INSPECTION OF
INDUSTRIAL VENTILATION SYSTEMS
STUDENT MANUAL
Prepared for:
U.S. Environmental Protection Agency
Stationary Source Compliance Division
Office of Air Quality Planning and Standards
Washington, D.C. 20460
Prepared by:
Crowder Environmental Associates, Inc.
2905 Province Place
Piano, Texas 75075
September 1990
-------
DISCLAIMER
This manual was prepared by Crowder Environmental Associates,
Inc. and Entropy Environmentalists, Inc. for the Stationary Source
Compliance Division of the U.S. Environmental Protection Agency.
It has been completed in accordance with EPA Contract Number 68-02-
4462, Work Assignment No. 174. The contents of this report are
reproduced herein as received from the authors. The opinions,
findings, and conclusions expressed are those of the authors and
not necessarily those of the U.S. Environmental Protection Agency.
Any mention of product names does not constitute endorsement by the
U.S. Environmental Protection Agency.
i
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ACKNOWLEDGEMENTS
This manual is a revised version of a manual originally
prepared by Crowder Environmental Associates, inc. for the U.s
EPA, Stationary Source Compliance Division (SSCD). it wa*
originally prepared under a subcontract to PEI Associates, inc
Entropy Environmentalists, Inc. has converted the manual into
standardized format developed by the EPA Work Assignment Manager^
Mr. Kirk Foster. The majority of the drawings and photography
included in the original manual have been redrawn and modified
Ms. Sherry Peeler, Pendragon Inc. with the assistance of Entropy
Environmentalists, Inc. ^
• •
11
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Table of Contents
Topic Page
Lesson 1: General Principles of Ventilation
Properties of standard air 1-1
Molecular weight 1-1
Equation of state 1-2
Density and specific volume 1-3
Specific gravity 1-4
Relative and absolute humidity 1-4
Dry-bulb, wet-bulb and dew-point temperatures 1-6
Enthalpy 1-7
Psychrometric chart 1-10
Continuity 1-12
Momentum 1-13
References 1-17
Lesson 2: Hood Systems
Hood types 2-1
Hood design principles 2-7
Factors affecting hood performance 2-8
Capture velocity 2-8
Cold flow into hoods 2-9
Hot flow into hoods 2-12
Hood pressure losses 2-12
Evaluation of hood performance 2-14
References 2-16
Lesson 3: Duet Systems
Duct pressure loss 3-1
Velocity pressure calculation method 3-6
Estimating hood flowrate 3-7
Transport velocity 3-8
Balancing duct systems 3-11
References 3-13
Lesson 4: Gas Cooling Systems
Dilution with ambient air 4-1
Quenching with water 4-4
Natural convection and radiation 4-6
References 4-8
iii
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Table of Contents (Continued)
Lesson 5: Fan Systems
Types of fans 5_^
Fan arrangements 5_7
Fan laws 5„9
Fan performance 5-11
Fan selection 5-17
Evaluation of fan performance 5-20
References 5-22
Lessosn 6: Measurement of Ventilation System Parameters
Measurement ports 6>1
Static pressure measurement 6_2
Temperature measurement 6_5
Flowrate measurement §_8
Fan speed measurement 6-14
Horsepower measurement 6-15
Use of grounding cables 6-16
References 6-18
Lesson 7: Ventilation System Inspection
Level 2 inspections 7_2
Level 3 inspections 7_3
Use of Flowcharts 7_5
References 7-6
Appendix A: Course Slides
Appendix B: Bibliography
Appendix C: Psychrometric Charts
iv
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LESSON 1
GENERAL PRINCIPLES OF VENTILATION
Level 2 and 3 inspections of ventilation systems require
interpretation of instrument readings and may, at times, require
measurement and calculation of performance parameters. To be able
to conduct these inspections effectively, it is important that you
have firm understanding of the basic information that affects the
behavior of air streams. The purpose of this lesson, divided into
two sections, is to give you that information. Section I defines
the various parameters that are important in ventilation system
evaluation and indicates techniques and information sources that
may be used in their determination. Section II presents the
fundamentals of fluid flow and includes a discussion of the
implications of continuity and momentum relationships. Consider-
able emphasis is placed on Bernoulli's equation, and it is used to
develop relationships for the pressures that exist in a flowing
system and for determining the velocity of an air stream.
I. PROPERTIES OF AIR AND AIR-WATER VAPOR MIXTURES
Standard air
Standard air is defined as air with a density of
0.075 lbm/ft3 and an absolute viscosity of 1.225 x 10"5
lbm/ft-sec. This is equivalent to dry air at a
temperature of 70"F and a pressure of 29.92 in. Hg.
Molecular weight
Atmospheric air is a mixture of dry air, water
vapor and various impurities. Dry air itself is also a
mixture of gases. Because of this, neither atmospheric
air nor dry air have a true molecular weight. However,
they do have an apparent molecular weight that can be
calculated from their composition. Assuming dry air
consists, by volume, of 78.09% nitrogen, 20.95% oxygen,
0.93% argon and 0.03% C02, its apparent molecular weight
may be calculated as:
Slides
1-4
Component
n2
°2
Ar
CO,
Volume
Fraction
0.7809
0.2095
0.0093
Ot QOQ?
1.0000
Molecular
Weight
x 28.016
x 32.000
X 39.944
X 44.010
lb/lb-mole
21.878
6.704
0.371
0-013
1-5
28.966
Lesson 1
General Principles
-------
The apparent molecular weight of dry air with this ^
composition is then 28.966 lb/lb-mole.
Suppose the compositional information were
available on a weight rather than a volume basis, if
dry air consisted, by weight, of 75.52% nitrogen
23.15% oxygen, 1.28% argon and 0.04% co2, its apparent
molecular weight would be determined as follows:
Component
N2
°2
Ar
CO,
weight
Fraction
0.7552
0.2315
0.0128
0.0005
1.0000
Molecular
Weight
x 28.016
X 32.000
x 39.944
X 44.010
lb/lb-mole
0.02696
0.00723
0.00032
O.OOOpj
0.03452
The aDDarent molecular weight of dry air with this
composition* If then 1/0.03452 = 28.969 lb/lb-mole.
When composition information is not available, dry air
is typically taken to have an apparent molecular weight
of 28.95 lb/lb-mole and sometimes approximated as 29
lb/lb-mole.
For wet air, the apparent molecular weight may be
calculated from the composition as shovmabove or by
combining the molecular weights of the dry air and the
water vapor on the basis or their respective volume
fraction or mole fraction:
XMWdry.ir) + (Cwter) (^ter)
MW,
wet air
(1 - C,
water
Equation of state
Equations of state relate the pressure, volume and
temperature properties of a pure substance or mixture
by semi—theoretical or empirical relationships. over
the range of temperature and pressure usually
encountered in ventila*-lon systems, these values may be 1—6
related by the ideal or perfect gas law:
PV « nRT (Eqn. 2)
where P - absolute pressure (lbf/ft2)
V ¦ gas volume (ft3)
n * num*>er of moles (lb-moles)
r « constant (1544.58 ft-lbJlb-mole-°R)
T *¦ absolute temperature (°R)
Lesson 1
1-2
General Principles
-------
Here, R is referred to as the universal gas Slides
constant, and its value depends on the units of the
other terms in the equation. Other values of R
include:
10.73 psia-ft3/lb-mole-°R
0.73 atm-ft3/lb-mole-°R
82.06 cm3-atm/g-mole-°K
8.31 x 103 kPa-m3/kg-mole-°K
A more useful form of the ideal gas law may be
developed by noting that PV/T = nR, and that, for a
given number of moles of a gas, nR is a constant.
Thus, at two different conditions for the same gas, we
may write:
P1 V1 = P2 V2
T, T2
(Egn. 3)
1-7
or
= V,
(Egn. 4)
Equation 1-4 allows volumes (or volume rates) to
be corrected from one set of temperature and pressure
conditions to another.
Another useful form of the ideal gas law may be
used to calculate the molar volume, V/n. For an ideal
gas at 70 °F and 29.92 in. Hg (14.7 psia), the molar
volume is given by:
V/n = RT/P =
10.73 psia-ft3
lb-mole - °R .
(530 °R)
(14.7 psia)
= 387 ft3/lb-mole
Density and specific volume 1-8
Density is the ratio of mass to the volume
occupied, e.g., lb/ft3 or g/cm3. Specific volume is the
volume occupied per mass and is equal to the inverse of
Lesson 1
1-3
General Principles
-------
density. Both of these quantities depend on the Slidej
temperature and pressure of the system. Using the
ideal gas law, and recognizing that the number of moles
is given by mass divided by molecular weight, density
(p) may be calculated from:
m
p = (Eqn. 5)
V
Density can also be determined from molecular
weight and molar volume:
MW _. 530 °R
P = '
387
1[ If 1 (Eqn. 6)
Jl T J L 29. 92J
where MW = molecular weight
T = absolute temperature (°R)
P = absolute pressure (in. Hg)
Values for the density and other properties of air
over a limited range of temperature are provided in
Table 1-1.
Specific gravity 1-9
Specific gravity is the ratio of the density of a
material to the density of some reference substance.
For gases that reference substance is frequently dry
air, while for liquids and solids it is usually water.
Referring to Equation 1-5, it can be seen that for an
ideal gas specific gravity is also given by the ratio
of the molecular weight of the gas to the molecular
weight of dry air.
Relative and absolute humidity
The state of an air-water vapor mixture is
completely defined by specifying the pressure,
temperature and humidity. The Gibbs-Dalton rule of l-io
partial pressures states that individual components in
a mixture exert a pressure that would be the same as
that exerted if the same mass of the component were
present alone in the same total volume and at the same
temperature. Thus, for an air-water vapor mixture:
Pair * Pinter = ^total (Eqn. 7)
Lesson 1
1-4
General Principles
-------
Table 1-1. Properties of Air (Danielson, 1973)
Temp.
(°F)
Specific
Heat at
Constant
Pressure
(cp>
Btu/lb-°F
Absolute
Viscosity
(u)
lb
hr-ft
Thermal
Conductivity
(*)
BtU
hr-ft-°F
Prandt
Number
(Cu/k)
dimen-
-sionless
Density
(P)
lb/ft3
0
0.240
0.040
0.0124
0.77
0.0863
20
0.240
0.041
0.0128
0.77
0.0827
40
0.240
0.042
0.0132
0.77
0.0794
60
0.240
0.043
0.0136
0.76
0.0763
80
0.240
0.045
0.0140
0.76
0.0734
100
0.240
0.047
0.0145
0.76
0.0708
120
0.240
0.047
0.0149
0.76
0.0684
140
0.240
0.048
0.0153
0.76
0.0662
160
0.240
0.050
0.0158
0.76
0.0639
180
0.240
0.051
0.0162
0.76
0.0619
200
0.240
0.052
0.0166
0.76
0.0601
250
0.241
0.055
0.0174
0.76
0.0538
300
0.241
0.058
0.0182
0.76
0.0521
350
0.241
0.060
0.0191
0.76
0.0489
400
0.241
0.063
0.0200
0.76
0.0460
450
0.242
0.065
0.0207
0.76
0.0435
500
0.242
0.067
0.0214
0.76
0.412
700
0.243
0.076
0.0243
0.76
0.0341
1,000
0.246
0.089
0.0283
0.77
0.0275
1,400
0.251
0.105
0.0328
0.80
0.0212
1,800
0.257
0.120
0.0360
0.85
0.0175 1
Lesson 1
1-5
General Principles
-------
Relative saturation is then defined as the ratio of Slides
the partial pressure of water vapor present to that
which would be present if the air were saturated: 1-11
Relative saturation = Pwater/PHater .t duration (Eqn. 8)
It should be noted that relative saturation is also
equal to the ratio of the corresponding mole fractions.
Relative humidity is simply relative saturation
multiplied by 100 to express it in percent.
Absolute or specific humidity is the weight of
water vapor per weight of dry air, usually expressed as
pounds of water per pound of dry air.
Dry-bulb, Wet-bulb and Dew-point temperatures 1-12
Temperature that is measured with a standard
thermometer, or an equivalent device, is termed dry-bulb
temperature. If you take that same standard thermometer
and place a porous wick over the sensing bulb, you will
have created a wet-bulb thermometer. As you move this
thermometer through the air, or place it in a moving air
stream, water from the wick will evaporate. When this
happens, the wick cools down and continues to cool until
the rate of energy transferred to the wick from the air
equals the rate of energy loss caused by the evaporating
water. The temperature of the bulb when the wet wick is
at equilibrium is termed wet-bulb temperature. Since
the rate of evaporation will depend on the moisture
content of the air, wet-bulb temperature provides an
indication of the humidity of the air.
It can also be shown that, for water only, the wet-
bulb temperature is essentially the same as the
adiabatic saturation temperature. Adiabatic processes
are simply those processes which occur without
exchanging heat with the surroundings. For example,
cooling of a gas stream by evaporating water is a
process that can usually be considered adiabatic. As
this process proceeds, the amount of moisture in the gas
stream and the gas stream temperature will always give
approximately the same wet-bulb temperature. We will
use this property later to estimate cooling water
requirements for evaporators.
Dew-point temperature is the temperature at which
condensation begins as moist air is gradually cooled.
Lesson 1
1-6
General Principles
-------
More precisely, it is the saturation temperature
corresponding to the absolute humidity.
Slides
Enthalpy 1-13
Enthalpy is a measure of the thermal energy of a
substance. Common units for ventilation work are Btu/lb
or calories/gram. The Btu, or British thermal unit, is
defined as the amount of heat necessary to raise one
pound of water from 59°F to 60°F at a pressure of one
atmosphere. The calorie is the amount of heat required
to raise one gram of water at one atmosphere from 14.5°C
to 15. 5°C.
The enthalpy of a substance at a given temperature
has no practical value except in relation to the
enthalpy at another temperature or condition. Since 1-14
enthalpy differences are proportional to temperature
differences, arbitrary datum temperatures may be chosen
to define enthalpy at any other temperature:
h - Cp(t-tref) (Eqn. 9)
where
h = enthalpy (Btu/lb)
C = heat capacity at constant pressure
(Btu/lb-°F)
t = temperature of substance (°F)
tref= reference temperature (°F)
For gases the reference temperature is typically 0 °F,
although this is highly variable, while for water it is
usually 32 °F.
The heat capacity, Cp, is a function of temperature
and is determined from tabulations in reference texts.
Values for air over a limited range of temperature are
provided in Table 1-1.
The enthalpy of water vapor is equal to the 1-15
enthalpy of the water plus the latent heat of
vaporization, A v:
vapor = ^water + A y (Eqn. 10)
Like C , the latent heat of vaporization is also a
function of temperature and can be found in reference 1-16
texts. The enthalpy of an air-water vapor mixture is
given by:
Lesson 1
1-7
General Principles
-------
^mixture ^dry air ^ ^water vapor^
(Egn. 11)
Slides
where 0 = absolute humiditv Mb /ih ^
- of
Usually we are interested in the enthalpy 1-17
difference, AH, between two temperature conditions,
since this represents the amount of heat that must be
added or removed in order to cause the change:
AH = h2 - h1
AH = (Cp)2(t2 - tref) - (CpMt, - tref) (Egn. 12)
For exact calculations, the heat capacity corres-
ponding to each temperature condition must be used.
Approximate results can be obtained, and the calcu-
lations considerable simplified, by using a heat
capacity averaged over the range of temperature change.
For air, this typically results in only a small error.
Assuming tref is the same for both enthalpies, our
relationship for enthalpy change then becomes:
An even simpler determination of enthalpy change
can be made by using tabulated values of enthalpy as a
function of temperature. One such tabulation has been
included as Table 1-2. Thus, if one were interested in
the amount of energy that must be removed in order to
cool an air stream from 500 °F to 100 °F, one need only
subtract the corresponding enthalpy values:
AH - hjoo - h100
« 106.7 Btu/lb - 9.6 Btu/lb
=97.1 BtU/lb
If enthalpy values from different information sources
are used in this manner, it must be remembered that the
choice of a reference
Lesson 1 General Principles
^ -1,)
(Egn. 13)
1-8
-------
Table 1-2. Enthalpies of Various Gases in Btu/Lb.
(Danielson, 1973)
Temp.
(°F)
co2
N2
h2o
°2
Air
100
5.8
6.4
17.8
8.8
9.6
150
17.6
20.6
40.3
19.8
21.6
200
29.3
34.8
62.7
30.9
33.6
250
40.3
47.7
85.5
42.1
45.7
300
51.3
59.8
108.2
53.4
57.8
350
63.1
73.3
131.3
64.8
70.0
400
74.9
84.9
154.3
76.2
82.1
450
87.0
97.5
177.7
87.2
94.4
500
99.1
110.1
201.0
99.5
106.7
600
124.5
135.6
248.7
123.2
131.6
800
176.8
187.4
346.4
171.7
182.2
1,000
231.9
240.5
447.7
221.6
234.1
1,200
289.0
294.9
552.9
272.7
287.2
1,400
347.6
350.5
661.3
324.6
341.5
1,600
407.8
407.3
774.2
377.3
396.8
1,800
469.1
464.8
889.8
430.4
452.9
2,000
531.4
523.0
1,003.1
484.5
509.5
2,200
594.3
582.0
1,130.3
538.6
567.1
2,400
658.2
642.3
1,256.8
593.5
625.0
2,500
690.2
672.3
1,318.1
621.0
654.3
3,000
852.3
823.8
1,640.2
760.1
802.3
8 Note: The enthalpies tabulated for H20 represent a gaseous system,
and the enthalpies do not include the latent heat of vaporization.
It is recommended that the latent heat of vaporization at 60 °F
(1,059.9 Btu/lb) be used where necessary.
Lesson 1
General Principles
-------
temperature is arbitrary and may not be the same for all Slid
tabulations. Before subtracting two enthalpy values,
you must confirm that they were determined for the same
reference temperature.
Psychometric chart 1-18
Many of the values for air-water vapor mixtures
that are used in ventilation calculations are avail-
able in a graphical representation known as a psycho-
metric chart. Two of these charts, covering a range of
temperatures, have been included in the Appendix of this
manual. As can be seen from perusing these figures, the
information contained on a given chart varies. Figure
1-1 is a chart schematic that shows the location of
possible information, and an explanation of each item
follows.
(Based on: Morse, 1965)
O Pendragon Inc. 1992
1. Dry-bulb temperature: The temperature of air
read on a standard thermometer is shown on
the chart by straight vertical lines. The
scale is at the bottom of the chart.
Lesson 1
1-10
General Principles
-------
2. Absolute humidity: The weight of water vapor Slides
per weight of dry air. On the chart these
lines are horizontal and at right angles to
the dry-bulb temperature lines.
3. Absolute humidity scale: The absolute
humidity at any point on the chart is read on
this scale.
4. Wet-bulb temperature: The temperature
indicated by a thermometer whose bulb is
covered by a wet porous wick and then exposed
to a stream of air. The lines are straight
and slope from upper left to lower right,
relative to the dry-bulb temperature lines.
The scale is on the curved line at the left
edge of the chart.
5. Specific volume: The volume of mixture per
weight of dry air. The lines are straight
and slope from upper left to lower right, at
a sharper angle than the wet-bulb temperature
lines. The value is located along each line.
6. Enthalpy: The heat energy contained in a
weight of dry air. The scale is located
beyond the left edge of the chart and is read
along extensions of the wet-bulb temperature
lines. On some charts this scale represents
the enthalpy at saturation only. Corrections
for non-saturated conditions are provided
along nearly vertical lines within the chart.
7. Dew-point temperature: The temperature at
which moisture begins to condense. The value
is read on the wet-bulb temperature scale
along a horizontal line of constant absolute
humidity.
8. Relative humidity: The ratio of the partial
pressure of water vapor in the air to the
partial pressure at saturation. The lines
are curved and extend from lower left to
upper right, relative to the dry-bulb
temperature lines. The value is located
along each line.
9. Vapor pressure: The pressure exerted by the
water vapor in the air. The scale is on the
Lesson 1 General Principles
1-11
-------
far right of the chart and is read along a
horizontal line of constant absolute humidity.
10. Sensible heat ratio: The ratio of the
sensible heat to the total heat of a process.
These values are typically used for
calculations related to conditioned air
supply and are not employed in this course.
In order to use the psychometric chart, one must
first locate the position on the chart that corres-
ponds to the conditions of the air stream. This is done
by knowing any two of the above quantities and locating
the point of intersection along their corresponding
lines. Once this point is determined, values for the
other quantities may be read from the appropriate
scales. In some cases the beginning and ending points
of a process may be located, and the changes in values
determined by subtracting the quantities corresponding
to each condition.
one value that is of interest but cannot be read
directly from the psychometric chart is the density of
the air-water vapor mixture. However, this can be
determined from values obtained from the chart, as
follows:
"War, - + ">
where 0 = absolute humidity (lb/lb. . )
u « specific volume (ft3/lbdry ajr)
II. PRINCIPLES OP FLUID PLOW 1_lg
Continuity
Consider the flow of fluid through a tube as shown 1-20
below:
© Pentagon Inc. 1992
Lesson 1
General Principles
1-12
-------
The mass rate of flow through the tube (e.g., lb/min) is Slides
given by G — pVA and the volume rate of flow (e.g.,
ft3/min) is given by Q = G/p - VA. Here, p is the fluid
density, V is the fluid velocity and A is the tube 1-21
cross-sectional area. If there is no accumulation or
removal of material between points 1 and 2, then we may
write:
G, - G2 (Eqn. 15)
or
p,V,A, - pzV2Aj, (Eqn. 16)
If the fluid is incompressible or, as is usually the
case in ventilation systems, the pressure is low enough
that the fluid may be considered incompressible, then py
= p2 and:
V^A1 « (Eqn. 17)
This relationship allows for the determination of
velocity change as a gas stream flows through ducts of
different diameter.
