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Full text of "ACGIH: Industrial Ventilation Manual"

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By Authority Of 

THE UNITED STATES OF AMERICA 

Legally Binding Document 



By the Authority Vested By Part 5 of the United States Code § 552(a) and 
Part 1 of the Code of Regulations § 51 the attached document has been duly 
INCORPORATED BY REFERENCE and shall be considered legally 
binding upon all citizens and residents of the United States of America. 
HEED THIS NOTICE : Criminal penalties may apply for noncompliance. 




Document Name: ACGIH: Industrial Ventilation Manual 

CFR Section(s) : 40 CFR 63 .2984(e) 

Standards Body: American Conference of Governmental Industrial 

Hygienists 



A Manual of Recommended Practice 



23rd Edition 



1998 



American Conference of Governmental Industrial Hygienists 
1330 Kemper Meadow Drive 
Cincinnati, Ohio 45240-1634 



Copyright© 1998 

by 

American Conference of Governmental Industrial Hygienists, Inc. 



Previous Editions 

Copyright© 1951, 1952, 1954, 1956, 1958, 1960, 1962, 1964, 1966, 
1968, 1970, 1972, 1974, 1976, 1978, 1980, 1982, 1984, 1986 



by 



Committee on Industrial Ventilation 
American Conference of Governmental Industrial Hygienists 



1st Edition — 1951 


13th Edition — 1974 


2nd Edition— 1952 


14th Edition — 1976 


3rd Edition — 1954 


15th Edition — 1978 


4th Edition — 1956 
5th Edition— 1958 
6th Edition — 1960 
7th Edition — 1962 
8th Edition— 1964 
9th Edition — 1966 


16th Edition — 1980 
17th Edition — 1982 
18th Edition— 1984 
19th Edition — 1986 
20th Edition — 1988 


10th Edition — 1968 


2 1st Edition— 1992 


11th Edition — 1970 


22nd Edition— 1995 


12th Edition — 1972 


23rd Edition — Metric —1998 



Third Printing 



ALL RIGHTS RESERVED. No part of this work covered by the copyright hereon may be reproduced or used in any form or by 
any means — graphic, electronic, or mechanical including photocopying, recording, taping, or information storage and retrieval 
systems — without written permission of the publisher. 

Published in the United States of America by 

American Conference of Governmental Industrial Hygienists, Inc. 

1330 Kemper Meadow Drive 

Cincinnati, Ohio 45240-1 634 



ISBN: 1-882417-22-4 
Printed in the United States 



CONTENTS 



DEDICATION . . . vii 

FOREWORD ix 

ACKNOWLEDGMENTS . . . . xi 

DEFINITIONS xiii 

ABBREVIATIONS xv 

CHAPTER 1 GENERAL PRINCIPLES OF VENTILATION 1-1 

1.1 Introduction 1-2 

1.2 Supply Systems 1-2 

1.3 Exhaust Systems 1-2 

1.4 Basic Definitions 1-3 

1.5 Principles of Air Flow 1-4 

1.6 Acceleration of Air and Hood Entry Losses 1-6 

1.7 Duct Losses 1-7 

1.8 Multiple-Hood Exhaust Systems 1-9 

1.9 Air Flow Characteristics of Blowing and Exhausting 1-10 

References 1-10 

CFIAPTER2 GENERAL INDUSTRIAL VENTILATION 2-1 

2.1 Introduction 2-2 

2.2 Dilution Ventilation Principles 2-2 

2.3 Dilution Ventilation for Health 2-2 

2.4 Mixtures — Dilution Ventilation for Health 2-6 

2.5 Dilution Ventilation for Fire and Explosion 2-7 

2.6 Fire Dilution Ventilation for Mixtures 2-8 

2.7 Ventilation for Heat Control 2-8 

2.8 Heat Balance and Exchange 2-8 

2.9 Adaptive Mechanism of the Body 2-9 

2.10 Acclimatization 2-10 

2.11 Acute Heat Disorders 2-10 

2.12 Assessment of Heat Stress and Heat Strain 2-11 

2.13 Worker Protection 2-13 

2.14 Ventilation Control 2-13 

2.15 Ventilation Systems 2-13 

2.16 Velocity Cooling 2-15 

2.17 Radiant Heat Control 2-15 

2.18 Protective Suits for Short Exposures 2-16 

2.19 Respiratory Heat Exchangers 2-16 

2.20 Refrigerated Suits 2-16 

2.21 Enclosures 2-16 

2.22 Insulation 2-16 

References 2-17 

CHAPTER 3 LOCAL EXHAUST HOODS 3-1 

3.1 Introduction 3-2 

3.2 Contaminant Characteristics 3-2 

3.3 Hood Types 3-2 

3.4 Hood Design Factors 3-2 

3.5 Hood Losses 3-15 



in 



IV 



Industrial Ventilation 



3.6 Minimum Duct Velocity 3-18 

3.7 Special Hood Requirements 3-18 

3.8 Push-Pull Ventilation 3-19 

3.9 Hot Processes 3-21 

References 3-23 

CHAPTER 4 AIR CLEANING DEVICES 4-1 

4.1 Introduction 4-2 

4.2 Selection of Dust Collection Equipment 4-2 

4.3 Dust Collector Types 4-3 

4.4 Additional Aids in Dust Collector Selection 4-22 

4.5 Control of Mist, Gas, and Vapor Contaminants 4-22 

4.6 Gaseous Contaminant Collectors 4-25 

4.7 Unit Collectors 4-25 

4.8 Dust Collecting Equipment Cost 4-25 

4.9 Selection of Air Filtration Equipment 4-28 

4.10 Radioactive and High Toxicity Operations 4-33 

4. 1 1 Explosion Venting 4-33 

References 4-34 

CHAPTERS EXHAUST SYSTEM DESIGN PROCEDURE 5-1 

5.1 Introduction 5-2 

5.2 Preliminary Steps 5-2 

5.3 Design Procedure 5-2 

5.4 Duct Segment Calculations 5-3 

5.5 Distribution of Air Flow 5-4 

5.6 Aids to Calculation 5-11 

5.7 Plenum Exhaust Systems 5-11 

5.8 Fan Pressure Calculations 5-11 

5.9 Corrections for Velocity Changes 5-12 

5.10 Sample System Design 5-13 

5.11 Different Duct Material Friction Losses 5-13 

5.12 Friction Loss for Non-Circular Ducts 5-13 

5.13 Corrections for Non-Standard Density 5-15 

5.14 Air Cleaning Equipment 5-32 

5.15 Evase' 5-32 

5.16 Exhaust Stack Outlets 5-33 

5.17 Air Bleed-Ins 5-35 

5.18 Optimum Economic Velocity 5-35 

5.19 Construction Guidelines for Local Exhaust Systems 5-35 

References 5-38 

CHAPTER 6 FANS 6-1 

6.1 Introduction 6-2 

6.2 Basic Definitions 6-2 

6.3 Fan Selection 6-6 

6.4 Fan Installation and Maintenance 6-21 

References 6-25 

CHAPTER 7 REPLACEMENT AND RECIRCULATED AIR 7-1 

7.1 Introduction 7-2 

7.2 Replacement Air . 7-2 

7.3 Replacement Air Distribution 7-4 

7.4 Replacement Air Flow Rate 7-5 



Contents 



7.5 Room Pressure 7-5 

7.6 Environmental Control 7-5 

7.7 Environmental Control Air Flow Rate 7-6 

7.8 Air Changes 7-6 

7.9 Air Supply Temperatures 7-6 

7.10 Air Supply Vs. Plant Heating Costs 7-8 

7. 1 1 Replacement Air Heating Equipment 7-9 

7.12 Cost of Heating Replacement Air 7-13 

7.13 Air Conservation 7-14 

7.14 Evaluation of Employee Exposure Levels 7-18 

References 7-20 

CHAPTER 8 VENTILATION ASPECTS OF INDOOR AIR QUALITY 8-1 

8.1 Introduction 8-2 

8.2 Dilution Ventilation for Indoor Air Quality 8-2 

8.3 HVAC Components and System Types 8-2 

8.4 HVAC Components, Functions and Malfunctions 8-4 

8.5 HVAC Component Survey Outline 8-8 

References 8-10 

CHAPTER 9 TESTING OF VENTILATION SYSTEMS 9-1 

9.1 Introduction 9-2 

9.2 Measurements of Volumetric Flow Rate 9-2 

9.3 Calibration of Air Measuring Instruments 9-6 

9.4 Pressure Measurement 9-11 

9.5 Pitot Traverse Method 9-15 

9.6 Corrections for Non-Standard Conditions 9-21 

9.7 Check-out Procedures 9-23 

References 9-27 

CHAPTER 10 SPECIFIC OPERATIONS 10-1 

BIBLIOGRAPHY 11-1 

APPENDICES 12-1 

A Threshold Limit Values for Chemical Substances in the Work 
Environment with Intended Changes for 1996-1997 

B Physical Constants/Conversion Factors 
INDEX I3-I 



This Edition is Dedicated in Memory of 

KNOWLTON J. CAPLAN, PE, C1H, CSP 
June 23, 1919 -April 11, 1997 



Knowlton J. Caplan, while at the Division of Occupational 
Health, Michigan Department of Health, supervised the 
preparation of a field manual on industrial ventilation. That 
manual became the basis of the first edition of Industrial 
Ventilation in 195 1 . For the next forty-six years, the Ventila- 
tion Committee has felt Caplan' s presence as we published 
the "Vent Manual." This 23 rd edition is no different. Although 
"Cap" has not been an active member of the Committee for 
the past eleven years, his presence was felt at almost every 
meeting. Frequently we punctuated discussions with a quota- 
tion from Cap or a reference to one of his published works. 
Because of his influence, we proudly dedicate this edition to 
Knowlton J. Caplan. 

During his 50-year career, Cap was a pioneer in the fields 
of industrial hygiene, industrial ventilation, and air pollution 
control. He conducted basic research on cyclone and fabric 
filter dust collectors and holds several patents for these de- 
vices. As an associate professor in the public health depart- 
ment of the University of Minnesota, Cap advised numerous 
Master's degree students in industrial hygiene, occupational 
health and air pollution control. He has been an instructor at 
the industrial ventilation conferences at Michigan State Uni- 
versity (thirty years) and the University of Washington (ten 
years). As an author of more than 70 technical papers, he was 
a frequent presenter at the American Industrial Hygiene Con- 
ference and the American Society of Heating Refrigeration 
and Air Conditioning Engineers (ASHRAE) meeting. In ad- 
dition he wrote chapters in Air Pollution by Stern, Industrial 
Hygiene and Toxicology by Patty, Uranium Production Tech- 
nology by Harrington and Rueble, and was the Associate 
Editor of Industrial Hygiene Aspects of Plant Operations: 
Volume 3 - Engineering Considerations in Equipment Selec- 



tion, Layout and Building Design by Crawley and Crawley. 

Besides his innovative ventilation design, Cap developed a 
method for testing laboratory fume hoods which won the Best 
Paper of the Year award of the Michigan Industrial Hygiene 
Society in 1982, which later became the basis for the 
ASHRAE Standard 110-1995, "Method of Testing Labora- 
tory Fume Hood Performance." Cap was a significant partici- 
pant in the development of the ANSI Standard Z9.5-1992, 
"American National Standard for Laboratory Ventilation." He 
was the first to employ "clean air islands" to supplement local 
exhaust ventilation where necessary. 

Cap has been active in several societies: ACGIH (Commit- 
tee on Industrial Ventilation), Air Pollution Control Associa- 
tion (Committee on Dust, Fume, and Mist Control), American 
Industrial Hygiene Association (Board of Directors, Air Pol- 
lution Control Committee), ASHRAE ( Industrial Ventila- 
tion, Industrial Process Air Cleaning), American National 
Standards Committee (Air Pollution Committee, Health and 
Safety Committee), American Board of Industrial Hygiene. 

Cap was born in St. Louis, Missouri. He earned his bache- 
lor's and master's degree in chemical engineering from Wash- 
ington University in the 1940s. He served in the 
Commissioned Corps of the U.S. Public Health Services. He 
worked as a chemical engineer and ventilation engineer at 
Ralston Purina Company and Mallinckrodt Chemical, Ura- 
nium Division. He also worked for the St Louis County 
Health Department and the Michigan Department of Health 
as an industrial hygienist. In addition, Cap did consulting 
work, primarily as a ventilation engineer for Industrial Health 
Engineering Associates (co-founder), Pace Incorporated, and 
Rust Environment and Infrastructure. 



FOREWORD 



Industrial Ventilation: A Manual of Recommended Practice 
is the outgrowth of years of experience by members of the 
ACGIH Industrial Ventilation Committee members and a com- 
pilation of research data and information on design, mainte- 
nance, and evaluation of industrial exhaust ventilation systems. 
The Manual attempts to present a logical method of designing 
and testing these systems. It has found wide acceptance as a guide 
for official agencies, as a standard for industrial ventilation 
designers, and as a textbook for industrial hygiene courses. 

The Manual is not intended to be used as law, but rather as a 
guide. Because of new information on industrial ventilation 
becoming available through research projects, reports from en- 
gineers, and articles in various periodicals and journals, review 
and revision of each section of the Manual is an ongoing Com- 
mittee project. The Manual is available as a hardbound book and 
on CD-ROM. In a constant effort to present the latest techniques 
and data, the Committee desires, welcomes, and actively seeks 
comments and suggestions on the accuracy and adequacy of the 
information presented herein. 

In this 23rd edition, the Committee has made a number of 
minor revisions. Chapter 5 includes updated duct calculation 
sheets designed to aid in calculations. The u 3 eye" duct friction 
charts have been replaced with tables to permit easier determi- 
nation of the duct friction factor. The metric supplement has been 



deleted and the Committee has developed a separate metric 
manual. 

This publication is designed to present accurate and 
authoritative information with regard to the subject matter 
covered. It is distributed with the understanding that nei- 
ther the Committee nor its members collectively or indi- 
vidually assume any responsibility for any inadvertent 
misinformation, omissions, or for the results in the use of 
this publication. 



COMMITTEE ON INDUSTRIAL VENTILATION 

R.T. Hughes, NIOSH, Ohio, Chair 

A.G. Apol, FEOH, Washington 

W.M. Geary, Retired, Michigan 

M.T. Davidson, The New York Blower Co., Indiana 

T.N. Do, NFESC, California 

Mrs. Norma Donovan, Editorial Consultant 

S.E. Guffey, U. of Washington, Washington 

G.S. Knutson, Knutson Ventilation Consultants, Minnesota 

G. Lanham, KBD/Technic, Ohio 

K. Mead, NIOSH, Ohio 

K.M. Paulson, NFESC, California 

O.P. Petrey, Phoenix Process Equipment Co., Kentucky 

A.L. Twombly, Pfeiffer Engineering Co. Inc., Kentucky 



ACKNOWLEDGMENTS 



Industrial Ventilation is a true Committee effort. It brings 
into focus in one source useful, practical ventilation data from 
all parts of the country. The Committee membership of indus- 
trial ventilation and industrial hygiene engineers represents a 
diversity of experience and interest that ensures a well- 
rounded, cooperative effort. 

From the First Edition in 1 95 1 , this effort has been success- 
ful as witnessed by the acceptance of the "Ventilation Man- 
ual" throughout industry, by governmental agencies, and as a 
worldwide reference and text. 

The present Committee is grateful for the faith and firm 
foundation provided by past Committees and members listed 
below. Special acknowledgment is made to the Division of 



Occupational Health, Michigan Department of Health, for 
contributing their original field manual which was the basis 
of the First Edition, and to Mr. Knowlton J. Caplan who 
supervised the preparation of that manual. 

The Committee is grateful also to those consultants who 
have contributed so greatly to the preparation of this and 
previous editions of Industrial Ventilation and to Mrs. Norma 
Donovan, Secretary to the Committee, for her untiring zeal in 
our efforts. 

To many other individuals and agencies who have made 
specific contributions and have provided support, sugges- 
tions, and constructive criticism, our special thanks. 



COMMITTEE ON INDUSTRIAL VENTILATION 



Previous Members 



A.G. Apol, 1984— present 

H.Ayer, 1962-1966 

RE. Bales, 1954-1960 

J. Baliff, 1950-1956; Chair, 1954-1956 

J.T. Barnhart, Consultant, 1986-1990 

J.C. Barrett, 1956-1976; Chair, 1960-1968 

J.L. Beltran, 1964-1966 

D. Bonn, Consultant, 1958-1968 

D.J. Burton, 1988-1970 

K.J. Caplan, 1974-1978; Consultant, 1980-1986 

W.M. Cleary. 1978-1993; Consultant, 1993-present; Chair, 

1978-1984 

L. Dickie, 1984-1994; Consultant 1968-1984 

B. Feiner, 1956-1968 

M. Franklin, 1991-1994 

S.E. Guffey, 1984-present 

G.M. Kama, 1950-1984; Chair, 1956-1960 

R.P, Hibbard, 1968-1994 

R.T. Hughes, 1976-present; Chair, 1989-present 

H.S.Jordan, 1960-1962 

J. Kane, Consultant, 1950-1952 



J. Kayse, Consultant, 1956-1958 

J.F.Keppler, 1950-1954, 1958-1960 

G.W. Knutson, Consultant, 1986-present 

J.J. Loeffler, 1980-1995; Chair, 1984-1989 

J.Lumsden, 1962-1968 

J.R. Lynch, 1966-1976 

G. Michaelson, 1958-1960 

K.M. Morse, 1950-1951; Chair, 1950-1951 

R.T.Page, 1954-1956 

K.M. Paulson, 1991 -present 

O.P. Petrey, Consultant, 1978-present 

G.S.Rajhans, 1978-1995 

K.E.Robinson, 1950-1954; Chair, 1952-1954 

A. Salazar, 1952-1954 

EL. Schall, 1956-1958 

M.M. Schuman, 1962-1994; Chair, 1968-1978 

J.C. Soet, 1950-1960 

A.L. Twombly, Consultant, 1986-present 

J. Willis, Consultant, 1952-1956 

R. Wolle, 1966-1974 

J.A. Wunderle, 1960-1964 



DEFINITIONS 



Aerosol: An assemblage of small particles, solid or liquid, 
suspended in air. The diameter of the particles may vary 
from 100 microns down to 0.01 micron or less, e.g., 
dust, fog, smoke. 

Air Cleaner: A device designed for the purpose of remov- 
ing atmospheric airborne impurities such as dusts, 
gases, vapors, fumes, and smoke. (Air cleaners include 
air washers, air filters, eletrostatic precipitators, and 
charcoal filters.) 

Air Filter, An air cleaning device to remove light particu- 
late loadings from normal atmospheric air before intro- 
duction into the building. Usual range: loadings up to 
3 grains per thousand cubic feet (0.003 grains per cubic 
foot). Note: Atmospheric air in heavy industrial areas 
and in-plant air in many industries have higher loadings 
than this, and dust collectors are then indicated for 
proper air cleaning. 

Air Horsepower. The theoretical horsepower required to 
drive a fan if there were no loses in the fan, that is, if its 
efficiency were 100 percent. 

Air, Standard: Dry air at 70 F and 29.92 in (Hg) barometer. 
This is substantially equivalent to 0.075 lb/ft3. Specific 
heat of dry air - 0.24 btu/lb/F. 

Aspect Ratio: The ratio of the width to the length; AR = 
W/L. 

Aspect Ratio of an Elbow: The width (W) along the axis 
of the bend divided by depth (D) in plane of bend; AR 
= W/D. 

Blast Gate: Sliding damper. 

Blow (throw): In air distribution, the distance an air stream 
travels from an outlet to a position at which air motion 
along the axis reduces to a velocity of 50 fpm. For unit 
heaters, the distance an air stream travels from a heater 
without a perceptible rise due to temperature difference 
and loss of velocity. 

Brake Horsepower: The horsepower actually required to 
drive a fan. This includes the energy losses in the fan 
and can be determined only by actual test of the fan. 
(This does not include the drive losses between motor 
and fan.) 

Capture Velocity: The air velocity at any point in front of 
the hood or at the hood opening necessary to overcome 
opposing air currents and to capture the contaminated 
air at that point by causing it to flow into the hood. 



Coefficient of Entry: The actual rate of flow caused by a 
given hood static pressure compared to the theoretical 
flow which would result if the static pressure could be 
converted to velocity pressure with 100 percent effi- 
ciency. It is the ratio of actual to theoretical flow. 

Comfort Zone (Average): The range of effective tempera- 
tures over which the majority (50% or more) of adults 
feel comfortable. 

Convection: The motion resulting in a fluid from the 
differences in density and the action of gravity. In heat 
transmission, this meaning has been extended to in- 
clude both forced and natural motion or circulation. 

Density: The ratio of the mass of a specimen of a substance 
to the volume of the specimen. The mass of a unit 
volume of a substance. When weight can be used with- 
out confusion, as synonymous with mass, density is the 
weight of a unit volume of a substance. 

Density Factor The ratio of actual air density to density 
of standard air. The product of the density factor and 
the density of standard air (0.075 lb/ft3) will give the 
actual air density in pounds per cubic foot; d x 0.075 = 
actual density of air, Ibs/ft3. 

Dust: Small solid particles created by the breaking up of 
larger particles by processes crushing, grinding, drill- 
ing, explosions, etc. Dust particles already in existence 
in a mixture of materials may escape into the air through 
such operations as shoveling, conveying, screening, 
sweeping, etc. 

Dust Collector An air cleaning device to remove heavy 
particulate loadings from exhaust systems before dis- 
charge to outdoors. Usual range: loadings 0.003 grains 
per cubic foot and higher. 

Entry Loss: Loss in pressure caused by air flowing into a 
duct or hood (inches H 2 0). 

Fumes: Small, solid particles formed by the condensation 
of vapors of solid materials. 

Gases: Formless fluids which tend to occupy an entire 
space uniformly at ordinary temperatures and pres- 
sures. 

Gravity, Specific: The ratio of the mass of a unit volume 
of a substance to the mass of the same volume of a 
standard substance at a standard temperature. Water at 
39.2 F is the standard substance usually referred to. For 
gases, dry air, at the same temperature and pressure as 
the gas, is often taken as the standard substance. 



xni 



XIV 



Industrial Ventilation 



Hood: A shaped inlet designed to capture contaminated air 
and conduct it into the exhaust duct system. 

Humidity, Absolute: The weight of water vapor per unit 
volume, pounds per cubic foot or grams per cubic 
centimeter. 

Humidity, Relative: The ratio of the actual partial pressure 
of the water vapor in a space to the saturation pressure 
of pure water at the same temperature. 

Inch of Water: A unit of pressure equal to the pressure 
exerted by a column of liquid water one inch high at a 
standard temperature. 

Lower Explosive Limit: The lower limit of flammability or 
explosibility of a gas or vapor at ordinary ambient 
temperatures expressed in percent of the gas or vapor in 
air by volume. This limit is assumed constant for tem- 
peratures up to 250 F. Above these temperatures, it 
should be decreased by a factor of 0.7 since explosibility 
increases with higher temperatures. 

Manometer: An instrument for measuring pressure; essen- 
tially a U-tube partially filled with a liquid, usually 
water, mercury or a light oil, so constructed that the 
amount of displacement of the liquid indicates the pres- 
sure being exerted on the instrument. 

Micron: A unit of length, the thousandth part of 1 mm or 
the millionth of a meter (approximately 1/25,000 of an 
inch). 

Minimum Design Duct Velocity: Minimum air velocity 
required to move the particulates in the air stream, fpm. 

Mists: Small droplets of materials that are ordinarily liquid 
at normal temperature and pressure. 

Plenum: Pressure equalizing chamber. 

Pressure, Static: The potential pressure exerted in all di- 
rections by a fluid at rest. For a fluid in motion, it is 
measured in a direction normal to the direction of flow. 
Usually expressed in inches water gauge when dealing 
with air. (The tendency to either burst or collapse the 
pipe.) 

Pressure, Total: The algebraic sum of the velocity pressure 
and the static pressure (with due regard to sign). 

Pressure, Vapor: The pressure exerted by a vapor. If a 
vapor is kept in confinement over its liquid so that the 
vapor can accumulate above the liquid, the temperature 
being held constant, the vapor pressure approaches a 
fixed limit called the maximum or saturated vapor pres- 



sure, dependent only on the temperature and the liquid. 
The term vapor pressure is sometimes used as synony- 
mous with saturated vapor pressure. 

Pressure, Velocity: The kinetic pressure in the direction of 
flow necessary to cause a fluid at rest to flow at a given 
velocity. Usually expressed in inches water gauge. 

Radiation, Thermal (Heat) Radiation: The transmission of 
energy by means of electromagnetic waves of very long 
wave length. Radiant energy of any wave length may, 
when absorbed, become thermal energy and result in an 
increase in the temperature of the absorbing body. 

Replacement Air: A ventilation term used to indicate the 
volume of controlled outdoor air supplied to a building 
to replace air being exhausted. 

Slot Velocity: Linear flow rate of contaminated air through 
slot, fpm. 

Smoke: An air suspension (aerosol) of particles, usually but 
not necessarily solid, often originating in a solid nu- 
cleus, formed from combustion or sublimation. 

Temperature, Effective: An arbitrary index which com- 
bines into a single value the effect of temperature, 
humidity, and air movement on the sensation of warmth 
or cold felt by the human body. The numerical value is 
that of the temperature of still, saturated air w r hich 
would induce an identical sensation. 

Temperature, Wet-Bulb: Thermodynamic wet- bulb tem- 
perature is the temperature at which liquid or solid 
water, by evaporating into air, can bring the air to 
saturation adiabatically at the same temperature. Wet- 
bulb temperature (without qualification) is the tempera- 
ture indicated by a wet-bulb psychrometer constructed 
and used according to specifications. 

Threshold Limit Values (TLVs): The values for airborne 
toxic materials which are to be used as guides in the 
control of health hazards and represent time-weighted 
concentrations to which nearly all workers may be 
exposed 8 hours per day over extended periods of time 
without adverse effects (see Appendix). 

Transport (Conveying) Velocity: See Minimum Design 
Duct Velocity. 

Vapor: The gaseous form of substances which are nor- 
mally in the solid or liquid state and which can be 
changed to these states either by increasing the pressure 
or decreasing the temperature. 



ABBREVIATIONS 



A area 

acfm flow rate at actual condition 

AHP air horsepower 

AR aspect ratio 

A s Slot area 

B barometric pressure 

blip brake horsepower 

bhp a brake horsepower, actual 

bhp s brake horsepower, standard air 

btu British thermal unit 

btuh btu/hr 

Ce coefficient of entry 

cfm cubic feet per minute 

CLR centerline radius 

D diameter 

df density factor 

ET effective temperature 

F degree, Fahrenheit 

F d duct entry loss coefficient 

F d elbow loss coefficient 

F cn entry loss coefficient 

fpm feet per minute 

fps feet per second 

F s slot loss coefficient 

ft 2 square foot 

ft 3 cubic foot 

g gravitational force, ft/sec/sec 

gpm gallons per minute 

gr grains 

h d duct entry loss 

h c overall hood entry loss 

h c! elbow loss 

h en entry loss 

HEP A high-efficiency particulate air filters 

h f loss in straight duct run 

H f duct loss coefficient 

hp horsepower 

hr hour 

h s slot or opening entry loss 

in inch 



in 2 square inch 

"wg inches water gauge 

lb pound 

lbm pound mass 

LEL lower explosive limit 

ME mechanical efficiency 

mg milligram 

min minute 

mm millimeter 

MRT mean radiant temperature 

MW molecular weight 

p density of air in lb/ft 3 

ppm parts per million 

psi pounds per square inch 

PWR power 

Q flow rate in cfm 

Q corr corrected flow rate at a junction 

R degree, Rankin 

RH relative humidity 

rpm revolutions per minute 

scfin flow rate at standard condition 

sfpm surface feet per minute 

sp gr specific gravity 

SP static pressure 

SP g0V .... higher static pressure at junction of 2 ducts 

SP h hood static pressure 

SP S Sp, system handling standard air 

STP standard temperature and pressure 

TLV Threshold Limit Value 

TP total pressure 

V velocity, fpm 

V d duct velocity 

VP velocity pressure 

VP d duct velocity pressure 

VP r resultant velocity power 

VP S slot velocity pressure 

V s Slot velocity 

V t duct transport velocity 

W watt 



Chapter 1 

GENERAL PRINCIPLES OF VENTILATION 



1.1 INTRODUCTION . 1-2 

1.2 SUPPLY SYSTEMS 1-2 

1.3 EXHAUST SYSTEMS 1-2 

1.4 BASIC DEFINITIONS 1-3 

1.5 PRINCIPLES OF AIR FLOW 1-4 

1 .6 ACCELERATION OF AIR AND HOOD ENTRY 

LOSSES 1-6 

Figure 1-1 SP, VP, and TP at a Point 1-4 

Figure 1-2 Measurement of SP, VP, and TP in a Pressurized 

Duct 1-4 

Figure 1-3 SP, VP, and TP at Points in a Ventilation System 1-5 

Figure 1 -4 Volumetric Flow Rates in Various Situations . . 1-5 



1.7 DUCT LOSSES 1-7 

1.7.1 Friction Losses 1-7 

1.7.2 Fitting Losses 1-9 

1.8 MULTIPLE-HOOD EXHAUST SYSTEMS 1-9 

1.9 AIR FLOW CHARACTERISTICS OF BLOWING 

AND EXHAUSTING 1-10 

REFERENCES 1-10 

Figure 1-5 Variation of SP, VP, and TP Through a Ventilation 

System 1-6 

Figure 1-6 Moody Diagram 1-8 

Figure 1-7 Blowing Vs. Exhausting 1-10 



1-2 



Industrial Ventilation 



1.1 INTRODUCTION 

The importance of clean uncontaminated air in the indus- 
trial work environment is well known. Modern industry with 
its complexity of operations and processes uses an increasing 
number of chemical compounds and substances, many of 
which are highly toxic. The use of such materials may result 
in particulates, gases, vapors, and/or mists in the workroom 
air in concentrations that exceed safe levels. Heat stress can 
also result in unsafe or uncomfortable work environments. 
Effective, well-designed ventilation offers a solution to these 
problems where worker protection is needed. Ventilation can 
also serve to control odor, moisture, and other undesirable 
environmental conditions. 

The health hazard potential of an airborne substance is 
characterized by the Threshold Limit Value (TLV* J ), The TLV 
refers to the airborne concentration of a substance and repre- 
sents conditions under which it is believed that nearly all 
workers may be exposed day after day without adverse health 
effects. The time-weighted average (TWA) is defined as the 
time-weighted average concentration for a conventional 8- 
hour workday and a 40-hour workweek which will produce 
no adverse health effects for nearly all workers. The 
TLV-TWA is usually used to determine a safe exposure level. 
TLVs are published annually by the American Conference of 
Governmental Industrial Hygienists (ACGIH); revisions and 
additions are made regularly as information becomes avail- 
able. Appendix A of this Manual provides the current TLV 
list for chemical substances as of the date of publication. 

Ventilation systems used in industrial plants are of two 
generic types. The SUPPLY system is used to supply air, 
usually tempered, to a work space. The EXHAUST system is 
used to remove the contaminants generated by an operation 
in order to maintain a healthful work environment. 

A complete ventilation program must consider both the 
supply and the exhaust systems. If the overall quantity of air 
exhausted from a work space is greater than the quantity of 
outdoor air supplied to the space, the plant interior will 
experience a lower pressure than the local atmospheric pres- 
sure. This may be desirable when using a dilution ventilation 
system to control or isolate contaminants in a specific area of 
the overall plant. Often, this condition occurs simply because 
local exhaust systems are installed and consideration is not 
given to the corresponding replacement air systems. Air will 
then enter the plant in an uncontrolled manner through cracks, 
walls, windows, and doorways. This typically results in: 1) 
employee discomfort in winter months for those working near 
the plant perimeter, 2) exhaust system performance degrada- 
tion, possibly leading to loss of contaminant control and a 
potential health hazard, and 3) higher heating and cooling costs. 
Chapter 7 of this Manual discusses these points in more detail. 

1.2 SUPPLY SYSTEMS 

Supply systems are used for two purposes: 1) to create a 



comfortable environment in the plant (the HVAC system); 
and 2) to replace air exhausted from the plant (the REPLACE- 
MENT system). Many times, supply and exhaust systems are 
coupled, as in dilution control systems (see Section 1 .3 and 
Chapter 2.) 

A well-designed supply system will consist of an air inlet 
section, filters, heating and/or cooling equipment, a fan, ducts, 
and register/grilles for distributing the air within the work 
space. The filters, heating and/or cooling equipment and fan 
are often combined into a complete unit called an airhouse or 
air supply unit. If part of the air supplied by a system is 
recirculated, a RETURN system is used to bring the air back 
to the airhouse. 

1.3 EXHAUST SYSTEMS 

Exhaust ventilation systems are classified in two generic 
groups: 1) the GENERAL exhaust system and 2) the LOCAL 
exhaust system. 

The general exhaust system can be used for heat control 
and/or removal of contaminants generated in a space by 
flushing out a given space with large quantities of air. When 
used for heat control, the air may be tempered and recycled. 
When used for contaminant control (the dilution system), 
enough outdoor air must be mixed with the contaminant so 
that the average concentration is reduced to a safe level. The 
contaminated air is then typically discharged to the atmos- 
phere. A supply system is usually used in conjunction with a 
general exhaust system to replace the air exhausted. 

Dilution ventilation systems are normally used for con- 
taminant control only when local exhaust is impractical, as 
the large quantities of tempered replacement air required to 
offset the air exhausted can lead to high operating costs. 
Chapter 2 describes the basic features of general ventilation 
systems and their application to contaminant and fire hazard 
control. 

Local exhaust ventilation systems operate on the principle 
of capturing a contaminant at or near its source. It is the 
preferred method of control because it is more effective and 
the smaller exhaust flow rate results in lower heating costs 
compared to high flow rate general exhaust requirements. The 
present emphasis on air pollution control stresses the need for 
efficient air cleaning devices on industrial ventilation sys- 
tems, and the smaller flow rates of the local exhaust system 
result in lower costs for air cleaning devices. 

Local exhaust systems are comprised of up to four basic 
elements: the hood(s), the duct system (including the exhaust 
stack and/or recirculation duct), the air cleaning device, and 
the fan. The purpose of the hood is to collect the contaminant 
generated in an air stream directed toward the hood. A duct 
system must then transport the contaminated air to the air 
cleaning device, if present, or to the fan. In the air cleaner, the 
contaminant is removed from the air stream. The fan must 
overcome all the losses due to friction, hood entry, and fittings 



General Principles of Ventilation 



1-3 



in the system while producing the intended flow rate. The duct 
on the fan outlet usually discharges the air to the atmosphere 
in such a way that it will not be re-entrained by the replace- 
ment and/or HVAC systems. In some situations, the cleaned 
air is returned to the plant. Chapter 7 discusses whether this 
is possible and how it may be accomplished. 

This Manual deals with the design aspects of exhaust 
ventilation systems, but the principles described also apply to 
supply systems. 

1.4 BASIC DEFINITIONS 

The following basic definitions are used to describe air flow 
and will be used extensively in the remainder of the Manual. 

The density (p) of the air is defined as its mass per unit 
volume and is normally expressed in pounds mass per cubic 
foot (lbm/ft 3 ). At standard atmospheric pressure (14.7 psia), 
room temperature (70 F) and zero water content, its value is 
normally taken to be 0.075 lbm/ft 3 , as calculated from the 
perfect gas equation of state relating pressure, density, and 
temperature: 



p-pRT 
where: 



[1.1] 

p = the absolute pressure in pounds per square foot 

absolute (psfa) 
p = the density, lbm/ft 3 
R = the gas constant for air and equals 53.35 ft- 

lb/lbm-degrees Rankine 
T = the absolute temperature of the air in degrees 

Rankine 

Note that degrees Rankine = degrees Fahrenheit + 459.7. 

From the above equation, density varies inversely with 
temperature when pressure is held constant. Therefore, for 
any dry air situation (see Chapter 5 for moist air calculations), 

pT-(pT) STD 
or 



Q = volumetric flow rate, cfm 
V = average velocity, fpm 

A = cross-sectional area, ft 2 

Given any two of these three quantities, the third can readily 
be determined. 

Air or any other fluid will always flow from a region of 
higher total pressure to a region of lower total pressure in the 
absence of work addition (a fan). There are three different but 
mathematically related pressures associated with a moving air 
stream. 

Static pressure (SP) is defined as the pressure in the duct 
that tends to burst or collapse the duct and is expressed in 
inches of water gage ( M wg). It is usually measured with a water 
manometer, hence the units. SP can be positive or negative 
with respect to the local atmospheric pressure but must be 
measured perpendicular to the air flow. The holes in the side 
of a Pitot tube (see Figure 9-9) or a small hole carefully drilled 
to avoid internal burrs that disturb the air flow (never 
punched) into the side of a duct will yield SP. 

Velocity pressure (VP) is defined as that pressure required 
to accelerate air from zero velocity to some velocity (V) and 
is proportional to the kinetic energy of the air stream. The 
relationship between V and VP is given by 

V = 1096 



or 



VP = P 



V 



1096 



[1.4] 



P = PSTD 



^ = 0.075^ 



[1.2] 



where: 

V = velocity, fpm 
VP = velocity pressure, "wg 

If standard air is assumed to exist in the duct with a density 
of 0.075 lbm/ft 3 , this equation reduces to 

V = 4005VVP 
or 



For example, the density of dry air at 250 F would be 



p = 0.075 



530 



460 + 250 



-0.056 lbm/ft 3 



The volumetric flow rate, many times referred to as "vol- 
umes," is defined as the volume or quantity of air that passes 
a given location per unit of time. It is related to the average 
velocity and the flow cross-sectional area by the equation 



Q-VA 
where; 



[1.3] 



VP: 



V 



4005 



[1.5] 



VP will only be exerted in the direction of air flow and is 
always positive. Figure 1-1 shows graphically the difference 
between SP and VP. 

Total pressure (TP) is defined as the algebraic sum of the 
static and velocity pressures or 

TP = SP + VP [1.6] 

Total pressure can be positive or negative with respect to 



1-4 



Industrial Ventilation 




FIGURE 1-1 . SP, VP, and TP at a point 

atmospheric pressure and is a measure of the energy content 
of the air stream, always dropping as the flow proceeds 
downstream through a duct. The only place it will rise is 
across the fan. 

Total pressure can be measured with an impact tube point- 
ing directly upstream and connected to a manometer. It will 
vary across a duct due to the change of velocity across a duct 
and therefore single readings of TP will not be representative 
of the energy content. Chapter 9 illustrates procedures for 
measurement of all pressures in a duct system. 

The significance of these pressures can be illustrated as 
follows. Assume a duct segment with both ends sealed was 
pressurized to a static pressure of 0.1 psi above the atmos- 
pheric pressure as shown in Figure 1-2. If a small hole 
(typically 1/16" to 3/32") were drilled into the duct wall and 
connected to one side ofa U-tube manometer, the reading 
would be approximately 2.77 "wg. Note the way the left-hand 
manometer is deflected. If the water in the side of the ma- 
nometer exposed to the atmosphere is higher than the water 
level in the side connected to the duct, then the pressure read 
by the gauge is positive (greater than atmospheric). Because 
there is no velocity, the velocity pressure is and SP = TP. A 
probe which faces the flow is called an impact tube and will 



measure TP. In this example, a manometer connected to an 
impact tube (the one on the right) will also read 2.77 "wg. 
Finally, if one side of a manometer were connected to the 
impact tube and the other side were connected to the static 
pressure opening (the center one), the manometer would read 
the difference between the two pressures. As VP - TP — SP, 
a manometer so connected would read VP directly. In this 
example, there is no flow and hence VP = as indicated by 
the lack of manometer deflection. 

If the duct ends were removed and a fan placed midway in 
the duct, the situation might change to the one shown on 
Figure 1-3. Upstream of the fan, SP and TP are negative (less 
than atmospheric). This is called the suction side. Down- 
stream of the fan, both SP and TP are positive. This is called 
the pressure side. Regardless of which side of the fan is 
considered, VP is always positive. Note that the direction in 
which the manometers are deflected shows whether SP and 
TP are positive or negative with respect to the local atmos- 
pheric pressure. 

1.5 PRINCIPLES OF AIR FLOW 

Two basic principles of fluid mechanics govern the flow of 
air in industrial ventilation systems: conservation of mass and 
conservation of energy. These are essentially bookkeeping 
laws which state that all mass and all energy must be com- 
pletely accounted for. A coverage of fluid mechanics is not in 
the purview of this manual; reference to any standard fluid 
mechanics textbook will show the derivation of these princi- 
ples. However, it is important to know what simplifying 
assumptions are included in the principles discussed below. 
They include: 

1. Heat transfer effects are neglected. If the temperature 
inside the duct is significantly different than the air 
temperature surrounding the duct, heat transfer will 
occur. This will lead to changes in the duct air tem- 
perature and hence in the volumetric flow rate. 

2. Compressibility effects are neglected. If the overall 
pressure drop from the start of the system to the fan is 
greater than about 20 "wg, then the density will change 




FIGURE 1-2. Measurement of SP, VP, and TP in a pressurized duct 



Genera! Principles of Ventilation 



1-5 



SUCTION S 



PRESSURE SIDE 



3000 fo 




( J 3000 forn 

01© 



F 

A 
N 



SP 

T 



± 
VP 
7" 



TF 



SP + VP = TP 

-1.1 + 0.56 - -0.54 

RESSURES BELOW 
ATMOSPHERIC 



FIGURE 1-3. SP, VP, and TP at points in a ventilation system 

change (see Chapter 5). 

3. The air is assumed to be dry. Water vapor in the air 
stream will lower the air density, and correction for 
this effect, if present, should be made. Chapter 5 
describes the necessary psychrometric analysis. 

4. The weight and volume of the contaminant in the air 
stream is ignored. This is permissible for the contami- 
nant concentrations in typical exhaust ventilation sys- 
tems. For high concentrations of solids or significant 
amounts of gases other than air, corrections for this 
effect should be included. 

Conservation of mass requires that the net change of mass 
flow rate must be zero. If the effects discussed above are 
negligible, then the density will be constant and the net change 
of volumetric flow rate (Q) must be zero. Therefore, the flow 
rate that enters a hood must be the same as the flow rate that 
passes through the duct leading from the hood. At a branch 
entry (converging wye) fitting, the sum of the two flow rates 
that enter the fitting must leave it. At a diverging wye, the 
flow rate entering the wye must equal the sum of the flow 
rates that leave it. Figure 1-4 illustrates these concepts. 

Conservation of energy means that all energy changes must 
be accounted for as air flows from one point to another. In 
terms of the pressures previously defined, this principle can 
be expressed as: 



TP 1 -TP 2 +h 1 



TP 

0. 76 



SP + VP 

0.20 + 0.56 

PRESSURES ABOVE 
ATMOSPHERIC 



© © 



X 



/* 



^ 



50 
fpm 



\ 
/ 

v 



3000 


fpm 










t u- 





© 



a. Qir, 02 



C"^\ 




© 



b. Q1 +02 = 03 



FIGURE 1-4. Volumetric flow rates in various situations, a. Flow through a 
hood; b. Flow through a branch entry 



1-6 



Industrial Ventilation 



C7> 



(./") 



cr> 



(1... 




- o 

_ 4 



FIGURE 1-5. Variation of SP, VP, and TP through a ventilation system 

SP^VP^SPz+VRj + h-, [1.7] 

where: 

subscript 1 = some upstream point 

subscript 2 = some downstream point 

h 1 = al l energy losses encountered by the air as 
it flows from the upstream to the down- 
stream point 

Note that, according to this principle, the total pressure 
must fall in the direction of flow. 

The application of these principles will be demonstrated by 







_. ..___ 












j I i i 


! i ! i 












I 


i 


; 


^ .... / . .,/_/ - 


/ 


./ 




..' /.. / .. JL _Z / ..A - 


^^f 


^Z..zl..Z 

_ J_ _ 


y //// 


kZkzk/Z-L 

i i 




i 











! i 


i 


\ 


! i 




. „i 








\ 


J I _ 



an analysis of the simple system shown in Figure 1-5. The 
normally vertical exhaust stack is shown laying horizontally 
to facilitate graphing the variation of static, total, and velocity 
pressures. The grinder wheel hood requires 300 cfin and the 
duct diameter is constant at 3.5 inches (0.0668 ft 2 area). 

1 .6 ACCELERATION OF AIR AND HOOD ENTRY 
LOSSES 



Air flows from the room (point 1 of Figure 1-5) through 
the hood to the duct (point 2 of Figure 1 -5) where the velocity 

can be calculated by the basic equation: 



General Principles of Ventilation 



1-7 



_ Q = 300 
A 0.0668 



:4490fpm 



This velocity corresponds to a velocity pressure of 1.26 "wg, 
assuming standard air. 

If there are no losses associated with entry into a hood, then 
applying the energy conservation principle (Equation 1.7) to 
the hood yields 



S^ + VP^ 



SP 2 +VP 2 



This is the well known Bernoulli principle of fluid mechan- 
ics. Subscript 1 refers to the room conditions where the static 
pressure is atmospheric (SPj - 0) and the air velocity is 
assumed to be very close to zero (VPi = 0). Therefore, the 
energy principle yields 

SP 2 = -VP 2 =-1.26 "wg 

Even if there were no losses, the static pressure must decrease 
due to the acceleration of air to the duct velocity. 

In reality, there are losses as the air enters the hood. These 
hood entry losses (h d ) are normally expressed as a loss coef- 
ficient (F d ) multiplied by the duct velocity pressure; so h d = 
FdVPcj (where VP d = VP 2 ). The energy conservation principle 
then becomes 

SP 2 --(VP 2 +h d ) [1-8] 

(See 3.5.1, 3,5.2, and Figure 5-1 for a discussion of h d and h e .) 

The absolute value of SP 2 is known as the hood static 
suction (SP h ). Then 

= VR 



[1.9] 



SP h - -SP 2 - vr 2 - 

For the example in Figure 1-5, assuming an entry loss coeffi- 
cient of 0.40, 

SP h = VP 2 + F d VP 2 

= 1.26 + (0.40) (1.26) 

= 1.26 + 0.50= 1.76 M wg 

In summary, the static pressure downstream of the hood is 
negative (less than atmospheric) due to two effects: 

1. Acceleration of air to the duct velocity; and 

2. Hood entry losses. 

From the graph, note that TP 2 = - h e , which confirms the 
premise that total pressure decreases in the flow direction. 

An alternate method of describing hood entry losses is by 
the hood entry coefficient (C e ). This coefficient is defined as 
the square root of the ratio of duct velocity pressure to hood 
static suction, or 



fyp" 
: VSP h 



[1.10] 



If there were no losses, then SP h = VP and C e = 1 .00. However, 
as hoods always have some losses, C e is always less than 1 .00. 
In Figure 1-5, 



VP 
SP h 



1.26 
1.76 



: 0.845 



An important feature of C e is that it is a constant for any given 
hood. It can, therefore, be used to determine the flow rate if 
the hood static suction is known. This is because 



Q = VA = 1096A /— - 1096A C '^i 



[1.11] 

P V P 

For standard air, this equation becomes 

Q = 4005AC 6 VSP^ [1-12| 

For the example in Figure 1-5, 

Q = 4005(0.0668)(0.845) Vt76 = 300 cfm 

By use of C e and a measurement of SP h , the flow rate of a 
hood can be quickly determined and corrective action can be 
taken if the calculated flow rate does not agree with the design 
flow rate. 

1.7 DUCT LOSSES 

There are two components to the overall total pressure 
losses in a duct run: 1) friction losses and 2) fitting losses. 

1.7.1 Friction Losses. Losses due to friction in ducts are 
a complicated function of duct velocity, duct diameter, air 
density, air viscosity, and duct surface roughness. The effects 
of velocity, diameter, density, and viscosity are combined into 
the Reynolds number (RJ, as given by 



FL = 



where: 



pdv 



[1.13] 



p = density, lbm/ft 3 

d = diameter, ft 

v = velocity, ft/sec 

[i = the air viscosity, lbm/s-ft 

The effect of surface roughness is typically given by the 
relative roughness, which is the ratio of the absolute surface 
roughness height (k), defined as the average height of the 
roughness elements on a particular type of material, to the 
duct diameter. Some standard values of absolute surface 
roughness used in ventilation systems are given in Table 1 - 1 . 

L. F. Moody (u ) combined these effects into a single chart 
commonly called the Moody diagram (see Figure 1-6). With 
a knowledge of both the Reynolds number and the relative 
roughness, the friction coefficient (f) y can be found. 



1-8 



Industrial Ventilation 



TABLE 1-1. Absolute Surface Roughness 



Duct Material 



Galvanized metal 

Biack iron 

Aluminum 

Stainless steel 

Flexible duct 
(wires exposed) 

Flexible duct 
(wires covered) 



Surface Roughness (k), feet 



0.00055 
0.00015 
0.00015 
0.00015 
0.01005 

0.00301 



The above roughness heights are design values. It should be noted that 
significant variations from these values may occur, depending on the 
manufacturing process. 



sionless) 
L = duct length, ft 
d = duct diameter, ft 
VP = duct velocity pressure, "wg 

There are many equations available for computer solutions 
to the Moody diagram. One of these is that of Churchill, 2) 
which gives accurate (to within a few percent) results over the 
entire range of laminar, critical, and turbulent flow, all in a 
single equation. This equation is: 



f = 8 



where: 



,ReJ 



12 



1/12 



+ (A + B) 



-3/2 



[1.15] 



Once determined, the friction coefficient is used in the 
Darcy-Weisbach friction coefficient equation to determine 
the overall duct friction losses: 



h f = f-VP 



[1.14] 



where: 



h f = friction losses in a duct, "wg 
f= Moody diagram friction coefficient (dimen- 



A = 4-2.457 In 



B: 



3.7D 



■(^f 



While useful, this equation is quite difficult to use without 
a computer. Several attempts have been made to simplify the 
determination of friction losses for specialized situations. For 



w|o 




OOO.OI 



REYNOLDS NUMBER Re = -8231 

FIGURE 1-6. Moody diagram (adapted from reference 1 .1) 



^qoo/^oqoos 



General Principles of Ventilation 



1-9 



many years, charts based on the Wright (l 3) equation have been 
used in ventilation system design: 



TABLE 1-2. Correlation Equation Constants 



:2.74 



(V/10 00) 

D 1 - 



,1.9 



[1.16] 



where: 



V = duct velocity, fpm 
D = duct diameter, inches 

This equation gives the friction losses, expressed as "wg per 
100 feet of pipe, for standard air of 0.075 lbm/ft 3 density 
flowing through average, clean, round galvanized pipe having 
approximately 40 slip joints per 100 feet (k = 0.0005 ft). 

