show compliance via the prescriptive method. In this prescriptive method, the designer must choose equipment with
required performance and obey a number of design requirements.
8–5
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These prescriptive requirements do not have to be met with the energy cost budget
method, which is detailed in Section 11. In the energy cost method, the building designers
must show that their design would have no greater energy cost than a building designed
under the prescriptive route.
Many of the Section 6 requirements apply to larger, multiple zone systems. The breadth of
this section may seem overwhelming to designers of simpler, single zone HVAC systems
that are typically used in one- or two-story buildings under 25,000 ft2.
8.3 ASHRAE Standard 62.1-2007
ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality, was briefly
introduced in Chapter 3 as the standard that sets minimum outside air ventilation rates
and requirements for exhaust. The standard also sets requirements to provide acceptable
indoor air quality during the building’s lifetime, and it requires documentation of the
design assumptions and that they are available for the system’s operation.
The standard includes requirements in the system planning that deal with the following
questions:
•
How much outside air is required in each space?
•
How will the differing requirements for each space be achieved?
•
When VAV systems are used, how will the required ventilation air volume be
maintained when the supply volume to a space is reduced?
•
How effectively is the ventilation air distributed to the occupants in the space?
•
What quality of air can be recirculated from one space to another space?
•
What ventilation is required when occupancy varies over time?
The standard includes specific construction requirements for:
•
Outdoor air intakes to minimize moisture problems due to rain and snow
•
Filtration requirements to prevent wet coils from excessive dirt collection
•
Drain pans’ slope and drainage arrangement to ensure that condensation
drains away
•
Access for maintenance and cleaning of coils
•
Duct construction
•
System startup and balancing
After construction, the standard has requirements for the system’s ongoing operation and
maintenance, including inspection and measurement of outdoor air flow.
8–6
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8.4 Other Codes and Standards
Several organizations other than ASHRAE produce codes and standards relating to HVAC
duct design. Included among these are the National Fire Protection Association (NFPA)
and the Sheet Metal, Air Conditioning Contractors’ National Association (SMACNA),
and American Conference of Governmental Industrial Hygienists (ACGIH).
NATIONAL FIRE PROTECTION ASSOCIATION
The National Fire Protection Association (NFPA) issues a wide range of standards. Three
of interest to HVAC designers are NFPA 90A, NFPA 90B and NFPA 96.2-4
NFPA 90A–Installation of Air Conditioning and Ventilating Systems applies to systems for
air movement in:
•
Structures over 25,000 ft3 in volume
•
Buildings of Type III, IV and V construction over three stories in height
regardless of volume
•
Buildings, spaces, occupants and processes not covered by other NFPA standards.
As stated in the standard, the purpose of NFPA 90A is “to prescribe minimum requirements for safety to life and property from fire.” The requirements of NFPA 90A are
intended to:
•
Restrict the spread of smoke through air duct systems in a building or into a
building from the outside
•
Restrict the spread of fire through air duct systems from the area of fire origin
whether it be within the building or from outside
•
Maintain the fire-resistive integrity of building components and elements
(such as floors, partitions, roofs, walls and floor/roof-ceiling assemblies)
affected by the installation of air duct systems
•
Minimize ignition sources and combustibility of the elements of the air duct
systems
•
Permit the air duct systems in a building to be used for the additional purpose
of emergency smoke control
NFPA 90A provides requirements for HVAC systems (equipment and air distribution),
integrating HVAC systems with building construction, controls and acceptance testing.
8–7
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Of particular interest with respect to duct design, Figure 8-2 shows required treatments of
penetrations of walls or partitions, and location of fire and smoke dampers. The requirement is that fire dampers be shown on the drawings.
The building’s architectural design will determine its fire separations and the requirements
for duct fire and smoke dampers. Fire dampers are a significant cost, and accessible access
doors must be provided for checking and servicing them. When laying out the ductwork,
choices can often be made to reduce the number of fire dampers and to position the access
doors to minimize costs.
NFPA 90B–Installation of Warm Air Heating and Air-Conditioning Systems applies to all
warm air heating and air-conditioning systems that serve: one- or two-family dwellings;
and spaces not exceeding 25,000 ft3 in volume in any occupancy (for example, light commercial).
Other systems are covered by NFPA 90A. Standard NFPA 90B addresses: system components; fire integrity of building construction; equipment; wiring; and controls.
NFPA 96–Installation of Equipment for the Removal of Smoke and Grease-Laden Vapors from
Commercial Cooking Equipment covers basic requirements for the design, installation and
use of exhaust system components including: hoods; grease removal devices; exhaust ducts;
dampers; air moving devices; auxiliary equipment; and fire extinguishing equipment for
the exhaust system and the cooking equipment used in commercial, industrial institutional, and similar cooking applications.
Other topics discussed in NFPA 96 include: duct systems; air movement; procedures for
use and maintenance of equipment; and minimum safety requirements for cooking equipment. This standard does not apply to installations for normal residential family use.
8–8
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Wall and Partition Penetrations and Smoke Dampers
Figure 8-2
8–9
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SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSN.
The SMACNA HVAC Duct Construction Standards cover: basic duct construction; fittings
and other construction; round, oval and flexible duct; hangers and supports; exterior components; casings; functional criteria for demonstrating equivalency; and duct sealing classifications.
Also included are highly valuable appendices useful in duct construction, and fibrous glass
duct construction standards.5
AMERICAN CONFERENCE OF GOVERNMENTAL INDUSTRIAL HYGIENISTS
The American Conference of Governmental Industrial Hygienists (ACGIH) publishes and
regularly updates Industrial Ventilation, A Manual of Recommended Practice. This book
includes general information on ventilation and numerous examples of industrial ventilation and the removal of contaminants from specific industrial processes. The 25th edition
was published in 2004.
8.5 Sources of Information
Many sources of information are available to HVAC designers:
•
ASHRAE produces an extensive range of standards, Handbooks and Advanced
Energy Design Guides which can be located at the website: www.ashrae.org.
•
National Fire Protection Association (NFPA), 1 Batterymarch Park, Quincy,
MA 02269-9101; 617/770-3000, Fax 617/770-0700; website: www.nfpa.org.
•
Sheet Metal and Air Conditioning Contractors’ National Association Inc.
(SMACNA), 4201 Lafayette Center Drive, Chantilly, VA 22021-1209;
703/803-2980, Fax 703/803-3732; website: www.smacna.org.
•
American Conference of Governmental Industrial Hygienists (ACGIH), Kemper Woods Center, 1330 Kemper Meadow Dr., Cincinnati, OH 45240;
513/742-2020, Fax 513/742-3355; website: www.acgih.org.
•
Air Movement and Control Association Inc. (AMCA), 30 West University
Drive, Arlington Heights, IL 60004-1893; 708/394-0150, Fax 708/253-0088;
website: www.amca.org.
In addition, each chapter of the ASHRAE Handbooks contains a detailed bibliography. An
extensive list of HVAC codes and standards is included in the ASHRAE Handbook–Fundamentals.6
8–10
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The Next Step
The next chapter will discuss some components in air systems including dampers, air filters, humidifiers, duct heaters and duct insulation.
Summary
In the private sector, each new construction or renovation project is normally governed by
state laws and/or local ordinances that require compliance with specific health, safety,
property protection and energy conservation regulations. These requirements are based on
existing design methods, and negotiation may be needed with the authorities to use new
design methods.
ASHRAE Standard 90.1 has become the legal requirement in the USA. The standard is a
consensus document with public review. It is adopted by the American National Standards
Institute (ANSI), and undergoes continuous improvement through addenda. The latest
printed edition was released in 2007.
The standard covers the building fabric and all permanent energy-using plant and equipment in the building. Section 6HVAC includes two methods of achieving compliance:
•
Prescriptive approach: Follow specific set of requirements, including: minimum
requirements for plant efficiency; when economizers and heat recovery must be
included; and insulation and control strategies to minimize wasting energy.
Simple rules are included for some small buildings.
•
Energy cost budget method: Design the building to have no greater energy cost
than a system designed under the prescriptive approach.
ASHRAE Standard 62-2007 sets out requirements for:
•
Ventilation with outside air and exhaust from polluted spaces
•
Design of the systems to facilitate correct operation through the life of the
building
•
Operations and maintenance requirements
•
Documentation requirements
The National Fire Protection Association (NFPA) has three standards applicable to
HVAC systems:
•
NFPA 90A–Installation of Air Conditioning and Ventilating Systems applies to
systems for air movement in larger buildings with the emphasis on life safety.
8–11
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The focus is on reducing the risk of fire and smoke and their effect when they
do occur.
•
NFPA 90B–Installation of Warm Air Heating and Air-Conditioning Systems
applies to smaller buildings.
•
NFPA 96–Installation of Equipment for the Removal of Smoke and Grease-Laden
Vapors from Commercial Cooking Equipment covers kitchen exhausts and fire
suppression.
The SMACNA book HVAC Duct Construction Standards covers the design, construction
and installation of galvanized ductwork in detail and other materials more generally.
The ACGIH book Industrial Ventilation, A Manual of Recommended Practice includes general information on ventilation and numerous examples of industrial ventilation and the
removal of contaminants from specific industrial processes.
Bibliography
1. ASHRAE. 2007. ASHRAE/IESNA Standard 90.1-2007, Energy Efficient Design of New Buildings
Except Low-Rise Residential Buildings. Atlanta, GA: ASHRAE.
2. ASHRAE. 2007. ASHRAE/IESNA Standard 62.1-2007, Ventilation for Acceptable Indoor Air
Quality.
3. NFPA. 2002. NFPA 90A–Installation of Air Conditioning and Ventilating Systems. Quincy, MA:
National Fire Protection Association.
4. NFPA. 2006. NFPA 90B–Installation of Warm Air Heating and Air-Conditioning Systems.
Quincy, MA: National Fire Protection Association.
5. NFPA. 2004. NFPA 96–Installation of Equipment for the Removal of Smoke and Grease-Laden
Vapors from Commercial Cooking Equipment. Quincy, MA: National Fire Protection
Association.
6. SMACNA. 2005. HVAC Duct Construction Standards. Chantilly – Metal and Flexible, VA: Sheet
Metal and Air Conditioning Contractors' National Association Inc.
ASHRAE HandbookFundamentals, codes and standards; HandbookHVAC Systems and Equipment, codes and standards relevant to specific systems and equipment; Handbook
Refrigeration, codes and standards; HandbookApplications, codes and standards relevant
to specific applications.
8–12
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Skill Development Exercises for Chapter 8
Complete these questions by writing your answers on the worksheet at the back of this
book.
8-1.
Combustibility and toxicity ratings are normally based on tests of:
a) New materials b) Old work c) Fibrous materials d) All of the above
8-2.
In the private sector, new construction is normally governed by:
a) State laws b) Local ordinances c) Codes d) All of the above
8-3.
Zone temperature controls are required for all systems, with special requirements for perimeter heating systems.
a) True b) False
8-4.
Which of the following standards applies to structures not exceeding 25,000 ft3
in volume?
a) NFPA 90A b) NFPA 90B c) NFPA 96 d) All of the above
8-5.
SMACNA HVAC Duct Construction Standards covers:
a) Basic duct construction b) Hangers and supports
c) Duct sealing classifications d) All of the above
8-6.
ASHRAE Standard 90.1 has a somewhat easier compliance route for many small
air-conditioned buildings.
a) True b) False
8-7.
Compliance with ASHRAE Standard 90.1, Section 6, assures a minimum level
of HVAC system performance.
a) True b) False
8-8.
HVAC designers must know which code compliance obligations affect their
designs.
a) True b) False
8-9.
HVAC systems are one of the most significant energy users in the types of buildings covered by ASHRAE Standard 90.1.
a) True b) False
8-10.
A very efficient HVAC system could have an overall efficiency greater than one.
a) True b) False
8–13
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Chapter 9
Air System Auxiliary
Components
Contents of Chapter 9
•
•
•
•
•
•
•
•
9.1 Dampers
9.2 Air Filters
9.3 Humidifiers
9.4 Duct Heaters
9.5 Duct Insulation
Summary
Bibliography
Skill Development Exercises for Chapter 9
9–1
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Study Objectives of Chapter 9
After completing this chapter, you should understand the function, selection and sizing of:
•
•
•
•
•
Dampers
Air filters
Humidifiers
Duct heaters
Duct insulation
9.1 Dampers
TYPES OF DAMPERS
Two damper arrangements
are used for air-handling system flow control: parallelblade and opposed-blade
(Figure 9-1). The linkages
shown in the figure are
attached to the blades. Moving the linkage upwards on
the parallel blade damper
opens the damper and lowering the linkage closes the
damper. Note that the ends
of the damper blades have
opposed grooves. This is so
that the grooves interlock
when the damper is closed to
improve the seal and provide
rigidity to the damper blade.
Having the linkage in the
airstream increases the
Figure 9-1 Parallel and Opposed Blade Dampers
damper resistance and, at
higher air speeds, can produce air noise. The preferable alternative, although a little more costly, is for the linkage to
be external and connected to the damper shafts.
The sheet metal blade section shown in Figure 9-1 is made by forming three grooves (one
at each edge and a central one around the shaft) and is thus called a triple-V blade. Blades
9–2
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are also made in aerofoil section, providing a lower resistance to airflow and lower noise
generation. The power to drive the linkage is from an actuator.
Optimum control of airflow is obtained with a linear relationship between air flow and the
degree to which the damper is open. For many years, the information about damper performance has been:
Parallel-blade dampers are adequate for two-position control
and can be used for modulating
control when they are the primary source of pressure drop
and directional air flow is not a
problem. Opposed-blade
dampers are preferable, because
they normally provide better
control. The characteristic
curves of installed parallel
blade dampers and installed
opposed blade dampers are
shown in Figures 9-2 and 9-3,
respectively. The parameter a
in both figures is the ratio of
the pressure drop across the
fully open damper at design
flow to the total subsystem
pressure drop, including fully
open control damper pressure
drop.
Figure 9-2
Installed Parallel Blade Dampers
Figure 9-3
lnstalled Opposed Blade Dampers
9–3
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These idealized curves are correct in concept but not realized in practice. Recent research,
particularly ASHRAE Research Report 1157 Flow Resistance and Modulating Characteristics of Control Dampers,1 shows that the performance of dampers is highly dependent on:
•
Construction. Differs from manufacturer to manufacturer for the same style,
triple-V or aerofoil. See Figure 9-4 where triple-V dampers from two manufacturers have very different performance curves in both arrangements.
•
Relative size of the damper to the duct or plenum and the arrangement. A simple
example is the situation where the damper is the same size as the duct, so the
airflow is relatively straight into the damper. In contrast, a small damper in a
large wall will have air coming from all directions into the damper, creating a
different flow characteristic. In Figure 9-4, the performance characteristic is
modified for an intake louver and damper to a damper and relief louver.
•
Location relative to other components including changes in duct direction. See Figure 9-5 for an example where the opposed blade damper characteristic is
degraded by being placed inside an inlet louver.
Figure 9-4
9–4
Two Parallel Blade Triple-V Dampers From Different Manufacturers
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Therefore, actual performance data must be
obtained from the manufacturers on their specific
dampers and the situational
conditions that influence
performance must be considered.
The one situation where
parallel blade dampers consistently provide more linear control is in the mixing
box, typically mixing outside air and return air to
provide supply air. The Figure 9-5 Effect of Inlet Louver on an Opposed
Blade Damper Characteristic
combination of three parallel bladed dampers working
in unison provides a more
linear control characteristic than when using opposed blade dampers.
Damper leakage is important, particularly where tight shutoff is required. For example, an
outdoor air damper must close tightly to prevent coils and pipes from freezing. Low leakage dampers are more costly and require larger operators because of the friction of the seals
in the closed position. Therefore, they should be used only when necessary, including any
location where the tight-closing damper will reduce energy consumption significantly. Literature from manufacturers expresses leakage rates when exposed to specific pressure differentials across the closed damper. Details about sizing dampers and leakage are given in
the ASHRAE course Fundamentals of HVAC Controls.
DAMPER OPERATORS
Damper operators are available using either electricity or compressed air as a power source:
•
Electric damper operators can be either unidirectional spring return or reversible. This type of operator is available with many options for rotational shaft
travel (expressed in degrees of rotation) and timing (expressed in the number of
seconds to move through the rotational range).
•
Pneumatic damper operators use air pressure to produce a linear motion of the
shaft through which a linkage moves the crank arm to open or close the dampers. Normally-open or normally-closed operation refers to the position of the
dampers when no air pressure is applied at the operator, the failed position.
Positive positioners are important for sequencing the damper with other
devices.
9–5
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DAMPER FUNCTIONS
Dampers have a wide variety of functions:
•
Shutoff dampers are used to regulate the air flow through a duct. When fully
closed, they shut off the flow aside from any leakage that may occur.
•
Balancing dampers are used to make final adjustments in the air flow through
a duct when the system is first being commissioned. In smaller ducts, balancing
dampers are often flat metal plates as they are just a variable resistance to be set
up by the balancing contractor. The balancing damper may be used to adjust
the total flow in a single duct system or to adjust the ratio of flows in systems
with multiple ducts.
•
In fire and smoke control, openings for ducts in walls and floors with fire resistance ratings should be protected by fire dampers and ceiling dampers as
required by local codes. Note that fire dampers are manufactured in two styles:
with the damper in the duct section and with the damper outside the duct section. Having the damper in the duct section may be required where space is
very tight, but the significant resistance must be allowed for particularly in
small ducts. Air transfer openings should also be protected.
A smoke damper can be used for either traditional smoke management (smoke
containment) or for smoke control. In smoke management, a smoke damper
inhibits the passage of smoke under the forces of buoyancy, stack effect and
wind. Generally, for smoke containment, smoke dampers should have low
leakage characteristics at elevated temperatures. However, smoke dampers are
only one of many elements (partitions, floors, doors, etc.) intended to inhibit
smoke flow. In smoke management applications, the leakage characteristics of
smoke dampers should be selected to be appropriate with the leakage of the
other system elements.
In a smoke control system, a smoke damper inhibits the passage of air that may
or may not contain smoke. Low leakage characteristics of a damper are not necessary when outside air is on the high-pressure side of the damper, as is the case
for dampers that shut off supply air from a smoke zone or that shut off exhaust
air from a nonsmoke zone. In these cases, moderate leakage of smoke-free air
through the damper does not adversely affect the control of smoke movement.
Smoke control supply air systems should be designed so that only smoke-free
air is on the high-pressure side of a smoke damper. These dampers should be
classified and labeled in accordance with UL-555 Standards.2,3,4
9–6
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9.2 Air Filters
The purpose of a filter is to remove contaminants from an airstream. Contaminants may
be gaseous, such as odors from the adjacent restaurant, or particulates from outside and
inside the building. Gaseous filtration is costly to install and maintain. Activated carbon
filters may be used for general organic vapor removal. In other situations, specific gaseous
compounds can be removed using filters containing chemicals to remove the specific contaminant. Gaseous filtration is a specialized field and is not covered in this course. The
most common situation is particulate removal. The characteristics of airstreams that most
affect the performance of an air filter include particle size and shape, mass, concentration
and electrostatic properties. The most important of these is particle size.
Particle size may be defined in numerous ways. Particles less than 2.5 µm (microns, or millionths of a meter) in diameter are generally referred to as fine, with those greater than 2.5
µm being considered as coarse.
From an industrial hygiene perspective, particles that are 5 µm or greater are considered
the nonrespirable fraction of dust, which means that they are filtered out in the nasal passage before reaching the lungs. Particles less than 5 µm are considered the respirable fraction. Particle size in this discussion refers to aerodynamic particle size (defined as the diameter of a unit-density sphere having the same gravitational settling velocity as the particle in
question). Therefore, larger particles with lower densities could be found in the lungs. Also
note that fibers are different than particles in that fiber shape, diameter and density all
affect where a fiber will settle in the body.5
Atmospheric dust is a complex mixture of smokes, mists, fumes, dry granular particles,
microorganisms, other biologically produced particles, and natural and synthetic fibers.
When suspended in air, this mixture is called an aerosol. A sample of atmospheric dust
usually contains soot, smoke, silica, clay, decayed animal and vegetable matter, organic
materials in the form of lint and plant fibers, and metallic fragments. It may also contain
living organisms, such as mold spores, bacteria and plant pollens, which may cause diseases
or allergic responses.
Major factors influencing filter design and selection include: degree of air cleanliness
required; specific particle size range or aerosols that require filtration; and aerosol concentration.
Note that filters are used to protect ductwork and equipment as well as occupied spaces.
Cooking facilities require a grease filter that both reduces the grease load in the duct and
also acts as a fire stop between the cooking surface and the ducting. Clothes dryers require
filters to reduce the buildup of fibers in the duct and on any exhaust screen. These situations are often quite specifically mandated in local building and fire codes.
9–7
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RATING FILTERS
In addition to criteria affecting the degree of air cleanliness, factors such as cost (initial
investment and maintenance), space requirements and air flow resistance have encouraged
the development of a wide variety of filters. Accurate comparisons of different filters can be
made only from data obtained by standardized test methods.
The three main operating characteristics that distinguish the various types of filters are efficiency, air flow resistance and dust-holding capacity:
•
Efficiency measures the filter’s ability to remove particulate matter from an airstream. Average efficiency during the life of the filter is the most meaningful
for most filters and applications. However, because the efficiency of many drytype filters increases with dust load, in applications with low dust concentrations, the initial (clean filter) efficiency should be considered for design.
•
Air flow resistance (or resistance) is the pressure drop across the filter at a given
air flow rate. The term pressure drop is used interchangeably with resistance.
•
Dust-holding capacity defines the weight of dust that a filter can hold when it is
operated at a specified air flow rate to some maximum resistance value or
before its performance drops seriously as a result of the collected dust.
In general, four types of tests, together with certain variations, determine filter efficiency:
9–8
•
Arrestance. A standardized synthetic dust consisting of various particle sizes is
fed into the filter, and the weight fraction of the dust removed is determined.
In the ASHRAE Standard 52.1 test, this type of efficiency measurement is
named synthetic dust weight arrestance to distinguish it from other efficiency
values. The synthetic dust used contains fibers and is generally coarser than
normally experienced dust, so the test is of limited value.
•
Dust spot efficiency. As defined in ASHRAE Standard 52.1, a standardized
atmospheric dust is passed into the filter, and the discoloration effect of the
cleaned air on filter paper targets is compared with that of the incoming air.
This type of measurement is called atmospheric dust spot efficiency.
•
Particle size removal efficiency test. ASHRAE Standard 52.2 details this method.
An optical particle counter measures the number of particles upstream and
downstream of the filter. The measurements are made for particles in the range
0.3 µm to 10 µm. Based on the results, filters are classified into 20 categories
called Minimum Efficiency Reporting Value (MERV). The MERV 1 filter is
the least efficient, typically collecting long fibers and particles over 10 µm. At
the other extreme are the MERV 17 to 20 filters used in industrial and medical
facilities to remove dusts of 0.3 µm at better than 99.97% efficiency.
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•
DOP Penetration Test. This is a U.S. Army specification test (MIL-STD-282)
using a chemical, DOP, producing a particle cloud of around 0.3 µm. The rating is based on the proportion of particles penetrating through the filter. This
test is used to rate very high efficiency filters, MERV 17 to 20.
Table 9-1 describes the common range of particulate filters, typical use, and their performance as tested under ASHRAE Standards 52.1 and 52.2. Read through the table’s Application Guidelines and MERV ratings so you understand the range of filter performance
and typical uses. This will provide context for the later discussion on filter selection.