Momentum
As a fluid flows through a duct, its momentum and
pressure may change. The magnitude of the change may be
determined by applying the relationship, force equals
rate of change of momentum, to a fluid element and then
integrating over the cross-section of the duct. If
frictional forces and compressibility effects are
neglected, the relationship that is obtained is referred 1-22
to a Bernoulli's equation:
Vz/2 + P lp + gz = constant (Eqn 18)
where V « fluid velocity
P = fluid pressure
p « fluid density
g ¦ gravitational acceleration
z » height above a reference datum
Although this equation strictly applies only to
incompressible, inviscid fluids, it is of significant
Lesson 1 General Principles
1-13
-------
importance because it relates the pressure at a point in
a fluid to its position and velocity and does so in a
rather simple way.
Since the conditions assumed in the development of
Bernoulli's equation are approximated in most ventilat
ion systems, it will be used to develop several useful
relationships. Rearranging gives:
V2/2g + P/pg + z = constant (Eqn. 19)
Iri this form, each term represents energy per unit
weight of fluid and has dimensions of length. Thus,
each term may be regarded as representing a contribution
to the total fluid head:
V2/2g = velocity head
P/Pg = pressure head
z = potential head
Consider the following situation, in which an open
tube has been inserted into a flowing fluid. The
pressure of the flowing fluid causes the fluid to rise
to a level, h, in the open tube.
©Pendragon Inc. 1992
1 2
i i
i i
i i
"1
i i
h
i i
i i
S-d
i
w » *
&
> i *—J-
~
Figure 1-3. Open Tube in Flowing Fluid
Lesson 1 General Principles
1-14
Slides
1-23
1-24
-------
Since the terms in Bernoulli's equation sum to a Slides
constant, we may write:
(V1)2/2g + Py/pq + 2, - (V2)2/2g + p2/pg + z2 (Eqn. 20)
The position of points 1 and 2 are on the same
level, so z1 = z2. Also, the fluid at point 2, just at
the entrance to the tube, is balanced by the fluid in 1-25
the tube, so the velocity at this point is zero.
Substituting gives:
(V1)2/2g + p,/pg = P2//>g (Eqn. 21)
or
(V1)2/2g = velocity head
Pi/pg - pressure head
P2/pg = total head
Expressing these energy heads as pressures, we may write
this relationship in its more common form:
VP + SP = TP (Eqn. 22)
where VP = velocity pressure
SP « static pressure
TP » total pressure
Thus, at any point in a flowing fluid, the total
pressure is the sum of the velocity pressure and the
static pressure. This relationship is illustrated in 1-26
Figure 1-4 for an air stream on either side of a fan.
Here, the manometers with one leg connected to ports
that are perpendicular to the flow streamlines and the
other leg open to the atmosphere measure static
pressure. The manometers with one leg connected to the
tubes facing into the flow and the other leg open to the
atmosphere measure total pressure, which is the sum of
the static and velocity pressures. The manometers with
one leg connected to the perpendicular ports and the
other leg connected to the tubes measure the difference
between total and static pressure, which is velocity
pressure. It should be noted that, while static and
total pressures are usually negative upstream of a fan
and positive downstream of a fan, velocity pressure is
always positive.
Lesson 1 General Principles
1-15
-------
Slides
© Pendragon inc. 1992
In the above development the velocity pressure, VP
was given by the term, V2/2g, in units of length of
fluid. In measuring velocity pressure, we typically use
the pressure of the flowing fluid to displace fluid in
a manometer, which we read in inches of water column.
Converting the units of the velocity pressure term so
that it has unit of inches of water gives:
VP 38 [ (V/60)2/2g] (pjpv) 12 (Eqn. 23)
where V * air velocity (ft/min)
pB = air density (lb/ft3)
pu = water density (lb/ft3)
g - acceleration of gravity (ft/sec2)
Substituting a water density at 70 °f of 62.302 lb/ft3 1-27
and a gravity acceleration at sea level of 32.174
ft/sec2 gives:
VP = p,(V/1096.7)2. (Eqn. 24)
Lesson 1
1-16
General Principles
-------
Since we usually measure the velocity pressure and
use that to calculate the air velocity, the more useful
form is:
V = 1096.7 (VP//0a)0*5 (Eqn. 25)
For standard air, pa = 0.075 lb/ft3 and our relationship
becomes:
V = 4005 (VP)0,5 (Eqn. 26)
References
ACGIH, Industrial Ventilation, Twentieth Edition, Cincinnati, 1988.
Coulson, J.M., and J.F. Richardson, Chemical Engineering, Volume
One, Second Edition, MacMillan, New York, 1964.
Danielson, J.A., ed., "Air Pollution Engineering Manual", Second
Edition, EPA AP-40, May 1973.
Himmelblau, D.M., Basic Principles and Calculations in Chemical
Engineering, Fourth Edition, Prentice-Hall, Englewood Cliffs, 1982.
Jorgensen, Robert, ed., Fan Engineering, Seventh Edition, Buffalo
Forge Company, Buffalo, 1970.
Morse, F.B., ed., Trane Air Conditioning Manual, The Trane Company,
La Crosse, 1965.
Lesson 1
1-17
General Principles
-------
LESSON 2
HOOD SYSTEMS
The hood constitutes one of the most important components of
the industrial ventilation system. When properly designed, it is
instrumental in containing or capturing contaminants released by
industrial processes. If the process is located outside, it is
directly responsible for preventing the release of emissions to the
atmosphere. When the process is located inside a building, it
serves to prevent the release of contaminants to the workspace and
to prevent fugitive emissions from building openings.
The goal of good hood design is high capture efficiency. The
importance of this in relation to the total system is illustrated
by the following equation:
pthood + (1 " Pthood)Ptcollector (Eqn. 1)
Here, Pt is the penetration, which is one minus the fractional
efficiency. Thus, even if the collector were 100 percent efficient
(Pt = 0), the overall penetration of the system can be significant
if the contaminants are not effectively captured by the hood.
To be able to conduct effective inspections of hood systems,
it is important to understand the concepts behind good hood design
and to know how to evaluate hood performance. In this lesson, we
will discuss the various types of hoods, the principles that govern
their design and the factors that affect their performance. We
will also discuss the pressure losses associated with flow into a
hood, and how knowledge of those losses can be used to estimate air
volume.
I. HOOD TYPES
All hoods can be classified as belonging to one of Slides
four types: (1) enclosure, (2) receiver, (3) exterior
and (4) push-pull. Enclosure hoods, as the name 2-4
implies, envelop the process to the maximum extent
possible. Typically, the designer begins by envisioning
a total enclosure around the process and then removes
portions of the hood only as much as is need to provide
material and worker access. Because of the nature of
this type of hood, it serves not to capture the 2-5
contaminants but rather to contain them and remove them
from within the enclosure. As a result, the quantity of
air flow required for a given process is usually the
least of the three hood types.
Lesson 2
2-1
Hood Systems
-------
Ficrure 2-1, a bucket elevator, is an example of an
enclosure-type hood. Buckets nountet cm * rotating
vertical belt are used to ti-ansfer materiai fromone
elevation to another. To reduce emissions from this
process, a housing is provided that completely encloses
the operation, and suction is provided to contain and
remove the contaminants. The onlyopenirigs »re tteC
necessary for receiving and transferring the material.
The ladle hood shown in Figure 2-2 is also an
enclosure-type hood. Molten metal is transferred to a
hot metal ladle from a rotating-dump torpedo car. The
ladle hood contains and removes the emissions as they
evolve from the transfer operation. Once the emissions
have subsided, the ladle hood is tilted out of the way
so that the ladle may be moved to other operations.
A receiving hood, sometimes thought of as one type
of exterior hood, is located adjacent to the point of
contaminant release and in orientation that allows it to
receive the emissions as they are ejected from the
process. Since the hood is located along the direction
of normal contaminant travel, the amount of capture
Slides
2-6
Figure 2-1. Bucket Elevator Enclosure
O Pendragon Inc. 1992
Lesson 2
2-2
Hood Systems
-------
Torpedo
Car
Ladel
Hood
Hot Metal
Ladle
L
'=0
Figure 2-2. Ladle Hood (Kemner et.al. 1984)
Slides
O Pendragon Inc. 1992
capability that must be provided by the air stream is
reduced.
The grinding wheel hood shown in Figure 2-3 is a
receiving-type hood. Material removed by the wheel has
a normal travel direction down and to the rear.
The hood is mounted in this location to take advantage
of this and reduce the necessary capture flow.
2-9
O Pendragon Inc. 1992
Lesson 2
2-3
Hood Systems
-------
The hood also extends around the top and sides of the
wheel to provide for enclosure of any contaminants that
follow the wheel motion.
The bag filling process shown in Figure 2-4 also
utilizes a receiving-type hood. To avoid interfer-
ences with the weighing scale, the hood is mounted above
the bag opening to take advantage of the normal vertical
travel of emissions and somewhat reduce the capture flow
that might be required with another orientation.
A hood that is mounted an extended distance away
from the contaminant source is referred to as an
exterior hood. The principal example of this type hood
is the overhead canopy typically employed with hot
sources. Because of the limited capability of hoods to
capture and draw-in contaminants from a long distance
away, this type of hood relies almost totally on the
normal movement of the buoyant plume to carry them into
the hood. Because the distance of plume travel may
sometimes be 20-40 feet, this type hood is particularly
subject to losses due to plume meander or cross-drafts
that carry the plume out from under the hood. It should
also be noted that as a buoyant plume rises it expands
because of entrainment of outside air, making exhaust
volume requirements the largest of the three hood types.
Slides
2-8
2-io
7
Bag
Figure 2-4. Bag Filling Hood
© Pendragon Inc. 1992
Lesson 2
2-4
Hood Systems
-------
One important use of overhead canopy hoods is in Slides
the control of emissions from electric-arc furnaces used
for steel production. A typical system is shown in
Figure 2-5, where a canopy hood is used in combination
with an enclosure-type hood at the furnace. During the
melting cycle, emissions are controlled from the hood
mounted on the furnace, with only a small amount of flow
drawn from the canopy to remove any contaminants that
might escape the furnace. During charging operations,
the roof swings off the furnace to provide access, and
all air flow is directed to the canopy to collect the
significant plume that usually results. In the tapping
cycle, the furnace tilts to pour the molten steel into
a ladle, disconnecting the furnace hood from the duct
through a break-flange arrangement. In the system
shown, air flow is then directed to an enclosure hood at
the ladle. In other systems, the air flow is again
directed to the canopy hood for contaminant capture and
removal.
Lesson 2 2-5 Hood Systems
-------
A variation of the exterior hood is the push-pull Slides
system shown in Figures 2-6 and 2-7. Here, a control-
led jet of air is directed across the contaminant source 2-n
and in the direction of the exterior hood. The exterior
hood primarily serves to receive the air jet and the 2-12
contaminants carried with it. The advantage of this
arrangement is that velocities can be maintained at a
higher level with distance from the jet source than can
be maintained with distance from the exhaust source,
considerably reducing the air volume requirements.
However, the open area and air flow of the hood must be
careful designed and the system well controlled to avoid
dispersing contaminants into the workspace. Also,
disturbance of the jet can occur if obstructions are
placed in its path, thereby reducing the system's
effectiveness.
Figure 2-6. Push-Pull Hood for Open Tank
© Pendragon inc. 1992
The various hood designs presented above serve not
only to illustrate the four design types but also to
indicate some of the variety of designs that are
employed. Unfortunately, an all-encompassing
presentation of hood designs is not possible, for the
variety of hood designs is as large, if not larger, than
the variety of industrial processes they control.
Lesson 2
2-6
Hood Systems
-------
Exhaust
Side
Baffle
Wall
Fugitive
/ / Emissions
Baffle
Wall
To Suction Fan
and Hood
Sample Location
Figure 2-7. Push-Pull Hood for Copper Converter
Slides
2-11
© Pendragon Inc. 1992
One important reference that provides information on the
design, air flow requirements and resistance
characteristics of a large number of industrial hoods
is, Industrial Ventilation, A Manual of Recommended
Practice, published by the American Conference of
Governmental Industrial Hygienists.
II. HOOD DESIGN PRINCIPLES
Although the design of hoods can be a complex
process that at times leans more toward art than
science, the basic principles that govern that design
are surprisingly simple and straight-forward:
1. Whenever possible, use an enclosure hood.
2. If an enclosure hood cannot be used, the hood
should be placed as close to the source as
possible and aligned with normal contaminant
flow.
3. To improve hood performance, duct take-offs
should also be placed in-line with normal
contaminant flow.
Adherence to these basic principles will result in a
hood system that gives high capture efficiency while
utilizing the minimum air flow necessary.
Lesson 2 2-7 Hood Systems
-------
Factors affecting hood perform*®®* Slides
Effective capture of contaminants by a hood system
relies on air flow toward the hood face. This air flow
must be sufficient to maintain control of the contamin 2-14
ants until they reach the hood. Of particular concern
is external air motion that may disturb this flow and
cause loss of the contaminant or require higher than
normal air velocities to maintain control, sources of
air motion that must be considered when designing and
placing hoods include:
• Room air currents associated with the workspace
ventilation system. These can become quite
large when windows and doors are opened.
Currents of as little as 50 feet/min may be
enough to affect the performance of some hoods.
• Thermal air currents fro® heat generating
equipment and processes. Even low heat
releases, such as those from an electric motor,
may be enough to disturb hood performance.
• Machinery motion. Rotating or reciprocating
machinery can be a source of significant air
currents.
• Material motion. Downward motion of material,
for example, will create a downward air current
that will make the upward notion of
contaminants more difficult to achieve.
• Operator movements. Rapid movements of an
operator can create air currents of 50-100
feet/min.
Capture velocity
Capture velocity is defined as that air velocity at 2-15
a point in front of a hood or at the hood face that is
necessary to overcome existing air currents and cause
the contaminated air to move into the hood. The needed
capture velocity will depend on both the direction and
velocity of the contaminants at the desired point of
capture, as well as the level of disturbing air currents
that must be overcome. An overhead canopy that relies
primarily on plume buoyancy to convey the contaminants
to the hood will require little capture velocity,
generally just enough to match the plume velocity at the
hood face. Contaminants generated by a high energy
process that results in rapid and random contaminant
motion will require quite high capture rates. A general 2-16
guide for appropriate capture velocities in several
Lesson 2
2-8
Hood Systems
-------
appropriate capture velocities in several situations is
provided in Table 2-1. Values at the low end of the
range would be appropriate when disturbing air currents
are low, the toxicity of the contaminants is low, or the
hood is large, resulting in a large air mass in motion.
The higher end of the range would be more appropriate
when air currents are high, the toxicity of the
contaminants is high, or the hood is small.
Table 2-1. Range of Capture Velocities
Type of Material Release
Capture Velocity
(feet/Minute)
With no velocity into
quiet air
50 - 100
At low velocity into
moderately still air
100 - 200
Active generation into
zone of rapid air motion
200 - 500
With high velocity into
zone of very rapid air
motion
500 - 2000
cold flow into hoods
For successful performance, most hood systems
rely totally or in part on the ability to provide
enough energy to capture a contaminant and draw it
into the hood, i.e., to develop the necessary capture
velocity. As previously indicated, the capability of
an exhaust flow to maintain velocity beyond the hood
face is considerably limited. This is illustrated in
Figure 2-8, which shows lines of constant velocity as
a function of distance from the hood face for both
flanged and unflanged hoods. In general, we see that
the capture velocity one hood diameter away from the
hood face is less that 10 percent of the velocity at
the hood face. Although the situation is improved
with the addition of a flange, the improvement is only
about 25 percent. In contrast, 10 percent of the face
velocity of a blowing jet would be found about thirty
diameters away. Thus, if one seeks to provide a large
capture velocity a significant distance from an
exhaust hood, very large and possibly impractical
Lesson 2
2-9
Hood Systems
-------
face velocities nay be required,
expecting a hood to provide high
several diameters away from the
expecting too much.
In other words, Slides
capture velocity
hood face may be 2-17
Law*
Figure 2-8. Hood Velocity Contours
(Percent of Hood Capture Velocity)
2-18
and
2-19
O Pendragon Inc. 1992
Figure 2-9 gives relationships for determining air
volume requirements to provide a desired capture
velocity a given distance from the hood face for several
hood configurations. For an existing hood, these same
equations could be used to estimate capture
Lesson 2
2-10
Hood Systems
-------
Hood Type
D—cHptton
Aepeel Ratio
Air How
;/
Slot
0.2otLms
0-3.7LVX
x>^-
Flanged Slot
0.2 or Lew
Q-2.6LVX
7
Plain Opening
0.2 or Graew
and Round
0-V(10X2+A)
A-WL («q. ft.)
Ranged Opening
O^orOteeier
and Round
0- 0.75V (10X2 ~ A)
x-^s
rT
Booth
To Suit Woifc
Q - VA - VWH
Canopy
To Suit Work
Q» 1.4 PVO
See VS-903
P * PiWMttr
D • Height Above Wort
Ki
Plain Multipt*
Stot Opening
2or Mom Slots
O^orGraater
Q-V(10X2+A)
Flanged Multiple
Stot Opening
2 or Mora Slots
0.2 or Greater
Q- 0.75V (10X2 «• A)
Figure 2-9. Air Flow Relationships for Various
Hood Types (ACGIH, 1988)
Slides
2-20
Lesson 2
2-11
Hood Systems
-------
velocity once the hood flowrate is known. A method for
estimating hood flowrate will be present later in this
lesson, and a method for measuring flowrate will be
discussed in Lesson 6.
Hot flow into hoods
Many hot sources utilize canopy hoods mounted over
and well above the source for capture and removal of
contaminants. As indicated previously, these hoods rely
more on the buoyancy of the hot plume to carry the
contaminants into the hood than on the ability to
generate a capture velocity. In general, the velocity
at the hood face need only be about the velocity of the
plume at that point. As a hot plume rises it expands
and cools by entraining outside air and its velocity
decreases. Thus, as the distance between the source and
the hood increases, the air volume required to capture
and remove it increases. Also, the slower moving upper
portion of the plume becomes more susceptible to being
disturbed by air currents, a particular problem with
high canopy hoods.
Hood pressure losses
To cause air to move into a hood it is necessary to
provide the energy needed to accelerate the air from
essentially zero velocity up to the velocity in the duct
connected to the hood and to overcome the entry
resistance of the hood itself. Consider the following
hood system:
Slides
2-21
2-22
Position 2
Position 1
Figure 2-10. Hood System
O Pendragon Inc. 1992
Lesson 2
2-12
Hood Systems
-------
Applying the Bernoulli equation from Lesson 1 to this
situation would indicate that the total pressure at
point 1 would equal the total pressure at point 2, or:
Slides
SP, + VP1 = SP2 + VP2 (Eqn. 2)
But, as noted, there is essentially no air motion at 2-23
point 1; therefore:
0 = SP2 + VP2 (Eqn. 3)
or
SP2 = -VP2 (2-4)
In reality, as air enters a hood or duct a "vena 2-24
contracta" forms, as illustrated in Figure 2-11.
Figure 2-11. Vena Contracta
Following the continuity relationship, this reduction
in effective cross-section within the vena contracta
causes an increase in velocity. Then, as the con-
tracta expands, the velocity decreases. This is not a
perfect process and results in an energy loss as static
pressure is converted to velocity pressure and then back
to static pressure. Therefore, we define the hood
static pressure loss, SPh, as:
SPh « -SP2 - VP2 + he (Eqn. 4)
where he * hood entry loss.
Hood entry loss includes entry and frictional losses
into the hood and entry losses from the hood into the
duct.
Lesson 2
2-13
Hood Systems
-------
The hood entry loss is usually expressed as some Slides
fraction of the velocity pressure in the duct attached
to the hood:
he = FhVP (Eqn. 5)
where Fh = hood loss factor
Hood losses may also be described by the hood entry
coefficient, Ce: 2-25
Ce - (VP/SPh)0*5 (Eqn. 6)
The hood entry coefficient can be related to the hood
loss factor by recalling that SPh = VP + he. Thus:
(Equation Set 2-7)
Ce - [VP/(VP + he)]0-5
Ce2 = VP/(VP + he)
he = VP/Ce2 - VP = [(1 - Ce2)/Ce2]VP
2-26
But he = FhVP; therefore:
Fh = (1 - Ce2)/Ce2 (Eqn. 8)
and
Ce - [1/(1 + Fh)]0-5 (Eqn. 9)
Values of Fh are given in Figure 2-12 for several hood
configurations.