The later work by Loeffler^ 1 4) presented equations for use 
in the "velocity pressure" calculation method. Using the 
standard values of surface roughness, equations were ob- 
tained that could be used with the Darcy-Weisbach equation 
in the form: 



■k: 



L VP - H f L VP 



[117] 



where the "12" is used to convert the diameter D in inches to 
feet. 

Simplified equations were determined for the flow of 
standard air through various types of duct material with good 
accuracy (less than 5% error). The equations thus resulting 
were: 



H f = 12i^ 
f D Q c 



[118] 



where the constant "a" and the exponents "b" and "c" vary 
as a function of the duct material as shown in Table 1-2. 
Note that no correlation was made with the extremely rough 
flexible duct with wires exposed. This equation, using the 
constants from Table 1-2 for galvanized sheet duct, were 
used to develop the friction Tables 5-5 and 5-6. Note that 
the value obtained from the chart or from equation 1.18 must 
be multiplied by both the length of duct and the velocity 
pressure. 

1.7.2 Fitting Losses. The fittings (elbows, entries, etc.) in 
a duct run will also produce a loss in total pressure. These 
losses are given in Chapter 5. 

The fitting losses are given by a loss coefficient (F) multi- 
plied by the duct velocity pressure. Thus, 

h en=F en VP [ 119 ] 

In contractions, entries, or expansions, there are several dif- 
ferent velocity pressures. The proper one to use with the loss 
coefficient will be identified where the coefficients are 
listed. 

In Figure 1-5, 15 feet of straight, constant diameter gal va- 



Duct Material 


k, Ft 


a 


b 


c 


Aluminum, black iron, 


0.00015 


0.0425 


0.465 


0.602 


stainless steel 










Galvanized sheet duct 


0.00051 


0.0307 


0.533 


0.612 


Flexible duct, fabric 


0.0035 


0.0311 


0.604 


0.639 


wires covered 











nized duct connects the hood to a fan inlet. Because the duct 
area is constant, the velocity, and therefore the velocity pres- 
sure, is also constant for any given flow rate. The energy 
principle is: 

SP 2 +VP 2 -SP 3 +VP 3 +h f 

where subscript 3 refers to the fan inlet location. Because VP 2 
= VP 3 , the losses will appear as a reduction in static pressure 
(there will, of course, be a corresponding reduction in total 
pressure). The friction loss can be found from Equation 1.17 
with the aid of Equation 1.18: 



/0.533 



H f = 0.0307 



Q 



0.612 



- 0.0307 



4490 Q 



300 u 



= 0.0828 



From Equation 1.17, h f = (0.0828)(15)(1.26) = 1.56 "wg. 
Using this in the energy principle, 



SP, = SP> 



-1.76 "wg- 1.56 "wg = -3.32 "wg 



Another 10 feet of straight duct is connected to the dis- 
charge side of the fan. The losses from the fan to the end of 
the system would be about 1.04 "wg. Because the static 
pressure at the end of the duct must be atmospheric (SP 5 = 0), 
the energy principle results in 

SP 4 =SP 5 +h f =0 "wg + 1.04 "wg = 1.04 "wg 

Therefore, the static pressure at the fan outlet must be 
higher than atmospheric by an amount equal to the losses in 
the discharge duct. 

1.8 MULTIPLE-HOOD EXHAUST SYSTEMS 

Most exhaust systems are more complicated than the pre- 
ceding example. It is usually more economical to purchase a 
single fan and air cleaner to service a series of similar opera- 
tions than to create a complete system for each operation. For 
example, the exhaust from 10 continuously used grinders can 
be combined into a single flow which leads to a common air 
cleaner and fan. This situation is handled similarly to a simple 
system, but with some provision to ensure that the air flow 
from each hood is as desired (see Chapter 5). 



1-10 



Industrial Ventilation 



30 D ■ 



r AN 



4-000 FPM AiR 
VELOCITY AT 
FACF OF BOTH 



BLOWING 

ana:::::" 



400 FPM 



"f 



/ 



APPROXIMATE! Y 10% OF PACT VD 
AT 30 DIA, AWAY FROM PRESSURE 
J FT OPFNING 



EXHAUSTING 



APPROXIMATELY 10% OF FACE VELOCITY 
AT ONE DIA. AWAY FROM EXHAUST 
OPENING. 



FIGURE 1-7. Blowing vs. exhausting 



1.9 AIR FLOW CHARACTERISTICS OF BLOWING AND 
EXHAUSTING 

Air blown from a small opening retains its directional effect 
for a considerable distance beyond the plane of the opening. 
However, if the flow of air through the same opening were 
reversed so that it operated as an exhaust opening handling 
the same volumetric flow rate, the flow would become almost 
non-directional and its range of influence would be greatly 
reduced. For this reason, local exhaust must not be contem- 
plated for any process that cannot be conducted in the imme- 
diate vicinity of the hood. Also, because of this effect, every 
effort should be made to enclose the operation as much as 
possible. Figure 1-7 illustrates the fundamental difference 
between blowing and exhausting. 

This effect also shows how the supply or replacement air 
discharge grilles can influence an exhaust system. If care is 



not taken, the discharge pattern from a supply grille could 
seriously affect the flow pattern in front of an exhaust hood. 

REFERENCES 

1.1. Moody, L.F.: Friction Factors for Pipe Flow. ASME 

Trans. 66:672 (1944). 

1.2. Churchill, S.W.: Friction Factor Equation Spans All 
Fluid Flow Regimes. Chemical Engineering, Vol. 84 

(1977). 

1.3. Wright, Jr., D.K.: A new Friction Chart for Round 
Ducts. ASHVE Trans., Vol. 51, Appendix 1, p. 312 
(1945). 

1.4. Loeffler, J.J.: Simplified Equations for HVAC Duct 
Friction Factors. ASHRAE J., p. 76 (January 1980). 



Chapter 2 

GENERAL INDUSTRIAL VENTILATION 



2.1 INTRODUCTION 2-2 

2.2 DILUTION VENTILATION PRINCIPLES 2-2 

2.3 DILUTION VENTILATION FOR HEALTH 2-2 

2.3.1 General Dilution Ventilation Equation 2-2 

2.3.2 Calculating Dilution Ventilation for 

Steady State Concentration 2-5 

2.3.3 Contaminant Concentration Buildup 2-5 

2.3.4 Rate of Purging 2-6 

2.4 MIXTURES— DILUTION VENTILATION 

FOR HEALTH 2-6 

2.5 DILUTION VENTILATION FOR FIRE 

AND EXPLOSION 2-7 

2.6 FIRE DILUTION VENTILATION FOR MIXTURES 2-8 

2.7 VENTILATION FOR HEAT CONTROL 2-8 

2.8 HEAT BALANCE AND EXCHANGE 2-8 

2.8.1 Convection 2-9 

2.8.2 Radiation 2-9 

2.8.3 Evaporation 2-9 

2.9 ADAPTIVE MECHANISM OF THE BODY 2-9 

2.10 ACCLIMATIZATION 2-10 

Figure 2-1 "K" Factors Suggested for Inlet and Exhaust 

Locations 2-4 

Figure 2-2 Contaminant Concentration Buildup 2-6 

Figure 2-3 Rate of Purging 2-6 

Figure 2-4 Heat Losses, Storage, and Temperature Relations 2-10 

Figure 2-5 Determination of Wet Bulb Globe Temperature 2-1 1 

Figure 2-6 Recommended Heat-Stress Alert Limits 2-13 



2.11 ACUTE HEAT DISORDERS 2-10 

2.11.1 Heatstroke 2-10 

2.11.2 Heat Exhaustion 2-10 

2.11.3 Heat Cramps and Heat Rash 2-11 

2. 1 2 ASSESSMENT OF HEAT STRESS AND HEAT 

STRAIN 2-11 

2.12.1 Evaluation of Heat Stress 2-11 

2.12.2 Evaluation of Heat Strain 2-12 

2.13 WORKER PROTECTION 2-13 

2.14 VENTILATION CONTROL 2-13 

2.15 VENTILATION SYSTEMS 2-13 

2.16 VELOCITY COOLING 2-15 

2.17 RADIANT HEAT CONTROL 2-15 

2. 1 8 PROTECTIVE SUITS FOR SHORT EXPOSURES 2- 1 6 

2.19 RESPIRATORY HEAT EXCHANGERS 2-16 

2.20 REFRIGERATED SUITS 2-16 

2.21 ENCLOSURES 2-16 

2.22 INSULATION 2-16 

REFERENCES 2-17 



Figure 2-7 Recommended Heat-Stress Exposure Limits . . 2-14 

Figure 2-8 Natural Ventilation 2-14 

Figure 2-9 Mechanical Ventilation 2-14 

Figure 2-10 Spot Cooling With Volume and Directional 

Control 2-16 

Figure 2-11 Heat Shielding 2-16 



2-2 



Industrial Ventilation 



2.1 INTRODUCTION 

"General industrial ventilation" is abroad term which refers 
to the supply and exhaust of air with respect to an area, room, 
or building. It can be divided further into specific functions 
as follows: 

1 . Dilution Ventilation is the dilution of contaminated air 
with un contaminated air for the purpose of controlling 
potential airborne health hazards, fire and explosive 
conditions, odors, and nuisance-type contaminants. 
Dilution ventilation also can include the control of 
airborne contaminants (vapors, gases, and particu- 
lates) generated within tight buildings. 

Dilution ventilation is not as satisfactory for health 
hazard control as is local exhaust ventilation. Circum- 
stances may be found in which dilution ventilation 
provides an adequate amount of control more eco- 
nomically than a local exhaust system. One should be 
careful, however, not to base the economical consid- 
erations entirely upon the first cost of the system since 
dilution ventilation frequently exhausts large amounts 
of heat from a building, which may greatly increase 
the energy cost of the operation. 

2. Heat Control Ventilation is the control of indoor at- 
mospheric conditions associated with hot industrial 
environments such as are found in foundries, laun- 
dries, bakeries, etc., for the purpose of preventing 
acute discomfort or injury. 

2.2 DILUTION VENTILATION PRINCIPLES 

The principles of dilution ventilation system design are as 
follows: 

L Select from available data the amount of air required 
for satisfactory dilution of the contaminant. The values 
tabulated on Table 2-1 assume perfect distribution and 
dilution of the air and solvent vapors. These values 
must be multiplied by the selected K value (see Section 
2.3.1). 

2. Locate the exhaust openings near the sources of con- 
tamination, if possible, in order to obtain the benefit 
of "spot ventilation." 

3. Locate the air supply and exhaust outlets such that the 
air passes through the zone of contamination. The 
operator should remain between the air supply and the 
source of the contaminant. 

4. Replace exhausted air by use of a replacement air 
system. This replacement air should be heated during 
cold weather. Dilution ventilation systems usually 
handle large quantities of air by means of low pressure 
fans. Replacement air must be provided if the system 
is to operate satisfactorily. 

5. Avoid re-entry of the exhausted air by discharging the 



exhaust high above the roof line or by assuring that no 
window, outdoor air intakes, or other such openings 
are located near the exhaust discharge. 

2.3 DILUTION VENTILATION FOR HEALTH 

The use of dilution ventilation for health has four limiting 
factors: 1) the quantity of contaminant generated must not be 
too great or the air flow rate necessary for dilution will be 
impractical; 2) workers must be far enough away from the 
contaminant source or the evolution of contaminant must be 
in sufficiently low concentrations so that workers will not 
have an exposure in excess of the established TLV; 3) the 
toxicity of the contaminant must be low; and 4) the evolution 
of contaminants must be reasonably uniform. 

Dilution ventilation is used most often to control the vapors 
from organic liquids with a TLV of 100 ppm or higher. In 
order to successfully apply the principles of dilution to such 
a problem, factual data are needed on the rate of vapor 
generation or on the rate of liquid evaporation. Usually such 
data can be obtained from the plant if any type of adequate 
records on material consumption are kept. 

2.3.1 General Dilution Ventilation Equation: The venti- 
lation rate needed to maintain a constant concentration at a 
uniform generation rate is derived by starting with a funda- 
mental material balance and assuming no contaminant in the 
air supply, 



Rate of Accumulation = 



: Rate of Generation - 
Rate of Removal 



or 



VdC = Gdt- Q'Cdt 

where: 

V= volume of room 

G = rate of generation 

Q - effective volumetric flow rate 

C = concentration of gas or vapor 
t= time 
At a steady state, dC = 

Gdt = Q'Cdt 



[2.1] 



P 2 P 2 

Gdt= Q'Cdt 

iu Jti 



At a constant concentration, C, and uniform generation rate, G, 
G(t 2 -t 1 ) = Q'C(t 2 -t 1 ) 



«-% 



[2.2] 



Due to incomplete mixing, a K value is introduced to the rate 
of ventilation; thus: 



General Industrial Ventilation 



2-3 



TABLE 2-1 . Dilution Air Volumes for Vapors 



The following values are tabulated using the TLV values shown in parentheses, parts per million. TLV values are subject to revision if further research or 
experience indicates the need. If the TLV value has changed, the dilution air requirements must be recalculated. The values on the table must be 
multiplied by the evaporation rate (pts/min) to yield the effective ventilation rate (Q') (see Equation 2.5). 



Liquid (TLV in ppm)* 



Ft 3 of Air (STP) Required for Dilution to TLV* 
Per Pint Evaporation 



Acetone (500) 

n-Amyl acetate (100) 

Benzene (0.5) 

n-Butanoi (butyl alcohol) (50) 

n-Buty! acetate (150) 

Butyl Cellosolve (2-butoxyethanol) (25) 

Carbon disulfide (10) 

Carbon tetrachloride (5) 

Cellosolve (2-ethoxyethanol) (5) 

Cellosolve acetate (2-ethoxyethyl acetate) (5) 

Chloroform (10) 

1-2 Dichloroethane (ethylene dichloride) (10) 

1-2 Dichloroethylene (200) 

Dioxane (25) 

Ethyl acetate (400) 

Ethyl alcohol (1000) 

Ethyl ether (400) 

Gasoline (300) 

Isoamyi alcohol (100) 

Isopropyl alcohol (400) 

Isopropyl ether (250) 

Methyl acetate (200) 

Methyl alcohol (200) 

Methyl n-butyl ketone (5) 

Methyl Cellosolve (2-methoxyethanol) (5) 

Methyl Cellosolve acetate (2-methoxyethyl acetate) (5) 

Methyl chloroform (350) 

Methyl ethyl ketone (200) 

Methyl isobutyl ketone (50) 

Methyl propyl ketone (200) 

Naphtha (coal tar) 

Naphtha VM&P (300) 

Nitrobenzene (1) 

n-Propyl acetate (200) 

Stoddard solvent (100) 

1,1,2,2-Tetrachloroethane (1) 

Tetrachloroethylene (25) 

Toluene (50) 

Trichloroethylene (50) 

Xylene (100) 



11,025 

27,200 
NOT RECOMMENDED 

88,000 

20,400 
NOT RECOMMENDED 
NOT RECOMMENDED 
NOT RECOMMENDED 
NOT RECOMMENDED 
NOT RECOMMENDED 
NOT RECOMMENDED 
NOT RECOMMENDED 

26,900 
NOT RECOMMENDED 

10,300 

6,900 

9,630 

REQUIRES SPECIAL CONSIDERATION 

37,200 

13,200 

11,400 

25,000 

49,100 
NOT RECOMMENDED 
NOT RECOMMENDED 
NOT RECOMMENDED 

11,390 

22,500 

64,600 

19,900 

REQUIRES SPECIAL CONSIDERATION 

REQUIRES SPECIAL CONSIDERATION 

NOT RECOMMENDED 

17,500 

30,000-35,000 

NOT RECOMMENDED 

159,400 

75,700 

90,000 

33,000 



*The tabulated dilution air quantities must be multiplied by the selected K value. 



r *See Threshold Limit Values 1997 Appendix A. 



2-4 



Industrial Ventilation 




RES-" AIR INLET 
BEST EXHAUST 
K - 1.0 MINIMUM 



i c 


PLENUM 


IX 


lllll 
) 





! 




1 

A 




Br 


S 


" AIR 


INLET 


BE 


SI 


" EXHAUST 


K 


= 


SO 


MINIMUM 



_LL 



IT 



BEST AIR INLET 
BEST EXHAUST 
K -1.5 MINIMUM 



o 



FAIR AIR INLET 
BEST EXHAUST 
K - 2.5 MINIMUM 



TT 



i 



/ I , 



J / s 



FAIR 

K =:: 2 TO 5 

REE. 2.2 



w 



c: 



I --J L 

J 



itttxtiix^^ 



t J T 



\V" T 




GOOD 

K = 1.5 TO 2 

REF. 2.2 



POOR 

K - 5 TO 10 

REF. 2.2 



NOTE: THE K FACTORS LISTED HERE CONSIDER ONLY THE INLET AND EXHAUST LOCATIONS 
AND ARE JUDGEMENTAL. TO SELECT THE K FACTOR USED IN THE EQUATION, THE 
NUMBER AND LOCATION OF THE EMPLOYEES, THE SOURCE OE THE CONTAMINANT, 
AND THE TOXICITY OF THE CONTAMINANT MUST ALSO BE CONSIDERED. 



IRICAN CONFERENCE 



OF GOVERNMENTS 



INDUSTRIAL 



("YGIENTSTS 



"K" FACTORS 

SUGGESTED FOR INLET 

AND EXHAUST LOCATIONS 



DATE 



1-88 



FIGURE 



2-1 



General Industrial Ventilation 



2-5 



where: 

Q = actual ventilation rate, cfm 
Q' = effective ventilation rate, cfm 
K = a factor to allow for incomplete mixing 

Equation 2.2 then becomes: 



[2.3] 



-(§)" 



[2.4] 



This K factor is based on several considerations: 

1. The efficiency of mixing and distribution of replace- 
ment air introduced into the room or space being 

ventilated (see Figure 2-1). 

2. The toxicity of the solvent. Although TLV and toxicity 
are not synonymous, the following guidelines have 
been suggested for choosing the appropriate K value: 

Slightly toxic material: TLV > 500 ppm 

Moderately toxic material: TLV<100-500 ppm 
Highly toxic material: TLV< 100 ppm 

3. A judgement of any other circumstances which the 
industrial hygienist determined to be of importance 
based on experience and the individual problem. In- 
cluded in these criteria are such considerations as: 

a. Duration of the process, operational cycle, and 
normal locations of workers relative to sources of 
contamination. 

b. Location and number of points of generation of the 
contaminant in the workroom or area. 

c . Seasonal changes in the amount of natural ventilation. 

d. Reduction in operational effectiveness of mechani- 
cal air moving devices. 

e. Other circumstances which may affect the concen- 
tration of hazardous material in the breathing zone 
of the workers. 

The K value selected, depending on the above considerations, 
ranges from 1 to 10. 

2.3.2 Calculating Dilution Ventilation for Steady-State 
Concentration: The concentration of a gas or vapor at a 
steady state can be expressed by the material balance equation 



Q' = - 



Therefore, the rate of flow of un contaminated air required to 
maintain the atmospheric concentration of a hazardous mate- 
rial at an acceptable level can be easily calculated if the 
generation rate can be determined. Usually, the acceptable 
concentration (C) expressed in parts per million (ppm) is 



considered to be the TLV. For liquid solvents, the rate of 
generation is 



CONSTANT xSGxER 
MW 



G = 

where: 

G = generation rate, cfm 
CONSTANT = the volume in ft 3 that 1 pt of liquid, when 
vaporized, will occupy at STP, ftVpt 
SG = Specific gravity of volatile liquid 
ER = evaporation rate of liquid, pts/min 
MW = molecular weight of liquid 
Thus, Q' = G -*■ C can be expressed as 



Q' = 



403x10 6 xSGxER 



MWxC 
EXAMPLE PROBLEM 



[2.5] 



Methyl chloroform is lost by evaporation from a tank at a 
rate of 1.5 pints per 60 minutes. What is the effective venti- 
lation rate (Q') and the actual ventilation rate (Q) required to 
maintain the vapor concentration at the TLV? 

TLV = 350 ppm, SG = 132, MW = 133.4, Assume K = 5 

Assuming perfect dilution, the effective ventilation rate (Q') is 

Q ,^ (403)(10 6 ) (132) (15/60) 
(133.4) (350) 

For incomplete mixing, the actual ventilation rate (Q) is 
(403) (10 6 ) (132) (15/60) (5) 



Q = 



(133.4) (350) 



2.3.3 Contaminant Concentration Buildup (see Figure 
2-2): The concentration of a contaminant can be calculated 
after any change of time. Rearranging the differential material 
balance results in 



dC 



= dt 

G-Q'C~ V 

which can be integrated to yield 



i 



G-Q'C 2 
G-Q'O, 



Q'(t 2 -ti) 

V 



[2.6] 



where subscript 1 refers to the initial condition and subscript 
2 refers to the final condition. If it is desired to calculate the 
time required to reach a given concentration, then rearranging 
t 2 -t t , or At, gives 



At = - 



Q' 



In 



G-Q'C 2 
G-Q'Ci 



[21] 



2-6 



Industrial Ventilation 




FIGURE 2-2. Contaminant concentration buildup 
if C ! = 0, then the equation becomes 
"G-CrCa 



Q' 



In 



[2.8] 



Note: the concentration C 2 is ppm or parts/10 6 (e.g., if C 2 = 
200 ppm, enter C 2 as 200 - 1 6 ). 

If it is desired to determine the concentration level (C 2 ) 
after a certain time interval, t 2 - 1] or At, and if C] =0, then 
the equation becomes 



1-e 



CTAt: 
V 



[2.9] 



G' 

Note: to convert C 2 to ppm, multiply the answer by 10 6 . 

EXAMPLE 

Methyl chloroform vapor is being generated under the 
following conditions: G = 1 .2 cftrt; Q' = 2000 crm; V - 100,000 
cu ft; C] = 0; K = 3. How long before the concentration (C 2 ) 
reaches 200 ppm or 200 - 1 6 ? 



At = - 



Q' 



ln Hr 



= 20.3min 



Using the same values as in the preceding example, what 
will be the concentration after 60 minutes? 



1-e^ v J 



x10 6 -419 ppm 



Q' 



2.3.4 Rate of Purging (see Figure 2-3): Where a quantity 
of air is contaminated but where further contamination or 
generation has ceased, the rate of decrease of concentration 



over a period of time is as follows: 
VdC = -Q'Cdt 



f C2 dc = _o; r 
J C1 c — c J t1 



dt 



In 



9L 
v 



(t 2 -ti) 



or, 



v>2 — \s-tQ 

EXAMPLE 



r Q'<t 2 - 

1 v 



T1) 



[2.10] 



In the room of the example in Section 2.3.3, assume that 
ventilation continues at the same rate (Q' = 2000 cfm) but that 
the contaminating process is interrupted. How much time is 
required to reduce the concentration from 100 (Cj) to 25 (C 2 ) 
ppm? 



t--t 2 =- — In 

1 2 Q , 



'c.^ 



1^1 



= 69.3 min 



In the problem above, if the concentration (Q) at t] is 100 
ppm, what will concentration (Cj) be after 60 minutes (At)? 

C 2 - C^ v J - 30.1 ppm 

2.4 MIXTURES— DILUTION VENTILATION FOR HEALTH 

In many cases, the evaporating liquid for which dilution 
ventilation rates are being designed will consist of a mixture 
of solvents. The common procedure used in such instances is 
as follows. 



o 

o 
o 



c 2 



T 1 TIME T" 2 

FIGURE 2-3. Rate of purging 



General Industrial Ventilation 



2-7 



When two or more hazardous substances are present, their 
combined effect, rather than that of either individually, should 
be given primary consideration. In the absence of information 

to the contrary, the effects of the different hazards should be 
considered as additive. That is, if the sum of the following 
fractions, 



"\ 



TLV, TLV, 



- + ... + . 



TLV n 



etc. 



[2.11] 



exceeds unity, then the threshold limit of the mixture should 
be considered as being exceeded. "C" indicates the observed 
atmospheric concentration and "TLV" the corresponding 
threshold limit. In the absence of information to the contrary, 
the dilution ventilation therefore should be calculated on the 
basis that the effect of the different hazards is additive. The 
air quantity required to dilute each component of the mixture 
to the required safe concentration is calculated, and the sum 
of the air quantities is used as the required dilution ventilation 
for the mixture. 

Exceptions to the above rule may be made when there is 
good reason to believe that the chief effects of the different 
harmful substances are not additive but independent, as when 
purely local effects on different organs of the body are pro- 
duced by the various components of the mixture. In such 
cases, the threshold limit ordinarily is exceeded only when at 
least one member of the series itself has a value exceeding 
unity, e.g., 



-or- 



TL^ TLV 2 

Therefore, where two or more hazardous substances are pre- 
sent and it is known that the effects of the different substances 
are not additive but act independently on the different organs 
of the body, the required dilution ventilation for each compo- 
nent of the mixture should be calculated and the highest cfm 
thus obtained used as the dilution ventilation rate. 

EXAMPLE PROBLEM 

A cleaning and gluing operation is being performed; 
methyl ethyl ketone (MEK) and toluene are both being re- 
leased. Both have narcotic properties and the effects are 
considered additive. Air samples disclose concentrations of 
150 ppm MEK and 50 ppm toluene. Using the equation given, 
the sum of the fractions [(150-200) + (50-50) - 1.75] is 
greater than unity and the TLV of the mixture is exceeded. 
The volumetric flow rate at standard conditions required for 
dilution of the mixture to the TLV would be as follows: 

Assume 2 pints of each is being released each 60 
min. Select a K value of 4 for MEK and a K value of 
5 for toluene; sp gr for MEK = 0.805, for toluene = 
0.866; MW for MEK = 72.1 , for toluene = 92.13. 



QforMEKJ 403 )( 0805 >( 106 )W( 2/6 °) = 3000cfm 
72.1x200 



rt x * . (403) (0.866) (1 6 ) (5) (2 / 60) _ „_ , 

Q for toluene = ^ ^ '— l -±-±± L = 12,627 cfm 

92.13x50 



Q for mixture = 3000 + 12,627 = 15,627 cfm 

2.5 DILUTION VENTILATION FOR FIRE AND 
EXPLOSION 

Another function of dilution ventilation is to reduce the 
concentration of vapors within an enclosure to below the 
lower explosive limit. It should be stressed that this concept 
is never applied in cases where workers are exposed to the 
vapor. In such instances, dilution rates for health hazard 
control are always applied. The reason for this will be appar- 
ent when comparing TLVs and lower explosive limits (LELs). 

The TLV of xylene is 100 ppm. The LEL of xylene is 1% 
or 10,000 ppm. An atmosphere of xylene safe-guarded against 
fire and explosion usually will be kept below 25% of the LEL 
or 2500 ppm. Exposure to such an atmosphere may cause 
severe illness or death. However, in baking and drying ovens, 
in enclosed air drying spaces, within ventilation ductwork, 
etc., dilution ventilation for fire and explosion is used to keep 
the vapor concentration to below the LEL. 

Equation 2.5 can be modified to yield air quantities to dilute 
below the LEL. By substituting LEL for TLV: 



Q; 



(403)(sp gr liquid)(100)(ER)(S f ) 
(MW liquid)(LEL)(B) 



(for Standard Air) [2-12] 



Note 1 . Since LEL is expressed in % (parts per 1 00) rather 
than ppm (parts per million as for the TLV), the 
coefficient of 1,000,000 becomes 100. 

2. Sf is a safety coefficient which depends on the 
percent of the LEL necessary for safe conditions. 
In most ovens and drying enclosures, it has been 
found desirable to maintain vapor concentrations 
at not more than 25% of the LEL at all times in all 
parts of the oven. In properly ventilated continuous 
ovens, a S f coefficient of 4 (25% of the LEL) is 
used. In batch ovens, with good air distribution, the 
existence of peak drying rates requires an S f coef- 
ficient of 10 or 12 to maintain safe concentrations 
at all times. In non-recirculating or improperly 
ventilated batch or continuous ovens, larger S f 
coefficients may be necessary. 

3. B is a constant which takes into account the fact 
that the lower explosive limit of a solvent vapor or 
air mixture decreases at elevated temperatures. B 
= 1 for temperatures up to 250 F; B - 0.7 for 
temperatures above 250 F. 



2-8 



Industrial Ventilation 



EXAMPLE PROBLEM 

A batch of enamel-dipped shelves is baked in a recirculat- 
ing oven at 350 F for 60 minutes. Volatiles in the enamel 
applied to the shelves consist of two pints of xylene. What 
oven ventilation rate, in cfm, is required to dilute the xylene 
vapor concentration within the oven to a safe limit at all times? 

For xylene, the LEL = 1 .0%; sp gr - 0.88; MW - 106; Sf 
- 10; B = 0.7. From Equation 2.12: 

Q J403)(0.88)(2/60)(100)(10) =159cfm 
(106)(10)(0.7) 

Since the above equation is at standard conditions, the air 
flow rate must be converted from 70 F to 350 F (operating 

conditions): 

Q A =(Q STp ) (Ratio of Absolute Temperature) 



: (Gstp) 



(460 F + 350 F) 
(460 F + 70 F) 



Q A =159^ 

A I 530 



- 243 cf m 

EXAMPLE PROBLEM 

In many circumstances, solvent evaporation rate is non- 
uniform due to the process temperature or the manner of 

solvent use. 

A 6 ft diameter muller is used for mixing resin sand on a 
10-minute cycle. Each batch consists of 400 pounds of sand, 
19 pounds of resin, and 8 pints of ethyl alcohol (the ethyl 
alcohol evaporates in the first two minutes). What ventilation 
rate is required? 

For ethyl alcohol, LEL - 3.28%; sp gr - 0.789; MW - 
46.07; Sf=4;B = l 

= (403)(0.789)(3/2 )( 100)(4) = ^^ 
(46.07)(3.28)(1) 

Another source of data is the National Board of Fire Un- 
derwriters' Pamphlet #86, Standard for Class A Ovens and 
FurnacesS 23) This contains a more complete list of solvents 
and their properties. In addition, it lists anddescribes anumber 
of safeguards and interlocks which must always be considered 
in connection with fire dilution ventilation. See also Refer- 
ence 2.4. 

2,6 FIRE DILUTION VENTILATION FOR MIXTURES 

It is common practice to regard the entire mixture as 
consisting of the components requiring the highest amount of 
dilution per unit liquid volume and to calculate the required 
air quantity on that basis. [This component would be the one 



with the highest value for sp gr/(MW)(LEL).] 

2.7 VENTILATION FOR HEAT CONTROL 

Ventilation for heat control in a hot industrial environment 
is a specific application of general industrial ventilation. The 
primary function of the ventilation system is to prevent the 
acute discomfort, heat-induced illness and possible injury of 
those working in or generally occupying a designated hot 
industrial environment. Heat-induced occupational illnesses, 
injuries, or reduced productivity may occur in situations 
where the total heat load may exceed the defenses of the body 
and result in a heat stress situation. It follows, therefore, that 
a heat control ventilation system or other engineering control 
method must follow a physiological evaluation in terms of 
potential heat stress for the occupant in the hot industrial 
environment. 

Due to the complexity of conducting a physiological evalu- 
ation, the criteria presented here are limited to general con- 
siderations. It is strongly recommended, however, that the 
NIOSH Publication No. 86-1 13, Criteria for a Recommended 
Standard, Occ upational Exposure to Hot Environm entsp- 5) be 
reviewed thoroughly in the process of developing the heat 
control ventilation system. 

The development of a ventilation system for a hot industrial 
environment usually includes the control of the ventilation air 
flow rate, velocity, temperature, humidity, and air flow path 
through the space in question. This may require inclusion of 
certain phases of mechanical air-conditioning engineering 
design which is outside the scope of this manual. The neces- 
sary engineering design criteria that may be required are 
available in appropriate publications of the American Society 
of Heating, Refrigeration and Air-Conditioning Engineers 
(ASHRAE) handbook series. 

2.8 HEAT BALANCE AND EXCHANGE 

An essential requirement for continued normal body func- 
tion is that the deep body core temperature be maintained 
within the acceptable range of about 37 C (98.6 F) ± 1 C 
(1.8 F). To achieve this, body temperature equilibrium re- 
quires a constant exchange of heat between the body and the 
environment. The rate and amount of the heat exchange are 
governed by the fundamental laws of thermodynamics of heat 
exchange between objects. The amount of heat that must be 
exchanged is a function of 1) the total heat produced by the 
body (metabolic heat), which may range from about 1 kilo- 
calorie (kcal) per kilogram (kg) of body weight per hour ( 1 . 1 6 
watts) at rest to 5 kcal/kg body weight/hour (7 watts) for 
moderately hard industrial work; and 2) the heat gained, if 
any, from the environment. The rate of heat exchange with 
the environment is a function of air temperature and hu- 
midity; skin temperature; air velocity; evaporation of 
sweat; radiant temperature; and type, amount, and charac- 
teristics of the clothing worn, among other factors. Respi- 



General Industrial Ventilation 



2-9 



ratory heat loss is of little consequence in human defenses 

againstheatstress. 

The basic heat balance equation is: 



C = 0.65V a ub (t a -t sk ) 
where: 



[2.14] 



AS = (M-W)±C±R-E 



[2.13] 



where: 



AS = change in body heat content 

(M - W) = total metabolism - external work per- 
formed 

C = convective heat exchange 

R = radiative heat exchange 

E = evaporative heat loss 

To solve the equation, measurement of metabolic heat 
production, air temperature, air water vapor pressure, wind 
velocity, and mean radiant temperature are required. 

The major modes of heat exchange between man and the 
environment are convection, radiation, and evaporation. 
Other than for brief periods of body contact with hot tools, 
equipment, floors, etc., which may cause burns, conduction 
plays a minor role in industrial heat stress. Because of the 
typically small areas of contact between either body surfaces 
or clothing and hot or cold objects, heat exchange by thermal 
conduction is usually not evaluated in a heat balance equation 
for humans. The effect of heat exchange by thermal conduc- 
tion in human thermal regulation is important when large 
areas of the body are in contact with surfaces that are at 
temperatures different from average skin temperature (nomi- 
nally 95 F), e.g., when someone is prone or supine for long 
periods. It is also important when even small body areas are 
in contact with objects that provide steep thermal gradients 
for heat transfer, e.g., when someone is standing on very cold 
or very hot surfaces. 

The equations for calculating heat exchange by convection, 
radiation, and evaporation are available in Standard Interna- 
tional (SI) units, metric units, and English units. In SI units 
heat exchange is in watts per square meter of body surface 
(W/m 2 ). The heat exchange equations are available in both 
metric and English units for both the seminude individual and 
the worker wearing conventional long-sleeved work shirt and 
trousers. The values are in kcal/h or British thermal units per 
hour (Btu/h) for the "standard worker" defined as one who 
weighs 70 kg (154 lbs) and has a body surface area of 1.8 m 2 
(19.4 ft 2 ). 

2.8. 1 Convection: The rate of convective heat exchange 
between the skin of a person and the ambient air immediately 
surrounding the skin is a function of the difference in tem- 
perature between the ambient air (t a ), the mean weighted skin 
temperature (t sk ) and the rate of air movement over the skin 
(V a ). This relationship is stated algebraically for the "standard 
worker" wearing the customary one-layer work clothing en- 
semble as: 



C = convective heat exchange, Btu/h 
V a = air velocity, fpm 

t a = air temperature, F 

t sk = mean weighted skin temperature, usually as- 
sumed to be 95 F 

When t a > 95 F there will be a gain in body heat from the 
ambient air by convection. When t a < 95 F, heat will be lost 
from the body to the ambient air by convection. 

2.8.2 Radiation: Infrared radiative heat exchange between 
the exposed surfaces of a person's skin and clothing varies as 
a function of the difference between the fourth power of the 
absolute temperature of the exposed surfaces and that of the 
surface of the radiant source or sink, the exposed areas and 
their emissivities. Heat is gained by thermal radiation if the 
facing surface is warmer than the average temperature of the 
exposed skin and clothing, and vice versa. A practical ap- 
proximation for infrared radiant heat exchange for a person 
wearing conventional clothing is: 

R = 15.0(tw-t sk ) [2.15] 

where: 

R = radiant heat exchange, Btu/h 
t w = mean radiant temperature, F 
t sk = mean weighted skin temperature 

2.8.3 Evaporation: The evaporation of water (sweat) or 
other liquids from the skin or clothing surfaces results in a 
heat loss from the body. Evaporative heat loss for humans is 
a function of air flow over the skin and clothing surfaces, the 
water vapor partial pressure gradient between the skin surface 
and the surrounding air, the area from which water or other 
liquids are evaporating and mass transfer coefficients at their 
surfaces. 



E = 2.4V a °- 6 ( Psk -p a ) 
where: 



[2.16] 



E = evaporative heat loss, Btu/h 
V a = air velocity, fpm 

p a = water vapor pressure of ambient air, mmHg 
Psk = water vapor pressure on the skin, assumed to be 
42 mm Hg at a 95 F skin temperature 

2.9 ADAPTIVE MECHANISM OF THE BODY 

Even people in generally good health can adjust physiologi- 
cally to thermal stress only over a narrow range of environ- 
mental conditions. Unrestricted blood flow to the skin, an 
unimpeded flow of dry, cool air over the skin surface and 
sweating are prime defenses in heat stress. Although heat 



2-10 



Industrial Ventilation 



600 



500 



( ^ 300 
b 

co 200 



m 100 



100 



HEAT LOSSES, STORAGE, AND TEMPERATURE 

RELATIONS LOR CLOTHED SUBJECT 



METABOLISM 




„L 



...J.„ 



200 

60 70 80 90 100 110 

DRY BULB TEMPERATURE, DEC. E 

FIGURE 2-4. Heat losses, storage, and temperature relations 

produced by muscle activity reduces the impact of cold stress, 
it can add substantially to the total challenge during heat 
stress. Diminished health status, medications, limited prior 
thermal exposure, among other factors, increase danger from 

thermal stresses. 

The reflex control of blood flow is the body's most effec- 
tive and important first line of defense in facing either cold or 
heat stress. Reducing blood flow to the skin of the hands, feet, 
fingers and toes is an important measure for reducing heat loss 
in a cold environment. Blood flow to the skin, however, 
increases many-fold during heat stress. Its effect is to increase 
rates of heat distribution in the body and maximize conduc- 
tive, convective, radiative and evaporative heat losses to the 
environment (Figure 2-4). Its cost is often to reduce perfusion 
of other organs, especially the brain, and reduce systemic 
arterial blood pressure, leading to reduced consciousness, 
collapse, heat exhaustion and other heat-induced illnesses. 

Reflex sweating during the physical activities of exercise, 
work and/or heat stress brings often large volumes of body 
water and electrolytes (salts) to the skin surface. Heat is lost 
when the water in sweat evaporates. Whether the electrolytes 
remain on the skin surface or are deposited in clothing, they 
are nonetheless permanently lost to the body. The electrolyte 
content of a typical American diet usually provides adequate 
electrolyte replacement for these losses. Electrolyte replace- 
ment fluids, however, may be necessary for people on salt-re- 
stricted diets and those who commonly sustain periods of 
prolonged and profuse sweating. It is essential for everyone 
that the lost body water and electrolytes are replaced in the 
same volume and proportion as lost in sweat. Muscle spasms, 
cramps, gastrointestinal disturbances and general malaise, 



among other signs and symptoms, commonly develop when 
they are not. 

2.10 ACCLIMATIZATION 

People in generally good health normally develop heat 
acclimatization in a week or so after intermittently working 
or exercising in high heat. Its effect is to improve the comfort 
and safety of the heat exposure. It occurs because of an 
increase in total circulating blood volume, an improved ability 
to maintain systemic arterial blood pressure during heat stress, 
and a developed ability to produce larger volumes of more 
dilute sweat at rates of production more precisely matched to 
the heat load. Heat acclimatization rapidly diminishes even 
after a day or so of discontinued activity in the heat. Most is 
lost after about a week. 

2.11 ACUTE HEAT DISORDERS 

A variety of heat disorders can be distinguished clinically 
when individuals are exposed to excessive heat. A brief 
description of these disorders follows. 

2.11.1 Heatstroke: Heat stroke (also known as "sun 
stroke") is a life-threatening condition which without excep- 
tion demands immediate emergency medical care and hospi- 
talization. Before medical care arrives, move the person to a 
shaded area, check for other injuries, ensure there is an 
unobstructed airway, remove or loosen clothing, and flood the 
body surface with free-flowing, tepid (not cold) water. Vig- 
orous fanning helps cooling. Heat stroke develops when body 
heat gains from exercise, work and/or a hot environment 
overwhelm normal thermoregulatory defenses. Charac- 
teristically, sweating has ceased, the skin is hot and dry, and 
deep body temperature is above about 104 F. The person may 
be either diaphoretic, semiconscious, unconscious or agitated, 
delirious and in convulsions. Demand medical care even if 
consciousness returns — lethal effects may develop in the next 
24 to 72 hours. 

2.11.2 Heat Exhaustion: Heat exhaustion (also called 
"exercise-induced heat exhaustion" and "heat syncope") most 
commonly occurs in people who are not heat acclimatized and 
who are in poor physical condition, obese, inappropriately 
dressed, and exercising or working energetically in the heat 
at unaccustomed and/or demanding tasks. It is characterized 
by lightheadedness, dizziness, vision disturbances, nausea. 
vague flu-like symptoms, tinnitus, weakness, and occasion- 
ally, collapse. The person's deep body temperature is typi- 
cally in a normal range or only slightly elevated; the skin is 
moist and cool but may be reddened by its high rate of blood 
flow. Heat exhaustion develops when there is reflex demand 
for blood flow to the skin to dissipate body heat and a 
simultaneous reflex demand for blood flow to exercising 
muscles to meet metabolic needs of increased activity. These 
peripheral distributions of blood volume reduce systemic 
arterial pressure and brain blood flow, causing most of the 



Genera! Industrial Ventilation 



2-11 



TABLE 2-2. Estimating Energy Cost of Work by Task Analysis' 2 6) 

A. Body position and movement kcal/min* 



Sitting 




0.3 


Standing 




0.6 


Walking 




2.0-3.0 


Walking uphill 


Add 0.8/meter rise 


B. Type of Work 


Average 
kcal/min 


Range 
kcal/min 


Hand work - light 


0.4 


0.2-1.2 


Hand work - heavy 


0.9 




Work one arm - light 


1.0 


07-2.5 


Work one arm - heavy 


1.7 




Work both arms - light 


1.5 


1.0-3.5 


Work both arms - heavy 


2.5 




Work whole body - light 


3.5 


2.5-15.0 


Work whole body - moderate 


5.0 




Work whole body - heavy 


7.0 




Work whole body - very heavy 


9.0 




C. Basal metabolism 


1.0 




D. Sample calculation ** 






Assembling work with heavy hand tools 






1 . Standing 


0.6 




2. Two-arm work 


3.5 




3. Basal metabolism 


1.0 




TOTAL 


5.1 kcal/min 



*For standard worker of 70 kg body weight (154 lbs) and 1.8 m body surface 
(19.4 ft 2 ). 

**Example of measuring metabolic heat production of a worker when performing 
initial screening. 



symptoms of heat exhaustion. Resting in a cool environment 
where there is free flowing, dry air usually remediates symp- 
toms quickly. Although heat exhaustion is debilitating and 
uncomfortable, it is not often a long-term health threat. There 
are considerable dangers, of course, for anyone operating 
machinery when consciousness is impaired because of heat 
exhaustion or for any other reason. 

2.11.3 Heat Cramps and Heat Rash: Heat cramps (also 
known as "muscle cramps") are spontaneous, involuntary, 
painful and prolonged muscle contractions that commonly 
occur in otherwise healthy people when both body water and 
electrolyte levels have not been restored after extended peri- 
ods of heavy sweating during exercise and/or heat stress. Full 
recovery can be expected in about 24 hours with the use of 
electrolyte replacement fluids and rest. Heat rash (also known 
as "prickly heat" or "miliaria rubia") is an acute, inflammatory 
skin disease characterized by small red, itchy or tingling 
lesions, commonly in areas of skin folds or where there is 



abrasive clothing. It commonly disappears when these areas 
are kept dry, unabraded and open to free flowing, dry air. 

2.12 ASSESSMENT OF HEAT STRESS AND HEAT 
STRAIN 

Heat Stress is defined by environmental measurements of 
air temperature, humidity, air flow rate, the level of radiant 
heat exchange and evaluation of a person's metabolic heat 
production rate from exercise and/or work. Heat stress is the 
load on thermoregulation. Heat Strain is defined as the cost 
to each person facing heat stress. Although all people working 
at the same intensity in the same environment face the same 
level of heat stress, each is under a unique level of heat strain. 
Almost any environmental thermal exposure will be comfort- 
able and safe for some, but endangering, even lethal to others. 
Because disabilities, danger and death arise directly from heat 
strain, no measure of heat stress is a reliable indicator of a 
particular person's heat strain or the safety of the exposure. 

2.12.1 Evaluation of Heat Stress: Dry-bulb air tempera- 
ture (DB: so-called "dry-bulb" temperature) is measured by 
calibrated thermometers, thermistors, thermocouples and 
similar temperature-sensing devices which themselves do not 
produce heat and which are protected from the effects of 
thermal conduction, evaporation, condensation and radiant 
heat sources and sinks. Relative humidity is evaluated psy- 
chrometrically as a function of the steady-state difference 
between dry-bulb temperature and that indicated by the tem- 
perature of a sensor covered with a freely evaporating, water- 
saturated cotton wick. Such a measure reports "NWB" 



NATURAL W. B. 
THERMOMETER 




D. B. THERMOMETER 
(USED ONLY OUTDOOR 
IN SUNSHINE) 



GLOBE 
THERMOMETER 



6" COPPER SHELL 
PAINTED MATTE BLACK 



FIGURE 2-5. Determination of wet-bulb globe temperature 



2-12 



Industrial Ventilation 



(natural wet-bulb temperature) when the wetted sensor is 
affected only by prevailing air movement, and "WB M (when 
it is exposed to forced convection). Free air movement is 
measured with an unobstructed anemometer. Infrared radiant 
"heat transfer" is typically measured by a temperature sensor 
at the center of a 6-inch, hollow, copper sphere painted flat 
("matte") black. Such a measure reports "GT" (globe tempera- 
ture) (Figure 2-5). A person's metabolic heat production is 
usually evaluated from an estimated level of average physical 
activity (Table 2-2). 

Although there are a number of different indices for evalu- 
ating heat stress, none is reliable as a sole indicator of heat 
strain for a specific person. Dry-bulb temperature is the least 
valuable measure of heat stress because it provides no infor- 
mation about ambient relative humidity, or heat exchange by 
convection or radiation, and gives no estimate of the metabo- 
lic heat production. Wet-bulb, globe temperature (WBGT) is 
often used as an index of heat stress. When there is a source 
of radiant heat transfer (solar radiation, hot surfaces of ma- 
chinery): 



WBGT - 0.7t nwb + 0.2 t g +0.1 t a 



where 

t nwb = natural wet-bulb temperature 
t g = globe temperature 
When radiant heat transfer is negligible; 

WBGT = 0.7 U h + 0.3 t n 



[2.17] 



[2.18] 



WBGT evaluates more factors contributing to heat stress than 
does dry-bulb temperature alone. It does not, however, effec- 
tively evaluate the importance of energy transfer from human 
skin by convection which is essential for the removal of heat 
from the skin surface and the formation of water vapor from 
secreted sweat. Nor does WBGT evaluate the importance of 
metabolic heat production in the heat stress. Under some 
environmental conditions, heat produced by metabolism is the 
predominant stressor. 

2.12.2 Evaluation of Heat Strain: The incidence and se- 
verity of heat strain will vary greatly among people exposed 
to the same level of heat stress. Paying attention to the early 
signs and symptoms of heat strain is the best first line of 
defense against debilitating heat-induced discomfort and in- 
juries. It is dangerous, inappropriate and irresponsible to 
consider a heat stress as safe for all when some exposed to it 
show heat strain signs and symptoms, while others do not. 
Acute heat strain is indicated by: 

Visible Sweating: Thermoregulatory reflexes nor- 
mally fine-tune with precision the rate of sweating to 
the rate at which body heat must be lost to maintain 
homeostasis. Normally, there is no liquid water on the 
skin surface in a tolerable heat stress because water 
brought to the skin surface by sweating readily forms 



invisible water vapor in the process of evaporative 
cooling. Although an all too common occurrence in 
the workplace, liquid sweat either on the skin surface, 
or soaked into clothing, is a sure sign of heat strain. It 
indicates the level of sweating required to keep body 
temperature in a normal range cannot be matched by 
the rate of water evaporation from the skin surface to 
the environment. It is necessary either to increase the 
air flow rate over skin and clothing surfaces, lower 
ambient temperature and relative humidity, reduce 
radiative heat gain, and/or reduce metabolic heat pro- 
duction if progressive heat disabilities are to be 
avoided. 