Table 9-1
Std. 52.2
MERV
Rating
Filter Types and Performance*
Approx. Std. 52.1
Results
Application Guidelines
Dust Spot
Efficiency
Arrestance
Typical
Controlled
Contaminant
20
19
18
n/a
n/a
n/a
n/a
n/a
n/a
17
n/a
n/a
Larger than 0.3 m
particles: Virus, all
combustion smoke,
sea salt, radon
progeny
Cleanrooms; pharHEPA/ULPA filters ranging from
maceutical manufac99.97% efficiency on 0.3 mm particles
turing; orthopedic
to 99.999% efficiency on 0.1 - 1.2 mm
surgery
particles
16
15
14
n/a
>95%
90-95%
n/a
n/a
>98%
13
80-90%
>98%
0.3-1.0 m particle size and all over
1 m: All bacteria,
most tobacco
smoke, sneeze
nuclei
Hospital in-patient
care; general surgery;
superior commercial
Bag Filters: Nonsupported (flexible
buildings
microfine fiberglass or synthetic media
12-36 in. deep, 6-12 pockets
12
11
10
70-75%
60-65%
50-55%
>95%
>95%
>95%
9
40-45%
>90%
8
7
6
30-35%
25-30%
<20%
>90%
>90%
85-90%
5
<20%
80-85%
4
3
2
<20%
<20%
<20%
75-80%
70-75%
65-70%
1
<20%
<65%
1.0-3.0 m particle size and all over
3.0 m: Legionella,
auto emissions,
welding fumes
Typical
Applications and
Limitations
Typical Air Cleaner/
Filter Type
Box Filters: Rigid style cartridge filters 6Hospital labs; better 12 in. deep may use lofted (air laid) or
commercial buildpaper (wet laid) media
ings; superior residential
3.0-10.0 m particle size and all over
10 m: Mold,
spores, cement dust
Commercial buildings; better residential; industrial
workplaces
>10.0 m particle
size: Pollen, dust
mites, sanding
dust, textile fibers
Minimum filtration; Throwaway Filters: Disposable fiberglass
residential; window or synthetic panel filters
air conditioners
Washable Filters: Aluminum mesh, latex
coated animal hair, foam rubber panels
Electrostatic Filters: Self-charging (passive) woven polycarbonate panel filter
Pleated Filters: Disposable extended surface 1-5 in. thick with cotton polyester
blend media, cardboard frame
Cartridge Filters: Graded density viscous
coated cube or pocket filters, synthetic
media
Throwaway Filters: Disposable synthetic media panel filters
*As related to ASHRAE Standard 52.1 and Standard 52.2
9–9
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MECHANISMS OF PARTICLE COLLECTION
Filters rely on five main principles or mechanisms:
•
Straining. The coarsest kind of filtration strains particles through a membrane
opening that is smaller than the particulate being removed. It is most often
observed as the collection of large particles and lint on the filter surface. The
mechanism is not adequate to explain the filtration of submicron aerosols
through fibrous matrices, which occurs through other mechanisms.
•
Direct interception. The particles follow a fluid streamline close enough to a
fiber that the particle contacts the fiber and remains there. The process is
nearly independent of velocity.
•
Inertial deposition. Particles in the airstream are large enough or of large enough
density that they cannot follow the fluid streamlines around a fiber; thus, they
cross over streamlines, contact the fiber and remain there. At high velocities
(where these inertia effects are most pronounced), the particle may not adhere
to the fiber because drag and bounce forces are so high. In this case, a viscous
coating applied to the fiber obtains the full benefit and is the predominant
mechanism in an adhesive-coated, wire screen impingement filter.
•
Diffusion. Very small particles have random motion about their basic streamlines (Brownian motion), which contributes to deposition on the fiber. This
deposition creates a concentration gradient in the region of the fiber, further
enhancing filtration by diffusion. The effects increase with decreasing particle
size and velocity. Do not exceed manufacturer’s recommended velocities.
•
Electrostatic effects. Particle or media charging can produce changes in the coagulation and collection of dust.
TYPES OF FILTERS
Common filters are broadly grouped as those using a fibrous media and electrically powered electrostatic filters. The fibrous filters can be subdivided into those with replaceable
panels and those with a renewable media which is moved across the airstream.
Panel filters. There are a variety of panel filters including viscous impingement filters,
dry-type extended-surface filters, and High Efficiency Particulate Air (HEPA) filters.
Viscous impingement filters are panel filters made of coarse fibers with high porosity. Glass
fibers, steel or aluminum mesh, and metal baffles are used for filter media. The filter
medium is coated with a viscous substance, such as filter oil (also known as adhesive), that
causes particles that impinge on the fibers to stick to them. Design air velocity through the
fiber is usually in the range of 200 to 800 fpm. These filters are characterized by low pressure drop, low cost and good efficiency on lint, but low efficiency on normal atmospheric
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dust. They are commonly made 0.5 in. to 4 in. thick. Unit panels are available in standard
and special sizes up to about 24 in. by 30 in. This filter is commonly used in residential
furnaces and air conditioning and is often used as a prefilter for higher efficiency filters.
Although viscous impingement filters usually operate in the range of 300 to 600 fpm, they
may be operated at higher velocities. The limiting factor, other than increased flow resistance, is the danger of blowing off agglomerates of collected dust and the viscous coating
on the filter. Do not exceed the manufacturer’s recommended velocities.
The loading rate of a filter depends on the type and concentration of the dirt in the air
being handled and the system’s operating cycle. Manometers, static pressure gauges and
pressure transducers are often installed to measure the pressure drop across the filter bank
and thereby indicate when the filter requires servicing. The final allowable pressure drop
may vary from one installation to another. But, in general, unit filters are serviced when
their operating resistance reaches 0.5 in. wg. Note that in systems with low air pressure
losses, the increase in filter pressure drop may seriously reduce the airflow as the filter
becomes loaded.
The manner of servicing unit filters depends on their construction and use. Disposable viscous impingement, panel-type filters are constructed of inexpensive materials and are discarded after one period of use. The sides of this design are usually a combination of cardboard and metal stiffeners. Permanent unit filters are generally constructed of metal to
withstand repeated handling. Various cleaning methods have been recommended for permanent filters; the most widely used involves washing the filter with steam or water (frequently with detergent) and then recoating it with its recommended adhesive by dipping
or spraying. Unit viscous filters are also sometimes arranged for in-place washing and
recoating.
Dry-type extended-surface filters use media made of random fiber mats or blanks of varying
thicknesses, fiber sizes and densities. Bonded glass fiber, cellulose fibers, wool felt, synthetics and other materials have been used commercially. The media in filters of this class are
frequently supported by a wire frame in the form of pockets, or V-shaped or radial pleats.
In other designs, the media may be self-supporting because of inherent rigidity or because
air flow inflates it into extended form, such as with bag filters. Pleating of the media provides a high ratio of media area to face area, thus allowing lower pressure drops.
The efficiency of dry media filters can be improved by the use of passive, electrostaticallycharged media. The charged media is manufactured in several ways, one being called electret. The charge increases the collection effect, particularly when the filter is clean. As the
dust builds up, the charged effect is reduced by varying degrees. Therefore, the lifetime
performance, not just the clean performance, must be considered in choosing this type of
filter. These are not electronic filters, although commercial literature sometimes implies
that they are.
9–11
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High Efficiency Particulate Air (HEPA) filters are made in an extended surface configuration of deep space folds of submicron glass fiber paper. Such filters operate at face velocities
near 250 fpm, with resistance rising from 0.5 to more than 2.0 in. wg or more over their
service life. Note that this large increase usually requires some type of control to ensure reasonably constant air flow. These filters are the standard for critical medical facilities, cleanrooms, nuclear and toxic-particulate applications.
Renewable media filters. The two types of renewable media filters are: moving-curtain
viscous impingement filters; and moving-curtain dry media filters. Automatic moving-curtain viscous filters are available in two main types. In one type, random-fiber media are furnished in roll form. Fresh media are fed manually or automatically across the face of the filter, while the dirty media are rewound onto a roll at the bottom. When the roll is
exhausted, the tail of the media is wound onto the take-up roll and the entire roll is thrown
away. A new roll is then installed and the cycle is repeated.
In moving-curtain dry media filters, random-fiber (nonwoven) dry media of relatively high
porosity are also used for general ventilation service. Operating duct velocities of about
200 fpm are generally lower than for viscous impingement filters.
Special automatic dry filters are also available. These are designed for the removal of lint in
textile mills and dry-cleaning facilities and the collection of lint and ink mist in press
rooms. The medium used is extremely thin and serves only as a base for the buildup of lint,
which then acts as a filter medium. The dirt-laden media are discarded when the supply
roll is used up. Another form of filter designed specifically for dry lint removal consists of a
moving curtain of wire screen, which is vacuum cleaned automatically at a position out of
the airstream. Recovery of the collected lint is sometimes possible with such a device.
Performance of renewable media filters is covered by ASHRAE Standards 52.1 and 52.2
(see also Table 9-1).
Higher MERV rated filters are normally provided with a lower MERV pre-filter. A low
cost MERV 6 filter may thus be used to collect the larger particles and extend the useful life
of a MERV 14 filter. The MERV 6 filter will likely need to be changed more frequently
than the MERV 14 filter.
Electronic filters. Electronic filters can be highly efficient, using electrostatic precipitation
to remove and collect particulate contaminants such as dust, smoke and pollen. The filter
consists of an ionization section and a collecting plate section. This filter should not be
confused with passive electrostatic unit filters.
In the ionization section, small diameter wires with a positive direct current potential of
between 6 and 25 kV DC are suspended equidistant between grounded plates. The high
voltage on the wires creates an ionizing field for charging particles. The positive ions cre-
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ated in the field flow across the airstream and strike and adhere to (charge) the particles,
which then pass into the collecting plate section.
The collecting plate section consists of a series of parallel plates equally spaced with a positive direct current voltage of 4 to 10 kV DC applied to alternate plates. Plates that are not
charged are at ground potential. As the particles pass into this section, they are forced to
the plates by the electric field on the charges they carry, and are removed from the airstream and collected by the plates. Figure 9-6 shows a typical electronic filter cell. Particulate retention is a combination of electrical and intermolecular adhesion forces and may be
augmented by special oils or adhesives on the plates.
Figure 9-6
Cross-section of an Ionizing Electronic Air Cleaner
FILTER SELECTION AND MAINTENANCE
To evaluate filters properly for a particular application, the following factors should be
considered:
•
•
•
•
•
•
•
Code requirements
Types of contaminants present indoors and outdoors
Sizes and concentrations of contaminants
Air cleanliness levels required in the system and space
Air filter efficiency needed to achieve cleanliness
Space available to install and access equipment
Life cycle costing, including:
Operating resistance to airflow (static pressure differential, fan power
Disposal or cleaning requirements of spent filters
Initial cost of selected system
Cost of replacement filters or cleaning
Cost of warehousing filter stock and change-out labor
9–13
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Savings  from reduction in housekeeping expenses, protection of valuable property and
equipment, ability to conduct dust-free manufacturing processes, improved working conditions and even health benefits  should be credited against the cost of installing and
operating an adequate filtration system. The capacity and physical size of the required unit
may emphasize the need for low maintenance cost. Operating costs, predicted life and efficiency are as important as initial cost because air cleaning is a continuing process. Panel filters do not have efficiencies as high as can be expected from extended-surface filters, but
their initial cost and upkeep are generally low. They require more careful attention than
the moving-curtain type if the resistance is to be maintained within reasonable limits.
If higher efficiencies are required, extended-surface filters or electronic filters should be
considered. The use of very fine glass fiber mats or other materials in extended-surface filters has made these available in the highest efficiency ranges. Initial costs of extended-surface filters are lower than for electronic types, but higher than for panel types. Operating
and maintenance costs of some extended-surface filters may be higher than for panel types
and electronic filters, but the efficiencies are always higher than for panel types, and the
cost/benefit ratio must be considered. The pressure drop of media-type filters is greater
than electronic filters and slowly increases during their useful life. The advantages are that
no mechanical or electrical services are required. The choice should be based on both initial and operating costs, as well as on the degree of cleaning efficiency and maintenance
requirements.
In selecting specific filters, designers should carefully evaluate the total media surface area.
Filters with more surface area generally have longer service life and lower pressure drops.
While electronic filters have a higher initial cost, they exhibit high initial efficiencies in
cleaning atmospheric air; this is largely because of their ability to remove fine particulate
contaminants. System resistance remains unchanged as particles are collected, and the
resulting residue has to be periodically washed off. The manufacturer must supply information on maintenance or cleaning, but this information is often not specific about cleaning frequencies. Industrial systems with automatic washing built in can be continuously
effective. The typical residential unit’s performance drops off in 10-20 days, so the panels
must be run through a dishwasher every two weeks to maintain performance. Not surprisingly, the typical problem is that the collection plates are not cleaned frequently enough to
maintain performance.
FILTER INSTALLATION
Many filters are available in units of convenient size for manual installation, cleaning and
replacement. A typical unit filter may be 20 to 24 in.2, from 1 to 40 in. thick, and or either
the dry or viscous impingement types. In large systems, the frames in which these units are
installed are bolted or riveted together to form a filter bank. Automatic filters are constructed in sections offering several choices of width up to 70 in. and generally range in
9–14
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height from 40 to 200 in., available in 4 to 6 in. increments. Several sections may be bolted
together to form a filter bank.
Several manufacturers provide side-loading filter sections for various types of filters. Filters
are changed from outside the duct, making service areas in the duct unnecessary and thus
saving cost and space.
Of course, the in-service efficiency of an air filter is sharply reduced if air leaks through
either bypass dampers or poorly designed frames. The higher the filter’s efficiency, the
more attention must be paid to the frame’s rigidity and sealing effectiveness. In addition,
high efficiency filters must be handled and installed with care.
Filters may be installed in the outdoor air intake ducts of buildings and residences and in
the recirculation and bypass air ducts. But they are always placed ahead of heating or cooling coils and other air-conditioning equipment in the system to protect the equipment
from dust. The dust captured in an outdoor air intake duct is likely to be mostly particulates of a greasy nature, while lint may predominate in dust from within the building.
Where high efficiency filters protect critical areas such as cleanrooms, the filters must be
installed as close to the room as possible to prevent the pickup of particles between the filters and the outlet. The ultimate is the so-called laminar flow room, in which the entire
ceiling or one entire wall becomes the filter bank.
The published performance data for all air filters are based on straight-through unrestricted air flow. Filters should be installed so that the face area is at right angles to the air
flow whenever possible, although a V-bank filter arrangement is often used. Eddy currents
and dead air spaces should be avoided; air should be distributed uniformly over the entire
filter surface. Baffles, diffusers or air blenders are occasionally necessary. Filters are sometimes damaged if significantly higher than recommended air velocities impinge directly on
the filter face.
An example is where a filter with design face velocity of 500 fpm is installed in a duct with
an air velocity of 1,000 fpm. The duct will need a tapered section to enlarge by a factor of
1.4 to fit the filter. The taper to the filter must be slow enough to let the air spread out to
cover the entire filter area. After the filter, the taper back down to duct size can be quite
short without significantly affecting filter performance.
Air filter installations that give unsatisfactory results can, in most cases, be traced to faulty
installation, improper maintenance or both. The most important requirements of a satisfactory and efficiently operating air filter installation are as follows:
•
The filter must have ample capacity for the amount of air and dust load it is
expected to handle. An overload of 10% to 15% is regarded as the maximum
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•
•
allowable. When air volume is subject to future increase, a larger filter bank
should be considered initially.
The filter must be suited to the operating conditions, such as degree of air
cleanliness required, amount of dust in the entering air, type of duty, allowable
pressure drop, operating temperatures and maintenance facilities.
The filter type should be the most economical for the specific application. The
initial installation cost should be balanced against efficiency and depreciation,
as well as the expense and convenience of maintenance.
The following recommendations apply to filters installed with central fan systems:
•
•
•
•
•
•
•
•
Duct connections to and from the filter should change size or shape gradually
to ensure even air distribution over the entire filter area.
Sufficient space should be provided in front of or behind the filter, or both,
depending on its type, to make it accessible for inspection and service. A distance of 20 to 40 in. may be required, depending on the filter chosen.
Access doors of convenient size should be provided to the filter service areas.
All doors on the clean-air side should be gasketed to prevent infiltration of
unclean air. All connections and seams of the sheet metal ducts on the clean-air
side should be sealed as airtight as possible. The filter bank must be sealed to
prevent bypass of unfiltered air, especially when high efficiency filters are used.
Electric lights should be installed in the plenum in front of and behind the air
filter.
Filters installed close to an air inlet should be protected from the weather by
suitable louvers or an inlet hood. A large mesh wire bird screen should be
placed in front of the louvers or in the hood.
Filters, other than electronic filters, should have permanent indicators to give a
warning when the filter resistance is exhausted or its value becomes too high, as
with automatic roll media filters.
Electronic filters should have an indicator or alarm system to signal when high
voltage is off or shorted out.
The required filter performance is generally not specified in codes for commercial, institutional, and large residential buildings. ASHRAE Standard 62.1-2007 requires a MERV 6
filter before any cooling coil that can run with a wetted surface and a MERV 6 filter in
dirty locations. Health facilities are normally covered by codes with very specific requirements. Remember that particulate filters do not remove odors and that gas phase filtration
is costly to install and maintain. If specific odors are near the site, such as an adjacent restaurant, or on the site, such as a loading dock, do not ignore them. It may be much more
cost effective to modify the layout of the building and air system so that the intake is well
away from the odor source.
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FILTER SAFETY REQUIREMENTS
Safety ordinances should be investigated when the filter installation is contemplated.
Combustible filtering media may not be permitted in accordance with some local regulations. Combustion of dust and lint on filtering media is possible, although the media itself
may not burn. This may cause a substantial increase in filter combustibility. Smoke detectors and fire sprinkler systems may be considered for filter bank locations. In some cases,
depending on the contaminant, hazardous material procedures must be followed during
removal and disposal of the spent filter.
9.3 Humidifiers
The purpose of a humidifier is to maintain, or increase, the relative humidity of the space
being conditioned. In cooler climates, the humidifier also offsets the low moisture content
of the incoming ventilation air. An example of the process is shown in Figure 9-7. The outside condition is 32F and 80% relative humidity and inside condition is 72 and 40% relative humidity. For every pound of outside air, the humidifier must add 0.0066 – 0.003 =
0.0036 pounds of moisture to bring the outside air up to the inside condition. A space
with an outside wall will also be losing moisture by diffusion through the building fabric
and infiltration of outside air. These losses must be offset by humidifying the supply air to
a higher moisture content than the room condition.
The moisture can be added by either evaporating water into the air or by injecting steam
into the air. When water evaporates in the air, the latent heat of evaporation (970 Btu/lb) is
provided by the air that is cooled. In this example, if the outside air is humidified by evaporation, the air would have to be heated to 88.4°F and then evaporative cooling would
reduce the temperature down to 72°F. The alternative is to put the same 970 Btu/lb into
water to produce steam and inject the steam into the air.
The heating energy for humidification can be a very significant part of the load in cold
weather. To keep energy use down, the relative humidity should be progressively lowered
as much as practical in cold weather.
Humidifiers must be installed where the air can absorb the water vapor and not be cooled
below the dewpoint, thereby causing condensation and potential rusting of steel duct and
water dripping from the joints. Note that humidifiers are usually downstream of the system filters, so any rusting of the duct can cause rust flakes that will not be collected by the
filters.
9–17
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Figure 9-7
Humidification of Outside Air
Several industrial/commercial humidifiers are depicted in Figure 9-8. In selecting the type
of humidifier, care should be taken to minimize growth of microbes and contamination
due to chemicals in the water that can become airborne. Note that a requirement of Standard 62-2007 is that the water for humidifiers “shall originate directly from a potable
source or from a source with equal or better water quality.”
Heated pan humidifiers. These units offer a broad capacity range and may be heated by
an electrical element, steam or hot-water coil (see Figure 9-8a). Electric heated pan humidifiers are usually provided with a low water level cutoff switch as a protection device for the
heating elements. Steam coils are commonly used in pan humidifiers. At steam pressures
above 15 psig, moisture carryover occurs because of splashing caused by nucleate boiling.
To prevent boiling over, baffle splash eliminators should be used according to manufacturers’ instructions. Eliminators are essential where steam pressure is greater than 15 psig.
9–18
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Figure 9-8
Humidifiers for Larger Systems
Steam humidifiers. Direct steam injection humidifiers cover a wide range of designs and
capacities. Because water vapor is steam at low pressure and temperature, the whole process
can be simplified by introducing steam directly into the air to be humidified. This method
is essentially an isothermal process because the air temperature remains constant as the
moisture is added in vapor form. The steam control valve may be modulating or two-position in response to a humidity controller. The steam may be either used from an external
source with enclosed grid, cup or jacketed dry steam humidifiers; or produced within the
humidifier, as in the self-contained type. When the steam is supplied from a separate
source at a constant supply pressure, it responds quickly to system demand.
9–19
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Steam units must be installed where the air can absorb the water vapor, otherwise condensation can occur in the duct. For proper psychrometric calculations, refer to the ASHRAE
Handbook–Fundamentals.
•
Enclosed steam grid humidifiers (see Figure 9-8b) should be used on low steam
pressures (under 12 psig) to prevent splashing of condensate in the duct. The
drain should be located on the side opposite the control valve. A drip leg 
dimension H, a minimum of 12 in.  should provide the pressure to flow the
condensate through the trap.
•
A cup or pot-type steam humidifier is usually attached under a system duct.
Steam is attached tangentially to the cup’s inner periphery by one or more
steam inlets, depending on the unit’s capability. The steam supply line should
have a suitable steam trap. There may be a tendency toward supersaturation
due to stratification along the bottom of the duct. Multiple units may be
required to produce satisfactory distribution. Under certain conditions, droplets of condensate may be injected into the airstream.
•
A jacketed steam humidifier uses an integral steam valve with a steam-jacketed
duct-traversing dispersing tube and condensate separator to prevent condensate from being introduced into the airstream (Figure 9-8b). An inverted
bucket-type steam trap is required to drain the separating chamber. This
humidifier may be used without the jacketed tube in nonducted installations.
These humidifiers inject steam directly from the boiler into the space or duct system. Some
boiler treatment chemicals can be discharged, which can affect indoor air quality. Care
should be taken to avoid contamination from boiler water or steam supply chemical treatment. Remember the requirement of Standard 62.1-2007 is that the water for humidifiers
“shall originate directly from a potable source or from a source with equal or better water
quality.”
•
A self-contained steam humidifier converts tap water to steam by electrical
energy using either the electrode boiler principle (Figure 9-8d) or resistance
heating (Figure 9-8e). This steam is injected into the duct system through a dispersion manifold, or the humidifier may be freestanding for nonducted applications.
Atomizing humidifiers with optional filter eliminator (Figures 9-8f, 8h). Centrifugal
atomizers use a high-speed disk that slings water through a fine comb to create a fine mist
that is introduced directly into the air where it is evaporated. The ability of the air to
absorb the moisture depends on temperature, air velocity and moisture content.
Where mineral fallout from hard water is a problem, optional filter eliminators may be
added to remove mineral dust from humidified air, or water demineralizers may be
installed.
9–20
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Additional atomizing methods use nozzles; one uses water pressure and the other uses both
air and water, as shown in Figure 9-8i. Mixing air and water streams at combined pressures
atomizes water into a fine mist, which is evaporated in the room or air duct.
Ultrasonic nozzles place air and water under pressure to atomize the water into a fine mist
(Figure 9-8g). Accurate psychrometric calculations must be made to ensure that the water
droplets are absorbed in the duct airstream.
Wetted element humidifier. Wetted element humidifiers (Figure 9-8j) have a wetted media,
sometimes in modular configurations, through or over which air is circulated to evaporate
water. This unit depends on air flow for evaporation; the rate varies with temperature,
humidity and velocity of the air.
9.4 Duct Heaters
Duct heaters may be either steam, water (hydronic) or electric. They are used for a variety
of purposes including preheating outside air, reheating and making up for heat loss when
the duct is run through an unconditioned space. When using electric in-duct heaters, the
liner must be removed from the area of the heater, and the duct insulated on the outside.
The required output of a duct heater is calculated using:
Btu/h = cfm  1.1  rise in temperature °F
or
kW = (cfm  1.1  rise in temperature °F) / 3,412
For example, a VAV box with maximum capacity of 600 cfm and 300 cfm flow when
reheating is to have a reheat coil installed. The coil design is to raise the air temperature
from 56°F to 82°F. The required capacity is:
cfm  1.1  rise in temperature °F = 300  1.1  (82-56) = 8580 Btu, or 2.5 kW
Electric heaters may also be combined with a VAV outlet for zone control. Here a control
with two minimum positions is used. If the space is too cold, the control closes off the air
flow to some preset amount (for example, 10%). If the space is still too cold, the control
opens up to a second preset amount (for example, 25%), and the heater is turned on. The
advantage of this arrangement is that the heater is required to heat only 25% of the maximum air flow. Most codes will not allow a design where 100% of the air flow is reheated.
9–21
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9.5 Duct Insulation
In all new construction (except low-rise residential buildings), air handling ducts and plenums installed as part of an HVAC system should be thermally insulated in accordance
with ASHRAE Standard 90.1 or local codes. The insulation used should not provide a fire
hazard and conforming to NFPA Standard 90A6 and NFPA Standard 90B7 is advised,
even if not code required.