Evaluation of hood performance
Measurements of the hood static pressure can be
used to estimate the flowrate at the hood. Recalling
the relationship between velocity and velocity pressure
from Lesson 1 and noting that the flowrate, Q, is equal
to the velocity, V, times the cross-sectional area, A,
gives:
Q - VA » 1096.7A(VP//7)0,5 (Eqn. 10) 2-27
Lesson 2
2-14
Hood Systems
-------
Plain Duel End
h,. 0.93 VP
Flanged Duct End
It* -0.49 VP
f Ieau2 times
Tapered Hood*
Flanged or untanged;
round, aqua* of
recsangulor.
• is fie major angle
on racanguwr hoods.
Entry Lou
8 Round Rectangular
IS* 015 VP 0.2SVP
30* 0.08 VP 0.16 VP
45* 0.06 VP 0.15 VP
60* 0.06 VP 0.17 VP
90* 0.15 VP 0.25 VP
120- 0.26 VP 0.35 VP
150* 0.40 VP 0.48 VP
r-R-D/2
Booth Ptu» Rounded Entrance
h» . 0.06 VP * 0.10 VP
Orifice Plus Flamed Dud
1.78 VP Orifice ~ 0.49 VP Duct
Figure 2-12. Hood Entry Loss Factors
Slides
Then, recalling that Ce = (VP/SPh)0*5, substitution gives:
Q = 1096.7ACe(SPh/p)0-5 (Eqn. 11)
To use this relationship to estimate hood flowrate,
a static pressure measurement would be made in the duct
connected to the hood and downstream of the vena
contracta, usually about 2 duct diameters from the duct
connection to the hood. Then the density of the air
stream would need to be estimated directly or from a
temperature measurement. Finally, the configuration of
the hood would be used to determine Ce from reference
information like Figure 2-12. Substituting these
values, along with the duct cross-sectional area, into
the above relationship would yield the flowrate
estimate. It should be noted that once an acceptable
hood static pressure has been determined, i.e., one that
results in adequate capture velocity, subsequent
inspections need only confirm this value. Techniques
for measuring static pressure and temperature will be
discussed in Lesson 6.
As a minimum, the performance of a hood should be
visually evaluated. If dusty material is being handled
Lesson 2
2-15
Hood Systems
-------
by a process, the amount of fugitive loss provides an
excellent indication of the effectiveness of the hood
system. Refraction lines due to the escape of gaseous
or vaporous contaminants may also be noticed. In
addition, the physical condition of the hood should be
assessed. Particular attention should be paid to any
modifications that have been made to the original hood
design or to any damage that it may have sustained that
may affect its performance. On movable hoods, the
connection between the hood system and the duct system
should be assessed to determine the "fit" of the
junction. Break-flanges should have a maximum gap of 1-
1% inches. Finally, the hood position should be
evaluated to assess the effects of cross-drafts or other
air motion on hood capture efficiency.
References
ACGIH, Industrial Ventilation, Twentieth Edition,
Cincinnati, 1988.
Hemeon, W.C.L., Plant and Process Ventilation, Second
Edition, Industrial Press, New York, 1963.
Kashdan, E.R., D.W. Coy, J.J. Spivey, T. Cesta and H.D. Goodfellow,
"Technical Manual: Hood System Capture of Process Fugitive Particulate
Emissions", EPA-600/7-86-016, April 1986.
Kemner, W., R. Gerstle and Y. Shah, "Performance Evaluation Guide for
Large Flow Ventilation Systems", EPA-340/1-84-012, May 1984.
Lesson 2
2-16
Hood Systems
-------
LESSON 3
DUCT SYSTEMS
Once contaminants from a process have been captured by the
hood system, it is the responsibility of the duct system to convey
these contaminants to the collection device and then convey the
cleaned air on to its discharge point. In designing duct systems,
much of the concern is in selecting proper size ducts and in being
sure the system is "balanced" so that the proper quantities of air
are drawn from each hood. This involves selecting the proper
transport velocity, choosing the duct sizes necessary to maintain
that velocity, determining the pressure loss or resistance of each
section of the duct system and then being sure that the resistance
of each branch entering a junction is the same.
As inspectors, the design of a duct system is of only limited
concern. However, it may be necessary to use some of the same
tools as the designer in order to accomplish our goals. For
example, suppose we wish to estimate the flow into a high canopy
hood. From Lesson 2, we know that this can be done by measuring
the static pressure in the duct just beyond the hood. In this
case, however, that would require obtaining a measurement from a
difficult, and perhaps dangerous, location to reach. Instead, we
could measure the static pressure at some more easily reached and
safer location and use the principles of duct resistance to
estimate the air flow we require.
In this lesson, we will apply Bernoulli's equation to
determine the relationship between pressures at different points in
a duct system. This relationship will then be modified to account
for losses associated with the fluid actually having viscosity, and
techniques for estimating these losses will be presented. Also,
transport velocity will be defined and its significance discussed.
Finally, the concepts behind balancing ventilation systems will be
presented.
I. DUCT PRESSURE L08S
Consider the following duct segment shown in Slides
Figure 3-1. Applying Bernoulli's equation to points 1
and 2, we can write: 3-2
TP, = TP2 (Eqn. 1)
However, because of friction between the gas stream and 3-3
the duct walls and because of non-ideal conversion
between static pressure and velocity pressure as the
Lesson 3
3-1
Duct Systems
-------
Figure 3-1. Example Duct Segment
Slides
3-4
O Pendiagon Inc. 1992
gas stream accelerates or decelerates, pressure losses
between the two points occur. Thus:
TP, = TP2 + hL (Eqn. 2)
or
SP, + VP, - SP2 + VP2 + hL (Eqn. 3)
If we assume that the velocity between points 1 and
2 is approximately constant, then VP, = VP2 and:
SP, = SP2 + hL (Eqn. 4)
Here, hL is the total pressure loss due to friction and
non-ideal pressure conversions. For calculation
purposes, we divide these losses into three categories:
(1) frictional losses in straight duct, (2) fitting
losses, and (3) acceleration losses. The frictional
loss in straight duct is rather self-explanatory and
simply involves the loss due to
Lesson 3
3-2
Duct Systems
-------
friction with the walls. Fitting losses occur when the
gas stream flows through elbows, entries, transitions
and other types of fittings. This loss results from
friction with the walls of the fitting and with
increased energy loss due to an increased level of
turbulence. Acceleration losses (or gains) are
associated with changes in the velocity of the gas
stream. Accelerating a gas stream requires the input of
energy, while decelerating a gas stream may result in a
gain of energy. The amount of pressure loss or gain
depends on the relative abruptness of the change.
Smoother accelerations result in lower losses and
smoother decelerations result in higher gains.
Because our interest in duct losses will usually be
over relatively short distances, our determination of
losses will be confined to those associated with
straight ducts and fittings. There are three techniques 3-6
in general use for estimating these losses: (1) the
equivalent length method, (2) the velocity pressure
method, and (3) the total pressure method. In the
equivalent length method, fitting losses are expressed
in length of equivalent straight duct. Losses are
calculated by adding the equivalent length for the
fittings to the actual length of straight segments and
then multiplying by a factor that expresses the pressure
loss per length of duct. Since the equivalent length of
fittings depends on the duct size, the amount of
information required to conduct an equivalent-length
calculation can be large.
In the velocity pressure and total pressure
methods, both straight duct and fitting losses are
expressed as a factor times the velocity pressure or
total pressure in the duct segment, respectively, since
the fitting loss factor does not vary with the duct
size, the amount of information required for calculation
is considerably reduced in comparison to the equivalent-
length approach.
The technique to calculate duct system losses that
is used in this course is the velocity pressure method.
With this method, frictional losses in straight duct are 3-7
expressed as:
hL1 = HfL*VP (Eqn. 5)
where = velocity pressure loss per foot of
duct
L « length of duct, feet
Slides
3-5
Lesson 3
3-3
Duct Systems
-------
The straight-duct loss factor can be determined from the Slides
following empirical equation or from Figure 3-1.
Hf - o.0307V°-B3/Q°-*12
- 0.4937/Q°*"7'd1'0^ (Eqn. 6)
where V = air velocity in feet/minute
Q = volumetric flowrate in cubic
feet/minute
D - duct diameter in inches
In Figure 3-2, a point is located using any two of
flowrate, velocity or duct diameter and the
corresponding loss factor read on the vertical
scale. For rectangular ducts, an equivalent diameter is
determined from:
- 1.3 (A X B) 0*625/^ + B)0-25 (Eqn. 7)
where = equivalent diameter in inches
A = length of one side in inches
B - length of adjacent side in inches
Using the equivalent diameter an equivalent velocity is
calculated and the loss factor chart entered using the
equivalent values.
Fitting losses in the velocity pressure method are 3_o
expressed as:
hLZ « F*VP (Eqn. 8)
where F is the fitting loss factor.
There are a variety of fittings used in ventilation
systems and factors for their resistance can be found in
a number of reference texts (e.g., ACGIH, 1988; SMACNA,
1977). The fitting of most interest for short distance
calculations is that for elbows. Fitting loss factors
for 90° round elbows are given in Table 3-1 in terms of
the ratio of the radius of turn to the duct diameter,
R/D. Most round elbows have an R/D of 1.5 or 2.0, with
2.0 being the most common. For rectangular elbows, loss
factors are given in Table 3-2 in terms of the duct
aspect ratio, W/D and the el bow R/D. Here, D is the
length of the side parallel to and centered on the
turning radius and W is the length of the adjacent side.
Resistances for other than 90° elbows are determined as
Lesson 3
3-4
Duct Systems
-------
10.0
u
3
o
o
o
®
Li.
o
o
T"
w
-------
percentage of the 90° elbow resistance. Thus, the
resistance of a 60° elbow is two-thirds that of a 90*
elbow, while the resistance of a 45* elbow is half.
Table 3-2. Loss Factors for 90° Round Elbows
R/D
Fraction of VP Loss
1.25
0.55
1.50
0.39
1.75
0.32
2.00
0.27
2.25
0.26
2.50
0.22
2.75
0.26
Table 3-2. Loss Factors for 90° Rectangular
R/D
0.25
0.50
1.0
2.0
3.0
4.0
0
1.5
1.32
1.15
1.04
0.92
0.86
0.5
1.36
1.21
1.05
0.95
0.84
0.79
1.0
0.45
0.28
0.21
0.21
0.20
0.19
1.5
0.28
0.18
0.13
0.13
0.12
0.12
2.0
0.24
0.15
0.11
0.11
o
H
•
o
0.10
3.0
0.24
0.15
0.11
0.11
0.10
0.10
Velocity pressure calculation method
In general, calculating the pressure loss from
one duct location to another requires determining the
loss of the straight sections and the loss of all the
fittings and then adding them together to get the
total loss. The specific step-wise procedure for a
segment, beginning at the hood, is as follows:
1. Determine the duct velocity and calculate the
velocity pressure.
esson 3
3-6
Duct Systems
-------
2. Determine the hood static pressure. Slides
3. Multiply the straight duct length by the 3-9
friction loss factor. 3-10
and
4. Determine the number and type of all fittings. 3-11
Multiply the loss factor for each type of
fitting by the number of that type and sum for
all types.
5. Add the results from Steps 3 and 4 and multiply
the result by the velocity pressure from
Step l.
6. Add the result from Step 5 to the hood static
pressure from Step 2. If there are other
losses (expressed in inches of water column),
add them also.
This calculation procedure gives the total energy, 3-12
expressed as static pressure, that is needed to move the
gas volume through the duct segment.
Estimating hood flowrate
In the introduction, it was noted that one could
measure the static pressure at some safer or more easily
reached location and use the principles of duct
resistance presented in this lesson to estimate the air
flow at a hood. Since this application is not
altogether obvious, the relationship for doing this will
be developed here.
Recall from Lesson 1 the relationship for the hood
entry loss coefficient, Ce:
C. « (VP/SPh)0-5 (Eqn. 9)
or
SPh = VP/Ce2 (Eqn. 10)
The hood static pressure could be determined by making
a measurement anywhere in the duct leading to the hood
and then correcting the measured value for the losses
between the measurement location and the hood. Assuming
losses due to both straight duct and fittings:
SP„ = SP - HfL"VP - SF'VP (Eqn. 11)
Lesson 3
3-7
Duct Systems
-------
Equating the two relationships for hood static pressure
gives:
Slides
VP/Ce2 - SP^ - HfL*VP - SF'VP (Eqn. 12)
or
VP = SPmn/ (1/C.2 + HfL + SF) (Eqn. 13)
Hf a trial velocity will have to be assumed. The
corresponding velocity pressure would be compared to the
calculated value from Equation 3-13 and the calculations
repeated until reasonable agreement is obtained. Once
an acceptable velocity pressure has been determined, the
flowrate can be calculated from:
q = VA = 1096.7A(VP/Pa) °"5 (Eqn. 14)
or, if standard air is involved:
Q - 4005A(VP)°*5 (Eqn. 15)
Transport velocity
The velocity maintained within a duct segment is 3-13
referred to as the transport velocity. Typical design
values are given in Table 3-3. For systems conveying
vapors, gases or smoke, the velocity chosen by the 3-14
designer is based on a compromise between fan energy
cost and duct cost. Large diameter ducts result in
lower pressure losses and reduced fan energy but cost
more than smaller diameter ducts. in general, this
economically optimum velocity is around 1000-2000
feet/minute.
For systems handling particles, a minimum velocity
is required to prevent settling. The value of this 3-15
minimum velocity increases as the size or density of the
particles increases. Should settling occur, the
resistance of the duct will increase due to the
reduction in effective flow area. if the deposited
material is dry and loose, then an equilibrium condition
will develop when the level of deposited material causes
the velocity to increase to that needed to re-entrain.
At this point the build-up will cease, but there will be
a decrease in volume in that segment due to the
increased resistance. If, however, the material is
sticky or tends to form solid cake, the deposition may
continue until the duct is completely plugged.
In addition to increasing duct resistance,
deposition also increases the weight of a duct and may
Lesson 3
3-8
Duct Systems
-------
cause the duct system supports to fail. Also, hardened
material deposited inside the duct may break loose as a
result of vibration and travel down the duct to the fan
or other equipment, causing damage.
Another concern of particle conveying systems is
abrasion of the duct surface, a potential that increases
with increase in transport velocity. This is a
particular problem wherever there is a change in
direction of the gas stream, such as at elbows or
entries. The particles traveling with the gas stream
Table 3-3. Range of Design Transport Velocities
Contaminant
Design Velocity
(feet/Minute)
Vapors, gases,
and smoke
Any (usually 1000
to 2000)
Fumes
1400 - 2000
Very fine, light
dust
2000 - 2500
Dry dust and
powders
2500 - 3500
Average industrial
dust
3500 - 4000
Heavy dusts
4000 - 4500
Heavy or moist
4500 and up
possess a certain amount of inertia that will tend to
carry them straight ahead as the gas stream turns.
When this material strikes the duct walls, erosion can
produce holes. If the duct is under negative
pressure, air will enter the system through the holes,
reducing the amount of air that enters at the hood,
possibly causing loss of capture efficiency. If the
duct is under positive pressure, fugitive emissions
may result as air exits the holes. Also, segments
that have reduced air flow because of holes may become
susceptible to build-up problems due to the loss of
transport velocity.
Somewhat related to transport velocity is the
problem of duct corrosion. As the gases within a duct
cool, condensation of moisture and/or acidic material
Lesson 3
3-9
Duct Systems
-------
can occur, depending on the gas stream composition.
This cooling may be the result of infiltration of
outside air or simply due to long residence times in
the duct, which can be exacerbated by low transport
velocities. Condensed material will tend to
accumulate along the bottom of the duct and cause its
first damage there. Visual inspections should
concentrate on this area when corrosion damage is
suspected.
The best technique for locating holes in a duct
system is a visual inspection. To make more effective
use of inspection time, areas on an elbow that are on
the outside of the turn, areas on a straight duct
opposite points of entry by other ducts, and areas
along the bottom of horizontal ducts should be
emphasized. On hot gas streams, a significant drop in
temperature between locations along a duct could also
be used to locate holes in negative pressure systems,
as could an increase in oxygen concentration if the
source is oxygen deficient. However, because of the
time and equipment investment needed to make such
measurements, they are not likely to be practical in
most situations.
Unfortunately, inspecting for material build-up
in a duct system cannot be done effectively without
making measurements, since the accumulation is not
visible from the outside. Measured static pressure
differences between locations along a duct could be
compared to expected values estimated using the
techniques described above. Significant differences
between measured and calculated values would indicate
the location of a build-up. Alternatively, measure-
ments of static pressure along a duct could be plotted
as a function of equivalent length. If there are no
obstructions, the measurements will produce a straight
line of constant slope, with the values decreasing
(becoming more negative) in the direction of gas flow.
Any unexplained deviation from this slope would
indicate an area of accumulation.
Finally, it should be noted that physical damage
to ducts can also increase their resistance. A duct
that has been partially collapsed acts much the same
as a duct that has accumulated material. The reduced
cross-sectional area causes the velocity to increase
as it moves through the. damaged area and then decrease
as it moves out of it, adding to the duct losses.
Also, depending on the nature of the damage, the
frictional resistance of the damaged section may be
Lesson 3
3-10
Duct Systems
-------
increased. Observations of physical damage should be
included in the visual inspection of a duct system.
Slides
Balancing duet systems
As indicated in the introduction, balancing a
branched duct system is the role of the designer and is 3-16
done to insure that the correct air volumes are drawn
from all hoods that are connected to a common system.
As inspectors, it is important to have some
understanding of how this is accomplished and the
limitations of the various techniques.
The fundamental rule in balancing a duct system is 3-17
that all ducts entering a junction must have equal
static pressure requirements. Consider the following, 3-18
two-hood duct system:
At the junction, the static pressure in the longer
duct that leads to Source 1 is 2.0 in. HzO, while the
static pressure in the duct leading to Source 2 is 1.5
Sp^ ¦ -1.0 Inch HgO
Sp - -2.0 Inches HgO
t
Oi
- Sp - -1.5 lnch«s H2O
t
ItW1 Itl
- Sph - -1.0 Inch H2O
t
o2
Hood 2
Figure 3-3. Example Duct System
© Pendragon inc. 1992
in. H20. Since it is not physically possible to
maintain two static pressures at one location, this
condition cannot prevail. If nothing is done, the
system will adjust itself by reducing the flowrate from
Source 1 and increasing the flowrate from Source 2 until
the same static pressure requirement results. However,
the reduced flowrate at Source 1 may not be sufficient
to prevent loss of contaminants. To prevent this
situation, the designer must adjust the resistances
Lesson 3
3-11
Duct Systems
-------
of each branch so that they equal each other at the Slides
junction, while maintaining design flowrates.
Static pressure balance at a junction can be 3-19
achieved through the re-design of one or both branches
or by inserting a damper into the duct with the lower
static pressure to raise its value up to that of the
other duct. In the re-design option, changes are made
in duct diameter, duct length, elbow radius, etc., or in
the air volume in order to achieve matching static
pressures.
Each approach to duct balancing has advantages and
disadvantages. Some of the characteristics of systems
balanced through design include: 3-20
1. Since the system resistance is fixed, air
volumes cannot be easily changed.
2. The system has limited flexibility for future
equipment changes or additions.
3. Because nothing protrudes into the gas stream,
there should be no unusual erosion or
accumulation problems.
4. Since balance may be achieved by making small,
but acceptable, changes in hood volume, the
total volume from a multiple source system may
be quite different than the original design.
5. The system information must be detailed, so
that branch losses may be determined
accurately.
6. The system must be installed exactly as it is
designed. Any deviation will change the
resistance of a branch and the flowrates in the
system.
Because of the need for flexibility in design and 3-21
installation, most industrial ventilation systems are
balanced by the use of dampers. The most popular
damper is the blast gate, which is simply a flat metal
blade inserted into the duct a sufficient distance to
produced the desired loss. Positioning of the blades
can be determined through trial-and-error measurements
on the installed system or by calculation of the amount
of resistance that must be added. Some of the
characteristics of systems balanced with dampers
include:
Lesson 3
3-12
Duct Systems
-------
1. Over a limited range, the air volumes may be
easily changed by changing the positions of the
dampers.
2. The system is flexible for future changes or
additions.
3. The damper may cause material accumulation in
the duct or may be eroded.
4. Since no volume changes are required for
balancing, the total volume will be the same as
the original design.
5. If balance is to be achieved by trial-and-error
positioning of the dampers, branches that
obviously have lower resistance need not be
calculated. However, determining the branch
that has the greatest resistance is critical.
6. Moderate variation from the design during
system installation is acceptable, particularly
if the system is to be balanced by trial-and-
error .
References
ACGIH, Industrial Ventilation, Twentieth Edition, Cincinnati, 1988.
Crowder, J.W., and K.J. Loudermilk, "Balancing of industrial
ventilation systems", JAPCA, 32, 115 (1982).
Kemner, W., R. Gerstle and Y. Shah, "Performance Evaluation Guide
for Large Flow Ventilation Systems", EPA-340/1-84-012, May 1984.