Discontinued Sweating: A hot, dry skin for someone 
exposed to heat stress is a dangerous sign. It indicates 
suppression of sweating, perhaps exacerbated by pre- 
scription or over-the-counter medications. The appear- 
ance of a hot, dry skin for someone in a heat stress 
demands immediate attention and corrective actions. 

Elevated Heart Rate: Short-term increases in heart 
rate are normal for episodic increases in work load. In 
a heat stress, however, a sustained heart rate greater 
than 160/min for those younger than about 35 years, 
or 140/min for those who are older, is a sign of heat 
strain. 

Elevated Deep Body Temperature: A sustained deep 
body temperature greater than 100.4 F is a sign of heat 
strain in someone exposed to heat stress. 

Decreased Systemic Arterial Blood Pressure: A fall 
in blood pressure of more than about 40 Torr in about 
3.5 minutes for someone working in a heat stress 
indicates a heat-induced disability. Reduced con- 
sciousness, feeling of weakness, vision disturbances, 
and other signs and symptoms are likely to follow. 

Personal Discomfort: Heat strain may be indicated in 
some heat-stressed individuals by severe and sudden 
fatigue, nausea, dizziness, lightheadedness, or faint- 
ing. Others may complain of irritability; mental con- 
fusion; clumsiness; forgetfulness; general malaise; the 
development of sometimes vague, flu-like symptoms; 
and paradoxical chills and shivering. 

Infrequent Urination: Urinating less frequently than 
normal and the voiding of a small volume of dark-col- 
ored urine is a sign of whole body dehydration. Dehy- 
dration compromises the body's ability to maintain a 
large enough circulating blood volume so that normal 
blood pressure is maintained in the face of the com- 
bined stressors of exercise and heat exposure. People 
who work or exercise in the heat need to develop the 
habit of drinking adequate volumes of water at fre- 
quent enough intervals to maintain the same patterns 
of urination they have when not heat stressed. Those 
who sweat heavily for long periods need also to discuss 



General Industrial Ventilation 



2-13 



with their physicians a possible need for using electro- 
lyte replacement fluids. 

2.13 WORKER PROTECTION 

There is improved safety, comfort and productivity when 
those working in the heat are: 

1. In generally good physical condition and not obese, 
are heat acclimatized, and are experienced in the heat 
stressing job. They also need to know how to select 
clothing and maintain whole body hydration and elec- 
trolyte levels to provide the greatest comfort and 
safety. 

2. In areas that are well- ventilated and shielded from 
infrared radiant heat sources. 

3. Knowledgeable about the effects of their medications 
on cardiovascular and peripheral vascular function, 
blood pressure control, body temperature mainte- 
nance, sweat gland activity, metabolic effects and 
levels of attention or consciousness. 

4. Appropriately supervised when there is a history of 
abuse or recovery from abuse of alcohol or other 

intoxicants. 

5. Provided accurate verbal and written instructions, fre- 
quent training programs and other information about 
heat stress and strain. 



6. Able to recognize the signs and symptoms of heat 
strain in themselves and others exposed to heat stress 
and know the appropriately effective steps for their 
remediation (Figures 2-6 and 2-7). 

2.14 VENTILATION CONTROL 

The control method presented here is limited to a general 
engineering approach. Due to the complexity of evaluating a 
potential heat stress-producing situation, it is essential that the 
accepted industrial hygiene method of recognition, evalu- 
ation, and control be utilized to its fullest extent. In addition 
to the usual time-limited exposures, it may be necessary to 
specify additional protection which may include insulation, 
baffles, shields, partitions, personal protective equipment, 
administrative control, and other measures to prevent possible 
heat stress. Ventilation control measures may require a source 
of cooler replacement air, an evaporative or mechanically 
cooled source, a velocity cooling method, or any combination 
thereof. Specific guidelines, texts, and other publications or 
sources should be reviewed for the necessary data to develop 
the ventilation system. 

2.15 VENTILATION SYSTEMS 

Exhaust ventilation can be used to remove excessive heat 
and/or humidity if a replacement source of cooler air is 
available. If it is possible to enclose the heat source, such as 
is the case with ovens or certain furnaces, a gravity or forced 



i — 104 

o 



40 




8 1 5 rr) i n . /h . \ 
♦30 min./h./ / 



>RAL 



\53_ _ ,„34_9_;__ "" _J_f_ 

METABOLIC HEAT 



_500kcn!/h, 
20Q0BLu/h. 

.,.380 wo Us 



C = CEiUNG LIMIT 

RAL = RECOMMENDED ALERT LIMIT 

*E0R "STANDARD WORKER" OF 70 kg (154 lbs) BODY WEIGHT AND 
1.8 m 2 (19.4 ft 2 ) BODY SURFACE, 



FIGURE 2-6. Recommended heat-stress alert limits, heat-unacclimatized workers 



2-14 



Industrial Ventilation 



1134 



95 



35 



— • C 



86 30 



25 




68 20; 



» 1 5) rnin. 
•30 min. 



1 



x 



— »45 min./h . 
~^<60 min./h J 



■REL 



"100 



200 



400 



800 



2J5 



300 4 -QQ 50 kcal/h. 

1 200 "1600 2000 Etu/h. 



349 " 



465 



1M0 Wotts 



METABOLIC HEAT 

C = CEILING LIMIT 

REL = RECOMMENDED EXPOSURE LIMIT 

*FOR 'STANDARD WORKER" OF 70 kq (154 lbs) BODY WEIGHT AND 
1.8 m 2 (19.4 ft 2 ) BODY SURFACE. " 



FIGURE 2-7. Recommended heat-stress exposure limits, heat-acclimatized workers 



air stack may be all that is necessary to remove excessive heat 
from the workroom. If a partial enclosure or local hood is 
indicated, control velocities should be used as described in 

Chapter 3. 

Many operations do not lend themselves to local exhaust. 
General ventilation may be the only alternative. To determine 
the required general ventilation, the designer must estimate 
the acceptable temperature or humidity rise. The first step in 
determining the required volumetric flow is to determine the 
sensible and latent heat load. Next, determine the volumetric 
flow to dissipate the sensible heat and the volumetric flow to 
dissipate the latent heat. The required general ventilation is 
the larger of the two volumetric flows. 



The sensible heat rise can be determined by the following: 
H s =Q s xpxc p xATx(60 min/hr) p.19] 

where: 

H s = Sensible heat gain, BTU/hr 

Q s = Volumetric flow for sensible heat, cfm 

p = Density of the air, lbm/ft 3 

c p = Specific heat of the air, BTU/lbm- F 

A T = Change in temperature, F 

For air c p - 0.24 BTU/lbm - F and p - 0.075 lbm/ft 3 ; 





{ 


) 


\ 


) 


*> ' 


D 









600 FPM 

TARGET VEL 



FIGURE 2-8. Good natural ventilation and circulation 



FIGURE 2-9. Good mechanically supplied ventilation 



Genera] Industrial Ventilation 



2-15 



consequently, the equation becomes 



Qi = 



H s =108xQ s xAT 



or 



Q 8 =H 8+ (t08xAT) 



[2.20] 



In order to use this equation, it is necessary to first estimate 
the heat load. This will include loads from the sun, people, 
lights, and motors, as well as other particular sources of heat. 
Of these, sun load, lights, and motors are all completely 
sensible. The people heat load is part sensible and part latent. 
In the case of hot processes which give off both sensible and 
latent heat, it will be necessary to estimate the amounts or 
percents of each. In using the above equation for sensible heat, 
one must decide the amount of temperature rise which will be 
permitted. Thus, in a locality where 90 F outdoor dry bulb 
may be expected, if it is desired that the inside temperature 
not exceed 100 F, or a 10-degree rise, a certain air flow rate 
will be necessary. If an inside temperature of 95 F is required, 
the air flow rate will be doubled. 

For latent heat load, the procedure is similar, although more 
difficult. If the total amount of water vapor is known, the heat 
load can be estimated from the latent heat of vaporization, 970 
BTU/lb. In a manner similar to the sensible heat calculations, 
the latent heat gain can be approximated by: 

H, ^Q^pxqxAhx^Omin/hOxOlb/yOOO grains) 

where: 

H 1 = Latent heat gain, BTU/hr 
Q 1 = Volumetric flow for latent heat, cftn 
p = Density of the air, Ibm/ft 3 
c 1 = Latent heat of vaporization, BTU/lbm 
Ah = Change in absolute humidity of the air, grains- 
water/lbm-dry air 

For air, q is approximately 970 BTU/lb and p = 0.075 
lbm/ft 3 . Consequently, the equation becomes 

Hi = 0.62xQ 1 xAh 

or 



Q 



K 



0.62 x Ah 



[2.21] 



If the rate of moisture released, M in pounds per hours, is 
known, then 

M = Q, x p x Ah x (1 lb / 7000 gr) x (60 min/ hr) 
= Q-|Xpx Ah -s- (116.7) 



or 



116.7xM 
pxAh 



[2.22] 



The value of the "grains-water per pound-air difference" is 
read from a psychrometric chart or table. It represents the 
difference in moisture content of the outdoor air and the 
conditions acceptable to the engineer designing the exhaust 
system. The air quantities calculated from the two equations 
above should not be added to arrive at the required quantity. 
Rather, the higher quantity should be used since both sensible 
and latent heat are absorbed simultaneously. Furthermore, in 
the majority of cases the sensible heat load far exceeds the 
latent heat load, so the design usually can be calculated on the 
basis of sensible heat alone. 

The ventilation should be designed to flow through the hot 
environment in a manner that will efficiently control the 
excess heat. Figures 2-8 and 2-9 illustrate this principle. 

2.16 VELOCITY COOLING 

If the air dry-bulb or wet-bulb temperatures are lower than 
95-1 00 F, the worker may be cooled by convection or evapo- 
ration. When the dry bulb temperature is higher than 95-100 
F, increased air velocity may add heat to the worker by 
convection. If the wet bulb temperature is high also, evapora- 
tive heat loss may not increase proportionately and the net 
result will be an increase in the worker's heat burden. Many 
designers consider that supply air temperature should not 
exceed 80 F for practical heat relief. 

Current practice indicates that air velocities in Table 2-3 
can be used successfully for direct cooling of workers. For 
best results, provide directional control of the air supply 
(Figure 2-10) to accommodate daily and seasonal variations 
in heat exposure and supply air temperature. 

2.17 RADIANT HEAT CONTROL 

Since radiant heat is a form of heat energy which needs no 
medium for its transfer, radiant heat cannot be controlled by 

TABLE 2-3. Acceptable Comfort Air Motion at the Worker 



Air Velocity, fpm* 


Continuous Exposure 
Air conditioned space 


50-75 


Fixed work station, general ventilation 
or spot cooling: Sitting 
Standing 


75-125 
100-200 


Intermittent Exposure, Spot Cooling or Relief Stations 


Light heat loads and activity 
Moderate heat loads and activity 
High heat loads and activity 


1000-2000 
2000-3000 
3000^000 



*Note: Velocities greater than 1000 fpm may seriously disrupt the performance of 
nearby local exhaust systems. Care must be taken to direct air motion to 
prevent such interference. 



2-16 



Industrial Ventilation 




FIGURE 2-10. Spot cooling with volume and directional control 

ventilation. Painting or coating the surface of hot bodies with 
materials having low radiation emission characteristics is one 
method of reducing radiation. 

For materials such as molten masses of metal or glass which 
cannot be controlled directly, radiation shields are effective. 
These shields can consist of metal plates, screens, or other 
material interposed between the source of radiant heat and the 
workers. Shielding reduces the radiant heat load by reflecting 
the major portion of the incident radiant heat away from the 
operator and by re-emitting to the operator only a portion of 
that radiant heat which has been absorbed. Table 2-4 indicates 
the percent of both reflection and emission of radiant heat 
associated with some common shielding materials. Addi- 
tional ventilation will control the sensible heat load but will 
have only a minimal effect, if any, upon the radiant heat load. 
See Figure 2-11. 

2.18 PROTECTIVE SUITS FOR SHORT EXPOSURES 

For brief exposures to very high temperatures, insulated 
aluminized suits and other protective clothing may be worn. 
These suits reduce the rate of heat gain by the body but provide 
no means of removing body heat; therefore, only short expo- 
sures may be tolerated. 



NO 

HEAT 

TO 

ROOM l 





J/1 

FIGURE 2-11. Heat Shielding 



,_ n nz- 




iri ! i 

(Tunis o~aro) 


> x 



TABLE 2-4. Relative Efficiencies of Common Shielding Materials 

Reflection of 

Radiant Heat Emission of 

Incident Upon Radiant Heat 

Surface of Shielding Surface from Surface 



Aluminum, bright 95 

Zinc, bright 90 

Aluminum, oxidized 84 

Zinc, oxidized 73 

Aluminum paint, new, clean 65 

Aluminum paint, dull, dirty 40 

Iron, sheet, smooth 45 

Iron, sheet, oxidized 35 

Brick 20 

Lacquer, black 10 

Lacquer, white 10 

Asbestos board 6 

Lacquer, flat black 3 



5 
10 
16 
27 
35 
60 
55 
65 
80 
90 
90 
94 
97 



2.19 RESPIRATORY HEAT EXCHANGERS 

For brief exposure to air of good quality but high tempera- 
ture, a heat exchanger on a half-mask respirator face piece is 
available. This device will bring air into the respiratory pas- 
sages at a tolerable temperature but will not remove contami- 
nants nor furnish oxygen in poor atmospheres. 

2.20 REFRIGERATED SUITS 

Where individuals must move about, cold air may be blown 
into a suit or hood worn as a portable enclosure. The usual 
refrigeration methods may be used with insulated tubing to 
the suit. It may be difficult, however, to deliver air at a 
sufficiently low temperature. If compressed air is available, 
cold air may be delivered from a vortex tube worn on the suit. 
Suits of this type are commercially available. 

2.21 ENCLOSURES 

In certain hot industries, such as in steel mills, it is imprac- 
tical to control the heat from the process. If the operation is 
such that remote control is possible, an air conditioned booth 
or cab can be utilized to keep the operators reasonably com- 
fortable in an otherwise intolerable atmosphere. 

2.22 INSULATION 

If the source of heat is a surface giving rise to convection, 
insulation at the surface will reduce this form of heat transfer. 
Insulation by itself, however, will not usually be sufficient if 
the temperature is very high or if the heat content is high. 



General Industrial Ventilation 



2-17 



REFERENCES 

2.1 U.S. Department of Health, Education and Welfare, 
PHS, CDC, NIOSH: The Industrial Environment—Its 
Evaluation and Control. Government Printing Office, 
Washington, DC (1973). 

2.2 U.S. Air Force: AFOSH Standard 161.2. 

2.3 National Board of Fire Underwriters: Pamphlet #86, 
Standards for Class A Ovens and Furnaces. 

2.4 Feiner, B.; Kingsley, L.: Ventilation of Industrial Ov- 



ens. Air Conditioning, Heating and Ventilating, pp. 82 
89 (December 1956). 

2.5 U.S. Department of Health and Human Services, PHS, 
CDC, NIOSH: Occupational Exposure to Hot Envi- 
ronments, Revised Criteria, 1986. 

2.6 American Conference of Governmental Industrial Hy- 
gienists, Inc.: 1997 Threshold Limit Values and Bio- 
logical Exposure Indices, p. 138, ACGIH, Cincinnati 
(1997). 



Chapter 3 

LOCAL EXHAUST HOODS 



3.1 INTRODUCTION 3-2 

3.2 CONTAMINANT CHARACTERISTICS 3-2 

3.2.1 Inertial Effects 3-2 

3.2.2 Effect of Specific Gravity 3-2 

3.2.3 Wake Effects 3-2 

3.3 HOOD TYPES 3-2 

3.3.1 Enclosing Hoods 3-2 

3.3.2 Exterior Hoods 3-2 

3.4 HOOD DESIGN FACTORS 3-2 

3.4.1 Capture Velocity 3-6 

3.4.2 Hood Flow Rate Determination 3-6 

3.4.3 Effects of Flanges and Baffles . . . 3-7 

3.4.4 Air Distribution 3-7 

3.4.5 Rectangular and Round Hoods 3-8 

3.4.6 Worker Position Effect 3-8 

3.5 HOOD LOSSES 3-15 

Figure 3-1 Hood Nomenclature Local Exhaust 3-3 

Figure 3-2 Effects of Specific Gravity 3-4 

Figure 3-3 Enclosure and Operator/Equipment Interface . . 3-5 

Figure 3-4 Point Suction Source 3-6 

Figure 3-5 Flow Rate as Distance From Hood 3-7 

Figure 3-6 Velocity Contours — Plain Circular Opening . . 3-8 

Figure 3-7 Velocity Contours — Flanged Circular Opening . 3-8 

Figure 3-8 Flow/Capture Velocity 3-9 

Figure 3-9 Flow/Capture Velocity 3-10 

Figure 3-10 Flow/Capture Velocity 3-11 

Figure 3-11 Hood Types 3-12 

Figure 3-12 Distribution Techniques — Slot Resistance and 

Fish Tail 3-13 



3.5.1 Simple Hoods 3-16 

3.5.2 Compound Hoods 3-16 

3.6 MINIMUM DUCT VELOCITY 3-18 

3.7 SPECIAL HOOD REQUIREMENTS 3-18 

3.7.1 Ventilation of Radioactive and High Toxicity 
Processes 3-18 

3.7.2 Laboratory Operations 3-19 

3.8 PUSH-PULL VENTILATION 3-19 

3.8.1 Push Jet 3-19 

3.8.2 Pull Hood 3-20 

3.8.3 Push-Pull System Design 3-21 

3.9 HOT PROCESSES 3-21 

3.9.1 Circular Fligh Canopy Hoods 3-21 

3.9.2 Rectangular High Canopy Hoods 3-22 

3.9.3 Low Canopy Hoods 3-23 

REFERENCES 3-23 

Figure 3-13 Distribution Techniques - Booth Canopy and 

Side-Drafts and Suspended Floods 3-14 

Figure 3-14 Worker Position Effect 3-15 

Figure 3-15 Airflow at the Vena Contracta 3-16 

Figure 3-16 Hood Loss Factors 3-17 

Figure 3-17 Simple Hood 3-18 

Figure 3-18 Compound Hood 3-18 

Figure 3-19 Jet Velocity Profile 3-20 

Figure 3-20 Dimensions Used to Design High-Canopy Hoods 

for Hot Sources 3-20 



3-2 



Industrial Ventilation 



3.1 INTRODUCTION 

Local exhaust systems are designed to capture and remove 
process emissions prior to their escape into the workplace 
environment. The local exhaust hood is the point of entry into 
the exhaust system and is defined herein to include all suction 
openings regardless of their physical configuration. The pri- 
mary function of the hood is to create an air flow field which 
will effectively capture the contaminant and transport it into 
the hood. Figure 3- 1 provides nomenclature associated with 
local exhaust hoods. 

3.2 CONTAMINANT CHARACTERISTICS 

3.2.1 Inertial Effects: Gases, vapors, and fumes will not 
exhibit significant inertial effects. Also, fine dust particles, 20 
microns or less in diameter (which includes respirable parti- 
cles), will not exhibit significant inertial effects. These mate- 
rials will move solely with respect to the air in which they are 
mixed. In such cases, the hood needs to generate an air flow 
pattern and capture velocity sufficient to control the motion 
of the contaminant- 1 ad en air plus extraneous air currents 
caused by room cross-drafts, vehicular traffic, etc. 

3.2.2 Effective Specific Gravity: Frequently, the location 
of exhaust hoods is mistakenly based on a supposition that the 
contaminant is "heavier than air" or "lighter than air." In most 
health hazard applications, this criterion is of little value (see 
Figure 3-2). Hazardous fine dust particles, fumes, vapors, and 
gases are truly airborne, following air currents, and are not 
subject to appreciable motion either upward or downward 
because of their own density. Normal air movement will 
assure an even mixture of these contaminants. Exception to 
these observations may occur with very hot or very cold 
operations or where a contaminant is generated at very high 
levels and control is achieved before the contaminant be- 
comes diluted. 

3.2.3 Wake Effects: As air flows around an object, a phe- 
nomenon known as "boundary layer separation" occurs. This 
results in the formation of a turbulent wake on the downstream 
side of the object similar to what is observed as a ship moves 
through the water. The wake is a region of vigorous mixing 
and recirculation. If the object in question is a person who is 
working with, or close to, a contaminant-generating source, 
recirculation of the contaminant into the breathing zone is 
likely. An important consideration in the design of ventilation 
for contaminant control is minimizing this wake around the 
human body and, to the extent possible, keeping contaminant 
sources out of these recirculating regions (see also Section 
3.4.6.) 

3.3 HOOD TYPES 

Hoods may be of a wide range of physical configurations 
but can be grouped into two general categories: enclosing and 
exterior. The type of hood to be used will be dependent on the 



physical characteristics of the process equipment, the con- 
taminant generation mechanism, and the operator/equipment 
interface (see Figure 3-3). 

3.3.1 Enclosing Hoods: Enclosing hoods are those which 
completely or partially enclose the process or contaminant 
generation point. A complete enclosure would be a laboratory 
glove box or similar type of enclosure where only minimal 
openings exist. A partial enclosure would be a laboratory hood 
or paint spray booth. An inward flow of air through the 
enclosure opening will contain the contaminant within the 
enclosure and prevent its escape into the work environment. 

The enclosing hood is preferred wherever the process con- 
figuration and operation will permit. If complete enclosure is 
not feasible, partial enclosure should be used to the maximum 
extent possible (see Figure 3-3). 

3.3.2 Exterior Hoods: Exterior hoods are those which are 
located adjacent to an emission source without enclosing it. 
Examples of exterior hoods are slots along the edge of the tank 
or a rectangular opening on a welding table. 

Where the contaminant is a gas, vapor, or fine particulate 
and is not emitted with any significant velocity, the hood 
orientation is not critical. However, if the contaminant con- 
tains large particulates which are emitted with a significant 
velocity, the hood should be located in the path of the emis- 
sion. An example would be a grinding operation (see Chapter 
10,VS-80-ll). 

If the process emits hot contaminated air, it will rise due to 
thermal buoyancy. Use of a side draft exterior hood (located 
horizontally from the hot process) may not provide satisfac- 
tory capture due to the inability of the hood-induced air flow 
to overcome the thermally induced air flow. This will be 
especially true for very high temperature processes such as a 
melting furnace. In such cases, a canopy hood located over 
the process may be indicated (see Section 3.9), 

A variation of the exterior hood is the push-pull system 
(Section 3.8). In this case, a jet of air is pushed across a 
contaminant source into the flow field of a hood. Contaminant 
control is primarily achieved by the jet. The function of the 
exhaust hood is to receive the jet and remove it. The advantage 
of the push-pull system is that the push jet can travel in a 
controlled manner over much greater distances than air can 
be drawn by an exhaust hood alone. The push-pull system is 
used successfully for some plating and open surface vessel 
operations but has potential application for many other proc- 
esses. However, the push portion of the system has potential 
for increasing operator exposure if not properly designed, 
installed, or operated. Care must be taken to ensure proper 
design, application, and operation. 

3.4 HOOD DESIGN FACTORS 

Capture and control of contaminants will be achieved by 



Loca! Exhaust Hoods 3-3 



"3UCT velocity 



DUCT VU OC! 



1> 



LACL 
VELOCITY 



N-. 



V. 



source: 



/<: 



X "At 

PLENUM \H 
4 VE'LOCEfY ~-^j 



\U 



SLOT 
VELOCiT N 



CAPTURE 
VELOCITY 



-i-^ 



SOURCE 



CAPTURE \/rC0C!'O 



FACE VELOCITY- 



SLOT VELOCITY- 



PLENUM VELOCITY- 



DUCT VELOCITY- 



AIR VELOCITY AY ANY POINT IN ERONT OE THE HOOD CM AT THE HOOD 
OPENING NECESSARY TO OVERCOME OPPOSING ASR CURRENTS AND TO 
CAPTURE THE CONTAMINATED AIR AT THAT POINT BY CAUSING i'i" TO EI.OW 
INTO THE HOOD. 

AIR VELOCITY AT THE HOOD OPENING. 

AIR VELOCITY THROUGH THE OPENINGS IN A SLOT- TYPE HOOD. IT IS 
USED PRIMARILY AS A MEANS OE OBTAINING UNIFORM AIR DISTRIBUTION 
ACROSS [THE PACE OF THE ROOD. 

AIR VELOCITY IN THE PLENUM. FOR GOOD AIR DISTRIBUTION 
WITH SLOT-TYPES OE HOODS, THE MAXIMUM PLENUM VELOCITY 
SHOULD BE 1/2 OE THE SLOT VELOCITY OR LESS. 

AIR VELOCITY THROUGH THE DUCT CROSS SECTION. WHEN SOLID MATERIAL If: 
PRESENT IN THE AIR STREAM, THE DUCT VELOCITY MUST BE EQUAL TO OR 
GREATER THAN THE MINIMUM AIR VELOCITY REQUIRED TO MOVE THE 
PARTICLES IN THE ASR STREAM. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRI AL H YGIEN1STS 



HOOD NOMENCLA TURF; 

LOCAL EXIT A LIST 



DATE 



4-96 I 



FIGURE 



'■j 



1 



3-4 Industrial Ventilation 




GOOD 



BAD 



LOCATION 



SOLVENT VAPORS IN HEALTH HAZARD CONCENTRATIONS ARE NOT APPRECIABLY HEAVIER THAN AIR. 
EXHAUST FROM THE FLOOR USUALLY GIVES FIRE PROTECTION ONLY. 



AM/ERIC AN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



EFFECTS OF 
SPECIFIC GRAVITY 



date 4-96 



FIGURE 



1 



3-2 



Loca! Exhaust Hoods 3-5 




E NIC LOSe THE OPERATION AS MUCH AS POSSIBLE. THE MORE COMPLETELY ENCLOSED THE 
SOURCE, THE LESS AIR REQUIRED EOR CONTROL. 



SLOT - 





GOOD 



BAD 



DIRECTION OF AIR FLOW 

LOCATE THE HOOD SO THE CONTAMINANT IS REMOVED AWAY FROM THE BREATHING 
ZONE OF THE OPERATOR. 



AMERIC AN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



ENCLOSURE AND OPERATOR 
EQUIPMENT INTERFACE 



DATE 



4-96 



FIGURE 



3 3 



3-6 



Industrial Ventilation 



TABLE 3-1 . Range of Capture Velocities* 3 x 3 2 > 



Condition of Dispersion of Contamination 



Example 



Capture Velocity, fpm 



Released with practically no velocity into quiet air. 
Released at low velocity into moderately still air. 

Active generation into zone of rapid air motion. 

Released at high initial velocity into zone at very rapid air motion. 



Evaporation from tanks; degreasing, etc. 50-1 00 

Spray booths; intermittent container filling; low speed 100-200 

conveyor transfers; welding; plating; pickling 

Spray painting in shallow booths; barrel filling; 200-500 

conveyor loading; crushers 

Grinding; abrasive blasting; tumbling 500-2000 



in each category above, a range of capture velocity is shown. The 
Lower End of Range 

1 . Room air currents minimal or favorable to capture. 

2. Contaminants of low toxicity or of nuisance value only. 

3. Intermittent, low production. 

4. Large hood-large air mass in motion. 



proper choice of values depends on several factors: 
Upper End of Range 

1. Disturbing room air currents. 

2. Contaminants of high toxicity. 

3. High production, heavy use. 

4. Small hood-local control only. 



the inward air flow created by the exhaust hood. Air flow 
toward the hood opening must be sufficiently high to maintain 
control of the contaminant until it reaches the hood. External 
air motion may disturb the hood-induced air flow and require 
higher air flow rates to overcome the disturbing effects. 
Elimination of sources of external air motion is an important 
factor in achieving effective control without the need for 
excessive air flow and its associated cost. Important sources 
of air motion are 

e Thermal air currents, especially from hot processes or 
heat-generating operations. 

• Motion of machinery, as by a grinding wheel, belt 
conveyor, etc. 

• Material motion, as in dumping or container filling. 
a Movements of the operator. 

• Room air currents (which are usually taken at 50 fpm 
minimum and may be much higher). 

• Rapid air movement caused by spot cooling and heating 
equipment. 

The shape of the hood, its size, location, and rate of air flow 
are important design considerations. 

3.4.1 Capture Velocity: The minimum hood-induced air 
velocity necessary to capture and convey the contaminant into 
the hood is referred to as capture velocity. This velocity will 
be a result of the hood air flow rate and hood configuration. 

Exceptionally high air flow hoods (example, large foundry 
side-draft shakeout hoods) may require less air flow than 
would be indicated by the capture velocity values recom- 
mended for small hoods. This phenomenon may be ascribed 
to: 



• The fact that the contaminant is under the influence of 
the hood for a much longer time than is the case with 
small hoods. 

• The fact that the large air flow rate affords considerable 
dilution as described above. 

Table 3-1 offers capture velocity data. Additional informa- 
tion is found in Chapter 10. 

3.4.2 Hood Flow Rate Determination: Within the bounds 
of flanges, baffles, adjacent walls, etc., air will move into an 
opening under suction from all directions. For an enclosure, 
the capture velocity at the enclosed opening(s) will be the 
exhaust flow rate divided by the opening area. The capture 
velocity at a given point in front of the exterior hood will be 
established by the hood air flow through the geometric surface 
which contains the point. 

As an example, for a theoretical unbounded point suction 
source, the point in question would be on the surface of a 
sphere whose center is the suction point (Figure 3-4). 

The surface area of a sphere is 4nX 2 . Using V = Q/A 

SURFACE OF 
SPHERE 

f POINT SUCTION 
SOURCE 



CAPTURE 
VELOCITY 




The presence of a large air mass moving into the hood. FIGURE 3-4. Point suction source 



Local Exhaust Hoods 



3-7 



(Equation 1 .3), the velocity at point X on the sphere's surface 
can be given by 

Q = V(4tcX 2 ) = 1257VX 2 [3.1] 

where: 

Q = air flow into suction point, cfm 

V = velocity at distance X, fpm 

A = 4kX2 = area of sphere, ft 2 

X = radius of sphere, ft 

Similarly, if an unbounded line source were considered, the 
surface would be that of a cylinder and the flow rate (neglect- 
ing end effects) would be 



Q-V(27tXL 2 ) = 6.28 VXL 



[3.2] 



where: 

L = length of line source, ft 

Equations 3.1 and 3.2 illustrate, on a theoretical basis, the 
relationship between distance, flow, and capture velocity and 
can be used for gross estimation purposes. In actual practice, 
however, suction sources are not points or lines, but rather 
have physical dimensions which cause the flow surface to 
deviate from the standard geometric shape. Velocity contours 
have been determined experimentally. Flow (33) for round 
hoods, and rectangular hoods which are essentially square, 
can be approximated by 

Q = V(10X 2 +A) [3.3] 

where: 

Q = air flow, cfm 

V = centerline velocity at X distance from hood, fpm 

X = distance outward along axis in ft. (NOTE: equa- 
tion is accurate only for limited distance of X, 
where X is within 1.5 D) 

A = area of hood opening, ft 2 

D = diameter of round hoods or side of essentially 
square hoods, ft 

Where distances of X are greater than 1.5 D, the flow rate 
increases less rapidly with distance than Equation 3.3 indi- 

cates. (3 - 4 - 3 - 5) 

It can be seen from Equation 3.3 that velocity decreases 
inversely with the square of the distance from the hood (see 
Figure 3-5.) 

Figures 3-6 and 3-7 show flow contours and streamlines 
for plane and flanged circular hood openings. Flow contours 
are lines of equal velocity in front of a hood. Similarly, 
streamlines are lines perpendicular to velocity contours. (The 
tangent to a streamline at any point indicates the direction of 
air flow at that point.) 

Flow capture velocity equations for various hood configu- 
rations are provided in Figures 3-8, 3-9, 3-10, and 3-11. 



3.4.3 Effects of Flanges and Baffles: A flange is a sur- 
face at and parallel to the hood face which provides a barrier 
to unwanted air flow from behind the hood. A baffle is a 
surface which provides a barrier to unwanted air flow from 
the front or sides of the hood. 

If the suction source were located on a plane, the flow area 
would be reduced (1/2 in both cases), thereby decreasing the 
flow rate required to achieve the same velocity. A flange 
around a hood opening has the same effect of decreasing the 
required flow rate to achieve a given capture velocity. In 
practice, flanging can decrease flow rate (or increase velocity) 
by approximately 25% (see Figures 3-6, 3-7, and 3-1 1). For 
most applications, the flange width should be equal to the 
square root of the hood area ( VA ). 

Baffles can provide a similar effect. The magnitude of the 
effect will depend on the baffle location and size. 

Figure 3-11 illustrates several hood types and gives the 
velocity/flow formulas which apply. 

A summary of other equations for hood velocity and the 
impact of cross-drafts on hood performance can be found in 
Reference 3.25. 

3.4.4 Air Distribution: Slot hoods are defined as hoods 
with an opening width-to-length ratio (W/L) of 0.2 or less. 
Slot hoods are most commonly used to provide uniform 



0.5 rrr 5 /3 NEEDE 
SOURCE 

O 




2 m J /s NEEDED 



SOURCE 
A'"C 




2X 



...-H 



LOCATION 



PLACE HOOD AS CLOSE TO THE SOURCE OF 
CONTAMINANT AS POSSIBLE. THE REQUIRED 
VOLUME VARIES WITH THE SQUARE OE THE 
DISTANCE FROM THE SOURCE, 

FIGURE 3-5. Flow rate as distance from hood 



3-8 



Industrial Ventilation 



























































































/ 




~W 
















( 














V 


-o 
o 




'O 
(.0 




or 




K 






A 






















V 






V 


y 




/ „ 


--^ 


1- ~ 




\ 















































































50 

% OF DIAMETER 



100 



FIGURE 3-6. Velocity contours — plain circular opening — % of opening 
velocity 

exhaust air flow and an adequate capture velocity over a finite 
length of contaminant generation, e.g., an open tank or over 
the face of a large hood such as a side-draft design. The 
function of the slot is solely to provide uniform air distribu- 
tion. Slot velocity does not contribute toward capture velocity. 
A high slot velocity simply generates high pressure losses. 
Note that the capture velocity equation (Figure 3-11) shows 
that capture velocity is related to the exhaust volume and the 
slot length, not to the slot velocity. 

Slot hoods usually consist of a narrow exhaust opening and 
a plenum chamber. Uniform exhaust air distribution across 
the slot is obtained by sizing slot width and plenum depth so 
that velocity through the slot is much higher than in the 
plenum. Splitter vanes may be used in the plenum; however, 
in most industrial exhaust systems, vanes are subject to cor- 
rosion and/or erosion and provide locations for material to 
accumulate. Adjustable slots can be provided but are subject 
to tampering and maladjustment. The most practical hood is 
the fixed slot and unobstructed plenum type. The design of 
the slot and plenum is such that the pressure loss through the 
slot is high compared with the pressure loss through the 
plenum. Thus, all portions of the slot are subjected to essen- 
tially equal suction and the slot velocity will be essentially 
uniform. 

There is no straightforward method for calculating the 
pressure drop from one end to the other of a slot-plenum 
combination. A very useful approximation, applicable to most 



hoods, is to design for a maximum plenum velocity equal to 
one-half of the slot velocity. For most slot hoods, a 2000 fpm 
slot velocity and 1000 fpm plenum velocity is a reasonable 
choice for uniformity of flow and moderate pressure drop. 
Centered exhaust take-off design results in the smallest prac- 
tical plenum size since the air approaches the duct from both 
directions. Where large, deep plenums are possible, as with 
foundry shake-out hoods, the slot velocity may be as low as 
1000 fpm with a 500 fpm plenum velocity, 

3.4.5 Rectangular and Round Hoods: Air distribution 
for rectangular and round hoods is achieved by air flow within 
the hood rather than by pressure drop as for the slot hood. The 
plenum (length of hood from face to tapered hood to duct 
connection) should be as long as possible. The hood take-off 
should incorporate a 60° to 90° total included tapered angle. 
Multiple take-offs may be required for long hoods. End 
take-off configurations require large plenum sizes because all 
of the air must pass in one direction. 

Figures 3-12 and 3-13 provide a number of distribution 
techniques. 

3.4.6 Worker Position Effect: The objective of industrial 
ventilation is to control the worker's exposure to toxic air- 
borne pollutants in a safe, reliable manner. As one of the main 




% OF DIAMETER 

FIGURE 3-7. Velocity contours — flanged circular opening — % of opening 
velocity 



Local Exhaust Hoods 



3-9 




-SOURCE 



r 




SOURCE 



y\ 



2X 




FREELY SUSPENDED HOOD 

Q = V(10X 2 + A) 



LARGE HOOD 

EARGE HOCD, X SMALL—MEASURE X 

PERPENDICULAR TO HOOD PACE, NOT LESS 
THAN 2X PROM HOOD EDGE. 



SOURCE -^ 



\ 

X 



* X -» 




SOURCE 



HOOD ON BENCH OR FLOOR 

Q = V(5X 2 + A) 



X— * 




ELANGE WIDTH > vV\ 



J_ 



IOOD WITH" WIDE FLANGE 

O - V 0,75(1 OX 2 -I- A) 



SUSPENDED HOODS 

(SMALL SIDE-DRAFT HOODS) 



Q - REQUIRED EXHAUST AIR FLOW, OEM. 

X = DISTANCE PROM HOOD FACE TO FARTHEST POINT OF CONTAMINANT RELEASE, FT. 

A = HOOD r ACE AREA, FT 2 . 

V - CAPTURE VELOCITY, FPM, AT DISTANCE X. 

NOTE: AIR FLOW RATE MUST INCREASE AS THE SQUARE OF DISTANCE OF "'HE SOURCE FROM THE HOOD. 
BAFFLING BY FLANGING OR BY PLACING ON BENCH, FLOOR, ECT. HAS A BENEFICIAL EFFECT. 



f^Q 




/45 MINIMUM 

1 

-a~\ |-a- - 0.4 D 



CANOPY HOOD 

= 1.4 PDV(P = PERIMETER OF TANK, FT). 

NOT RECOMMENDED IE WORKERS VilJST BEND OVER SOURCE, v RANGES 

FROM 50 TO 500 FPM DEPENDING ON CROSSDRAFTS. SIDE CURTAINS ON TWO OR THREE SIDES TO 

CREATE A SEMI-BOOTH OR BOOTH ARE DESIRABLE. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



FLO W /CAPTURE VELOCITY 



DATE 



1-88 



FICURF 



3 8 



3-10 



Industrial Ventilation 



^2L 



T/: 



^#— x--£ 



SOURCE 






FREELY SUSPENDED SLOT 

O - G.7LVX 



SLOT 



SOURCE- 



- x / — FLANGE (WIDTH > -Jk) 



\^ 



FLANGED SLO r 

- Z61.VX 




TANK 



Ua_ w -*-J 



TANK 



SLOT ON TANK 

= CLW 

ONE HALF IN EACH SLOT IF" 

SLOTS ON BOTH SIDES 



Q - REQUIRED EXHAUST FLOW RATE, CFIvi 
X - DISTANCE, HOOD FACE TO FARTHEST 

POINT OF SOURCE (USUALLY ON 

CENTERLINE OF HOOD), FT. 
V - CAPTURE VELOCITY AT DISTANCE X, FPM. 
L - LENGTH, OF HOOD, SLOT, TABLE, TANK, ETC.. F 
W - WIDTH, OF TABLE, TANK. ETC., FT. 
A = HOOD FACE AREA. FT. 



JVNGED SLO r 

Q - CLW 




FLANGED SLO r 

- CLW 



L - LENGTH OF SLOT, FT. 

W - WIDTH OF TABLE OR TANK, F :. 

C - CONSTANT, VARIES FROM SO YO SOO. 
USUAL CHOICE iS ISO TO 250. 
FLANGED SI...OTS REQUIRE LOWEST 
EXHAUST. SEE CHAPTER 10. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



FLO W '/CAPTURE VELOCITY 



DATE J _ QQ 



FIGURE 



Q — 9 



o 



Local Exhaust Hoods 3-11 



SOURCE CLOSE TO GRILLE 



/-' 




SOURCE EAR 
FROM GRILLE 



m. 



=■-■■■■ V A 




0-0 OX" -i A)V 



■ IMRAR TO BOOTH 



R TO SUSPENDED HOOD 



DOWNDRAFT HOODS 



NOT R ROOM MENDED FOR HOT OR HEAT PRODUCING OPERATIONS IF DOWNDRAFT AREA 

IS LARGE, SEE "CAPTURE VELOCITY" IN THIS SECTION. 



-^ 



" ' 


""■'■■■■ ■ — — ~ ■■ 




i 

E 

r ""'" 

1 


_ __;__; 




/ V A ! 






t 






T -= ::B tuA \ 


• i 



ANGLE BAFFLE 



M I N 



V 



i 



) 

A 



BOOTH- TYPE HOODS 



O-AV (A--tFACE AREA, FTO V=FACE VELOCITY, FPM). 

BAFFLES ARE OPTIONAL FOR AIR DISTRIBUTION; NOT RROUIRED IF A WATEF 

IF OTHER MEANS FOR DISTRIBUTION IS PROVIDED. 
S VARIES FROM 4 INCHES TO 8 INCHES, DEPENDING ON SIZE O r BOOTH. 
T VARIES FROM 6 INCHES TO 12 INCHES, DEPENDING ON SIZE OF BOOTH. 
INCREASE THE NUMBER OF PANELS WITH SIZE OF BOOTH. 



WALL BOOTH OR 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYG.lENi.STS 



FLO W/ CAPTURE \ t.'LOClTY 



DATE 



1-88 



FIGURE 



3 10 



3-12 Industrial Ventilation 



HOOD TYPE 






A = WL (ft ) 




// 




y^ 






H 


^ 




y^ 



w 






DESCRIPTION 



SLOT 



FLANGED SLOT 



PLAIN OPENING 



FLANGED OPENING 



BOOTH 



CANOPY 



PLAIN MULTIPLE 

SLOT OPENING 

2 OR MORE SLOTS 



FLANGED MULTIPLE 

SLOT OPENING 
2 OR MORE SLOTS 



ASPECT RATIO, W/L 



0.2 OR LESS 



0.2 OR LESS 



0.2 OR GREATER 
AND ROUND 



0.2 OR GREATER 
AND ROUND 



TO SUIT WORK 



TO SUIT WORK 



0.2 OR GREATER 



0.2 OR GREATER 



AIR FLOW 



Q = 3.7 LVX 



Q = 2.6 LVX 



Q = V(10X -FA) 



Q = 0.75V(10X + A) 



Q = VA = VWH 



Q = 1.4 PVD 
SEE FIG. VS-99-03 
P - PERIMETER 
D - HEIGHT 

ABOVE WORK 



Q = V(10X + A) 



Q = 0.75V(10X + A) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HOOD TYPES 



DATE 4-96 



FIGURE 



3-11 



Local Exhaust Hoods 3-13 



INSIDE RADIUS MORE IMPORTANT 

THAN OUTSIDE 



A 



'" A 



SLOT VELOCITY 2000 FPM 
OR HIGHER. 



VELOCITY- 1 



MAX. PLENUM 
2 SLOT VELOCIT 



:2 Mih 



r 




~V— 



SECTION 



SLOPE FOR DRAINING IS DESIRABLE — 
SLOPE DOES NOT AID IN DISTRIBUTION 



DISTRIBUTION BY SLOT RESISTANC. 



rs 





TANK 



■-Q 



H 



DISTRIBUTION BY FISH TAIL 

WITH LOW PLENUM VELOCITIES AND HIGH SLOT VELOCITIES, GOOD DISTRIBUTION IS OBTAINED. 
SLOTS OVER 10 FEET TO 12 FEET IN LENGTH USUALLY NEED MULTIPLE TAKE-OFFS. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL ITYGIENISTS 



DISTRIB UTION TECHNIQ UE 



'Q 



DATE ]-gQ 



T 



'I CURE 



3-12 



3-14 Industrial Ventilation 



BAF 



[Mil 

\ SnA., JL 



DiSTRIBUTlON BY BAFFLE 
SBB FIG. 3 A3 



^ 






LONG B001 



FF0N T' MIF "R ~ " [ AK 






BOO I'M CANO. 

.AMF PRINCIPLES APPLY TO CANOPY TYPE ) 



SLOT VELOCITY 2000 FPM OR HICHFR 




/V 


a 


# 


/ 


^\ 



30" to 45' 
A. 



JA 



DISTRIBUTION BY 
A.OT (OR BAFAi.ES) 



DISTRIBUTION BY TAPER 



SIDE- DRAFTS AND SUSPENDED HOODS 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DIS TRIE UTION TECHNIQ UES 



DATE 



1-88 \ 



FIGURE 



u — 1 t J 



Local Exhaust Hoods 



3-15 



engineering controls, local exhaust ventilation is designed to 
be near the point of contaminant generation. Often, considera- 
tion is not given to how the workers will position themselves 
with respect to the air flow. Studies* 3 - 6 " 3 - 9 ) show that the posi- 
tion of the worker with respect to the flow direction is an 
important parameter in determining the breathing zone con- 
centration. 

Figure 3-14, Position 2, shows a worker oriented with his 
back to the air flow. Immediately downstream of the worker, 
a zone of reverse flow and turbulent mixing occurs due to 
boundary layer separation. Contaminant released into this 
region (e.g., from a hand-held or proximal source) will be 
mixed into the breathing zone resulting in exposure. Figure 
3-14, Position 1, shows a worker oriented at 90° to the flow 
direction; here, the reverse flow zone forms to the side and 
there is less opportunity for the entrainment of contaminant 
into the breathing zone. 

Studies suggest that this phenomenon is important when 
large booth-type hoods are employed or in situations where 
there is a reasonably uniform air flow. Exposure studies (3J0) 
using a tailor's mannequin to simulate an operator in a booth- 
type hood used for the transfer of powders showed, in ail 
cases, that exposures for Position 1 were less than those in 
Position 2 by at least a factor of 2000. 

A second case study (3 6) reported women who used a spray 
and brush application of a chloroform-based adhesive were 
significantly exposed despite working in a ventilated booth. 
A 50% reduction in exposure was found when the workers 
stood side-on to the air flow (Position 1). Subsequent modifi- 
cation of spray practices resulted in a determination that a 30° 
angle to the air flow and holding the nozzle in the downstream 
hand seemed optimal. No alterations to the actual design or 
air flow of the booth were needed to achieve acceptable 
exposure levels. 

The preceding discussion assumes that the worker is not in 
the wake of an upstream object and that the contaminant 
source has negligible momentum. In cases where the contami- 
nant source has significant momentum (e.g., high-pressure 
compressed air paint spray operations), the effect of position 
on exposure may be reversed — i.e., Position 1 inFigure3-14 
may produce higher exposures. This is associated with the 
deflection of the spray upstream of the worker and subsequent 
recirculation through the breathing zone. Further research and 
field studies are needed to evaluate the tendency for reverse 
flow to occur in more complex situations. Although the im- 
portance of boundary layer separation effects with smaller 
local exhaust hoods has not been thoroughly explored, three 
studies O- 11 - 3 - 13 ) suggest that the 90° orientation is beneficial 
even in this instance. It is recommended that the side orienta- 
tion (i.e., Position 1) be the preferred orientation in situations 
where feasible. Down-draft configurations may provide simi- 
lar benefits under certain conditions. 

It is recommended that the side orientation (i.e., Position 





- 






Source 

(i 


Airflow -4— 


\x ,v \\ 




y~- yy..y 




Position // 1 

'" \ 


Airflow — it-— 


^ J} &> Source 

V<p ^ 









Position jj2 
FIGURE 3-14. Worker position effect 

1) be investigated as a preferred work practice where feasible. 
It is important to assess the exposure with personal sampling 
pumps to confirm the benefits of one position versus another 
as other factors may complicate the issue. 

3.5 HOOD LOSSES 

Plain duct openings, flanged duct openings, canopies, and 
similar hoods have only one significant energy loss. As air 
enters the duct, a vena contracta is formed and a small energy 
loss occurs first in the conversion of static pressure to velocity 
pressure (see Figure 3-1 5.) As the air passes through the vena 
contracta, the flow area enlarges to fill the duct and velocity 
pressure converts to static pressure. At this point, the uncon- 
trolled slow down of the air from the vena contracta to the 
downstream duct velocity results in the major portion of the 
entry loss. The more pronounced the vena contracta, the 
greater will be the energy loss and hood static pressure. 

Compound hoods are hoods which have two or more points 
of significant energy loss and must be considered separately 
and added together to arrive at the total loss for the hood. 
Common examples of hoods having double entry losses are 
slot-type hoods and multiple-opening, lateral draft hoods 
commonly used on plating, paint dipping and degreasing 
tanks, and foundry side-draft shakeout ventilation. 

The hood entry loss (h e ) can be expressed, therefore, in 
terms of hood loss coefficients (F f ) which, when multiplied 
by the slot or duct velocity pressure (VP), will give the entry 
loss (h e ) in inches of water. The hood static pressure is equal 



3-16 



Industrial Ventilation 






/ 


\ 












^ 






VP 


Q . 