In low-rise residential buildings, any ducts run through unconditioned areas such as roof
and crawl spaces should be sealed and insulated. A reflective surface on external insulation
significantly reduces heat gain in places with a high radiant temperature, such as outdoors
and in residential roof spaces. In addition to thermal insulating properties, insulation also
provides some degree of sound control.
Duct insulation may be either inside or outside the metal duct. The advantages of inside
duct insulation are that it can be applied by the duct fabricator in one operation and it provides sound attenuation. The disadvantage is that it may be more likely to provide a breeding ground for pathogenic bacteria and fungus.
The advantage of outside insulation of the metal duct is that there is no pressure drop or
breeding ground for bacteria or fungus, and the ductwork around equipment is cleanable
as required by Standard 62.1. The disadvantage is that outside insulation is typically
installed by a different contractor than the duct fabricator.
Ducts that are likely to operate at temperatures below the surrounding dewpoint must be
protected against condensation. Typically, external insulation is used with a vapor retarder
on the outside to minimize moisture entry. This vapor retarder must be well sealed, with
care taken around hangers and flanges. Leaks allow moisture in to wet the insulation and
further lower the insulation value.
The heat transmission (U-factor) for uninsulated sheet metal ducts is affected by air velocity, emittance and duct shape. An approximate value of 1.0 Btu/h·ft2·°F may be used. For
insulated ducts, the heat transmission is reduced by a factor of about 4 for typical 1 in.
insulation and about 8 for 2 in. insulation. A method for determining heat loss or gain for
ducts is given in the HandbookFundamentals.
The Next Step
The noise generated by an air system can range from being very beneficial to very detrimental to the occupied environment. The next chapter starts with an introduction to noise
and a discussion of the benefits and disadvantages of system-produced noise. The rest of
the chapter is devoted to system design to achieve the desired sound criteria.
9–22
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Summary
Parallel-blade and opposed-blade are the standard control and shutoff damper designs.
They may have the linkage attached to the blades in the airstream, or with the linkage driving the damper shafts and out of the airstream. Blade form may be flat for balancing use,
triple-V, or aerofoil for lowest resistance.
Damper performance depends on the particular manufacturer’s design, size relative to the
duct, and any other components or changes in duct shape or direction. Ideally, the air flow
would vary directly with change in damper angle. In practice, this is often not achieved
near fully open or fully closed.
The one situation where parallel blade dampers consistently provide more linear control is
in the mixing box, typically mixing outside air and return air to provide supply air.
Particulate air filters are covered by ASHRAE Standard 52.1 for lower efficiency filters and
ASHRAE Standard 52.2, which grades the full range based on particle size efficiency into
MERV 1 (inefficient) to MERV 20 (super efficient). Due to the different methods of particle capture, the performance varies with face velocity.
Panel and extended surface filters feature fiber panels using fibers of varying materials,
diameter and packing in varying depths to provide a complete range of filtration efficiency.
Performance may be enhanced by using electrostatic fibers or by coating them with a viscous fluid. Filters for finer particles are normally protected from the high mass of larger
particles with a lower performance prefilter.
With renewable media filters, replaceable rolls or a continuous belt of filter media are
moved across the airstream when loaded.
Electronic filters electrically charge and then collect the particles. Their clean efficiency is
high but drops quite quickly. Industrial units are often provided with automatic washing
equipment.
Filter selection depends on many factors including: code requirements, the required filtration performance, filter cost and replacement costs as well as the energy cost in driving the
air through the filter.
Filters take up space within the air-handling unit and also for access. Poor mounting
frames and difficult access causing filter distortion can both seriously reduce the actual performance of the filters.
The purpose of a humidifier is to maintain, or increase, the relative humidity of the space
being conditioned. In cooler climates, the humidifier also offsets the low moisture content
of the incoming ventilation air. The moisture can be added by either having water evapo-
9–23
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rate into the air or by injecting steam into the air, with both methods requiring the latent
heat of evaporation, 970 Btu/lb. To keep energy use down, the relative humidity should be
progressively lowered as much as practical in cold weather.
Humidifiers must be installed where the air can absorb the vapor and will not be cooled
below the dewpoint, thereby causing condensation and potential rusting of steel duct and
water dripping from the joints.
In selecting the type of humidifier, biological and chemical contamination must also be
considered.
Humidifiers are either direct steam injection, steam produced locally from boiling water,
atomizing, or wetted fabric in the airstream. In all cases, the humidification should be to
potable standards.
Duct heaters may be either steam, water (hydronic) or electric. The required output of a
duct heater is calculated using: Btu/h = cfm  1.1  rise in temperature °F
In all new construction (except low-rise residential buildings), air handling ducts and plenums installed as part of an HVAC system should be thermally insulated in accordance
with ASHRAE Standard 90.1 or local codes. The insulation used should not provide a fire
hazard and should conform to NFPA Standards 90A and 90B, even if not code required.
Interior duct insulation provides greater sound reduction, increased air flow resistance, and
more biological contamination problems. The disadvantages of exterior duct insulation are
that it is typically installed by a different contractor than the duct fabricator and sealing
against condensation moisture can be difficult in restricted spaces. Leakage reduces system
performance and maximum rates should be specified.
The heat transmission (U-factor) for uninsulated sheet metal ducts is affected by air velocity, emittance and duct shape. An approximate value of 1.0 Btu/h·ft2·°F may be used for
bare metal, 0.25 Btu/h·ft2 for 1 in. of insulation, and 0.125 Btu/h·ft2 for 2 in. of insulation.
9–24
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Bibliography
1. ASHRAE. 2004. Research Report (RP-1157) Flow Resistance and Modulating Characteristics of
Control Dampers.
2. UL. 2006. UL-555, Standard for Fire Dampers. Northbrook, IL: Underwriters Laboratories.
3. UL. 2006. UL-555C, Standard for Ceiling Dampers. Northbrook, IL: Underwriters Laboratories.
4. UL. 1999. UL-555S, Standard for Smoke Dampers. Northbrook, IL: Underwriters Laboratories.
5. NIOSH. 1973. The Industrial Environment–Its Evaluation and Control. Washington, DC: US
Government Printing Office.
6. NFPA. 2002. NFPA 90AInstallation of Air-Conditioning and Ventilating Systems. Quincy, MA:
National Fire Protection Association.
7. NFPA. 2006. NFPA 90BInstallation of Warm Air Heating and Air-Conditioning Systems.
Quincy, MA: National Fire Protection Association.
ASHRAE. 2007. ASHRAE/IESNA Standard 90.1-2007, Energy Efficient Design of New Buildings
Except Low-Rise Residential Buildings. Atlanta, GA: ASHRAE.
ASHRAE. 2007. ANSI/ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air
Quality. Atlanta, GA: ASHRAE.
ASHRAE. 1992. ASHRAE Standard 52.1-1992, Gravimetric and Dust-Spot Procedures for Testing
Air-Cleaning Devices Used in General Ventilation for Removing Particulate Matter.
ASHRAE. 2007. ASHRAE Standard 52.2-2007, Method of Testing General Ventilation AirCleaning Devices for Removal Efficiency by Particle Size.
ASHRAE Handbook-Fundamentals: dampers, humidity calculations, filters, and duct heat losses
and gains; Handbook-Systems and Equipment: contaminants and filtration
9–25
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Skill Development Exercises for Chapter 9
Complete these questions by writing your answers on the worksheets at the back of this
book.
9–26
9-1
In a smoke control system:
a) A smoke damper inhibits the passage of air that may or may not contain
smoke
b) Moderate leakage of smoke-free air through the damper does not adversely
affect the control of smoke movement
c) Design the system so that only smoke-free air is on the high-pressure side of a
smoke damper, unless the smoke control damper is on the return air
d) All of the above e) None of the above
9-2.
Particles less than _____ in diameter are referred to as the fine mode.
a) 0.75 µm b) 7.5 µm c) 75 µm d) None of the above
9-3.
From an industrial hygiene perspective, particles with an aerodynamic particle
size of _____ or greater are considered the nonrespirable fraction of dust.
a) 5 µm b) 10 µm c) 15 µm d) None of the above
9-4.
____________ measures the ability of the filter to remove particulate matter
from an airstream.
a) Efficiency b) Air flow resistance c) Dust-holding capacity
d) All of the above e) None of the above
9-5.
Different types of filters are distinguished by:
a) Efficiency b) Air flow resistance c) Dust-holding capacity
d) All of the above e) None of the above
9-6.
Filters collect particles by:
a) Straining b) Inertial deposition c) Electrostatic effects
d) All of the above e) None of the above
9-7.
In panel filters, the accumulating dust load causes pressure drop to:
a) Decrease to the filtration load rating, then increase
b) Increase to the filtration load rating, then decrease
c) No effect, remains constant
d) None of the above
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9-8.
Electronic filters, which if maintained properly by regular cleaning, have relatively constant pressure drop and efficiency.
a) True b) False
9-9.
Important requirements of a satisfactory and efficiently operating air filter installation include:
a) Ample capacity for the amount of air and dust load it is expected to handle
b) Suited to the operating conditions
c) Economical for the specific application
d) All of the above e) None of the above
9-10.
Duct heaters may be:
a) Steam b) Water c) Electric d) All of the above
9-11.
The performance of particulate filters is categorized in Standard 52.2 into 20
MERV ratings, with MERV 1 being a coarse screen and MERV 20 being the
high rating filter for demanding cleanroom situations. To control a buildup of
dirt on wet cooling coils, Standard 62 requires what MERV rating filter be
installed before cooling coils that can run wet?
a) MERV 2 b) MERV 6 c) MERV 10 d) MERV 14
9-12.
The parallel blade damper deflects the air in one direction as the air passes
through. This usually makes the performance of a parallel blade damper more
sensitive to location than an opposed blade damper in the same location.
a) True b) False
9–27
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Chapter 10
Sound and Vibration in
Air Systems
Contents of Chapter 10
•
•
•
•
•
•
•
10.1 Fundamentals of Sound
10.2 Sound and Vibration Sources
10.3 Sound Attenuation
10.4 Vibration Control
Summary
Bibliography
Skill Development Exercises for Chapter 10
10–1
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Study Objectives of Chapter 10
After completing this chapter, you should be able to:
•
Explain the fundamentals of sound and sound transmission that are relevant to
air system design
•
List and explain the major sound sources and absorbers found in an air system
•
Understand the calculation of sound reduction methods
10.1 Fundamentals of Sound
Sound is vibration; the vibration of fluids such as air and water and the vibration of solids
such as framed walls and rotating equipment. In conversations, sound is often considered
as being what we hear in the air around us. The typical human hearing range is from to 20
to 20,000 cycles per second, or Hertz (Hz). In contrast, vibration is restricted to motion we
can feel or see, the very low end of our hearing range and lower frequencies. Thus, someone standing beside an air-handling unit and touching it might say, “This unit is noisy and
it vibrates a lot.” They are commenting on one physical phenomena, sound, but they are
differentiating between the higher frequencies they can hear and the lower frequencies they
can feel.
In this chapter, the words sound and vibration will be discussed as they are commonly used.
But remember that sound and vibration are not separate phenomena. For example, consider a drum in a room with stud and gypsum board walls. When the drum is played, you
can hear the sound. If you place your hand on the wall, you feel it vibrating. The vibration
is set up by the low frequency sound from the drum. You are hearing the higher frequencies and feeling the lower frequencies of the sound produced by the drum.
Sound behavior is very different from, and more complex than, other types of energy.
Therefore, this chapter will provide general advice and an introduction to sound assessment. In the next chapter, sound assessment and troubleshooting existing situations will be
covered.
Sound is generated by a vibrating surface or a turbulent fluid stream. In an air-handling
system, the fans with their motors, drives and blades are the common surface generator.
The air flowing through contorted duct paths and leaking through holes produces turbulent stream noise. The generated sound can be transmitted as air-borne sound and structure-borne sound. Following are several terms about sound:
10–2
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•
Speed. The speed of a longitudinal wave in a fluid medium is a function of the
medium’s density and modulus of elasticity. In air at room temperature, the
speed of sound is about 1,100 ft/s; in water, about 5,000 ft/s.
•
Frequency. Frequency is the number of oscillations (or cycles) per second completed by a vibrating object. The international unit for frequency is the Hertz
(Hz). Natural sounds have a range of frequencies. For example, when talking,
the voice frequencies typically cover the range from 200 Hz to 8,000 Hz (8
kHz), with the hardwired telephone about 300 Hz to 3 kHz.
•
Sound Pressure. The human ear and microphones are sensitive to the pressure
changes in a sound. The threshold of excellent youthful hearing at 1 kHz has
been adopted as the international standard as 20 µPa. Sound pressure measurements are all related to this base level. Unlike heat and light, adding two equal
sound pressures does not produce twice the sound pressure. Thus, a different
scale called the decibel is used.
•
Decibel. The sound pressure decibel is defined as 10 times the base 10 logarithm of the ratio of the actual sound to the reference (whisper) sound. Lp = 10
log(p/pref). A buzzing insect at 3 feet is perceptible at a level of about 20 dB,
often referred to as a sound level of 20 dB. To be correct, the level should
include the reference level, 20 dB re 20 µPa. Table 10-1 shows the sound pressure and sound pressure level (dB) for a wide range of sources.
Table 10-1 Typical Sound Pressures and Pressure Levels
10–3
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•
Octave band. Each particular sound will have varying pressures at different frequencies. Thus, a deep voice will have relatively more sound power at lower
frequencies than a high pitched voice. To represent this variation, sound measurements are done in octave bands. An octave band is a frequency band with
an upper frequency limit of twice the lower frequency limit. The standard center frequencies for octave bands are 63 Hz, 125 Hz, 250 Hz, 500 Hz, 1kHz, 2
kHz, 4 kHz, 8 kHz and 16 kHz. Our ears increase in sensitivity up to about 1
kHz and then drop off again above 8 kHz. The sensitivity decreases with age.
Thus, it is possible for a child to complain about a 20 kHz noise that is inaudible to many adults due to its high frequency.
A meter calibrated to behave like the human ear is called A-weighted and its response curve
is shown in Figure 10-1. Also shown is the almost, but not flat, C-weighting curve. The
simplest sound level meters have the A-weighting built in. More sophisticated meters have
the C-weighting and octave band filters. These measure the sound pressure level at individual octave bands. This ability to measure the levels in the octave bands can help detect
the source of a noise. Figure 10-2 shows the frequency range of some sources with their
descriptive characteristic. Figure 10-3 shows the descriptive words for various sounds and
their likely equipment cause.
Figure 10-1
10–4
A and C Sound Pressure Curves
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Figure 10-2
Figure 10-3
Mechanical System Component Ranges of Predominant Sound
Frequency Ranges of Likely Sources of Sound-Related Complaints
10–5
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•
Wavelength. Wavelength is the distance between successive rarefactions or compressions due to the sound wave. Wavelength, speed and frequency are interrelated by the following equation:
 = c f
(10-1)
where,
  = wavelength, ft
c = speed of sound, ft/s
f = frequency, Hz
•
Noise. The first and simplest definition of noise is any unwanted sound. The
second definition is that noise is broadband sound with no distinguishable frequency characteristics, such as the sound of a waterfall. The second definition
is appropriate when one sound is used to mask another, as when controlled
sound radiated into a room from a well-designed air-conditioning system is
used to mask or hide low-level intrusive sounds from adjacent spaces to
increase privacy. This controlled sound is called noise, but not in the context of
unwanted sound. Rather, it is a broadband, bland sound that is frequently
unobtrusive; it is sometimes called white noise.
A unobtrusive steady sound (no pulsing or on/off) has: a spectrum that sounds about equal
loudness in each frequency band; no detectable specific tones; and a loudness that is relatively quiet in the particular environment.
A series of standard curves of roughly equal loudness called noise criteria (NC) is shown in
Figure 10-4. These curves are commonly used to specify the maximum loudness of a noise
in each frequency band. Thus, specifying NC35 for an office is to require the loudness in
each frequency band to be no higher than defined by the NC35 line at any frequency.
There is no requirement for the noise to have a spectrum matching the curve. Thus, a noise
that peaks at the NC curve at a high frequency will sound hissy. An example of an NC43
hissy sound is shown by the bold line in Figure 10-4. A sound that peaks at the NC curve at
a low frequency will sound boomy, or rumbly. In addition, the total sound is, on average,
lower than the curve, so the noise may not mask as well as anticipated.
There are two conflicting aspects to the required level of noise. First, in many situations,
enough background noise is needed to mask, at least partially, unwanted noises. Thus, in a
large open office, the background noise should be loud enough to mask nearby phone conversations. Another example is the downtown hotel where the noise from the HVAC system can mask traffic noise.
10–6
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Noise Criteria, NC Curves
Figure 10-4
10–7
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The second issue is having the noise level low enough to avoid discomfort and obstruction
of activities. In a concert hall, a very low noise level from the HVAC system is required so
every nuance of the music can be heard. Offices should be quiet enough to comfortably
talk to another person or on the telephone. These more noise/less noise needs lead to different ideals in apparently similar situations, offices for example. Here are ASHRAE suggested levels for office situations:
Rooms with frequent teleconferencing: NC20 to NC25
Private offices and conference rooms: NC25 to NC35
Open-plan office areas: NC35 to NC40
Corridors, hallways: NC35 to NC45
The teleconferencing situation requires minimum background noise to maximize the quality of the microphone pickup. This is in contrast to the open-plan office requiring a masking background level.
The use of the NC curves is far from satisfactory. Unfortunately, the modern methods of
defining noise criteria more effectively using Room Criteria Method (RC) and RC Mark II
are not simple to explain or use. These two methods set out to quantify overall level and
the effect of higher and lower levels in the higher and lower frequency bands.
10–8
•
Sound intensity. Sound intensity is a measure of the power in a sound per square
meter. Sound intensity follows the inverse square law; that is, sound intensity
varies inversely as the square of distance from the source. This is reasonably
true for acoustical sources outdoors but indoors, reflections significantly mask
the effect. Sound intensity level is expressed in dB, with a reference quantity of
10-12 W/m2.
•
Sound power and sound power level. A fundamental characteristic of an acoustic
source is its ability to radiate energy, whether weak and small in size (a cricket)
or strong and large (a compressor). Some energy input excites the source,
which radiates some fraction of this energy in the form of sound. Because unit
power radiated through a sphere one meter in diameter yields unit intensity,
the power reference base, established by international agreement, is 1 picowatt
(pW) which is 10-12 W.
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10.2 Sound and Vibration Sources
Sound and vibration sources are either actively powered such as fans and compressors or
passive (aerodynamic) such as turbulence and vibration in ductwork. In an air system, a fan
generates sound that propagates out of its supply outlet, its inlet, and the casing. The
mechanical engineer should make preliminary equipment selections as soon as possible to
allow for a preliminary noise analysis and to determine the probable required sizes of the
mechanical rooms.
Each type of HVAC system has its own set of layout and operating features that determine
which noise and vibration control measures are most effective. Therefore, the choice of one
system type over another should not be made without considering the cost of controlling
noise and vibration. For example, consider a midrise office building with a built-up penthouse fan system, which would generate most of its noise and vibration in the penthouse
and therefore, would require the most attention there. Some care should also be taken in
designing the supply duct takeoff and return air opening at each floor.
Conversely, a system using water-source heat pump units distributed throughout the
building’s ceiling plenums would require less concern for noise and vibration control at the
central plant, but much more care in the selection, placement and installation of the dozens of noise and vibration sources (the heat pump units).
Figure 10-5 shows several problems that can create sound and vibration issues. The problems have been grouped under four headings that will be considered later in more detail.
Figure 10-6 shows an optimal mechanical room with the features grouped under the same
four headings.
The most obvious difference between the two figures is the optimal room is larger and
more massive. Unfortunately, extra space and masonry walls are often not readily available
to the air-system designer. The challenge is then to choose the equipment and layout in the
space available to be as satisfactory as possible. As you look at the two situations, consider
how you could re-layout the equipment in the first figure to avoid several of the problems
even if none, or only some, of the additional space were available.
10–9
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Figure 10-5
Mechanical Room with Common Sound Problems
Supply air path to occupied space
2. The counterclockwise rotation of the fan's discharge airstream is forced to change its spin direction at the downstream elbow. The turbulence generated at the change can produce unstable flow with a very high, fluctuating pressure drop, thereby resulting in fan instability that is heard as rumble.
3. Problem 2 is aggravated if the elbow's turning vanes do not have long trailing edges to straighten the airflow and
control the turbulence.
4. The duct sound trap (silencer) is too close to the elbow. This compounds the turbulence problem.
Breakout sound from supply duct
11. Ductwall vibration in the duct silencer (or any other part of the trunk duct system) touching the drywall partition
can cause the partition to act as a sounding board and radiate low frequency noise into the occupied space.
5. Rectangular ductwork and duct silencers do not control the rumble produced by turbulent airflow.
12. Suspending the dropped ceiling from the supply duct causes the ceiling to be a sound radiator.
Return air path
7. The lack of a duct silencer in a mechanical room return air opening allows fan noise to travel into the ceiling cavity, then through the lightweight acoustical ceiling into the occupied space.
Breakout sound from mechanical room
1. AHU panel vibration “couples” (vibrates in sympathy) to the lightweight, flexible gypsum wall just a few inches
away. This coupling lets low frequency noise pass easily through the wall.
6. The AHU's air inlet is too close to the wall. This causes two acoustical problems: unstable fan operation leading
to surge and rumble, and direct exposure of the inlet noise to the mechanical room wall.
8. The unit is resting on thin cork/neoprene isolation pads that are too stiff to adequately isolate the fan vibration.
9. The poorly isolated unit is resting on a relatively flexible floor slab without sufficient structural support. This
arrangement allows unit vibration to enter the slab.
10. The chilled water piping is rigidly attached to the slab above, thereby letting unit vibration enter the slab.
10–10
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Figure 10-6
Mechanical Room with Optimal Acoustical Features
Supply air path to occupied space
2. Use of a horizontal discharge AHU eliminates the need for a turbulence-producing airflow.
3. Gradual transition at AHU outlet minimizes turbulence.
4. Duct silencer is far enough away from AHU outlet to avoid excessive regenerated noise and turbulence.
Breakout sound from supply duct
5. Circular ductwork controls the transmission of low frequency noise and rumble into the occupied space.
11. The supply trunk duct does not touch the wall. A 1/2 in. gap surrounding the duct is filled with a non-hardening
sealant.
12. Ceiling not suspended from supply duct.
Return air path
7. The return air duct silencer controls AHU noise via the return air path.
Breakout sound from mechanical room
1. Keeping a minimum 2 ft clearance reduces coupling between AHU and wall. Masonry wall provides excellent low
frequency sound isolation as long as it is well sealed at the ends and against the ceiling slab.
6. The large clearance at the AHU inlet keeps the unit away from the wall and avoids excessive inlet turbulence.
8. The unit is resting on high-deflection, steel spring vibration isolators.
9. The floor assembly supporting the unit has a housekeeping pad and at least one major beam under the unit. Additional stiffness and mass both help to control the transmission of unit vibration into the slab.
10. The chilled water pipes are suspended by vibration isolation hangers.
10–11
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10.3 Sound Attenuation
Having introduced the issues in Figures 10-5 and 10-6, the four general challenges are:
•
Supply air path to the occupied space
•
Breakout sound from the supply duct
•
Return air path
•
Breakout sound from the mechanical room
SUPPLY AIR PATH TO THE OCCUPIED SPACE
For air systems, the critical sound generator is the fan. Fan noise depends on fan design,
the volume flow, the total fan pressure, and inlet and outlet (system) conditions. Most
HVAC fans are made in a range of sizes, and several fans in the range will provide the
required flow and total pressure. In general, the fan with the highest efficiency will be the
quietest. Metaphorically, think of the situation as the smallest fan screaming as it rotates
very fast to achieve the duty, the efficient fan just quietly doing its job, and the very large
fan grumbling that it really is bored with so little to do. Needless to say, choose the efficient
fan so the mechanical room does not have to be built to contain the screams or grumbling.
The sound power produced by an installed fan depends on the fan and the inlet and outlet
conditions, or system effects. Therefore, the first step in choosing a fan is to assess the available space and layout possibilities so the most suitable type and fan arrangement can be
chosen. For centrifugal fans, this is particularly important at the outlet as the air velocity is
much higher at the outside of the volute. Note that poor supply duct arrangements with
rectangular duct can lead to severe duct rumble in adjacent occupied spaces. Installations
such as shown on the right in Figure 10-7 create significant noise and pressure drop which
can be largely avoided by having the fan connected as shown on the left of the figure.