SMACNA, HVAC Duet Systen Design, First Edition, Vienna (Virginia),
1977.
Lesson 3
3-13
Duct Systems
-------
LESSON 4
GAS COOLING SYSTEMS
Many processes generate gas streams at temperatures that are
too high for some air pollution control devices to accept. Because
of this, it is necessary to employ some type of cooling device to
reduce gas temperature. Since the gases must pass through the
cooling device, it is an integral part of the ventilation system.
This equipment will add to the resistance of the system and may
change the volume and composition of the gases. If it is not
functioning properly, it can affect the performance of other
components in the system. Thus, the inspector should be familiar
with these devices and be able to evaluate their performance in the
field.
In this lesson methods of cooling industrial exhaust streams
will be described, and the fundamentals governing their performance
will be presented. Parameters that affect this performance will be
discussed, along with procedures for conducting field inspections.
The most commonly used methods for cooling gases in industrial
ventilation systems are: (1) dilution with ambient air, (2)
quenching with water, and (3) natural convention and radiation from
ductwork. In a limited number of cases, forced convection systems
using air or water for cooling, may be encountered. The following
discussion will focus on the three more common cooling methods.
I. DILUTION WITH AMBIENT AIR
Cooling gases by dilution with ambient air is the Slides
simplest method that can be employed. With this
technique, hot gases from a process are cooled by adding 4-3
ambient air in sufficient quantity to produce a mixture
with the desired temperature. The fundamental
relationship governing performance may be developed
through a heat balance on the dilution system:
n^h, + m2h2 = mjhj (Eqn. 1)
where
m1 « mass flowrate of hot gases
h, = enthalpy of hot gases
1&2 = mass flowrate of dilution air
h2 = enthalpy of dilution air
m3 = mass flowrate of gas mixture
(m, + mj)
h3 = enthalpy of gas mixture
Lesson 4
4-1
Gas Cooling Systems
-------
Recalling from Lesson 1 that enthalpy is a function of
the temperature of the gas, we see that the quantity and
temperature of the hot gases and the desired temperature
of the mixture are the parameters that control "the
design. The success of the design will depend on the
quantity of dilution air supplied in relation to these
parameters.
Dilution cooling is used extensively when hot gases
discharged from a process are collected in an exterior
hood, such as a canopy. One such system is illustrated
in Figure 4-1. In this case, the amount of air volume
needed to insure capture and removal of the plume is
generally sufficient to cool the gas stream to desirable
temperatures. The technique may
Figure 4-1. Canopy Hood System
C Pendragon Inc. 1992
also be appropriate when enclosure_°5 5eceivJ"9 ho©ds
are used, if the hot gas stream is small. Here, the
dilution air could be introduced through a branch into
the hot duct or by a combination of capture and
introduced air. However, when the hot gas stream is
large, the quantity of dilution air becomes excessively
large! making the cost of the downstream exhaust system
and control device uneconomical*
One problem with dilution cooling has already been
noted, and that is that large air volumes nay be
required to accomplish the desired temperature
Lesson 4
4-2
Gas Cooling Systems
-------
reduction. Problems with temperature control may also Slides
be experienced if the quantity of dilution air is not
regulated. This would only be possible if the dilu-
tion air were introduced through a branch duct where a
damper could be used to modulate the flow. Here, a
feed-back control system could be used to maintain some
pre-set temperature. When the temperature is not
controlled, corrosion problems are possible if acid or 4-5
moisture dew points are reached.
The inspection of a dilution cooling system is
rather straight-forward. The system should first be
visually inspected to evaluate the integrity of the
system and to determine if there are any indications of
corrosion problems. Next, if the temperature of the
mixed gas stream is being monitored, this value should
be noted and evaluated in terms of compatibility with
downstream equipment, particularly the control device,
and with regard to moisture or acid condensation
potential. If controlled introduction of the dilution
air is employed, the condition and operation of the
controller and damper should be evaluated and the set-
point temperature compared with any monitored values.
If temperature is not monitored, measurements in the
mixed gas stream may be necessary to complete the
evaluation.
Measurements on a mixed gas stream that employs a
branch duct to introduce the cooling air could also be
used to estimate the volume of air coming from the
process and the volume used for dilution. To accomp-
lish this, it would be necessary to measure the
volumetric flowrate and temperature of the mixed stream
and the temperature of the hot gas stream and the
dilution air (usually ambient air). Next, the
volumetric flowrate of the mixture would be converted to
a mass flowrate using the density corresponding to the
measured temperature, and the enthalpies of the three
streams would be determined based on their temperature.
These values would then be substituted into the heat
balance shown previously to determine the mass flowrate
of the hot gas and the dilution air, as follows:
"hoAot + - »h«t) dilution - <4~2)
or
®hot = — ^dilution) / (khot "* ^dilution) (4—3)
and
®dilution * ®»ix ®hot ^
Lesson 4
4-3
Gas Cooling Systems
-------
The mass flowrates would then be converted to Slides
volumetric flowrates using the densities corresponding
to their respective temperatures. ^ Although this
procedure is somewhat involved,^ it is considerably
simpler and less time consuming than measuring
volumetric flowrate in all three streams, if that
information is needed. The method^ for measuring
volumetric flowrate will be discussed in Lesson 6.
II. QUENCHING WITH WATER
When the volume of hot gases is large and the
amount of air needed to capture them is small, some
cooling method other than dilution with ambient air is
needed. Since evaporation of water requires a large
amount of heat, the gas under these circumstances can be
effectively cooled by simply spraying water into the hot
gas stream. The fundamental relationship governing 4_6
performance may again be developed through a heat
balance on the evaporation system:
mgas (kgas jn — ^gas out) ~ ®nater ^nater vapor ^water) (4 5)
Again we see that the quantity and temperature of the
hot gases and the desired outlet temperature are the
parameters that control the design. The success of the
design will depend on the quantity of water supplied in
relation to these parameters and the efficiency of its
evaporation.
The design of a quench chamber will, in general,
depend on how critical it is that all of the water be
evaporated. In systems where a wet scrubber is
employed, it is only important that the temperature of
the gas stream be reduced below the vaporization
temperature of the scrubbing liquid. Since liquid
carry-over is not a concern, the efficiency of the
evaporation equipment need not be particularly high.
Indeed, the evaporative cooler may be little more than
a series of spray nozzles mounted in the duct leading to
the scrubber.
In systems where water carry-over is a concern, a 4-7
separate piece of equipment is likely to be used for
evaporative cooling. Here, the gas stream velocity will
be reduced to about 500-700 feet/minute, and the liquid
introduced through a series of spray nozzles at
pressures ranging from 50 to 150 psig in order to
a^sson 4
4-4
Gas Cooling Systems
-------
produce a fine spray. Where any water carry-over is Slides
undesirable, as in a fabric filter system, even lower
velocities will be used and liquid pressures as high as
400 psig may be employed. In some cases, air-atomized
nozzles may be used to reduce the droplet size produced
and enhance evaporation. These techniques will likely
be accompanied by an elaborate control system to assure
that all of the spray water is evaporated before exiting
the cooler. This is sometimes referred to as "dry
bottom" operation.
In addition to water carry-over, another problem
associated with evaporative coolers is corrosion. Since
the coolers utilize water sprays, the potential for
moisture and condensed acid corrosion is significant,
particularly in systems where all of the water is not
evaporated. The designer usually seeks to mitigate
these problems through the use of appropriate corrosion
resistant materials and linings.
Temperature control may also be a problem with
evaporative coolers. This can occur if the rate of
water addition is not controlled or if the control
system is unable to respond to the rate of change of the
gas stream temperature. Temperature control can also be
a problem if the water atomization is not efficient.
This can occur as a result of changes in nozzle
performance because of erosion or plugging.
Inspection of evaporative coolers begins by first
conducting a visual inspection to evaluate system
integrity and indications of corrosion. If the outlet
temperature is being monitored, this value should be
noted and evaluated in terms of compatibility with
downstream equipment. If temperature is not monitor-
ed, it may be necessary to measure it; however, this
measurement may be complicated by the presence of water
droplets in the gas stream. Techniques for dealing with
this situation will be discussed in Lesson 6.
Next, some determination of the quantity of cooling
water used should be made. The most desirable way of
doing this would be to read the value indicated by a
flowmeter mounted on the delivery line or in the control
room. Since it is likely that such a meter will not be
available or may not be functioning properly if it is
available, another method to evaluate water flowrate is
needed. One technique would be to observe the delivery
pressure on a gauge mounted at the cooler or on the
pump. The quantity of water delivered varies with the
square-root of the pressure. However, to be certain
that any observed changes in pressure are due only to a
Lesson 4
4-5
Gas Cooling Systems
-------
change in water flowrate, the condition of the nozzles Slides
must also be evaluated. If the nozzles are plugged, the
pressure gauge will indicate an increase in pressure.
If this condition were not known, the increase in
pressure night be interpreted as an increase in water
flow, when the flowrate has probably decreased.
Similarly, eroded nozzles would cause a decrease in
pressure, which might be interpreted as resulting from
decreased water flow.
Plugged nozzles can be determined by observing the
spray pattern during operation. If a viewport is not
available, this will require plant personnel to remove
an inspection plate, if one is installed. Eroded
nozzles may also exhibit a different spray pattern, so
a physical inspection of the nozzle may be needed to
distinguish between the two problems. Since this should
also be done by observing from outside the cooler, the
results may not be very rewarding. To improve your
chances of observing a damaged nozzle, you should ask
plant personnel to bring you a new one from their spare-
parts inventory for comparison.
Finally, if the water used in the evaporative
cooler is recycled, its quality should be evaluated.
Water containing large particles would be of concern
because of the potential for plugging or eroding the
nozzles. If the water contains small particles, there
would be concerns about passing these more difficult to
collect particles on to the control device after the
water has been evaporated. The quality of the water can
be evaluated by having plant personnel draw a sample
into a clear plastic container that you provide. The
sample should be well mixed by shaking and then allowed
to settle. If the rate of settling is fairly rapid,
then the water contains large particles. If the
settling rate is very slow, the water contains fine
particles.
III. NATURAL CONVECTION AND RADIATION
When a hot gas stream flows through a duct, the
duct becomes hot and heats the surrounding air. As the
air near the duct becomes heated, convection currents
develop that carry the heat away. This phenomenon is
referred to as natural convection and may be aided by a
small amount of forced convection due to wind motion.
Heat may also be transferred from the duct surface by
direct radiation. This behavior can be exploited to
produce significant amounts of cooling by providing a
section of duct that has large surface area. This is
typically done by arranging the duct in a series of
Lesson 4
4-6
Gas Cooling Systems
-------
vertical columns, as shown in Figure 4-2. Since the Slides
velocities through the columns are usually below the
necessary transport velocity for particles, the base of
alternate columns are joined across a hopper for
collection and removal of any settled dust.
The fundamental relationship governing the
performance of the cooler may again be developed through 4-8
a heat balance:
mMt(hMcin - lij = UADT (3-6)
gas * gas in gas out7 m x '
where U = overall heat transfer coefficient
A = heat transfer area
DT = log-mean temperature difference
4-9
In the English system, the overall heat transfer
coefficient, U, has units of Btu/hr-f t2-°F, and
represents the ability to transfer heat. It is a
function of a number of parameters, including the duct
diameter, the nature of the duct surfaces, the thermal
conductivity of the metal, the velocity of the hot
gases, the wind speed and the temperature difference
between the duct and the ambient air. The log-mean
temperature difference, DTm, is simply a means of
calculating an average difference when that difference
varies along the length of the duct. Once again we see
4-10
Lesson 4
4-7
Gas Cooling Systems
-------
that the quantity and temperature of the hot gases and
the desired outlet temperature are the parameters that
control the design. The success of the design will
depend on all the parameters that affect the heat
transfer coefficient and on the total surface area
provided.
There are a number of problems associated with
convection and radiation cooling systems. First,
because of the need to provide large amounts of surface
area for heat transfer, the size of the cooler is
usually quite large, requiring significant amounts of
plant area for its installation. Also, because the
velocities in the cooler are generally below the
transport velocity for particles, the system must be
cleaned continually to avoid build-up in the hopper
sections. Finally, because there is essentially no
control on the cooling process, it is not possible to
control the outlet temperature. Depending on the
temperature of the gases coming from the process, this
could result in outlet temperatures that exceed the
limitations of downstream equipment or that decrease
into the range of moisture or acid condensation. This
latter situation could lead to corrosion of the cooler
surfaces, resulting in infiltration of outside air or
the escape of fugitive emissions.
Because of the nature of cooler design and
operation, the items that can be inspected on these
systems is limited. As with the other devices, the
convection and radiation cooler should first be visually
inspected to evaluate the integrity of the system and to
determine if there are any indications of corrosion
problems. Particular attention should be paid to the
dust removal doors, which should be checked for leakage.
Next, evaluate the dust removal operation by having
plant personnel open a few of the doors. if the
temperature of the outlet gas stream is being monitored,
this value should be noted and evaluated in terms of
compatibility with downstream equipment and with regard
to moisture or acid condensation potential. If
temperature is not monitored, measurements in the outlet
gas stream may be necessary to complete the evaluation.
References
Danielson, J.A., ed., "Air pollution Engineering Manual". Second
Edition, EPA AP-40, May 1973. a
Segal, R., and J. Richards, "Inspection Techniques for Evaluation
of Air Pollution Control Equipment", Volume II, EPA-340/1-85-022b
September 1985. '
Lesson 4 4~8
Gas Cooling Systems
-------
LESSON 5
FAN SYSTEMS
One of the most critical parts of an industrial ventilation
system is the air mover or fan. Its function is to cause the
desired amount of air to move through the system by overcoming
resistances in the hoods, ducts, coolers, collection devices,
stacks and any other equipment present. The fan is also one of the
most complex pieces of equipment in the ventilation system. Its
performance depends on the type of fan employed, the parameters of
its operation, the characteristics of the system it is used in, and
the properties of the gas stream it operates on. To be able to
effectively inspect fan systems, the inspector must be familiar
with how these various factors influence fan performance.
In this lesson the types of fans usually employed in
industrial ventilation systems will be described, and the design
and operating characteristics that affect their performance will be
discussed. Next, the way a fan interacts with the rest of the
ventilation system to determine the air volume that is moved will
be described, and the manner by which changes in either the fan or
the system characteristics interact to change the flowrate will be
discussed. Also, installation conditions that affect fan
performance will be described and techniques to evaluate their
effect will be presented. Finally, the procedures used by
designers in selecting a fan will be described, and techniques that
can be used to evaluate the performance of an existing fan,
involving variations on these procedures, will be presented.
I. TYPES OF FANS
A fan can be generally characterized on the basis Slides
of its location in the ventilation system with respect
to the control device, as shown in Figure 5-1. Fans 5-1
located before the control device are referred to as
forced draft because they force or push the air through
the collector. In this location, the fan acts on the 5-2
dirty gas stream and may be subject to increased wear
and require a higher level of maintenance, thereby
increasing operating costs. The control device,
however, will only have to withstand the pressure
required to push the gas stream through the device and
on to the stack. As a result, the collector will
require less structural reinforcing and will likely be
somewhat cheaper.
Lesson 5 5-1 Fans Systems
-------
Figure 5-1. Fan Locations
Slides
© Pendragon Inc. 1992
Fans located after the control device are referred
to as induced draft because they induce or pull the flow
through the collector. In this location, the fan acts
on the cleaned gas stream and would be less subject to
wear. This would likely require a lower level of
maintenance, reducing operating costs. However, the
control device will now have to be structurally
reinforced to withstand essentially the entire negative
pressure of the system and will probably be more
expensive.
Fans designs can be classified as either axial or
centrifugal. A special class of fan that employs a 5-3
centrifugal wheel mounted in an axial arrangement will
sometimes be encountered in industrial ventilation
systems. As will be discussed later, its performance is
determined by the wheel design, rather than the
orientation, and it generally behaves like other
centrifugal fans.
Axial fans are used to move large volumes of air
against low resistances. They may be used for general
ventilation or in low resistance industrial ventilation
systems; however, "they are not used very
Lesson 5
5-2
Fans Systems
-------
often in air pollution control systems. Occasionally, an
axial fan will be used in combination with the more
common centrifugal fan to provide extra energy to
overcome resistances.
There are three basic types of axial fans: (1)
propeller, (2) tubeaxial and (3) vane axial. The
propeller fan is used for moving air against very low
static pressures, such as would be encountered in
general room ventilation. Their performance is very
sensitive to resistance and a small increase will cause
a significant reduction in flow.
Tubeaxial and vaneaxial fans are shown in Figure 5-
2. The tubeaxial or ducted fan is essentially a
propeller-type fan mounted in a cylindrical housing and
is capable of moving air against pressures less than
about 2 in. H20. Since the motor is typically mounted
inside the housing, it is not generally used on
contaminant-containing gas streams.
Slides
5-4
VMKMMi
Figure 5-2. Tubeaxial and Vaneaxial Fans
(ACGIH, 1988)
Lesson 5
5-3
Fans Systems
-------
The vaneaxial fan is similar in construction but
uses airfoil-style blades and straightening vanes on the
inlet and outlet to improve efficiency, locates the
motor on the outside of the casing and provides an
enclosure to protect the drive system. It is capable of
developing pressures up to about 8 in. HjO.
The principal fan used in air pollution control
systems is the centrifugal fan. The basic design of the
centrifugal fan, as illustrated in Figure 5-3, employs
a fan wheel or impeller mounted inside a scroll-type
housing. Air is draw into the inside of the impeller
and then forced out through the housing. In general,
centrifugal fans are distinguished by the design of the
impeller. There are three basic impeller types: (1)
forward curved, (2) radial and (3) backward inclined.
The backward inclined impeller may use a standard
single-thickness blade or an airfoil blade.
Slides
5-5
Scroll Sldo
Scroll Piaca
Side Sheet
Sid* Plata
Support
Stlffenera
Scroll
Bind
Scnll Shoot
Wrapper
Wheel Cons
Retaining Ring
Wrap Shoot
Scroll Back
Inlot Collar
Inlet Slaava
Inlet Band
Inlet Rim
Wheel Rbn
Flange—
Inlet Plata
Figure 5-3. Centrifugal Fan Components
(ACGIH. 1988)
Lesson 5
5-4
Fans Systems
-------
Forward curved impellers, shown in Figure 5-4, have Slides
blades that curve into the direction of rotation.
Commonly referred to as "squirrel-cage blowers", they 5-7
are constructed of lightweight materials and usually
have 24 to 64 closely-spaced blades. For a given duty,
these impellers are the smallest of all the centrifugal
types and operate at the lowest speeds. As a result,
they are quiet in operation but are only able to develop
low to moderate static pressures. Because of this they
are not commonly used in air pollution control systems.
As shown in the accompanying performance chart, the 5-8
highest mechanical efficiency (ME) is developed at a
point to the right of peak static pressure (SP), at
about 60 percent of the wide open volume, and the
horsepower requirement (HP) rises continually toward the
free delivery volume. Since particles may adhere to the
closely-spaced blades and cause a balance problem, their
use should be limited to clean gas streams.
Fan Whsel
Fan Housing
Forward Curved
Fan Blades
t"
2
9
1
ss
Ui
Gas Flow, SCF/min.
Figure 5-4. Forward Curved Wheel Design
and Performance (Modifed Drawing Based
on AC6IH. 1988)
Lesson 5
5-5
Fans Systems
-------
Radial impellers, shown in Figure 5-5, have 6 to 10 Slides
blades that extend either straight out from the hub (R)
or in a radial direction (M). They are the simplest of 5-9
all the centrifugal fans and the least efficient, but
they are capable of developing the highest static
pressures. For a given duty, they operate at moderate
speeds. Highest mechanical efficiency is developed just
to the left of peak static pressure, at about 30-40
percent of the wide open volume, and the horsepower 5-10
requirement again rises continually toward the free
delivery volume. The radial blade shape is generally
resistant to material build-up and may be used in
systems handling either clean or dirty air. There are
a variety of impeller designs, ranging from "high
efficiency, minimum material" to "heavy impact
resistant". Impellers in this latter category usually
have no inlet plate or backplate in order to minimize
locations of potential material build-up.
Fan Wheal
Fan Housing
Radial
Fan Bladas
£¦"" '
f"*
8.
§ £
£ -¦
w
O ..
Gas Row, SCF/min.
Figure 5-5. Radial Wheel Design and Performance
(Modified Drawing Based on ACGIH. 1988)
Lesson 5
5—6
Fans Systems
-------
The backward inclined impellers have 9 to 16 Slides
hollow airfoil-style blades that incline or curve away
from the direction of rotation. They have the highest 5-11
efficiency of all the centrifugal fans and, for a given
duty, will operate at the highest speed. Highest
mechanical efficiency is developed to the right of peak
static pressure, at about 50-60 percent of the wide open
volume. A unique characteristic of the backward 5-12
inclined impeller is that the horsepower requirement
reaches a maximum value near the point of peak
efficiency and then declines toward the free delivery
volume. For this reason backward inclined fans are
sometimes referred to as "non-overloading", since any
variation from the optimum operating point due to a
change in system resistance will result in a reduction
in operating horsepower. One negative feature of the
airfoil design is the use of hollow blades. These
blades can erode and accumulate material inside the
blade, causing a balance problem. They should,
therefore, be limited to clean air applications.