, _^"^ 














- ^ ^ 




-—-..._ 






TP 




\ 






SP 

















FIGURE 3-15. Airflow at the vena contracta 

to the hood entry loss plus the velocity pressure in the duct. 
The hood entry loss represents the energy necessary to over- 
come the losses due to air moving through and into the duct. 
The velocity pressure represents the energy necessary to 
accelerate the air from rest to duct velocity (see Chapter 1, 
Section 1.6, "Acceleration of Air and Hood Entry Losses.") 
This may be expressed as: 



SP h -h P +VP H 



SP h ^(F s )(VP s ) + (F d )(VP d )-fVP d 



[3.4] 



[3.5] 



where: 

h e = overall hood entry loss = h s + h d , "wg 

h s = slot or opening loss = (F S )(VP S ), "wg 

h d = duct entry loss = (F d )(VP d ), "wg 

F s = loss coefficient for slot 

F d = loss coefficient for duct entry 

VP S = slot or opening velocity pressure, "wg 

VP d = duct velocity pressure, "wg 

One exception can occur when the slot velocity (or other hood 
entry velocity is higher than is the duct velocity. In such case, 
the acceleration velocity pressure used in determining SP is 
the higher slot or opening velocity pressure. 

Figures 3-16 and 5-13 give hood entry loss coefficients for 
several typical hood types. 



3.5.1 Simple Hoods: A simple hood is shown in Figure 
3-17. If the hood face velocity for such a simple hood is less 
than 1000 fpm, h s will be negligible and the loss will be 
dependent on h d only. If the hood face velocity is greater than 
1000 fpm, both h s and h h should be considered. Face velocities 
greater than 1000 fpm will usually only occur with relatively 
small hood face areas (0.25 to 0.50 ft 2 ). 



EXAMPLE PROBLEM 



_ Q 



Given: Face Velocity (V f ) =_rL = 250 fpm 



Duct Velocity (V d ) =_9_ = 3000 fpm 



VP d =f^ 



^4005 



= 0.56 "wg 



Fd = 0.25 as shown in Figure 5-12 
SPh = h d + VP d 

= (0.25)(0.56) + 0.56 
= 0.70"wg 

3.5.2 Compound Hoods: Figure 3-18 illustrates a double 
entry loss hood. This is a single slot hood with a plenum and 
a transition from the plenum to the duct. The purpose of the 



Local Exhaust Hoods 3-17 



HOOD TYPE 



DESCRIPTION 




:) L.A!N OPENING 



HOOD ENTRY LOSS (F ,,) 



FLANGED OPENING 





PAPER OR CONE 
HOOD 



0.9: 



0.4S 



SEE CHAPTER 10 



BELL MOUTH 
INLET 




<£_\L>^ 



Hi 




ORIFICE 



"1 " 



TYPICAL. GRINDING 
HOOD 



4= 



0.04 



SEE CHAPTER 10 



(STRAIGHT TAKEOFF) 
0.65 



(TAPERED TAKEOFF) 
0.40 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HOOD LOSS FACTORS 



DATE 



4-96 | nGURE 3-~ 



3-18 



Industrial Ventilation 



m. 



.,05$ 



■'AOL 



FIGURE 3-17, Simple hood 

plenum is to give uniform velocity across the slot opening. 
Air enters the slot, in this case a sharp-edged orifice, and loses 
energy due to the vena contracta at this point. The air then 
continues through the plenum where the greater portion of the 
slot velocity is retained because the air stream projects itself 
across the plenum in a manner similar to the ''blowing" supply 
stream shown in Figure 1-7. (The retention of velocity in the 
plenum is characteristic of most local exhaust hoods because 
of the short plenum length.) In the case of very large hoods or 
exhausted closed rooms, however, the velocity loss must be 
taken into account. Finally, the air converges into the duct 
through the transition where the second significant energy 
loss occurs. For this type hood, both h s and h h must be 
considered. 

EXAMPLE PROBLEM 

Given; Slot Velocity (V s ) = 2000 fpm 

Duct Velocity (V d ) = 3500 fpm 

(Vd is greater than V s ; therefore, use VPd as 
the acceleration VP) 



VP = 



V c 



0.25 "wg 



,4005, 
F s for slot = 1.78 from Figure 5-12 

VP H = f-^M = °- 76 " w 9 

F d = 0.25 as shown in Figure 5-13 
SP h = hs + hd + VPd 

= (1.78)(0.25)+ (0.25) (0.76) + 0.76 
= 1.40 "wg 

3.6 MINIMUM DUCT VELOCITY 

The velocity pressure, VP d , utilized to determine hood 
losses in the previous examples is determined from the air 
velocity in the duct immediately downstream of the hood to 
duct connection. This velocity is determined by the type of 



material being transported in the duct. 

For systems handling particulate, a minimum design veloc- 
ity is required to prevent settling and plugging of the duct. On 
the other hand, excessively high velocities are wasteful of 
power and may cause rapid abrasion of ducts. (3M ~ 32i) Mini- 
mum recommended design velocities are higher than theoreti- 
cal and experimental values to protect against practical 
contingencies such as: 

1 . Plugging or closing one or more branch will reduce the 
total flow rate in the system and correspondingly will 
reduce the velocities in at least some sections of the 
duct system. 

2. Damage to ducts, by denting for example, will increase 
the resistance and decrease the flow rate and velocity 
in the damaged portion of the system. 

3. Leakage of ducts will increase flow rate and velocity 
downstream of the leak but will decrease air flow 
upstream and in other parts of the system. 

4. Corrosion or erosion of the fan wheel or slipping of a 
fan drive belt will reduce flow rates and velocities. 

5. Velocities must be adequate to pick up or re-entrain 
dust which may have settled due to improper operation 
of the exhaust system. 

The designer is cautioned that for some conditions such as 
sticky materials, condensing conditions in the presence of 
dust, strong electrostatic effects, etc., velocity alone may not 
be sufficient to prevent plugging, and other special measures 
may be necessary. 

Some typical duct velocities are provided in Table 3-2. The 
use of minimum duct velocity is treated in detail in Chapter 5. 

3.7 SPECIAL HOOD REQUIREMENTS 

3.7.1 Ventilation of Radioactive and High Toxicity 
Processes: Ventilation of radioactive and high toxicity 
processes requires a knowledge of the hazards, the use of 
extraordinarily effective control methods, and adequate main- 
tenance which includes monitoring. Only the basic principles 



TRANS 
LOSS 



j / SL0! lOSS 

L SLOT VSlOCOv 

FIGURE 3-18. Compound hood 



Local Exhaust Hoods 



3-19 



TABLE 3-2. Range of Minimum Duct Design Velocities 



Nature of Contaminant 



Examples 



Design Velocity 



Vapors, gases, smoke 

Fumes 

Very fine light dust 

Dry dusts & powders 

Average industrial dust 

Heavy dusts 
Heavy or moist 



Ail vapors, gases, and smoke 

Welding 

Cotton lint, wood flour, litho powder 

Fine rubber dust, Bakelite molding powder dust, jute lint, cotton dust, 
shavings (light), soap dust, leather shavings 

Grinding dust, buffing lint (dry), wool jute dust (shaker waste), coffee 
beans, shoe dust, granite dust, silica flour, general material handling, 
brick cutting, clay dust, foundry (general), limestone dust, packaging and 
weighing asbestos dust in textile industries 

Sawdust (heavy and wet), metal turnings, foundry tumbling barrels and 
shake-out, sand blast dust, wood blocks, hog waste, brass turnings, cast 
iron boring dust, lead dust 

Lead dusts with small chips, moist cement dust, asbestos chunks from 
transite pipe cutting machines, buffing lint (sticky), quick-lime dust 



Any desired velocity 

(economic optimum velocity 

usually 1 000-2000 fpm) 

2000-2500 

2500-3000 

3000^000 

3500-4000 



4000-4500 



4500 and up 



can be covered here. For radioactive processes, reference 
should be made to the standards and regulations of the nuclear 
regulatory agencies. 

Local exhaust hoods should be of the enclosing type with 
the maximum enclosure possible. Where complete or nearly 
complete enclosure is not possible, control velocities from 50 
to 100% higher than the minimum standards in this manual 
should be used. If the enclosure is not complete and an 
operator must be located at an opening, such as in front of a 
laboratory hood, the maximum control velocity should not 
exceed 125 fpm. Air velocities higher than this value will 
create eddies in front of the operator which may pull contami- 
nant from the hood into the operator's breathing zone. Re- 
placement air should be introduced at low velocity and in a 
direction that does not cause disruptive cross drafts at the hood 
opening. 

3.7.2 Laboratory Operations: Glove boxes should be 
used for high-activity alpha or beta emitters as well as highly 
toxic and biological materials. The air locks used with the 
glove box should be exhausted if they open directly to the 
room. 

For low-activity radioactive laboratory work, a laboratory 
fume hood may be acceptable. For such hoods, an average 
face velocity of 80-100 fpm is recommended. See Section 

10.35, VS-35-01, -02, -04, and -20. 

For new buildings, it is frequently necessary to estimate the 
air conditioning early — before the detailed design and equip- 
ment specifications are available. For early estimating, the 
guidelines provided in Section 10.35 for hood air flow and 
replacement air flow can be used. 



3.8 PUSH-PULL VENTILATION 

Push-pull ventilation consists of a push nozzle and an 
exhaust hood to receive and remove the push jet. Push-pull 
is used most commonly on open surface vessels such as 
plating tanks (3 22) but may be effectively used elsewhere (see 
VS-70-10). The advantage of push-pull is that a push jet will 
maintain velocity over large distances, 20-30 ft or more, 
whereas the velocity in front of an exhaust hood decays very 
rapidly as the distance from the hood increases. Properly used, 
the push jet intercepts contaminated air and carries it relatively 
long distances into the exhaust hood, thus providing control 
where it may be otherwise difficult or impossible. 

3.8.1 Push Jet: Ambientair is entrained in the push jet and 
results in a jet flow at the exhaust hood several times greater 
than the push nozzle flow rate. The jet velocity will decay with 
distance from the nozzle. The entrainment ratio for a long thin 
slot-(or pipe) type nozzle may be approximated by: (3 - 23) 



Ox 



'12. 



t \ 
ax 

v b °y 



+ 0.41 



[3.6] 



The velocity ratio may be approximated by: 



1.2 



(3.23) 



V n lax _., 
+ 0.41 



[3.7] 



where: 

Q = the push nozzle supply flow 

Q x = the jet flow rate at a distance x from the nozzle 

V = the push nozzle exit air velocity 

V x = the peak push jet velocity at a distance x 



3-20 



Industrial Ventilation 



o 

-y 




PEAK VELOCITY 



JET 
CENTERLINE 



VELOCITY -*- 
(a) FREE PLANE JET 




WALL 
OR 
SURFACE 



VELOCITY ~&— 
(b) PLANE WALL JET 

FIGURE 3-19. Jet velocity profile 

a = a coefficient characteristic of the nozzle (0.13 

for slots and pipes) 
x = distance from the nozzle 
b = the slot width* 

[*.If the nozzle is freely suspended (free plane jet), b 
is equal to one-half the total slot width. If the nozzle 
is positioned on or very near a plane surface (wall jet), 
b is equal to the full slot width. For pipes with holes, 
b is the width of a slot with equivalent area.] 

Typical jet velocity profiles are shown in Figure 3-19. 

Obstructions in the jet path should be minimized near the 
jet. Objects with small cross-sections, such as parts hangers, 
will cause serious problems; however, large flat surface ob- 
jects should be avoided. At further distances from the nozzle 
where the jet has expanded, larger objects may be acceptable 
if they are located within the jet. 

The nozzle may be constructed as a long thin slot, a pipe 
with holes or individual nozzles. The total nozzle exit area 
should not exceed 50% of the nozzle plenum cross-sectional 



area to assure even flow distribution. Slot width can range 
from 0.125-0.25 inch for short push length such as plating 
tanks (4-8 ft). Hole size should be 0.25 inch on 3 to 8 diameter 
spacing. The nozzle momentum factor, which is proportional 
to nozzle exit flow per foot of nozzle length times nozzle exit 
velocity (Q x V ), must be sufficient to result in an effective 
jet but not so strong that the exhaust hood is overpowered. A 
Q o V range should be approximately 50,000-75,000 per foot 
of nozzle length for short distances of 4-8 feet. 

3.8.2 Pull Hood: The pull hood will accept and remove the 
push jet flow. The same design considerations regarding flow 
distribution, hood entry losses, etc., used for a normal pull- 
only hood should be used. The hood pull flow should be 
approximately 1 .5-2.0 times the push flow which reaches the 
hood. If design criteria specifying pull flow rate are not 
available, Equation 3.6 can be used. 

The hood opening height should be the same as the width 
of the expanded jet, if possible. However, smaller opening 
heights are acceptable if the hood flow rate meets the 1 .5- 2.0 
times jet flow criteria. 




HYPOTHETICAL 
POINT SOURCE 



FIGURE 3-20. Dimensions used to design high-canopy hoods for hot 
sources (Ref. 3.24) 



Local Exhaust Hoods 



3-2! 



Each push-pull application will necessitate special atten- 
tion. Wherever possible, a pilot system should be evaluated 
prior to final installation. 

3.8.3 Push-Pull System Design: Specific design criteria 
have been developed experimentally for plating, cleaning, or 
other open surface vessels and are provided in VS-70-10,VS- 
70-1 1, and VS-70-12. Where such specific design criteria are 
not available, the criteria provided in Sections 3.8.1 and 3.8.2 
can be used. When designing with Equation 3.7, a push jet 
velocity (V x ) of 1 50-200 fpm at the exhaust hood face should 
be specified. 

3.9 HOT PROCESSES 

Design of hooding for hot processes requires different 
considerations than design for cold processes.^ 3 24) When sig- 
nificant quantities of heat are transferred to the air above and 
around the process by conduction and convection, a thermal 
draft is created which causes an upward air current with air 
velocities as high as 400 fpm. The design of the hood and 
exhaust rate must take this thermal draft into consideration. 

3.9.1 Circular High Canopy Hoods: As the heated air 
rises, it mixes turbulently with the surrounding air. This 
results in an increasing air column diameter and volumetric 
flow rate. The diameter of the column (see Figure 3-20) can 
be approximated by: 



D =0.5X° £ 



[3-8] 



where: 

D c = column diameter at hood face 

X c = y + z = the distance from the hypothetical point 
source to the hood face, ft 

y = distance from the process surface to the hood 
face, ft 

z = distance from the process surface to the hypo- 
thetical point source, ft 

"z" can be calculated from: 



Z=(2A S ) 1138 
where: 



[3.9] 



A» = diameter of hot source, ft. 



The velocity of the rising hot air column can be calculated 
from: 



V f -8(A S ) 033 
where: 



(At) - 42 

X / 25 



[3.10] 



V f = velocity of hot air column at the hood face, fpm 

A s - area of the hot source, ft 2 

At= the temperature difference between the hot 



source and the ambient air, F 
X c = y + z = the distance from the hypothetical point 
source to the hood face, ft. 

The diameter of the hood face must be larger than the 
diameter of the rising hot air column to assure complete 
capture. The hood diameter is calculated from: 



D f = D c = 0.8y 
where: 



[3.11] 



D f = diameter of the hood face, ft 
Total hood air flow rate is 

Q t =V f A c+ V r (A f -A c ) [3i12 ] 

where: 

Q t = total volume entering hood, cfm 
V f = velocity of hot air column at the hood face, fpm 
A c = area of the hot air column at the hood face, ft 2 
V r = the required air velocity through the remaining 

hood area, fpm 
A f = total area of hood face, ft 2 

EXAMPLE PROBLEM 

Given: 4.0 ft diameter melting pot (D a ) 

1000 F metal temperature 

100 F ambient temperature 

Circular canopy hood located 10 ft above pot (y) 

Calculate x c : 

x c = y + z = y + (2D s ) 1138 

x c =10 + (2x4) 1 - 138 

x c = 10.7 ft 

Calculate the diameter of the hot air column at the hood 
face: 

D c = 0.5 xc 088 

D c = 0.5(20.7) 088 

D c = 7.2 ft 

Calculate the velocity of the hot air column at the hood face: 



V f = 8(A S 



(At) 



(x) u 



A s = 0.257tD c ^ 

A s = 0.25tc(4.2) 2 

A s =12.6 ft 2 

At =1000 -100 = 900 F 



3-22 



Industrial Ventilation 



V f - 8(1.26)' 
V f =(8)(231) 



0.33 (900) u 



(20.7) 025 
(17.4) 



(2.1 3) 

V f = 151fpm 
Calculate diameter of hood face: 

D f = D c + 0.8y 

D f =7.2 = 0.8 (10) 

Df= 15.2 ft 
Calculate total hood airflow rate 

Qf = VfA c + V r (Af -Ac) 
Ac = 0.25tiD c 2 
A c = 0.25tc(7.2) 2 
A c = 41 ft 2 

A f - 0.25rcDf 2 

Af=0.257i(7.2) 2 

A f = 181 ft 2 

Qf = 151(41) + 100(181 -41) 

Qf= 10,290 cfm 

3,9.2 Rectangular High Canopy Hoods: Hot air col- 
umns from sources which are not circular may be better 
controlled by a rectangular canopy hood. Hood air flow 
calculations are performed in the same manner as for circular 
hoods except the dimensions of the hot air column at the hood 
(and the hood dimensions) are determined by considering 
both the length and width of the source. Equations 3.8, 3.9, 
and 3.11 are used individually to determine length and width 
of the hot air column and the hood. The remaining values are 
calculated in the same manner as for the circular hood. 

EXAMPLE PROBLEM 

Given: 2.5 ft x 4 ft rectangular melting furnace 

700 F metal temperature 

80 F ambient temperature 

Rectangular canopy hood located 8 ft above 
furnace (y) 

Calculate X c for each furnace dimension. 

X c2 .5 = y + Z 2 .5 = y + (2D S 2.5) 1 - 138 

= 8 + (2x2.5) 1 138 

= 14.2 ft 



Xc4 = 8 + (2x4) 1 - 138 
= 18.7ft 
Calculate the width of the hot air column at the hood face. 

D C 2.5 - 0.5 Xc2.5 " 

= 0.5(14.2)° 88 

= 5.2 ft 
D C 4.o = 0.5(18.7)°- 88 

= 6.6 ft 
Calculate the velocity of the hot air column at the hood face. 

.x0.42 



V f -8(A S )' 



,0.33 (At) u 



v0.25 



(X c ) u 
A s = 2.5x4= 10 ft 2 

At = 700 - 80 = 620 F 
X c = x c2 .5 = 14.2ft 



Note: X c25 is used rather than x c4Q as it is smaller and 
as such will yield a slightly larger V r which results in a 
margin of safety. 



V = 8(10)' 



= 8 (2.1) 



n 33 (620)° 



(14.2) - 25 
(14.9) 



1.9 



= 132fpm 
Calculate hood face dimensions. 
Hood width = D C 2.5 + 0.8y 
= (5.2) + 0.8(8) 
= 11.6 ft 
Hood length = D C 4.o + 0.8y 
= 6.6 + 0.8(8) 
= 13.0 ft 
Calculate the total hood air flow rate. 

Qf = VfA c + V r (Af-A c ) 
A c = (D C 2.5)(Dc4.o) 

= (5.2)(6.6) 

= 34 ft 2 
Af = (hood length)(hood width) 

= (11.6)(13.0) 

= 151 ft 2 



Local Exhaust Hoods 



3-23 



Qf=(151)(34) + 100(151 -34) 
= 5134 + 11,700 
= 16,834 cfm 

3.9.3 Low Canopy Hoods: If the distance between the 
hood and the hot source does not exceed approximately the 
diameter of the source or 3 ft, whichever is smaller, the hood 
may be considered a low canopy hood. Under such conditions, 
the diameter or cross-section of the hot air column will be 
approximately the same as the source. The diameter or side 
dimensions of the hood therefore need only be 1 ft larger than 
the source. 

The total flow rate for a circular low canopy hood is 



[3.13] 



Q t -4.7(D f ) 233 (At) a41 

where: 

Q t = total hood air flow, cfm 
D f = diameter of hood, ft 

At = difference between temperature of the hot 
source, and the ambient, F. 

The total flow rate for a rectangular low hood is 

5i-6.2b 133 At a42 
L 

where: 

Q t = total hood air flow, cfm 
L = length of the rectangular hood, ft 
b = width of the rectangular hood, ft 
At = difference between temperature of the hot 
source and the ambient, F. 

REFERENCES 

3.1. Brandt, A.D.: Industrial Health Engineering. John 
Wiley and Sons, New York (1947). 

3.2. Kane, J.M.: Design of Exhaust Systems. Health and 
Ventilating 42:68 (November 1946). 

3.3. Dalla Valle, J.M.: Exhaust Hoods. Industrial Press, 

New York (1946). 

3.4. Silverman, L.: Velocity Characteristics of Narrow Ex- 
haust Slots. J. Ind. Hyg. Toxicol. 24:267 (November 
1942). 

3.5. Silverman, L.: Center-line Characteristics of Round 
Openings Under Suction. J. Ind. Hyg. Toxicol. 24:259 
(November 1942). 

3.6. Piney, M.; Gill, F.; Gray, C; et al.: Air Contaminant 
Control : the Case History Approach — Learning From the 
Past and Looking to the Future. In: Ventilation '88, J. H. 
Vincent, Ed., Pergammon Press, Oxford, U.K. (1989). 

3.7. Ljungqvist, B.: Some Observations on the Interaction 



Between Air Movements and the Dispersion of Pollu- 
tion. Document D8: 1979. Swedish Council for Build- 
ing Research, Stockholm, Sweden (1979). 

3.8. Kim, T.; Flynn, M.R.: Airflow Pattern Around a 
Worker in a Uniform Freestream. Am. Ind. Hyg. As- 
soc. J. 52:(7): 187-296 (1991). 

3.9. George, D.K.; Flynn, M.R.; Goodman, R.: The Impact 
of Boundary Layer Separation on Local Exhaust De- 
sign and Worker Exposure. Appl. Occup. Env. Hyg. 

5:501-509(1990). 

3.10. Heriot, N.R.; Wilkinson, J.: Laminar Flow Booths for 
the Control of Dust. Filtration and Separation 
16:2:159-164(1979). 

3.11. Flynn, MR.; Shelton, W.K.: Factors Affecting the 
Design of Local Exhaust Ventilation for the Control 
of Contaminants from Hand-held Sources. Appl. Oc- 
cup. Env. Hyg. 5:707-714 (1990). 

3.12. Turn Suden, K.D.; Flynn, M.R.; Goodman, R.: Com- 
puter Simulation in the Design of Local Exhaust 
Hoods for Shielded Metal Arc Welding. Am. Ind. Hyg. 
Assoc. J, 5 1(3): 115-1 26 (1990). 

3.13. American Welding Society: Fumes and Gases in the 
Welding Environment. F. Y. Speight and H. C. Camp- 
bell, Eds. AWS, Miami, FL (1979). 

3.14. American Society of Mechanical Engineers: Power 
Test Code 19.2.4: Liquid Column Gages. ASME 
(1942). 

3.15. Hemeon, W.C.L.: Plant and Process Ventilation. In- 
dustrial Press, New York (1963). 

3.16. Alden, J.L.: Design of Industrial Exhaust System. 
Industrial Press, New York (1939). 

3.17. Rajhans, G.S.; Thompkins, R.W.: Critical Velocities 
of Mineral Dusts. Canadian Mining J. (October 1967). 

3.18. Djamgowz, O.T.; Ghoneim, S.A.A.: Determining the 
Pick-Up Air Velocity of Mineral Dusts. Canadian 
Mining! (July 1974). 

3.19. Baliff, J.L.; Greenburg, L.; Stern, A.C.: Transport 
Velocities for Industrial Dusts — An Experimental 
Study. Ind. Hyg. Q. (December 1948). 

3.20. Dalla Valle, J.M.: Determining Minimum Air Veloci- 
ties for Exhaust Systems. Heating, Piping and Air 
Conditioning (1932). 

3.21. Hatch, T.F.: Economy in the Design of Exhaust Sys- 
tems. 

3.22. Hughes, R.T.: Design Criteria for Plating Tank Push- 
Pull Ventilation. In: Ventilation '85. Elsevier Press, 
Amsterdam (1986). 

3.23. Baturin, V.V.: Fundamentals of Industrial Ventilation. 
Pergamon Press, New York (1972). 



3-24 Industrial Ventilation 



3.24. U.S. Public Health Service: Air Pollution Engineering John Wiley & Sons, New York (1989). 

Manual. Publication No. 999-AP-40 (1973). 326 graconnier, R: Bibliographic Review of Velocity 

3.25. Burgess, W.A.; Ellenbecker, M.J.; Treitman, R.D.: Fields in the Vicinity of Local Exhaust Hoods. Am. 
Ventilation for Control of the Work Environment, Ind. Hyg. Assoc. J., 49(4): 185-1 98 (1988). 



Chapter 4 

AIR CLEANING DEVICES 



4.1 INTRODUCTION 4-2 

4.2 SELECTION OF DUST COLLECTION EQUIPMENT 4-2 

4.2.1 Contaminant Characteristics 4-2 

4.2.2 Efficiency Required . 4-2 

4.2.3 Gas Stream Characteristics 4-3 

4.2.4 Contaminant Characteristics 4-3 

4.2.5 Energy Considerations 4-3 

4.2.6 Dust Disposal 4-3 

4.3 DUST COLLECTOR TYPES . . 4-3 

4.3.1 Electrostatic Precipitators 4-3 

4.3.2 Fabric Collectors 4-9 

4.3.3 Wet Collectors 4-17 

4.3.4 Dry Centrifugal Collectors 4-18 

4.4 ADDITIONAL AIDS IN DUCT COLLECTOR 

SELECTION 4-22 

4.5 CONTROL OF MIST, GAS, AND VAPOR 
CONTAMINANTS 4-22 

4.6 GASEOUS CONTAMINANT COLLECTORS . . . 4-25 

4.6.1 Absorbers 4-25 

4.6.2 Adsorbers 4-25 



4.6.3 Thermal Oxidizers 4-25 

4.6.4 Direct Combustors 4-25 

4.6.5 Catalytic Oxidizers 4-25 

4.7 UNIT COLLECTORS 4-25 

4.8 DUST COLLECTING EQUIPMENT COST 4-25 

4.8.1 Price Versus Capacity 4-25 

4.8.2 Accessories Included 4-25 

4.8.3 Installation Cost 4-28 

4.8.4 Special Construction 4-28 

4.9 SELECTION OF AIR FILTRATION EQUIPMENT . 4-28 

4.9.1 Straining 4-28 

4.9.2 Impingement 4-32 

4.9.3 Interception 4-32 

4.9.4 Diffusion 4-32 

4.9.5 Electrostatic 4-32 

4. 1 RADIOACTIVE AND HIGH TOXICITY 

OPERATIONS 4-33 

4.11 EXPLOSION VENTING 4-33 

REFERENCES 4-34 



Figure 4-1 Dry Type Dust Collectors — Dust Disposal .... 4-4 Figure 4-11 

Figure 4-2 Dry Type Dust Collectors — Discharge Valves . . 4-5 

Figure 4-3 Dry Type Dust Collectors — Discharge Valves . . 4-6 Figure 4-12 

Figure 4-4 Electrostatic Precipitator, High Voltage Design . 4-7 

Figure 4-5 Electrostatic Precipitator, Low Voltage Design . 4-8 Figure 4-13 

Figure 4-6 Performance Versus Time Between Reconditionings Figure 4-14 

— Fabric Collectors 4-10 Figure 4-15 

Figure 4-7 Fabric Collectors 4-14 Figure 4-16 

Figure 4-8 Air Flow Through Fabric Collectors 4-15 Figure 4-17 

Figure 4-9 Fabric Collectors Pulse Jet Type 4-16 

Figure 4-10 Wet Type Collector (for Gaseous Contaminant) 4-19 



Wet Type Dust Collectors (for Particulate 

Contaminants) 4-20 

Wet Type Dust Collector (for Particulate 

Contaminants) 4-21 

Dry Type Centrifugal Collectors 4-23 

Range of Particle Size 4-24 

Unit Collector (Fabric— Shaker Type) 4-29 

Cost Estimates of Dust Collecting Equipment . 4-30 

Comparison Between Various Methods 

of Measuring Air Cleaning Capability 4-32 



4-2 



Industrial Ventilation 



4.1 INTRODUCTION 

Air cleaning devices remove contaminants from an air or 
gas stream. They are available in a wide range of designs to 
meet variations in air cleaning requirements. Degree of re- 
moval required, quantity and characteristics of the contami- 
nant to be removed, and conditions of the air or gas stream 
will all have a bearing on the device selected for any given 
application. In addition, fire safety and explosion control must 
be considered in all selections. (See NFPA publications.) 

For particulate contaminants, air cleaning devices are di- 
vided into two basic groups: AIR FILTERS and DUST COL- 
LECTORS. Air filters are designed to remove low dust 
concentrations of the magnitude found in atmospheric air. 
They are typically used in ventilation, air-conditioning, and 
heating systems where dust concentrations seldom exceed 1 .0 
grains per thousand cubic feet of air and are usually well 
below 0.1 grains per thousand cubic feet of air. (One pound 
equals 7000 grains. A typical atmospheric dust concentration 
in an urban area is 87 micrograms per cubic meter or 0.038 
grains per thousand cubic feet of air.) 

Dust collectors are usually designed for the much heavier 
loads from industrial processes where the air or gas to be 
cleaned originates in local exhaust systems or process stack 
gas effluents. Contaminant concentrations will vary from less 
than 0.1 to 100 grains or more for each cubic foot of air or 
gas. Therefore, dust collectors are, and must be, capable of 
handling concentrations 100 to 20,000 times greater than 
those for which air filters are designed. 

Small, inexpensive versions of all categories of air cleaning 
devices are available. The principles of selection, application, 
and operation are the same as for larger equipment. However, 
due to the structure of the market that focuses on small, 
quickly available, and inexpensive equipment, much of the 
available equipment is of light duty design and construction. 
One of the major economies of unit collectors implies recir- 
culation, for which such equipment may or may not be 
suitable. For adequate prevention of health hazards, fires, and 
explosions, application engineering is just as essential for unit 
collectors as it is for major systems. 

4.2 SELECTION OF DUST COLLECTION EQUIPMENT 

Dust collection equipment is available in numerous designs 
utilizing many different principles and featuring wide vari- 
ations in effectiveness, first cost, operating and maintenance 
cost, space, arrangement, and materials of construction. Con- 
sultation with the equipment manufacturer is the recom- 
mended procedure in selecting a collector for any problem 
where extensive previous plant experience on the specific dust 
problem is not available. Factors influencing equipment se- 
lection include the following: 

4.2.1 Contaminant Characteristics: Contaminants in ex- 
haust systems cover an extreme range in concentration and 



particle size. Concentrations can range from less than 0.1 to 
much more than 100,000 grains of dust per cubic foot of air. 
In low pressure conveying systems, the dust ranges from 0.5 
to 100 or more microns in size. Deviation from mean size (the 
range over and under the mean) will also vary with the 
material. 

4.2.2 Efficiency Required: Currently, there is no accepted 
standard for testing and/or expressing the "efficiency" of a 
dust collector. It is virtually impossible to accurately compare 
the performance of two collectors by comparing efficiency 
claims. The only true measure of performance is the actual 
mass emission rate, expressed in terms such as mg/m 3 or 
grains/ft 3 . Evaluation will consider the need for high effi- 
ciency-high cost equipment requiring minimum energy such 
as high voltage electrostatic precipitators, high effi- 
ciency-moderate cost equipment such as fabric or wet collec- 
tors, or the lower cost primary units such as the dry centrifugal 
group. If either of the first two groups is selected, the combi- 
nation with primary collectors should be considered. 

When the cleaned air is to be discharged outdoors, the 
required degree of collection can depend on plant location; 
nature of contaminant (its salvage value and its potential as a 
health hazard, public nuisance, or ability to damage property); 
and the regulations of governmental agencies. In remote 
locations, damage to farms or contribution to air pollution 
problems of distant cities can influence the need for and 
importance of effective collection equipment. Many indus- 
tries, originally located away from residential areas, failed to 
anticipate the residential building construction which fre- 
quently develops around a plant. Such lack of foresight has 
required installation of air cleaning equipment at greater 
expense than initially would have been necessary. Today, the 
remotely located plant must comply, in most cases, with the 
same regulations as the plant located in an urban area. With 
the present emphasis on public nuisance, public health, and 
preservation and improvement of community air quality, 
management can continue to expect criticism for excessive 
emissions of air contaminants whether located in a heavy 
industry section of a city or in an area closer to residential 
zones. 

The mass rate of emission will also influence equipment 
selection. For a given concentration, the larger the exhaust 
volumetric flow rate, the greater the need for better equip- 
ment. Large central steam-generating stations might select 
high efficiency electrostatic precipitators or fabric collectors 
for their pulverized coal boiler stacks while a smaller indus- 
trial pulverized fuel boiler might be able to use slightly less 
efficient collectors. 

A safe recommendation in equipment selection is to select 
the collector that will allow the least possible amount of 
contaminant to escape and is reasonable in first cost and 
maintenance while meeting all prevailing air pollution regu- 
lations. For some applications even the question of reasonable 



Air Cleaning Devices 4-3 



cost and maintenance must be sacrificed to meet established 
standards for air pollution control or to prevent damage to 
health or property. 

It must be remembered that visibility of an effluent will be 
a function of the light reflecting surface area of the escaping 
material. Surface area per pound increases inversely as the 
square of particle size. This means that the removal of 80% 
or more of the dust on a weight basis may remove only the 
coarse particles without altering the stack appearance. 

4.2.3 Gas Stream Characteristics: The characteristics of 
the carrier gas stream can have a marked bearing on equip- 
ment selection. Temperature of the gas stream may limit the 
material choices in fabric collectors. Condensation of water 
vapor will cause packing and plugging of air or dust passages 
in dry collectors. Corrosive chemicals can attack fabric or 
metal in dry collectors and when mixed with water in wet 
collectors can cause extreme damage. 

4.2.4 Contaminant Characteristics: The contaminant 
characteristics will also affect equipment selection. Chemi- 
cals emitted may attack collector elements or corrode wet type 
collectors. Sticky materials, such as metallic buffing dust 
impregnated with buffing compounds, can adhere to collector 
elements, plugging collector passages. Linty materials will 
adhere to certain types of collector surfaces or elements. 
Abrasive materials in moderate to heavy concentrations will 
cause rapid wear on dry metal surfaces. Particle size, shape, 
and density will rule out certain designs. For example, the 
parachute shape of particles like the "bees wings" from grain 
will float through centrifugal collectors because their velocity 
of fall is less than the velocity of much smaller particles 
having the same specific gravity but a spherical shape. The 
combustible nature of many finely divided materials will 
require specific collector designs to assure safe operation. 

4.2.5 Energy Considerations: The cost and availability 
of energy makes essential the careful consideration of the total 
energy requirement for each collector type which can achieve 
the desired performance. An electrostatic precipitator, for 
example, might be a better selection at a significant initial cost 
penalty because of the energy savings through its inherently 
lower pressure drop. 

4.2.6 Dust Disposai: Methods of removal and disposal of 
collected materials will vary with the material, plant process, 
quantity involved, and collector design. Dry collectors can be 
unloaded continuously or in batches through dump gates, 
trickle valves, and rotary locks to conveyors or containers. 
Dry materials can create a secondary dust problem if careful 
thought is not given to dust-free material disposal or to 
collector dust bin locations suited to convenient material 
removal. See Figures 4-1, 4-2, and 4-3 for some typical 
discharge arrangements and valves. 

Wet collectors can be arranged for batch removal or con- 



tinual ejection of dewatered material. Secondary dust prob- 
lems are eliminated, although disposal of wet sludge can be a 
material handling problem. Solids carry-over in waste water 
can create a sewer or stream pollution problem if waste water 
is not properly clarified. 

Material characteristics can influence disposal problems. 
Packing and bridging of dry materials in dust hoppers, floating 
or slurry forming characteristics in wet collectors are exam- 
ples of problems that can be encountered. 

4.3 DUST COLLECTOR TYPES 

The four major types of dust collectors for particulate 
contaminants are electrostatic precipitators, fabric collectors, 
wet collectors, and dry centrifugal collectors. 

4.3.1 Electrostatic Precipitators: In electrostatic precipi- 
tation, a high potential electric field is established between 
discharge and collecting electrodes of opposite electrical 
charge. The discharge electrode is of small cross-sectional 
area, such as a wire or a piece of flat stock, and the collection 
electrode is large in surface area such as a plate. 

The gas to be cleaned passes through an electrical field that 
develops between the electrodes. At a critical voltage, the gas 
molecules are separated into positive and negative ions. This 
is called "ionization" and takes place at, or near, the surface 
of the discharge electrode. Ions having the same polarity as 
the discharge electrode attach themselves to neutral particles 
in the gas stream as they flow through the precipitator. These 
charged particles are then attracted to a collecting plate of 
opposite polarity. Upon contact with the collecting surface, 
dust particles lose their charge and then can be easily removed 
by washing, vibration, or gravity. 

The electrostatic process consists of: 

1. Ionizing the gas. 

2. Charging the dust particles. 

3. Transporting the particles to the collecting surface. 

4. Neutralizing, or removing the charge from the dust 
particles. 

5. Removing the dust from the collecting surface. 

The two basic types of electrostatic precipitators are "Cot- 
trell," or single-stage, and "Penny," or two-stage (see Figures 
4-4 and 4-5). 

The "Cottrell," single-stage, precipitator (Figure 4-4) com- 
bines ionization and collection in a single stage. Because it 
operates at ionization voltages from 40,000 to 70,000 volts 
DC, it may also be called a high voltage precipitator and is 
used extensively for heavy-duty applications such as utility 
boilers, larger industrial boilers, and cement kilns. Some 
precipitator designs use sophisticated voltage control systems 
and rigid electrodes instead of wires to m inimize maintenance 
problems. 



4-4 Industrial Ventilation 



Collector 




Bag or 
collector sock 



Covered 
tote box 
or drum 



Covered drum or pail 
for dust removal 





Collector 



Vent to collector 
or inlet duct 




Covered 
tote box 
or drum 



Pug mill, sluice, 
pneumatic conveyor 
X| / or screw conveyor 




Enclosure 



Disposable ba 
or tote box 




Collapsed 

XI / bag 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DRY TYPE DUST COLLECTORS 
DUST DISPOSAL 



DATE 



3-97 1 FIGURE 4-1 



Air Cleaning Devices 4-5 



Hopper 



Handle 



For intermilfenf manual dumping 
where dust loads are light. 



% 



X 



\ 



/ 



^ / 

^ / 



Hopper 



DUST DOOR 



Rubber gasket 



Similar to dust door but designed 
for direct attachment to dusf chute, 
external pipe or canvas connection. 




Hopper 




DUST GATE 



For intermittent, manual dumping where 
dust loads are light. Flange for connection 
to dust disposal chute. 



SLIDE GATE 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DRY TYPE DUST COLLECTORS 
DISCHARGE VALVES 



DATE 1 _qq 



FIGURE 



4-2 



4-6 



Industrial Ventilation 



Hopper 




Curtain 



For continuous rernovai of collected dust where hopper 
is under negative pressure. Curtain is kept closed by 
pressure differential until collected material builds up 
sufficient height to overcome pressure. 



Hopper 



TRICKLE VALVE 



Motor driven multiple blade rotary valve provide air 
lock while continuously dumping collected materia!. 
Can be used with hoppers under either positive or 
negative pressure. Flanged for connection fo dust 
disposal chute. 




Gate 



ROTARY LOCK 



Gate 




Motor driven, double gate valve for continuous 
removal of collected dust. Gates are sequenced 
so only one is open at a time in order to provide 
air seal. Flanged for connection to dust disposal 
chute. 



DOUBLE DUMP VALVE 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DRY TYPE DUST COLLECTORS 
DISCHARGE VALVES 



DATE 



1-88 



FIGURE 



4-3 



Air Cleaning Devices 4-7 



ionizer Wires 



Inlet Nozzle 



Distribution Plates 



Plates 



High Voltage Rectifier 




^ Wire Tensioning Weights 



Hoppers 







/ 


- Collec 


tion 


plates 
















IT 










n l 






"" J 


+ 
















<r" 














To collect difficult dusts 


Airflow —- 




o 


7 








o4\ 




1 


2" 




Change treatment time 

1. Lengthen passage 

2. Lower velocities 








i 




« -~ 




i 




i 




3. Closer plate spacing 


Discharge e 


ec 


trode 


J 








t?—- -.^ 












r 










~i r 






~V1 


+ 












Firsi 


field 






s 


econd field 













AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



ELECTROSTATIC PRECIPITATOR 

HIGH VOLTAGE DESIGN 

(40,000 TO 75,000 VOLTS) 



DATE 1-88 



FIGURE 



4-4 



4-8 Industrial Ventilation 



Trash screen 

and distribution 

baffle 



Side access door 



Power pack 



Air flow 



Spray nozzle 
header 




C~3 



ZE 



ZH 



VtE 



r\ 



1/ 



Ionizer 
wire ' 



XM 







Insulator 



Plates 



Air flow- 



T-r~r 



-r 



¥ 



Discharge electrode- 



Collection plates 
(Grounded) 




- Grounded plates 

+ Charged plates 



+ r 

' T 



0.25" 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



ELECTROSTATIC PRECIPITATOR 

LOW VOLTAGE DESIGN 

(11,000 TO 15,000 VOLTS) 



DATE 1~88 J FIGURE ~--~~-^- 



Air Cleaning Devices 4-9 



The "Penny," or two-stage, precipitator (Figure 4-5) uses 
DC voltages from 1 1,000 to 14,000 volts for ionization and 
is frequently referred to as a low voltage precipitator. Its use 
is limited to low concentrations, normally not exceeding 
0.025 grains per cubic foot. It is the most practical collection 
technique for the many hydrocarbon applications where an 
initially clear exhaust stack turns into a visible emission as 
vapor condenses. Some applications include plasticizer ov- 
ens, forge presses, die-casting machines, and various welding 
operations. Care must be taken to keep the precipitator inlet 
temperature low enough to insure that condensation has al- 
ready occurred. 

For proper results, the inlet gas stream should be evaluated 
and treated where necessary to provide proper conditions for 
ionization. For high voltage units, a cooling tower is some- 
times necessary. Low voltage units may use wet scrubbers, 
evaporative coolers, heat exchangers or other devices to con- 
dition the gas stream for best precipitator performance. 

The pressure drop of an electrostatic precipitator is ex- 
tremely low, usually less than 1 "wg; therefore, the energy 
requirement is significantly less than for other techniques. 

4.3.2 Fabric Collectors: Fabric collectors remove par- 
ticulate by straining, impingement, interception, diffusion, 
and electrostatic charge. The "fabric" may be constructed of 
any fibrous material, either natural or man-made, and may be 
spun into a yarn and woven or felted by needling, impacting, 
or bonding. Woven fabrics are identified by thread count and 
weight of fabric per unit area. Non-woven (felts) are identified 
by thickness and weight per unit area. Regardless of construc- 
tion, the fabric represents a porous mass through which the 
gas is passed unidirectionally such that dust particles are 
retained on the dirty side and the cleaned gas passes on 
through. 

The ability of the fabric to pass air is stated as "permeabil- 
ity" and is defined as the cubic feet of air passed through one 
square foot of fabric each minute at a pressure drop of 0.5 
"wg. Typical permeability values for commonly used fabrics 
range from 25 to 40 cfm, 

A non-woven (felted) fabric is more efficient than a woven 
fabric of identical weight because the void areas or pores in 
the non-woven fabric are smaller. A specific type of fabric 
can be made more efficient by using smaller fiber diameters, 
a greater weight of fiber per unit area, and by packing the 
fibers more tightly. For non-woven construction, the use of 
finer needles for felting also improves efficiency. While any 
fabric is made more efficient by these methods, the cleanabil- 
ity and permeability are reduced. A highly efficient fabric that 
cannot be cleaned represents an excessive resistance to air 
flow and is not an economical engineering solution. Final 
fabric selection is generally a compromise between efficiency 
and permeability. 

Choosing a fabric with better cleanability or greater perme- 



ability but lower inherent efficiency is not as detrimental as 
it may seem. The efficiency of the fabric as a filter is mean- 
ingful only when new fabric is first put into service. Once the 
fabric has been in service any length of time, collected par- 
ticulate in contact with the fabric acts as a filter aid, improving 
collection efficiency. Depending on the amount of particulate 
and the time interval between fabric reconditioning, it may 
well be that virtually all filtration is accomplished by the 
previously collected particulate — or dust cake — as opposed 
to the fabric itself. Even immediately after cleaning, a residual 
and/or redeposited dust cake provides additional filtration 
surface and higher collection efficiency than obtainable with 
new fabric. While the collection efficiency of new, clean 
fabric is easily determined by laboratory test and the informa- 
tion is often published, it is not representative of operating 
conditions and therefore is of little importance in selectingthe 
proper collector. 

Fabric collectors are not 100% efficient, but well-designed, 
adequately sized, and properly operated fabric collectors can 
be expected to operate at efficiencies in excess of 99%, and 
often as high as 99.9% or more on a mass basis. The ineffi- 
ciency, or penetration, that does occur is greatest during or 
immediately after reconditioning. Fabric collector ineffi- 
ciency is frequently a result of by-pass due to damaged fabric, 
faulty seals, or sheet metal leaks rather than penetration of the 
fabric. Where extremely high collection efficiency is essen- 
tial, the fabric collector should be leak tested for mechanical 
leaks. 

The combination of fabric and collected dust becomes 
increasingly efficient as the dust cake accumulates on the 
fabric surface. At the same time, the resistance to air flow 
increases. Unless the air moving device is adjusted to com- 
pensate for the increased resistance, the gas flow rate will be 
reduced. Figure 4-6 shows how efficiency, resistance to flow, 
and flow rate change with time as dust accumulates on the 
fabric. Fabric collectors are suitable for service on relatively 
heavy dust concentrations. The amount of dust collected on a 
single square yard of fabric may exceed five pounds per hour. 
In virtually all applications, the amount of dust cake accumu- 
lated in just a few hours will represent sufficient resistance to 
flow to cause an unacceptable reduction in air flow. 

In a well-designed fabric collector system, the fabric or 
filter mat is cleaned or reconditioned before the reduction in 
air flow is critical. The cleaning is accomplished by mechani- 
cal agitation or air motion, which frees the excess accumula- 
tion of dust from the fabric surface and leaves a residual or 
base cake. The residual dust cake does not have the same 
characteristics of efficiency or resistance to air flow as new 
fabric. 

Commercially available fabric collectors employ fabric 
configured as bags or tubes, envelopes (flat bags), rigid 
elements, or pleated cartridges. Most of the available fabrics, 
whether woven or non-woven, are employed in either bag or 



4-10 



Industrial Ventilation 



£ a> 



>- D < 

-Q en- 






>- -C -t: 

u u « 

c c ^ 

"o - o 

0) 1- 3 



Collection efficiency 




Time, from last reconditioning 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



PERFORMANCE VS TIME 

BETWEEN RECONDITIONINGS 

FABRIC COLLECTORS 



DATE 



1-88 



FIGURE 



4-6 



Air Cleaning Devices 4-1 .1 



envelope configuration. The pleated cartridge arrangement 
uses a paper-like fiber in either a cylindrical or panel configu- 
ration. It features extremely high efficiency on light concen- 
trations. Earlier designs employed cellulose-based media. 
Today, more conventional media, such as polypropolene or 
spun-bonded polyester, are frequently used. 

The variable design features of the many fabric collectors 
available are: 

1 . Type of fabric (woven or non-woven). 

2. Fabric configuration (bags or tubes, envelopes, car- 
tridges). 

3. Intermittent or continuous service. 

4. Type of reconditioning (shaker, pulse-jet, reverse-air). 

5. Housing configuration (single compartment, multiple 
compartment). 

At least two of these features will be interdependent. For 
example, non-woven fabrics are more difficult to recondition 
and therefore require high-pressure cleaning. 

A fabric collector is selected for its mechanical, chemical, 
and thermal characteristics. Table 4-1 lists those charac- 
teristics for some common filter fabrics. 

Fabric collectors are sized to provide a sufficient area of 
filter media to allow operation without excessive pressure 
drop. The amount of filter area required depends on many 
factors, including: 

1. Release characteristics of dust. 

2. Porosity of dust cake. 

3. Concentration of dust in carrier gas stream. 

4. Type of fabric and surface finish, if any. 

5. Type of reconditioning. 

6. Reconditioning interval. 

7. Air flow pattern within the collector. 

8. Temperature and humidity of gas stream. 

Because of the many variables and their range of variation, 
fabric collector sizing is a judgment based on experience. The 
sizing is usually made by the equipment manufacturer, but at 
times may be specified by the user or a third party. Where no 
experience exists, a pilot installation is the only reliable way 
to determine proper size. 

The sizing or rating of a fabric collector is expressed in 
terms of air flow rate versus fabric media area. The resultant 
ratio is called "air to cloth ratio" with units of cfm per square 
foot of fabric. This ratio represents the average velocity of the 
gas stream through the filter media. The expression "filtration 
velocity" is used synonymously with air to cloth ratio for 
rating fabric collectors. For example, an air to cloth ratio of 
7: 1 (7 cfm/sq ft) is equivalent to a filtration velocity of 7 fpm. 

Table 4-2 compares the various characteristics of fabric 



collectors. The different types will be described in detail later 
Inspection of Table 4-2 now may make the subsequent dis- 
cussion more meaningful. The first major classification of 
fabric collectors is intermittent or continuous duty. Intermit- 
tent-duty fabric collectors cannot be reconditioned while in 
operation. By design, they require that the gas flow be inter- 
rupted while the fabric is agitated to free accumulated dust 
cake. Continuous-duty collectors do not require shut down for 
reconditioning. 