Figure 10-7
10–12
Improved Air Flow by Design of Fan Delivery
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Having selected the type of fan, next choose the most efficient fan at the desired duty.
With VAV systems, aim for maximum efficiency at 80% flow because the system will
rarely be at 100% flow. Use the fan manufacturer’s data, which has been obtained by a recognized test method; normally, AMCA Standard 301 Methods for Calculating Fan Sound
Ratings from Laboratory Data. Note, manufacturers do not test every size of every fan at
every duty to produce the data. Limited testing is done and the results are interpolated to
produce the extensive data tables. As a result, do not assume particular data are correct to
better than 3 dB.
Having chosen the fan and obtained the inlet and outlet sound spectrums, the effect of the
sound travel along the duct and into the occupied space can be assessed. The elements
along the duct to consider are plenum, straight duct, silencers, branches, elbows, outlets
and the occupied space.
Plenum
Plenums  lined cavities  are often used to reduce the fan noise entering the distribution
system (Figure 10-8). Significant pressure loss occurs through the plenum and this disadvantage must be balanced against other ways of ducting the air and providing sound control.
Figure 10-8
Schematic of a Plenum Chamber
10–13
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Sound attenuation can be calculated using Equation 10-2:
 Q cos  1 –  A 
- + --------------- 
TL = -10log 10 S out  --------------S A 
 4r 2
(10-2)
Sout = area of plenum outlet, ft2
S
r
Q
A
= total inside surface area of plenum minus inlet and outlet areas, ft2
= distance between the centers of inlet and outlet of plenum, ft
= directivity factor, which may be taken as 4
= average absorption coefficient of plenum lining

= angle of vector representing r to long axis l of duct (see equation below)
l
l
cos  = - = ------------------------------r
2
2
2
l + rv + rh
(10-3)
l = length of plenum, ft
rv = vertical offset between axes of plenum inlet and outlet, ft
rh = horizontal offset between axes of plenum inlet and outlet, ft
Note that absorbency is the average in the plenum calculated as the sum total of each material’s sound-absorbent coefficient times its area divided by the total area. Typical absorbencies are shown in Table 10-2. Note that increasing the fiberglass sound absorbency is effective at the low frequencies but not at higher frequencies.
Table 10-2 Sound Absorption Coefficients of Selected Plenum Materials
63
Octave Midband Frequency (Hz)
125
250
500 1000 2000
Non-Sound-Absorbing Material
Concrete
0.01 0.01 0.01 0.02
Bare sheet metal
0.04 0.04 0.04 0.05
Sound-Absorbing Material (Fiberglass Insulation Board)
0.05 0.11 0.28 0.68
25 mm, 48 kg/m3
0.02
0.05
0.02
0.05
0.03
0.07
0.90
0.93
0.96
50 mm, 48 kg/m3
0.10
0.17
0.86
1.00
1.00
1.00
1.00
75 mm, 48 kg/m3
0.30
0.53
1.00
1.00
1.00
1.00
1.00
100 mm, 48 kg/m3
0.50
0.84
1.00
1.00
1.00
1.00
0.97
Note: The 63 Hz values are estimated from higher frequency values.
10–14
4000
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Straight duct
Ductwork transmits and attenuates the fan noise. If the air velocity is high, aerodynamic
noise can also be generated. Table 10-3 provides recommended maximum air velocities for
specific NC targets. Note that the table indicates a decreasing maximum velocity as the air
moves from the main duct to branches to outlets. This progressive reduction in velocity
occurs automatically if the design method of constant pressure drop per unit length (e.g.,
0.1 in. wg/100 ft) is used. This method lines up reasonably well with the Table 10-3 criteria for NC35 spaces.
Table 10-3 Recommended Maximum Duct Velocities for Design NC Conditions
Main Duct Location
Design
RC (N)
Maximum Airflow
Velocity (m/s)
Rectangular
Circular
Duct
Duct
In shaft or above drywall ceiling
45
35
25
17.8
12.7
8.6
25.4
17.8
12.7
Above suspended acoustic ceiling
45
35
25
12.7
8.9
6.1
22.9
15.2
10.2
Duct located within occupied space
45
35
25
10.2
7.4
4.8
19.8
13.2
8.6
Notes:
1. Branch ducts should have airflow velocities of about 80% of the values listed.
2. Velocities in final runouts to outlets should be 50% of the values or less.
3. Elbows and other fittings can increase airflow noise substantially, depending on the type.
Thus, duct airflow velocities should be reduced accordingly.
10–15
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Sheet metal duct provides some sound attenuation. Table 10-4 shows the attenuation per
foot for some square ducts based on tests done with 10-ft lined sections. The data show the
attenuation for bare metal, with 1 in. of fiberglass insulation board and with 2 in. of fiberglass insulation board. In each case, the clear passage is maintained. So with 2 in. of fiberglass, the sheet metal outside dimensions will be increased by 4 in. on each side. Look at
the table and note that:
•
The attenuation peaks in the 1 kHz to 2 kHz range
•
Increasing the absorber thickness improves low frequency attenuation but has
no effect on higher frequency attenuation
•
Attenuation is much greater in ducts with a high perimeter to area (smaller or
long and narrow), a factor that will be mentioned again regarding passive
silencers
Table 10-4 Square Duct Attenuation Data
Table 10-5 shows the attenuation data for a selection of round ducts. Again the peak attenuation is in the 1-2 kHz range, and increasing the absorber thickness is only effective for
lower frequencies. Also note that the attenuation of lower frequencies in bare round ducts
is very small, a tenth of the attenuation in rectangular ducts. This is due to the inability of
the round duct metal to vibrate, absorb and radiate-out the low frequencies.
10–16
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Table 10-5
Round Duct Attenuation Data
The final connection to ceiling outlets is often conveniently achieved using nonmetallic
insulated flexible duct. The flexible duct allows for dimensional flexibility between the
ductwork installer and the ceiling layout. Their length is normally in the range of 3 to 6
feet to provide the flexibility but restrict the higher pressure loss of the ducting. The installation should keep the duct straight with long radius bends. Any abrupt bend, or offset at
the outlet, can produce unacceptable air noise and flow restriction. The attenuation for
typical lined flexible ducts is shown in Table 10-6.
Table 10-6 Insertion Loss for Lined Flexible Ducts
10–17
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Silencers
At times, the attenuation in the ductwork is not sufficient or the sound must be contained
in the mechanical room, so an extra attenuation unit called a silencer or muffler is included
in the duct run. The silencer has a much higher attenuation than the same length of
ductwork. The actual attenuation required for the silencer is based on the complete sound
assessment from fan to listener. The required insertion for the silencer is the calculated levels at each frequency arriving at the listener less the desired maximum level. There are three
types of silencer: dissipative silencers, reactive silencers and active silencers.
Dissipative silencers are passive and use sound-absorbing media similar to duct lining. They
take advantage of the fact that attenuation is greater in smaller ducts. Therefore, they are
designed with narrower air paths and, consequently, have a higher air velocity and resistance through them. A schematic section through an absorptive silencer is shown in Figure
10-9. Note that friction is increased due to the increased surface area and the air velocity is
higher due to the obstruction of the duct cross-section. To avoid the absorptive material
being eroded, these silencers have a perforated metal cover over the absorber.
Figure 10-9
Cross-section of Rectangular Dissipative Silencer
The performance of a silencer is defined in two ways. First, it absorbs sound, and the manufacturer provides a dB insertion loss for each octave band. These insertion losses are simply deducted from the entering sound power levels to give the reduced sound power level.
Second, the silencer may generate noise due to the air velocity over the restricted passage
and absorber cover sheet. This regenerated noise is given as sound power levels in each
10–18
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octave band. Unfortunately, the sound powers cannot be arithmetically added due to the
logarithmic nature of the dB unit. The combined effect can be calculated, but it is easier to
use Table 10-7 which is adequately accurate for most HVAC situations.
Table 10-7 Combining Two Sound Levels
Difference between two levels to
be combined
Number of decibels to be added
to highest level to obtain
combined level
0 to 1
2 to 4
5 to 9
<9
3
2
1
0
Table 10-8 provides an example to show the effect of adding a high pressure silencer. First,
the dB insertion loss is deducted and then the regenerated sound is combined using data
from Table 10-7 to produce an output sound power in each octave band.
Table 10-8 Silencer Attenuation and Sound Regeneration
If you review the numbers above, you can see that the regenerated silencer noise is predominant at only the two lowest frequencies.
Reactive silencers are constructed only of metal with cavities that act as tuned resonators to
cancel sounds at specific frequencies. They have the advantage of no material that can shed
fibers, but their performance is lower than a dissipative unit for ventilation systems, as
shown in Figure 10-10. The reactive silencer is particularly valuable for reducing low frequency sound from equipment such as compressors and reciprocating engines; a car muffler is a common example.
10–19
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Figure 10-10
Performance Comparison of Dissipative and Reactive Silencers
Active silencers have a microphone to measure the noise and a loudspeaker that delivers the
same sound but exactly out of phase. So the incoming noise and speaker noise cancel each
other. This is often referred to as a noise canceling system. Noise canceling works well at
low frequencies, so a combination of active and dissipative silencers can provide a very
effective full spectrum silencer combination. The system, particularly the pickup microphone, needs a steady, non-turbulent airflow at a velocity of below 1,500 fpm for good
performance.
General guidelines for locating duct silencers in duct systems with fans, for minimum
static pressure drop with maximum acoustical performance, are as follows:
Centrifugal and axial fans:
•
From fan discharge  1.0 duct diameter for every 1,000 fpm
•
From fan intake  0.75 duct diameter for every 1,000 fpm
Duct elbows:
•
Three duct diameters, or equivalent, both upstream and downstream
Mixing boxes and VAV terminals:
•
10–20
One duct diameter upstream or downstream
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Branches
At a branch, the sound energy is divided between the branches in proportion to their cross-sectional areas. Table 10-9 shows the attenuation into a branch duct based on the branch
cross-sectional area as compared to the sum of all branch areas.
Table 10-9
Elbows
Elbows provide some attenuation as the sound changes direction. Table 10-10 has data for
square and radiused elbows. The attenuation of a lined square elbow is often used in return
air sound traps where two lined square bends are connected to form a U. The trap is
mounted through a wall with the open ends facing upwards. Where the local codes allow
the return air to pass from the occupied room into the corridor, this is an economical
method of allowing free return air passage but preventing normal speech being heard
through the sound trap.
Table 10-10 Square Elbow Attenuation
10–21
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Outlets
At the supply outlet, the sound transmission is affected in two ways. First, some of the
sound is reflected back into the duct. Second, some noise is generated by the outlet diffuser
or grill.
The Occupied Space
When the sound enters the occupied space, it spreads out through the space and reflects
back and forth being attenuated (dying away). Thus, the sound level is the sum of the
direct sound from the inlet and reflected sound from the surfaces in the space. In a typical
room with furniture, carpet or dropped ceiling, and blinds, the reflectivity is quite low and
the reflected sound is not obvious. In a newly finished room with flat walls with no furniture or finishes, the average reflectivity may be high and the room is reverberant, or echoes.
Sound level measurements should not be made before furnishing unless allowance is made
for the reverberant situation.
The dB sound values calculated as entering through the diffuser are sound power levels.
We hear and measure sound pressure levels, so the power must be converted to pressure:
Lp = Lw + A – B
(10-4)
where,
Lp = sound pressure level at specified distance from sound source, dB
Lw = sound power level of sound source, dB
A, B see Tables 10-11 and 10-12
Table 10-11 Values for A in Equation 10-4
Room
Volume
(m3)
42
71
113
170
283
425
10–22
63
125
4
3
3
2
1
1
2
Value for A (dB)
Octave Midband Frequency (Hz)
250
500
1000
2000
1
0
1
2
1
0
1
2
1
2
3
1
2
3
4
2
3
4
5
3
4
5
6
4000
2
3
4
5
6
7
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Table 10-12 Values for B in Equation 10-4
Distance from
Sound Source (m)
Value for B (dB)
0.9
1.2
1.5
1.8
2.4
3.0
4.0
4.9
6.1
5
6
7
8
9
10
11
12
13
From tests in real rooms, the decrease in sound level is about 3 dB for every doubling of
distance from the source.
BREAKOUT SOUND FROM THE SUPPLY DUCT
In the last section, the changes in sound from the main sound generator to the occupied
space through ductwork and fittings were covered. As the duct passes through other spaces,
breakout sound is radiated from the ductwork. The two sources of breakout sound are fan
noise in the duct and airstream turbulence.
For fan noise, tables of transmission loss (attenuation) from inside the duct to outside are
used to estimate the radiated sound power. Then other tables are used to assess the sound
pressure at a distance from the duct. If the duct is above a false ceiling, an additional attenuation for the ceiling must be factored into the final sound pressure.
Note, a suspended ceiling must not be directly, or indirectly, hung from the ductwork.
This requirement should be clearly stated in the contract for the ceiling installation.
Duct breakout noise due to airstream turbulence is a low frequency rumble, typically 16 to
100 Hz. It is produced by dramatic changes in airflow direction near the fan and by large
(over 48 in.) unreinforced duct walls near the fan. Ideally, this problem is resolved by modifying the source. In some cases, rectangular duct vibration can be adequately damped by
adding reinforcement or bonding drywall sheets to the duct. Encasing the duct is difficult
due to the low frequency, which requires both a massive enclosure and deep absorber.
Because round duct walls are very stiff, rumble breakout is not a problem with round duct.
However, a length of round duct will often carry the turbulence. Thus, a round main with
rectangular branches may experience the rumble in the branches.
10–23
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RETURN AIR PATH
Sound travels back along the return air path in just the same way as it travels along the supply air path. The manufacturer should provide the fan inlet sound spectrum, which will be
lower than the supply spectrum. But using the supply spectrum will be a conservative
choice if specific data are not available.
An open air return situation was depicted in Figure 10-5. The mechanical room will act as
a sound plenum. As concrete and gypsum board have very low absorbencies, the room will
not act as an effective silencer. Lining the walls is generally not as practical, or effective, as
providing a lined duct sound trap or silencer at the intake to the mechanical room, as
depicted in Figure 10-6.
BREAKOUT SOUND FROM THE MECHANICAL ROOM
During the initial building planning stages, the location of plant and the noise separations
should be carefully considered. Arranging for the mechanical rooms to be well away from
noise sensitive areas is the first choice. Then choose a location with as few transmission
paths as possible.
Basement space can be a good choice as the surrounding structure is massive and the closest noise sensitive areas are above the ceiling and not adjacent. Be careful to allow adequate
space for plant replacement. On the penthouse or roof, the structure is usually relatively
light and vibration isolation can be a challenge.
It is tempting to place the mechanical room in the building core surrounded by elevators,
washrooms, power and communications equipment rooms, and service riser shafts. Surrounded by non-critical areas is good, but the restrictions on how ductwork gets into and
out of the mechanical room may create convoluted high-velocity ductworkthat generates
duct noise as well as requiring a noisier fan.
Economic pressure to maximize a building’s rentable space has resulted in less space being
available for the HVAC system and other building services. This reduction in room size
often forces the mechanical engineer to select small, inefficient equipment or to shoehorn
properly sized equipment into a restricted space. Both options can lead to excessive noise.
To minimize the possibility of this problem, mechanical rooms should be sized as follows.
A mechanical room housing a fan or AHU with an unducted intake should have a floor
area of 1015 ft2 for each 1,000 cfm of air flow. This allows adequate space for proper air
flow into the fan, for low-noise supply duct fittings, and for duct silencers, if required.
All HVAC equipment rooms should have a floor area large enough to allow a clearance of
at least 2 feet around all equipment, with more clearance at the piping, drive and filter
areas. Building codes require a minimum 3-ft clearance in some cases.
10–24
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These are all internal considerations. Now consider the situation outside the building.
Noise can, and will, emanate from the building intakes and exhausts. Assess whether these
noise sources will be limited by code, the neighborhood or client use. Using the required
levels, assess what measures must be taken to control the sound. Just as a sound trap was
required to prevent the return air from being an objectionable sound path, so a sound trap
may be needed to reduce intake or exhaust noise.
The noise from the mechanical room has two sources: radiated noise and equipment vibration. Radiated noise will pass to the surrounding spaces through any air gaps (flanking
paths) around partition walls, doors, pipes, ducts or conduits. Figure 10-11 shows typical
details for sealing around ducts and pipes. The caulking used must be a non-hardening
acoustic sealant that does not transmit the vibration and can flex as needed.
Figure 10-11
Duct, Conduit and Pipe Penetration Details
Because sealing every leak is an arduous and often poorly executed task, avoid having any
ducts, pipes or conduits passing through the mechanical room. This also prevents the relocation of the air-handling unit because a pipe riser has already been installed in a “convenient” location for the installer. Doors can be weathersealed to avoid noise leakage around
their frames. If the mechanical room is at negative pressure, an outward opening door
should be chosen so that it is sucked against the seals.
Having sealed all the leaks, sound can still radiate through the doors, walls, floor and ceiling. Usually the doors and walls are the main challenge, especially if the walls are framed
and gypsum boarded. The transmission may be substantially increased by installing the
AHU closer than 2 feet from the wall. If the AHU is close to the wall, the air between the
vibrating casing and the wall can act as a pulsing plunger to actively vibrate the wall.
10–25
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Wall constructions have sound transmission class (STC) ratings. The ratings are the mean
attenuation over the speech frequencies from 125 Hz and up, but they tend to overrate the
performance at low frequencies. For low frequency attenuation, mass is the important factor, so masonry is better than frame and gypsum board even for the same STC.
Having the air intake facing a wall can also produce wall vibration. It is wise to ensure that
the intake is as far from the wall as the intake is high. If the intake into the fan is too close
to the wall, there will also be system effect, because the air is not flowing evenly into the fan
inlet. As a result, the fan will run faster (and noisier) and produce turbulence in the fan
which produces low frequency surging.
10.4 Vibration
Having dealt with the radiated sound, now consider the vibration produced by the equipment. Vibration is conducted through solid materials and is generally a low frequency
issue. Effective vibration isolation requires all components to have a flexible connection
between the vibrating equipment and the building fabric. This flexible connection allows
the equipment to vibrate, but transmits little vibration energy. Isolation depends on the
equipment, the weight of the equipment including the base, and the mass and stiffness of
the supporting structure.
In general, the objective is to have an isolator mount that deflects substantially more than
the deflection of the supporting structure. As structures are being designed to be lighter
and more equipment is being roof mounted, the challenges of effective vibration control
are increasing.
Ductwork should have a flexible, typically rubberized canvas, connection. Conduit needs a
loop of flexible conduit and pipes should have resilient hangers. The main equipment will
need some form of vibration reduction mount, which will depend on the mass and stiffness
of the supporting floor, ceiling or roof and the weight and vibration produced by the
equipment. For a heavy concrete slab on grade, a resilient pad will often be adequate.
At the other extreme is the rooftop AHU on a light metal truss roof where structural reinforcement and sophisticated anti-vibration measures may be needed to avoid vibration
problems. In general, position rooftop units as close to columns as possible to minimize
flexing the roof structure and get professional advice in conjunction with the structural
designer.
The ASHRAE HandbookApplications has general suggestions as to the type of vibration
isolation to use in a variety of cases. However, it may be best to contact a manufacturer for
specific proposals.
10–26
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Correcting a noise or vibration problem usually costs much more than preventing one.
The real costs are not only the direct payments to the retrofitting contractor; they also
include the time required to coordinate the investigation and retrofit, as well as the loss of
goodwill from the complaining occupants. Therefore, in most cases, the slight extra cost
for prevention (usually about 1% of the total HVAC system cost) is money well spent.
Specifying quiet equipment and adding noise control materials to an HVAC system are
necessary parts of the design process because they help control noise and vibration. Calculations can be used to estimate the sound levels in a room or to select noise control materials to achieve a design goal. Comparing manufacturers’ sound data can help in the selection of quiet equipment. But design decisions based on such work lose their value if the
equipment and materials are not integrated into a properly designed, and installed, system.
The Next Step
The next chapter covers system startup when the system is set working, adjusted and balanced to work as designed. It also covers some diagnostic methods for problem solving
once the system is running.
Summary
Sound is vibration in solids and fluids. Typically, we can hear sound in the air in the frequency range of 200 to 20 kHz and feel the sound vibrations at lower frequencies. The
human ear and microphones sense sound pressure, measured in decibels (dB). Because the
decibel is a logarithmic unit, adding sound powers is not simple and adding two equal
sound powers increases the dB level +3 dB. Sound power is typically measured in octave
band levels. The ear is less sensitive at lower frequencies, so a sound of equal loudness
through the audible range will have a decreasing sound power as the frequency band rises.
This is approximately shown in the commonly used NC scale.
Noise may be unwanted due to annoyance or wanted to mask other obtrusive sounds. To
be unobtrusive, noise must have a reasonably equal loudness at audible frequencies, and
not have specific detectable tones or a pulsing quality. Thus, choosing sound levels in
spaces will often be a compromise between a low level for easy hearing and a higher level to
provide privacy.
Equipment produces sound power that radiates away in all directions. The sound power
passing through an area of 1 m2 is the sound intensity. Both sound power and intensity are
measured in dB relative to a reference level.
10–27
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Sound and vibration are produced by equipment and the turbulent passage of air, called
regenerated noise. Fans normally produce minimum noise at their maximum efficiency, so
selecting the fan size for maximum efficiency typically is the quietest choice. Fan noise is
increased by poor inlet and outlet arrangements. For minimum noise, design for minimum turbulence in the system.
The sound power produced by a fan is attenuated by the ductwork and fittings. The flat
sheet metal of rectangular duct vibrates, attenuating the sound. Round duct cannot vibrate
significantly and does little attenuation. Attenuation can be increased by lining the duct
and including purpose-made silencers. Duct lining is typically 1 or more inches thick.
Increasing above 1 in. improves absorption for lower frequencies.
Silencers may be dissipative containing absorbent materials, reactive with tuned cavities, or
active where sound exactly out of phase is added to cancel the lower frequencies. The final
connection to an outlet is often made with a short length of lined duct. The outlet produces some regenerated sound and diffuser selection is made both on the required air flow
performance and required sound level in the space.
The sound pressure level in the space is calculated by taking the sound power from the fan
with additions due to regenerated noise and deductions by attenuation to find the sound
power into the space. This sound power is then used to calculate the resulting sound pressure level in the space.
Sound is radiated from sheet metal ducts by vibration of the sheet metal. Round duct is
very stiff and breakout sound is small. The flat sheets of rectangular duct can more easily
vibrate, so breakout sound is greater. As a result, the in-duct attenuation is much higher
than for round duct. Rectangular duct with larger unsupported sheets can vibrate seriously
at low frequencies, setting up a resonance rumble with the fan.
The sound from the mechanical room travels along the return path and is often controlled
by a sound trap at the entry to the mechanical room.
Locating the mechanical room away from sound-sensitive areas and in more massive structural areas can greatly assist in reducing sound and vibration challenges from the system.
The mechanical room should be designed to contain the sound and vibration, which is
both a design and installation issue.
Vibration, usually low frequency, is transmitted from the equipment to the building fabric. Isolation is achieved by using specifically designed anti-vibration mountings and fittings, which reduce this transfer to acceptable levels. Equipment mounts are best chosen
with the manufacturer’s assistance.
The vibration mount must deflect substantially more than the structure with the plant
load, so light structures such as steel truss roofs can be a challenge and may require equip-
10–28
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ment relocation or structural stiffening. Dealing with light structures and other sound and
vibration mitigation are usually less than 1% of the mechanical system cost if included in
the design. Remedial work can be very difficult and expensive.
Bibliography
Blazier, Charles and Warren, ed. 1998. Application of Manufacturers Sound Data. Atlanta, GA:
ASHRAE.
Schaffer, M.E. 1991, 2005. A Practical Guide to Noise and Vibration Control in HVAC Systems.
Atlanta, GA: ASHRAE.
AMCA. 2006. AMCA Standard 301-2006, Methods for Calculating Fan Sound Ratings from
Laboratory Data.
ASHRAE Handbook–Fundamentals: sound and vibration fundamental data; ASHRAE Handbook–
Applications: practical information on system design for sound and vibration control, and
testing, adjusting and balancing information on equipment to use and how to test for
sound and vibration performance in buildings and outside.
10–29
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Skill Development Exercises for Chapter 10
Complete these questions by writing your answers on the worksheets at the back of this
book.