The standard backward inclined impellers, shown in
Figure 5-6, are identical in design to the airfoil-type,
except they employ single-thickness blades. Peak
efficiency is slightly less than the airfoil design but
is still developed to the right of peak static pressure
and at about 50-60 percent of the wide open volume. The
horsepower requirement again exhibits the non-
overloading characteristic. Because the blades are
single thickness, they can be used in gas streams with
light dust loadings. However, they should not be used
in heavy loading situations that could cause build-up on
the blade surfaces.
Fan Arrangements
Fans are constructed with different bearing
locations and motor mounting capabilities, generally
referred to as "arrangements". Knowledge of the various
arrangements is desirable when it is useful to speak the
language of fan systems.
There are ten basic fan arrangements (see AMCA,
1979) and three of the most common of these are shown in
Figure 5-7. Here, SW or DW refers to single- or double-
width fans, respectively. As the name implies, double-
width fans have an impeller that is twice as wide as the
corresponding single-width version and a capacity that
is also approximately double. Also shown are SI and DI,
which refer to single- or double-inlet, respectively.
Lesson 5
5-7
Fans Systems
-------
Slides
Fan Wheel
Fan Housing
Backward Curved
Fan Blades
-t"
Q.
X
kT
©
s
::t
>»
-•g
O)
52
o
•X
w
O
O
2
©
2
s
Figure 5-6. Backward inclined standard Wheel
Design and performance
© Pendragon Inc. 1992
The more common single-inlet fans have the air
inlet on only one side, °PPoS^e drive. Double-
inlet fans have air inlets on both sides.
Arrangement 1 has two st»af^ gearings mounted on the
pedestal with the impeller overhung on the end of the
shaft. For a V-belt drive, the motor would be mounted
on a separate base adjacent to the pedestal and at
ground level, with pulleys or sneaves on both the motor
shaft and the fan shaft. Arrangement 3 also has two
shaft bearings, but they are loc*ted on either side of
the impeller and supported W the fan housing. The
drive arrangement would be the same as Arrangement 1.
Arrangement 9 is the same as Arrangement 1, but has the
motor mounted on the outside ot the fan base with a V-
belt drive system. Most of tne other fan arrangements
not shown in Figure 5-7, are asicaliy variations on
these three.
5-13
— — - -
Lesson 5 Fans Systems
-------
Figure 5-7. Fan Arrangements
(Modified Drawing Based on AMCA, 1979)
Slides
O Pandragon Inc. 1992
II. 7AM LAWff -
The fan'laws relate the performance variables for 5-14
any homologous series of fans. A homologous series is
simply a range of fan sizes where all of the dimen-
sional parameters are proportional. At the same
relative point of operation on any two performance
curves in a homologous series, the mechanical
efficiencies will be equal. Under these conditions, the
following relationships apply:
Lesson 5 5-9 Fans Systems
-------
q2 « Q^size^size^'Crpn^/rpm,) "J Slides
P2 - P1(size2/size1)2(rpn2/rp»1)2(p2yPi) (Egn.l)
bhp2 = bhp1 (size2/size1)5 (rpfflg/rp»i)3 (Pz/Py)
The performance variables inVO}ve^ ^i°^Sure
(Q), fan size (size), rotational spee? 'nP the
(P), gas density (p) and horsepower (bhp) • Here,
pressure may be represented by total pressure, sta
pressure, velocity pressure, fan 1 ned
static pressure. These latter two terms will be defa-
later in this lesson.
As indicated, these expressions rely on the
performance curves being homologous and apply
the same relative point of rating. Under tiurbulent flow
conditions, which occur in most air pollution control
systems, two performance conditions will be at the same
relative point of rating if the pressures and flowrates
at these two conditions are related by:
P2 - PJQz/Qt)2 2>
As will be seen in the next section* this is the same as
saying that the two performance points must lie along
the same system curve.
In actual practice, the fan laws are typically used
to determine the effect of changing only one variable at
a time and are most often applied to a single fan size.
The most common variable of interest is fan speed. For
determining the effect of changing fan speed while
operating on the same gas stream (ri " r2), the fan laws
become:
q2 « Q, (rpn^/rpm,)
p2 - P^rpmj/rpm,)2
(Eqn. 3)
bhp2 - bhp, (rpmj/rpm,) 3
Referring to the original equations, it is
interesting to note that if only c»f»9es in gas density
are involved, pressure capabilities and power
requirements change proportionally* while flowrate i8
unaffected, *»®havior is sometimes characterized bv
stating that -« j.~» *¦*.** — .
it moves volumes of air not masses or air.
ments change proportionaxa-j» wane nowrat® i«
ted. This behavior is sometimes characterized bv
that "a fan is a constant volume machine*1 i a
« vniiiwM of air not masses of air. *'
Lesson 5
5-10
Fans Systems
-------
III. FAN PERFORMANCE
Slides
Performance graphs for centrifugal fans were
presented in Figures 5-4 through 5-6. One of the
relationships shown on these figures was that between
pressure developed and air volume moved, labeled SP.
This relationship is sometimes referred to as the fan
curve or fan characteristic. For a particular fan
turning at a given rpm, there is one and only one fan
curve. It represents all of the pressure/air volume
combinations that that fan is capable of when opera-
ting at that one rpm. These range from low air flow
delivered against high pressure (upper left) to high air
flow delivered against low pressure (lower right) . What
determines which condition a fan will operate at is how
this curve interacts with the ventilation system
characteristics, as represented by the system curve.
Normalized duct system curves for three systems are 5-15
shown in Figure 5-8. Plotted here is the percentage of
duct system resistance as a function of the percent of
duct system flowrate. This is simply a normalized
version of a P verses Q plot. The system curves shown
follow the general relationship characteristic of
turbulent systems:
P, (Qj/Q,)2
(Egn. 4)
Design Resistance
/ Lower Resistance
Calculated Qas Flow Rate,
ArrangementA
20 40 60 80 100 120 140 160 180 200 220
Percent of Calculated System Qas Flo* Rate (SCFM) »
Figure 5-8. Normalized Duct Curves
O Pendragon Inc. 1992
Lesson 5
5-11
Fans Systems
-------
In practice, the system curve is developed by first
determining the resistance or static pressure for one
flowrate through the system, using the techniques
discussed in Lessons 2 and 3. Other points on the
curve are then determined using the above relationship.
Thus, if the design point for System A were at 100
percent volume and 100 percent resistance, increasing
the flowrate to 120 percent of the design flow would
increase the resistance to 144 percent of the design
resistance. Likewise, decreasing the flowrate to 50
percent of the design value would decrease the
resistance to 25 percent of the design resistance. Note
that, on a percentage basis, the same relationships also
hold for Systems B and C.
The point of intersection of the system curve with
the fan curve determines the actual fan performance.
This is shown in Figure 5-9, where a normalized fan
curve has been plotted with the system curves from the
previous figure. Here, the 100 percent design volume of
System A has been arbitrarily selected to intersect at
Point 1 with the 60 percent free delivery volume of the
fan. Unless actions are taken to change either the fan
curve or the system curve, the performance delivered
will be that indicated by the intersection point.
Slides
5-15
5-16
Parcant ot Fan WidaOpan Gas Flow Rato(SCFM)
15 30 45 60 75 90 105 120 135 150 165
20 40 60 00 100 120 140 160 180 200 220
Percent of Calcutatad Sysrnn Ga* Flow Rait (SCFM) ~
Fan Characteristic
CwvaatRPM'x*
%
- 16 5
£
481
§
«•
I
Figure 5-9. interaction of System curves
and Fan Curves
O Pendragon Inc. 1992
Lesson 5
5-12
Fans Systems
-------
One way to change the flowrate would be to change Slides
the system. This could be done by closing or opening a
damper, producing a system with more or less resis-
tance and changing the system curve. For example,
referring again to Figure 5-9, the flowrate could be
decreased to 80 percent of the design volume by closing
a damper until the more resistant System B curved is
obtained, shifting the intersection to Point 2.
Likewise, the flowrate could be increased to 120 percent
of the design value by opening a damper and shifting the
intersection to Point 3.
Changes in flowrate could also be produced by
changing the fan speed, shifting the fan curve. This is 5-17
illustrated in Figure 5-10, where a new fan curve
Piout of Fan Wid» Open Qtt Row R*a»
•o
100
1
•g
80
01
O
GO
"3
c
§
40
©
Q.
20
l 1 1 1 j
RPMY
RPM X \
y
•
A3 ^ Fin Charactantac
\ / Curva al RPM "x"
•' ' ¦ l 1 1 1
x
2
&
<
150 !
100 ;
so
20 40 60 80 100 120 140 ISO lao 200 220
Parcanto* Cateutaad Sycwn Gas Flow RN (SCFM)-
Figure 5-10. Effect of Increased Fan Speed
O Pendragon Inc. 1992
representing a 10 percent increase in speed has been
added. At this new speed, the point of operation shifts
to Point 2. Since flowrate is proportional to fan
speed, this 10 percent increase in speed produces a
corresponding 10 percent increase in volume. However,
following the fan laws, this 10 percent increase in
speed will require a 33 percent increase in operating
horsepower, which may be beyond the capabilities of the
existing motor.
According to the fan laws, changing gas density
will shift the fan curve. However, since gas density 5-18
Lesson 5
5-13
Fans Systems
-------
also affects the system resistance, the system curve Slides
will also be shifted. This is illustrated in Figure 5-
11 for a density change from 0.0375 lb/ft3 to 0.075
lb/ft3. As previously indicated, the new operating
point will deliver the same air volume but at double the
resistance and double the horsepower requirement.
Fan performance can also be affected by the manner
in which the fan is installed in the system. Most
manufacturers in the United States and Canada rate the
performance of their fans from tests made in accordance
with the Air Movement and Control Association (AMCA)
Standard 210, "Test Code for Air Moving Devices". The
test set-up prescribed by this standard is designed to
produce inlet and outlet flows that are as uniform as
possible. This condition insures consistency and
reproducibility of test results and permits the fan to
develop its maximum performance, in any installation
where this uniform flow condition does not exist, fan
performance will be reduced.
Installation conditions that affect fan perfor- 5-19
mance are referred to as "system effects". The three
most common causes of deficient performance are:
(1) non-uniform inlet flow, (2) swirl at the fan inlet,
and (3) improper outlet connections. These conditions
alter the aerodynamic characteristics of the fan such
that it does not operate at its rated performance.
Duct SycMm Curva A
& 0075 Mt3 Dantity
at Fan Mat
DuctSyttam Curva A
9 0.0375 Danoty
•I Fan MM
\ Dan«i«y di 0.0375
\ RMo -ij" 0375 ¦°5
20 40 60 >0 100 120 140 160 tlO 200
Paroart ol Otitt 9r«w» Row (CfM)
Figure 5-11. Influence of Gas Density
©Pendragon Inc. 1992
Lesson 5
5-14
Fans Systems
-------
The influence of system effects on fan performance is Slides
shown in Figure 5-12. Here, the solid system curve has
been determined without consideration of system effects
and performance corresponding to Point 1 is anticipated.
However, because of system effects the effective fan
curve is the dashed line and performance corresponding
to Point 3 is obtained.
Figure 5-12. System Effects for Outlet Ducts
(Modified Drawing Based on ACGIH, 1988)
O Pendragon inc. 1992
Lesson 5
5-15
Fans Systems
-------
This deficient performance could be prevented by
calculating the system effect loss, adding it to the
system resistance and selecting a fan to operate at
Point 2.
Slides
System effect loss factors for some common
centrifugal fan outlet duct installation conditions are
given in Figures 5-13 and 5-14. These factors are in
numbers of velocity pressures lost, so the addition to
the system static pressure would be equal to the loss
factor times the appropriate velocity pressure. Loss
factors for other situations can be found in ACGIH's
Industrial Ventilation or in AMCA's "Fan Application
Manual", Part 1.
SO 75
Pweantol Efacto* Dud Lwigtti, L
Figure 5-13. System Effect Loss Factors
O Pendragon Inc. 1992
Lesson 5
5-16
Fans Systems
-------
Slides
Pareant of Effective Duet Langlh to Bbow
Figure 5-14. System Effect Factors for
Outlet Duct Conditions Shown in Figure 5-13
© Pendragon Inc. 1992
System effect losses in Figures 5-13 and 5-14 are
expressed in terms of the percentage of effective duct
that is present at the fan outlet and the blast area-
outlet area ratio. Effective duct length is one
diameter for each 1000 fpm duct velocity, with a mini-
mum of 2.5 diameters. The ratio of blast area to outlet
area can be determined from manufacturer's literature or
estimated as 0.7.
IV. FAN SELECTION
Selecting a fan is usually the responsibility of
the ventilation system designer. However, since the
inspector may wish to apply variations of the
Lesson 5
5-17
Fans Systems
-------
selection technique in evaluating the performance of an Slides
existing fan, it is important that the methods used by
the designer be understood.
Fan selection is typically done using ratings 5-20
tables published by manufacturers for their products.
An example of one of these tables is shown in 5-21
Figure 5-18. In general, the ratings table is entered
along the row corresponding to the design volume and
down the column corresponding to the design static
pressure, including system effects. The point of
intersection indicates the rpm that the fan would have
to turn to deliver the required performance and the
horsepower that would be needed to drive it. Shaded
regions are usually included on the chart to indicate
areas of good mechanical efficiency.
Ratings tables indicate the performance of a fan
when operating on air having a density of 0.075 lb/ft3.
Since a given system may be handling air of a different
density, some adjustments are involved before entering
the table. Remember, density does not affect fan
volume, but it does influence static pressure conditions
and horsepower requirements. The specific procedure
involved in fan selection is as follows:
1. Determine the design air volume at actual
conditions.
2. Calculate the fan static pressure at actual
conditions, including system effects. Fan
static pressure is defined as:
FSP = SP^ - SPin - VPin (Eqn. 5)
In calculating fan static pressure, the sign of
the static pressure is important and must be
included. Some manufacturers rate their fans
on fan total pressure. Fan total pressure is
defined as:
FTP = TP^ - TPin (Eqn. 6)
The sign of the total pressure is again
important and must be included.
3. Correct the fan static pressure to an
equivalent value for standard air:
FSPjqyjvslent * ^^actuil (Eqn. 7)
Lesson 5
5-18
Fans Systems
-------
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Figure 5-15. Example Fan Ratings Table
Slides
4. Enter the ratings table at the actual volume
and the equivalent fan static pressure.
Determine the rpm and horsepower requirements.
5. Correct the horsepower requirement to the
conditions of actual operation:
bhP.ctu.l = bhPequiv.l«,t^.ctu.l/°-075) 8>
Since density varies inversely with temperature,
corrections for operating conditions could also be made
using a ratio of absolute temperatures. Also, because
the exact input parameters may not be contained in the
ratings table, linear interpolation between the nearest
values may be required.
In selecting the appropriate motor for a fan, the
designer must give consideration to the possible air
densities the fan may have to handle. For example, a
system operating at elevated temperature would require
the horsepower as determined above. However, when
starting the system it may be necessary to handle colder
higher-density air, requiring more horsepower.
Lesson 5
5-19
Fans Systems
-------
Should the system be located outdoors in an area that Slides
has extreme low temperatures, the horsepower requirement
for start-up could be considerable.
An alternate way of dealing with this situation
would be to install the horsepower required for normal
operation and then use an inlet or outlet damper,
together with an amperage control system, during start-
up. When the fan is started on cold air, the amperage
control system would sense a high current flow and close
the damper to prevent or reduce air flow into the fan.
The fan turning through the restricted air flow would
heat it up, reducing its density and reducing the
current draw. The amperage control system would sense
this and open the damper a little bit to allow some air
flow from the hot process. This scenario would continue
until the damper was fully open and the system was
operating at design conditions.
V. EVALUATION OF FAN PERFORMANCE
The fan laws and the techniques involved in fan
selection may be used by the inspector to estimate the
air volume the fan is delivering. For example, if the
air volume delivered by an existing fan were known, but
a subsequent inspection determined that the fan speed
had been changed, the new air volume could be estimated
from:
Q2 = Qi(rpmj/rpm,) (Eqn. 9)
Initial estimates of air volume could be made using
measurements of the fan operating parameters, together
with the appropriate ratings table. This approach would
apply only to V-belt or variable-speed drive fans, for
which general ratings tables are available. Direct
drive fans are specially constructed to deliver the
required volume when turned at the speed of the motor
and do not have published performance tables. It should
be noted, however, that because of our inability to make
exact measurements of some of the parameters and because
of a lack of precision in the ratings tables, these
techniques will yield only a rough estimate of flowrate.
The most satisfactory technique for estimating fan
performance from the ratings tables would be to use
measurements of fan speed and fan static pressure. The
specific procedure is as follows:
Lesson 5
5-20
Fans Systems
-------
1. Measure or estimate fan rpm. To obtain as much Slides
accuracy as possible, measurement of the rpm is
preferred. However, if this is not possible, an
estimate can be used. Techniques for deter-
mining fan rpm will be discussed in Lesson 6.
2. Determine fan static pressure for operation on
standard air. This would involve measuring inlet
and outlet static pressures and estimating inlet
velocity pressure. Because of turbulence levels
near the inlet and outlet of a fan, it may be
difficult to get acceptable readings. To avoid
these turbulence problems, the measurements could
be made some distance away from the fan, where
acceptable readings can be obtained. Since static
pressure losses are small in the larger ducts
usually found at the fan, the error introduced by
this should be minimal. If inlet or outlet
dampers are present, the loss introduced by these
fittings would have to be estimated and included
in the determination of the respective static
pressure. Likewise, any losses due to system
effects should be included. Air density could be
estimated from a measurement of temperature, and
velocity pressure could be estimated based on the
expected velocity at the inlet. The estimated fan
static pressure would then be given by:
estimated = O.OTSfSP^ - SPin - VPjn) /PKtM{ (Eqn. 10)
3. Enter the ratings table for the fan at the column
corresponding to the estimated fan static
pressure. Proceed down this column until the row
containing the measured fan speed is located. Read
along this row to determine the estimated
flowrate. Interpolation between values in the
ratings table nay be necessary.
If the air volume determined from this procedure
gives a significantly different velocity pressure than
that assumed in Step 2, the velocity pressure should be
re-estimated and the procedure repeated until a
reasonable agreement is obtained.
A less satisfactory technique for estimating fan
performance from the ratings tables would be to use
measurements of fan static pressure and operating
Lesson 5
5-21
Fans Systems
-------
horsepower. Following the same general procedures
outlined above, the fan static pressure would be
estimated and used to enter the ratings table. The
row corresponding to the estimated operating horse-
power would then be located and used to determine the
estimated flowrate. Techniques for determining these
parameters, as well as the other parameters involved
in the evaluation of fan performance, will be
discussed in Lesson 6.
As a minimum, the fan inspection should include
an evaluation of the condition of the fan. This would
include a visual determination of the condition of the
fan housing to assess any indications of corrosion, an
evaluation of the inspection door seal for leakage, an
assessment of the condition of isolation sleeves used
to dampen vibration between the fan and the inlet and
outlet ducts to determine if there are any leaks, and
an evaluation of any vibration or belt squeal. Belt
squeal during operation indicates that the belts are
slipping on the pulleys. This can result m the loss
of 200-300 rpm, with a corresponding loss in ait
volume. A fan that is vibrating severely represents
a significant safety hazard. If this condition should
be encountered, the inspection should be terminated
immediately and plant personnel notified of the
condition. If the fan is not operating, an inspection
of the condition of the fan wheel would also be useful
to identify any build-up or corrosion problems.
References
ACGIH, Industrial Ventilation, Twentieth Edition, Cincinnati, 1988.
AMCA, "Fan Systems", Pan Application Manual, Part 1, Publication
201, Arlington Heights (Illinois), 1979.
Kemner, W., R. Gerstle and Y. Shah, "Performance Evaluation Guide
for Large Flow Ventilation Systems'*, EPA-340/1-84-012, May 1984.
Lesson 5
5-22
Fans Systems
-------
LESSON 6
MEASUREMENT OF VENTILATION SYSTEM PARAMETERS
In previous lessons, the use of various parameters to evaluate
the performance of ventilation system components has been sugges-
ted. In this lesson, the methods available for making these
measurements will be discussed and recommendations on the most
appropriate techniques and procedures will be made. Where
appropriate, additional techniques for estimating some parameters
will be given.
I. MEASUREMENT PORTS
Slides
When a ventilation system is first inspected, it is
unlikely that measurement ports will be available. If 6-1
some ports are available, they are not likely to be in
the locations needed or of an appropriate size. The 6-2
most likely port to be found on a ventilation system is
a 3 or 4 inch diameter sampling port located on or near 6-3
the stack. Although a port in this location may be
useful for some inspection measurements, ports of this
size should, m general, be avoided. They present
difficulties in sealing under both positive and negative
conditions, and they may be quite difficult to open
because of the large thread area.