Intermittent-duty fabric collectors may use a tube, car- 
tridge, or envelope configuration of woven fabric and will 
generally employ shaking or vibration for reconditioning. 
Figure 4-7 shows both tube and envelope shaker collector 
designs. For the tube type, dirty air enters the open bottom of 
the tube and dust is collected on the inside of the fabric. The 
bottoms of the tubes are attached to a tube sheet and the tops 
are connected to a shaker mechanism. Since the gas flow is 
from inside to outside, the tubes tend to inflate during opera- 
tion and no other support of the fabric is required. 

Gas flow for envelope-type collectors is from outside to 
inside; therefore, the envelopes must be supported during 
operation to prevent collapsing. This is normally done by 
inserting wire mesh or fabricated wire cages into the enve- 
lopes. The opening of the envelope from which the cleaned 
air exits is attached to a tube sheet and, depending on design, 
the other end may be attached to a support member or canti- 
levered without support. The shaker mechanism may be lo- 
cated in either the dirty air or cleaned air compartments. 

Periodically (usually at 3- to 6-hour intervals) the air flow 
must be stopped to recondition the fabric. Figure 4-8 illus- 
trates the system air flow characteristics of an intermittent- 
duty fabric collector. As dust accumulates on the fabric, 
resistance to flow increases and air flow decreases until the 
fan is turned off and the fabric reconditioned. Variations in 
air flow due to changing pressure losses is sometimes a 
disadvantage and, when coupled with the requirement to 
periodically stop the air flow, may preclude the use of inter- 
mittent collectors. Reconditioning seldom requires more than 
two minutes but must be done without air flow through the 
fabric. If reconditioning is attempted with air flowing, it will 
be less effective and the flexing of the woven fabric will allow 
a substantial amount of dust to escape to the clean air side. 

The filtration velocity for large intermittent- duty fabric 
collectors seldom exceeds 6 fpm and normal selections are in 
the 2-4 fpm range. Lighter dust concentrations and the ability 
to recondition more often allow the use of higher filtration 
velocities. Ratings are usually selected so that the pressure 
drop across the fabric will be in the 2-5 "wg range between 
start and end of operating cycle. 

With multiple-section, continuous-duty, automatic fabric 
collectors, the disadvantage of stopping the air flow to permit 
fabric reconditioning and the variations in air flow with dust 
cake build-up can be overcome. The use of sections or com- 



TABLE 4-1 . Characteristics of Filter Fabrics* 





Example 
Trade Name 


Max. Temp. F 




Resistance to Physical Action 






Resistance to Chemicals 






Generic 




























Names 


Fabrics** 


Continuous 


Intermittent 


Dry Heat 


Moist Heat 


Abrasion 


Shaking 


Flexing 


Mineral Acid 


Organic Acid 


Alkalies 


Oxidizing 


Solvents 




Cotton 


Cotton 


180 


— 


G 


G 


F 


G 


G 


P 


G 


F 


F 


E 


a" 


Polyester 


Dacron (1) 
Fortret 2 ) 
Vycron {3) 
Kodel< 4 > 
Enka 


























C 

< 




Polyester* 3 * 


275 


— 


G 


F 


G 


E 


E 


G 


G 


F 


G 


E 


6i 

5' 


Acrylic 


Orion* 1 ) 
Acriian (6) 
Creslan^ 
Dralon T^ 


























S3 




Zefran 


275 


285 


G 


G 


G 


G 


E 


G 


G 


F 


G 


E 




Modacrylic 


Dynel* 10 ' 






























VerelW 


160 


— 


F 


F 


F 


P-F 


G 


G 


G 


G 


G 


G 




Nylon 


Nylon 




























(Polyamide) 


6,6( U6 > 






























Nylon 6^ 11 - 12 > 


225 


— 


G 


G 


E 


E 


E 


P 


F 


G 


F 


E 






Nomex (11) 


400 


450 


E 


E 


E 


E 


E 


P-F 


E 


G 


G 


E 




Polymide 


P_84( 18 ) 


500 


580 


E 


P 


G 


G 


E 


P-F 


G 


F 


G 


E 




Polypropylene 


Hercuion (13) 
Reevon (14 > 






























Vectra< 15 > 


200 


250 


G 


F 


E 


E 


G 


E 


E 


E 


G 


G 




Teflon 


Teflon 




























(flurocarbon) 


TFE^ 
Teflon 


500 


550 


E 


E 


P-F 


G 


G 


E 


E 


E 


E 


E 






FEPd) 


450 


— 


E 


E 


P-F 


G 


G 


E 


E 


E 


E 


E 




Expanded 


Rastex 


500 


550 


E 


E 


P-F 


G 


G 


E 


E 


E 


E 


E 




PFTE 






























Vinyon 


Vinyon< 16) 






























Clevyltt 17 > 


350 


— 


F 


F 


F 


G 


G 


E 


E 


G 


G 


P 




Glass 


Glass 


500 


600 


E 


E 


P 


P 


F 


E 


E 


F 


E 


E 




Fiberglass 


Fiberglass 091 


550 


550 


E 


E 


P 


P 


G 


G 


G 


G 


E 


G 





*E - excellent; G = good; F = fair; P = poor 

** Registered Trademarks 

(1) Du Pont; (2) Celanese; (3) Beaunit; (4) Eastman; (5) American Enka; (6) Chemstrand; (7) American Cyanamid; (8) Farbenfabriken Bayer AG; (9) Dow Chemical; (10) Union Carbide; (11) Allied Chemical; (12) Firestone; (13) Hercules; 
(14) Alamo Polymer; (15) National Plastic; (16) FMC; (17) Societe Rhovyl; (18) Lenzing; (19) Huyglas 



Air Cleaning Devices 



4-13 



partments, as indicated in Figure 4-7, allows continuous 
operation of the exhaust system because automatic dampers 
periodically remove one section from service for fabric recon- 
ditioning while the remaining compartments handle the total 
gas flow. The larger the number of compartments, the more 
constant the pressure loss and air flow. Either tubes or enve- 
lopes may be used and fabric reconditioning is usually accom- 
plished by shaking or vibrating. 

Figure 4-8 shows air flow versus time for a multiple-section 
collector. Each individual section or compartment has an air 
flow versus time characteristic like that of the intermittent 
collector, but the total variation is reduced because of the 
multiple compartments. Note the more constant air flow 
characteristic of the five-compartment unit as opposed to the 
three-compartment design. Since an individual section is out 
of service only a few minutes for reconditioning and remain- 
ing sections handle the total gas flow during that time, it is 
possible to clean the fabric more frequently than with the 
intermittent type. This permits the multiple-section unit to 
handle higher dust concentrations. Compartments are recon- 
ditioned in fixed sequence with the ability to adjust the time 
interval between cleaning of individual compartments. 

One variation of this design is the low-pressure, reverse-air 
collector which does not use shaking for fabric recondition- 
ing. Instead, a compartment is isolated for cleaning and the 
tubes collapsed by means of a low pressure secondary blower, 
which draws air from the compartment in a direction opposite 
to the primary air flow. This is a "gentle" method of fabric 
reconditioning and was developed primarily for the fragile 
glass cloth used for high-temperature operation. The reversal 
of air flow and tube deflation is accomplished very gently to 
avoid damage to the glass fibers. The control sequence usually 
allows the deflation and re-inflation of tubes several times for 
complete removal of excess dust. Tubes are 6-11 inches in 
diameter and can be as long as 30 feet. For long tubes, stainless 
steel rings may be sewn on the inside to help break up the dust 



cake during deflation. A combination of shaking and reverse 
air flow has also been utilized. 

When shaking is used for fabric reconditioning, the filtra- 
tion velocity usually is in the 1-4 fpm range. Reverse-air 
collapse-type reconditioning generally necessitates lower fil- 
tration velocities since reconditioning is not as complete. 
They are seldom rated higher than 3 fpm. The air to cloth ratio 
or filtration velocity is based on net cloth area available when 
a compartment is out of service for reconditioning. 

Reverse-jet, continuous-duty fabric collectors may use en- 
velopes or tubes of non-woven (felted) fabric, pleated car- 
tridges of non-woven mat (paper-like) in cylindrical or panel 
configuration, or rigid elements such as sintered polyethyl- 
ene. They differ from the low-pressure reverse-air type in that 
they employ a brief burst of high-pressure air to recondition 
the fabric. Woven fabric is not used because it allows exces- 
sive dust penetration during reconditioning. The most com- 
mon designs use compressed air at 80-1 00 psig, while others 
use an integral pressure blower at a lower pressure but higher 
secondary flow rate. Those using compressed air are generally 
called pulse-jet collectors and those using pressure blowers 
are called fan-pulse collectors. 

All designs collect dust on the outside and have air flow 
from outside to inside the fabric. All recondition the media 
by introducing the pulse of cleaning air into the opening where 
cleaned air exits from the tube, envelope, or cartridge. In many 
cases, a venturi shaped fitting is used at this opening to 
provide additional cleaning by inducing additional air flow. 
The venturi also directs or focuses the cleaning pulse for 
maximum efficiency. 

Figure 4-9 shows a typical pulse-jet collector. Under nor- 
mal operation (air flow from outside to inside), the fabric 
shape will tend to collapse; therefore, a support cage is 
required. The injection of a short pulse of high-pressure air 
induces a secondary flow from the clean air compartment in 



Table 4-2. Summary of Fabric-Type Collectors and Their Characteristics 





INTERRUPTABLE OPERATION 
Light to Moderate Loading 


INTERRUPTABLE OPERATION 
Heavy Loading 


CONTINUOUS OPERATIONS 
Any Loading 


Fabric Reconditioning 
Requirement 


Intermittent 


Continuous 


Type of Reconditioning 


Shaker 


Shaker 


Reverse Air 
(Low Pressure) 


Reverse Pulse - (High Pressure) 
Pulse Jet of Fan Pulse 


Collector Configuration 


Single Compartment 


Multiple Compartment 
with inlet or outlet dampers for each 


Single Compartment 


Fabric Configuration 


Tube, Cartridge or Envelope 


Tube or Envelope 


Tube 


Tube or Envelope 


Pleated Cartridge 


Type of Fabric 


Woven 


Woven 


Non-Woven (Felt) 


Non-Woven 


Air Flow 


Highly Variable 


Slightly Variable 


Virtually Constant 


Virtually Constant 


Normal Rating 
(filtration velocity, fpm) 


1 to 6 fpm 


1 to 3 fpm 


1 to 3 fpm 


5 to 12 fpm 


<1 to 7 fpm 



4-14 Industrial Ventilation 



— Motor driven vibrator 



n 



j 



/ 



vvvvv 



Clean 

air 
outlet 



Dusty 

air 
inlet 



TUBE TYPE 



Baffle 



6 

q 

ft 



Clean , ft 
air > 

ou ' let Iftfi 



Dusty 

air 

inlet 



Motor driven 
vibrator 




Dust outlet 



ENVELOPE TYPE 



ft _Com&J_2\ 



{ ~\\ m\ 



n — j: 



Screen rapping 
mechanism ' 



ft" Com p. 2*^ 



?i rF^e 



7m h ) 



1/ 



inlet pipes 



Comp.3 " \ 



( v( m{ * 



-J^ 



Clean air side 
Three position outlet valves 



-QjPJU&AJN 






Reverse air flow 



Compartments 1,2, and 3 
under air-- load. Compartment 
4 closed off for fabric 
cleaning. 



^L 



MULTIPLE SECTION CONTINUOUS AUTOMATIC 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



FABRIC COLLECTORS 



DATE 



1-88 | FIGURE 4-7 



Air Cleaning Devices 



4-15 



— One cycle *-| 

TIME 

INTERMITTENT DUTY FABRIC COLLECTOR 







c 




C 




c 




o o o 




-f- -+- -+- 




\_ "O 


r\ ~° 


"\ ~° 




\. c 


^\ c 


\. c 




^\ o 


\. o 


\. o 




\^ u 


\. o 


\. O 




] Q 


1 CD 


I <u 






q: 




OH 




QC 


F 




V 








k 






o 




5 




o 




_i 




^ . 


u 













£/ 



— One cycle — 
TIME 

3 Compartment 



-J 



o 


^"\. o 


"~"-\ o 


^--~. O 




o 


^"\- ° 


^"^\ u 


^"^\ o 


^^-~~^ 


a) 


^--, CD 


^^~t ^ 


^--, CD 


~--, 


VC 




ce 




cc 




(X 





TJ 

C 



— One cycle 

TIME 

5 Compartment 



Note: 



MULTIPLE SECTION, CONTINUOUS DUTY FABRIC COLLECTOR 



The flow variation has been exaggerated. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DATE 



AIR FLOW THROUGH 
FABRIC COLLECTORS 



1-88 



FIGURE 



4-8 



4-16 Industrial Ventilation 



Dirty air inlet 



Reverse air 
jet nozzles 




Clean air outlet 



Clean air outlet 



Fiber envelope 



Rotary valve 



Dust outlet 



Collection Pail 



Dirty air inlet 




Reverse jel piping 



Solenoid valves & controls 



Fabric element 



Dusi hopper 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



FABRIC COLLECTORS 
PULSE JET TYPE 



DATE 1-88 1 



FIGURE 



4-9 



Air Cleaning Devices 4-17 



a direction opposite to the normal air flow. Reconditioning is 
accomplished by the pulse of high-pressure air which stops 
forward air flow, then rapidly pressurizes the media, breaking 
up the dust cake and freeing accumulated dust from the fabric. 
The secondary or induced air acts as a damper, preventing 
flow in the normal direction during reconditioning. The entire 
process, from injection of the high-pressure pulse and initia- 
tion of secondary flow until the secondary flow ends, takes 
place in approximately one second. Solenoid valves which 
control the pulses of compressed air may be open for a tenth 
of a second or less. An adequate flow rate of clean and dry 
compressed air of sufficient pressure must be supplied to 
ensure effective reconditioning. 

Reverse-jet collectors normally clean no more than 10% of 
the fabric at any one time. Because such a small percentage 
is cleaned at any one time and because the induced secondary 
flow blocks normal flow during that time, reconditioning can 
take place while the collector is in service and without the 
need for compartmentation and dampers. The cleaning inter- 
vals are adjustable and are considerably more frequent than 
the intervals for shaker or reverse-air collectors. An individual 
element may be pulsed and reconditioned as often as once a 
minute to every six minutes. 

Due to this very short reconditioning cycle, higher filtration 
velocities are possible with reverse-jet collectors. However, 
with all reverse-jet collectors, accumulated dust that is freed 
from one fabric surface may become reintrained and redepo- 
sited on an adjacent surface, or even on the original surface. 
This phenomenon of redeposition tends to limit filtration 
velocity to something less than might be anticipated with 
cleaning intervals of just a few minutes. 

Laboratory tests (4t) have shown that for a given collector 
design redeposition increases with filtration velocity. Other 
test work (42) indicates clearly that redeposition varies with 
collector design and especially with flow patterns in the dirty 
air compartment. EPA-sponsored research^ 3) has shown that 
superior performance results from downward flow of the dirty 
air stream. This downward air flow reduces redeposition since 
it aids gravity in moving dust particles toward the hopper. 

Filtration velocities of 5-12 fpm are normal for reverse-jet 
collectors. The pleated cartridge type of reverse-jet collector 
is limited to filtration velocities in the 7 fpm range. The pleat 
configuration may produce very high approach velocities and 
greater redeposition. 

4.3.3 Wet Collectors: Wet collectors, or scrubbers, are 
commercially available in many different designs, with pres- 
sure drops from 1.5 M wg to as much as 100 "wg. There is a 
corresponding variation in collector performance. It is gener- 
ally accepted that, for well-designed equipment, efficiency 
depends on the energy utilized in air to water contact and is 
independent of operating principle. Efficiency is a function 
of total energy input per cfm whether the energy is supplied 



to the air or to the water. This means that we 11 -designed 
collectors by different manufacturers will provide similar 
efficiency if equivalent power is utilized. 

Wet collectors have the ability to handle high -temperature 
and moisture-laden gases. The collection of dust in a wetted 
form minimizes a secondary dust problem in disposal of 
collected material. Some dusts represent explosion or fire 
hazards when dry. Wet collection minimizes the hazard; 
however, the use of water may introduce corrosive conditions 
within the collector and freeze protection may be necessary 
if collectors are located outdoors in cold climates. Space 
requirements are nominal. Pressure losses and collection ef- 
ficiency vary widely for different designs. 

Wet collectors, especially the high-energy types, are fre- 
quently the solution to air pollution problems. It should be 
recognized that disposal of collected material in water without 
clarification or treatment may create water pollution prob- 
lems. 

Wet collectors have one characteristic not found in other 
collectors — the inherent ability to humidify. Humidification, 
the process of adding water vapor to the air stream through 
evaporation, may be either advantageous or disadvantageous 
depending on the situation. Where the initial air stream is at 
an elevated temperature and not saturated, the process of 
evaporation reduces the temperature and the volumetric flow 
rate of the gas stream leaving the collector. Assuming the fan 
is to be selected for operation on the clean air side of the 
collector, it may be smaller and will definitely require less 
power than if there had been no cooling through the collector. 
This is one of the obvious advantages of humidification; 
however, there are other applications where the addition of 
moisture to the gas stream is undesirable. For example, the 
exhaust of humid air to an air-conditioned space normally 
places an unacceptable load on the air conditioning system. 
High humidity can also result in corrosion of finished goods. 
Therefore, humidification effects should be considered before 
designs are finalized. While all wet collectors humidify, the 
amount of humidification varies for different designs. Most 
manufacturers publish the humidifying efficiency for their 
equipment and will assist in evaluating the results. 

Chamber or Spray Tower: Chamber or spray tower collec- 
tors consist of a round or rectangular chamber into which 
water is introduced by spray nozzles. There are many vari- 
ations of design, but the principal mechanism is impaction of 
dust particles on the liquid droplets created by the nozzles. 
These droplets are separated from the air stream by centrifugal 
force or impingement on water eliminators. 

The pressure drop is relatively low (on the order of 0.5-1 .5 
M wg), but water pressures range from 10-400 psig. The high 
pressure devices are the exception rather than the rule. In 
general, this type of collector utilizes low-pressure supply 
water and operates in the lower efficiency range for wet 



4-18 



industrial Ventilation 



collectors. Where water is supplied under high pressure, as 
with fog towers, collection efficiency can reach the upper 
range of wet collector performance. 

For conventional equipment, water requirements are rea- 
sonable, with a maximum of about 5 gpm per thousand scfm 
of gas. Fogging types using high water pressure may require 
as much as 10 gpm per thousand scfm of gas. 

Packed Towers: Packed towers (see Figure 4-10) are es- 
sentially contact beds through which gases and liquid pass 
concurrently, counter-currently, or in cross-flow. They are 
used primarily for appl ications involving gas, vapor, and mist 
removal. These collectors can capture solid particulate matter, 
but they are not used for that purpose because dust plugs the 
packing and requires unreasonable maintenance. 

Water rates of 5-1 gpm per thousand scfm are typical for 
packed towers. Water is distributed over V-notched ceramic 
or plastic weirs. High temperature deterioration is avoided by 
using brick linings, allowing gas temperatures as high as 1600 
F to be handled direct from furnace flues. 

The air flow pressure loss for a four foot bed of packing, 
such as ceramic saddles, will range from 1.5-3.5 "wg. The 
face velocity (velocity at which the gas enters the bed) will 
typically be 200-300 fpm. 

Wet Centrifugal Collectors: Wet centrifugal collectors (see 
Figure 4-11) comprise a large portion of the commercially 
available wet collector designs. This type utilizes centrifugal 
force to accelerate the dust particle and impinge it upon a 
wetted collector surface. Water rates are usually 2~~5 gpm per 
thousand scfm of gas cleaned. Water distribution can be from 
nozzles, gravity flow or induced water pickup. Pressure drop 
is in the 2-6 M wg range. 

As a group, these collectors are more efficient than the 
chamber type. Some are available with a variable number of 
impingement sections. A reduction in the number of sections 
results in lower efficiency, lower cost, less pressure drop, and 
smaller space. Other designs contain multiple collecting 
tubes. For a given air flow rate, a decrease in the tube size 
provides higher efficiency because the centrifugal force is 
greater. 

Wet Dynamic Precipitator: The wet dynamic precipitator 
(see Figure 4-1 2) is a combination fan and dust collector. Dust 
particles in the dirty air stream impinge upon rotating fan 
blades wetted with spray nozzles. The dust particles impinge 
into water droplets and are trapped along with the water by a 
metal cone while the cleaned air makes a turn of 180 degrees 
and escapes from the front of the specially shaped impeller 
blades. Dirty water from the water cone goes to the water and 
sludge outlet and the cleaned air goes to an outlet section 
containing a water elimination device. 

Orifice Type: In this group of wet collector designs (see 



Figure 4-12), the air flow through the collector is brought in 
contact with a sheet of water in a restricted passage. Water 
flow may be induced by the velocity of the air stream or 
maintained by pumps and weirs. Pressure losses vary from 1 
"wg or less for a water wash paint booth to a range of 3-6 "wg 
for most of the industrial designs. Pressure drops as high as 
20 "wg are used with some designs intended to collect very 
small particles. 

Venturi: The venturi collector (see Figure 4-11) uses a 
venturi-shaped constriction to establish throat velocities con- 
siderably higher than those used by the orifice type. Gas 
velocities through venturi throats may range from 
12,000-24,000 fpm. Water is supplied by piping or jets at or 
ahead of the throat at rates from 5-15 gpm per thousand scfm 
of gas. 

The collection mechanism of the venturi is impaction. As 
is true for all well-designed wet collectors, collection effi- 
ciency increases with higher pressure drops. Specific pressure 
drops are obtained by designing for selected velocities in the 
throat. Some venturi collectors are made with adjustable 
throats allowing operation over a range of pressure drops for 
a given flow rate or over a range of flow rates with a constant 
pressure drop. Systems are available with pressure drops as 
low as 5 "wg for moderate collection efficiency and as high 
as 100 "wg for collection of extremely fine particles. 

The venturi itself is a gas conditioner causing intimate 
contact between the particulates in the gas and the multiple 
jet streams of scrubbing water. The resulting mixture of gases, 
fume-dust agglomerates, and dirty water must be channeled 
through a separation section for the elimination of entrained 
droplets as shown in Figure 4-11. 

4.3.4 Dry Centrifugal Collectors: Dry centrifugal collec- 
tors separate entrained particulate from an air stream by the 
use or combination of centrifugal, inertial, and gravitational 
force. Collection efficiency is influenced by: 

1. Particle size, weight and shape. Performance is im- 
proved as size and weight become larger and as the 
shape becomes more spherical 

2. Collector size and design. The collection of fine dust 
with a mechanical device requires equipment designed 
to best utilize mechanical forces and fit specific appli- 
cation needs. 

3. Velocity. Pressure drop through a cyclone collector 
increases approximately as the square of the inlet 
velocity. There is, however, an optimum velocity that 
is a function of collector design, dust characteristics, 
gas temperature and density. 

4. Dust concentration. Generally, the performance of a 
mechanical collector increases as the concentration of 
dust becomes greater. 



Air Cleaning Devices 4-19 



Distributor 



Suitable packing 
media 



Air inlet 



Water in 




Steel cylindrical 
jacket 



Corrosion lining 
where required 



Support plate 



Water Outlet 



PACKED TOWER 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



WET TYPE COLLECTOR 
FOR GASEOUS CONTAMINANT) 



DATE 



1-88 | FIGURE 4-10 



4-20 



Industrial Ventilation 




WET CENTRIFUGAL 



Symbols 


Parts 


A 


Clean air outlet. 




B 


Entrainment separator. 




C 


Water inlet. 




D 


Impingement plates. 




E 


Dirty air inlet. 




F 


Wet cyclone for collecting 
material. 


heavy 


G 


Water and sludge drain. 






VENTURI SCRUBBER 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



WET TYPE DUST COLLECTORS 
(FOR PARTICULATE 

CONTAMINANTS) 



DATE 



1-88 



FIGURE 



4-11 



Air Cleaning Devices 4-21 



Entrapment 




Water spray nozzle. 



Water and 
sludge outlet. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



WET TYPE DUST COLLECTOR 

(FOR PARTICULATE 

CONTAMINANTS) 



DATE 1-88 



FIGURE 



4-12 



4-22 



industrial Ventilation 



Gravity Separators: Gravity separators consist of a cham- 
ber or housing in which the velocity of the gas stream is made 
to drop rapidly so that dust particles settle out by gravity. 
Extreme space requirements and the usual presence of eddy 
currents nullify this method for removal of anything but 
extremely coarse particles. 

Inertial Separators: Inertial separators depend on the in- 
ability of dust to make a sharp turn because its inertia is much 
higher than that of the carrier gas stream. Blades or louvers 
in a variety of shapes are used to require abrupt turns of 120° 
or more. Well -designed inertial separators can separate parti- 
cles in the 10-20 micron range with about 90% efficiency. 

Cyclone Collector: The cyclone collector (see Figure 4-13) 
is commonly used for the removal of coarse dust from an air 
stream, as a precleaner to more efficient dust collectors, and/or 
as a product separator in air conveying systems. Principal 
advantages are low cost, low maintenance, and relatively low 
pressure drops (in the 0.75-1.5 "wg range). It is not suitable 
for the collection of fine particles. 

High Efficiency Centrifugals: High-efficiency centrifugals 
(see Figure 4-13) exert higher centrifugal forces on the dust 
particles in a gas stream. Because centrifugal force is a 
function of peripheral velocity and angular acceleration, im- 
proved dust separation efficiency has been obtained by: 

1 . Increasing the inlet velocity. 

2. Making the cyclone body and cone longer. 

3. Using a number of small diameter cyclones in parallel. 

4. Placing units in series. 

While high-efficiency centrifugals are not as efficient on 
small particles as electrostatic, fabric, and wet collectors, their 
effective collection range is appreciably extended beyond that 
of other mechanical devices. Pressure losses of collectors in 
this group range from 3-8 "wg. 

4.4 ADDITIONAL AIDS IN DUST COLLECTOR SELECTION 

The collection efficiencies of the five basic groups of air 
cleaning devices have been plotted against mass mean particle 
size (Figure 4- 14). The graphs were found through laboratory 
and field testing and were not compiled mathematically. The 
number of lines for each group indicates the range that can be 
expected for the different collectors operating under the same 
principle. Variables, such as type of dust, velocity of air, water 
rate, etc., will also influence the range for a particular appli- 
cation. 

Deviation lines shown in the upper right hand corner of the 
chart allow the estimation of mass mean material size in the 
effluent of a collector when the inlet mean size is known. 
Space does not perm it a detai led explanation of how the slopes 
of these lines were determined, but the following example 
illustrates how they are used. The deviation lines should not 



be used for electrostatic precipitators but can be used for the 
other groups shown at the bottom of the figure. 

Example: A suitable collector will be selected for a lime 
kiln to illustrate the use of the chart. Referring to Figure 4- 14, 
the concentration and mean particle size of the material leav- 
ing the kiln can vary between 3 and 10 grains per cubic foot, 
with 5-10 microns the range for mass mean particle size. 
Assume an inlet concentration of 7.5 grains per cubic foot and 
a mean inlet size of 9 microns. Projection of this point 
vertically downwardly to the collection efficiency portion of 
the chart will indicate that a low-resistance cyclone will be 
less than 50% efficient; a high-efficiency centrifugal will be 
60-80% efficient and a wet collector, fabric arrester and 
electrostatic precipitator will be 97% efficient or more. A 
precleaner is usually feasible for dust concentrations over 5 
grains per cubic foot unless it is undesirable to have the 
collected dust separated by size. For this example a high-ef- 
ficiency centrifugal will be selected as the precleaner. The 
average efficiency is 70% for this group, therefore the effluent 
from this collector will have a concentration of 7.5 (1.00 - 
0.70) = 2.25 grains per cubic foot. Draw a line through the 
initial point with a slope parallel to the deviation lines marked 
"industrial dust." Where deviation is not known, the average 
of this group of lines normally will be sufficiently accurate to 
predict the mean particle size in the collector effluent. A 
vertical line from the point of intersection between the 2.25 
grains per cubic foot horizontal and the deviation line to the 
base of the chart will indicate a mean effluent particle size of 
6.0 microns. 

A second high-efficiency centrifugal in series would be less 
than 50% efficient on this effluent. A wet collector, fabric 
arrester, or electrostatic would have an efficiency of 94% or 
better. Assume that agood wet collector will be 98% efficient. 
The effluent would then be 2.25 (1 .00 - 0.98) = 0.045 grains 
per cubic foot. Using the previous deviation line and its 
horizontal intersection of 0.045 grains per cubic foot yields a 
vertical line intersecting the mean particle size chart at 1.6 
microns, the mean particle size of the wet collector effluent. 

In Table 4-3, an effort has been made to report types of dust 
collectors used for a wide range of industrial processes. While 
many of the listings are purely arbitrary, they may serve as a 
guide in selecting the type of dust collector most frequently 
used. 

4.5 CONTROL OF MIST, GAS, AND VAPOR 
CONTAMINANTS 

Previous discussion has centered on the collection of dust 
and fume or particulate existing in the solid state. Only the 
packed tower was singled out as being used primarily to collect 
mist, gas, or vapor. The character of a mist aerosol is very similar, 
aerodynamically, to that of a dust or fume aerosol, and the mist 
can be removed from an air stream by applying the principles 
that are used to remove solid particulate. 



Air Cleaning Devices 4-23 



\l r 




LOW PRESSURE CYCLONE 





HIGH EFFICIENCY CENTRIFUGALS 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DRY TYPE 
CENTRIFUGAL COLLECTORS 



DATE 



3-97 



FIGURE 



4-13 



4-24 



Industrial Ventilation 











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RANGE OF PARTICLE SIZES, CONCENTRATION, k COLLECTOR PERFORMANCE 

COMPILED BY S. SYLVAN APRIL 1952 : COPYRIGHT 1952 AMERICAN AIR FILTER CO. INC 

ACKNOWLEDGEMENTS OF PARTIAL SOURCES OF DATA REPORTED : 

1 FRANK W.G. - AMERICAN AIR FILTER - SIZE AND CHARACTERISTICS OF AIR BORNE SOLIDS - 1931 

2 FIRST AND DRINKER - ARCHIVES OF INDUSTRIAL HYGIENE AND OCCUPATIONAL MEDICINE - APRIL 1932 
J TAFT INSTITUTE AND AAF LABORATORY TE-ST DATA- l9«l-"©3 

4RCVEPu5E COLLAPSE CLOTH CLEANING ADDED 106 4 




AMERICAN CONFERENCE 
OF GOVERNMENTAL 

TMnTTOTinT A T TTA^nTTTiATTOrpO 


RANGE OF PARTICLE SIZE 


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DATE i-/3/3 


FIGURE 4 — ^4 



Air Cleaning Devices 



4-25 



Standard wet collectors are used to collect many types of 
mists. Specially designed electrostatic precipitators are fre- 
quently employed to collect sulfuric acid or oil mist. Even 
fabric and centrifugal collectors, although not the types pre- 
viously mentioned, are widely used to collect oil mist gener- 
ated by high speed machining. 

4.6 GASEOUS CONTAMINANT COLLECTORS 

Equipment designed specifically to control gas or vapor 
contaminants can be classified as: 

1. Adsorbers 

2. Thermal oxidizers 

3. Direct com bus tors 

4. Catalytic oxidizers 

4.6.1 Absorbers: Absorbers remove soluble or chemi- 
cally reactive gases from an air stream by contact with a 
suitable liquid. While all designs utilize intimate contact 
between the gaseous contaminant and the absorbent, different 
brands vary widely in configuration and performance. Re- 
moval may be by absorption if the gas solubility and vapor 
pressure promote absorption or chemical reaction. Water is 
the most frequently used absorbent, but additives are fre- 
quently required. Occasionally other chemical solutions must 
be used. Packed towers (Figure 4-14) are typical absorbers. 

4.6.2 Adsorbers: Adsorbers remove contaminants by col- 
lection on a solid. No chemical reaction is involved as adsorp- 
tion is a physical process where molecules of a gas adhere to 
surfaces of the solid adsorbent. Activated carbon or molecular 
sieves are popular adsorbents. 

4.6.3 Thermal Oxidizers: Thermal oxidizers, or after- 
burners, may be used where the contaminant is combustible. 
The contaminated air stream is introduced to an open flame 
or heating device followed by a residence chamber where 
combustibles are oxidized producing carbon dioxide and 
water vapor. Most combustible contaminants can be oxidized 
at temperatures between 1000 and 1500 F. The residence 
chamber must provide sufficient dwell time and turbulence to 
allow complete oxidation. 

4.6.4 Direct Combustors: Direct combustors differ from 
thermal oxidizers by introducing the contaminated gases and 
auxiliary air directly into the burner as fuel. Auxiliary fuel, 
usually natural gas or oil, is generally required for ignition 
and may or may not be required to sustain burning. 

4.6.5 Catalytic Oxidizers: Catalytic oxidizers may be 
used where the contaminant is combustible. The contami- 
nated gas stream is preheated and then passed through a 
catalyst bed which promotes oxidation of the combustibles to 
carbon dioxide and water vapor. Metals of the platinum 
family are commonly used catalysts which will promote 



oxidation at temperatures between 700 and 900 F. 

To use either thermal or catalytic oxidation, the combusti- 
ble contaminant concentration must be below the lower ex- 
plosive limit. Equipment specifically designed for control of 
gaseous or vapor contaminants should be applied with caution 
when the air stream also contains solid particles. Solid par- 
ticulates can plug absorbers, adsorbers, and catalysts and, if 
noncombustible, will not be converted in thermal oxidizers 
and direct combustors. 

Air streams containing both solid particles and gaseous con- 
taminants may require appropriate control devices in series. 

4.7 UNIT COLLECTORS 

Unit collector is a term usually applied to small fabric 
collectors having capacities in the 200-2000 cfm range. They 
have integral air movers, feature smal 1 space requirements and 
simplicity of installation. In most applications cleaned air is 
recirculated, although discharge ducts may be used if the 
added resistance is within the capability of the air mover. One 
of the primary advantages of unit collectors is a reduction in 
the amount of duct required, as opposed to central systems, 
and the addition of discharge ducts to unit collectors negates 
that advantage. 

When cleaned air is to be recirculated, a number of precau- 
tions are required (see Chapter 7). 

Unit collectors are used extensively to fill the need for dust 
collection from isolated, portable, intermittently used or fre- 
quently relocated dust producing operations. Typically, a 
single collector serves a single dust source with the energy 
saving advantage that the collector must operate only when 
that particular dust producing machine is in operation. 

Figure 4- 1 5 shows a typical unit collector. Usually they are 
the intermittent-duty, shaker-type in envelope configuration. 
Woven fabric is nearly always used. Automatic fabric clean- 
ing is preferred. Manual methods without careful scheduling 
and supervision are unreliable . 

4.8 DUST COLLECTING EQUIPMENT COST 

The variations in equipment cost, especially on an installed 
basis, are difficult to estimate. Comparisons can be mislead- 
ing if these factors are not carefully evaluated. 

4.8.1 Price Versus Capacity: All dust col lector prices per 
cfm of gas will vary with the gas flow rate. The smaller the 
flow rate, the higher the cost per cfm. The break point, where 
price per cfm cleaned tends to level off, will vary with the 
design. See the typical curves shown on Figure 4- 16. 

4.8.2 Accessories Included: Careful analysis of compo- 
nents of equipment included is very important. Some collector 
designs include exhaust fan, motor, drive, and starter. In other 
designs, these items and their supporting structure must be 
obtained by the purchaser from other sources. Likewise, while 



4-26 



Industrial Ventilation 



Table 4-3. Dust Collector Selection Guide 









( 


Collector Types Used in Industry 








Concen- 


Particle 


Dry Cen- 






Low-Volt 


Hi-Volt 




Operation 


tration 


Sizes 


trifugal 


Wet 


Fabric 


Electro- 


Electro- 


See 




Notel 


Notew 


Collector 


Collector 


Collector 


static 


static 


Remark No. 


CERAMICS 


















a. Raw product handling 


light 


fine 


S 








N 


N 


1 


b. Fettling 


light 


fine- 
medium 


S 


S 





N 


N 


2 


c. Refractory sizing 


heavy 


coarse 


N 


s 





N 


N 


3 


d. Glaze & vitr, enamel spray 


moderate 


medium 


N 








N 


N 




CHEMICALS 
















49 


a. Material handling 


light- 
moderate 


fine- 
medium 


S 








N 


N 


4 


b. Crushing, grinding 


moderate- 
heavy 


fine- 
coarse 





s 





N 


N 


5 


c. Pneumatic conveying 


very 
heavy 


fine- 
coarse 





s 





N 


N 


6 


d. Roasters, kilns, coolers 


heavy 


mid- 
coarse 











N 


N 


7 


COAL, MINING AND POWER PLANT 














49 


a. Material handling 


moderate 


medium 





s 





N 


N 


8 


b. Bunker ventilation 


light 


fine 


S 


s 





N 


N 


9 


c. Dedusting, air cleaning 


heavy 


medium- 
coarse 


S 








N 


N 


10 


d. Drying 


moderate 


fine 


N 








N 


N 


11 


FLY ASH 


















a. Coai burning-chain grate 


light 


fine 


S 


s 





N 





12 


b. Coal burning-stoker fired 


moderate 


fine- 
coarse 


S 


s 





N 







c. Coal burning-pulverized 
fuel 


moderate 


fine 


S 


s 





N 





13 


d. Wood burning 


varies 


coarse 


S 


s 





N 


s 


14 


FOUNDRY 


















a. Shakeout 


light- 
moderate 


fine 


N 








N 


N 


15 


b. Sand handling 


moderate 


fine- 
medium 


N 








N 


N 


16 


c. Tumbling mills 


heavy 


medium- 
coarse 


N 


s 





N 


N 


17 


d. Abrasive cleaning 


moderate- 
heavy 


fine- 
medium 


N 


s 





N 


N 


18 


GRAIN ELEVATOR, FLOUR AND FEED MILLS 












49 


a. Grain handling 


light 


medium 





s 





N 


N 


19 


b. Grain dryers 


light 


coarse 


S 


s 





N 


N 


20 


c. Flour dust 


moderate 


medium 





s 





N 


N 


21 


d. Feed mill 


moderate 


medium 





s 





N 


N 


22 


METAL MELTING 
















50 


a. Steel blast furnace 


heavy 


varied 


N 





s 


N 


S 


23 


b, Steel open hearth 


moderate 


fine- 
coarse 


N 





s 


N 


S 


24 


c. Steel electric furnace 


light 


fine 


N 


s 





N 


S 


25 


d. Ferrous cupola 


moderate 


varied 


N 








N 




26 


e. Non-ferrous reverberatory 


varied 


fine 


N 


s 





N 


N 


27 


f. Non-ferrous crucible 


light 


fine 


N 


s 





N 


N 


28 


METAL MINING AND ROCK PRODUCTS 
















a. Material handling 


moderate 


fine- 
medium 


N 








N 


N 


29 


b. Dryers, kilns 


moderate 


medium- 
coarse 











N 





30 


c. Rock dryer 


moderate 


fine- 
medium 


N 


s 


s 


N 


S 


31 


d. Cement kiln 


heavy 


fine- 
medium 


N 


N 





N 


S 


32 



Air Cleaning Devices 



4-27 









Collector Types Used in Industry 








Concen- 


Particle 


Dry Cen- 






Low-Volt 


Hi-Volt 




Operation 


tration 


Sizes 


trifugal 


Wet 


Fabric 


Electro- 


Electro- 


See 




Notel 


Notew 


Collector 


Collector 


Collector 


static 


static 


Remark No. 


e. Cement grinding 


moderate 


fine 


N 


N 





N 


N 


33 


f. Cement clinker cooler 


moderate 


coarse 





N 





N 


N 


34 


METAL WORKING 
















49 


a. Production grinding, 
scratch brushing, abrasive 


light 


coarse 











N 


N 


35 


















cutoff 


















b. Portable and swing frame 


light 


medium 


S 








N 


N 




c. Buffing 


light 


varied 


S 








N 


N 


36 


d. Tool room 


light 


fine 


S 


S 


S 


N 


N 


37 


e. Cast iron machining 


moderate 


varied 











S 


N 


38 


PHARMACEUTICAL AND FOOD PRODUCTS 
















a. Mixers, grinders, weighing, 


light 


medium 











N 


N 


39 


blending, bagging, 


















packaging 


















b. Coating pans 


varied 


fine- 
medium 


N 








N 


N 


40 


PLASTICS 
















49 


a. Raw material processing 


(See comments under 





S 





N 


N 


41 




Chemicals) 
















b, Plastic finishing 


light- 
moderate 


varied 


S 


S 





N 


N 


42 


c. Extrusion 


light 


fine 


N 


S 


N 





N 




RUBBER PRODUCTS 
















49 


a. Mixers 


moderate 


fine 


S 





s 


N 


N 


43 


b. Batchout roils 


light 


fine 


S 





s 


S 


N 




c. Talc dusting and dedusting 


moderate 


medium 


S 


S 





N 


N 


44 


d. Grinding 


moderate 


coarse 











N 


N 


45 


WOODWORKING 
















49 


a. Woodworking machines 


moderate 


varied 





S 





N 


N 


46 


b. Sanding 


moderate 


fine 


s 


S 





N 


N 


47 


c. Waste convevinq, hoqs 


heavy 


varied 





S 


s 


N 


N 


48 


Note 1 : Light: less than 2 gr/ft 3 ; Moderage: 2 to 5 gr/ft 3 ; Heavy: 


5 gr/ft 3 and up. 












Note 2: Fine: 50% less than 5 microns; Medium: 50% 5 to 15 microns; Coarse: 50% 15 microns and larger 








Note 3: = often; S = seldom; N = 


= never. 

















Remarks Referred to in Table 4-3 



1 . Dust released from bin filling, conveying, weighing, mixing, pressing 11 . 
forming. Refractory products, dry pan and screen operations more 

severe. 12. 

2. Operations found in vitreous enameling, wall and floor tile, pottery. 

3. Grinding wheel or abrasive cut-off operation. Dust abrasive. 1 3. 

4. Operations include conveying, elevating, mixing, screening, weighing, 
packaging. Category covers so many different materials that recom- 14. 
mendation will vary widely. 15. 

5. Cyclone and high efficiency centrifugals often act as primary collectors 1 6, 
followed by fabric or wet type. 1 7 

6. Cyclones used as product collector followed by fabric arrester for high ig. 
over-all collection efficiency, 

7. Dust concentration determines need for dry centrifugal; plant location, 
product value determines need for final collectors. High temperatures 

are usual and corrosive gases not unusual. 19 

8. Conveying, screening, crushing, unloading. 20. 

9. Remove from other dust producing points. Separate collector usually. 

10. Heavy loading suggests final high efficiency collector for all except 21. 

very remote locations. 



Difficult problem but collectors will be used more frequently with air 
pollution emphasis. 

Public nuisance from boiler blow-down indicates collectors are 
needed. 

Large installations in residential areas require electrostatic in addition 
to dry centrifugal. 

Cyclones used as spark arresters in front of fabric collectors. 
Hot gases and steam usually involved. 
Steam from hot sand, adhesive clay bond involved. 
Concentration very heavy at start of cycle. 
Heaviest load from airless blasting due to higher cleaning speed. 
Abrasive shattering greater with sand than with grit or shot. Amounts 
removed greater with sand castings, less with forging scale removal, 
least when welding scale is removed. 
Operations such as car unloading, conveying, weighing, storing. 
Collection equipment expensive but public nuisance complaints be- 
coming more frequent. 

Operations include conveyors, cleaning rolls, sifters, purifiers, bins 
and packaging. 



4-28 



Industrial Ventilation 



22. Operations include conveyors, bins, hammer mills, mixers, feeders 
and baggers. 

23. Primary dry trap and wet scrubbing usual. Electrostatic is added where 
maximum cleaning required. 

24. Use of this technique declining. 

25. Air pollution standards will probably require increased usage of fabric 
arresters. 

26. CAUTION! Recent design improvements such as coke-less, plasma- 
fired type, have altered emission characteristics. 

27. Zinc oxide loading heavy during zinc additions, Stack temperatures 
high. 

28. Zinc oxide plume can be troublesome in certain plant locations. 

29. Crushing, screening, conveying involved. Wet ores often introduce 
water vapor in exhaust air. 

30. Dry centrifugals used as primary collectors, followed by final cleaner. 

31 . Industry is aggressively seeking commercial uses for fines. 

32. Collectors usually permit salvage of material and also reduce nuisance 
from settled dust in plant area. 

33. Salvage value of collected material high. Same equipment used on 
raw grinding before calcining. 

34. Coarse abrasive particles readily removed in primary collector types. 

35. Roof discoloration, deposition on autos can occur with cyclones and 
less frequently with high efficiency dry centrifugal. Heavydutyairfilters 
sometimes used as final cleaners. 

36. Linty particles and sticky buffing compounds can cause piuggage and 
fire hazard in dry collectors. 



Remarks Referred to in Table 4-3 (continued) 

37. Unit collectors extensively used, especially for isolated machine tools. 

38. Dust ranges from chips to fine floats including graphitic carbon. Low voltage 



ESP applicable only when a coolant is used. 

39. Materials vary widely. Collector selection depends on salvage value, 
toxicity, sanitation yardsticks. 

40. Controlled temperature and humidity of supply air to coating pans 
makes recirculation desirable. 

41 . Plastic manufacture allied to chemical industry and varies with opera- 
tions involved. 

42. Operations and collector selection similar to woodworking. See Item 
13. 

43. Concentration is heavy during feed operation. Carbon black and other 
fine additions make collection and dust-free disposal difficult. 

44. Salvage of collected material often dictates type of high efficiency 
collector. 

45. Fire hazard from some operations must be considered. 

46. Bulking material. Collected material storage and bridging from splin- 
ters and chips can be a problem. 

47. Dry centrifugals not effective on heavy concentration of fine particles 
from production sanding. 

48. Dry centrifugal collectors required. Wet or fabric collectors may be 
used for final collectors. 

49. See NFPA publications for fire hazards, e.g., zirconium, magnesium, 
aluminum, woodworking, plastics, etc. 



dust storage hoppers are integral parts of some dust collector 
designs, they are not provided in other types. Duct connec- 
tions between elements may be included or omitted. Recircu- 
lating water pumps and/or settling tanks may be required but 
not included in the equipment price. 

4.8.3 Installation Cost: The cost of installation can equal 
or exceed the cost of the collector. Actual cost will depend on 
the method of shipment (completely assembled, sub-assem- 
bled or completely knocked down), the location (which may 
require expensive rigging), and the need for expensive sup- 
porting steel and access platforms. Factory installed media 
will reduce installation cost. The cost can also be measurably 
influenced by the need for water and drain connections, 
special or extensive electrical work, and expensive material 
handling equipment for collection material disposal. Items in 
the latter group will often also be variable, decreasing in cost 
per cfm as the flow rate of gas to be cleaned increases. 

4.8.4 Special Construction: Prices shown in any tabula- 
tion must necessarily assume standard or basic construction. 
The increase in cost for corrosion resisting material, special 
high-temperature fabrics, insulation, and/or weather protec- 
tion for outdoor installations can introduce a multiplier of one 
to four times the standard cost, 

A general idea of relative dust collector cost is provided in 
Figure 4-16. The additional notes and explanations included 



in these data should be carefully examined before they are 
used for estimating the cost of specific installations. For more 
accurate data, the equipment manufacturer or installer should 
be asked to provide estimates or a past history record for 
similar control problems utilized. Table 4-4 lists other char- 
acteristics that must be evaluated along with equipment cost. 

Price estimates included in Figure 4-16 are for equipment 
of standard construction in normal arrangement. Estimates for 
exhausters and dust storage hoppers have been included, as 
indicated in Notes 1 and 2, where they are normally furnished 
by others. 

4.9 SELECTION OF AIR FILTRATION EQUIPMENT 

Air filtration equipment is available in a wide variety of 
designs and capability. Performance ranges from a simple 
throwaway filter for the home furnace to the "clean room" in 
the electronics industry, where the air must be a thousand 
times as clean as in a hospital surgical suite. Selection is based 
on efficiency, dust holding capacity, and pressure drop. There 
are five basic methods of air filtration. 

4.9.1 Straining: Straining occurs when a particle is larger 
than the opening between fibers and cannot pass through. It 
is a very ineffective method of filtration because the vast 
majority of particles are far smaller than the spaces between 
fibers. Straining will remove lint, hair, and other large parti- 



Air Cleaning Devices 4-29 



Shaker 



Motor 



Fan Impeller 




Envelopes 



Air inlet 



Funnel hopper 



Dust drum 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



UNIT COLLECTOR 
(FABRIC- SHAKER TYPE) 



date 1-88 



FIGURE 



4-15 



4-30 



Industrial Ventilation 



o 
en 



m 
O 
o 

LiJ 

> 



100 



75 



50 



25 

























\ 

A 


























c \ 




n . 










B \ 






































v \ \ 


F 
\\ 

\ 
























""\B 














A 










E 










^' % ^^ 
















■ c 




















G 












""*■"■*«,_ *J 


N,^*** 


G 


=^ 


"f 




E 




Ks 

















100 



000 



CFM IN THOUSANDS 



C. 
D. 
E. 
F. 
G. 



High voltage precipitator (minimum cost range) 

Continuous duty high temperature fabric collector (2.0:1) 

Continuous duty reverse pulse (8:1) 

Wet collector 

Intermittent duty fabric collector (2.0:1) 

Low voltage precipitator 

Cyclone 



Note 1: Cost based on collector section only. Does not include ducts, dust disposal 
devices, pumps, exhausters or other accessories not an integral part of the 
collector. 