10-1.
Fundamentals of sound important to the HVAC designer include:
a) Sound pressure levels in occupied spaces
b) Sound power levels produced by equipment
c) Sound intensity in supply and return sound paths
d) All of the above e) a and b above
10-2.
The audible frequency range extends from about:
a) 20 kHz to 20 MHz b) 2 Hz to 20 Hz
c) 20 Hz to 20 kHz d) None of the above
10-3.
Typically, the cost of preventing sound and vibration problems in an HVAC
system is approximately __________ of the total system cost.
a) 3% b) 2% c) 1% d) None of the above
10-4.
A mechanical room housing a fan or AHU with an unducted intake should have
a floor area of __________ for each 1,000 cfm of air flow.
a) 2 to 4 ft2 b) 5 to 7.5 ft2 c) 7 to 10 ft2 d) None of the above
10-5.
All HVAC equipment rooms should have a floor area large enough to allow a
clearance of at least __________ feet around all equipment.
a) 1.0 b) 1.5 c) 2.0 d) None of the above
10–30
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10-6.
Sound-absorbing material can be arranged in a duct system by:
a) Lining fan suction and discharge plenums
b) Lining ducts with sound-absorbing material
c) Lining duct sections close to elbows
d) All of the above
10-7.
Duct silencers should be located _________ duct diameters for every 1,000 fpm
from fan discharges.
a) 0.1 b) 0.5 c) 1.0 d) None of the above
10-8.
Resonant silencers are often used in medical facilities where biological decontamination may be required.
a) False b) True
10-9.
Choosing a silencer is dependent on:
a) air resistance b) regenerated noise c) space availability
d) attenuation e) all of the above
10-10.
Advantages of round duct are its inherent resistance to vibration and sound
break-out.
a) True b) False
10–31
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Chapter 11
Air System Startup and
Diagnostics
Contents of Chapter 11
•
•
•
•
•
•
•
•
11.1 Introduction
11.2 Design Considerations
11.3 Air Volumetric Measurement Methods
11.4 Balancing Procedures for Air Distribution Systems
11.5 Noise and Vibration Diagnostics
Summary
Bibliography
Skill Development Exercises for Chapter 11
11–1
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Study Objectives of Chapter 11
After completing this chapter, you should be able to list and explain the major steps
involved in starting up an air system, and diagnose common problems associated with
starting up an air system.
11.1 Introduction
Testing, adjusting and balancing (TAB) are required to change an installed air system into
a properly working air system. Historically, TAB was done on each air handling system
with little reference to the associated cooling, heating and control systems. The TAB of
each system is now often part of the commissioning process. To quote from the ASHRAE
HandbookApplications:
“COMMISSIONING is a quality assurance process of the installation of the
systems in a building. It is a process for achieving, verifying, and documenting
the performance of each system to meet the operational needs of the building
within the capabilities of the documented design and specified equipment
capacities, according to the owner’s functional criteria. It is a process that
ensures the quality of the installation. Successful commissioning includes the
preparation of manuals and training of operation and maintenance personnel.
The result of commissioning should be fully functional systems that can be
properly operated and maintained throughout the useful life of the building.
All efforts related to commissioning should be specified in the contract documents.”
Under commissioning, TAB is extended to making sure the air handling system itself is
working effectively, and also that it is working effectively with the other parts of the building’s HVAC systems. Because the noise generated by the air system may be important or
critical, a section on checking sound power levels and some suggestions on deciding the
sources of noise problems are also included.
The use and environment in a building are dynamic, change with time, and must be rebalanced accordingly. The designer must consider initial and supplementary testing and balancing requirements during design and specification writing. Complete and accurate operating and maintenance instructions and manuals that include design intent and how to
test, adjust and balance the building systems are essential.
This chapter focuses on TAB, leaving the more comprehensive commissioning to other
courses. This is accomplished by: checking installations for conformity to design; measuring and establishing the system’s air quantities as required to meet design specifications;
and recording and reporting the results.
11–2
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The following terms and definitions are used in this chapter:
•
Test. To determine quantitative performance of equipment.
•
Balance. To proportion flows within the distribution system (submains,
branches and terminals) according to specified design quantities.
•
Adjust. To regulate the specified fluid flow rate and air patterns at the terminal
equipment (for example, reducing fan speed or resetting a damper position).
•
Procedure. An approach and execution of a sequence of work operations to
yield repeatable results.
•
Report forms. Test results summary and data sheets arranged for collecting test
data in logical order for submission and review. The data sheets should also
form the permanent record to be used as the basis for any future testing, adjusting and balancing.
•
Terminal. A point where the controlled medium (fluid or energy) enters or
leaves the distribution system. In air systems, these may be variable air or constant volume boxes, registers, grilles, diffusers, louvers and hoods.
11.2 Design Considerations
Testing, adjusting and balancing begin as design functions, with most of the devices
required for adjustments being integral parts of the design and installation. To ensure that
proper balance can be achieved, the engineer should show and specify a sufficient number
of dampers, valves, flow measuring locations and flow balancing devices. These must be
properly located in required straight lengths of pipe or duct for accurate measurement.
They must also be located to provide access.
The testing procedure depends on the system’s characteristics and layout. The interaction
between individual terminals varies with the system pressures, flow requirements and control devices.
11–3
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11.3 Air Volumetric Measurement Methods
The pitot-tube, or digital anemometer, traverse is a generally accepted method of measuring air flow in duct systems. Other methods of measuring air flow at individual terminals
are described by the various terminal manufacturers. The primary objective is to establish
measurement procedures that provide reliable, repeatable results.
In critical situations, laboratory tests, data and techniques prescribed by equipment and air
terminal manufacturers must be reviewed and corroborated for accuracy, applicability and
repeatability of the results. Conversion factors that correlate field data with laboratory
results must be developed to predict the equipment’s actual field performance.
AIR DEVICES
Generally, K-factors of air diffuser manufacturers should be checked for accuracy by field
measurement, comparing actual flow measured by pitot-tube traverse to actual measured
velocity. Air diffuser manufacturers usually base their volumetric test measurements on
readings obtained using a deflection vane anemometer. The velocity is multiplied by an
empirical effective area to obtain the air diffuser’s delivery. Accurate results are obtained by
measuring at the vena contracta with the probe of the deflection vane anemometer.
The methods advocated for measuring the air flow of troffer-type terminals are similar to
the methods described for air diffusers. The capture hood is frequently used to measure
device air flows, primarily of diffusers and slots. K-factors should be established for hood
measurements with varying flow rates and deflection settings. If the air does not fill the
measurement grid, the readings will require establishing a correction factor (similar to the
K-factor).
Rotating vane anemometers are commonly used to measure air flow into sidewall grilles.
Effective areas (K-factors) should be established with the face dampers fully open and
deflection set uniformly on all grilles. Correction factors1 are required when measuring air
flow in open ducts, such as damper openings and fume hoods.
DUCT FLOW
Most procedures for testing, adjusting and balancing air-handling systems rely on measuring volumes in the ducts rather than at the terminals.
The preferred method of duct volumetric flow measurement is the pitot-tube traverse average. Care should be taken to obtain the maximum straight run to the traverse station. Test
holes should be located as shown in the ASHRAE Handbook–Fundamentals and ASHRAE
Standard 1112 to obtain the best duct velocity profile. Where factory-fabricated volume
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measuring stations are used, the measurements should be checked against a pitot-tube traverse for field calibration.
The power input to a fan’s driver should be used only as a guide to indicate its delivery. It
may be used to verify performance determined by a reliable method (such as a pitot-tube
traverse of the system’s main), considering system effects that may be present. The flow
rate from some fans is not proportional to the power needed to drive them. In some cases
(as with forward-curved blade fans), the same power is required for two or more flow rates.
The backward-curved blade centrifugal fan is the only type with suitable characteristics;
flow rate that varies directly with the power input up to the point of maximum horsepower.
If an installation has an inadequate straight length of ductwork or no ductwork to allow a
pitot-tube traverse, multiple face velocities across the coil using the vane anemometer and
determining the K-factor may be read. The velocity readings must be taken exactly as prescribed by Sauer and Howell,1 using procedures from the air flow measurements at coil
faces.
MIXTURE PLENUMS
Approach conditions are often so unfavorable that the air quantities comprising a mixture
(such as outdoor air and return air) cannot be determined accurately by volumetric measurements. In such cases, the temperatures of the mixing airstreams and of the mixture can
be used to assess the proportions with reasonable accuracy.
Q mt m = Q ot o + Q r t r
(11-1)
where,
Qm = mixture air quantity, 100%
Qo = outside air quantity, as a % of mixture
Qr = return air quantity, as a % of mixture
tm = temperature of outside air and return air mixture, °F
to = outdoor air temperature, °F
tr = return air temperature, °F
This method generally provides acceptable results if the difference between to and tr is
greater than 20°F.
11–5
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PRESSURE MEASUREMENTS
The pressures involved with air measurements are barometric pressure, static pressure,
velocity pressure, total pressure and differential pressure. Pressure measurement for field
evaluation of air-handling system performance should be taken as recommended in
ASHRAE Standard 111 and analyzed together with the manufacturers’ fan curves and system effect as described in Chapter 6. When measured in the field, pressure readings, air
quantity and power input often do not agree with the manufacturers’ certified performance curves unless proper correction is made.
Pressure drops through equipment such as dampers or filters should not be used to measure air flow. Pressure is an acceptable means of establishing flow volumes only where it is
required by, and performed in accordance with, recommendations of the manufacturer
certifying the equipment.
STRATIFICATION
Normal design should minimize conditions causing air turbulence to produce the least
friction, resistance and consequent pressure losses in the system. However, under certain
conditions, air turbulence is desirable and necessary. For example, two airstreams of different temperatures can stratify in smooth, uninterrupted flow conditions. In this situation,
mixing should be promoted in the design.
The return and outside airstreams at the inlet side of the air-handling unit tend to stratify
where enlargement of the inlet plenum or casing size decreases the air velocity. Without a
deliberate effort to mix the two airstreams, stratification can exist and be carried throughout the system filters, coils, eliminators, fans and ducts. Stratification can cause damage by
freezing coils and rupturing tubes. It can also affect the temperature control in plenums,
spaces or both.
Stratification can be reduced by adding vanes to break up and mix the two airstreams. No
solution to stratification problems is guaranteed; each condition must be evaluated by field
measurements and experimentation. In extreme situations, static air mixers could be effectively used to assist mixing of airstreams at significantly different temperatures.
11–6
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11.4 Balancing Procedures for Air Distribution Systems
General procedures for testing and balancing are described here, although no one established procedure is applicable to all systems. Instrumentation for field testing and balancing of air systems and checking sound pressure levels are:
•
Manometers to measure the differential pressure across a pitot tube
•
Pitot tubes in various lengths, as required
•
Digital anemometer to measure air velocity and often also temperature
•
Tachometer to measure rotational speed; either direct contact, self-timing type,
or strobe light
•
Clamp-on ammeter with voltage scales (RMS type)
•
Deflecting vane anemometer to measure air velocity
•
Rotating vane anemometer to measure air velocity
•
Flow hood to capture the air from a diffuser and funnel it through an orifice for
flow measurement
•
Dial thermometers (2 in. diameter minimum and 1°F graduations minimum)
and glass stem thermometers (1°F graduations minimum)
•
Sling psychrometer to establish wet- and dry-bulb temperatures
•
Etched stem thermometer (30° to 120°F in 0.1°F increments)
•
Digital or analog hygrometers
•
Digital thermometers
•
Sound level meter, ideally with octave band filters
The instrumentation must be evaluated periodically to verify its accuracy and repeatability
prior to use in the field. Note that balancing is somewhat of an art. An experienced person
can do a very good job with limited tools, while a novice can become very frustrated when
damper changes produce unexpected results. For duct runs that contain no automatic flow
control devices such as VAV or dual duct boxes, the following is a simple, and logical, balancing methodology.
Start with all dampers open; except that the main duct dampers should be partially closed
if the fan is the forward-curve or radial-blade type to prevent motor overload. Now choose
the longest run of ductwork and start balancing at the farthest outlets from the fan. An
example is shown in Figure 11-1. Four outlets are shown with their initial air flows and
required air flows. Considering the last two outlets A and B, they both need the same flow
of 130 cfm. Balance B to have the same flow as A. The actual flow does not matter. Now
11–7
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consider outlet C, which requires 260 cfm, twice the flow of outlet A. Balance C to twice
the flow through A. Again, the actual flow does not matter; the ratio of C (adjusted) to A
(not adjusted). Next outlet D, again balance to twice outlet A.
Figure 11-1
Duct with 4 Outlets to be Balanced
Now outlets B, C and D are all in the correct flow ratio to outlet A. If the flow into the
branch is made correct, then the flows in A, B, C and D will be correct because their flows
are in the correct ratios. Do the same for all outlets on a branch. Where two branches
divide, adjust the branch flows to be in the desired ratio. The result is a system in the right
ratio. Finally, adjust the fan flow to the desired flow. To minimize energy waste, this will
be done by slowing the fan, not by closing dampers. All the outlets will have the correct
flow and there will be the minimum damper resistance in the system.
In practice, an experienced balancer can measure several outlets and decide which to throttle and by how much to get the desired flows. If you are starting balancing, the above
method can help avoid frustration.
PRELIMINARY PROCEDURE FOR AIR BALANCING
Before operating the system, the following steps should be performed:
11–8
•
Obtain as-built design drawings and specifications and become thoroughly
acquainted with the design intent.
•
Obtain copies of approved shop drawings of all air-handling equipment,
including performance curves, outlets (supply, return and exhaust), and temperature control diagrams.
•
Compare design to installed equipment and field installation.
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•
Walk the system from the air-handling equipment to terminal units to determine variations of installation from design.
•
Check dampers (both volume and fire) for correct and locked position, and
temperature control for completeness of installation before starting fans.
•
Prepare report test sheets for both fans and outlets. Obtain manufacturers’ outlet factors and recommended testing procedure. A summation of required outlet volumes permits a cross-checking with required fan volumes.
•
Determine best locations in main and branch ductwork for most accurate duct
traverses.
•
Place all outlet dampers in the open position. The main duct dampers should
be partially closed if the fan is the forward-curve or radial-blade type to prevent
motor overload.
•
Prepare schematic diagrams of system as-built ductwork and piping layouts to
facilitate reporting.
•
Check filters for cleanliness and proper installation (no air bypass). Checking
cleanliness is particularly important if the system has been running during construction. If the specifications require, establish a procedure to simulate dirty
filters.
•
For VAV systems, develop a plan to simulate diversity. Zones with different
geographic orientations experience peak loads at different times through the
day. For setting the system working, VAV terminals should not all be set at
maximum design volume, but at a lower value to simulate peak system load
which is less than the sum of all terminal peak volumes.
EQUIPMENT AND SYSTEM CHECK
•
Place all fans (supply, return and exhaust) in operation and immediately check
the following items:
Motor amperage and voltage to guard against overload
Fan rotation, speed and direction
Operability of static pressure limit switch
Automatic dampers for proper position.
Air and water resets operating to deliver required temperatures
Equipment anti-vibration mounts should be free with no metal-to-metal contact. On many smaller systems, the flexibility of the mounts can be checked by
pushing against the equipment and noting the movement. All pipes, ducts and
electrical conduits should have the required anti-vibration mounts and flexible
connections specified.
11–9
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Air leaks in the casing and around the coils and filter frames should be checked
by moving a bright light along the outside of the joints against the ducts while
observing the darkened interior of the casing. Any leaks should be sealed. Note
points where piping enters the casing to ensure that escutcheons are tight. Do
not rely on pipe insulation to seal these openings because the insulation may
shrink. In prefabricated units, check that all panel-fastening holes are filled to
prevent whistling.
•
Traverse the main supply ductwork whenever possible. All main branches
should also be traversed where duct arrangement permits. Selection of traverse
points and method of traverse should be as follows:
Traverse each main or branch after the longest possible straight run for the duct
involved.
For test hole spacing, refer to the ASHRAE Handbook–Fundamentals.
Traverse using a pitot tube and manometer where velocities are over 600 fpm.
Below this velocity, use a recently calibrated thermal anemometer.
Note temperature and barometric pressure to determine if they need to be corrected for standard air quantity. Corrections are normally insignificant below
2,000-ft elevation. However, where accurate results are desirable, corrections
would be justified.
Proportionally adjust branch dampers until each has the proper air volume.
Each branch and main duct should be within 10% of its design air flow rate to
be considered in balance.
After establishing the total air being delivered, adjust the fan speed to obtain
the design air flow, if necessary. Check power and speed to see that motor
power, and/or critical fan speed, have not been exceeded.
With the supply, return and exhaust fans operating at or near design speed and
delivering close to design air flow rates, set the minimum outdoor and return
air ratio. This can be done by measuring the mixture temperature with thermometers in the return air, outdoor air louver and filter section. As an approximation, the temperature of the mixture may be calculated from Equation
11-1.
The greater the temperature difference between hot and cold air, the more accurate the
results can be. Take the temperature at many points in a uniform traverse to be sure that no
stratification exists.
After the minimum outdoor air damper has been set for the proper percentage of outdoor
air, take another traverse of mixture temperatures and install baffling if the variation from
the average is more than 5°F. Remember that stratified mixed air temperatures vary greatly
11–10
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with the outdoor temperature in cold weather, while return air temperature has only a
minor effect.
•
Carefully set the system for balance using the following procedures:
Adjust the system with mixing dampers positioned for minimum outdoor air.
When adjusting multizone or double-duct constant volume systems, establish
the ratio of the design volume through the cooling coil to total fan volume to
achieve the desired diversity factor. Keep the proportion of cold air to total air
constant during the balance. However, check each zone or branch with this
component on full cooling. If the design calls for full flow through the cooling
coil, the entire system should be set to full flow through the cooling side while
making tests. Perform the same procedure for the hot air side.
•
Balance the terminal outlets in each control zone in proportion to each other.
The following steps may be followed to balance the terminals:
Once the preliminary fan quantity is set, proportion the terminal outlet balance from the outlets into the branches to the fan. Concentrate on proportioning the flow rather than the absolute quantity. As changes are made to the fan
settings and branch dampers, the outlet terminal quantities remain proportional. Branch dampers should be used for major adjusting; terminal dampers,
if used, are for trim or minor adjustment only. It may be necessary to install
additional sub-branch dampers to decrease the use of terminal dampers that
create objectionable noise.
Normally, several passes through the entire system are necessary to obtain
proper outlet values; unlike the method starting at the terminals where multiple passes are only required in some branches and one achieves minimum overall resistance.
The tested outlet air quantity may be an indicator of duct leakage.
With total design air established in the branches and at the outlets, perform the
following: take new fan motor amperage readings; find static pressure across
the fan; read and record static pressure across each component (intake, filters,
coils and mixing dampers); and take a final duct traverse.
Each outlet should be within 10% of the design air flow rate, to the nearest 25
cfm, to be considered in balance.
11–11
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DUAL-DUCT SYSTEMS
Most constant volume dual-duct systems are designed to handle a portion of the total system’s supply through the cold duct and smaller air quantities through the hot duct. Balancing should be accomplished as follows:
•
Check the leaving air temperature at the nearest terminal to verify that the hot
and cold damper inlet leakage is not greater than the maximum allowable leakage established.
•
Check apparatus and main trunks.
•
Determine if the static pressure at the end of the system (the longest duct run)
is at or above the minimum required for mixing box operation. Proceed to the
extreme end of the system and check the static pressure drop across the last
three boxes with a manometer. The drop across the box should exceed the minimum static pressure recommended by the manufacturer. Additional static
pressure is required for the low-pressure distribution system downstream of the
box.
•
Proportionately balance the diffusers or grilles on the low-pressure side of the
box, as described for low-pressure systems.
•
Change the control settings to full heating, and make certain that the controls
and dual-duct boxes function properly. Spot-check the air flow at several diffusers. Check for stratification.
•
If the engineer has included a diversity factor in selecting the main apparatus,
it will not be possible to get full flow from all boxes simultaneously.
VARIABLE AIR VOLUME SYSTEMS
The general procedure for balancing a VAV system is:
11–12
•
Determine the required maximum air volume to be delivered by the supply
and return air fans. Diversity of load usually means that the volume will be
somewhat less than the outlet total.
•
Obtain fan curves and request information on surge characteristics from the
fan manufacturer.
•
If an inlet vortex damper control is to be used, obtain the fan manufacturer’s
data pertaining to the de-aeration of the fan when used with the damper. If
speed control is used, find the maximum and minimum speeds that can be
used on the project.
•
Obtain the minimum and maximum operating pressures for terminal or variable volume boxes to be used on the project.
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•
Construct a theoretical system curve, including an approximate surge area. The
system curve starts at the minimum inlet static pressure of the boxes, plus system loss at minimum flow, and terminates at the design maximum flow. The
operating range using an inlet vane damper is between the surge line intersection with the system curve and the maximum design flow. When variable speed
control is used, the operating range is between: the minimum speed that can
produce the necessary minimum box static pressure at minimum flow still in
the fan’s stable range; and the maximum speed necessary to obtain maximum
design flow.
•
Position the terminal boxes to the proportion of maximum fan air volume to
total installed terminal maximum volume.
•
Set the fan to operate at approximate design speed (increase about 5% for a
fully open inlet vane damper).
•
Check a representative number of terminal boxes. If a wide variation in static
pressure is encountered, or if the air flow at a number of boxes is below minimum at maximum flow, check every box.
•
Run a total air traverse with a pitot tube.
•
Increase the speed if static pressure and/or volume are low. If the volume is correct, but the static pressure is high, reduce the speed. If the static pressure is
high or correct, but the volume is low, go over all terminals and adjust them to
the proper volume.
•
Run the previous four steps with the return or exhaust fan set at design and
measured by a pitot-tube traverse, and with the system set on minimum outdoor air.
•
Proportion the outlets, and verify the design volume with the VAV box on the
maximum flow setting. Verify the minimum flow setting.
•
Set the terminals to minimum, and adjust the inlet vane or speed controller
until minimum static pressure and air flow are obtained.
•
The temperature control personnel, the balancing personnel and the design
engineer should agree on the final placement of the sensor(s) for the static pressure controller. This sensor must be placed in a representative location in the
supply duct to sense static pressure(s) in the system.
•
Check the return air fan speed or its inlet vane damper that tracks or adjusts to
the supply fan air flow to ensure proper outside air volume.
•
On systems with economizers, operate the system on 100% outside air
(weather permitting), and check supply and return fans for proper power and
static pressure.
11–13
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INDUCTION SYSTEMS
Most induction systems use high-velocity air distribution. Balancing should be accomplished as follows:
•
Perform steps outlined under the basic procedures common to all systems for
apparatus and main trunk capacities.
•
Determine the primary air flow at each terminal unit by reading the unit plenum pressure with a manometer and locating the point on the charts (or
curves) of air quantity versus static pressure supplied by the unit manufacturer.
•
Normally, about three complete passes around the entire system are required
for proper adjustment. Make a final pass without adjustments to record the
final result.
•
To provide the quietest possible operation, adjust the fan to run at the slowest
speed that provides sufficient nozzle pressure to all units with minimum throttling of all unit and riser dampers.
•
After balancing each induction system with minimum outdoor air, reposition
to allow maximum outdoor air, and check power and static pressure readings.
REPORT INFORMATION
To be of value to the consulting engineer and the owner’s maintenance department, the
air-handling report should consist of at least the following items:
11–14
•
Design:
Total air quantity to be delivered
Fan static pressure
Motor power
Percent of outside air under minimum conditions
Fan speed
Power required to obtain this air quantity at design static pressure
•
Installation:
Equipment manufacturer (indicate model number and serial number)
Size of unit installed
Arrangement of the air-handling unit
Fan class
Nameplate power, nameplate voltage, phase, frequency and full-load amperes
of the motor installed
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•
Field tests:
Fan speed
Power readings (voltage, amperes of all phases at motor terminals)
Total pressure differential across unit components
Fan suction and fan discharge static pressure (equals fan total pressure)
Plot of actual readings on manufacturer’s fan performance curve to show the
installed fan operating point
Measured air flow rate
The initial static pressures must be accurately established for the air treatment equipment
and the duct system so that the variation in air quantity due to filter loading can be calculated. This ensures that the total fan air quantity will never be less than the minimum
requirements. This also serves as a check of dirt loading in coils, because the design air
quantity for peak loading of the filters has already been calculated.