The most useful port size for inspections is l%-2 6-4
inch diameter, and this size is needed only if mea-
surements of velocity pressure are anticipated. For the
more routine measurements of temperature and static
pressure, ports of inch will accommodate most
measurement probes. The larger inspection ports will
require the installation of a pipe stub with a threaded
plug for closing. The smaller ports should simply be
drilled and then covered with duct tape when not in
use. Because of the potential for fire or explosion
from sparks and because of possible damage to downstream
equipment, the inspector should not request that ports
be installed while the equipment is running. Rather,
the locations and sizes needed should be marked for
plant personnel, so that they may install them the next
time the system is shut down.
Ports of a proper size may already be installed in
some locations and used by the plant for continuous
Lesson 6
6-1
Measurements
-------
monitoring of certain parameters. In general, these Slides
ports should be avoided by the inspector, if they must
be used, they should be opened only by plant personnel. 6-5
Never open a port that was not placed there for your
exclusive use. Plant monitoring ports may be connecte
d to controllers that initiate equipment shutdown if the
signal from them is lost.
Measurement ports are subject to the accumulation 6-6,
of material that may cause them to become plugged, even 6-7,
if they are on the clean side of the control device. and
Before using any port, it should be cleaned out with a 6-8
non-sparking rod to assure unobstructed access to the
gas stream. Also, while making measurements the port
should be sealed to prevent flow in or out around the
probe. Flow into or out of the port may cause an
interference with the measurements being made. For
inspection ports, the best sealing technique is to
insert the probe through a rubber stopper and then place
that stopper into or against the port. For the larger
stack-sampling ports, a rubber sanding disc may be used
to cover the opening. The probe, equipped with a rubber
stopper, would then be inserted through the center of
the sanding disc, using the stopper to complete the
seal.
Finally, the inspector should not make heroic
efforts to reach existing ports and should not have
ports installed in locations that cannot be reached and
used in safety. This should include consideration of
hazards to walking and climbing, as well as the
potential for exposure to inhalation, vision, hearing
and fire and burn hazards.
II. 8TATIC PRESSURE MEASUREMENTS
Static pressure measurements must be made with a 6-9
square-ended probe placed at a right-angle to the flow
direction. If measurements of velocity pressure are
also being made, the static pressure ports on a standard
commercial pitot tube that is oriented into the oncoming
flow may also be used, as could one leg of the s-type
pitot if it is turned at a right-angle to its normal
position. These two pitot tubes will be discussed
later in this lesson. The purpose of the probe
orientation is to be sure that no component of velocity
pressure is impacting the probe during static pressure
measurements.
Lesson 6
6-2
Measurements
-------
The area between the probe and the port opening Slides
should be sealed to avoid errors associated with flow
into or out of the duct. Errors resulting from improper 6-10
sealing can be as large as 10-30 percent. Flow into the
duct can result in an aspiration effect at the end of
the probe that can increase (make more negative) the
negative pressures being measured, while flow out of the
duct can add a component of velocity pressure to the
measurement of positive pressures. To further mitigate
this problem, the probe should be extended well into the
duct while making measurements.
There are two widely used technigues for sensing
the pressures measured by the probe: (1) a U-tube
manometer or (2) a Magnehelic* pressure gauge. The U- 6-11
tube manometer is a reference instrument that is
available in a flexible or slack-tube configuration,
shown in Figure 6-1, to enhance its portability. The
manometer is eguipped with magnets at the top and bottom
to facilitate temporary mounting and is eguipped with
threaded connectors that are used to seal the manometer
during transport.
Lesson 6
6-3
Measurements
-------
The manometer indicates the static pressure by Slides
displacing the fluid in the tube. In making static
pressure measurements, one leg of the manometer is
connected to the probe and the other is left open to the
atmosphere. The height difference between the levels in
the two columns is the pressure in height of fluid,
usually expressed in inches of water. One of the
principal difficulties with the U-tube manometer relates
to the fluid. If the pressure in the duct exceeds the
capacity of the manometer, fluid will either be drawn
into the duct or blown out onto the inspector. Also,
the inspector must remember to close the seals when
transporting the manometer to prevent loss of fluid and
to open them before making a measurement.
The Magnehelic* pressure gauge, is a product of 6-12
Dwyer Instruments, Inc. It senses pressure difference
by deflecting a silicone rubber diaphragm and then
translating that deflection to a needle indication
through a magnetic linkage. Although not as accurate as
the U-tube manometer, it is much more forgiving, making
it easier to use in field situations. The Magnehelic
is accurate to within 2 percent of full scale and has a
high resistance to shock and vibration. It is available
in over 70 ranges, from 0-0.25 in. H20 to 0-20 psig.
The most useful ranges for ventilation system inspection
are 0-5, 0-20 and 0-50 in. H20. For inspection of high
pressure drop wet scrubber systems, a 0-100 in. H20
range may be needed.
Except for the 0-0.25 and 0-0.50 in. H20 ranges,
the Magnehelic* may be used in any orientation and can
accept pressures up to 15 psig without being harmed.
This property allows gauges with different ranges to be
combined in one instrument package (as shown in Figure
6-2), with the gauge giving the most readable indication 6-13
used for recording the measurement.
Because of the silicone rubber diaphragm, the
ambient temperature range is limited to 20 °F to 140 °F.
This lower limitation can be accommodated when
conducting inspections in cold environments by keeping
the gauge in a location that is within the range and
then taking it out briefly for making the measurement.
For extended use under cold conditions, gauges with a
lower temperature limit of -65 °F are available on
special order.
Lesson 6
6-4
Measurements
-------
Slides
The Magnehelic* is not a reference instrument, so
its calibration should be checked periodically. The
simplest way of doing this is to check its indications
against a U-tube manometer. Using a laboratory squeeze-
bulb equipped with check valves, pressures from -4 0 to
+40 in. H20 can be easily generated. The Magnehelic*
indications should be plotted against those of the
manometer to check for accuracy and linearity. Gauges
that give inaccurate or non-linear indications should be
discarded. Also, while using the gauge its zero should
be checked frequently and adjusted as needed using the
set-screw on the front plate. Adjusting the zero will
not affect the calibration of the gauge.
III. TEMPERATURE MEASUREMENT 6-14
There are several techniques available for
measuring temperature, including: (1) mercury
thermometers, (2) dial thermometers, (3) thermistors and
(4) thermocouples. Unfortunately, each of these
techniques has some limitation when applied to the
inspection of industrial ventilation systems. The
mercury thermometer is constructed of glass and is
subject to breakage, with resulting exposure to a toxic
material. Also, both the mercury and dial
Lesson 6
6-5
Measurements
-------
thermometers have a limited probe extension, making the Slides
measurement of temperatures across large ducts
impossible. Since locations near the wall of a hot duct
will be cooler than near the center, measurements made
there may not be representative of actual temper-
atures. The thermistor, which measures temperature
through the change in resistance of a fine wire sensor,
is easy to use but its response becomes non-linear over
some part of its temperature range, making data
interpretation difficult. Finally, the potentiometer
used to measure the output of a thermocouple is not yet
available in an intrinsically-safe construction and
cannot be used in areas where explosive or ignitable
materials may be present.
Despite its limitations, the thermocouple is 6-15
recommended as the primary method for measuring
temperatures in the inspection of industrial ventila-
tion systems. in situations where explosive or
ignitable materials may be present, use of the dial
thermometer is suggested, but the inspector should be
aware of the potential problems in obtaining
representative measurements on large hot ducts. The
thermocouple sensor is formed by joining two wires made
of different metals or alloys. If the junctions at the
ends of these two wires are then held at different
temperatures, an electric current flows in the wire
loop. This current is produced by an electromotive
force whose value depends on the difference in
temperature between the junctions.
The electromotive force generated by a thermo-
couple is measured with a potentiometer, a variety of
metals and alloys are used in the construction of
thermocouples, providing for measurements over differ-
ent temperature ranges. The most common thermocouple,
and the one recommended for use in inspections, is Type
K. The Type K thermocouple has a temperature range of -
400 °F to +2,300 °F and is constructed with a positive
wire of chromel and a negative wire of alumel. Most
hand-held potentiometers are calibrated for certain
thermocouple types and internally convert the measured
electromotive force to a temperature indication.
The thermocouple/potentiometer is not a reference
instrument and must be calibrated against a National
Institute of Standards and Technology (NIST) traceable
thermocouple to assure accuracy. Since the equipment
required to do this is expensive and not likely to be
available to the inspector, it may be necessary to
Lesson 6
6-6
Measurements
-------
send the unit to a specialized laboratory for calibra- Slides
tion. For most inspection situations, however, high
accuracy is not required. In these cases, an accept-
able evaluation of instrument accuracy may be conduct-
ed by checking its response in an ice bath and a boiling
water bath. Under frequent use, this check should be
done on a weekly basis. For less frequent use, it
should be done prior to taking the instrument into the
field.
There are several potential sources of error in 6-16
making temperature measurements. One of these, use of
an unrepresentative location, has already been
mentioned. With the thermocouple this problem can be
avoided by making measurements at several locations
across the duct cross-section and averaging them. This
can be done through a formal procedure, such as that to
be discussed for making velocity pressure measurements,
or it can be performed with random locations and mental
averaging. The formal procedure will, of course, give
more accurate averages. To reach locations well within 6-17
the duct, the thermocouple wire will need to be
supported. One of the more satisfactory techniques for
doing this is to thread the wire through a small
diameter copper tube, allowing the junction to protrude
out the end.
Lesson 6
6-7
Measurements
-------
Problems can also occur from the cooling of the probe Slides
due to air infiltration through the port or through
leaks into the duct upstream of the measurement point.
The former problem can be avoided by sealing the port in
the manner described in the section on static pressure 6-18
measurements. In addition, if a copper tube is used to
support the thermocouple, it could be bent slightly so
that it extends into the oncoming gas stream. To avoid
problems from upstream leaks, the area near the
measurement location should be inspected for holes in
the duct or leaks in inspection hatches or expansion
joints. If these are found to exist, the measurement
location should be changed to an area where these leaks
will have mixed into the flow. If this is not possible,
the number of measurement points used to obtain an
average should be increased.
Measurements downstream of evaporative coolers or 6-19
wet scrubbers can be complicated due to the presence of
water droplets. As these droplets impact on the sensor,
the temperature will vary between the dry-bulb and wet-
bulb values. However, since the degree of wetting will
not be known and cannot be controlled, the exact
condition of the measurement cannot be ascertained.
Under these conditions, the most reasonable option is to
shield the sensor from the water droplets. It should be
realized, however, that doing this will likely slow the
response of the sensor, requiring longer times to make
the measurements.
IV. FLOWRATE MEASUREMENT
Measurement of gas flowrate in ducts is accomp-
lished by first measuring the average velocity pressure 6-20
and temperature of the gas stream and then calculating
the average velocity. The flowrate is obtained by
multiplying this average velocity by the duct cross-
sectional area. Procedures for conducting this
measurement are contained in 40CFR60, Appendix A, 6-21
Methods 1 through 4, as part of the procedures for
conducting compliance sampling. Since the level of
accuracy required for inspection of industrial
ventilation systems is not as high as that needed for
compliance sampling, some variances to these methods
will be employed to expand their application and
simplify the procedures and calculations, as follows:
Lesson 6
6-8
Measurements
-------
1. Method l limits the technique to ducts larger than Slides
12 inches diameter. For inspection purposes, the
procedures will be applied to ducts of all sizes.
To minimize errors, a pitot tube smaller than 5/16
inch 0.D. should be used in ducts smaller than 12
inches diameter.
2. Method 1 prohibits the location of measurement
points within 1 inch of the wall for ducts larger
than 24 inches diameter and within 0.5 inch of the
wall for ducts smaller than 24 inches diameter.
For inspection purposes, measurement points will
be at the locations prescribed by the location
procedures, with no adjustments made.
3. Method 2 requires the determination of the
apparent dry molecular weight using Method 3 and
the moisture content using Method 4 in order to
calculate the gas velocity and flowrate. For 6-22
inspection purposes, an apparent dry molecular
weight of 28.95 and a moisture content of zero
will be assumed.
4. Method 2 requires the measurement of the absolute
stack pressure in order to calculate the gas
velocity and flowrate. For inspection purposes,
an absolute stack pressure of 29.92 in. Hg will be
assumed.
With these changes, the procedures for determining
flowrate become comparable to those recommended by
ACGIH.
Measurement of velocity pressure can be made with 6-23
either a S-type (Staubscheid) pitot tube or a standard
pitot tube, both of which are shown in Figure 6-4. The
S-type pitot is preferred when there are particles in
the gas stream that could plug the static pressure holes
of a standard pitot. If the construction of a standard
pitot conforms to Section 2.7 of Method 2, it does not
have to be calibrated and may be assigned a pitot
coefficient, c , of 0.99. An s-type pitot may be 6-24
assigned a pitot coefficient of 0.84 if its construction
conforms to Section 4.1 of Method 2. Since most all of
the commercially available pitot tubes conform to these
requirements, they will not be repeated here.
Procedures for calibrating an S-type pitot may be found
in Section 4 of Method 2.
Lesson 6
6-9
Measurements
-------
For highest accuracy, pressures measured with the
pitot tube should be read using an inclined manometer.
The inclined manometer is similar to the U-tube
manometer previously discussed, except that the first
inch is inclined to improve the ability to read low
pressures accurately. Since standard air flowing at
4005 feet/minute has a velocity pressure of 1 in. h20,
most velocity pressure readings will be made in this
inclined area. If less accuracy is acceptable, a 0-2
in. H,0 Magnehelic® pressure gauge could be substituted
for the manometer.
Average velocity pressure is determined by 6-25
averaging the square roots of velocity pressures
measured at prescribed locations in the duct.
Temperature is measured at the same time by attaching a
thermocouple to the pitot tube, and its average is
calculated arithmetically. The number of locations that
are used to compute the averages depends on the degree
of accuracy desired. Figure 6-5 provides guidance in
determining the minimum number of locations, based on
the upstream and downstream distances to flow
obstructions. In general, any condition other than
straight duct constitutes an obstruction. In using this
chart, the number of locations is first determined for
the distance
Lesson 6
6-10
Measurements
-------
Slides
50
I 40
a.
2 30
0.5
i
3
z
E
20
| 10
2
Duct Diameters Upstream from Flow Disturbance (Distance A)
1.0 1.5 2.0 2.5
i i i i T
^Higher Number is for Rsctanguiar Stacks or Ducts
T
16
I
Disturbance
Measurement
r--- Sita
~ ID
Disturbance
Slack Diameter .0.61 m (24 in.)
12
2
8 or 9* -
Stack Diameter - 0.30 to 0.61 m (12-24 in.)
_J I I ' » I I
3 4 5 6 7 8 9
Duct Diameters Downstream from Flow Disturbance (Distance B)
T
10
Figure 6-5. Minimum Number of Traverse Points
(40 CFR 60, Appendix A)
downstream from an obstruction by reading vertically
upward from the lower x-axis and then for the distance
upstream from an obstruction by reading vertically
downward from the upper x-axis. The larger of these two
numbers is the minimum number of locations or traverse
points. The number of locations for rectangular ducts
is based on equivalent diameter calculated from:
= 2LW/(L + W)(6-1)
where L and W are the lengths of adjacent
sides of the duct.
For circular ducts, the number of traverse points 6-26
determined from Figure 6-5 is divided by two to and
determine the number of measurement points on each 6-27
diameter. The location of each traverse point is given
in Table 6-1 as a percentage of duct diameter
Lesson 6
6-11
Measurements
-------
Table 6-1. Location of Traverse Points*
in Circular Ducts
Point Number
Number of
4
Points on a
6
Diameter
8
1
6.7
4.4
3.2
2
25.0
14.6
10.5
3
75.0
29.6
19.4
4
93.3
70.4
32.2
5
85.4
67.7
6
95.6
80.6
7
89.5
8
96.8
BssasBn&sBBEH^BnoaBiHaHHBBnB^EBaBBsaHi^EBBssEaassaa^ei^^BnsBns^^aaxsssBS^HBB
" Note:Values are expressed in percent of duct
diameter from inside wall to traverse point.
from the inside wall to the point location. Locations
for other numbers of points may be found in Table 1-2
of EPA Method 1. With rectangular ducts, the cross-
section is divided into a grid of equal rectangular
areas and measurements are made in the center of each
grid element. The grid configuration should be either
3x3, 4x3 or 4x4, depending on the number of
measurement locations needed.
In most ducts the direction of the gas flow is
essentially parallel to the duct walls. However,
downstream of such devices as cyclones, inertial
demisters or ducts with tangential entry, a swirling
or cyclonic motion may be encountered. When high
accuracy is desired, it should be determined whether
the degree of cyclonic flow is enough to cause
significant error in the measurements. The procedure
for accomplishing this is as follows:
1. Level and zero the manometer.
2. Connect an S-type pitot to the manometer.
3. Place the pitot tube at each traverse point so
that the openings are perpendicular to the duct
cross-sectional plane. In this position, each
tube should be reading static pressure and the
indication of the manometer should be zero.
Lesson 6
6-12
Measurements
-------
4. If the differential pressure is not zero, rotate Slides
the pitot tube until a zero reading is obtained
and record the resulting angle. 6-28
5. Calculate the average of the absolute values of
the angles, including those angles that were
zero (no rotation required) . If the average is
greater than 20 degrees, the flow conditions at
that location are not acceptable.
Prior to making any measurements of velocity
pressure, a leak check should be conducted, as follows:
1. Blow through the pitot impact opening until 3
in. HjO velocity pressure registers on the
pressure gauge, then close-off the opening. The
pressure should remain stable for at least 15
seconds.
2. Repeat the above step for the static opening,
except using suction to obtain -3 in. H,0. Once
an acceptable location has been identified, the
velocity pressure and temperature measurements
are performed and the averages are calculated.
The average gas velocity is then determined
from:
V = 2.9C (AP0-5) (T )0-5 (6-2)
" ' averagev average' v '
where V ® average gas velocity (feet/sec)
= pitot coefficient (dimensionless)
VP - velocity pressure (in. H20)
T ¦ gas temperature (°R)
The average gas flowrate is then determined from:
Q = 6OVA (6-3)
where Q = average gas flowrate (ft3/min)
A - duct cross-sectional area (ft2)
Other methods are available for determining gas
velocity and these are typically applied to measure-
ments at the hood face. To determine flowrate with
these devices, it would be necessary to make several
velocity measurements across the face of the hood,
determine an average and then multiply it by the area of
the hood opening. The accuracy of this technique
Lesson 6
6-13
Measurements
-------
will depend primarily on the number of measurement
locations used.
One of the more common velocity measuring
instruments is the swinging vane anemometer. Inside the
instrument case is an aluminum vane which deflects the
pointer on the scale in proportion to the velocity. Air
flows through the probe and connecting tube into the
case and then through the channel which contains the
vane.
A second type of anemometer is the rotating vane
instrument. This anemometer has a small lightweight
propeller that rotates as air flows through the
instrument. The instrument is calibrated in feet and
has to be used with a timing device to determine the
velocity.
A third type of anemometer is the "hot wire. •• The
probe of this instrument is provided with a wire element
that is heated with current from batteries in the
instrument case. As air flows over the element, its
temperature changes from what it was in still air and
the accompanying resistance change is translated into
velocity on the indicating scale of the instrument.
V. FAN SPEED MEASUREMENT 6-29
Techniques available for the measurement of fan
speed include: (1) standard tachometers, (2) strobe-
tachometers and (3) phototachometers. Strobetach-
ometers and phototachometers are expensive instruments
that are not likely to be available to the inspector.
Also, phototachometers require reflective tape to be
placed on the drive shaft and this can only be done when
the shaft is not moving.
The recommended technique for measuring fan speed
is the standard tachometer, and the easiest location for
making the measurement is on the end of the shaft,
through the access hole in the belt guard. If no access
hole is provided, the inspector should request the
assistance of plant personnel. Under no circumstances
should the inspector attempt to obtain access to the
shaft end by enlarging the mesh covering on the belt
guard. An alternative measurement location is on the
shaft, using the roller attachment supplied with the
tachometer. However, using this method requires
knowledge of the shaft diameter in
Lesson 6
6-14
Measurements
-------
order to calculate the rotational speed from the Slides
tachometer reading.
An estimate of the fan speed can be obtained by 6-30
measuring the diameter of the fan and motor sheaves and and
using their ratio, as follows: 6-31
Fan rpm = MS(MD/FD) (6-4)
where MD = motor sheave diameter
FD = fan sheave diameter
MS = motor speed (rpm)
Motors are generally available in the nominal speeds of
600, 1200, 1800, 2400 and 3600 rpm. The actual speed is
somewhat less than the nominal value and is stamped on
the motor nameplate.
VI. HORSEPOWER MEASUREMENT
Determination of operating horsepower is an
involved process that is not likely to be performed very
often. Also, because of the procedures and measurements
required, it should be performed only by plant personnel
and never by the inspector. The procedure plant
personnel should use is as follows:
1. Prepare a graph like that shown in Figure 6-6, 6-32
with horsepower on the x-axis and amperage on the
y-axis.