Note 2: Price of high voltage precipitator will vary substantially with applications 
and efficiency requirements. Costs shown are for fly ash aplications 
where velocities of 200 to 300 fpm are normal. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



COST ESTIMATES OF 
DUST COLLECTING EQUIPMENT 



DATE 



1-88 



FIGURE 



4-16 



TABLE 4-4. Comparison of Some Important Dust Collector Characteristics 





Higher efficiency 














Max. Temp. F 




Range on Particles 
Greater than Mean 


Pressure 


H 2 Gal per 




Sensitivity to cfm Change 




Standard 
Construction 


Type 


Size in Microns 


Loss Inches 


1,000 cfm 


Space 


Pressure 


Efficiency 


Humid Air Influence 


Note 4 


Electrostatic 


0.25 


1 /2 


_ 


Large 


Negligible 


Yes 


Improves Efficiency 


500 


Fabric 














^ 




Intermittent— Shaker 


0.25 


3-6 " 




— 


Large 


As cfm 


Negligible 


May Make 




See Table 4-1 


Continuous— Shaker 


0.25 


3-6 


► Note 1 


— 


Large 


As cfm 


Negligible 


Reconditioning 


► 




Continuous— Reverse Air 


0.25 


3-6 


— 


Large 


As cfm 


Negligible 


Difficult 






Continuous— Reverse Pulse 


0.25 


3-6 ^ 




— 


Moderate 


As cfm 


Negligible 


j 






Glass, Reverse flow 
Wet: 
Packed Tower 


0.25 


3-8 


_ 


Large 


As cfm 


Negligible 




550 


1-5 


1.5-3.5 


5-10 


Large 


As cfm 


Yes -| 




-v 


Wet Centrifugal 


1-5 


2.5-6 


3-5 


Moderate 


As (cfm) 2 


Yes 








Wet Dynamic 


1-2 


Note 2 


V 2 to 1 


Small 


Note 2 


No 


> None 




► Unlimited 


Orifice Types 


1-5 


2V2-6 


10-40 


Small 


As cfm or less 


Varies with 
design j 








Higher Efficiency: 


















Fog Tower 


0.5-5 


2-4 


5-10 


Moderate 


As (cfm) 2 


Slightly 


None 


Note 3 


Venturi 


0.5-2 


10-100 


5-15 


Moderate 


As (cfm) 2 


Yes 




Unlimited 


Dry Centrifugal: 














^ 




Low Pressure Cyclone 


20-40 


0.75-1.5 


— 


Large 


As (cfm) 2 


Yes 


May Cause 


750 


High Eft. Centrifugal 


10-30 


3-6 


— 


Moderate 


As (cfm) 2 


Yes 


^condensation 


750 


Dry Dynamic 


10-20 


Note 2 


_ 


Small 


Note 2 


No 


J and plugging 





Note 1: Pressure loss is that for fabric and dust cake. Pressure tosses associated with outlet connections to be added by system designer. 

Note 2: A function of the mechanical efficiency of these combined exhausters and dust collectors. 

Note 3: Precooling of high temperature gases will be necessary to prevent rapid evaporation of fine droplets. 

Note 4: See NFPA requirements for fire hazards, e.g., zirconium, magnesium, aluminum, woodworking, etc. 



as 
5° 

o 

fD 

< 



4-32 



Industrial Ventilation 



cles. 



4.9.2 Impingement: When air flows through a filter, it 
changes direction as it passes around each fiber. Larger dust 
particles, however, cannot follow the abrupt changes in direc- 
tion because of their inertia. As a result, they do not follow 
the air stream and collide with a fiber. Filters using this 
method are often coated with an adhesive to help fibers retain 
the dust particles that impinge on them. 

4.9.3 Interception: Interception is a special case of im- 
pingement where a particle is small enough to move with the 
air stream but, because its size is very small in relation to the 
fiber, makes contact with a fiber while following the tortuous 
air flow path of the filter. The contact is not dependent on 
inertia and the particle is retained on the fiber because of the 
inherent adhesive forces that exist between the particle and 
fiber. These forces, called van der Waals (J. D. van der Waals, 
1837-1923) forces, enable a fiber to trap a particle without 
the use of inertia. 

4.9.4 Diffusion: Diffusion takes place on particles so 
small that their direction and velocity are influenced by 
molecular collisions. These particles do not follow the air 
stream, but behave more like gases than particulate. They 
move across the direction of air flow in a random fashion. 
When a particle does strike a fiber, it is retained by the van 
der Waals forces existing between the particle and the fiber. 
Diffusion is the primary mechanism used by most extremely 
efficient filters. 

4.9.5 Electrostatic: A charged dust particle will be at- 
tracted to a surface of opposite electrical polarity. Most dust 
particles are not electrically neutral; therefore, electrostatic 
attraction between dust particle and filter fiber aids the col- 
lection efficiency of all barrier- type air filters. Electrostatic 
filters establish an ionization field to charge dust particles so 
that they can be collected on a surface that is grounded or of 
opposite polarity. This concept was previously discussed in 
Section 4.3.1. 

Table 4-5 shows performance versus filter fiber size for 
several filters. Note that efficiency increases as fiber diameter 
decreases because more small fibers are used per unit volume. 



Note also that low velocities are used for high-efficiency 
filtration by diffusion. 

The wide range in performance of air filters makes it 
necessary to use more than one method of efficiency testing. 
The industry-accepted methods in the United States are 
ASHRAE Arrestance, ASHRAE Efficiency, and DOP. For 
ASHRAE Arrestance, a measured quantity of 72% stand- 
ardized air cleaner test dust, 23% carbon black, and 5% cotton 
lint is fed to the filter. The efficiency by weight on this specific 
test dust is the ASHRAE Arrestance. ASHRAE Efficiency is 
a measure of the ability of a filter to prevent staining or 
discoloration. It is determined by light reflectance readings 
taken before and after the filter in a specified test apparatus. 
Atmospheric dust is used for the test. Both ASHRAE tests are 
described in ASHRAE Publication 52-76.^ 

In a DOP Test, 0.3 micron particles of dioctylphthalate 
(DOP) are drawn through a HEPA (High Efficiency Particu- 
late Air) filter. Efficiency is determined by comparing the 
downstream and upstream particle counts. To be designated 
as a HEPA filter, the filter must be at least 99.97% efficient, 
i.e., only three particles of 0.3 micron size can pass for every 
ten thousand particles fed to the filter. Unlike both ASHRAE 
tests, the DOP test is not destructive, so it is possible to repair 
leaks and retest a filter that has failed. 

The three tests are not directly comparable; however, Fig- 
ure 4-1 7 shows the general relationship. Table 4-6 compares 
several important characteristics of commonly used air filters. 
Considerable life extension of an expensive final filter can be 
obtained by the use of one or more cheaper, less efficient, 
prefilters. For example, the life of a HEPA filter can be 
increased 25% with athrowaway prefilter. If the throwaway 
filter is followed by a 90% efficient extended surface filter, 
the life of the HEPA filter can be extended nearly 900%o. This 
concept of "progressive filtration" allows the final filters in 
clean rooms to remain in place for ten years or more. 

The European Committee on the Construction of Air Han- 
dling Equipment has developed a method for testing air filters 
in general ventilation. Although their method, called 
Eurovent 4/5, is based directly on ASHRAE Standard 52-76, 
some wording and definitions have been amended to suit the 
needs of Eurovent. Eurovent 4/5 aims to establish a uniform 



1 

ARRESTANCE 
52-76 

EFFICIENCY 
52-76 

DOP 


60 70 80 90 95 


60 


























i i 


70 80 90 c 










5 


.0 30 40 5 


5 






1 1 




i 






J 


\ 




10 




20 30 40 50 6 


70 80 90 95 99.99 




i 




i 




zn 





























FIGURE 4-17. Comparison between various methods of measuring air cleaning capability. 



Air Cleaning Devices 



4-33 



comparative testing procedure for air filters having volumet- 
ric flow rates greater than 0.236 mVs (500 cfm) and an average 
dust spot efficiency up to 98%. 

The wide range of filter efficiency is segregated into 14 
grades of filters from EU 1 to EU 14. 

4.10 RADIOACTIVE AND HIGH TOXICITY OPERATIONS 

There are three major requirements for air cleaning equipment 
to be utilized for radioactive or high toxicity applications: 

1. High efficiency 

2. Low maintenance 

3. Safe disposal 

High efficiency is essential because of extremely low tol- 
erances for the quantity and concentration of stack effluent 
and the high cost of the materials handled. Not only must the 
efficiency be high, it must also be verifiable because of the 
legal requirement to account for all radioactive material. 

The need for low maintenance is of special importance 
when exhausting any hazardous material. For many radioac- 
tive processes, the changing of bags in a conventional fabric 
collector may expend the daily radiation tolerances of 20 or 
more persons. Infrequent, simple, and rapid maintenance 
requirements are vital. Another important factor is the desir- 
ability of low residual buildup of material in the collector 
since dose rates increase with the amount of material and 
reduce the allowable working time. 

Disposal of radioactive or toxic materials is a serious and 
very difficult problem. For example, scalping filters loaded 
with radioactive dust are usually incinerated to reduce the 
quantity of material that must be disposed of in special burial 
grounds. The incinerator will require an air cleaning device, 
such as a wet collector of very special design, to avoid 
unacceptable pollution of air and water. 

With these factors involved, it is necessary to select an air 
cleaning device that will meet efficiency requirements with- 
out causing too much difficulty in handling and disposal. 

Filter units especially designed for high efficiency and low 
maintenance are available. These units feature quick 
changeout through a plastic barrier which is intended to 
encapsulate spent filters, thereby eliminating the exposure of 
personnel to radioactive or toxic material. A filtration effi- 
ciency of 99.97% by particle count on 0.3 micron particles is 
standard for this type of unit. 

For further information on this subject, see Reference 4.5. 

4.11 EXPLOSION VENTING 

There is a wide range of dusts which are combustible and 
capable of producing an explosion. Explosions occur when 
the right concentration of finely divided dust is suspended in 
air and exposed to a sufficient source of ignition. A dust 
collector, by its very operation, maintains a cloud of finely 



TABLE 4-5. Media Velocity vs. Fiber Size 







Media 






Filter Size 


Velocity 


Filtration 


Filter Type 


(microns) 


(fpm) 


Mechanism 


Panel Filters 


25-50 


250-625 


Impingement 


Automatic Roil Filters 


25-50 


500 


Impingement 


Extended Surface Filters 


0.75-2.5 


20-25 


Interception 


HEPA Filters 


0.5-6.3 


5 


Diffusion 



divided particles suspended in air. If a source of ignition 
initiates the combustion of the dust cloud, the gases in the 
cloud will rapidly expand due to heat developed during the 
combustion. If a dust collector vessel constructs this expan- 
sion, a rapid pressure buildup inside the collector casing will 
cause a violent rupture. When dust particles are know to be 
combustible, precautions for an explosion must be taken and 
suitable protection provided to reduce the risk of property 
damage and personal injury. 

To begin taking precautions, sources of possible ignition 
must be identified and controlled to minimize the risk of a 
dust cloud explosion. Usual causes of explosions include 
static by minimizing the ignition sources such as static dis- 
charge, hot surfaces on machinery and sparks and flames from 
processes. After identifying possible sources of ignition, pre- 
ventive measures should be taken. Static grounding of the 
equipment and spark traps are typical preventive measures. 
The addition of an inert gas to replace oxygen in a dust 
collector can prevent an explosion by ensuring the minimum 
oxygen content required for ignition is never reached. Inerting 
can be very effective in closed loop systems but is not eco- 
nomical in typical local exhaust systems because of the con- 
stant loss of expensive inerting gas. Should ignition occur, 
protective measures must be taken to limit the damage. Typi- 
cal protective measures include explosion suppression, explo- 
sion containment, and explosion venting. 

Explosion suppression requires the early detection of an 
explosion, usually within the first 20 milliseconds. Once 
ignition is detected, an explosion suppression device injects 
a pressurized chemical suppressant into the collector to dis- 
place the oxygen and impede combustion. These are typically 
used in conjunction with fast acting isolation valves on the 
inlet and outlet ducts. These systems can be very useful when 
toxic dusts are being handled. 

Explosion containment uses specialized dust collectors 
designed to withstand the maximum pressure generated and 
contain the explosion. Most pressure capabilities of commer- 
cially available dust collectors are not sufficient to contain an 
explosion in progress. 

Explosion venting, the most common protection, is af- 
forded by fitting pressure relief vents to the collector housing. 



4-34 



Industrial Ventilation 



TABLE 4-6. Comparisor 


\ of Some Important Air Filter Characteristics* 












Pressure Drops "wg 
(Notes 1& 2) 


ASHRAE Performance 
(Note 4) 


Face Velocity 
fpm 


Maintenance 
(Note 5) 


Type 


Initial 


Final 


Arrestance 


Efficiency 


Labor 


Material 


Low/Medium Efficiency 
















1. Glass Throwaway 
(2" deep) 


0,1 


0.5 


77% 


NA 
Note 6 


300 


High 


High 


2. High Velocity 
(permanent units) 
(2" deep) 


0.1 


0.5 


73% 


NA 
Note 6 


500 


High 


Low 


3. Automatic 
(viscous) 


0.4 


0.4 


80% 


NA 
Note 6 


500 


Low 


Low 


Medium/High Efficiency 
















1. Extended Surface 
(dry) 


0.15-0.60 


0.5-1.25 


90-99% 


25-95% 


300-625 


Medium 


Medium 


2. Electrostatic: 
















a. Dry Agglomerator/ 
Roll Media 


0.35 


0.35 


NA 
Note 7 


90% 


500 


Medium 


Low 


b. Dry Agglomerator/ 
Extended Surface 
Media 


0.55 


1.25 


NA 
Note 7 


95%+ 


530 


Medium 


Medium 


c. Automatic Wash 
Type 


0.25 


0.25 


NA 
Note 7 


85-95% 


400-600 


Low 


Low 


Ultra High Efficiency 
















1.HEPA 


0.5-1.0 


1.0-3.0 


Note 3 


Note 3 


250-500 


High 


High 



Note 1: Pressure drop values shown constitute a range or average, whichever is applicable, 

Note 2: Final pressure drop indicates point at which filter or filter media is removed and the media is either cleaned or replaced. Ail others are 
cleaned in place, automatically, manually or media renewed automatically. Therefore, pressure drop remains approximately constant. 

Note 3: 95-99.97% by particle count, DOP test. 

Note 4: ASHRAE Standard 52-76 defines (a) Arrestance as a measure of the ability to remove injected synthetic dust, calculated as a percentage on a weight 
(b) Efficiency as a measure of the ability to remove atmospheric dust determined on a light-transmission (dust spot) basis. 

Note 5: Compared to other types within efficiency category. 

Note 6: Too low to be meaningful. 

Note 7: Too high to be meaningful. 



.is and 



As pressure increases quickly leading up to an explosion, a 
relief vent opens to allow the rapidly expanding gases to 
escape. This effectively limits the maximum pressure build- 
up to less than the bursting pressure of the vessel. The neces- 
sary area for such a relief vent is a function of the vessel 
volume, vessel strength, the opening pressure ofthe relief vent 
and the rate of pressure rise characteristic of the dust in 
question. Most standard dust collectors will require reinforc- 
ing to withstand the reduced maximum pressure experienced 
during an explosion. 

To choose the most reliable, economical and effective 
means of explosion control, an evaluation ofthe specifics of 
the exhaust system and the degree of protection required is 
necessary. 

NAPA 68-1994, Guide for Explosion Venting,(4.6) is the 
most commonly recognized standard and should be studied 



and thoroughly familiar to anyone responsible for the design 
or evaluation of dust collectors applied to potentially explo- 
sive dusts. 

REFERENCES 

4.1. Leith, D.; First, M.K.W.; Feldman, H.: Performance 
of a Pulse- Jet at High Velocity Filtration II, Filter Cake 
Redeposition. J. Air Pollut. Control Assoc. 28:696 
(July 1978). 

4.2. Beake, E.: Optimizing Filtration Parameters. J. Air 
Pollut. Control Assoc. 24:1150 (1974). 

4.3. Leith, D.; Gibson, D.D.; First, M.W.: Performance of 
Top and Bottom Inlet Pulse-Jet Fabric Filters. J. Air 
Pollut. Control Assoc. 24:1150 (1974). 

4.4. American Society of Heating, Refrigerating and Air- 
Conditioning Engineers: Method of Testing Cleaning 



Air Cleaning Devices 



4-35 



Devices Used in General Ventilation for Removing 
Particulate Matter. ASHRAE Pub. No. 52-76. 
ASHRAE, Atlanta, GA (May 1976). 

4.5. National Council on Radiation Protection and Meas- 
urement: NCRP Report No. 39, Basic Radiation Pro- 



tection Criteria. NCRP Report No. 39. NCRP Publi- 
cations, Bethesda, MD (January 1971). 

4.6. National Fire Protection Association: Guide for Ex- 
plosion Venting. NFPA 68-1978. NFPA, Quincy, MA 

(1978). 



Chapter 5 

EXHAUST SYSTEM DESIGN PROCEDURE 



5.1 INTRODUCTION 5-2 

5.2 PRELIMINARY STEPS 5-2 

5.3 DESIGN PROCEDURE 5-2 

5.4 DUCT SEGMENT CALCULATIONS 5-3 

5.5 DISTRIBUTION OF AIR FLOW 5-4 

5.5.1 Balance by Design Method 5-4 

5.5.2 Blast Gate Method 5-10 

5.5.3 Choice of Methods 5-10 

5.5.4 Balance by Design Procedure 5-10 

5.5.5 Blast Gate Procedure 5-10 

5.5.6 System Redesign 5-10 

5.6 AIDS TO CALCULATION 5-11 

5.7 PLENUM EXHAUST SYSTEMS 5-11 

5.7.1 Choice of Systems 5-11 

5.7.2 Design 5-11 

5.8 FAN PRESSURE CALCULATIONS 5-11 

5.8.1 Fan Total Pressure 5-11 

5.8.2 Fan Static Pressure 5-12 

5.8.3 Completion of the Example on Figure 5-3 . . 5-12 

5.9 CORRECTIONS FOR VELOCITY CHANGES . . . 5-12 

5.9.1 Branch Entries to Main Ducts 5-12 

5.9.2 Contractions and Expansions 5-13 

5.10 SAMPLE SYSTEM DESIGN 5-13 

5. 1 1 DIFFERENT DUCT MATERIAL FRICTION 

LOSSES 5-13 

Figure 5-1 System Duct Calculation Parameter Location . . 5-3 

Figure 5-2 Problem 1 5-5 

Figure 5-3 Velocity Pressure Method Calculation Sheet ... 5-6 

Figure 5-4 Plenum Vs. Conventional System 5-8 

Figure 5-5 Types of Plenums 5-9 

Figure 5-6 Branch Entry Velocity Correction 5-13 

Figure 5-7 Expansions and Contractions 5-14 

Figure 5-8 Problem 2 5-16 

Figure 5-9 Balanced Design Method Calculation Sheet . . .5-17 

Figure 5-10 Blast Gate Method Calculation Sheet 5-23 

Figure 5-11 System Layout 5-28 

Figure 5-12 Psychrometric Chart for Humid Air 5-30 

Figure 5-13 Hood Entry Loss Coefficients 5-40 

Figure 5-14 Duct Design Data Elbow Losses 5-41 

Figure 5- 1 5 Duct Design Data (Branch Entry Losses/ 

Weather Cap Losses) 5-42 

Figure 5-16 Duct Design Data (Static Pressure 

Regains/Losses) 5-47 



5.12 FRICTION LOSS FOR NON-CIRCULAR DUCTS .5-13 

5.13 CORRECTIONS FOR NONSTANDARD DENSITY 5- 1 5 

5.13.1 Variable Temperature and/or Different 

Altitude 5-27 

5.13.2 Elevated Moisture 5-27 

5.13.3 Psychrometric Principles 5-27 

5.13.4 Density Determination 5-28 

5.13.5 Hood Flow Rate Changes with Density . . . .5-28 

5.14 AIR CLEANING EQUIPMENT 5-32 

5.15 EVASE' DISCHARGE 5-32 

5.16 EXHAUST STACK OUTLETS 5-33 

5.16.1 Stack Considerations 5-34 

5.17 AIRBLEED-INS 5-35 

5.18 OPTIMUM ECONOMIC VELOCITY 5-35 

5.19 CONSTRUCTION GUIDELINES FOR LOCAL 
EXHAUST SYSTEMS 5-35 

5.19.1 Materials 5-35 

5.19.2 Construction 5-35 

5.19.3 System Details 5-38 

5.19.4 Codes 5-38 

5.19.5 Other Types of Duct Materials 5-38 

5.19.6 Testing 5-38 

REFERENCES 5-38 



Figure 5-17 Psychrometric Chart — Normal Temperature . . 5-54 

Figure 5-18 Psychrometric Chart — Low Temperatures . . . 5-55 

Figure 5-19 Psychrometric Chart for High Temperatures . . 5-56 

Figure 5-20 Psychrometric Chart for Very High 

Temperatures 5-57 

Figure 5-21 Principles of Duct Design Elbows 5-58 

Figure 5-22 Heavy Duty Elbows 5-59 

Figure 5-23 Cleanout Openings 5-60 

Figure 5-24 Blast Gates 5-61 

Figure 5-25 Principles of Duct Design 5-62 

Figure 5-26 Principles of Duct Design Branch Entry .... 5-63 

Figure 5-27 Principles of Duct Design Fan Inlets 5-64 

Figure 5-28 Airflow Around Buildings 5-65 

Figure 5-29 Effective Stack Height and Wake Downwash . 5-66 

Figure 5-30 Stackhead Designs 5-67 



5-2 



Industrial Ventilation 



5.1 INTRODUCTION 

The duct system that connects the hoods, air cleaning 
device(s), and fan must be properly designed. This process is 
much more involved than merely connecting pieces of duct. 
If the system is not carefully designed in a manner which 
inherently ensures that the design flow rates will be realized, 
contaminant control may not be achieved. 

The results of the following design procedure will deter- 
mine the duct sizes, material thickness, and the fan operating 
point (system flow rate and required pressure) required by the 
system. Chapter 6 describes how to select a fan based on these 
results. 

5.2 PRELIMINARY STEPS 

Coordinate design efforts with all personnel involved, in- 
cluding the equipment or process operator as well as mainte- 
nance, health, safety, fire, and environmental personnel. The 
designer should have, at a minimum, the following data 
available at the start of the design calculations: 

1. A layout of the operations, workroom, building (if 
necessary), etc. The available location(s) for the air 
cleaning device and fan should be determined. An 
important aspect that must be considered at this time 
is to locate the system exhaust point (where the air exits 
the system) so that the discharged air will not re-enter 
the work space, either through openings in the building 
perimeter or through replacement air unit intakes. (See 
Figures 5-28 and 5-29.) 

2. A line sketch of the duct system layout, including plan 
and elevation dimensions, fan location, air cleaning 
device location, etc. Number, letter, or otherwise iden- 
tify each branch and section of main duct on the line 
sketch for convenience. The examples show hoods 
numbered and other points lettered. 

Locate the fan close to pieces of equipment with high 
losses. This will facilitate balancing and may result in 
lower operating costs. 

Flexible duct is susceptible to sagging and excessive 
bending, which increases static pressure losses. Usually, 
these additional System Pressure (SP) losses cannot be 
predicted accurately. Use hard duct whenever possible 
and keep flexible duct lengths as short as possible. 

3 . A design or sketch of the desired hood for each opera- 
tion with direction and elevation of outlet for duct 
connection. 

4. Information about the details of the operation(s), spe- 
cifically toxicity, ergonomics, physical and chemical 
characteristics, required flow rate, minimum required 
duct velocity, entry losses, and required capture ve- 
locities. 

5. Consider the method and location of the replacement 



air distribution devices on the hood's performance. The 
type and location of these fixtures can dramatically 
lower contaminant control by creating undesirable 
turbulence at the hood (see Chapter 7). Perforated 
plenums or perforated duct provide better replacement 
air distribution with fewer adverse effects on hood 
performance. 

5.3 DESIGN PROCEDURE 

All exhaust systems are comprised of hoods, duct seg- 
ments, and special fittings leading to an exhaust fan. A com- 
plex system is merely an arrangement of several simple 
exhaust systems connected to a common duct. There are two 
general classes of duct system designs: tapered systems and 
plenum systems. The duct in a tapered system gradually gets 
larger as additional flows are merged together, thus keeping 
duct velocities nearly constant. If the system transports par- 
ticulate (dust, mist, or condensable vapors), the tapered sys- 
tem maintains the minimum velocity required to prevent 
settling. The duct in a plenum system (see Section 5.7) is 
generally larger than that in a tapered system, and the velocity 
in it is usually low. Any particulate in the air stream can settle 
out in the large ducts. Figures 5-4 and 5-5 illustrate design 
alternatives. Regardless of which system is used, the follow- 
ing procedure will result in a workable system design. 

1 . Select or design each exhaust hood based on the tox- 
icity, physical, and chemical characteristics of the 
material and the ergonomics of the process and deter- 
mine its design flow rate, minimum duct velocity, and 
entry losses (see Chapters 3 and 10). Note that mini- 
mum duct velocity is only important for systems trans- 
porting particulate, condensing vapors, or mist and to 
prevent explosive concentrations building up in the 
duct (see Section 5.18 for a discussion on economic 
velocities for non-particulate systems). 

2. Start with the duct segment that has the greatest 
number of duct segments between it and the fan. A duct 
segment is defined as the constant diameter round (or 
constant area rectangular) duct that separates points of 
interest such as hoods, entry points, fan inlet, etc. 

3. Determine the duct area by dividing the design flow 
rate by the minimum duct velocity. Convert the resul- 
tant cross- sectional area into a tentative duct diameter. 
A commercially available duct size (see Table 5-8) 
should be selected. If solid particulates or condensable 
vapors are being transported through the system, a 
minimum velocity is required (see Chapters 3 and 1 0). 
If the tentative duct diameter is not a standard size, 
select the next smaller size to ensure that the actual 
duct velocity is equal to or greater than the minimum 
required. 

4. Using the line sketch, determine the design length for 
each duct segment and the number and type of fittings 



Exhaust System Design Procedure 5-3 



(elbows, entries, and other special fittings) needed 
Design length is the centerline distance along the duct 
(the distance between the intersection of the center- 
lines of the straight duct components). 

5. Calculate the pressure losses for the duct segments that 
merge at a common junction point. (See Section 5.4 
for the details on how to calculate these losses.) 

6. Directly after each junction point, there must be one 
and only one SP, regardless of the path taken to reach 
that point. If not ensured by the design process, the 
system will "self-balance" by reducing the flow rate in 
the higher-resistance duct segment(s) and increasing 
the flow rate in the lower-resistance duct segment(s) 
until there is a single SP in the duct downstream of 
each junction point. 

SP balance at any junction point can be achieved by 
either one of two fundamental design methods: 1) 
Adjust the flow rate through the hood(s) until the SPs 
at each junction point are the same. 2) Increase the 
resistance in the low resistance duct segment(s) by 
means of some artificial device such as a blast gate, 
orifice plate, or other obstruction in the segment. 

Section 5.5 discusses the details of these procedures. 

7. Select both the air cleaning device and fan based upon 
final calculated system flow rate, temperature, moisture 



condition, contaminant loading, physical and chemical 
characteristics, and overall system resistance. 

8. Check the duct sizes designed against the available 
space and resolve any interference problems. (For 
example, will the elbow size desired actually fit in the 
available space?) This may cause a redesign of part of 
the system. 

9. Determine the material type and thickness (gauge) for 
each duct segment based on the air stream characteristics. 

5.4 DUCT SEGMENT CALCULATIONS 

The Velocity Pressure (VP) Method is based on the fact 
that all frictional and dynamic (fitting) losses in ducts and 
hoods are functions of the velocity pressure and can be 
calculated by a loss coefficient multiplied by the velocity 
pressure. Loss coefficients for hoods, straight ducts, elbows, 
branch entries, contractions, and expansions are shown in 
Figures 5-13 through 5-1.6. Figure 5-1 shows the application 
of these coefficients. For convenience, loss coefficients for 
round elbows and entries are also presented on the calculation 
sheet (see Figure 5-3). 

Friction data for this method are presented as Tables 5-5 
and 5-6. These tables give the loss coefficients per foot of 
galvanized and commercial steel, aluminum, PVC, and stain- 
less steel duct. The equations for these tables are listed on 




duct 3 



h e, - Fe, VP 2 



(1) See 3.5.1 and 3.5.2 

(2) See 5.9.1 



FIGURE 5-1 . System duct calculation parameter location 



5-4 



Industrial Ventilation 



these tables and also on the calculation sheet (see Figure 
5-3). (51) These equations and the resultant tables have been 
designed to be no more than 4% different from the "exact" 
values of the Colebrook- White equation and were designed 
to err on the high side of the normal velocity range of exhaust 
ventilation systems. 

For convenience, two data sets determined from the same 
equations were used to generate the friction tables. These 
tables are possible because, for a specific diameter, the friction 
loss coefficient changes only slightly with velocity. Each table 
lists the friction coefficient as a function of diameter for six 
different velocities. The error in using these data with veloci- 
ties plus or minus 1000 fpm is within 6%. If desired, a linear 
interpolation between velocity values can be performed. 

In Chapter 1, an equation was presented for flexible duct 
with the wires covered. No data are presented here for this 
type of material due to the wide variability from manufacturer 
to manufacturer. Perhaps an even more important reason is 
that these data are for straight duct losses, and flexible duct, 
by its very nature, is seldom straight. Typically, bends in 
flexible duct can produce extremely large losses which cannot 
be predicted easily. Be very careful to keep the flexible duct 
as straight and as short as possible. 

The following steps will establish the overall pressure loss 
of a duct segment that starts at a hood. Figure 5-2 shows a 
simple one-hood ventilation system. The use of a calculation 
sheet can be very beneficial when performing the calculations 
manually. Figure 5-3 shows the details of the calculations for 
each component of the system. There is also a profile through 
the system showing the magnitude and relationships of total, 
static, and velocity pressures on both the "suction" and the 
"pressure" sides of the fan on Figure 5-2. It should be noted 
that VP is always positive. Also, while total and static pressure 
may be either negative or positive with respect to atmospheric 
pressure, Total Pressure (TP) is always greater than SP (TP = 
SP + VP). 

NOTE: The numbers in the problems presented in this 
chapter were generated using one of the available 
computer programs (see Section 5.6). The values pre- 
sented in the calculation sheets may be different from 
those determined by other methods. 

1 . Determine the actual velocity by dividing the flow rate 
by the area of the commercial duct size chosen. Then 
determine the corresponding velocity pressure from 
Table 5-7 or the equations in Chapter 1 . In the example, 
the diameter chosen was 4" (line 5), the actual velocity 
is given on line 7 and the VP corresponding to this 
actual velocity is given on line 8. 

2. Determine the hood static pressure from the equations 
in Chapter 3. In this example, there are no slots, so the 
duct entry loss is as given on lines 17 through 22. 



3. Multiply the design duct length by the loss coefficient 
from the tabulated data of Tables 5-5 or 5-6 (lines 23 
through 25.) The use of galvanized sheet metal duct 
was assumed throughout this chapter. 

4. Determine the number and type of fittings in the duct 
segment For each fitting type (see Figures 5-13, 5-14, 
5-15, and 5-16), determine the loss coefficient and 
multiply by the number of fittings (there were none in 
this example.) 

5. Add the results of Steps 3 and 4 above and multiply by 
the duct VP. This is the actual loss in inches of water 
for the duct segment (given on line 34). 

6. Add the result of Step 5 to the hood suction. If there 
are any additional losses (expressed in inches of 
water), such as for an air cleaning device, add them in 
also. This establishes the cumulative energy required, 
expressed as static pressure, to move the design flow 
rate through the duct segment (line 37). Note that the 
value on line 37 is negative. 

The calculations listed in the last three columns of Figure 
5-3 will be discussed in Section 5.8.3. 

5.5 DISTRIBUTION OF AIR FLOW 

As discussed previously, a complex exhaust system is 
actually a group of simple exhaust systems connected to a 
common main duct. Therefore, when designing a system of 
multiple hoods and branches, the same rules apply. In a 
multiple branch system, however, it is necessary to provide a 
means of distributing air flow between the branches either by 
balanced design or by the use of blast gates. 

Air will always take the path of least resistance. A natural 
balance at each junction will occur; that is, the exhaust flow 
rate will distribute itself automatically according to the pres- 
sure losses of the available flow paths. The designer must 
provide distribution such that the design air flow at each hood 
will never fall below the minimum s listed in Chapter 3 and/or 
Chapter 10. To do so, the designer must make sure that all 
flow paths (ducts) entering a junction will have equal calcu- 
lated static pressure requirements. 

To accomplish this, the designer has a choice of two 
methods. The object of both methods is the same: to obtain 
the desired flow rate at each hood in the system while main- 
taining the desired velocity in each branch and main. 

The two methods, labeled Balance by Design Method and 
Blast Gate Method, are outlined below. Their relative advan- 
tages and disadvantages can be found in Table 5-1. 

5.5.1 Balance by Design Method: This procedure (see 
Section 5.10) provides for achievement of desired air flow (a 
"balanced" system) without the use of blast gates. It is often 
called the "Static Pressure Balance Method." In this type of 
design, the calculation usually begins at the hood farthest from 



Exhaust System Design Procedure 5-5 



V 16" diorn wheel, 6000 SFPM 



Fabric 
Collector 



Vertical discharge cap - 

(Fig. 5-30,5 51) f q 



409- 

408- 
5*407™ 

3 406 

{/) 

^405- 
"P.404- 

..a 
<C 

403 

402- 
401- 




De tails of Operation 



m n 

IN w. 



HOOD 
NO. 



REQUIRED AIR- 



PR! N~ 



FLOW, cfm 



1 6'' Dicirneier Grinding 
wheel, 2" Wide 



A 



80- R 



"1Q 



30 



Dimensions 



No. of Bror 


ich 




Straight 


CFM 






or Main 






Run, Ft 


Required 


Elbows 


Eotries 


ob 

be 






15 


390 
390 






cd 






1 


590 







ef 






10 


390 


--- 




fa 




S 


took Head 


390 







AMERIC AN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



PROBLEM 1 



DATE 



1-88 



FIGURE ^. 



5-6 



Industrial Ventilation 



VELOCITY PRESSURE METHOD CALCULATION SHEET 

Problem #1 Class Designer Date 


1 


Duct Segment Identification 


a-b 


b-c 


c-d 


e-f 




2 


Target Volumetric Flow Rate cfm 


390 


390 


390 


390 




3 


Minimum Transport Velocity fpm 


4000 










4 


Maximum Duct Diameter inches 


4.23 




4.5 


4.5 




5 


Selected Duct Diameter inches 


4 




4.5 


4.5 




6 


Duct Area ft 2 


0.0873 




0.1104 


0.1104 




7 


Actual Duct Velocity fpm 


4469 




3531 


3531 




8 


Duct Velocity Pressure "wg 


1.25 




0.78 


0.78 




9 


H 


D 

S 
U 
C 
T 
I 


N 


S 
L 
O 
T 
S 


Maximum Slot Area ft 2 












10 


Slot Area Selected ft 2 












11 


Slot Velocity fpm 












12 


Slot Velocity Pressure "wg 












13 


Slot Loss Coefficient 












14 


Acceleration Factor (0 or 1) 












15 


Slot Loss per VP (13 + 14) 












16 


Slot Static Pressure (12 x 15) 












17 


Duct Entry Loss Coefficient (Fig. 5-13) 


.65 




.5 






18 


Acceleration Factor (1 or 0) 


1 




1 






19 


Duct Entry Loss per VP (17+18) 


1.65 




1.5 






20 


Duct Entry Loss (8x19) "wg 


2.05 




1.17 






21 


Other Losses "wg 




2.0 








22 


Hood Static Pressure (16 + 20 + 21) "wg 


2.05 


2.0 


1.17 






23 


Straight Duct Length feet 


15 




1 


10 




24 


Friction Factor (H f ) 


0.0703 




0.0620 


0.0620 




25 


Friction Loss per VP (23 x 24) 


1.05 




0.06 


0.62 




26 


Number of 90 deg. Elbows 












27 


Elbow Loss Coefficient 












28 


Elbow Loss per VP (26x27) 












29 


Number of Branch Entries (1 or 0) 












30 


Entry Loss Coefficient 












31 


Branch Entry Loss per VP (29 x 30) 












32 


Special Fitting Loss Coefficients 












33 


Duct Loss per VP (25 + 28 + 31 + 32) 


1.05 




0.06 


0.62 




34 


Duct Loss (33 x 8) "wg 


1.31 




0.05 


0.48 




35 


Duct Segment Static Pressure Loss (22 + 34) "wg 


3.36 


2.0 


1.22 


0.48 




36 


Other Losses (VP-VP r , etc.) » wq 












37 


Cumulative Static Pressure "wg 


-3.36 


-5.36 


-6.58 


+0.48 




38 


Governing Static Pressure "wg 












39 


Corrected Volumetric Flow Rate cfm 












40 


Corrected Velocity fpm 












41 


Corrected Velocity Pressure "wg 












42 


Resultant Velocity Pressure "wg 













Exhaust System Design Procedure 



5-7 



Temperature Remarks: 
Elevation 
















1 


Pertinent Information 
From Chapter 5 
















2 
















3 
















4 


O — O 


















5 


SPg ov 
















6 


^corr ^desiqm/ op 

V or duct 
















7 
















8 


VP ' = § VP ^-| VP2 
















9 
















10 
















11 


Straight Duct Friction Loss 
V b 
















12 
















13 
















14 
















15 


Duct Material 


a 


b 


c 
















16 


Galvanized 


0.0307 


0.533 


0.612 
















17 


Black iron. Aluminum, 
PVC, Stainless steel' 


0.0425 


0.465 


0.602 
















18 
















19 


Flexible (fabric covered 
wires) 


0.0311 


0.604 


0.639 
















20 
















21 


Fan Static Pressure 

FANSP-SP out -SP in -VP in 
















22 
















23 
















24 


Branch Entry Loss Coefficients 
















25 
















26 


Anale 


Loss Coefficients 


15° 
30° 

45° 


0.09 
0.18 
0.28 
















27 
















28 
















29 


















30 


90° Round Elbow Loss Coefficients 
















31 
















32 


R/D 


Loss Coefficients 
















33 


1.5 
2.0 
2.5 


0.24 
0.19 
0.17 
















34 
















35 
















36 


60° elbow - 2/3 loss 
45° elbow = 1 / 2 loss 
30° elbow = 1/3 loss 

Adapted from Michigan Industrial 
Ventilation Conference (8/96) 
















37 
















38 
















39 
















40 
















41 














42 













5-8 



Industrial Ventilation 



Size for balance and 
transport velocity. 




TAPERED DUCT SYSTEM 

Maintains transport velocity 



\— Space collectors and fans to keep plenum- 
\size as small as practical. 



To belt drive 
fan. 



\i 



Hopper of dry 
collectors can 
discharge into 
duct. 




Air enters fabric collectors 
through hopper. 

Separate duct for 

other types. 



To fan 




^^ 




Air lock if 
required. 



^o Si « <« 5 °^° ««> 'p^ qx o^j 




Cleanout door every 10' 
Branch ducts 

EXTENDED PLENUM SYSTEM 

Self cleaning type 



NOTE: Design plenum velocities are at the most 1/2 the branch duct 
design velocities ond typically less than 2000 fpm. 



Branch ducts 



M i 



AMERICAN CONFERENCE 
OF GOVERNMENTAL 
1 INDUSTRIAL HYGIENISTS 



PLENUM vs 
CONVENTIONAL SYSTEM 



DATE 



1-88 



FIGURE 



5-4 



Exhaust System Design Procedure 5-9 



pSize plenum for 1500 - 2000 fpm. 
U. 

r \ 

\ \ 

f\ \- Drog chain 



■Size plenum for 1500 - 2000 fpm 




\ 







1. Self cleaning main - drag chain 2. Self cleaning mam bell conveying 



Size plenum for 1 500 y 

to 2000 fpm ! 

Deck plate 




10 collector 
and fan. 



& 



i r s 



Size for 
convenience 



\ K 



\ 



3. Under floor - manual cleaning 



4. Large plenum -- manual cleaning 



P i e n u m 

7 
— Hopper 



P! en urn 



/ 



A 



[Hopper 




_ 



V- To collector 



00 



xl mv 

1.1 1 



and fan 



J 



Pneumatic cleaning duct. Size for -^ 
balance and transport velocity. 
5. Hopper duct - with pneumatic cleaning 

Reference 5.3 



NOTE: Design plenum velocities ore 0!. [he most 1/2 the branch due 
design velocities and typically less than 2000 fpm. 



ERIC AN CONFERENCE 
OF GOVERNMENTAL 
INDUSTRIAL HYG1ENISTS 



TYPES OF PLENUMS 



DATE 



./ - 88 



FIGURE 



5- 5 



5-10 



Industrial Ventilation 



the fan (in terms of number of duct segments) and proceeds, 
segment by segment, to the fan. At each junction, the static 
pressure necessary to achieve desired flow in one stream must 
equal the static pressure in the joining air stream. The static 
pressures are balanced by suitable choice of duct sizes, elbow 
radii, etc., as detailed below. 

5.5.2 Blast Gate Method: The design procedure depends 
on the use of blast gates which must be adjusted after instal- 
lation to achieve the desired flow at each hood. At each 
junction, the flow rates of two joining ducts are achieved by 
blast gate adjustment which results in the desired static pres- 
sure balance. 

It is a common practice to design systems on the assumption 
that only a fraction of the total number of hoods will be used 
at a time and the flow to the branches not used will be shut 
off with dampers. For tapered system designs, where particu- 
late is transported, this practice may lead to plugging in the 
main duct due to settled particulate. 

5.5.3 Choice of Methods: The Balance by Design 
Method is normally selected where highly toxic materials are 
controlled to safeguard against tampering with blast gates and 
consequently subjecting personnel to potentially excessive 
exposures. This method is mandatory where explosives, ra- 
dioactive dusts, and biologicals are exhausted because the 
possibility of accumulations in the system caused by a blast 
gate obstruction is eliminated. 

5.5.4 Balance by Design Procedure: The pressure loss 
of each duct segment is calculated from an exhaust hood to 
the junction with the next branch based on hood design data, 
fittings, and total duct length. At each junction, the SP for each 
parallel path of air flow must be the same. Where the ratio of 
the higher SP to the lower SP is greater than 1 .2, redesign of 



the branch with the lower pressure loss should be considered. 
This may include a change of duct size, selection of different 
fittings, and/or modifications to the hood design. Where static 
pressures of parallel paths are unequal, balance can be ob- 
tained by increasing the air flow through the run with the 
lower resistance. This change in flow rate is calculated by 
noting that pressure losses vary with the velocity pressure and 
therefore as the square of the flow rate, so: 



Q 



Corrected 



Design tac7 



[5.1] 



where the "governing" SP is the desired SP at the junction 
point and the "duct" SP is that calculated for the duct segment 
being designed. 

5.5.5 Blast Gate Procedure: Data and calculations in- 
volved are the same as for the balanced design method except 
that the duct sizes, fittings, and flow rates are not adjusted; 
the blast gates are set after installation to provide the design 
flow rates. It should be noted that a change in any of the blast 
gate settings will change the flow rates in all of the other 
branches. Readjusting the blast gates during the system bal- 
ancing process sometimes can result in increases to the actual 
fan static pressure and increased fan power requirements. 

Recent work (52) describes a method whereby blast gate 
settings can be made by means of pressure readings instead 
of by velocity readings. The biggest advantage of this method 
is that the process of resetting the insertion depths need not 
be a repetitive procedure. 

5.5.6 System Redesign: Many ventilation systems are 
changed after installation (processes are changed; operations 
are relocated; additional equipment is added to the production 
floor; etc.). When such changes occur, the effect of the pro- 



TABLE 5-1. Relative Advantages and Disadvantages of the Balance by Design Method and Blast Gate Method 



Balance by Design Method 



Blast Gate Method 



1 . Flow rates cannot be changed easily by workers or at the whim of the 
operator. 

2. There is little degree of flexibility for future equipment changes or 
additions. The duct is "tailor made" for the job. 

3. The choice of exhaust flow rates for a new operation may be incorrect. 
In such cases, some duct revision may be necessary. 

4. No unusual erosion or accumulation problems will occur. 

5. Duct will not plug if velocities are chosen correctly. 

6. Total flow rate may be greater than design due to higher air 
requirements. 

7. The system must be installed exactly as designed, with all obstructions 
cleared and length of runs accurately determined. 



1 . Flow rates may be changed relatively easily. Such changes are 
desirable where pickup of unnecessary quantities of material may affect 
the process. 

2. Depending on the fan and motor selected, there is somewhat greater 
flexibility for future changes or additions, 

3. Correcting improperly estimated exhaust flow rates is relatively easy 
within certain ranges. 

4. Partially closed blast gates may cause erosion thereby changing 
resistance or causing particulate accumulation. 

5. Duct may plug if blast gate insertion depth has been adjusted 
improperly. 

6. Balance may be achieved with design flow rate; however, the net 
energy required may be greater than for the Balance by Design Method. 

7. Moderate variations in duct layout are possible. 



Exhaust System Design Procedure 5-11 



posed change(s) to the ventilation system should be calcu- 
lated. Often, systems are altered without adequate design, 
resulting in catastrophic changes to some hood flow rates. The 
result is that worker safety and health are jeopardized. 

5.6 AIDS TO CALCULATIONS 

As an alternative to performing these calculations manu- 
ally, programmable calculators and computers can be used to 
provide assistance with the design of systems. The ACGIH 
Industrial Ventilation Committee does not recommend any 
specific hardware or software. Many firms have developed 
their own software, and software packages are available com- 
mercially. Many of these software packages are available 
through ACGIH. 

5.7 PLENUM EXHAUST SYSTEMS 

Plenum systems differ from the designs illustrated earlier 
(see Figures 5-4 and 5-5). Minimum transport velocities are 
maintained only in the branch ducts to prevent settling of 
particulate matter; the main duct is oversized and velocities 
are allowed to decrease far below normal values, many times 
below 2000 fpm. The function of the main duct is to provide 
a low-pressure loss path for air flow from the various branches 
to the air cleaner or the fan. This helps to maintain balanced 
exhaust in all of the branches and often provides a minimum 
operating power. 

Advantages of the plenum-type exhaust system include: 

1. Branch ducts can be added, removed, or relocated at 
any convenient point along the main duct. 

2. Branch ducts can be closed off and the flow rate in the 
entire system reduced, provided minimum transport 
velocities are maintained in the remaining branches. 

3. The main duct can act as a primary separator (settling 
chamber) for large particulate matter and refuse mate- 
rial which might be undesirable in the air cleaner or 
fan. 

Limitations of this design include: 

1. Sticky, linty materials, such as buffing dust, tend to 
clog the main duct. It may be expected that greatest 
difficulty will be encountered with the drag chain type 
of cleaning, but the other types will be susceptible to 
buildup as well. 

2. Materials that are subject to direct or spontaneous 
combustion must be handled with care. Wood dust has 
been handled successfully in systems of this type; 
buffing dust and lint are subject to this limitation and 
are not recommended. Explosive dusts such as mag- 
nesium, aluminum, titanium, or grain dusts should not 
be handled in systems of this type. 

5.7.1 Choice of Systems: Various types of plenum ex- 
haust systems are used in industry (see Figure 5-5). They 



include both self-cleaning and manual-cleaning designs. Self- 
cleaning types include pear-shaped designs which incorporate 
a drag chain conveyor in the bottom of the duct to convey the 
dust to a chute, tote box, or enclosure for disposal. Another 
self-cleaning design uses a rectangular main with a belt con- 
veyor. In these types, the conveyors may be run continuously 
or on periodic cycles to empty the main duct before consid- 
erable buildup and clogging occur. A third type (S - 3) of self- 
cleaning design utilizes a standard conveying main duct 
system to remove the collected material from a hopper-type 
of main duct above. Such a system is usually run continuously 
to avoid clogging of the pneumatic air circuit. Manual-clean- 
ing designs may be built into the floor or may be large 
enclosures behind the equipment to be ventilated. Experience 
indicates that these should be generously oversized, particu- 
larly the underfloor designs, to permit added future exhaust 
capacity as well as convenient housekeeping intervals. 

5.7.2 Design: Control flow rates, hoods, and duct sizes for 
all branches are calculated in the same manner as with tapered 
duct systems. The branch segment with the greatest pressure 
loss will govern the static pressure required in the main duct. 
Other branches will be designed to operate at this static 
pressure or locking dampers can be used to adjust their 
pressure loss to the same static pressure as the governing 
branch. Where the main duct is relatively short or where the 
air cleaners or fans can be spaced along the duct, static 
pressure losses due to air flow in the main duct can be ignored. 
For extremely long ducts, it is necessary to calculate the static 
pressure loss along the main in a manner similar to that used 
in the balanced and blast gate methods. Design plenum ve- 
locities are at most one-half the branch velocity design duct 
velocities and typically less than 2000 fpm. Duct connections 
to air cleaners, fans, and discharge to outdoors are handled in 
the normal manner. 

5.8 FAN PRESSURE CALCULATIONS 

Exhaust system calculations are based on static pressure; 
that is, all hood static pressures and balancing or governing 
pressures at the duct junctions are given as static pressures 
which can be measured directly as described in Chapter 9. 
Most fan rating tables are based on Fan Static Pressure. An 
additional calculation is required to determine Fan Static 
Pressure before selecting the fan. 