•
Terminal Outlets:
Outlet by room designation and position
Outlet manufacture and type
Outlet size (using manufacturer’s designation to ensure proper factor)
Manufacturer’s outlet factor. Where no factors are available, or field tests indicate the listed factors are incorrect, a factor must be determined in the field by
traverse of a duct leading to a single outlet.
Design air quantity and the required velocity
Test velocities and resulting air quantity
Adjustment pattern for every air terminal
•
Additional Information (if applicable):
Air-handling units: belt number and size; drive and driven sheave size; belt
position on adjusted drive sheaves (bottom, middle and top); motor speed
under full load; motor heater size; filter type and static pressure at initial use,
anticipated full load, or time to replace; variations of velocity at various points
across the face of the coil; and existence of vortex or discharge dampers, or
both.
Distribution system: unusual duct arrangements; branch duct static readings in
double-duct and induction system; ceiling pressure readings where plenum
ceiling distribution is being used, and tightness of ceiling; relationship of
building to outdoor pressure under both minimum and maximum outdoor
air; and induction unit manufacturer and size (including required air quantity
11–15
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and plenum pressures for each unit) and a test plenum pressure and resulting
primary air delivery from the manufacturer’s listed curves.
All equipment nameplates visible and easily readable.
Many independent firms have developed detailed procedures and forms suitable to their
own operations and the area in which they function. These procedures are often available
for information and evaluation on request.
11.5 Noise and Vibration Diagnostics
A well designed and installed air system will, hopefully, not provide any problems. A frustrating feature of human behavior is that once a noise and vibration problem occurs, the
levels needed to satisfy occupants are lower than when no problem has occurred. Even
when no problems exist, it may be necessary to verify sound levels. To effectively verify
that the design requirements have been met, you need two pieces of information:
•
the required sound levels in terms of A weighting, NC or RC when the space is
fully furnished for anticipated operation
•
the measuring method or parameters
Ideally, these will have been clearly identified, and agreed with the client, before checking
is started. Note that the levels in a just-finished space before furnishing are likely to be
about 4 dB higher than the levels when the space is furnished. Having the measuring
method defined is important because the results can be very different depending on the
measurement locations. A specification that required all measurements to be taken at 4 feet
from the floor and 3 feet from any flat surface would provide measurements relevant to
seated occupants. It would also avoid any higher levels that may occur close to walls and in
corners due to sound reflections.
The next challenge is that the sound level meter will likely be accurate to ±3 dB even when
recently calibrated. If your client will check the measurements, be aware that their meter
will also have ±3 dB accuracy. Depending on your relations with the client, it may be prudent to suggest that the client’s meter be used and that a member of the client’s staff be
present so that there is no later argument about the readings. Remember that equipment
will only measure sound pressure levels. The quality of the sound cannot be measured; it is
an issue of human perception and humans vary.
Traffic noise and occupant activity can make it very difficult to obtain useful sound pressure readings. It may be prudent to carry out the tests at night or on a weekend to reduce
these challenges. Ideally, the test will be done after the system has been balanced as this
may significantly modify results.
11–16
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A basic procedure for doing a sound check is:
•
Obtain project system specifications and required sound pressure levels and
measurement methodology.
•
Visually check the system for compliance and any obvious likely problems.
•
Turn the system on and listen to it. Can you hear specific equipment, duct
leaks or duct vibrations?
•
Check that the sound measuring equipment is correctly calibrated.
•
In each measurement area, use an A-weighted meter to find the spot with the
highest reading and do detailed measurements in that location. Note the location so truly comparative measurements can be obtained. In situations where
background noise is significant, it may be necessary to assess the background
noise by taking measurements first with the equipment off and then with the
equipment running.
•
Note any specific noises even if the measurements indicate an acceptable sound
pressure level.
If there is a problem to be resolved, be systematic in your approach. Remember that every
sound and vibration has a specific source, a specific path from that source to the specific listener. If the overall sound level is excessive, check which frequencies are high. Lower frequencies indicate equipment generated noise and higher frequencies are regenerated air
noise. Can you identify the source just by carefully listening? As the ear is very sensitive to
changes in sound, have someone else turn off one piece of equipment at a time as you listen. If it is vibration, can you feel where it is most obvious in the space? You will likely be
able to hear/feel the change as the offending item is turned off.
Decide whether sound, vibration or both are the problem. If you sense that vibration is the
source of the problem, then that is most likely although it may be due to light-weight panels vibrating due to sound pressures. The ASHRAE HandbookApplications suggests the
following for determining if vibration is the problem:
•
If a sound level meter is available, check C-weighted and overall (unweighted,
or linear) readings. If the difference is greater than 6 dB, or if the slope of the
acoustic spectrum is greater than 5 to 6 dB per octave at low frequencies, vibration is likely a contributing factor.
•
If excessive noise is found close to the equipment and/or main ductwork, airborne noise is probably the contributor.
•
If the affected area is remote from the source equipment, no problem is apparent in intermediary spaces, and noise does not appear to be coming from the
duct system or diffusers, structure-borne noise is probably the cause.
11–17
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Sound leakage between adjacent spaces can be checked by listening in one space while
someone shouts in the adjacent space. If easily heard, look for air leakage paths, seal any
found and recheck. If none are found, check the wall performance.
If the offending equipment is the supply fan, identify where the offending sound enters the
space. While moving around the space, can you detect where the sound is entering? Is it
from the outlets, the ceiling, above the ceiling, walls, or floor? If it is from the outlets, is the
air flow correct? Can you take out an outlet and recheck to find out if it is outlet-generated
noise?
The Next Step
The final chapter takes you through the calculations for a sample air system design.
Summary
Testing, adjusting and balancing (TAB) are required to change an installed air handling
system into a properly working air system. Where TAB is part of commissioning, it must
be integrated into the total commissioning process.
TAB must be considered during design so that adequate dampers and access are available
to the contractor. The actual procedure depends on whether TAB is part of a comprehensive commissioning process and on the system type and components used.
The best volume flow readings are obtained from duct traverses in straight lengths of duct
using a recently calibrated digital anemometer or pitot-tube. The flow from most devices
can be measured using flow hoods.
Most procedures for testing, adjusting and balancing air-handling systems rely on measuring volumes in the ducts rather than at the terminals. The preferred method of duct volumetric flow measurement is the digital anemometer or pitot-tube traverse average. This
requires care to ensure the flow is reasonably even (straight section of duct) and that the
measurement locations are representative. Fan speed and pressure, or pressure across coils
or filters should only be used as a check, not to assess air flow. The inlet and outlet conditions of a fan (system effect) can markedly influence measurements and performance.
The temperatures of mixing air streams can be used to estimate relative flows by using
Equation 11-1. This method provides acceptable results if the temperature difference
between the airstreams is greater than 20°F, but it loses accuracy when the temperature difference is less than 20°F.
11–18
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Check that the outside air and return air mix without significant stratification. Mixing
vanes can be added to promote better mixing.
A variety of analog and digital equipment is available for measuring. The actual process
depends on the system. A logical progressive method from farthest outlet to fan was demonstrated. Effective air balancing depends on gathering the design information, preparing
the balancing procedure, and then following it while recording data as you go.
Having ensured that the supply fan cannot overload, start it and check that it and other
equipment are operating correctly. Then adjust balancing dampers and fan speed to
achieve correct air flows. With the correct flows, check that the outside air supply is correct
and adjust as necessary.
Check that the static pressure for the last three boxes is adequate, but not excessive. Check
that each box leakage is acceptable at full cooling and full heating. Adjust to correct flow
and balance outlets from the box.
VAV systems are more complex to TAB. The VAV boxes must normally be set to a fixed
value to set up the fan static pressure control. Once this is achieved, the performance of
each box must be set and the terminal outlet volumes balanced.
Induction systems operate at a high supply air pressure and the system must be balanced to
provide adequate pressure to each terminal. Due to noise generation, minimal damper use
is preferred.
Comprehensive records of all test data should be made as the tests are conducted. These
records are invaluable should any dispute arise and for future checking and retrofitting.
The data must be recorded in a predetermined format and indexed for later access.
Firms specializing in TAB have well-established procedures for doing the work and recording results.
Occupant expectations generally increase when noise and vibration problems arise, so a
well designed and installed air handling system can save a lot of aggravation.
Sound pressure levels and measurement methods should be agreed with the client before
checking begins. When checking, remember that equipment has limited accuracy and
record specifically where measurements were taken.
If problems exist, be very systematic in tracking down the source. The human ear is your
best tool for detecting changes. Having equipment turned on and off while you listen will
often enable you to identify the problem source. Methods for deciding whether the problem is noise, vibration or both were outlined.
11–19
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Fundamentals of Air System Design
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Bibliography
1. Sauer, H., Howell, R. 1990. Principles of Heating, Ventilating and Air-Conditioning. Atlanta,
GA: ASHRAE.
2. ASHRAE. 1988. ANSI/ASHRAE Standard 111, Practices for Measurement, Testing, Adjusting and
Balancing of Building Heating, Ventilation, Air-Conditioning and Refrigeration Systems.
Atlanta, GA: ASHRAE.
ASHRAE Guideline 0–2005, The Commissioning Process. Atlanta, GA: ASHRAE.
ASHRAE Guideline 1–1996, The HVAC Commissioning Process. Atlanta, GA: ASHRAE.
ASHRAE Handbook–Fundamentals: measurement and instruments; Handbook–Applications:
commissioning, TAB, sound and vibration control
11–20
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Skill Development Exercises for Chapter 11
Complete these questions by writing your answers on the worksheets at the back of this
book.
11-1.
The _______________ is the generally accepted method of measuring air flow
in duct systems.
a) Anemometer b) Psychrometer
c) Digital anemometer or pitot-tube traverse
d) All of the above
11-2.
Most procedures for establishing accurate air flows in air-handling systems rely
on measuring air volumes at the terminals.
a) True b) False
11-3.
Pressure drops through equipment such as dampers or filters may be used to
measure air flow.
a) True b) False
11-4.
_______________ is a necessary instrument for air balancing.
a) Pitot-tube or digital anemometer b) Flow hood
c) Digital or dial thermometers d) All of the above
11-5.
Air leaks in casings and around coils and filter frames should be checked by:
a) Moving a bright light along the outside of joints against the duct
b) Red smoke introduced into the airstream
c) Halogen sniffers d) All of the above
11-6.
__________ dampers should be used for major adjusting and __________
dampers for trim, or minor, adjustment only.
a) Branch, terminal b) Terminal, branch
c) Branch, subbranch d) All of the above
11-7.
The ___________ is frequently used to measure diffusers and slot air flows.
a) Anemometer b) Pitot-tube
c) Capture hood d) All of the above
11-8.
__________ are commonly used to measure air flow into sidewall grilles.
a) Rotating vane anemometers b) Pitot-tubes
c) Flow hoods d) All of the above
11–21
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11-9.
Air volume measurements taken at terminals are generally more reliable than
those obtained in the ducts.
a) True b) False
11-10.
Pressures involved with air measurements include:
a) Barometric pressure b) Static pressure
c) Velocity pressure d) All of the above
11-11.
Under certain conditions, air turbulence is desirable and necessary.
a) True b) False
11-12.
To avoid your check on sound pressure levels being challenged, it is best to have
agreed on _____________.
a) The allowable levels
b) The method for determining reading locations
c) An adjustment to the readings if the spaces are unfurnished
d) Whether the levels are to be assessed against an A weighting, NC, RC or
other scale
e) All of the above
11-13.
The most useful item for initial assessment of the cause of noise and vibration
problems is:
a) A complainer b) Your ears
c) A high quality sound meter with octave band filters
d) All of the above
11-14.
When recording sound pressure levels during a check to establish that levels are
within specification, it is valuable to note any noticeable tones even when the
levels are within specification because:
a) The tone indicates that the system causes significant vibration
b) The occupants may complain about the tone
c) The client’s sound level meter may be more sensitive to the tone than yours
d) All of the above
11–22
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Air System Startup and Diagnostics
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Chapter 12
An Actual Duct Design
Problem
Contents of Chapter 12
•
•
•
•
•
•
•
12.1 Introduction
12.2 Duct Design Procedure
12.3 The Building and System
12.4 Working Through the Problem
12.5 Conclusion
Bibliography
Skill Development Exercises for Chapter 12
12–1
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12.1 Introduction
This final chapter leads you through the duct design for one floor of an office building.
Duct design was the focus of Chapter 7, but information from other chapters will also be
used. You should complete the exercises as you go through the chapter, so you become
more comfortable with the choices and calculations. Compare your results to the results
given, and become familiar with the process. You may work in an office where all the duct
sizing is done with software, but you will still be able to produce better designs if you
understand the process and critically review the results the software produces.
This and other systems that you work on in the “real world” will undoubtedly require
return, outside and exhaust ducts that you would then design following these same principles.
Obviously, not every contingency nor every challenge that you will encounter in practice
has been covered. Rather, you have been exposed to the fundamentals of duct design and
provided a few examples of their application. Understanding these fundamentals and their
application will help you solve more advanced problems on your own.
12.2 Duct Design Procedure
Before beginning the duct design problem, review the proper steps leading to the design of
a good supply duct system:
12–2
•
From your load calculations, supply air temperature, space design temperature
and outside air ventilation requirements, calculate the maximum amount of air
to be supplied to each space. For variable volume systems, also calculate the
minimum volumes. Make preliminary adjustments for heat gain and loss and
for leakage.
•
Based on your knowledge of the building plans and intended occupancy, select
the size and type of air terminals and outlets, and then locate them on the plan.
•
Sketch a duct system that will connect the outlets and any VAV, or bypass, terminals to the supply fan discharge. Indicate the amount of air at each outlet,
terminal and duct segment.
•
Assign a number to each fitting and segment of straight duct so that they can
be readily identified in the tabulations of size and static pressure losses.
•
Size the ducts and calculate the static pressure that the supply fan will have to
overcome. To minimize resistance, use round ducts where feasible, then square
ducts and, finally, rectangular ducts that are as close to square as possible.
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•
Layout the system in detail, giving attention to height and width restrictions.
Be sure to allow space for joints, supports and any external insulation in addition to light fixtures, sprinklers and other building systems. Assess the sound
performance and add attenuation as required. Recalculate sizes and losses if
there are any changes.
•
Complete the design by specifying duct materials, pressure classification, seal
classification and any insulation or liner.
12.3 The Building and System
In the northern hemisphere, an intermediate floor of a 100 ft  200 ft multi-story office
building is to be conditioned (Figure 12-1). The space has been divided with full partitions
into individual perimeter offices and an open interior office. The 20 ft  50 ft core area of
toilets and mechanical equipment room (MER) shown in the middle is unconditioned.
The single-zone draw-through air handling unit (AHU) serves a VAV cooling-only system
for this floor only. The AHU will be installed in the MER. The design is to be based on a
supply air temperature of 54°F and maximum space temperature of 75°F. All duct is to be
constructed with galvanized steel.
The architect will lower the ceiling for a short distance near the MER to allow 24-in.-high
supply duct, up to the first take-off. The remainder of the duct is to be a maximum of 16
in. internal height which allows for duct joints, support and insulation.
The perimeter offices will have 4-ft slot diffusers handling a maximum of 120 cfm each.
Where the air flow to an individual office exceeds this amount, there will be two diffusers
per office. The interior will have 24 in.  24 in. perforated lay-in diffusers.
The ceiling space is a return air plenum and the MER is a mixed air plenum.
12–3
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12–4
Plan of Building
Figure 12-1
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An Actual Duct Design Problem
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SYSTEM PARAMETERS AND DESIGN ASSUMPTIONS
You previously determined from your load calculations that:
•
Each square foot of enclosed office has a load of 16 Btu/ft2
•
Each lineal foot of perimeter office has a peak wall heat load of:
North exposure: 83 Btu/ft run of wall
East exposure: 314 Btu/ft run of wall
South exposure: 222 Btu/ft run of wall
West exposure: 360 Btu/ft run of wall
For example, a 10 ft wide  12 ft deep north perimeter office will require cooling for the
floor area and the 10 feet of wall.
Design the duct system for peak air flow, allowing for no diversity. You may assume zero
leakage for calculating air flows.
The supply duct between the AHU and VAV terminals will be wrapped with insulation.
Between the VAV terminals and diffusers, the duct is to be lined. Use a roughness factor of
1.3 to compensate for the additional resistance of duct liner. In all cases, show the inside
dimensions of your recommended duct sizes.
Using the equal friction method, size the duct as follows:
•
Supply fan to VAV boxes: 0.20 in. SP loss per 100 ft
•
VAV terminals to diffusers: 0.10 in. SP loss per 100 ft
•
First segment of duct leaving the fan: 24 in. high
•
16 in. high thereafter, until
•
Duct can be reduced to 12 in.  12 in., and then
•
As near square as practical through the end of the run, including the duct
downstream of the VAV terminals
•
Use only integers for equivalent round duct sizes and only even rectangular
sizes
•
30° transitions
•
Do not add transitions beyond those shown
Summary information on VAV terminals, excerpted from the manufacturer’s catalog data,
is provided. Use the given static pressure loss listed for each terminal without further correction.
12–5
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The AHU manufacturer has provided the following data:
•
Mixing damper loss: 0.120 in. SP
•
Wet cooling coil loss: 0.450 in. SP
•
Dirty filters loss (filters and accumulated dirt): 1.0 in. SP
Allow 0.375 in. SP loss for the fan discharge plenum (before entering the straight duct section), which includes any system effect upon the fan. Allow for a 0.15 in. SP loss through
the ceiling return plenum. Add 0.006 in. SP loss for each diffuser connection, but do not
add any length for them in your calculations.
Air distribution SP losses:
•
Perimeter slot diffusers: 0.0625 in.
•
Interior 24 in.  24 in. diffusers: 0.075 in.
•
Return grilles: 0.050 in.
You do not need to show balancing dampers, fire dampers or any other air distribution
devices; nor do you need to calculate any losses through them for this problem.
This problem is “open book” and the data you will need have already been provided or will
be provided as you need it. Most of the process is dealt with in Chapter 7 and you will need
Equation 1-8:
V 2
p v =  ------------ 
 4005 
12–6
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An Actual Duct Design Problem
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12.4 Working Through The Problem
1. From your load calculations, supply air temperature, space design temperature and
outside air ventilation requirements, calculate the maximum amount of air to be supplied to each space. For variable volume systems, also calculate the minimum volumes. Make preliminary adjustments for heat gain and loss and for leakage.
The cooling loads for this floor have been calculated. In many buildings, these loads will be
the same for all floors except the top floor with added roof cooling load, the main floor
with entrance loads, and the basement with different floor and wall loads. The loads are
shown in Table 12-1 in two parts. The first is a Btu load per square foot. The second is the
additional Btu load for each foot run of perimeter wall. In this example, the exterior offices
will need cooling for their floor area plus the relevant orientation wall load(s) times feet of
wall.
The supply temperature is given as 54°F and the space design temperature as 75°F.
Now you can fill in the required cfm values to offset these cooling loads in Table 12-1.
Help is available in Chapter 1, Section 5. When you are done, turn the page and check
your answers in Table 12-2.
Table 12-1 CFM to Offset Btu/h Loads, Partial
Per square foot
North wall/foot run
East wall/foot run
South wall/foot run
West wall/foot run
Btu/h
16
83
314
222
360
CFM
12–7
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Did you remember the formula from Section 1.5 at the end of Chapter 1?
Btu/h = cfm  1.1  rise in temperature °F
Rearranging gives:
cfm = (Btu/h)/(1.1  rise in temperature °F)
= (Btu/h)/(1.1 (75-54))
= (Btu/h)/ 23.1
Table 12-2 CFM to Offset Btu/h Loads, Completed
Per square foot
North wall/foot run
East wall/foot run
South wall/foot run
West wall/foot run
Btu/h
16
83
314
222
360
CFM
0.7
3.6
13.6
9.6
15.6
We are going to use these values to calculate the individual office volumes. When we add
these calculated volumes, what will they total?
The North and South wall lengths are 17  10 + 2  15 = 200 ft
The East and West wall lengths are 7  10 + 2  15 = 100 ft
Total air-conditioned area = 200  100 – core area 20  50 = 19,000 ft2
System volume = Walls N + E + S + W + per square foot
= (200  3.6) + (100  13.6) + (200  9.6) +(100  15.6) + (19,000
 0.7) = 18,860 cfm
When we calculate the supply volumes to each zone, they should add up to about 18,860
cfm. If they do not, there is an error that must be found.
This process of using a global figure  the whole floor cfm  to check the sum of calculated
zone cfms is a very useful way of avoiding errors.
Now use these cfm values to calculate the required cfm for the typical spaces shown in
Table 12-3.
12–8
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Table 12-3
Calculate Sample Required Office Cooling Volumes
Wall Run
Wall Run
Floor Area
Total CFM
North 1012
East 1012
South 1012
West 1012
NE corner 1515
SW corner 1515
The cfm you calculated, as instructed, does not make allowance for heat gains or losses or
for air leakage. In this problem, the supply duct is in the ceiling plenum, so the surrounding temperature will be close to room temperature. The duct is stated as being insulated, so
heat gain will be small.
Leakage will reduce the actual volume supplied by the fan from reaching the conditioned
spaces. The leakage will cool the plenum, so it is not a load on the plant. This is in contrast
to a duct run through an unconditioned roof space where leakage is lost to the system. In
this system, leakage is accounted for by an increase in fan capacity and static. So it does not
need to show in the duct sizing calculations.
Did your calculations match the numbers shown in the first six lines of Table 12-4? Now
you have the required maximum supply air volumes for each space.
Table 12-4 Required Room Cooling Volumes (CFM)
Wall Run
North 1012
East 1012
South 1012
West 1012
NE corner 1515
SW corner 1515
SE corner 1515
NW corner 1515
103.6
1013.6
109.6
1015.6
153.6
159.6
159.6
153.6
Wall Run
Floor Area
Total CFM You Calculated
Interior 3250 ft2
10120.7
10120.7
10120.7
10120.7
15150.7
15150.7
15150.7
15150.7
32500.7
120
220
180
240
416
536
506
446
2275
Interior 2160 ft2
21600.7
1512
Interior 2385 ft2
23850.7
1670
1513.6
1515.6
1513.6
1515.6






12–9
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Fundamentals of Air System Design
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2. Based on your knowledge of the building plans and intended occupancy, select the
size and type of air terminals and outlets, and then locate them on the plan.
In this case, you were told that slot diffusers are being used in the perimeter offices and 24
in.  24 in. perforated face diffusers in the interior offices.
The exercise is to decide on a layout and size of diffusers for an open office area. The ADPI
information and representative manufacturers’ data are shown in Tables 12-5 and 12-6. In
Table 12-6, the throw is given as three numbers; for example, 3/5/10 which is the distance
from the diffuser (throw) before the jet velocity falls to 150 fpm/100 fpm/50 fpm. For situations where occupants are sitting in a fixed location, the 50 fpm throw (T50) is normally
chosen. For perforated, louvered ceiling diffusers, L is the distance from the diffuser to the
nearest wall.
When laying out diffusers, the ADPI gives guidance on the diffuser-to-wall distance. With
more than one diffuser, you must also consider the diffuser-to-diffuser distance. The air
from each diffuser will meet and drop as indicated in Figure 12-2. To achieve the same
downward velocity on the occupants, the diffusers should be approximately twice as far
apart as they are from the wall.
Table 12-5 ADPI Data for Perforated, Louvered Ceiling Diffusers
Terminal
Device
Perforated,
louvered ceiling
diffusers
12–10
Room Load
(Btu/h×ft2)
T50/L for
Maximum
ADPI
11  50
2.0
Maximum
For ADPI
ADPI
Greater Than
96
Range of
T50/L
90
1.4  2.7
80
1.0  3.4
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Table 12-6
Neck
Size
88
24  24 in. Perforated Face Diffuser Performance
Neck Velocity (fpm)
400
500
600
700
Flow Rate (cfm)
Throw (ft)
Static Pressure (in. wg)
NC
178
3/5/10
0.028
14
222
4/7/12
0.045
20
267
5/8/13
0.061
26
311
6/9/14
0.087
32
1010 Flow Rate (cfm)
Throw (ft)
Static Pressure (in. wg)
NC
278
3/6/12
0.031
16
347
5/7/14
0.05
24
417
6/9/15
0.068
30
486
7/10/16
0.096
35
1212 Flow Rate (cfm)
Throw (ft)
Static Pressure (in. wg)
NC
400
4/6/13
0.034
19
500
5/8/15
0.054
26
600
6/10/17
0.075
32
700
7/11/18
0.105
38
Figure 12-2
Diffuser Spacing
In this exercise, the space is 3,250 ft2, requiring 2,275 cfm. First, what are the space
dimensions? Knowing the office sizes, the space works out to 43 by 76 feet. Now, how
many diffusers, what neck size, and what volume through each would you choose? Write
down how you would tackle the challenge before turning to the next page.