2. Disconnect the motor from the fan and measure the 6-33
amperage when running at no load. Mark this
reading as point "a" on the y-axis. Divide the no-
load amperage by two and mark this value as point
MbM on the y-axis.
3. Read the full load amperage from the motor
nameplate and draw a horizontal line across the
graph at this value. Read the rated horsepower
from the nameplate and draw a vertical line at
this value until it intersects the full load
amperage line. Call this intersection point"cM.
Draw a straight line from Mbw to Hc".
4. Divide the rated horsepower by two and draw a
vertical line at this value until it intersects
the line from "b" to "cw. Call this intersection
point "d". Draw a smooth curve through points
"a", "dM and "c".
Lesson 6
6-15
Measurements
-------
! i
J
s
: |
1
i i
I j
i
i
1 !
i
i
1 |
—i- Nam
i An
II
1 !
ips
i 1
r
i|
*
c i
j 1
i
! I
I i «
:
No Loads
Amps
i
i |
J
J j
idL
1
|
!
Q-; i
f
i i
1 i ;
9 i
CL'
i i»'i 1 1 }
1 1
£
1 M )
! I i
"ZL
1/2 No Lo
Amps
Z !
#VI f i
s 1
! ! !
z j
Horsepower
Figure 6-6. Determination of Operating Horsepower
(SMACNA, 1967)
Slides
5. Connect the motor to the fan and measure the
amperage when running at load. Read the horsepower from
the curve "a-d-c".
Once the relationship between amperage and
horsepower is determined, it could be used in subsequent
inspections, provided the motor has not been changed.
Estimates of operating horsepower can be made in
two ways. Perhaps the simplest is to have plant
personnel measure the amperage while running at load.
This value would then be divided by the full load
amperage from the nameplate and then multiplied by the
rated horsepower to obtain the estimated operating
horsepower. Another method would require plant
personnel to measure both voltage and amperage while
running at load. For three phase motors, these values
would then be substituted into:
bhp= a^Afe/TASfe-s)
where V= voltage
A= amperage
f= power factor
e= motor efficiency
Lesson 6
6-16
Measurements
-------
For single phase motors, the square root of 3 is Slides
replaced by one. If the power factor were also measured
and the motor efficiency determined from the
manufacturer, then a good value of operating horse-
power could be determined with this relationship.
However, because of the time involved, we can usually
only estimate these parameters and thereby obtain an
estimate of operating horsepower. In the absence of
other information, a combined factor of 0.80-0.85 should
be used for the product of power factor times motor
efficiency.
VI. USE OF GROUNDING CABLES
When working with portable instruments in areas where
potentially explosive or ignitable materials are 6-34
present, all metal probes should be grounded to the duct
to avoid a static discharge. The most satisfactory
technique is to use a stranded cable with a pipe clamp
attached to one end and a spring-loaded jaw clamp on the
other, as shown in Figure 6-7. The pipe clamp is firmly
attached to the probe and the jaw clamp attached to the
duct, usually at a flange or support. Care should be
taken to assure a good connection at the duct and that
all paint and rust has been penetrated. One way to
check the connection would be to measure the resistance
between the probe and the duct using an explosion-proof
ohmmeter.
Figure 6-7. Grounding Cable
Lesson 6
6-17
Measurements
-------
If the resistance is less that 3 ohms, the connection is
good. Guidance on when to use grounding cables is
provided by the following list:
1. When the moisture content of the gas stream is
low.
2. When the gas stream velocity across the probe
is high.
3. When the gas stream contains a relatively high
mass concentration of small-sized particles.
4. When there is the possibility of dust deposits
in the bottom of the duct.
5. When there is any question about the need for a
grounding cable.
References
ACGIH, Industrial Ventilation, Twentieth Edition, Cincinnati, 1988.
Richards, J., "Air Pollution Source Field Inspection Notebook",
Revision 2, USEPA, APTI, June 1988.
Segal, R., and J. Richards, "Inspection Techniques for Evaluation
of Air Pollution Control Equipment", Volume II, EPA-340/l-85-022b,
September 1985.
SMACNA, Manual for the Balancing and Adjustment of Air Distribution
Systems, First Edition, Vienna (Virginia), 1967.
USEPA, "Determination of Stack Gas Velocity and Volumetric Flow
Rate", 40CFR60, Appendix A, Method 2.
USEPA, "Sample and Velocity Traverses for Stationary Sources",
40CFR60, Appendix A, Method 1.
Lesson 6
6-18
Measurements
-------
LESSON 7
VENTILATION SYSTEM INSPECTION
The efforts involved in inspecting air pollution sources are
generally categorized according to a system of "levels", as
follows:
Visual evaluation of stack opacity and
fugitive emissions from off the plant site.
On site evaluation of the control system relying on
plant instruments for the values of any inspection
parameters.
Similar to Level 2, but relying on measurements by
the inspector to determine missing or inaccurate
inspection parameters.
Similar to Level 3, but including the development of
a process flowchart, determination of measurement
port locations and evaluation of safety hazards and
protective equipment needs. If the process or
control equipment do not change, this level of
inspection would only be conducted once.
The inspection level that is actually utilized is dictated by
the individual situation and based on the judgement of the inspec-
tor. For example, if a Level 1 inspection indicates no problems,
the inspector may elect to terminate the inspection and proceed to
another facility. Or, if in the course of a Level 2 inspection,
critical information is needed to complete the evaluation, the
inspector may elect to proceed to Level 3, making on-site
measurements to obtain the data.
Level 1
Level 2
Level 3
Level 4
These same four levels can also be applied to the inspection
of industrial ventilation systems. Level 1 would be limited to an
off-site evaluation of fugitive emissions from hoods and ducts,
and the additional items in a Level 4 would focus on port locations
and safety issues. Most inspections will be conducted at Level 2,
with the occasional need for a Level 3 evaluation, and would follow
the same general approach used with control devices.
In previous lessons, inspection points and procedures have been
discussed that can be used in both Level 2 and Level 3 inspections.
In this lesson, those items will be organized according to level
and, in some cases, expanded. The purpose of this is to provide
both a review of the inspection points and a "check-list" for
Lesson 7
7-1
Inspection
-------
conducting field evaluations. Under the Level 3 category, only
those additional items will be listed. Level 3 inspections would
also include the Level 2 items.
I. LEVEL 2 INSPECTIONS Slides
Hoods 7-2
1. Capture efficiency: visual evaluation of 7-3
fugitive losses as indicated by escaping dust or
refraction lines.
2. Physical condition: hood modifications or
damage that could affect performance; evidence
of corrosion.
3. Fit of "swing-away" joints: evaluation of gap
distance between hood system and duct system on
movable hoods.
4. Hood position/cross-drafts: location of hood
relative to point of contaminant generation;
effect of air currents on contaminant capture.
Ducts 7-4
1. Physical condition: indications of corrosion,
erosion or physical damage; presence of fugitive
emissions.
2. Position of emergency dampers: emergency by-
pass dampers should be closed and not leaking.
3. Position of balancing dampers: a change in
damper positions will change flowrates; mark
dampers with felt pen to document position for
later inspections.
4. Condition of balancing dampers: damper blades
can erode, changing system balance; have plant
personnel remove a few dampers to check their
condition.
Lesson 7
7-2
Inspection
-------
Coolers
Slides
1. Physical condition: indications of corrosion, 7-5
erosion or physical damage; presence of fugitive
emissions.
2. Outlet temperature: observe plant instruments
to determine cooler effectiveness; if controller
is used, compare to set-point value.
3. Spray pattern/nozzle condition: indications of
effective atomization on evaporative coolers.
4. Water flowrate: observe plant flow meters or
pressure gauges to evaluate changes in water
flowrate on evaporative coolers.
Fans 7-6
1. Physical condition: indications of corrosion.
2. Vibration: indications of balance problems due
to material build-up or wheel erosion or
corrosion; severely vibrating fans are a safety
hazard.
3. Belt squeal: squealing belts under normal
operation indicate a loss of air volume.
4. Fan wheel build-up/corrosion: internal
inspection of non-operating fans.
5. Condition of isolation sleeves: check vibration
isolation sleeves for holes.
6. Rotation direction: check rotation direction
with direction marked on fan housing.
II. LEVEL 3 INSPECTIONS 7-7
Hoods
1. Estimated volume: estimate flowrate using SPh, 7-8
temperature and hood configuration.
2. Actual volume: determine flowrate by measuring
VP and temperature.
Lesson 7 7-3 Inspection
-------
Ducts
Slides
1. Change in gas temperature: measure tempera- 7-9
ture change across a section of duct to evaluate
air infiltration.
2. Change in static pressure: measure static
pressure change across a duct section to
evaluate duct deposits; compare measurements to
calculations for clean duct.
3. Air volume: determine flowrate by measuring VP
and temperature.
Coolers 7-10
1. Inlet and outlet temperatures: measure inlet
and outlet temperature to evaluate cooling
effectiveness.
2. Water requirement: estimate water requirement
for evaporative coolers using inlet and outlet
temperatures and enthalpy relationships; compare
to actual use information supplied by plant
personnel or indicated by flow meter.
3. Water turbidity: perform settling test on \rtar
sample gathered by plant personnel to evaluate
particle size of solids.
4. Air volume: estimate air volume in dilution
cooling systems using measured temperatures and
enthalpy relationships; could also be determined
by measuring VP and temperature.
Fans 7-11
1. Volume changes: estimate new flowrate using
known performance (Q, rpm and q) and new rpm.
2. Estimated volume: estimate flowrate using rpm,
FSP, temperature and ratings table; could also
be done using bhp, FSP, temperature and ratings
table.
Lesson 7
7-4
Inspection
-------
III. USE OF FLOWCHARTS
One of the first steps in solving essentially any
technical problem is to draw a picture. This is
especially true in the inspection of air pollution
control equipment. Problems which result in excessive
emissions are rarely due to simple failures of a single
component, but are instead usually due to combinations
of problems affecting the entire system. The inspection
flowchart is a valuable tool in sorting out the usually
complex and sometimes conflicting data. Additional
advantages include:
• Improves inspector's ability to communicate the
results of an inspection to plant personnel and
to inspection supervisors.
• Organizes inspection data making anomalous trends
easier to identify.
• Reduces inspection report preparation time.
Stack Cap
@@ 0
j'j 11 t r
'Inc. /
^Primary '
'Chamber/'
Wast*
Faad
Inc.
Second a/y
Chamfer
Y///M
' Ping*
Figure 7-1. Example flowchart
© Pendragon Inc. 1992
Lesson 7 7-5 Inspection
-------
Because they can serve many purposes, there are slides
many levels of sophistication in flowchart prepara-
tion. Flowcharts for air pollution control equipment
inspections should be relatively simple. One example is 7-14
shown in Figure 7-1. Here, major equipment items are
shown as simple blocks, while items such as the fan,
pumps and the stack are represented with standard
equipment symbols. The gas flow path and important
material and utility streams are included and labeled
for easy reference. Parameter monitoring locations are
shown using standard instrument symbols.
Once the flowchart has been prepared, it can be
duplicated and used to record data on individual 7-12
inspections. Recorded data should be examined to be
sure that static pressures decrease from the inlet of 7-13
the system to the fan and from the fan to the stack.
Likewise, in hot systems temperature should decrease
toward the fan and away from the fan. A slight
temperature increase across the fan may occur because of
compression of the gas stream. Data that do not follow
expected trends should be re-evaluated.
References
J. Richards, "Flowchart preparation for air pollution source
inspection", USEPA, SSCD, September 1989.
Lesson 7
7-6
inspection
-------
APPENDIX A
SLIDES
-------
Slide 1-1
VENTILATION
FUNDAMENTALS
Slide 1-2
PROPERTIES OF AIR
AND AIR-WATER
VAPOR MIXTURES
PRINCIPLES OF
FLUID FLOW
A-l
-------
Slide 1-3
PROPERTIES OF AIR
AND AIR-WATER VAPOR
MIXTURES
Slide 1-4
STANDARD AIR
Density = 0.075 lb/ft3
Absolute Viscosity = 1.225 x 10-Slb|n /ft-sec
Equivalent to dry air at 70 °F and 29.92 in. Hg
A-2
-------
Slide 1-5
MOLECULAR WEIGHT
^^mixture *" S (* component) (^W
component^
MWdry air = aPProx- 28.95 lb / lb-mole
MWwet..r = <1 " * water^ (««wdiyJ +
(x.„J
Slide 1-6
IDEAL GAS LAW
P = absolute pressure (lb{ /ft2)
V = gas volume (ft3)
n = number of moles (lb-moles)
R = constant (1545 ft-lbf /lb-mole-°R)
T = absolute temperature (°R)
A-3
-------
Slide 1-7
IF MOLES DON'T
CHANGE...
PV/ T = nR = constant
p,v1/T = p2v2/ts
or
V,=V2(P2/P,)CT,/T2)
Slide 1-8
Density =
Weight / Unit Volume
Specific Volume =
Volume / Unit Weight = 1 / Density
For an ideal gas:
Density (p) =
(MW / 387) (530 / T) (P / 29.92)
A—4
-------
Slide 1-9
Specific Gravity = Pga8/ pdryair
For an ideal gas:
Specific Gravity = MWgas/MWdrya|r
Slide 1-10
G1BBS-DALTON RULE
OF PARTIAL PRESSURE
^total— Rdrv alr^"
total ndry air "water
A-5
-------
Slide 1-11
Relative Saturation =
Pwater ^water at saturation
Relative Humidity =
Relative Saturation x 100
Absolute Humidity =
Wei3ht»,t.r/Wei9h«drv.lr
Slide 1-12
Dry Bulb Temperature
Wet Bulb Temperature
Dew Point Temperature
A-6
-------
Slide 1-13
ENTHALPY
The measure of the thermal
energy of a substance
Slide 1-14
h = Cp(t-tref)
h
= enthalpy (Btu / lb)
cP
= heat capacity at constant
pressure (Btu / lb-°F)
t
= temperature of substance (°F)
*ref
= reference temperature (°F)
for dry air usually equals 0 °F
for water usually equals 32 °F
A-7
-------
Slide 1-15
^water vapor" ^water \
X = latent heat of vaporization
(Btu/Ib)
Slide 1-16
FOR AN AIR-WATER
VAPOR MIXTURE:
h (Btu/lbdiy J = hdiyjlr ++ (h^^)
= absolute humidity (Ib^ /Ib^J
A-8
-------
Slide 1-17
USUALLY INTERESTED IN
ENTHALPY DIFFERENCE
AH=hj-h,
MCpUta-tJ-tCpUV*,.,)
= (CP)lvg (t, -t,)
Slide 1-18
PSYCHROMETRIC
CHART
A-9
-------
Slide 1-19
PRINCIPLES OF
FLUID FLOW
Slide 1-20
A-10
© Pendragon Inc. 1992
-------
Slide 1-21
IF NO ACCUMULATION
OR REMOVAL...
G = G
\*i \*2
or
p,v,A = pyA
or, if p1 » p2 V1 A, = YA
Slide 1-22
BERNOULLI'S
EQUATION
For an Incompressible, Invlscld fluid:
V2/2 + P/p+ gZ = constant
V = fluid velocity
P = fluid pressure
p = fluid density
g = acceleration of gravity
Z = elevation of fluid
A-ll
-------
Slide 1-23
REARRANGING GIVES...
V2 /2g + P£g + Z = constant
V2 /2g= velocity head (ft)
P/pg = pressure head (ft)
Z - potential head (ft)
Slide 1-24
2
•y
S,
w
S*
V
W
V
S*
n
? h
^ I
I
V
—*- i
if pa
HI
uuutv
A-12 © Pendragon Inc. 1992
-------
Slide 1-25
SUBSTITUTING GIVES,
V,' /2g + P, /pg = P2/pg
V,2 /2g= velocity pressure
P, = static pressure
P2 £>g = total pressure
Slide 1-26
Discharge Side
intake Side
3500 FPM a
3500 FPM
IkU
¦1.0 + 0.6
Sp + Vp
5.0 + 0.6 - 1.1
Sp + Vp • Tp
A-13
© Pendragon Inc. 1992
-------
Slide 1-27
VELOCITY CAN BE
DETERMINED FROM
VELOCITY PRESSURE
V= 1096.7 (Vp/p)°5
V = velocity (ft/min)
Vp = velocity pressure (in. H,0)
p = density (lb/ft3)
For Standard Air(p = 0.075 lb/ft3)
V = 4005 (Vp)
A-14
-------
Slide 2-1
HOOD
SYSTEMS
Slide 2-2
The Goal of Good Hood
Design is High Capture
Efficiency
A-15
-------
Slide 2-3
PENETRATION =
1 - FRACTIONAL EFFICIENCY
Pt + (1 - pt ) Pt
hood * Tiood ' collector
Slide 2-4
TYPES OF HOODS
• Enclosures
• Receiving
• Exterior
A-16
-------
Slide 2-5
HOOD DESIGN
PRINCIPLES
• Enclose whenever possible
• If can't enclose, place hood close to source
• Locate duct take-offs in line with normal
contaminent flow
Slide 2-6
Air Handling
Duct
A-17
© Pendragon Inc. 1992
-------
Slide 2-7
Ladel
© Pendragon Inc. 1992
Slide 2-8
A-18
© Pendragon Inc. 1992
-------
Slide 2-9
© Pendragon Inc. 1992
Slide 2-10
&
t
[l PcT-1
Hot Source
Dg-
y
*r
Hypothetleal
Point Source —
A-19
© Pendragon Inc. 1992
-------
Slide 2-11
HOOD DESIGN
PRINCIPLES
• Enclose whenever possible
• If can't enclose, place hood close to source
• Locate duct take-offs in line with normal
contaminent flow
Slide 2-12
A-20
O Pendmgon Inc. 1992
-------
Slide 2-13
Jat Exhaust
Side Side
© Pendragon Inc. 1992
Slide 2-14
FACTORS AFFECTING
HOOD PERFORMANCE
• Thermal air currents
• Motion of machinery
• Material motion
• Movement of operator
• Room air currents
A-21
© Pendragon Inc. 1992
-------
Slide 2-15
CAPTURE VELOCITY
Air velocity in front of hood or at hood
face necessary to overcome air currents
and cause air to move into hood.
Slide 2-16
RANGE OF CAPTURE
VELOCITIES (ft/min)
Type of Material Release
Capture Velocity
With no velocity into quiet air
50- 100
At low velocity into moderately
100 - 200
still air
Active generation into zone of
200-500
rapid air motion
With high velocity into zone of
500 -2000
very rapid air motion
A-22
-------
Slide 2-17
FOR COLD FLOW
INTO HOODS
Capture velocity decreases
with distance from hood face.
Slide 2-18
© Pendragon Inc. 1992
A-23
© Pendragon Inc. 1992
-------
Slide 2-19
© Pendragon Inc. 1992
Slide 2 - 20a
Hood TVp*
Description
As pact Ratio
Air Flow
Slot
0.2 or Less
Q « 3.7 LVX
Flanged Slot
Odorless
Q - 2.6 LVX
W
>#
A»WL(*q. ft.)
Plain Opening
0.2 or Greater
and Round
O-V(10X2*A)
Flanged Opening
0.2 or Greater
and Round
O-0.75V (10X2 ~A)
A-24
C Pendragon Inc. 1992
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Slide 2 - 20b
Hood Type
Description
Booth
Canopy
Plain Multiple
Slot Opening
2 or Mora Slots
Flanged Multiple
Slot Opening
2 or Mora Slots
Aspect Ratio
To Suit Work
To Suit Work
0.2 or Graaw
0.2 or Gtmmt
Air Flow
O . VA . VWH
O. 1.4 PVD
See VS-903
P - Pehwef
D > Height Above Work
O.V(10X2*A)
O-0.75V (10X2+A)
Slide 2-21
FOR HOT FLOW
INTO HOODS
• As plume rises it expands, cools and
slows down.
• Long rise distances make plume more
subject to movement by air currents.
• Because of distance between source and
hood, air volumes are usually large.
A-25
-------
Slide 2-22
Position
Position 2
Slide 2-23
© Pendragon Inc. 1992
But there is no air motion
at point 1; therefore:
0 = SP2+ VP2
or
SP2 = -VP2
A-26
O Pendragon Inc. 1992
-------
Slide 2-24
Vena Contracta
Sph = -Sp2 = Vp2 + he
he = hood entry loss = F^Vp
Fh = hood entry loss factor
Slide 2-25
HOOD ENTRY
COEFFICIENT
Ce = (VP/SPJ"
A-27
-------
Slide 2-26
C AND F. ARE RELATED
e h
\ = [(1 - C.»)/C e'J VP
therefore:
Fh = (1 - C*)/C*
Slide 2-27
ESTIMATING HOOD VOLUME
Q = VA = 1096.7A (VP/p)
= 1096.7AC (SPh/p)
6 n
Q = volumetric flow rate (ft3 /min)
V =s hood face velocity (fp /min)
A = hood inlet area (ft3)
p = air density (lb/ft3)
A—28
-------
Slide 3-1
DUCT
SYSTEMS
Slide 3-2
A-29
© Pendragon Inc. 1992
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Slide 3-3
Because of friction and
non-ideal conversion
between SP and VP...