5.8.1 Fan Total Pressure (FTP) is the increase in total 
pressure through or across the fan and can be expressed by 
the equation: 



FTP = TP utlet-TPintet 



[5-2] 



Some fan manufacturers base catalog ratings on Fan Total 
Pressure. To select a fan on this basis the Fan Total Pressure 
is calculated noting that TP = S'P + VP: 



FTP = (SPoutlet + VPoutlet) - (SPinlet + VPjnlet) 



[5.3] 



5-12 



Industrial Ventilation 



5.8.2 Fan Static Pressure: The Air Movement and Con- 
trol Association Test Code defines the Fan Static Pressure 
(FSP) as follows: "the static pressure of the fan is the total 
pressure diminished by the fan velocity pressure. The fan 
velocity pressure is defined as the pressure corresponding to 
the air velocity at the fan outlet. M(54) Fan Static Pressure can 
be expressed by the equation: 



FSP = FTP-VPiniet 



or 



FSP = SP, 



outlet - 



-SPinlet-VPj 



inlet 



[5.4] 



[5.5] 



In selecting a fan from catalog ratings, the rating tables 
should be examined to determine whether they are based on 
Fan Static Pressure or Fan Total Pressure. Fan system effects 
(see Chapter 6) should also be considered when selecting a 
fan. The proper pressure rating can then be calculated keeping 
in mind the proper algebraic signs; i.e., VP is always positive 
(+), SPi n i et is usually negative (-), and SP outlet is usually 
positive (+). 

5.8.3 Completion of the Example on Figure 5-3: To de- 
termine the Fan Static Pressure and Fan Total Pressure, note 
that the second column adds the fabric pressure drop through 
the bags in the collector. Column 3 adds the losses from the 
clean air plenum to the fan inlet, and the last column deter- 
mines the pressure losses through the stack. 

The FSP and FTP can be calculated from these values. At 
the outlet of the fan, the SP must be 0.48 "wg. At the inlet to 
the fan, the SP is -6.56 "wg. The VP at both locations is 0,78 
"wg. From Equation 5.3, the system FTP = (0.48 + 0.78) - 
(-6.56 + 0.78) - 7.04 "wg. From Equation 5.5, the FSP = 0.48 
- (-6.56)- 0.78 ■ 



■ 6.26 "wg. 



5.9 CORRECTIONS FOR VELOCITY CHANGES 

Variations in duct velocity occur at many locations in 
exhaust systems because of necessary limitations of available 
standard duct sizes (area) or due to duct selections based on 
balanced system design. As noted earlier, small accelerations 
and decelerations are usually compensated automatically in 
the system where good design practices and proper fittings are 
used. There are times, however, when special circumstances 
require the designer to have a knowledge of the energy losses 
and regains which occur since these may work to his advantage 
or disadvantage in the final performance of the system. 

5.9.1 Branch Entries to Main Ducts: Sometimes the fi- 
nal main duct velocity exceeds the higher of the two velocities 
in the branches entering the main. If the difference is signifi- 
cant, additional static pressure is required to produce the 
increased velocity. A difference of 0.10 "wg or greater be- 
tween the main VP and the resultant VP of the two branches 
should be corrected. 

At any junction point, energy must be conserved. The 



energy entering each of the two air streams would be Q(TP) 
= Q(SP+VP). The first law of thermodynamics states that the 
sum of these must equal the energy leaving, or 

Qi(VPi+SPi) + Q 2 (VP 2 +SP 2 ) = Q3(VP 3 +SP 3 ) + Losses 
Note that the overall losses would be: 

Losses - ^Q^P-i + F 2 Q 2 VP 2 

where the subscripts refer to the ducts shown in Figure 5-6. 
In this manual, F, is considered to be zero and F 2 is given on 
Figure 5-15. Assuming we are balanced and the junction 
losses are included such that SP, = SP 2 and Q 3 = Q, + Q 2 (see 
Figure 5-6), there might be an additional change in static 
pressure due to the acceleration or deceleration of the gas 
stream. The following equation shows this effect: 



SP 3 +VP 3 -SP 1 + 



£M» 1+ 






VP, 



The last two terms on the right are defined as the resultant 
velocity pressure, VP r ; this can be simplified to 



VP r = |fHVP 1 + 






VP 9 



[5.6] 



where: 

VP r = resultant velocity pressure of the combined 

branches 
Q., = flow rate in branch #1 
Q 2 = flow rate in branch #2 
Q 3 = combined flow rate leaving the junction 

Note that the above equation is valid for all conditions, 
including merging different density gas streams, as long as 
the velocity pressures include the density effects. Also note 
that, if the flow rate through one branch was changed to 
balance at the branch entry, the velocity pressure and cor- 
rected flow rates should be used in Equation 5.6. 

The resultant velocity pressure (VP r ) is computed using 
Equation 5.6. If VP 3 is less than VP r , a deceleration has 
occurred and SP has increased. If VP 3 is greater than VP r , an 
acceleration has occurred, and the difference between VP 3 and 
VP r is the necessary loss in SP required to produce the increase 
in kinetic energy between VP 3 and VP r . The correction is 
made as follows: 

SP 3 = SPi - (VP 3 - VP r ) [5.7] 

where: 

SP 3 = SP in main #3 

SP! = SP at branch #1 - SP at branch #2 

VP 3 = velocity pressure in main #3 

It should be noted that many designers believe a conserva- 
tive approach to fan selection would be to ignore any correc- 
tion if VP r is larger than VP 3 . 



Exhaust System Design Procedure 



5-13 



A simpler equation for VP r was used in prior editions of 
this manual: 



VR 



Qi+Qa 



4005(A 1 +A 2 



This equation gives acceptable results (less than a 4% error) 
when the velocities of the two merging air streams are within 
500 fpm of each other. 

EXAMPLE 




Duct. No. 


Dia. 


Area 





V 


VP 


SP 


0) 


10 


0.545 


1935 


3550 


0,79 


-2.11 


(2) 


A 


0.087 


340 


3890 


0.94 


~2.11 


Main (3) 


10 


0.545 


2275 


4170 


1.08 


..... 



FIGURE 5-6. Branch entry velocity correction 



With the data shown, 

(1935)(0.79) (340)(0.94) 



VP r = 



2275 



2275 



:0.81"wg 



SP 3 = SP 1 -(VP 3 -VP r ) = -2.11 -(108 -0.81) 
= - 2.1 1 -0.27 = - 2.38 "wg 

Therefore, in this situation, an additional -0.27 "wg should 
be added to the junction SP to account for losses in pressure 
due to acceleration of the air stream. 

5.9.2 Contractions and Expansions: Contractions are 
used when the size of the duct must be reduced to fit into tight 
places, to fit equipment, or to provide a high discharge veloc- 
ity at the end of the stack. Expansions are used to fit a 
particular piece of equipment or to reduce the energy con- 
sumed in the system by reducing velocity and friction. Expan- 
sions are not desirable in transport systems since the duct 
velocity may become less than the minimum transport veloc- 
ity and material may settle in the ducts. 

Regain of pressure in a duct system is possible because 
static pressure and velocity pressure are mutually convertible. 
This conversion is accompanied by some energy loss. The 
amount of this loss is a function of the geometry of the 
transition piece (the more abrupt the change in velocity, the 
greater the loss) and depends on whether air is accelerated or 
decelerated. Loss is expressed as a loss coefficient multiplied 
by the velocity pressure in the smaller area duct of the transi- 
tion piece. One minus the loss factor is the efficiency of the 
energy conversion or regain. 



A perfect (no loss) contraction or expansion would cause 
no change in the total pressure in the duct. There would be an 
increase or decrease in static pressure corresponding exactly 
to the decrease or increase in velocity pressure of the air. In 
practice, the contraction or expansion will not be perfect, and 
there will be a change in total pressure (see Figure 5-7). In 
each example, total pressure and static pressure are plotted to 
show their relationship at various points in each system. See 
Figure 5-16 for design data. 

5.10 SAMPLE SYSTEM DESIGN 

A discussion of the calculations for either tapered duct 
method can best be done by a typical example using the 
exhaust system shown in Figure 5-8. Calculation sheets illus- 
trate the orderly and concise arrangement of data and calcu- 
lations (see Figures 5-9 and 5-10). The procedure outlined in 
Section 5.3 was used to develop the design. Each column is 
for a constant diameter duct segment that starts at a hood, 
junction point, air cleaning device, fan, or transition point. 

The problem considered is a foundry sand-handling and 
shake-out system. A minimum conveying velocity of 3500 
fpm is used throughout the problem except in ducts where 
excess moisture or dust loading increases that value. The 
operations, hood designations on the diagram, VS-print ref- 
erences, and required flow rates are presented in Table 5-2. 

5.11 DIFFERENT DUCT MATERIAL FRICTION LOSSES 

The friction loss table, Table 5-5, provides average values 
for galvanized sheet metal duct material (0.0005 feet equiva- 
lent sand grain roughness, where the roughness height repre- 
sents the average height of the roughness elements of the 
material). Table 5-6 provides the same information for black 
iron and other materials possessing a roughness height of 
0.00015 feet. Recent research indicates that an equivalent 
sand grain roughness factor of 0.0003 feet more accurately 
reflects the losses incurred in new HVAC galvanized duct 
systems. However, past experiences in industrial ventilation 
applications successfully reinforce the application of the 
0.0005 feet equivalent sand grain roughness factor. This may 
be due to inherently shorter duct runs and dustier environ- 
ments which are common within industrial ventilation appli- 
cations. The values in both tables can be used with no 
significant error for the majority of designs but special con- 
siderations may be desired if environmental conditions could 
significantly affect the duct design parameters. If the design 
requires special material, operates at a non-standard density, 
or is very hot, the duct material manufacturer should be 
consulted for the anticipated friction loss. 

5.12 FRICTION LOSS FOR NON-CIRCULAR DUCTS 

Round ducts are preferred for industrial exhaust systems 
because they provide a more uniform air velocity to resist 
settling of material and an ability to withstand higher static 



5-14 



Industrial Ventilation 



'L^U^\ 




EXAMPLE 1 — DUCT LOCATED ON SUCTION SIDE 
OF FAN 

Velocity changes as indicated. Since all the duct is on 
the suction side of the fan, TP at the fan inlet (point F) is 
equal to VP at the fan inlet plus the total duct resistance up 
to that point. This equals -4.2 SP since static pressure on 
the suction side of the fan is always negative. The duct 
system is the same as was used in Example 2 and therefore 
has the same overall resistance of 3.2. If it is again as- 
sumed that the inlet and discharge of the fan are equal 
areas, the total pressure across the fan will be the same as 
in Example 2 and, in each case, the fan will deliverthe same 
air horsepower when handling equal volumes of air. 

Static pressure conversion between B and C follows 
contraction formula (Figure 5-16). There must be sufficient 
SP at B to furnish the additional VP required at C. In 
addition, the energy transfer between these two points is 
accompanied by a loss of 0.3. Since SP at B = -2, SP at C 
= -2.0 + (-1 .0) + (-0.3) = -3.3 "wg. 

Static pressure regain between D and E follows the 
regain formulae (Figure 5-16). If there were no loss in the 
transition piece, the difference of 1 "wg in velocity pressure 
would be regained as static pressure at E, and SP at that 
point would be -2.8. However, the transition is only 60% 
efficient (0.4 loss) so the SP at E = -2.8 + (-0.4) = -3.2. 



30% loss 



N)% loss 




-l-4 
. -i. 


n 


i j 




A 77 ^""--^ 


^ 








-I- 2 


"--4$ 


"2.9 
,2.2 


2.4 




2 














Q.T 




\^ 



' A ionospheric pressure 

EXAMPLE 2 — DUCT LOCATED ON DISCHARGE 
SIDE OF THE FAN 

Velocity changes as indicated. The duct is located on 
the discharge side of the fan. Total pressure at the fan 
discharge (point A) is equal to the velocity pressure at the 
discharge end of the duct (point F) plus the accumulated 
resistances. These add up to 1 .0 + 1 .0 + 0.4 + 0.5 + 0.3 
+ 1.0 = 4.2. 

Static pressure regain between D and E follows the 
regain formulae (Figure 5-16). If there were no energy 
loss in the transition piece, static pressure at D would be 
because the difference in VP of 1 would show up as 
static pressure regain. However, the transition is only 
60% efficient which means a loss of 0.4, so SP at point D 
= + 0.4 = 0.4. 

Conversion of static pressure into velocity pressure 
between B and C follows contraction formulae (Figure 
5-16). There must be sufficient static pressure at B to 
furnish the additional velocity pressure required at C. In 
addition, transformation of energy between these two 
points is accompanied by a loss of 0.3. Since SP at C = 
0.9, SP at B = 0.9 + 0.3 + 1 .0 = 2.2. Since there is no duct 
on the suction side of the fan, total pressure against which 
the fan is operating is 4.2". 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
iNDUSTRiAL HYGIENISTS 



DATE 



EXPANSIONS 

AND CONTRACTIONS 



1 - 88 



I 



'IGURE 



b-Y 



Exhaust System Design Procedure 5-15 



TABLE 5-2. Details of Operation 



No. 



Hood No. VS-Print 



Minimum 
Exhaust, cfm 



1. Vibrating Shakeout 
4' x 6' grate 


1 


20-02 


9600 


2. Shakeout hopper 


2 


20-03 


960 


3. Vibrating pan feeder 
24" wide 


3 


20-03 


700 


4. Incline sand belt 
24" x 28" long 


5 




700 


5. Magnetic pulley 








6. Tramp iron box 








7. Bucket elevator 
24" x 30" casing 


7a(iower) 
7b(upper) 


50-01 


250 
250 


8. Vibrating screen 24 ft 2 


8 


99-01 


1200 


9. Sand bin 600 ft 3 
1 8" x 20" opening 


9 


50-10 


500 


10. Waster sand box 44" x 
54", 6" clearance 


10 


99-03 
(V = 150fpm) 


1225 


11. Sand weigh hopper 


11 


60-02 


900 


12. Sandmul!er6'dia. 


12 


60-02 




13. Wet dust collector 
(includes fan) 









DIMENSIONS 



No. of 
Branch 
or Main 



CFM 
Required, 
Minimum 



Straight 
Run, ft Elbows 



Entries 



1-A 


9600 


13 


1-90° 




2-B 


960 


3 


1-60° 


1-30° 


3-B 


700 


4 


1-90°+1-60° 


1-30° 


B-A 


1660 


18 


2-90° 


1-30° 


A-C 


11,260 


34 






5-D 


700 


7 


1-30°+1-60° 


1-30° 


7a-D 


250 


5 






D-C 


950 


14 


1-90°+1-60° 


1-30° 


C-E 


12,210 


6.5 






8-F 


1200 


11 


2-90° 




9-F 


500 


4 


1-90°+1-60° 


1-30° 


F-G 


1700 


5 






7b-G 


250 


15 


1-60° 


1-30° 


G-E 


1950 


6 


1-60° 


1-30° 


E-H 


14,160 


3,5 






10- J 


1225 


6 


1-45° 




12- J 


900 


2.5 


1-30° 


1-30° 


J-H 


2125 


8 


1-90°+1-60° 


1-30° 


H-K 


16,285 


9 


2-45° 




13 


16,285 








14-L 


16,285 


20 







pressure. At times, however, the designer must use other duct 

shapes. 

Rectangular duct friction can be calculated by using Table 
5-5 or 5-6 in conjunction with Table 5-9 to obtain rectangular 
equivalents for circular ducts on the basis of equal friction 
loss. It should be noted that, on this basis, the area of the 
rectangular duct will be larger than the equivalent round duct; 
consequently, the actual air velocity in the duct will be re- 
duced. Therefore, it is necessary to use care to maintain 
minimum transport velocities. 

Occasionally, the designer will find it necessary to estimate 
the air handling ability of odd-shaped ducts. The following 
procedure^ 5) will be helpful in determining the frictional 
pressure losses for such ducts. The wetted perimeter in the 
following discussion is the inside perimeter of the odd-shaped 
duct corresponding to the cross-sectional area. 

1. Find duct cross-sectional area, ft 2 A 

2. Find wetted perimeter, ft P 

3. Calculate hydraulic radius, ft R (R - A/P) 

4. Convert R to inches r(r= 12R) 

5. Calculate equivalent diameter, in D (D = 4r) 

6. Use the proper friction table based on the equivalent 
diameter and flow rate (or velocity). 

5.13 CORRECTIONS FOR NONSTANDARD DENSITY 

Fan tables and exhaust flow rate requirements assume a 
standard air density of 0.075 Ibm/ft 3 , which corresponds to 
sea level pressure, no moisture, and 70 F. Changes in air 
density can come from several factors including elevation, 
temperature, internal duct pressure, changes in apparent molecu- 
lar weight (moisture content, gas stream constituents, etc.), and 
amount of suspended particulate. Where appreciable variation 
occurs, the change in air density must be considered. 

Factors for different temperatures and elevations are listed 
in Table 5-10. Correction for temperatures between 40 F and 
100 F and/or elevations between -1 ,000 feet and +1 ,000 feet 
are seldom required with the permissible variations in usual 
exhaust system design. 

Similarly, if internal duct pressures vary by more than 20 
M wg from standard pressure, the density will change by over 
5%. If there is excessive moisture in the airstream, the density 
will decrease. Suspended particulate is assumed to be only a 
trace impurity in industrial exhaust systems. If there are 
significant quantities of particulate in the duct system, this 
addition to the air stream density should be addressed. This 
field is called material conveying and is beyond the scope of 
this manual. 

Many times, the system designer is confronted with a 
combination of these five means of changing density. If so, 
then the density of the air stream should be determined and, 
if the density is more than 5% different from standard density, 



5-16 Industrial Ventilation 





,<- \ 




S 




3 








10 


rfr" 


$} Sc' ^_ 


\| 






< 




r 






\ 7 iai 


< \ $4 7 


7^ 


-J 






^J 




D 








J5- 


.--''9 


S 


^ 


t 


114 
13 




K wj 


LDb» 




\ 


^ 


C^ A> 




1 








E 


3 














"tU 



Plan View 



All elbows - t radius = 2.0 D (5 piece) 

Branch entries - 30° 

All duct lengths are £ to t 




Elevation View 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



PROBLEM 2 



DATE 



1-88 8 



FIGURE 



5-8 



Exhaust System Design Procedure 5-17 



VELOCITY PRESSURE METHOD CALCULATION SHEET 

Problem #2 Class Designer Date 


1 


Duct Segment Identification 


2-B 


3-B 


B-A 


1-A 


1-A 


2 


Target Volumetric Flow Rate 




cfm 


960 


700 


1711 


9600 


9600 


3 


Minimum Transport Velocity 




fpm 


4000 


4000 


4000 


4000 


4000 


4 


Maximum Duct Diameter 




inches 


6.63 


5.66 


8.86 


20.97 




5 


Selected Duct Diameter 




inches 


6 


5.5 


8 


20 


18 


6 


Duct Area 




ft 2 


0.1963 


0.1650 


0.3491 


2.1817 


1.7671 


7 


Actual Duct Velocity 




fpm 


4889 


4243 


4902 


4400 


5432 


8 


Duct Velocity Pressure 




"wg 


1.49 


1.12 


1.50 


1.21 


1.84 


9 


H 

O 
D 

S 

U 

c 

T 
I 


N 


S 
L 
O 

T 
S 


Maximum Slot Area 




ft 2 








4.8 


4.8 


10 


Slot Area Selected 




ft 2 








4.8 


4.8 


11 


Slot Velocity 




fpm 








2000 


2000 


12 


Slot Velocity Pressure 




"wg 








0.25 


0.25 


13 


Slot Loss Coefficient 








1.78 


1.78 


14 


Acceleration Factor 




(0or1) 














15 


Slot Loss per VP 


(13 + 14) 










1.78 


1.78 


16 


Slot Static Pressure 


(12x15) 










0.44 


0.44 


17 


Duct Entry Loss Coefficient 


(Fig. 5-13) 




.25 


.25 




.25 


.25 


18 


Acceleration Factor 




(1 or 0) 


1 


1 




1 


1 


19 


Duct Entry Loss per VP 


(17 + 18) 




1.25 


1.25 




1.25 


1.25 


20 


Duct Entry Loss 


(8x 19) 


"wg 


1.86 


1.40 




1.51 


2.30 


21 


Other Losses 




"wg 












22 


Hood Static Pressure 


(16 + 20 + 21) 


"wg 


1.86 


1.40 




1.95 


2.74 


23 


Straight Duct Length 




feet 


3 


4 


18 


13 


13 


24 


Friction Factor (H f ) 


.0425 


.0478 


.0299 


.0098 


.0110 


25 


Friction Loss per VP 


(23 x 24) 




.13 


.19 


.54 


.13 


.14 


26 


Number of 90 deg. Elbows 


.67 


1.67 


2 


1 


1 


27 


Elbow Loss Coefficient 


.19 


.19 


.19 


.19 


.33 


28 


Elbow Loss per VP 


(26 x 27) 




.13 


.32 


.33 


.19 


.33 


29 


Number of Branch Entries 




(1 or 0) 


1 


1 


1 






30 


Entry Loss Coefficient 


.18 


.18 


.18 






31 


Branch Entry Loss per VP 


(29 x 30) 




.18 


.18 


.18 






32 


Special Fitting Loss Coefficients 












33 


Duct Loss per VP 


(25 + 28 + 31 + 32) 




.43 


.69 


1.10 


.32 


.47 


34 


Duct Loss 


(33 x 8) 


"wg 


.65 


.77 


1.64 


.38 


.87 


35 


Duct Segment Static Pressure Loss 


(22 + 34) 


M wg 


-2.51 


-2.18 


»164 


-2.34 


-3.61 


36 


Other Losses (VP-VP n etc.) 




"wq 






.10 






37 


Cumulative Static Pressure 




"wg 


-2.51 


-2.18 


-4.25 


-2.34 


-3.61 


38 


Governing Static Pressure 




"wg 




-2.51 






-4.25 


39 


Corrected Volumetric Flow Rate 




cfm 




751 






10416 


40 


Corrected Velocity 




fpm 




4551 






5894 


41 


Corrected Velocity Pressure 




"wg 




1.29 






2.16 


42 


Resultant Velocity Pressure 




"wg 


1.40 


1.40 


2.07 




2.07 



FIGURE 5-9. Balanced design method 



5-18 



Industrial Ventiiation 





Temperature 
Elevation 








Remarks: 






A-C 


7a-D 


5-D 


5-D 


D-C 


D-C 


C-E 


1 


Pertinent Information 
From Chapter 5 


12127 


250 


700 


700 


955 


955 


13130 


2 


4000 


3500 


3500 


3500 


3500 


3500 


4000 


3 


23.58 


3.62 


6.05 




7.07 




24.53 


4 


s°\ , /■<* 






22 


3.5 


6 


5.5 


7 


6 


24 


5 


I ^*gov 




2.6398 


0.0668 


0.1963 


0.1650 


0.2673 


0.1963 


3.1416 


6 


^corr "" ^design \j nn 

V or duct 


4594 


3742 


3565 


4243 


3573 


4864 


4179 


7 


1.32 


0.87 


0.79 


1.12 


0.80 


1.47 


1.09 


8 


VP r = -§JVP 1+ <| VP 2 
















9 
















10 
















11 


Straight Duct Friction Loss 
V b 
















12 
















13 
















14 
















15 


Duct Material 


a 


b 


c 
















16 


Galvanized 


0.0155 


0.533 


0.612 




1.0 


.25 


.25 








17 


Black iron. Aluminum, 
PVC, Stainless steel 


0.662 


0.645 


0.602 




1 


1 


1 








18 




2.0 


1.25 


1.25 








19 


Flexible (fabric covered 
wires) 


0.0186 


0.604 


0.639 




1.74 


.99 


1.40 








20 
















21 


Fan Static Pressure 

FANSP = SP 0Ut -SP in -VP in 




1.74 


.99 


1.40 








22 


34 


5 


7 


7 


14 


14 


6.5 


23 


.0087 


.0840 


.0436 


.0478 


.0361 


.0425 


.0079 


24 


Branch Ent**v Loss Cnpfficipnts 


.30 


.42 


.30 


.33 


.51 


.60 


.05 


25 


Angle 


Loss Coefficients 






1 


1 


1.67 


1.67 




26 


15° 
30° 
45° 


0.09 
0.18 
0.28 






.19 


.19 


.19 


.19 




27 






.19 


.19 


.32 


.32 




28 






1 


1 


1 


1 




29 








.18 


.18 


.18 


.18 




30 


90° Round Elbow Loss Coefficients 

(5 piece) 






.18 


.18 


.18 


.18 




31 
















32 


R/D 


Loss Coefficients 


.30 


.42 


.67 


.70 


1.01 


1.10 


.05 


33 


1.5 

2.0 
2.5 


0.24 
0.19 
0.17 


.39 


.37 


.53 


.79 


.80 


1.61 


.06 


34 


.39 


2.11 


1.52 


2.19 


.80 


1.61 


.06 


35 












.41 




36 


60° elbow -2/3 loss 
45° elbow = 1/2 loss 
30° elbow = 1/3 loss 

! Adapted from Michigan Industrial 
Ventilation Conference (8/96) 


-4.64 


-2.11 


-1.52 


-2.19 


-2.99 


-4.21 


-4.70 


37 




-2.19 








-4.64 




38 




255 








1003 




39 




3817 








5109 




40 




.91 








1.63 




■41 


1.34 


1.06 




1.06 




1.34 


1.15 


42 













FIGURE 5-9. Balanced design method (continued) 



Exhaust System Design Procedure 



5-19 



VELOCITY PRESSURE METHOD CALCULATION SHEET 

Problem #2 (Continued) Class Designer Date 


1 


Duct Segment Identification 


8-F 


9-F 


9-F 


F-G 


7b-G 


2 


Target Volumetric Flow Rate cfm 


1200 


500 


500 


1730 


250 


3 


Minimum Transport Velocity fpm 


3500 


3500 


3500 


3500 


3500 


4 


Maximum Duct Diameter inches 


7.93 


5.12 




9.52 


3.62 


5 


Selected Duct Diameter inches 


7 


5 


4.5 


9 


3.5 


6 


Duct Area ft 2 


0.2673 


0.1364 


0.1104 


0.4418 


0.0668 


7 


Actual Duct Velocity fpm 


4490 


3667 


4527 


3916 


3742 


8 


Duct Velocity Pressure "wg 


1.26 


0.84 


1.28 


0.96 


0.87 


9 


H 
O 

D 

S 

U 

c 

T 
I 


N 


S 

L 
O 
T 
S 


Maximum Slot Area ft 2 












10 


Slot Area Selected ft 2 












11 


Slot Velocity fpm 












12 


Slot Velocity Pressure "wg 












13 


Slot Loss Coefficient 












14 


Acceleration Factor (0 or 1) 












15 


Slot Loss per VP (13+14) 












16 


Slot Static Pressure (12x15) 












17 


Duct Entry Loss Coefficient (Fig. 5-13) 


.5 


.25 


.25 




1.0 


18 


Acceleration Factor (1 or 0) 


1 


1 


1 




1 


19 


Duct Entry Loss per VP (17 + 18) 


1.5 


1.25 


1.25 




2.0 


20 


Duct Entry Loss (8x19) "wg 


1.89 


1.05 


1.60 




1.74 


21 


Other Losses "wg 












22 


Hood Static Pressure (16 + 20 + 21) "wg 


1.89 


1.05 


1.60 




1.74 


23 


Straight Duct Length feet 


11 


4 


4 


5 


15 


24 


Friction Factor (H f ) 


.0354 


.0543 


.0608 


.0263 


.0840 


25 


Friction Loss per VP (23 x 24) 


.39 


,22 


.24 


.12 


1.26 


26 


Number of 90 deg. Elbows 


2 


1.67 


1.67 




.67 


27 


Elbow Loss Coefficient 


.19 


.19 


.19 




.19 


28 


Elbow Loss per VP (26 x 27) 


.38 


.32 


.32 




.13 


29 


Number of Branch Entries (1 or 0) 




1 


1 




1 


30 


Entry Loss Coefficient 




.18 


.18 




.18 


31 


Branch Entry Loss per VP (29 x 30) 




.18 


.18 




.18 


32 


Special Fitting Loss Coefficients 












33 


Duct Loss per VP (25 + 28 + 31 + 32) 


.77 


.71 


.74 


.12 


1.57 


34 


Duct Loss (33 x 8) "wg 


.97 


.60 


.95 


.12 


1.37 


35 


Duct Segment Static Pressure Loss (22 + 34) "wg 


2.86 


1.65 


2.55 


.12 


3.11 


36 


Other Losses (VP-VP r> etc.) « wq 












37 


Cumulative Static Pressure "wg 


-2.86 


-1.65 


-2.55 


-2.98 


-3.11 


38 


Governing Static Pressure "wg 






-2.86 


-3.11 




39 


Corrected Volumetric Flow Rate cfm 






530 


1767 




40 


Corrected Velocity fpm 






4801 


3999 




41 


Corrected Velocity Pressure "wg 






1.44 


1.00 




42 


Resultant Velocity Pressure "wg 


1.31 




1.31 


.98 


,98 



FIGURE 5-9. Balanced design method (continued) 



5-20 



Industrial Ventilation 





Temperature 
Elevation 








Remarks: 






G-E 


G-E 


E-H 


12- J 


10- J 


10- J 


J-H 


1 


Pertinent Information 
From Chapter 5 


2017 


2017 


15308 


900 


1225 


1225 


2274 


2 


3500 


3500 


4000 


4500 


3500 


3500 


4500 


3 


10.28 




26.49 


6.06 


8.01 




9.57 


4 












10 


9 


26 


6 


8 


7 


9 


5 


q =q i sp g°v 

^corr ^design ^/ en 

V or duct 


0.5454 


0.4418 


3.687 


0.1963 


0.3491 


0.2673 


0.4418 


6 


3698 


4566 


4152 


4584 


3509 


4584 


5147 


7 


0.85 


1.32 


1.07 


1.31 


0.77 


1.31 


1.65 


8 


VP ' = % Vp i + T5T VPz 
















9 
















10 
















11 


Straight Duct Friction Loss 
V b 
















12 
















13 
















14 
















15 


Duct Materia! 


a 


b 


c 
















16 


Galvanized 


0.0307 


0.533 


0.612 








.25 


.25 


.25 




17 


Black iron, Aluminum, 
PVC, Stainless steel 


0.0425 


0.465 


0.602 








1 


1 


1 




18 








1.25 


1.25 


1.25 




19 


Flexible (fabric covered 
wires) 


0.0311 


0.604 


0.639 








1.64 


.96 


1.64 




20 
















21 


Fan Static Pressure 

FANSP-SP out -SP in -VP in 








1.64 


.96 


1.64 




22 


6 


6 


3.5 


11 


6 


6 


8 


23 


.0232 


.0260 


.0072 


.0427 


.0307 


.0354 


.0258 


24 




.14 


.16 


.03 


.47 


.18 


.21 


.21 


25 


Branch Entry Loss Coefficients 


.67 


.67 




.33 


.5 


.5 


1.67 


26 


Anale 


Loss Coefficients 


15° 
30° 
45° 


0.09 
0.18 
0.28 


.19 


.19 




.19 


.19 


.19 


.19 


27 


.13 


.13 




.06 


.10 


.10 


.32 


28 


1 


1 




1 






1 


29 






.18 


.18 




.18 






.18 


30 


90° Round Elbow Loss Coefficients 

(5 piece) 


.18 


.18 




.18 






.18 


31 
















32 


R/D 


Loss Coefficients 


.45 


.47 


.03 


.71 


.28 


.31 


.61 


33 


1.5 
2.0 
2.5 


0.24 
0.19 
0.17 


.38 


.60 


.03 


.94 


.21 


.41 


1.01 


34 


.38 


.60 


.03 


2.58 


1.17 


2.05 


1.01 


35 




.34 










.18 


36 


60° elbow - 2/3 loss 


-3.49 


-4.05 


-4.73 


-2.58 


-1.17 


-2.05 


-3.77 


37 


45° elbow = 1/2 loss 




-4.70 








-2.58 


-4.73 


38 


30° elbow = 1/3 loss 

Adapted from Michigan Industrial 
Ventilation Conference (8/96) 




2178 








1374 


2492 


39 




4930 








5140 


5641 


40 




1.52 








1.65 


1.98 


41 




1.15 


1.20 


1.47 




1.47 


1.20 


42 













FIGURE 5-9. Balanced design method (continued) 



Exhaust System Design Procedure 



5-21 



VELOCITY PRESSURE METHOD CALCULATION SHEET 

Problem #2 (Continued) Class Designer Date 


1 


Duct Segment Identification 


H-K 


K-FAN 


FAN-L 






2 


Target Volumetric Flow Rate cfm 


17800 


17800 


17800 






3 


Minimum Transport Velocity fpm 


4500 


2600 


2600 






4 


Maximum Duct Diameter inches 


26.93 


35.4 








5 


Selected Duct Diameter inches 


26 


35.5 


34 






6 


Duct Area ft 2 


3.6870 


6.8736 


6.3050 






7 


Actual Duct Velocity fpm 


4828 


2590 


2823 






8 


Duct Velocity Pressure "wg 


1.45 


0.42 


0.50 






9 


H 



D 

S 

u 
c 

T 
I 


N 


S 
L 
O 
T 
S 


Maximum Slot Area ft 2 












10 


Slot Area Selected ft 2 












11 


Slot Velocity fpm 












12 


Slot Velocity Pressure "wg 












13 


Slot Loss Coefficient 












14 


Acceleration Factor (0 or 1 ) 












15 


Slot Loss per VP (13 + 14) 












16 


Slot Static Pressure (12 x 15) 












17 


Duct Entry Loss Coefficient (Fig. 5-13) 












18 


Acceleration Factor (1 or 0) 












19 


Duct Entry Loss per VP (17 + 18) 












20 


Duct Entry Loss (8 x 19) "wg 












21 


Other Losses "wg 




4.5 


.08 






22 


Hood Static Pressure (16 + 20 + 21) "wg 




4.5 


.08 






23 


Straight Duct Length feet 


9 


2 


20 






24 


Friction Factor (H f ) 


.0071 


.0051 


.0053 






25 


Friction Loss per VP (23 x 24) 


.06 


.01 


.11 






26 


Number of 90 deg. Elbows 


1 










27 


Elbow Loss Coefficient 


.19 










28 


Elbow Loss per VP (26 x 27) 


.19 










29 


Number of Branch Entries (1 or 0) 












30 


Entry Loss Coefficient 












31 


Branch Entry Loss per VP (29 x 30) 












32 


Special Fitting Loss Coefficients 












33 


Duct Loss per VP (25 + 28 + 31 + 32) 


.25 


.01 


.11 






34 


Duct Loss (33 x 8) "wg 


.37 





.05 






35 


Duct Segment Static Pressure Loss (22 + 34) "wg 


.37 


4.5 


.13 






36 


Other Losses (VP-VP r , etc.) » wq 


.25 










37 


Cumulative Static Pressure "wg 


-5.35 


-9.85 


.13 






38 


Governing Static Pressure "wg 












39 


Corrected Volumetric Flow Rate cfm 












40 


Corrected Velocity fpm 












41 


Corrected Velocity Pressure "wg 












42 


Resultant Velocity Pressure "wg 













FIGURE 5-9. Balanced design method (continued) 



5-22 



Industrial Ventilation 



NOTES 

1. Balancing at B: SP ratio = -2.51 /-2.1 8 = 1.15, so the 
flow rate through the lower resistance run can be 
corrected. From Equation 5.1, 



2. 



Q 



corrected 



= 700 



-2.51 



'-2.18 
From Equation 5.6, 



= 751 cfm 



VP r = 



960 : 1.49 + ^i 129 = 1 40 "wg 



1711 



1711 



Note that the numbers for 3-B reflected the corrected 
values. The velocity pressure of 1.29 "wg corre- 
sponds to the new velocity after correcting the flow 
rate to 751 cfm (751/0.165 = 4551 fpm, up from 4243 
fpm). 

The VP in duct B-A is 1 .50, while the VP r at A was 
1 .40, so there was an acceleration at B of 0, 1 0. This 
is reflected in the overall pressure drop to point A 
(2.51 + 0.10 + 1.64 = 4.25 "wg.) Column 19 is a 
convenient place to enter this additional SP drop. 

Balancing at A: Initial SP ratio at A (-4.25/-2.34 = 
1.82) is too high to allow flow rate increase, so 1-A 
was redesigned with an 18" dia. duct to result in an 
acceptable SP ratio (-4.25 ■*■ -3.61 =1.18). Then 



rf corrected 



:9600j^^ = 10416 cfm 
-3.61 



segment from A to C was clearly less than VP r , 
correction to the SP was made. 



no 



From Equation 5.6, VP r = 2.07 "wg. As the VP in the 



3. This same procedure was followed at each of the 
other junction points until reaching the collector. At 
junction points J and H, the SP ratio was slightly over 
the 1.20 recommended. However, the flows were 
increased using Equation 5.1 anyway because re- 
ducing the diameter would have resulted in unac- 
ceptably high duct velocities and SP drops. 

4. K is the inlet to that collector and the fan is labeled 
FAN. Assuming the collector loss of 4.5 "wg is the 
"flange-to-flange" loss, and 2 feet of straight duct 
separates the fan from the collector, it might be 
advisable to use a non-standard diameter duct equal 
to the fan's inlet to connect the two devices. That is 
why a 35.5" diameter was chosen here. 

5. The duct loss calculations forthe stack (FAN-L) show 
an overall loss of 0.05 "wg. However, there is an 
additional effect to consider. Most centrifugal fans 
have an exit area virtually the same as the inlet area. 
If so, then there would be an additional acceleration 
from approximately 0.42 "wg to the duct VP of 0.50 
"wg, or 0.08 "wg. This acceleration "loss" is included 
in line 19 of the FAN-L column. 

6. The above FSP calculation assumes that there are 
no losses due to the contraction from the fan exit to 
the 32" diameter duct. This is usually a small enough 
loss to ignore. For instance, if the contraction half- 
angle was 1 5 degrees with a loss factor of 0.08 (see 
Figure 5-16), the maximum error in ignoring this loss 
would be less than 0.02 "wg. 



FIGURE 5-9. Balanced design method (continued) 



Exhaust System Design Procedure 



5-23 



VELOCITY PRESSURE METHOD CALCULATION SHEET 

Problem #2 With Blast Gates Class Designer Date 


1 


Duct Segment Identification 


1-A 


2-B 


3-B 


B-A 


A-C 


2 


Target Volumetric Flow Rate cfm 


9600 


960 


700 


1660 


11260 


3 


Minimum Transport Velocity fpm 


4000 


4000 


4000 


4000 


4000 


4 


Maximum Duct Diameter inches 


20.97 


6.63 


5.66 


8.72 


22.72 


5 


Selected Duct Diameter inches 


20 


6 


5.5 


8 


22 


6 


Duct Area ft 2 


2.1917 


0.1963 


0.1650 


0.3491 


2.6398 


7 


Actual Duct Velocity fpm 


4400 


4889 


4243 


4756 


4265 


8 


Duct Velocity Pressure "wg 


1.21 


1.49 


1.12 


1.41 


1.13 


9 


H 


D 

S 

U 

c 

T 
I 

O 
N 


S 

L 
O 
T 
S 


Maximum Slot Area ft 2 


4.8 










10 


Slot Area Selected ft 2 


4.8 










11 


Slot Velocity fpm 


2000 










12 


Slot Velocity Pressure "wg 


.25 










13 


Slot Loss Coefficient 


1.78 










14 


Acceleration Factor (0 or 1) 













15 


Slot Loss per VP (13+ 14) 


1.78 










16 


Slot Static Pressure (12x15) 


.44 










17 


Duct Entry Loss Coefficient (Fig. 5-13) 


.25 


.25 


.25 






18 


Acceleration Factor (1 or 0) 


1 


1 


1 






19 


Duct Entry Loss per VP (17 + 18) 


1.25 


1.25 


1.25 






20 


Duct Entry Loss (8x19) "wg 


1.51 


1.86 


1.40 






21 


Other Losses "wg 












22 


Hood Static Pressure (16 + 20 + 21) "wg 


1.95 


1.86 


1.40 






23 


Straight Duct Length feet 


13 


3 


4 


18 


34 


24 


Friction Factor (H f ) 


0.0098 


0.0425 


0.0478 


0.0299 


0.0088 


25 


Friction Loss per VP (23 x 24) 


.13 


.13 


.19 


.54 


.30 


26 


Number of 90 deg. Elbows 


1 


.67 


1.67 


2 




27 


Elbow Loss Coefficient 


.19 


.19 


.19 


.19 




28 


Elbow Loss per VP (26 x 27) 


.19 


.13 


.32 


.38 




29 


Number of Branch Entries (1 or 0) 




1 


1 






30 


Entry Loss Coefficient 




.18 


.18 






31 


Branch Entry Loss per VP (29 x 30) 




.18 


.18 






32 


Special Fitting Loss Coefficients 












33 


Duct Loss per VP (25 + 28 + 31 + 32) 


.32 


.44 


.69 


.92 


.30 


34 


Duct Loss (33 x 8) "wg 


.38 


.65 


.77 


1.30 


.34 


35 


Duct Segment Static Pressure Loss (22 + 34) "wg 


2.33 


2.51 


2.17 


1.30 


.34 


36 


Other Losses (VP-VP n etc.) .- wa 








0.07 




37 


Cumulative Static Pressure "wg 


-2.33 


-2.51 


-2.17 


-3.88 


-4.22 


38 


Governing Static Pressure "wg 


-3.88 




-2.51 






39 


Corrected Volumetric Flow Rate cfm 


GATE 




GATE 






40 


Corrected Velocity fpm 












41 


Corrected Velocity Pressure "wg 












42 


Resultant Velocity Pressure "wg 


1.24 


1.34 


1.34 


1.24 


1.14 



FIGURE 5-10. Blast gate method 



5-24 



Industrial Ventilation 





Temperature 
Elevation 








Remarks: 






5-D 


7a-D 


D-C 


C-E 


8-F 


9-F 


F-G 


1 


Pertinent Information 
From Chapter 5 


700 


250 


950 


12210 


1200 


500 


1700 


2 


3500 


3500 


3500 


4000 


3500 


3500 


3500 


3 


6.05 


3.62 


7.05 


23.65 


7.93 


5.12 


9.44 


4 












6 


3.5 


7 


22 


7 


5 


9 


5 


^**corr ^design \j on 

V ^ r duct 


0.1963 


0.0668 


0.2673 


2.6398 


0.2673 


0.1364 


0.4418 


6 


3565 


3742 


3555 


4625 


4490 


3667 


3848 


7 


0.79 


0.87 


0.79 


1.33 


1.26 


0.84 


0.92 


8 


vp ' = t vp ^ + § vp * 
















9 
















10 
















11 


Straight Duct Friction Loss 

V b 
H f - a Q, 
















12 
















13 
















14 
















15 


Duct Material 


a 


b 


c 
















16 


Galvanized 


0.0155 


0.533 


0.612 


.25 


1.0 






.5 


.25 




17 


Black iron, Aluminum, 
PVC, Stainless steel 


0.662 


0.645 


0.602 


1 


1 






1 


1 




18 


1.25 


2.0 






1.5 


1.25 




19 


Flexible (fabric covered 
wires) 


0.0186 


0.604 


0.639 


.99 


1.74 






1.89 


1.05 




20 
















21 


Fan Static Pressure 

FANSP = SP out -SP in -VP jn 


.99 


1.74 






1.89 


1.05 




22 


7 


5 


14 


6.5 


11 


4 


5 


23 


0.0436 


0.0840 


0.0361 


0.0087 


0.0354 


0.0543 


0.0264 


24 


Branch Ent»*v f.n«s f'neffirients 


.30 


.42 


.51 


.06 


.39 


.22 


.13 


25 


Angle 


Loss Coefficients 


1 




1.67 




2 


1.67 




26 


15° 
30° 

45° 


0.09 
0.18 
0.28 


.19 




.19 




.19 


.19 




27 


.19 




.32 




.38 


.32 




28 


1 




1 






1 




29 




.18 




.18 






.18 




30 


90° Round Elbow Loss Coefficients 

(5 piece) 


.18 




.18 






.18 




31 
















32 


R/D 


Loss Coefficients 


.67 


.42 


1.01 


.06 


.77 


.71 


.13 


33 


1.5 
2.0 
2.5 


0.24 
0.19 
0.17 


.53 


.37 


.79 


.08 


.97 


.60 


.12 


34 


1.52 


2.11 


.79 


.08 


2.85 


1.65 


.12 


35 








.19 








36 


60° elbow -2/3 loss 
45° elbow = 1/2 loss 
30° elbow =1/3 loss 

Adapted from Michigan Industrial 
Ventilation Conference (8/96) 


-1.52 


-2.11 


-2.90 


-4.49 


-2.85 


-1.65 


-2.97 


37 


-2.11 




-4.22 






-2.85 


-3.11 


38 


GATE 




GATE 






GATE 


GATE 


39 
















40 
















41 


.81 


.81 


1.14 


1.26 


1.13 


1.13 


.92 


42 













FIGURE 5-10. Blast gate method (continued) 



Exhaust System Design Procedure 5-25 



VELOCITY PRESSURE METHOD CALCULATION SHEET 

Problem #2 With Blast Gates Class Date 


1 


Duct Segment Identification 


7b-G 


G-E 


E-H 


12- J 


10- J 


2 


Target Volumetric Flow Rate cfm 


250 


1950 


14160 


900 


1225 


3 


Minimum Transport Velocity fpm 


3500 


3500 


4000 


4500 


3500 


4 


Maximum Duct Diameter inches 


3.62 


10.1 


25.48 


6.06 


8.01 


5 


Selected Duct Diameter inches 


3.5 


10 


24 


6 


8 


6 


Duct Area ft 2 


0.0668 


0.5454 


3.1416 


0.1963 


0.3491 


7 


Actual Duct Velocity fpm 


3742 


3575 


4507 


4584 


3509 


8 


Duct Velocity Pressure "wg 


0.87 


0.80 


1.27 


1.31 


0.77 


9 


H 


D 

S 
U 
C 

T 
I 


N 


S 

L 

T 
S 


Maximum Slot Area ft 2 












10 


Slot Area Selected ft 2 












11 


Slot Velocity fpm 












12 


Slot Velocity Pressure "wg 












13 


Slot Loss Coefficient 












14 


Acceleration Factor (0 or 1) 












15 


Slot Loss per VP (13 + 14) 












16 


Slot Static Pressure (12x15) 












17 


Duct Entry Loss Coefficient (Fig. 5-13) 


1.0 






.25 


.25 


18 


Acceleration Factor (1 or 0) 


1 






1 


1 


19 


Duct Entry Loss per VP (17+18) 


2.00 






1.25 


1.25 


20 


Duct Entry Loss (8x19) "wg 


1.74 






1.64 


.96 


21 


Other Losses "wg 












22 


Hood Static Pressure (16 + 20 + 21) "wg 


1.74 






1.64 


.96 


23 


Straight Duct Length feet 


15 


6 


3.5 


11 


6 


24 


Friction Factor (H f ) 


0.0840 


0.0233 


0.0078 


0.0427 


0.0307 


25 


Friction Loss per VP (23 x 24) 


1.26 


.14 


.03 


.47 


.18 


26 


Number of 90 deg. Elbows 


.67 


.67 




.33 


.50 


27 


Elbow Loss Coefficient 


.19 


.19 




.19 


.19 


28 


Elbow Loss per VP (26 x 27) 


.13 


.13 




06 


.10 


29 


Number of Branch Entries (1 or 0) 


1 






1 




30 


Entry Loss Coefficient 


.18 






.18 




31 


Branch Entry Loss per VP (29 x 30) 


.18 






.18 




32 


Special Fitting Loss Coefficients 












33 


Duct Loss per VP (25 + 28 + 31 + 32) 


1.57 


.94 


.03 


.71 


.28 


34 


Duct Loss (33 x 8) "wg 


1.37 


.75 


.03 


.94 


.21 


35 


Duct Segment Static Pressure Loss (22 + 34) "wg 


3.11 


.75 


.03 


2.58 


1.17 


36 


Other Losses (VP-VP r , etc.) » wa 












37 


Cumulative Static Pressure "wg 


-3.11 


-3.86 


-4.52 


-2.58 


-1.17 


38 


Governing Static Pressure "wg 




-4.49 






-2.58 


39 


Corrected Volumetric Flow Rate cfm 




GATE 






GATE 


40 


Corrected Velocity fpm 












41 


Corrected Velocity Pressure "wg 












42 


Resultant Velocity Pressure "wg 


.92 


1.26 


1.29 


1.00 


1.00 



FIGURE 5-10. Blast gate method (continued) 



5-26 



Industrial Ventilation 





Temperature 
Elevation 


Remarks: 




J-H 


H-K 


K-FAN 


FAN-L 








1 


Pertinent Information 
From Chapter 5 


2125 


16285 


16285 


16285 








2 


4500 


4500 


2600 


2600 








3 


9.31 


25.76 


33.9 










4 




9 


26 


35.5 


34 








5 


0.4418 


3.6870 


6.8736 


6.305 








6 


^corr- ^design 






fCD 




4810 


4417 


2369 


2583 








7 


v ^ r duct 


1.44 


1.22 


0.35 


0.42 








8 


VP = -?! VP + ?L. VP 
Vr ' Q 3 VKl+ Q 3 Vr 2 
















9 
















10 
















11 


Straight Duct Friction Loss 
V b 
















12 
















13 
















14 
















15 


Duct Material 


a 


b 


c 
















16 


Galvanized 


0.0307 


0.533 


0.612 
















17 


Black iron, Aluminum, 
PVC, Stainless steel 


0.0425 


0.465 


0.602 
















18 
















19 


Flexible (fabric covered 
wires) 


0.0311 


0.604 


0.639 
















20 






4.50 


.08 








21 


Fan Static Pressure 

FANSP-SP out -SP in -VP in 






4.50 


.08 








22 


8 


9 




20 








23 


0.0259 


0.0072 




0.0053 








24 




.21 


.06 




.11 








25 


Branch Entry Loss Coefficients 

Anale Loss Coefficients 


1.67 


1.0 












26 


15° 
30° 
45° 


0.09 
0.18 
0.28 


.19 


.19 












27 


.32 


.19 












28 


1 














29 




.18 














30 


90° Round Elbow Loss Coefficients 


.18 














31 
















32 




V v k 1 ^^/ 


.70 


.25 




.11 








33 


R/D 


Loss Coefficients 


1.5 
2.0 
2.5 


0.24 
0.19 
0.17 


1.02 


.31 




.04 








34 


1.02 


.31 


4.50 


.12 








35 


.44 














36 


60° elbow = 273 loss 
45° elbow = 1/2 loss 
30° elbow = 1/3 loss 

Adapted from Michigan Industrial 
Ventilation Conference (8/96) 


-4.04 


-4.83 


-9.33 


.12 








37 


-4.52 














38 


GATE 














39 
















40 
















41 


1.29 














42 













FIGURE 5-10. Blast gate method (continued) 



Exhaust System Design Procedure 5-27 



corrections to the system design by the following means 
should be made. 