12–11
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Here is one way of tackling the challenge. Based on the ADPI, the throw should be
between 2  L (ideal) down to 1.4  L.
12–12
•
One diffuser: Not suitable because the room is not near square, so the throw
will be excessive in the short dimension and inadequate in the long dimension.
•
Two diffusers: Throw across the narrow distance is 43/2 = 21.5 ft. Minimum
throw is 21.5  1.4 = 30 ft. None of the diffuser choices have a T50 throw
exceeding 18 ft. Therefore, two rows of diffusers are needed.
•
Four diffusers: L will be 43/4 = 11 ft (see Figure 12-1) across the short room
dimension, but 76/4 = 19 across the long span. Too much of a difference.
•
Six diffusers as 2 rows of 3: L will be 43/4 = 11 ft (see Figure 12-1) across the
short room dimension, but 76/6 = 13 across the long span, close to the short
dimension L of 11 ft. Ideal throw is the average L, 12  2 = 24, down to 12 
1.4 = 16.8. The flow through each diffuser is 2,275/6 = 379 cfm. You could
choose the 10  10 neck, although the throw is short, giving an ADPI of only
85%.
•
Fifteen diffusers as 3 rows of 5: L will be 43/6 = 7.2 ft across the short room
dimension and 76/10 = 7.6 across the long span, close to the short dimension
L of 7.2 ft. Ideal throw is the average L, 7.4  2 = 14.8, down to 7.4  1.4 =
10.4. The flow through each diffuser is 2,275/10 = 228 cfm. An 8  8 neck diffuser at 222 cfm has a throw of 12 ft in the range required. This is a better solution than six diffusers, providing an ADPI over 90%, at a higher construction
cost.
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3. Sketch a duct system that will connect the outlets and any VAV, or bypass, terminals to the supply fan discharge. Indicate the amount of air at each outlet, terminal
and duct segment.
There are two somewhat interrelated tasks here: zoning and laying out duct routes. Zoning
is choosing which spaces are to be grouped together and served by a VAV terminal. If
funding were not an issue, every space could have a VAV terminal and its own thermostat.
Typically, several spaces with the same use and load profile are grouped and controlled by
a thermostat in one of the spaces. How do you choose the group size? With experience! A
very general rule-of-thumb is that the more variable the load, the fewer spaces should form
a zone. Next is the issue of balancing the outlet volumes. A very long supply duct with
many outlets after the VAV terminal means that the outlets close to the terminal will have
a substantially higher pressure drop and damper noise may be an issue.
Table 12-7 has the manufacturer’s VAV terminal data and Figure 12-3 shows the floor layout with the main duct runs chosen and air flows to each space. How would you group the
spaces? Again, write down your choices before looking at the table below.
Table 12-7 Available VAV Terminal Units
Inlet
Diameter (in.)
6
8
10
12
14
Maximum
CFM
400
700
1000
1700
2400
Static Pressure
Loss (in. wg)
0.11
0.07
0.08
0.12
0.17
12–13
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12–14
Main Duct Layout and Air Volumes
Figure 12-3
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Table 12-8 and Figure 12-4 show one solution for this exercise. It is almost certainly different from yours; there is no ‘right’ answer. Look at the choices and note that they do obey
the suggestions of fewer spaces per zone the greater the range of load and the number of
outlets per zone is limited.
Table 12-8 Zone VAV Unit Choice
Zone
1
2
3
4
5
6
7
8
9
10
11
12
13
14
15
16
17
18
19
Office
Description
NE Corner
East Perimeter
East Perimeter
SE Corner
South Perimeter
South Perimeter
South Perimeter
South Perimeter
SW Corner
West Perimeter
West Perimeter
NW Perimeter
North Perimeter
North Perimeter
Interior
Interior
Interior
Interior
Interior
TOTAL
Zone Air
Flow (cfm)
416
880
660
506
720
720
900
720
536
720
960
446
1080
960
2275
1512
1512
1670
1670
18,863
VAV Inlet
Diameter
(in.)
VAV Max.
Air Flow
(cfm)
Max. as %
of Capacity
8
10
8
8
10
10
10
10
8
10
10
8
12
10
14
12
12
12
12
700
1000
700
700
1000
1000
1000
1000
700
1000
1000
700
1700
1000
2400
1700
1700
1700
1700
59%
88%
94%
72%
72%
72%
90%
72%
77%
72%
96%
64%
64%
96%
95%
89%
89%
98%
98%
Do you remember how we calculated the total supply air flow for the floor on page 12-8?
We used the total wall lengths and total floor area to calculate the flow as 18,860 cfm. Our
total for the zones is almost identical, so our zone calculations are likely correct.
12–15
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12–16
Sections A, B and C Ducts and Fittings Labeled
Figure 12-4
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4. Assign a number to each fitting and segment of straight duct so they can be readily
identified in the tabulations of size and static pressure losses.
If you are using software for duct sizing, check whether there are any specific rules for
numbering duct sections and fittings. Some require that you do the numbering from outlets back to a branch point and that you make sure that all outlets downstream of a branch
are numbered before numbering upstream of the branch.
The location of each zone VAV box (number in a square), duct section (A1, A2) and fitting (FB2, FB3) have been added to Figure 12-4 for three duct runs:
A – from the mechanical room to the south offices then eastwards to serve the south
east portion of the building
B – from the mechanical room to the north offices and then east to serve the north east
portion of the building
C – a major branch from duct run B
Do not be concerned with the details portion of the building west of the mechanical room.
5. Size the ducts and calculate the static pressure that the supply fan will have to overcome. To minimize resistance, use round ducts where feasible, then square ducts and,
finally, rectangular ducts that are as close to square as possible.
Having chosen the zones and VAV terminals, you can now size the ducts. Ignore the
length of fittings as their coefficients are based on having the duct loss accounted for in the
duct run. The procedure is:
•
Work out the flow in each section of duct
•
Look up the required round size, noting the velocity if using round ducts
•
Convert to rectangular, as necessary, to fit the space restrictions, allowing for
duct supports, sealing and insulation
The ducts are to be sized using a fixed static pressure drop of 0.2 in./100 ft for the main
runs of duct to the VAV terminals and then 0.1 in./100 ft for the runs from VAV terminal
to outlet.
Lengths for each segment for you to size are given in Table 12-9.
12–17
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Table 12-9
Sizing Branch C
Duct Size
Segment
Length
(ft)
B1
20
B2
10
B3
65
B4
20
C5
25
C6
25
Flow
(cfm)
Static
Pressure
(in./100 ft)
Diameter
(in.)
Width
(in.)
Height
(in.)
In this problem, we are going to use rectangular ductwork with the following guidelines. In
the “real world,” you may be able to use round or oval duct:
•
First segment of duct (A1 and B1) leaving the fan: 24 in. high
•
16 in. high thereafter, until duct can be reduced to 12 in.  12 in., and then
•
Use a height of 12 in. downstream of the VAV terminals
•
Use only integers for round duct sizes and only even rectangular sizes
•
30° transitions (reduction in duct size). Do not add transitions not shown
Work out the duct sizes from the fan to the end of branch C in Table 12-9 using the sizing
chart in Figure 12-5 and the Round-to-Rectangular chart shown in Table 12-10.
12–18
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Sizing Chart for Round Ducts
Figure 12-5
12–19
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Table 12-10 Round to Rectangular Duct Size Conversion
You will need the air velocity in the ducts to assess the pressure drop through the various
fittings. If you were going to use round ducts, you could note the velocity as you note the
size on the chart. When you convert from round to rectangular duct, the conversion table
will be on the basis of the same flow and pressure drop. As the rectangular duct is a less efficient shape, the velocity will be lower than if a round duct had been used. You can calculate the velocity by dividing the flow (in cfm) by the lengths of the sides (in feet). On a calculator, this is most simply done as: cfm  144 / in. side 1 / in. side 2 = fpm.
12–20
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Table 12-11 Chosen Duct Sizes
SP
Diameter
(in./100 ft)
(in.)
Duct Size
Width
(in.)
Height
(in.)
Segment
Length
(ft)
Flow
(cfm)
A1
A2
A3
A4
A5
A6
A7
A8
20
10
30
40
30
10
10
18
8664
5482
2606
1886
1166
660
660
440
0.2
0.2
0.2
0.2
0.2
0.2
0.1
0.1
27
23
18
16
13
10
12
9
28
28
16
14
12
12
12
6
24
16
16
16
12
12
12
12
B1
B2
B3
B4
B5
B6
B7
B8
20
10
65
20
15
10
20
18
10,199
7017
4531
3571
1296
880
880
440
0.2
0.2
0.2
0.2
0.2
0.2
0.1
0.1
29
25
21
20
14
12
13
10
32
32
24
20
12
12
12
8
24
16
16
16
16
12
12
12
C5
C6
25
25
2275
1140
0.1
0.1
19
15
24
15
12
12
The fitting coefficients shown in Figures 12-6 through 12-9 are for use without roughness
or other correction. They include all static and dynamic gains and losses.
12–21
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Figure 12-6
12–22
Fitting Coefficient SR4-1
Figure 12-7
Fitting Coefficient CR3-1
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Figure 12-8
SR5-14 Fitting Coefficient
Figure 12-9
SR5-13 Fitting
Coefficient
12–23
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Now calculate the pressure drops through the fittings in duct section C in Table 12-12.
With the fitting loss, you can then calculate the total static pressure the fan has to produce
to offset the pressure drops around the section C air system.
Table 12-12 Duct Run C- Dynamic Fitting Losses, Partial
Now compare your calculations with those shown in Table 12-13. Where differences exist,
see if you can work out why.
12–24
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Table 12-13 Duct Run C - Fitting Dynamic Losses, Completed
Once you have completed this comparison, look at Tables 12-14 and 12-15 which show
the comparable calculations for sections A and B. Did you notice that the total pressure
losses were highest for section C? The totals are A 2.723 in. wg, B 2.958 in. wg, and C 3.22
in. wg. The pressure loss in section C is 0.497 in. wg greater than section A. Thus, the fan
must produce the extra 0.497 in. wg for a flow to C of 2275 cfm, just 12% of the total
flow.
Producing extra pressure takes energy; let’s consider how much:
Air horsepower = (cfm  total pressure)/6358
kW = horsepower  0.746
Typical equipment efficiencies are: fan 65% to 70% (say 68%); drive 98% to 99% (say
99%); and motor 86% to 90+% (say 86%).
Actual kW, typically = (cfm  static pressure)/(6358  0.68  0.99  0.86)
= (cfm  static pressure)/3680
In this case, an increase in pressure drop of 0.497 in. requires an additional:
18,862  0.497/3680 = 2.55 kW
If the system runs for only 70 hours a week and electricity cost averages $0.1/kWh, the
annual cost of just 0.497 in. extra static pressure is 2.55  70  52  0.1 = $928. For continuous running, the annual cost rises to $2,228.
12–25
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Table 12-14 Duct Run A
Table 12-15 Duct Run B
12–26
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Look back at Tables 12-13 and
12-14. Where are the extra pressure losses in section C compared
to section A? They are in the fittings FB2 and FB4. In both
cases, the fitting has a high resistance. They can both be changed
to fittings with substantially
lower pressure drops. In the case
of fitting FB2, having it made up
of two CR3-1 elbows with guide
vanes (Figure 12-6 on page 1222) will substantially reduce the
resistance (Figure 12-10). Fitting
CR3-1, with r/W =1 and H/W =
16/32 = 0.5, the coefficient is
0.25. This reduces the pressure
loss from 0.243 to 0.243  0.25
= 0.061, a reduction of 0.18.
Figure 12-10
New Fitting FB2
In the same way, the fitting FB4
can be made as a duct that splits (SR5-1), so the coefficient drops to 0.25 (Figure 12-11).
The result of these two changes is to reduce section C to 2.812 in. wg and section B to
2.776. The changes have reduced the difference in losses to less than 0.1 in. wg and
reduced the calculated extra annual fan power cost from $928 to under $200.
Always review your duct designs to check whether a particular fitting has a significantly
high drop and also whether a particular section has a significantly higher pressure loss.
Always mark on your drawings where you require particular fittings so that the contractor
does not degrade your design intent by installing what is most convenient for them.
12–27
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12–28
SR5-1 Fitting Coefficient
Figure 12-11
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6. Layout the system in detail, giving attention to height and width restrictions. Be
sure to allow space for joints and any external insulation in addition to light fixtures,
sprinklers and other building systems. Assess the sound performance and add attenuation as required. Recalculate sizes and losses if there are any changes.
This is a very important step and one that is all too easy to do unsatisfactorily. At this stage,
you must make allowances for all the electrical, plumbing, fire protection, communication
and other services that also need space. Often, a coordination meeting to sort out routing
between the designers can save time, embarrassment and costly changes. On more complex
buildings such as hospitals, 3-dimensional drawings and models are frequently cost effective for construction and a great help to the subsequent maintenance of the facility.
When specifying holes through the structure, be careful to allow for a little misalignment,
and remember you need holes for the outside diameter of the ducts, not the nominal diameter. External insulation is typically installed after the duct is installed. Therefore, it
requires extra space, particularly if you require a continuous vapor barrier for a cold duct.
Recalculating sizes can be critical if significant extra resistances have been added by including sound attenuation. Note that an increase may be for the system overall and require a
simple increase in fan pressure, with the resulting increase in fan noise. Where sound attenuation has been added for just one or two zones, a significant increase in pressure drop may
occur for those branches. This can inadvertently create significant energy waste.
7. Complete the design by specifying duct materials, pressure classification, seal classification and any insulation or liner.
This final step of completing the specification is reasonably straightforward. On more
complex projects requiring special materials, fabrication or duct sealing (such as kitchen
exhausts, laboratory exhausts and duct extraction), be careful that the specification and
drawings clearly define the requirements for each run of ductwork.
With external insulation, be clear where a vapor barrier is required. For liner, be very clear
about both the method and spacing of fixing and any surface protection required. This is a
situation where copying the manufacturer’s suggested specification can be very effective,
particularly if there are any challenges with the installation. It is always easier to negotiate
with a contractor if you have the material manufacturer on your side.
Now that you have worked through this example, see what you have learned by doing the
Skill Development Exercises.
12–29
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12.5 Conclusion
This course has introduced you to air systems design. If you have worked through the
course examples and exercises, you should understand the process. This course does not
include all the detailed information you will need on fan performance, equipment resistances, duct fitting coefficients, sound attenuator performance, ventilation air requirements and costs, to name a few. For this information, you will need other texts and manufacturers’ information.
The ASHRAE Handbooks contain information on air systems, duct design, sound and
vibration control, as well as information for specific comfort and industrial systems. The
Sheet Metal and Air Conditioning Contractors National Association (SMACNA) produces several texts on duct system design, duct fabrication standards and installation
requirements.
For industrial system design, particularly for the exhaust of contaminants, the American
Conference of Governmental Industrial Hygienists (ACGIH) produces a detailed system
design guide. Then there are local Code requirements, which may incorporate the
National Fire Protection Association (NFPA) codes on design to deal with fire and smoke.
If you are doing your designs with a computer program, you must practice with it before
you attempt a project with an imminent deadline. Even the simplest software can require
inputs in a particular format or order. If you collect your input data in a format that
matches the input requirements, it will be much quicker to input, errors are less likely, and
finding errors will be quicker.
If you are new to your company and someone else has already been using the program, ask
them to run a demonstration for you, and ask which things they have found tricky with the
program. Finally, when you have your computer output, run a common-sense check on
the results: Are the sizes sensible for the airflows?
Finally, use information from the manufacturers. Their staffs have been trained in how
their products work, so ask them for the information you need. Always ask more than one
manufacturer, so you can make comparisons and increase your knowledge and understanding. If at first you do not receive the information you need, ask them again. Remember, manufacturers would much rather that you specified their products correctly for suitable applications, so they do not have to become involved in trying to resolve mistakes.
Good air system design is somewhat of an art, much helped by experience. We wish you
well in your designs.
12–30
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Bibliography
ASHRAE Handbooks (www.ashrae.org)
Handbook-Fundamentals: Duct design, sound and vibration, ventilation, energy estimating
Handbook-Systems and Equipment: Air-conditioning and heating systems, air-handling
equipment
Handbook-Applications: Comfort and industrial applications, sound and vibration control, fire and smoke management
SMACNA (www.smacna.org)
NFPA (www.nfpa.org)
ACGIH. 2008. Industrial Ventilation: A Manual of Recommended Practice for Design. 26th
ed. (www.acgih.org)
12–31
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Skill Development Exercises for Chapter 12
Complete these questions by writing your answers on the worksheets at the back of this
book.
12-1.
In the example you worked through in this chapter, you were told this is “an
intermediate floor” of the multi-story office building. Assuming that the layout
and use of each floor is the same and excluding the basement, would you expect
the loads to be the same on each floor?
a) Yes b) No c) Inadequate information to make an assumption
12-2.
Due to a serious design error, you are being pressured to dramatically reduce the
main duct sizes. You are asked how much the ducts would be reduced in size if
the system were designed to provide the supply air at 41F instead of 54F while
retaining the space design temperature of 75F. As an example, you suggest that
a 16  24 in. duct could be reduced to ______ although the insulation thickness may need to be increased.
a) 12  12 in. b) 16  12 in. c) 16  16 in. d) 16  20 in. e) 16  30 in.
12-3.
One open area in the example office floor will be used initially for general storage and printing, with a room load of 17 Btu/h ft2. In the future, it will be developed for special use, but no layout has been established and no ceiling is to be
installed at this stage. The ceiling is the flat concrete floor above and temporary
sidewall diffusers blowing across the space will be used. Knowing the tenant
organization, you believe part of the space will be used as temporary office space
before any fitting out is done. Therefore, you want to choose sidewall grilles that
will provide reasonable comfort. Using the table below, which range of T50/L
would you choose?
a) 0.8 to 2.2 b) 1.5 to 2.1 c) 0.9 to 1.8
Air Diffusion Performance Index (ADPI) Selection Guide
Room Load T50/L for
(Btu/h ft2) Max. ADPI
80
1.8
60
1.8
High Sidewall
Grilles
40
1.6
20
1.5
Terminal
Device
12–32
Maximum
For ADPI
ADPI
Greater Than
68
72
70
78
70
85
80
Range of
T50/L
1.5 - 2.2
1.2 - 2.3
1.0 - 1.9
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12-4.
Due to a misunderstanding, the electrician has run a conduit across the route for
your main supply duct. The sheet metal contractor can narrow the 8,000 cfm,
24  24 duct down to 24  18 to get past the conduit using a fitting. The fitting
coefficient is 0.3. You tell the owner that it can be done at slightly less cost than
having the electrician reroute the conduit. However, as the system will be running 24 hours a day all year and electricity costs are already $0.12 per kWh, the
ductwork change will add $___ to the annual electricity bill, as this problem
adds to the total resistance of the ductwork system.
a) $51 b) $96 c) $133 d) $171
12-5.
You are in a design office and notice a chart pinned on the wall (see below).
What pressure drop (in. wg/100 ft) was used to produce this chart? (Use Figure
12-5.)
a) 0.1/100 b) 0.08/100 c) 0.05/100 d) 0.2/100 e) 0.4/100
12-6.
Duct Size
(in.)
Max. Flow
(cfm)
6
8
10
12
14
120
250
440
700
1100
You are designing a dead-end length of supply duct to provide air at regular
intervals in a pedestrian tunnel between the office and somewhat dirty manufacturing buildings. The architect would like an unpainted galvanized steel spiral
steel duct with diffusers mounted from it. You warn the architect that:
a) The unpainted surface will attract dust.
b) The unsealed duct will likely produce dirt smudging at joint leaks making it
look bad.
c) The leakage of air will reduce the conditioning air in the tunnel.
12–33
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12–34
12-7.
The system is working well and a near-peak design day is experienced. You are
showing a visitor round the control room and they ask why the return temperature is up at 78F when the system is meant to be maintaining 75F. You explain
that:
a) Unfortunately, this was a lowest bid DDC system and the sensor accuracy is
not particularly good or reliable.
b) The air from the occupied spaces at 75F gets warmed by heat from the
lights between the occupied spaces and the return air sensor at the air-handling
unit inlet.
c) The energy from the supply fan increases the return temperature several
degrees.
12-8.
You are in a design office and notice a chart pinned on the wall (see below).
What limiting velocity was used to produce the chart? (Use Figure 12-5.)
a) 1000 fpm b) 1500 fpm c) 2000 fpm d) 2500 fpm e) 3000 fpm
Duct Size
(in.)
Max. Flow
(cfm)
28
30
32
36
11,000
12,000
15,000
18,000
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12-9.
An 18 in. high  36 in. wide supply duct carrying 4500 cfm runs along the ceiling before turning upwards to rise a floor and run back along the corridor above
(see below). What is the total pressure loss in the pair of smooth elbows assuming they are type CR3-1 smooth radius elbow fittings with r/W of 1.5?
a) 0.00875 in. wg b) 0.0125 in. wg c) 0.0175 in. wg d) 0.025 in. wg
12–35
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 1 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 1:
Question
1-1
1-2
1-3
1-4
1-5
1-6
1-7
1-8
1-9
1-10
1-11
Answer
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In the figure below, Area A1 = 2 ft2, Area A2 = 1.25 ft2, and velocity V1 = 1,000 fpm. Calculate V2 (fpm).
a) 1,600 fpm b) 625 fpm c) 1,406 fpm d) 2,569 fpm
1-2.
The total pressure at a certain point in a system is determined to be 5 in. wg, and the static
pressure at that point is determined to be 2 in. wg. What is the velocity pressure (in. wg) at
that point?
a) 21 in. wg b) 7 in. wg c) 3 in. wg d) 2 in. wg
1-3.
Which of the following is the most correct definition of static pressure regain?
a) As the velocity of an airstream decreases, the static pressure increases.
b) As the velocity of an enclosed airstream decreases due to friction, the static pressure
increases.
c) Friction reduces static pressure while velocity pressure increases with reduction in
duct size.
1-4.
An air handling system is determined to have a 6 in. pressure drop through the system at a
flow of 8,000 cfm. What is the system constant?
a) 1.5 b) 1,333 c) 3,265 d) 4,000
1-5.
The product of fluid pressure and specific volume is ______?
a) Internal energy b) Reynolds number c) Kinetic energy
d) Flow work e) Viscosity
1-6.
What does a water manometer measure?
a) Velocity b) Pressure c) Temperature d) All of the above
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1-1.
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
1-7.
Fan pressures are typically indicated in what units?
a) in. wg b) in. Hg c) cfm d) None of the above
1-8.
If the cross-sectional area of a duct decreases in size, the velocity of an airstream passing
through the duct will increase.
a) True b) Cannot tell c) False
1-9.
Air is passing through a length of inaccessible duct with a constant cross-sectional area. You
suspect that there is a serious leak in the duct. The velocity pressure drops from 0.85 in. wg
to 0.60 in. wg along the suspect section of duct. Approximately what percentage of air is
being lost through the leak?
a) 50% b) 31% c) 16% d) 11%
1-10.
In an air-conditioning system, 3000 cfm of outside air at 34°F is drawn in over a heater and
delivered into the building at 74°F. What volume of air is delivered?
a) 6,529 cfm b) 1,378 cfm c) 2,775 cfm d) 3,243 cfm
1-11.
In an air-conditioning system, 30,000 cfm of return air at 78°F is mixing with 4,600 cfm of
outside air at 95°F. What is the approximate resulting volume and temperature?
a) 36,400 cfm, 79.2°F b) 36,400 cfm, 80.3°F
c) 34,600 cfm, 80.3°F d) 34,600 cfm, 79.2°F
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 2 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 2:
Question
2-1
2-2
2-3
2-4
2-5
2-6
2-7
2-8
2-9
2-10
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This is the symbol for:
a) Centrifugal fan b) Axial fan c) Diffuser d) None of the above
2-2.
This ductwork is _____________, and the dimension of the side shown is _________.
a) Dropping, 20 b) Dropping, 12 c) Rising, 12 d) None of the above
2-3.
This is the symbol for a flexible duct:
a) True b) False
c) Cannot be determined from the information given.