TP, = T P2 + hL
or
SP, + VP, = SP2 + VP2+ 1^
Slide 3-4
If velocity is constant
between points 1 and 2...
SP, * SP2 + hL
A-30
-------
Slide 3-5
TYPES OF LOSSES
• Frictional losses
• Fitting losses
• Acceleration losses
Slide 3-6
WAYS TO
ESTIMATE LOSSES
• Equivalent length
• Velocity pressure
• Total pressure
A-31
-------
Slide 3-7
FRICTIONAL LOSSES
hL1 = H,L (VP)
hL1 = frictional loss (in. H20)
Hf = frictional loss factor (Vp /ft)
L = duct length (ft)
VP = velocity pressure (in. H20)
Slide 3-8
FITTING LOSSES
hL2=F (VP)
h,^ = fitting loss (in. H20)
F = fitting loss factor (VP)
VP » velocity pressure (in. H20)
A-32
-------
Slide 3-9
VELOCITY PRESSURE
CALCULATION METHOD
1. Determine duct velocity and VP.
2. Determine hood static pressure.
3. Multiply straight duct length by
friction loss factor.
Slide 3-10
VELOCITY PRESSURE
CALCULATION METHOD (cont'd)
4. Determine number and type of
fittings. Multiply fitting loss
factor by number of that type and
sum for all types.
5. Add results for steps 3 and 4 and
multiply by duct VP from step 1.
A-33
-------
Slide 3-11
VELOCITY PRESSURE
CALCULATION METHOD (cont'd)
6. Add result from step 5 to the hood
static pressure.
7. Add result from step 6 to any other
losses expressed in inches H20.
Slide 3-12
This calculation gives the total
energy required, expressed as
static pressure, to move the gas
volume through the duct segment.
A-34
-------
Slide 3-13
transport velocity
That duct velocity required
to prevent build-up
Slide 3-14
RANGE OF DESIGN
VELOCITIES
Contaminant
Design Velocity (ft/min)
Vapors, gases
Usually 1000 - 2000
Fumes
1400 • 2000
Very fine, light dust
2000 - 2500
Dry dust and powders
2500 - 3500
Average Industrial dust
3500 - 4000
Heavy dusts
4000 - 4500
Heavy or moist
4500 and up
A-35
-------
Slide 3-15
BUILD-UP CONCERNS
• Increase duct resistance
• Duct pluggage
• Vibration breakup
• Increased weight
Slide 3-16
DISTRIBUTION OF FLOW IN
BRANCHED DUCTWORK
All ducts entering a junction must have
equal static pressure requirements.
A—36
-------
Slide 3-17
&
Sp m -1.0 inch HoO
h *
Sp - -2.0 lnch«s H20
t
°1
- Sp » -1.5 Inches H2O
Itl"00"1 Ifl
¦ Sp ¦ -1.0 Inch H2O
t
°2
Hood 2
Slide 3-18
TECHNIQUES TO
OBTAIN BALANCE
* Through design
* With dampers
A-37
© Pendragon Inc. 1992
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Slide 3-19
CHARACTERISTICS OF
BALANCE THROUGH DESIGN
•Volume not easily changed
•Little flexibility
•No unusual erosion or build-up problems
•Final volume may be greater than design
•All branch losses identified
•Must be installed exactly as designed
Slide 3-20
CHARACTERISTICS OF
BALANCE WITH DAMPERS
•Volumes easily changed
•Great flexibility
•Dampers erode or cause build-up
•Balance achieved with design volume
•May miss branch of greatest resistance
• Moderate variation in installation allowed
A-38
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Slide 3-21
© Pendragon Inc. 1992
A—39
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Slide 4-1
GAS
COOLING
Slide 4-2
METHOD FOR
COOLING GASES
• Dilution with cooler gases
• Quenching with water
• Natural convection and radiation
• Forced convection
A—40
-------
Slide 4-3
Dilution with Cooler Gases
m.
T, = ambient
m3
T. = hot
Damper
m3
T3 = mix
From Heat Balance:
mihn+ m2hT2= m3hT3
mh = (Ib/min) (Btu/lb) = Btu/min
m3 = m1 + m2
Slide 4-4
© Pendragon Inc. 1992
PROBLEMS WITH
DILUTION
* Large air volume
* Lack of temperature control
* Corrosion potential
A—41
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Slide 4-5
400
380
360
340
u. 320
L- 300
O 280
| 2M
° 240
220
200
0.01 0.1 1.0 10 100 1000
SO3 (H2SO4) In Flue Gas, ppm
€ Pendragon Inc. 1992
Slide 4-6
QUENCHING WITH WATER
Heat loss from gas stream is taken
up by evaporating water...
mga, ^gaa in " ^g«» ouP = m*mter^water vapor ^ water^
A—42
-------
Slide 4-7
PROBLEMS WITH
QUENCHING
• Water carryover
• Temperature control
• Corrosion
Slide 4-8
NATURAL CONVECTION
AND RADIATION
Heat lost from gas stream is
taken up by atmosphere
m (h nM - h ) = U A AT
gas * gas in gas our
A-43
-------
Slide 4-9
PROBLEMS WITH
CONVECTION AND
RADIATION COOLING
• Large size
• Dust settling
• Lack of temperature control
¦ Corrosion potential
A-44
-------
Slide 5-1
FAN
SYSTEMS
Slide 5-2
Forced Draft Fan
(Dirty)
d
Pollutant
Source
Control Device
{under
Stack
Forced Draft Fan
(Clean)
Pollutant
Source
Control Oevioe
{under
auction)
Fan Stack
A-45
© Pendragon Inc. '1992
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Slide 5-3
TYPES OF FANS
• Axial
• Centrifugal
• Special
Slide 5-4
AXIAL FANS
• Propeller
• Tubeaxial
• Vaneaxial
A-46
-------
Slide 5-5
Scroll Side
Scroll Piece
Side Sheet
Side Plate
Inlet
Inlet Cone
Inlet Bell
Inlet Rare
Inlet Nozzle
Venturi
Backplate
Hub Disk
Hubplate
Webplate
Blades
Housing
Scroll Housing
Volute
Casing
Blast Area
Outlet
Discharge
Outlet Area
Inlet Collar
Inlet Sleeve
Inlet Band
t
— Rim
Shroud
Wheel Ring
Wheel Cone
Retaining Ring
Inlet Rim
Wheel Rim
—Range
Inlet Plate
Impeller
Wheel
Rotor
Scroll
Band
Scroll Sheet
Wrapper
Wrap Sheet
Scroll Back
Slide 5-6
CENTRIFUGAL
* Forward curved
* Radial
* Backward inclined
• standard blade
- airfoil blade
A-47
-------
Slide 5-7
Forward Curved
• Has 24-64 shallow blades
• Efficiency less than backward inclined
• Smallest of all centrifugal fans
• Operates at lowest speed
Slide 5-8
Fan Wheel
Fan Housing
Forward Curved
Fan Blades
t--
a
X
kT
$
8.
a
O)
fc-
O
Z
o ..
>>
u
c
9
£
3=
UJ
A-48
O Pendragon Inc. 1992
-------
Slide 5-9
Radial
—~
• Has 6-10 blades
• Least efficient
• Narrowest of all centrifugal fans
• Operates at medium speed
Slide 5-10
Fan Wheel
Fan Housing
Radial
Fan Blades
0
a
5
® 1
-------
Slide 5-11
Backward Inclined
Airfoil Blade
• Has 9-16 blades
• Most efficient
• Operates at highest speed
Slide 5-12
Fan Wheel
Fan Housing
Backward Curved
Fan Blades
i i i
t-'
a.
x
w~
®
8.
C3
M
w
o
X
w
Q
O
::t
s
0
1
LU
Gas Flow. SCF/min.-
A-50
© Pendragon Inc. 1992
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Slide 5-13
Houaing
.Scaring*
FtnWhMl
ii
Fmn <9*mH
Support
ARR. 1 SWSI for baft drive or attract
connection. Impotlor ovortiung.
Two tearing* on I
Slide 5-14
FAN LAWS
Q, = Q2 (size2 /size,)3 (rpm2 /rpm,)
P2 = P, (size^size^frpnyrpm,) *(p2Pl)
bhp2 s bhp, (size2 /size,)' (rpm* /rpm1)3 (p/p,)
A-51
-------
Slide 5-15
160
140
©
cc
E
o>
120
c/>
•a
a
5
3
J2
(8
o
'o
c
100
80 -
60 -
8 40
IB
a.
20
I I I I I
Higher Resistance |
144%
i i i r—"
Design Resistance
Lower Resistance
Calculated Gas Flow Rate.
Arrangement A
X
20 40 60 80 100 120 140 160 180 200 220
Percent of Calculated System Gas Row Rate (SCFM) >¦
© Pendragon Inc. 1992
Slide 5-16
Percent of Fan Wide Open Gas Flow Rata (SCFM)
15 30 45 60 75 90 105 120 135 1 50 165
160
I 140
2
120
M 100
¦= 80
a
~ 60
8 40
o
Q.
20
I I I I J
/X
U I I I 1 —1— I
' Jy
¦ 3 -
yA >— Fan Characteristic
/ \ Curve at RPM Y
o
5
*»
m
o
6
112
96
80
64
48
32
16
e
iS
¦s
s.
20 40 60 80 100 120 140 160 180 200 220
Peicant of Catarialad System Gas Flow Rale (SCFM) >
A-52
O Pendragon Inc. 1992
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Slide 5-17
Percent of Fan Wide Open Gas Flow Rale (SCFM)
15 30 45 60 75 90 105 120 135 150 165
I I I 1 ! —
X
2
a.
S
<
c
150 |
m
5
100
o
X
J
50 £
a
&
20 40 60 80 100 120 140 160 180 200 220
Percent of Calculated Systsm Gas Flow Rait (SCFM) >¦
© Pendragon Inc. 1992
Slide 5-18
Fan Pressure Curve
90075 RM1t3
"*"* N
\
Duet Sysism Curve A
9 0.075 to/It3 Density
at Fan Met
Duct Sysism Curve A
9 0.0375 Density
at Fan Met
Density 0.0375
S 0.0575
" 0.075
0.5
20 40 60 80 100 120 140 160 180 200
Percent of Duet System Volume Ftew (CFM)
A-53
O Pendragon Inc. 1992
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Slide 5-19
© Pendragon Inc. 1992
Slide 5-20
FAN SELECTION
1.
Determine volume required at actual
conditions
2.
Calculate fan static pressure (FSP) at
actual conditions- FSP = SP - s P - VP
out In In
3.
Correct FSP to equivalent value for
standardair-
FSP#=FSPA(0.075/pa)
A-54
-------
Slide 5-21
FAN SELECTION
4.
Enter ratings table at actual volume
and equivalent FSP
5.
Determine rpm and bhp
6.
Correct bhp to actual conditions
bhpa = bhp#(pa/0.075)
A-55
-------
Slide 6-1
MEASUREMENT OF
VENTILATION SYSTEM
PARAMETERS
Slide 6-2
MEASUREMENT PORTS
FOR PORTABLE
INSPECTION EQUIPMENT
A-56
-------
Slide 6-3
A-57
-------
Slide 6-5
Slide 6-6
|
A—58
-------
Slide 6-7
A-59
-------
Slide 6-9
MEASUREMENT OF
PRESSURE
• Manometer
• Differential pressure gauge
Slide 6-10
%
..p®
StaHe Pr—u
A-60
-------
Slide 6-11
A—61
-------
Slide 6-13
Slide 6-14
TECHNIQUES FOR THE
MEASUREMENT OF
TEMPERATURE
1.
Mercury thermometer
2.
Dial-type thermometer
3.
Thermister
4.
Thermocouple (battery powered)
A—62
-------
Slide 6-15
THERMOCOUPLES
1. Calibrate probe and meter against
an NIST traceable thermocouple.
2. Check ice point and boiling point
values prior to each day.
Slide 6-16
TEMPERATURE
MEASUREMENT ERRORS
1. Unrepresentative measurement
location
2. Cooling of the probe due to air
infiltration through the port
3. Impaction of water droplets
A—63
-------
Slide 6-17
Slide 6-18
A-64
-------
Slide 6-19
Slide 6-20
MEASUREMENT OF
GAS FLOW RATE
A-65
-------
Slide 6-21
40CFR60 APPENDIX A
METHODS 1 ¦ 4
Slide 6-22
A—66
-------
Slide 6-23
Slide 6-24
(cf». = "
(AorB)»E|(Cp). -(Cp).
-------
50
Duct Diameters Upstream from Row Disturbance (Distance A)
0.5 1.0 1.5 2.0 2.5
(A
I«
®
®
2 30
o
®
Is 20
2
E
10 -
J 1 1 1 1—
3Higher Number is (or Ractanguiar Stacks or Ducts
_1S_
^Disturbance
Maaauramant
?— Site
B
{(Disturbance
Stack Diamatar .0.61 m (24 in.)
12
1
8or9* —
1
Stack Diameter • 0.30 to 0.61 m (12-24 in.) -J
I I I I I
3 4 5 6 7 8 9
Duct Diameters Downstream from Flow Disturbance (Distance B)
10
Slide 6-26
A—68
-------
Slide 6-27
LOCATION OF TRAVERSE POINTS IN CIRCULAR STACKS
ft
t*
««
••
«*
i
I4J 17
44
U
U
Lt
14
J
14
u
I.T
vt
t
K4 MJ
14J
1 »J
u
IT
ft.7
kj
mj
as
tw
1U
MJ
tu
T
ms
m
•A
MJ
sj
&j
ha
IU
141
•
MJ
*JI
a*
1TJ
1U
HI
tij
1M
•
MA
MJ
ni
MJ
JU
MJ
MJ
IU
If
H4
MJ
1M
tij
«i
MJ
»J
I7J
1>
M
«4
1M
•II
MJ
IU
tj
trj
mi
«.i
7*4
»4
M.7
tu
II
MJ
•TJ
•U
MJ
14
11
•U
tij
«u
71J
nj
•7.7
¦.1
m.1
MJ
IU
nj
1t
M«4|
MjI
W.I
MJ
ru
1?
14
*«¦ MM
WWM
• MM
MIM
mj
mj
M4
MS
If
at
n
M
Slide 6-28
Cheeking for Cyclonic Row
Duct
MM
Mil «f duct)
A-by
C Pendragon Inc. 1992
-------
Slide 6-29
MEASUREMENT OF
FAN SPEED
1. Manual tachometer
2. Phototachometer
3. Strobetachometer
4. Sheeve ratio calculation
Slide 6-30
A—70
-------
Slide 6-31
RPM = MS(MD/FD)
Slide 6-32
-------
Slide 6-33
Slide 6-34
A-72
-------
Slide 7-1
VENTILATION SYSTEM
INSPECTION
Slide 7-2
LEVEL 2
A-73
-------
Slide 7-3
HOODS
• Capture efficiency
• Condition
• Fit of "swing-away" joints
• Hood position/cross drafts
Slide 7-4
DUCTS
• Condition
• Position of emergency dampers
• Position of balancing dampers
• Conduction of balancing dampers
A-74
-------
Slide 7-5
COOLERS
• Condition
• Outlet temperature
• Spray pattern / nozzle condition
• Water flowrate
Slide 7-6
FANS
• Condition
• Vibration
• Belt squeal
• Fan wheel build-up/corrosion
• Condition of isolation sleeves
• Rotation direction
A-75
-------
Slide 7-7
LEVEL 3
Slide 7-8
HOODS
• Estimated volume using SPh, p and
configuration
• Estimated volume using VP^ and p
A-76
-------
Slide 7-9
DUCTS
• Change in gas temperature
• Change in static pressure
• Actual volume using VP^ and p
Slide 7-10
COOLERS
• Inlet and outlet temperatures
• Estimated water requirement using
temperature and enthalpy relationships
• Estimated air flow using temperatures
and enthalpy relationships
• Water turbidity
A-77
-------
Slide 7-11
FANS
•Estimated volume using known performance (Q, rpm
and p) and new rpm
•Estimated volume using rpm, FSP, temperature and
ratings tables
•Estimated volume using FSP, bhp, temperature and
ratings tables
- Estimate bhp from voltage snd amperage
measurements
- Estimate bhp from amperage measurement and
name plate full-load ratings
A-78
-------
>
I
u>
O
«P
&
a
?
3
Tempeiatme decrease due to air
infillralion across mechanical collector
iTemp-eialuie decrease
dus to nit inlittralion
across fabric fMter
O? rise due to air
InliHraticn across fan
J Of tlw due toak
T IrflHratton across
I lebrie fitter
'"1 J Tempoiatm« decreajo
* f dug la a* inrffflrafion
across f«n
A O2 fts4 diMloalr
| frniHrafton across
' mechanlcaf collfrctcu
Mechanical
Collector
Fabtic
FiKef
Equivalent Linear Feet
in
h->
M-
cl
ID
I
H1
(v>
Far*
Stack
Discharge
-------
Mechanical Fabric Filter
Collector
Fan
Equivalent Linear Feel
-------
Slide 7-14
© Pendragon Inc. 1992
-------
APPENDIX B
BIBLIOGRAPHY
B-l
-------
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- Instructor's Guide". EPA-450/2-80-003, Februaiy 19S0.
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(Illinois j, 1979.
.AMCA. "Field Performance Measurements". Fan Application Manual. Pan 3, Publication
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.ANSI. "Fundamentals Governing the Design and Operation of Local Exhaust Systems".
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ASHRAE, Equipment Guide and Data Book, New York. 1972.
ASHRAE. Handbook or Fundamentals, Atlanta, 1985.
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-------
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_ _ , uc Staub "How to estimate system total pressure requiremer.ts .
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w A and J Murrow, 'Evaluation of hoods for low volume-high velocity exhaust
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, D n TrMtman Ventilation for Control of the Work
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x m Flvnn Prediction and measurement of velocity into
uiM ®- =" »
c—%raar -" ""*s ~
Ventilating, 60(6), 85 (lSto*). ^
crowder. J.W, and KJ. Loudennilk, 'Balancing of industrial venulauon systems , JAPCA.
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ifi 15J7 (10
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B-4
-------
Yallc, J.M.. ana 7. Hatch. "Studies in the -^r.gr. : jcaJ exhaus: hoods . .4J v/£
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ventilation systems", AIHA Journal, 44, 752 (1983).
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Journal. 47,'448 (1986).
First. M.W.. and L Silverman, "Airfoil pitometer", I&EC, 42.301 (1950).
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~ for Exhaust Systems for Open Surface Tanks", NIOSH Publication 75-108,w1975.
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suction", Ann. Occup. Hyg. ', 20,141.(1977).
Fletcher. B.. "Effect of flanges on the velocity in front of exhaust ventilation hoods". Ann.
Occup. Hyg., 21,265 (1978).
Fletcher. B.. and A Johnson, "Velocity profiles around hoods and slots and the effects of an
adjacent plane",w4nn. Occup. Hyg., 25,365 (1986).
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hoods", Ann. Occup. Hyg., 30,497 (1986).
Flvnn. M.. and. MJ. Ellenbecker, "Empirical validation of theoretical velocity fields into
Ganged circular hoods", AIHA Journal, 48,380 (1987).
Flvnn. M.. and MJ. Ellenbecker, "The potential flow solution for air flow into a flanged
circular hood", AIHA Journal, 46,318 (1985).
Garrison. R.. "Centerline velocity gradients for plain and flaneed local exhaust inlets", AIHA
Journal. 42, 739 (1981).
3arrison. R., "Velocitv calculation for local exhaust inlets—graphical desien", AIHA Journal,
44.941 (1983).'
B-5
-------
Garrison. R.. "Velocity calculation for local exhaust inlets-empirical design equat ons".
AJHA Journal. 44, 937 (1983).
_ , t\ nv»rr "Static oressurc and velocirv characteristics of circular nozziss
GamStaWghanv=loc,t^oCwvolSn= Klvemilation'.AJHA Journal. 41.803 (1980).
_ , j-v T3v»rr "Static nressure. velocity and noise characteristics of rectangular
GamS'exhaust ventilation". AJHA Journal 41. 855
(1980).
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^ . r and R T "Industrial Health and Safety Criteria
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Guffev S.E., and J. Hickey, "Equations for redesign of existing ventilation systems\AIK-i
' Journal, 44,819 (1983).
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"(1983).
Hama G.M. "A calibrating wind tunnel for air measuring instrumems\^fr&gih«m«, <1.
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pilot xubt",AlHA Journal. 28.204 (1967).
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of -air - h —1
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B-6
-------
—ernscr.. vV.C.L.. Plan: ana Process Ventilation. Ssccnc Edition. L-.austnai Press. Nev.
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107(1946).
Maichaire. J.B.. "Design of industrial exhaust systems using a programmable calculator".
Ann. Occup. Hyg., 24. 217 (1981).
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Air Conditioning, Octooer 1969. p. 70.'
B-7
-------
McDermoru KJ.. Handbook of Ventilation for Contaminant Control. Ann Arbor Science.
Ann Arbor. 1977.
Mclnnes. R.G.. B.R. Hobbs and S.V. Caponc. "Guide for Inspecting Capture Systems anc
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65. ' '
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-------
APPENDIX C
PSYCHROMETRIC CHARTS
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