2-4.
This symbol shows
a) A blanked-off duct, with a top dimension of 12
b) A return air duct, with a side dimension of 18
c) A supply air duct, with a side dimension of 18
d) None of the above
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
2-5.
The shown dimension of this duct is 24:
a) True b) False
c) Cannot be determined from the information given.
2-6.
A filter that uses a liquid as an adhesive is a:
a) Carbon filter b) Electrostatic filter c) Viscous filter
d) All of the above e) None of the above
2-7.
An air handling unit may be used to:
a) Move air b) Mix air c) Heat air d) All of the above e) None of the above
2-8.
This is the symbol for:
a) Manually operated damper b) Electrically controlled damper
c) Manual damper d) All of the above e) None of the above
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This is the symbol for:
a) Pneumatically operated damper b) Inline psychrometric observation device
c) Fire damper d) All of the above e) None of the above
2-10.
This is the symbol for:
a) Temperature relay b) Test station c) Remote bulb thermostat
d) All of the above e) None of the above
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 3 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 3:
Question
3-1
3-2
3-3
3-4
3-5
3-6
3-7
3-8
3-9
3-10
3-11
3-12
3-13
Answer
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The human body uses which of the following heat transfer mechanisms:
a) Radiation b) Convection c) Evaporation d) All of the above
3-2.
The perception of comfort relates to:
a) Individual physical condition
b) Body heat exchange with the surroundings
c) Physiological characteristics d) All of the above e) None of the above
3-3.
Which of the following would be within the acceptable range of temperature and humidity
for human comfort when wearing light summer clothing?
a) 72°F, 20% rh b) 70°F, 65% rh c) 80°F, 30% rh d) All of the above
3-4.
In a system with 8,000 annual wet-bulb degree hours above 66°F, with a 60% indoor relative humidity desired, and 56 hours of cooling system operation per week, the energy used
will be _______ Btu  106 per year per 1,000 cfm.
a) 51 b) 25 c) 36 d) None of the above
3-5.
The ____________________ Procedure for determining the required ventilation rate is
based on knowledge of the contaminants being generated within the space and the capability
of the ventilation air supply to limit them to acceptable levels.
a) Indoor Air Quality b) Ventilation Rate c) Contaminant Mitigation
d) All of the above e) None of the above
3-6.
Many designers have adopted a minimum total supply air flow of _____ for office applications.
a) 0.1 to 0.3 cfm/ft2 b) 0.6 to 0.8 cfm/ft2 c) 0.2 to 2.0 cfm/ft2
d) All of the above e) None of the above
3-7.
The airstream velocity at the end of the throw is called:
a) Terminal velocity b) Primary velocity c) Airstream velocity
d) All of the above e) None of the above
3-8.
_________________ air distribution systems create relatively uniform air conditions in the
occupied zone.
a) Unidirectional b) Local c) Mixing d) All of the above
e) None of the above
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
3-9.
The stagnant region of a Group B mixing outlet in a heating only system is ___________
the stagnant region of a Group A mixing outlet.
a) Larger than b) The same as c) Smaller than d) All of the above
3-10.
In displacement systems, the outlets are frequently located:
a) At or near the floor level b) In the walls c) In the ceiling
d) A and B e) None of the above
3-11.
Smudging is most likely to occur from dirt particles held in suspension in:
a) The room air b) The supply air c) The return air
d) All of the above e) None of the above
3-12.
The fan horsepower for under-floor supply systems can often be less than required for a ceiling supply mixing system due to which of the following?
a) Much cooler supply air
b) The low resistance to air flow in the plenum
c) The insulating value of the floor and carpet
3-13.
The under-floor supply systems work well for large open areas and the most effective control
is a thermostat in the return duct. True or false?
a) True b) False
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 4 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 4:
Question
4-1
4-2
4-3
4-4
4-5
4-6
4-7
4-8
Answer
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External offices with windows will have different thermal characteristics than windowless
rooms in the interior of the building:
a) True b) False
4-2.
In a building with a natatorium, the air pressure gradients within the building should
____________________:
a) Draw air from the natatorium into the rest of the building
b) Draw air into the natatorium from the rest of the building
c) Relieve the natatorium air intake
d) All of the above e) None of the above
4-3.
Which of the following is an advantage of an all-air system?
a) Additional duct clearance is not required
b) Air balancing in large systems is less difficult
c) Vertical shaft space is not required
d) All of the above e) None of the above
4-4.
Single-duct, single-zone systems can respond simultaneously to more than one set of space
conditions, in more than one area at a time:
a) True b) False
4-5.
In air-and-water systems, the air supply generally has a constant volume:
a) True b) False
4-6.
Evaporative coolers____________________:
a) Evaporate water into an airstream
b) Exchange sensible heat for latent heat
c) Can be either direct or indirect
d) All of the above e) None of the above
4-7.
An air economizer can achieve energy savings when _______:
a) The outdoor air enthalpy is lower than the supply air enthalpy
b) The outdoor air enthalpy is higher than the supply air enthalpy, but lower than the
return air enthalpy
c) Both of the above d) None of the above
4-8.
A minimum height of _________ above the roof surface is recommended for locating outside air louvers where light snowfall is expected:
a)1.0 ft b) 2.5 ft c) 4.0 ft d) All of the above e) None of the above
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 5 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 5:
Question
5-1
5-2
5-3
5-4
5-5
5-6
5-7
Answer
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Natural ventilation systems are most applicable when the building will produce a significant
stack effect:
a) True b) False
5-2.
Care must be taken in exhaust systems to minimize:
a) Corrosion b) Dissolution c) Melting d) All of the above
e) None of the above
5-3.
All other things being equal, ductwork is least subject to condensation corrosion when the
runs are:
a) Long and horizontal b) Short and vertical
c) Direct to the terminal discharge
d) All of the above e) None of the above
5-4.
Kitchen air pressure should be kept ______________ relative to other areas.
a) Positive b) Neutral c) Negative d) All of the above e) None of the above
5-5.
Smoke movement is driven by:
a) Stack effect b) Buoyancy c) Expansion d) All of the above
e) None of the above
5-6.
To prevent smoke infiltration on a fire floor, a pressurized stairwell must maintain a
_________________ pressure difference across a closed stairwell door.
a) Positive b) Neutral c) Negative d) All of the above e) None of the above
5-7.
Health facility ventilation requires:
a) Little need for accurate control of temperature and humidity
b) Free movement of air between departments
c) Removal of airborne microorganisms
d) All of the above e) None of the above
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 6 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 6:
Question
6-1
6-2
6-3
6-4
6-5
6-6
6-7
6-8
6-9
6-10
Answer
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A fan is delivering 6,000 cfm at a pressure of 1.5 in. wg at a rotational speed of 750 rpm. If
the fan speed is reduced to 600 rpm, how much air will the fan deliver, and at what pressure?
a) 4,800 cfm, 1.2 in. wg b) 4,800 cfm, 0.96 in. wg
c) 3,840 cfm, 0.96 in. wg d) 3,840 cfm, 1.2 in. wg.
e) None of the above
6-2.
Given a fan operating at 4,000 cfm, 3 in. wg total pressure, and 2.5 hp, what is the fan total
efficiency?
a) 85% b) 80% c) 75% d) All of the above e) None of the above
6-3.
Given a fan operating at 4,000 cfm, using 1.5 hp, what is the fan total efficiency?
a) 85% b) 75% c) 65% d) None of the above
6-4.
What is one effective duct length for a duct with a duct velocity of 4,000 fpm and an area of
200 in.2?
a) 80 ft b) 3.3 ft c) 5.34 ft d) None of the above
e) Cannot be determined from the information given
6-5.
What is one effective duct length for a duct with a duct velocity of 2,000 fpm and an area of
225 in.2?
a) 3.5 ft b) 3.0 ft c) 52.3 ft d) None of the above
e) Cannot be determined from the information given
6-6.
A rectangular duct is 10 in. high and 20 in. wide. What is the equivalent duct diameter of
this duct?
a) 200 in.2 b) 254 in. c) 16 in. d) None of the above
e) Cannot be determined from the information given
6-7.
For any given system, the system effect factor is constant across the range of flow volumes of
the fan:
a) True b) False c) Cannot be determined from the information given
6-8.
A fixed fan system is drawing 3 hp to deliver 10,000 cfm. If the air flow requirement can be
reduced to 7,000 cfm by decreasing the fan speed, the horsepower requirement will be
reduced to:
a) 2.1 hp b) 1.0 hp c) 0.44 hp d) All of the above e) None of the above
f) Cannot be determined from the information given
6-9.
The ___________________ is the highest efficiency centrifugal fan design.
a) Radial b) Forward-curved c) Backward-inclined, backward-curved
d) All of the above e) None of the above
6-10.
Power roof ventilators ___________ :
a) Usually operate without discharge ductwork b) Operate at low pressure
c) Operate at high volume d) All of the above e) None of the above
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 7 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 7:
Question
7-1
7-2
7-3
7-4
7-5
7-6
7-7
Answer
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As depicted in the figure below, this system consists of a fan, ductwork and outlets. The duct is
the same size as the fan outlet, so no system effect factor needs to be added. The outlets are
VAV boxes with a 1 in. wg pressure drop at 2,500 cfm. The fan speed will be adjusted to deliver
5,000 cfm. The Co value of the elbow is 0.2. What is the total pressure drop in the system?
a) 3.2 in. wg b) 1.8 in. wg c) 1.6 in. wg d) None of the above
7-2.
Air duct system design must consider:
a) Noise b) Duct leakage, heat gains and heat losses
c) Fire and smoke control d) All of the above e) None of the above
7-3.
Duct sizing and construction specifications are generally stated in terms of the use of:
a) Galvanized steel b) Aluminum c) Fiberglass reinforced plastic
d) All of the above e) None of the above
7-4.
Generally the most efficient and economical ducts for air systems are:
a) Rectangular b) Oval c) Round
d) All of the above e) None of the above
7-5.
Duct systems of rectangular fibrous glass are generally limited to:
a) 2,400 fpm and ±2 in. wg b) 4,000 fpm and ±3 in. wg
c) 1,000 fpm and ±3 in. wg d) None of the above
7-6.
Compression of flexible ducts significantly decreases air flow resistance.
a) True b) False
c) Cannot be determined from the information given
7-7.
Sealant systems have been developed that can substitute for mechanical joining of ductwork.
a) True b) False
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 8 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 8:
Question
8-1
8-2
8-3
8-4
8-5
8-6
8-7
8-8
8-9
8-10
Answer
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Combustibility and toxicity ratings are normally based on tests of:
a) New materials b) Old work c) Fibrous materials d) All of the above
8-2.
In the private sector, new construction is normally governed by:
a) State laws b) Local ordinances c) Codes d) All of the above
8-3.
Zone temperature controls are required for all systems, with special requirements for perimeter heating systems.
a) True b) False
8-4.
Which of the following standards applies to structures not exceeding 25,000 ft3 in volume?
a) NFPA 90A b) NFPA 90B c) NFPA 96 d) All of the above
8-5.
SMACNA HVAC Duct Construction Standards covers:
a) Basic duct construction b) Hangers and supports
c) Duct sealing classifications d) All of the above
8-6.
ASHRAE Standard 90.1 has a somewhat easier compliance route for many small air-conditioned buildings.
a) True b) False
8-7.
Compliance with ASHRAE Standard 90.1, Section 6, assures a minimum level of HVAC
system performance.
a) True b) False
8-8.
HVAC designers must know which code compliance obligations affect their designs.
a) True b) False
8-9.
HVAC systems are one of the most significant energy users in the types of buildings covered
by ASHRAE Standard 90.1.
a) True b) False
8-10.
A very efficient HVAC system could have an overall efficiency greater than one.
a) True b) False
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 9 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 9:
Question
9-1
9-2
9-3
9-4
9-5
9-6
9-7
9-8
9-9
9-10
9-11
9-12
Answer
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Worksheets for All Chapters.fm Page 24 Wednesday, October 16, 2019 3:20 PM
In a smoke control system:
a) A smoke damper inhibits the passage of air that may or may not contain smoke
b) Moderate leakage of smoke-free air through the damper does not adversely affect the
control of smoke movement
c) Design the system so that only smoke-free air is on the high-pressure side of a smoke
damper, unless the smoke control damper is on the return air
d) All of the above e) None of the above
9-2.
Particles less than _____ in diameter are referred to as the fine mode.
a) 0.75 μm b) 7.5 μm c) 75 μm d) None of the above
9-3.
From an industrial hygiene perspective, particles with an aerodynamic particle size of _____
or greater are considered the nonrespirable fraction of dust.
a) 5 μm b) 10 μm c) 15 μm d) None of the above
9-4.
____________ measures the ability of the filter to remove particulate matter from an airstream.
a) Efficiency b) Air flow resistance c) Dust-holding capacity
d) All of the above e) None of the above
9-5.
Different types of filters are distinguished by:
a) Efficiency b) Air flow resistance c) Dust-holding capacity
d) All of the above e) None of the above
9-6.
Filters collect particles by:
a) Straining b) Inertial deposition c) Electrostatic effects
d) All of the above e) None of the above
9-7.
In panel filters, the accumulating dust load causes pressure drop to:
a) Decrease to the filtration load rating, then increase
b) Increase to the filtration load rating, then decrease
c) No effect, remains constant
d) None of the above
9-8.
Electronic filters, which if maintained properly by regular cleaning, have relatively constant
pressure drop and efficiency.
a) True b) False
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
9-9.
Important requirements of a satisfactory and efficiently operating air filter installation
include:
a) Ample capacity for the amount of air and dust load it is expected to handle
b) Suited to the operating conditions
c) Economical for the specific application
d) All of the above e) None of the above
9-10.
Duct heaters may be:
a) Steam b) Water c) Electric d) All of the above
9-11.
The performance of particulate filters is categorized in Standard 52.2 into 20 MERV ratings,
with MERV 1 being a coarse screen and MERV 20 being the high rating filter for demanding cleanroom situations. To control a buildup of dirt on wet cooling coils, Standard 62
requires what MERV rating filter be installed before cooling coils that can run wet?
a) MERV 2 b) MERV 6 c) MERV 10 d) MERV 14
9-12.
The parallel blade damper deflects the air in one direction as the air passes through. This
usually makes the performance of a parallel blade damper more sensitive to location than an
opposed blade damper in the same location.
a) True b) False
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 10 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 10:
Question
10-1
10-2
10-3
10-4
10-5
10-6
10-7
10-8
10-9
10-10
Answer
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Fundamentals of sound important to the HVAC designer include:
a) Sound pressure levels in occupied spaces
b) Sound power levels produced by equipment
c) Sound intensity in supply and return sound paths
d) All of the above e) a and b above
10-2.
The audible frequency range extends from about:
a) 20 kHz to 20 MHz b) 2 Hz to 20 Hz
c) 20 Hz to 20 kHz d) None of the above
10-3.
Typically, the cost of preventing sound and vibration problems in an HVAC system is
approximately __________ of the total system cost.
a) 3% b) 2% c) 1% d) None of the above
10-4.
A mechanical room housing a fan or AHU with an unducted intake should have a floor area
of __________ for each 1,000 cfm of air flow.
a) 2 to 4 ft2 b) 5 to 7.5 ft2 c) 7 to 10 ft2 d) None of the above
10-5.
All HVAC equipment rooms should have a floor area large enough to allow a clearance of at
least __________ feet around all equipment.
a) 1.0 b) 1.5 c) 2.0 d) None of the above
10-6.
Sound-absorbing material can be arranged in a duct system by:
a) Lining fan suction and discharge plenums
b) Lining ducts with sound-absorbing material
c) Lining duct sections close to elbows
d) All of the above
10-7.
Duct silencers should be located _________ duct diameters for every 1,000 fpm from fan
discharges.
a) 0.1 b) 0.5 c) 1.0 d) None of the above
10-8.
Resonant silencers are often used in medical facilities where biological decontamination may
be required.
a) False b) True
10-9.
Choosing a silencer is dependent on:
a) air resistance b) regenerated noise c) space availability
d) attenuation e) all of the above
10-10.
Advantages of round duct are its inherent resistance to vibration and sound break-out.
a) True b) False
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10-1.
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 11 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 11:
Question
11-1
11-2
11-3
11-4
11-5
11-6
11-7
11-8
11-9
11-10
11-11
11-12
11-13
11-14
Answer
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The _______________ is the generally accepted method of measuring air flow in duct systems.
a) Anemometer b) Psychrometer
c) Digital anemometer or pitot-tube traverse
d) All of the above
11-2.
Most procedures for establishing accurate air flows in air-handling systems rely on measuring air volumes at the terminals.
a) True b) False
11-3.
Pressure drops through equipment such as dampers or filters may be used to measure air
flow.
a) True b) False
11-4.
_______________ is a necessary instrument for air balancing.
a) Pitot-tube or digital anemometer b) Flow hood
c) Digital or dial thermometers d) All of the above
11-5.
Air leaks in casings and around coils and filter frames should be checked by:
a) Moving a bright light along the outside of joints against the duct
b) Red smoke introduced into the airstream
c) Halogen sniffers d) All of the above
11-6.
__________ dampers should be used for major adjusting and __________ dampers for
trim, or minor, adjustment only.
a) Branch, terminal b) Terminal, branch
c) Branch, subbranch d) All of the above
11-7.
The ___________ is frequently used to measure diffusers and slot air flows.
a) Anemometer b) Pitot-tube
c) Capture hood d) All of the above
11-8.
__________ are commonly used to measure air flow into sidewall grilles.
a) Rotating vane anemometers b) Pitot-tubes
c) Flow hoods d) All of the above
11-9.
Air volume measurements taken at terminals are generally more reliable than those obtained
in the ducts.
a) True b) False
11-10.
Pressures involved with air measurements include:
a) Barometric pressure b) Static pressure
c) Velocity pressure d) All of the above
11-11.
Under certain conditions, air turbulence is desirable and necessary.
a) True b) False
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11-1.
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
11-12.
To avoid your check on sound pressure levels being challenged, it is best to have agreed on
_____________.
a) The allowable levels
b) The method for determining reading locations
c) An adjustment to the readings if the spaces are unfurnished
d) Whether the levels are to be assessed against an A weighting, NC, RC or other scale
e) All of the above
11-13.
The most useful item for initial assessment of the cause of noise and vibration problems is:
a) A complainer b) Your ears
c) A high quality sound meter with octave band filters
d) All of the above
11-14.
When recording sound pressure levels during a check to establish that levels are within specification, it is valuable to note any noticeable tones even when the levels are within specification because:
a) the tone indicates that the system causes significant vibration
b) the occupants may complain about the tone
c) the client’s sound level meter may be more sensitive to the tone than yours
d) all of the above
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Answer and Work Sheets for Chapter 12 Exercises
Please complete by selecting one answer for each question. For your records, you may wish to
show your work on the following pages.
To receive full continuing education credit, all questions must be answered and submitted to
edu@ashrae.org (preferred method) or faxed to 678-539-2161. Please provide your student ID
number and the SDL number.
Your student ID number is the last 5 digits of your Social Security number or other unique five
digit number you create. The SDL number is the last 5 digits of the ISBN number (found under
the copyright information of this book).
The correct answers will be provided upon completion of each chapter’s exercises.
Insert Answers for Chapter 12:
Question
12-1
12-2
12-3
12-4
12-5
12-6
12-7
12-8
12-9
Answer
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In the example you worked through in this chapter, you were told this is “an intermediate
floor” of the multi-story office building. Assuming that the layout and use of each floor is the
same and excluding the basement, would you expect the loads to be the same on each floor?
a) Yes b) No c) Inadequate information to make an assumption
12-2.
Due to a serious design error, you are being pressured to dramatically reduce the main duct
sizes. You are asked how much the ducts would be reduced in size if the system were
designed to provide the supply air at 41F instead of 54F while retaining the space design
temperature of 75F. As an example, you suggest that a 16  24 in. duct would be reduced
to ______ although the insulation thickness may need to be increased.
a) 12  12 in. b) 16  12 in. c) 16  16 in. d) 16  20 in. e) 16  30 in.
12-3.
One open area in the problem office floor will be used initially for general storage and printing, with a room load of 17 Btu/h ft2. In the future, it will be developed for special use, but
no layout has been established and no ceiling is to be installed at this stage. The ceiling is the
flat concrete floor above and temporary sidewall diffusers blowing across the space will be
used. Knowing the tenant organization, you believe part of the space will be used as temporary office space before any fitting out is done. Therefore, you want to choose sidewall grilles
that will provide reasonable comfort. Using the table below, which range of T50/L would
you choose?
a) 0.8 to 2.2 b) 1.5 to 2.1 c) 0.9 to 1.8
Air Diffusion Performance Index (ADPI) Selection Guide
Room Load T50/L for
(Btu/h ft2) Max. ADPI
80
1.8
60
1.8
High Sidewall
Grilles
40
1.6
20
1.5
Terminal
Device
Maximum
For ADPI
ADPI
Greater Than
68
72
70
78
70
85
80
Range of
T50/L
1.5 - 2.2
1.2 - 2.3
1.0 - 1.9
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12-1.
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
12-4.
Due to a misunderstanding, the electrician has run a conduit across the route for your main
supply duct. The sheet metal contractor can narrow the 8,000 cfm, 24  24 duct down to 24
 18 to get past using a fitting as shown below. The fitting coefficient is 0.3. You tell the
owner that it can be done at the slightly less cost than having the electrician reroute their
conduit. However, as the system will be running 24 hours a day all year and electricity costs
are already $0.12 per kilowatt hours the ductwork change will add $___ to the annual electricity bill as this problem adds to the total resistance of the ductwork system.
a) $51 b) $96 c) $133 d) $171
12-5.
You are in a design office and notice a chart pinned on the wall (see below). What pressure
drop (in in.wg/100 ft) was used to produce this chart? (Use Figure 12-5).
a) 0.1/100 b) 0.08/100 c) 0.05/100 d) 0.2/100 e) 0.4/100
12-6.
Duct Size
(in.)
Max. Flow
(cfm)
6
8
10
12
14
120
250
440
700
1100
You are designing a dead-end length of supply duct to provide air at regular intervals in a
pedestrian tunnel between the office and somewhat dirty manufacturing buildings. The
architect would like an unpainted galvanized steel spiral steel duct with diffusers mounted
from it. You warn the architect that:
a) The unpainted surface will attract dust
b) The unsealed duct will likely produce dirt smudging at joint leaks making it look bad
c) The leakage of air will reduce the conditioning air in the tunnel
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The system is working well and a near-peak design day is experienced. You are showing a visitor round the control room and they ask why the return temperature is up at 78F when the
system is meant to be maintaining 75F. You explain that:
a) Unfortunately, this was a lowest bid DDC system and the sensor accuracy is not
particularly good or reliable.
b) The air from the occupied spaces at 75F gets warmed by heat from the lights between
the occupied spaces and the return air sensor at the air-handling unit inlet.
c) The energy from the supply fan increases the return temperature several degrees
12-8.
You are in a design office and notice a chart pinned on the wall (see below). What limiting
velocity was used to produce the chart? (Use Figure 12-5).
a) 1000 fpm b) 1500 fpm c) 2000 fpm d) 2500 fpm e) 3000 fpm
Duct Size
(in.)
Max. Flow
(cfm)
28
30
32
36
11,000
12,000
15,000
18,000
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12-7.
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Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
Work Sheet for Fundamentals of Air System Design (I-P), Second Edition
12-9.
An 18 in. high  36 in. wide supply duct carrying 4500 cfm runs along the ceiling before
turning upwards to rise a floor and run back along the corridor above (see below). What is
the total pressure loss in the pair of smooth elbows assuming they are type CR3-1 smooth
radius elbow fittings with r/W of 1.5?
a) 0.00875 in. wg b) 0.0125 in. wg c) 0.0175 in. wg d) 0.025 in. wg
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EvaluationForm.fm Page 1 Monday, August 11, 2014 4:01 PM
Course Title: Fundamentals of Air System Design (I-P), Second Edition
On a scale of 1 to 5, circle the number that corresponds to your feeling about the statements
below. (1 = strongly agree, 5 = strongly disagree, 3 = undecided)
COURSE CONTENT
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ASHRAE LEARNING INSTITUTE
Self-Directed Learning Course Evaluation Form
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1791 Tullie Circle
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Fax: 404/321-5478
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Product Code: 98036 10/19
Errata noted in the list dated 07/18/19 have been corrected.
Air System Design I-P.indd 2
ISBN 978-1-933742-45-8
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