The density variation equations of Chapter 1 (Section 1.4) 
demonstrate that if temperature increases or absolute pressure 
decreases, the density will decrease. For the mass flow rate at 
the hood(s) to remain the same, the flow rate must change if 
density changes. It is helpful to remember that a fan connected 
to a given system will exhaust the same volume flow rate 
regardless of air density. The mass of air moved, however, 
will be a function of the density. 

5.13.1 Variable Temperature and/or Different Altitude: 

Consider an exhaust system at sea level where 5000 cfm 
of air at 70 F is drawn into a hood. The air is then heated to 
600 F and the density of the air leaving the heater becomes 
0.0375 lbm/ft 3 . The flow rate downstream of the heater would 
be 10,000 actual cubic feet per minute (acfm) at the new 
density of 0.0375 lbm/ft 3 . This is true because the 50% 
decrease in density must correspond to a twofold increase in 
the volume flow rate since the mass flow rate has remained 
constant. 

If this temperature effect is ignored and a fan selected for 
5000 cfm is placed in the system, the hood flow rate will be 
well below that required to maintain contaminant control. The 
exact operating point of such a system would have to be 
recalculated based upon the operating point of the incorrectly 
sized fan. 

5.13.2 Elevated Moisture: When air temperature is under 
100 F, no correction for humidity is necessary. When air 
temperature exceeds 100 F and moisture content is greater 
than 0.02 lbs H 2 per pound of dry air, correction is required 
to determine fan operating RPM and power. Correction coef- 
ficients may be read from the psychrometric charts such as 
those illustrated in Figures 5-17 through 5-20. 

5.13.3 Psychrometric Principles: The properties of 
moist air are presented on the psychrometric chart at a single 
pressure. These parameters define the physical properties of 
an air/water vapor mixture. The actual gas flow rate and the 
density of the gas stream at the inlet of the fan must be known 
in order to select the fan. The psychrometric chart provides 
the information required to calculate changes in the flow rate 
and density of the gas as it passes through the various exhaust 
system components. These properties are: 

• Dry-Bulb Temperature is the temperature observed 
with an ordinary thermometer. Expressed in degrees 
Fahrenheit, it may be read directly on the chart and is 
indicated on the bottom horizontal scale. 

e Wet-Bulb Temperature is the temperature at which 
liquid or solid water, by evaporating into air, can bring 
the air to saturation adiabatically at the same 
temperature. Expressed in degrees Fahrenheit, it is read 
directly at the intersection of the constant enthalpy line 



with the 100% saturation curve. 

Dew Point Temperature is that temperature at which 
the air in an air/vapor mixture becomes saturated with 
water vapor and any further reduction of dry bulb 
temperature causes the water vapor to condense or 
deposit as drops of water. Expressed in degrees 
Fahrenheit, it is read directly at the intersection of the 
saturation curve with a horizontal line representing 
constant moisture content. 

Percent Saturation curves reflect the mass of moisture 
actually in the air as a percentage of the total amount 
possible at the various dry bulb and moisture content 
combinations. Expressed in percent, it may be read 
directly from the curved lines on the chart. 

Density Factor is a dimensionless quantity which 
expresses the ratio of the actual density of the mixture 
to the density of standard air (0.075 lbm/ft 3 ). The lines 
representing density factor typically do not appear on 
low-temperature psychrometric charts when relative 
humidity or percent saturation curves are presented. A 
method of calculating the density of the gas defined by 
a point on the chart (when density factor curves are not 
presented) is discussed in Section 5.13.4. 

Moisture Content, or weight of water vapor, is the 
amount of water that has been evaporated into the air. 
In ordinary air, it is very low pressure steam and has 
been evaporated into the air at a temperature 
corresponding to the boiling point of water at that low 
pressure. Moisture content is expressed in grains of 
water vapor per pound of dry air (7000 grains = one 
pound) or pounds of water vapor per pound of dry air 
and is read directly from a vertical axis. 

Enthalpy (Total Heat) as shown on the psychrometric 
chart is the sum of the heat required to raise the 
temperature of a pound of air from F to the dry-bulb 
temperature, plus the heat required to raise the 
temperature of the water contained in that pound of air 
from 32 F to the dew point temperature, plus the latent 
heat of vaporization, plus the heat required to superheat 
the vapor in a pound of air from the dew point 
temperature to the dry-bulb temperature. Expressed in 
British Thermal Units per pound of dry air, it is shown 
by following the diagonal wet-bulb temperature lines. 

Humid Volume is the volume occupied by the air/vapor 
mixture per pound of dry air and is expressed in cubic 
feet of mixture per pound of dry air. it is most important 
to understand the dimensions of this parameter and 
realize that the reciprocal of humid volume is not 
density. Humid volume is the parameter used most 
frequently in determining flow rate changes within a 
system as a result of mixing gases of different properties 



5-28 



Industrial Ventilation 



Elbow: (^R=1.5D (4 piece) 




30' 



FIGURE 5-11. System layout 

or when evaporative cooling occurs within the system. 

5.13.4 Density Determination: When the quality of an 
air/vapor mixture is determined by a point on a psychrometric 
chart having a family of density factor curves, all that must 
be done to determine the actual density of the gas at the 
pressure reference for which the chart is drawn is to multiply 
the density factor taken from the chart by the density of 
standard air (0.075 lbm/ft 3 ). Should relative humidity curves 
be presented on the chart in lieu of density factor curves, 
information available through dimensional analysis must be 
used to determine the actual density of the mixture. This can 
be done quite easily as follows: The summation of one pound 
of dry air plus the mass of the moisture contained within that 
pound of dry air divided by the humid volume will result in 
the actual density of the mixture. 



1 + W 
HV 



[5.8] 



where: 



p= density of the mix (lbm/ft 3 ) 
W = moisture content (lbm H 2 0/lbm dry air) 
HV = humid volume (ft 3 mix/lbm dry air) 

5. 13.5 Hood Fiow Rate Changes with Density: I f the 

density of the air entering a hood is different from standard 
density due to changes in elevation, ambient pressure, tem- 
perature, or moisture, the flow rate through the hood should 
be changed to keep the mass flow rate the same as for standard 
air. This can be accomplished by multiplying the hood flow 
rate required for standard air by the ratio of the density of 
standard air to the actual ambient density. 

The example shown in Figure 5-11 illustrates the effect of 
elevated moisture and temperature and a method of calculation: 

EXAMPLE 

GIVEN: The exit flow rate from a 60" x 24' dryer is 16,000 
scfm plus removed moisture. The exhaust air temperature is 
500 F. The drier delivers 60 tons/hr of dried material with 



H 



\/ ran 

Wet: collector 



capacity to remove 5% moisture. Required suction at the dryer 
hood is - 2.0 f, wg; minimum conveying velocity must be 4000 
fpm(see Figure 5-11). 

It has been determined that the air pollution control system 
should include a cyclone for dry product recovery and a 
high-energy wet collector. These devices have the following 
operating characteristics: 

• Cyclone: Pressure loss is 4.5 "wg at rated flow rate of 
35,000 scfm. The pressure loss across any cyclone 
varies directly with any change in density and as the 
square of any change in flow rate from the rated 
conditions. 

• High-Energy Wet Scrubber. The manufacturer has 
determined that a pressure loss of 20 "wg is required in 
order to meet existing air pollution regulations and has 
sized the collector accordingly. The humidifying 
efficiency of the wet collector is 90%. 

NOTE: As a practical matter, a high energy scrubber 
as described in this example would have essentially 
100% humidifying efficiency. The assumption of 90% 
humidifying efficiency along with a high pressure drop 
allows discussion of multiple design considerations in 
one example and was therefore adopted for instruc- 
tional purposes. 

• Fan: A size #34 "XYZ" fan with the performance shown 
in Table 5-3 has been recommended. 

REQUIRED: 

Size the duct and select fan RPM and motor size. 

SOLUTION: 

Stepl 

Find the actual gas flow rate that must be exhausted from 
the dryer. This flow rate must include both the air used for 
drying and the water, as vapor, which has been removed from 



Exhaust System Design Procedure 5-29 



the product. Since it is actual flow rate, it must be corrected 
from standard air conditions to reflect the actual moisture, 
temperature, and pressures which exist in the duct. 

Step 1A: 

Find the amount (weight) of water vapor exhausted. 

Dryer discharge = 60 tons/hr of dried material (given) 

Since the dryer has capacity to remove 5% moisture, the 
dryer discharge is 95% x dryer feed rate. 

60 tons/hr dried material = (0.95) (dryer feed) 
60 tons/hr 



dryer feed = - 



0.95 



= 63.2 tons/hr 



Moisture removed = (feed rate) - (discharge rate) 
= 63.2 tons/hr - 60 tons/hr 
= 6400 Ibs/hr or 106.7 Ibm/min 

Step 1 B 

Find the amount (weight) of dry air exhausted. 

Dry air exhausted = 16,000 scfm at 70 F and 
29.92 "Hg (0.075 lbs/ft 3 density) 

Exhaust rate, Ibs/min 



(16,000 scfm)(0.075 Ibs/fT) 
1200 Ibs/min dry air 



StepIC 



Knowing the water-to-dry air ratio and the temperature of 
the mixture, it is possible to determine other quantities of the 
air-to-water mixture. This can be accomplished by the use of 
psychrometric charts (see Figures 5-17 to 5-20) which are 
most useful tools when working with humid air. 

W= 0.089 lbs H 2 0/lb dry air 

Dry bulb temperature = 500 F (given) 



The intersection of the 500 F dry-bulb temperature line and 
the 0.089 lbs H 2 0/lb dry air line can be located on the 
psychrometric chart (see Figure 5-12). Point #1 completely 
defines the quality of the air and water mixture. Other data 
relative to this specific mixture can be read as follows: 



Dew Point Temperature: 
Wet-Bulb Temperature: 
Humid volume, ft 3 of mix/lb of dry air: 
Enthalpy, BTU/lb of dry air; 
Density factor, df: 



122 F 

145 F 

27.5 ft 3 /lb dry air 

235 BTU/lb dry air 

0.53 



StepID 



Find actual gas flow rate, (acfm). 

Exhaust flow rate, acfm = (humid volume)(weight of dry 
air/min). Humid volume, HV, was found in Step 1C as 27.5 
ftVlb. Weight of dry air/min was found in Step IB as 1200 
lb/min. Exhaust flow rate = (27.5 ft 3 /lb)(1200 Ib/min) = 
33,000 acfm. 

Step 2 

Size the duct. Minimum conveying velocity of 4000 fpm 
was given. Suction at the dryer exit of -2.0 "wg corresponds 
to hood suction. 

The duct area equals the actual flow rate divided by the 
minimum duct velocity, or A - 33,000 -*■ 4,000 - 8.25 ft 2 . A 
38" diameter duct with a cross-sectional area of 7.876 ft 2 
should be chosen as this is the largest size available with an 
area smaller than calculated. Then the actual duct velocity 
would be 33,000 actual ftVmin - 7.876 ft 2 - 4,190 fpm. 

Step 2A 

The velocity pressure in the duct cannot be found using the 
equation VP = (V -*- 4005) 2 , as this equation is for standard air 



TABLE 5-3. Fan Rating Table 


Fan size No. 34 Inlet diameter = 34" Max. safe rpm = 1700 


CFM 


20" SP 


22" SP 


24" SP 


26" SP 


28" SP 


30" SP 


32" SP 


34" SP 


36" SP 


38" SP 


40" SP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


14688 
16524 
18360 


1171 
1181 
1191 


73.3 
81.8 
90.2 


1225 
1234 
1244 


81.4 
90.2 
99.5 


1277 
1286 
1294 


89.8 
98.8 
108 


1326 
1335 
1344 


98.3 
107 
118 


1374 
1382 
1391 


107 
116 
127 


1421 
1428 
1437 


116 
126 
137 


1466 
1472 
1481 


125 
135 
146 


1510 
1516 
1524 


134 
145 
157 


1552 
1557 
1565 


143 
155 
167 


1594 
1600 
1606 


153 
165 
178 


1634 
1639 
1645 


162 
175 
188 


20196 
22032 
23868 


1204 
1217 
1230 


99.9 
110 
120 


1256 
1268 
1282 


109 
120 
131 


1306 
1318 
1331 


119 
130 
142 


1354 
1366 
1378 


129 
141 
154 


1400 
1412 
1424 


139 
151 
165 


1446 
1456 
1468 


149 
162 
176 


1490 
1499 
1511 


160 
173 
187 


1532 
1542 
1553 


170 
184 
199 


1574 
1584 
1594 


181 
196 
211 


1615 
1624 
1633 


191 
207 
223 


1654 
1663 
1672 


202 
218 
235 


25704 
27540 
29376 


1245 
1261 
1277 


131 
143 
156 


1296 
1311 
1327 


143 
156 
169 


1345 
1359 
1374 


155 
168 
182 


1391 
1406 
1421 


167 
181 
196 


1437 
1450 
1465 


179 
193 
209 


1481 
1494 
1508 


191 
206 
222 


1524 
1537 
1550 


203 
219 
236 


1565 
1578 
1591 


215 
232 
249 


1606 
1618 
1631 


227 
245 
263 


1645 
1658 
1670 


239 
258 
277 


1683 
1695 


252 
271 


31212 
33048 
34884 


1295 
1313 
1331 


170 
184 
198 


1344 
1361 
1379 


184 
198 
214 


1391 
1407 
1425 


197 
213 
229 


1436 
1453 
1469 


211 
228 
245 


1480 
1496 
1513 


225 
242 
260 


1523 
1538 
1555 


239 
257 
276 


1564 
1580 
1595 


253 
272 
291 


1605 
1620 
1635 


268 
287 
307 


1644 
1659 
1674 


282 
302 
323 


1683 
1697 


297 
317 







5-30 



Industrial Ventilation 



- Humid Volume - ft J /lb Dry Air 
Density Factor-Mixture 



1 A ■ 



1 30 • 



120 • 




200 500 

Dry Bulb Temperature, F 



FIGURE 5-12. Psychrometric chart for humid air (see Figures 5-17 through 5-20) 



only. The actual velocity pressure in the duct is given by 

VPactual - (df)(VP s td) 

where: df - density factor 

As the density factor was determined in Step 1C, the actual 
velocity pressure in the duct will be 

VP = (df)(VP s td) = (0.53)(1 .09 "wg) = 0.58 "wg 

Step 3 

Calculate the pressure loss from A to B and determine static 
pressure at Point B. 

The data from Figure 5-14 and Table 5-5 can be used 
directly. The static pressure loss through the duct can be found 
by multiplying the length of duct by the friction coefficient, 
adding the elbow loss coefficient, and multiplying the result 
by the duct velocity pressure: 

SP loss = [(0.0045)(30) + 0.27][0.58] = (0.405)(0.58) 
= 0.23 "wg 

Then the static pressure at the inlet to the cyclone should 
be -2.23 "wg (hood suction plus friction and fitting losses). 

Step 4 

The pressure loss of the cyclone is provided by the manu- 
facturer. In this example, the cyclone pressure loss is 4.5 "wg 
at a rated flow of 35,000 scfm. The pressure loss through a 
cyclone, as with duct, varies as the square of the change in 
flow rate and directly with change in density. 

Therefore, the actual loss through the cyclone would be 



(4.5) 



33,000 



35,000 



(0.53)- -2.12 "wg 



and the static pressure at the cyclone outlet would be -4.35 
"wg. 

Step 5 

The calculation from Point C to D is the same as from A to 
B in Step 3. Thus, the static pressure at the wet collector inlet 
would be 

-4.35 - (0.0045)(1 5)(0.58) = -4.39 "wg 

NOTE: Information for Steps 6 and 7 which involve 
calculation of changes in flow rate, density, etc., across 
the wet collector should be provided by the equipment 
manufacturer. 

Step 6 

An important characteristic of wet collectors is their ability 
to humidify a gas stream. The humid ifi cation process is 
generally assumed to be adiabatic (without gain or loss of heat 
to the surroundings). Therefore, water vapor is added to the 
mixture, but the enthalpy, expressed in BTU/lb dry air, re- 
mains unchanged. During the process of humidification, the 
point on the psychrometric chart that defines the quality of 
the mixture moves to the left, along a line of constant enthalpy, 
toward saturation. 

All wet collectors do not have the same ability to humidify. 
If a collector is capable of taking an air stream to complete 
adiabatic saturation, it is said to have a humidifying efficiency 
of 100%. The humidifying efficiency of a given device may 
be expressed by either of the following equations: 

Tin^-^xlOO 



T-T„ 



where: 



Exhaust System Design Procedure 5-31 



n n = humidifying efficiency, % 
"I" - dry-bulb temperature at collector inlet, F 
T - dry-bulb temperature at collector outlet, F 
T s = adiabatic saturation temperature, F 



or 



TV 



W-W r 



w-w c 



5-xlOO 



where: 



Wj = moisture content in lb H 2 0/lb dry air at inlet 
W = moisture content in lb H 2 0/lb dry air at outlet 
W, 



s - moisture content in lb H 2 0/lb dry air at adiabatic 

saturation conditions 



Step 6A 



Find the quality of the air to water mixture at Point 2, the 
collector outlet. 

Humidifying Efficiency == 90% (given). Dry -bulb Tem- 
perature at Collector Inlet = 500 F (given). Adiabatic satura- 
tion temperature = 145 F from inspection of Psychrometric 
Chart. 



90 = 



, (500 -t ) 
"(500-145) 



x100 



where: 

to =180F 

Then the air leaving the collector will have a dry-bulb tem- 
perature of 1 80 F and an enthalpy of 235 BTU/lb of dry air as 
the humidifying process does not change the total heat or 
enthalpy. 

The point of intersection of 180 F dry bulb and 235 BTU/lb 
dry air on the psychrometric chart defines the quality of the 
air leaving the collector and allows other data to be read from 
the chart as follows: 



Dew Point Temperature 


143 F 


Wet-Bulb Temperature 


145 F 


Humid Volume, ft 3 /lb dry air 


20.5 ft 3 /lb dry air 


Enthalpy, BTU/lb dry air 


235 BTU/lb dry air 


Density factor, df 


0.76 



Step 7 

What is the exhaust flow rate in acfm and the density factor 
at the collector outlet? 

Step 7A 

Exhaust flow rate = (humid volume)(weight of dry 
air/min). Humid Volume from Step 6 is 20.5 ftVlb dry air. 
Weight of dry air/min from Step IB is 1200 Ibs/min. Flow 
rate = (20.5 ft 3 /lb)(1200 lbs/min) - 24,600 acfm. 



As the wet collector loss was stated to be 20 "wg, the static 
pressure at the wet collector outlet would be -24.39 "wg. 

Step 7B 

On low-pressure exhaust systems, where the negative pres- 
sure at the fan inlet is less than 20 "wg, the effect of the 
negative pressure is usually ignored. However, as the pres- 
sures decrease, or the magnitude of negative pressures in- 
creases, it is understood that gases expand to occupy a larger 
volume. Unless this larger volume is anticipated and the fan 
sized to handle the larger flow rate, it will have the effect of 
reducing the amount of air that is pulled into the hood at the 
beginning of the system. From the characteristic equation for 
the ideal gas laws, PQ = wRT (where w = the mass flow rate 
in Ibm/min), the pressure flow rate relationship is 



or 



Up to this point, the air has been considered to be at standard 
atmospheric pressure which is 14.7 psia, 29.92 "Hg or 407 
"wg. The pressure within the duct at Point E is -24.4 "wg and 
minus or negative only in relation to the pressure outside the 
duct which is 407 "wg. Therefore, the absolute pressure 
within the duct is 407 "wg -24.4 "wg = 382.6 "wg. 

407 Q 9 



P1Q1 


= P2Q2 


Ri_ 
p 2 


Q 2 
Qi 



382.6 24,600 cfm 

Q2 = 26,170 acfm 

Step 7C 

Pressure also affects the density of the air. From PQ = wRT 
the relationship 



(w 2 - Q 2 )RT 2 P 2 



can be derived. Density factor is directly proportional to the 
density and the equation can be rewritten 



Rl_ 
p 2 


df 2 




Substitute 




407 




0.76 



382.6 df 2 
(d^ was determined to be 0.76 in Step 6.) 

df 2 = 0.71 
Step 7D 

The duct from the wet collector to the fan can now be sized. 



5-32 



Industrial Ventilation 



The flow rate leaving the wet collector was 26,170 acfm. As 
the fan selected has a 34-in. diameter inlet (area = 6.305 ft 2 ), 
it is logical to make the duct from the wet collector to the fan 
a 34-in. diameter. Thus, the velocity through the duct would 
be 26,170-6.305 =4,151 fpra.TheVP would be (0.71)(4151 
- 4005) 2 - (0.71 )( 1.07) = 0.76 "wg. 

Step 7E 

The duct pressure loss, based on 26,170 cfm and a 34-in. 
diameter duct, would be (0.0052)(5)(0.76) = 0.02 "wg. There- 
fore, the SP at the fan inlet would be -24.41 "wg. 

Step 8 

Calculate the pressure loss from fan discharge F to stack 
discharge G. Since the air is now on the discharge side of the 
fan, the pressure is very near atmospheric. No pressure cor- 
rection is needed. The flow rate and density factor are 24,600 
acfm and 0.76, respectively. 

Assuming that the fan discharge area is nearly the same as 
at the fan inlet, the same 34-in. diameter duct would result in 
a velocity of 3902 fpm. The velocity pressure would be 
(0.76)(3902 - 4005) 2 - 0.72 "wg. 

From Table 5-5, the friction coefficient is 0.0052 and the 
frictional pressure loss for the 30 ft. high stack would be 
(0.0052)(30)(0.72) = 0.1 1 "wg. As the static pressure at the 
exit of the stack must be atmospheric, the static pressure at 
the fan exit will be positive. 

Step 9 

Determine actual fan static pressure. 

Actual FSP = SPout - SPm - VP in 

= + 0.11 -(-24.41) -0.76 
= 23.76 "wg 

Step 10 

Determine equivalent fan static pressure in order to enter 
fan rating table. Equivalent fan static pressure is determined 
by dividing the actual fan static pressure by the density factor 
at the fan inlet. This is necessary since fan rating tables are 
based on standard air. 



Equivalent FSP ■ 



Step 1 1 



23.76 
0.71 



33.46 "wg 



Select fan from rating table using the equivalent fan SP and 
the fan inlet flow rate. Interpolating the fan rating table (Table 
5-3) for 26,200 cfm at 33.5 "wg yields a fan speed of 1559 
RPMat217BHP. 

Step 12 

Determine the actual required fan power. Since actual 



density is less than standard air density, the actual required 
power is determined by multiplying by the density factor, or 
(217 BHP)(0.71) = 154 BHP. If a damper is installed in the 
duct to prevent overloading of the motor, at cold start the 
motor need only be a 200 HP (see Chapter 6). 

5.14 AIR CLEANING EQUIPMENT 

Dusts, fumes, and toxic or corrosive gases should not be 
discharged to the atmosphere. Each exhaust system handling 
such materials should be provided with an adequate air cleaner 
as outlined in Chapter 4. As a rule, the exhaust fan should be 
located on the clean air side of such equipment. An exception 
is in the use of cyclone cleaners where the hopper discharge 
is not tightly sealed and better performance is obtained by 
putting the fan ahead of the collector. 

5.15 EVASE DISCHARGE 

An evase discharge is a gradual enlargement at the outlet 
of the exhaust system (see Figure 5-16). The purpose of the 
evase is to reduce the air discharge velocity efficiently; thus, 
the available velocity pressure can be regained and credited 
to the exhaust system instead of being wasted. Practical 
considerations usually limit the construction of an evase to 
approximately a 10° angle (5° side angle) and a discharge 
velocity of about 2000 fpm (0.25 "wg velocity pressure) for 
normal exhaust systems. Further streamlining or lengthening 
the evase yields diminishing returns. 

It should be noted, however, that for optimum vertical 
dispersion of contaminated air, many designers feel that the 
discharge velocity from the stack should not be less than 
3000-3500 fpm. When these considerations prevail, the use 
of an evase is questionable. 

The following example indicates the application of the 
evase fitting. It is not necessary to locate the evase directly 
after the outlet of the fan. It should be noted that, depending 
upon the evase location, the static pressure at the fan discharge 
may be below atmospheric pressure, i.e., negative (-), as 
shown in this example. 

EXAMPLE 



Duct No. 


Dia. 


Q 


V 


VP 


SP 


1 Fan Inlet 


20 


8300 


3800 


0.90 


-7,27 


2 Fan Discharge = 




8300 


3715 


0.86 




16.5x19.5 












3 Round Duct Connection 


20 




3800 


0.90 




4 Evase Outlet 


28 




1940 


0.23 






To calculate the effect of the evase, see Figure 5-16 for 
expansion at the end of the duct where the Diameter Ratio, 
D4-D3 = 28-20 = 1 .4 and Taper length L/D - 40-20 = 2. 

R = 0.52 x 70% (since the evase is within 5 
diameters of the fan outlet) 



Exhaust System Design Procedure 5-33 



VP3 = 0.9 as given 

SP4 = (since the end of the duct is at atmospheric 
pressure) 

SP3 = SP 4 » R(VP 3 ) 

= 0-(0.52)(0.70)(0.90") 
= -0.33 "wg 

FSP = SPoutlet - SPjnlet - VPjnlet 

= -0.33 - (-7.27) - 0.9 = 6.04 "wg 

5.1 6 EXHAUST STACK OUTLETS 

The final component of the ventilation system is the ex- 
haust stack, an extension of the exhaust duct above the roof. 
There are two reasons for the placement of an exhaust stack 
on a ventilation system. First, the air exhausted by a local 
exhaust system should escape the building envelope. Second, 
once it has escaped the building envelope, the stack should 
provide sufficient dispersion so that the plume does not cause 
an unacceptable situation when it reaches the ground. This 
brief description of stack design will address only the first 
concern. 

When placing an exhaust stack on the roof of a building, 
the designer must consider several factors. The most impor- 
tant is the pattern of the air as it passes the building, Even in 
the case of a simple building design with a perpendicular 
wind, the air flow patterns over the building can be complex 
to analyze. Figure 5-28a shows the complex interaction be- 
tween the building and the wind at height H. A stagnation 
zone is formed on the upwind wall. Air flows away from the 
stagnation zone resulting in a down draft near the ground. 
Vortices are formed by the wind action resulting in a recircu- 
lation zone along the front of the roof or roof obstructions, 
down flow along the downwind side, and forward flow along 
the upwind side of the building. 

Figure 5-28b shows a schematic of the critical zones 
formed within the building cavity. A recirculation zone is 
formed at the leading edge of the building. A recirculation 
zone is an area where a relatively fixed amount of air moves 
in a circular fashion with little air movement through the 
boundary. A stack discharging into the recirculation zone can 
contaminate the zone. Consequently, all stacks should pene- 
trate the recirculation zone boundary. 

The high turbulence region is one through which the air 
passes; however, the flow is highly erratic with significant 
downward flow. A stack that discharges into this region will 
contaminate anything downwind of the stack. Consequently, 
all stacks should extend high enough that the resulting plume 
does not enter the high turbulence region upwind of an air 
intake. 

Because of the complex flow patterns around simple build- 
ings, it is almost impossible to locate a stack that is not 
influenced by vortices formed by the wind. Tall stacks are 
often used to reduce the influence of the turbulent flow, to 



release the exhaust air above the influence of the building and 
to prevent contamination of the air intakes. Selection of the 
proper location is made more difficult when the facility has 
several supply and exhaust systems and when adjacent build- 
ings or terrain cause turbulence around the facility itself. 

When locating the stack and outdoor air inlets for the air 
handling systems, it is often desirable to locate the intakes 
upwind of the source. However, often there is no true upwind 
position. The wind in all locations is variable. Even when 
there is a natural prevailing wind, the direction and speed are 
constantly changing. If stack design and location rely on the 
direction of the wind, the system will clearly fail. 

The effect of wind on stack height varies with speed: 

9 At very low wind speeds, the exhaust jet from a vertical 
stack will rise above the roof level resulting in 
significant dilution at the air intakes. 

• Increasing wind speed will decrease plume rise and 
consequently decrease dilution. 

e Increasing wind speed will increase turbulence and 
consequently increase dilution. 

The prediction of the location and form of the recirculation 
cavity, high turbulence region and roof wake is difficult. 
However, for wind perpendicular to a rectangular building, 
the height (H) and the width (W) of the upwind building face 
determine the airflow patterns. The critical dimensions are 
shown in Figure 5-28b. According to Wilson/ 5 6) the critical 
dimensions depend on a scaling coefficient (R) which is given 
by: 



R 



:B 0.67 xB 0.33 



[5.9] 



where B s is the smaller and B, is the larger of the dimensions 
H and W. When B! is larger than 8B S , use B, = 8 B s to calculate 
the scaling coefficient. For a building with a flat roof, Wil- 
son (5 7) estimated the maximum height (H c ), center (X c ), and 
lengths (L c ) of the recirculation region as follows: 



H c = 0.22 R 
X c = 0.5 R 
L c = 0.9 R 



[5.10] 
[5.11] 
[5.12] 



In addition, Wilson estimated the length of the building 
wake recirculation region by: 



L r =1.0R 



[5.13] 



The exhaust air from a stack often has not only an upward 
momentum due to the exit velocity of the exhaust air but 
buoyancy due to its density as well. For the evaluation of the 
stack height, the effective height is used (see Figure 5 -29a). 
The effective height is the sum of the actual stack height (H s ), 
the rise due to the vertical momentum of the air, and any wake 
downwash effect that may exist. A wake downwash occurs 
when air passing a stack forms a downwind vortex. The vortex 



5-34 



Industrial Ventilation 



will draw the plume down, reducing the effective stack height 
(see Figure 5-29b). This vortex effect is eliminated when the 
exit velocity is greater than 1 .5 times the wind velocity. If the 
exit velocity exceeds 3000 fpm, the momentum of the exhaust 
air reduces the potential downwash effect. 

The ideal design extends the stack high enough that the 
expanding plume does not meet the wake region boundary. 
More realistically, the stack is extended so that the expanding 
plume does not intersect the high turbulence region or any 
recirculation cavity. According to Wilson, < 5 - 6 > the high turbu- 
lence region boundary (Z 2 ) follows a 1:10 downward slope 
from the top of the recirculation cavity. 

To avoid entrapment of exhaust gas into the wake, stacks 
must terminate above the recirculation cavity. The effective 
stack height to avoid excessive reentry can be calculated by 
assuming that the exhaust plume spreads from the effective 
stack height with a slope of 1 :5 (see Figure 5-28b). The first 
step is to raise the effective stack height until the lower edge 
of the 1 :5 sloping plume avoids contact with all recirculation 
zone boundaries. The zones can be generated by rooftop 
obstacles such as air handling units, penthouses or architec- 
tural screens. The heights of the cavities are determined by 
Equations 5.10, 5. 1 1 and 5.12 using the scaling coefficient for 
the obstacle. Equation 5.13 can be used to determine the 
length of the wake recirculation zone downwind of the obsta- 
cle. 

If the air intakes, including windows and other openings, 
are located on the downwind wall, the lower edge of the plume 
with a downward slope of 1:5 should not intersect with the 
recirculation cavity downwind of the building. The length of 
the recirculation cavity (L r ) is given by Equation 5.13. If the 
air intakes are on the roof, the downward plume should not 
intersect the high turbulence region above the air intakes. 
When the intake is above the high turbulence boundary, 
extend a line from the top of the intake to the stack with a 
slope of 1:5. When the intake is below the high turbulence 
region boundary, extend a vertical line to the boundary, then 
extend back to the stack with a slope of 1:5. This allows the 
calculation of the necessary stack height. The minimum stack 
height can be determined for each air intake. The maximum 
of these heights would be the required stack height. 

In large buildings with many air intakes, the above proce- 
dure will result in very tall stacks. An alternate approach is to 
estimate the amount of dilution that is afforded by stack 
height, distance between the stack and the air intake and 
internal dilution that occurs within the system itself. This 
approach is presented in the "Airflow Around Buildings" 
chapter in the Fundamentals volume of the 1993 ASHRAE 
Handbook.™ 

5.16.1 Stack Considerations: 

l . Discharge velocity and gas temperature influence the 
effective stack height. 



2. Wind can cause a downwash into the wake of the stack 
reducing the effective stack height. Stack velocity 
should be at least 1 .5 times the wind velocity to prevent 
downwash. 

3. A good stack velocity is 3000 fpm because it: 

• Prevents downwash for winds up to 2000 fpm (22 
mph). Higher wind speeds have significant dilution 
effects. 

• Increases effective stack height. 

• Allows selection of a smaller centrifugal exhaust 
fan to provide a more stable operation point on the 
fan curve (see Chapter 6). 

• Provides conveying velocity if there is dust in the 
exhaust or there is a failure of the air cleaning 
device. 

4. High exit velocity is a poor substitute for stack height. 
For example, a flush stack requires a velocity over 
8000 fpm to penetrate the recirculation cavity bound- 
ary. 

5. The terminal velocity of rain is about 2000 fpm. A 
stack velocity above 2600 fpm will prevent rain from 
entering the stack when the fan is operating. 

6. Locate stacks on the highest roof of the building when 
possible. If not possible, a much higher stack is re- 
quired to extend beyond the wake of the high bay, 
penthouse, or other obstacle. 

7. The use of an architectural screen should be avoided. 
The screen becomes an obstacle and the stack must be 
raised to avoid the wake effect of the screen. 

8. The best stack shape is a straight cylinder. If a drain is 
required, a vertical stack head is preferred (see Figure 
5-30). In addition, the fan should be provided with a 
drain hole and the duct should be slightly sloped 
toward the fan. 

9. Rain caps should not be used. The rain cap directs the 
air toward the roof, increases the possibility of reentry, 
and causes exposures to maintenance personnel on the 
roof. Moreover, rain caps are not effective. A field 
study (59) with a properly installed standard rain cap 
showed poor performance. A 1 2-inch diameter stack 
passed 16% of all rain and as high as 45% during 
individual storms. 

10. Separating the exhaust points from the air intakes can 
reduce the effect of reentry by increasing dilution. 

11. In some circumstances, several small exhaust systems 
can be manifolded to a single exhaust duct to provide 
internal dilution thereby reducing reentry. 

12. A combined approach of vertical discharge, stack 
height, remote air intakes, proper air cleaning device, 
and internal dilution can be effective in reducing the 



Exhaust System Design Procedure 5-35 



consequences of reentry. 

13. A tall stack is not an adequate substitute for good 
emission control. The reduction achieved by properly 
designed air cleaning devices can have a significant 
impact on the potential for reentry. 

5.17 AIR BLEED-INS 

Bleed-ins are used at the ends of branch ducts to provide 
additional air flow rates to transport heavy material loads as 
in woodworking at saws and jointers or at the ends of a main 
duct to maintain minimum transport velocity when the system 
has been oversized deliberately to provide for future expan- 
sion. Some designers use bleed-ins also to introduce addi- 
tional air to an exhaust system to reduce air temperature and 
to assist in balancing the system. 

EXAMPLE 




End cap bleed-in (see sketch). Consider it to be an orifice 
or slot. From Figure 5- 13, h e - 1.78 VP. 

1 . Calculate SP for branch duct to junction (X). 

2. Determine flow rate in main duct according to design 
or future capacity or determine Q bleed-in directly 
from temperature or moisture considerations. 

3. Q bleed-in = (Q main duct) - (Q branch) 

4. SP bleed-in = SP branch as calculated - (h e + 1 VP) = 
(1.78 + 1.0) VP 

5. VP, bleed-in 



SP 



SP 



(1.78 + 1.0) 2.78 
6. Velocity, bleed-in from VP and Table 5-7a. 



7. Area bleed- in = 



Q bleed -in 
V bleed -in 



5.18 OPTIMUM ECONOMIC VELOCITY 

In systems which are intended to carry dust, a minimum 
conveying velocity is necessary to ensure that the dust will 
not settle in the duct. Also, when a system is installed in a 
quiet area, it may be necessary to keep velocities below some 
maximum to avoid excessive duct noise. When axial flow fans 
are used, duct velocities of 1000 to 1500 fpm are preferred. 
In a gas or vapor exhaust system installed in a typical factory 



environment where none of these restrictions apply, the ve- 
locity may be selected to yield the lowest annual operating 
cost. 

To determine the optimum economic velocity, the system 
must first be designed at any assumed velocity and the total 
initial costs of duct material, fabrication, and installation 
estimated/ 5 10 > 

This optimum economic velocity may range from under 
2000 fpm to over 4000 fpm. Lengthy expected service periods 
and system operating times tend to lower the optimum while 
high interest rates and duct costs tend to raise the optimum. 
In general, a velocity of 2500 to 3000 fpm will not result in 
equivalent total annual costs much in excess of the true 
optimum. 

5.19 CONSTRUCTION GUIDELINES FOR LOCAL 
EXHAUST SYSTEMS 

Ducts are specified most often for use in the low static 
pressure range (-10 "wg to +10 "wg), but higher static pres- 
sures are occasionally encountered. The duct conveys air or 
gas which is sometimes at high temperatures and often con- 
taminated with abrasive particulate or corrosive aerosols. 
Whether conditions are mild or severe, correct design and 
competent installation of ducts and hoods are necessary for 
proper functioning of any ventilation system. The following 
minimum specifications are recommended. 

Exhaust systems should be constructed with materials suit- 
able for the conditions of service and installed in a permanent 
and workman-like manner. To minimize friction loss and 
turbulence, the interior of all ducts should be smooth and free 
from obstructions — especially at joints. 

5.19.1 Materials: Ducts are constructed of black iron, 
which has been welded, flanged, and gasketed; or of welded 
galvanized sheet steel unless the presence of corrosive gases, 
vapors, and mists or other conditions make such material 
impractical. Arc welding of black iron lighter than 18 gauge 
is not recommended. Galvanized construction is not recom- 
mended for temperatures exceeding 400 F. The presence of 
corrosive gases, vapor, and mist may require the selection of 
corrosive resistant metals, plastics, or coatings. It is recom- 
mended that a specialist be consulted for the selection of 
materials best suited for applications when corrosive atmos- 
pheres are anticipated. Table 5-4 provides a guide for selec- 
tion of materials for corrosive conditions. 

5.19.2 Construction: 

1. There are four classifications for exhaust systems on 
noncorrosive applications: 

Class 1 — Light Duty: Includes no n abrasive applica- 
tions (e.g., replacement air, general ventilation, gase- 
ous emissions control). 

Class 2 — Medium Duty: Includes applications with 



5-36 



Industrial Ventilation 



moderately abrasive particulate in light concentrations 
(e.g., buffing and polishing, woodworking, grain 
dust). 

Class 3 — Heavy Duty: Includes applications with 
high abrasive in low concentrations (e.g., abrasive 
cleaning operations, dryers and kilns, boiler breeching, 
sand handling). 

Class 4 — Extra Heavy Duty: Includes applications 
with highly abrasive particles in high concentrations 
(e.g., materials conveying high concentrations of par- 
ticulate in all examples listed under Class 3 — usually 
used in heavy industrial plants such as steel mills, 
foundries, mining, and smelting). 

2. For most conditions, round duct is recommended for 
industrial ventilation, air pollution control, and dust 
collecting systems. Compared to non-round duct, it 
provides for lower friction loss, and its higher struc- 
tural integrity allows lighter gauge materials and fewer 
reinforcing members. Round duct should be con- 
structed in accordance with the Reference 5.1 1. Metal 
thickness required for round industrial duct varies with 
classification, static pressure, reinforcement, and span 
between supports. Metal thicknesses required for the 
four classes are based on design and use experience. 

3. Rectangular ducts should only be used when space 
requirements preclude the use of round construction. 
Rectangular ducts should be as nearly square as possi- 
ble to minimize resistance, and they should be con- 
structed in accordance with Reference 5.12. 

4. For many applications, spiral wound duct is adequate 
and less expensive than custom construction. How- 
ever, spiral wound duct should not be used for Classes 
3 and 4 because it does not withstand abrasion well. 
Elbows, branch entries, and similar fittings should be 
fabricated, if necessary, to achieve good design. Spe- 
cial considerations concerning use of spiral duct are as 
follows: 

A. Unless flanges are used for joints, the duct should 
be supported close to each joint, usually within 2 
in. Additional supports may be needed. See Refer- 
ence 5.11. 

B. Joints should be sealed by methods shown to be 
adequate for the service. 

C. Systems may be leak tested after installation at the 
maximum expected static pressure. The acceptable 

leakage criteria, often referred to as leakage class, 
should be carefully selected based on the hazards 
associated with the contaminant. 

5. The following formula (5 11} can be used for specifying 
ducts to be constructed of metals other than steel. For 
a duct of infinite length, the required thickness may be 
determined from: 



i-f 



(1-v 2 ) 
035714 p K } (52 + D) 



where: 



t = the thickness of the duct in inches 

D = the diameter of the duct in inches 

p = the intensity of the negative pressure on the 
duct in psi 

E = modulus of elasticity in psi 

v = Poisson's ratio 

The above equation for Class 1 ducts incorporates a 
safety factor which varies linearly with the diameter 
(D), beginning at 4 for small ducts and increasing to 8 
for duct diameters of 60 in. This safety factor has been 
adopted by the sheet metal industry to provide for lack 
of roundness; excesses in negative pressure due to 
particle accumulation in the duct and other manufac- 
turing or assembly imperfections unaccounted for by 
quality control; and tolerances provided by design 
specifications. 

Additional metal thickness must be considered for 
Classes 2, 3 and 4. The designer is urged to consult the 
Sheet Metal and Air Conditioning Contractors Na- 
tional Association (SMACNA) standards for complete 
engineering design procedures. 

6. Hoods should be a minimum of two gauges heavier 
than straight sections of connecting branches, free of 
sharp edges or burrs, and reinforced to provide neces- 
sary stiffness. 

7. Longitudinal joints or seams should be welded. All 
welding should conform to the standards established 
by the American Welding Society (AWS) structural 
code. (513) Double lock seams are limited to Class 1 
applications. 

8. Duct systems subject to wide temperature fluctuations 
should be provided with expansion joints. Flexible 
materials used in the construction of expansion joints 
should be selected with temperature and corrosion 
conditions considered. 

9. Elbows and bends should be a minimum of two gauges 
heavier than straight lengths of equal diameter and 
have a centerline radius of at least two and preferably 
two and one-half times the pipe diameter (see Figure 
5-21). Large centerline radius elbows are recom- 
mended where highly abrasive dusts are being con- 
veyed. 

10. Elbows of 90° should be of a five piece construction 
for round ducts up to 6 in. and of a seven piece 
construction for larger diameters. Bends less than 90° 
should have a proportional number of pieces. Prefab- 
ricated elbows of smooth construction may be used 



Exhaust System Design Procedure 5-37 



TABLE 5-4. Typical Physical and Chemical Properties of Fabricated Plastics and Other Materials 



Chemical Type 


Trade 
Names 


Max. Opr. 
Temp., F 


Flam- 
mability 








Resistance to 








Gasoline 


Mineral 
Oil 


Strong 
Alk. 


Weak 
Alk. 


Strong 
Acid 


Weak 
Acid 


Salt 
Solution 


Solvents 


Urea Formaldehyde 


Beetle 
Plaskon 

Sylplast 


170 


Self Ext 


Good 


Good 


Unac. 


Fair 


Poor 


Poor 


— 


Good 


Melamine 
Formaldehyde 


Cymel 

Plaskon 

Resimene 


210-300 


Self Ext. 


Good 


Good 


Poor 


Good 


Poor 


Good 


— 


Good 


Phenolic 


Bakelite 
Durite 
Durez G.E. 
Resinox 


250-450 


Self Ext. 


Fair 




Poor 


Fair 


Poor 


Fair 




Fair 


Alkyd 


Plaskon 


— 


Self Ext. 


Good 


_ 


Unac. 


Poor 


— 


Good 


— 


Good 


Silicone 


Bakelite GE 


550 


— 


Good 


Good 


— 


— 


Good 


Good 


— 


Unac. 


Epoxy 


Epiphem 
Araldite 
Maraset 
Renite 
Tool Plastik 
Epon Resin 


50-200 


Self Ext. 


Good 




Good 


Good 


Good 


Good 




Good 


Cast Phenolic 


Marblette 


__ 


Self Ext. 


— 


— 


Unac. 


Fair 


Good 


Good 


— 


Good to 
Unac. 


Ally! & Polyester 


Laminae 

Bakelite 

Plaskon 

Glykon 

Paraplex 


300^50 


Self Ext. 






Poor 


Fair 


Poor 


Fair 




Fair 


Acrylic 


Lucite 
Plexiglas 
Wasco line 


140-200 


0.5-2.0 
in/min 


— 


— - 


— 


Good 


Unac. 


Good 


— . 


Good to 
Unac. 


Polyethylene 


Tenite 
Irrathene 


140-200 


Slow 
Burning 


— 


— 


— 


— 


_ 


— 


— 


Unac. 


Tetrafluoroethylene 


Teflon 


500 


Non-FI. 


Good 


— 


Good 


Good 


Good 


Good 


— 


Good 


Chlorotriffuoroethylene 


KelF 






















Polyvinyl Formal & 
Butyral 


Vinylite 

Butacite 

Saflex 

Butvar 

Formuare 




Slow 
Burning 


Good 


Good 


Good 


Good 


Unac. 


Unac. 




Unac. 


Vinyl Chloride 
Polymer 
& Copolymer 


Krene 
Batelite Vinyl 
Dow pvc 
Vygen 


130-175 


Slow 
Burning 






Good 


Good 


Good 


Good 




Unac. 


Vinylidene Chloride 


Saran 


160-200 


Self Ext. 


Good 


Good 


Good 


Good 


Good 


Good 


— 


Fair 


Styrene 


Bakelite 

Catalin 

Styron 

Dylene 

Luxtrex 


150-165 


0.5-2.0 
in/min 


Unac. 


Fair 


Good 


Good 






Good 


Poor 


Polystyrene 
Reinforced with 
Fibrous Glass 

Cellulose Acetate 


Celanese 

Acetate 

Tenite 


Thermo 
Plastic 


0.5-2.0 
in/min 


Unac. 
Good 


Fair 
Good 


Good 
Unac. 


Good 
Unac. 


Unac. 


Fair 


Good 


Poor 
Poor 


Nylon 


Plaskon 

Zytel 

Tynex 


250 


Self Ext. 


Good 


Good 


Good 


Good 


Unac. 


Good 


— 


Good 


Glass 


Pyrex 


450 


Non-FI. 


Good 


Good 


Good 


Good 


Good 


Good 


Good 


Good 



NOTE: Each situation must be thoroughly checked for compatibility of materials during the design phase or if usage is changed. 



5-38 



Industrial Ventilation 



(see Figure 5-22 for heavy duty elbows). 

11. Where the air contaminant includes particulate that 
may settle in the ducts, clean-out doors should be 
provided in horizontal runs, near elbows, junctions, 
and vertical runs. The spacing of clean -out doors 
should not exceed 12 ft for ducts of 12 in. diameter and 
less but may be greater for larger duct sizes (see Figure 
5-23). Removable caps should be installed at all ter- 
minal ends, and the last branch connection should not 
be more than 6 in. from the capped end. 

12. Transitions in mains and sub-mains should be tapered. 
The taper should be at least five units long for each one 
unit change in diameter or 30° included angle (see 
Figure 5-25). 

.13. All branches should enter the main at the center of the 
transition at an angle not to exceed 45° with 30° pre- 
ferred. To minimize turbulence and possible particu- 
late fall out, connections should be to the top or side 
of the main with no two branches entering at opposite 
sides (see Figure 5-26). 

34. Where condensation may occur, the duct system 
should be liquid tight and provisions should be made 
for proper sloping and drainage. 

15. A straight duct section of at least six equivalent duct 
diameters should be used when connecting to a fan (see 
Figure 5-27). Elbows or other fittings at the fan inlet 
will seriously reduce the volume discharge (see Fig- 
ures 6-23, 6-24 and AMCA 20 1< 5 ,4 >). The diameter of 
the duct should be approximately equal to the fan inlet 
diameter. 

16. Discharge stacks should be vertical and terminate at a 
point where height or air velocity limit re-entry into 
supply air inlets or other plant openings (see Figures 
5-28 and 5-29). 

5.19.3 System Details: 

1 . Provide duct supports of sufficient capacity to carry 
the weight of the system plus the weight of the duct half 
filled with material and with no load placed on connect- 
ing equipment. [See SMACNA standards/ 5 - 11 - 5 - 12 )] 

2. Provide adequate clearance between ducts and ceil- 
ings, walls and floors for installation and maintenance. 

3. Install fire dampers, explosion vents, etc., in accord- 
ance with the National Fire Protection Association 
Codes and other applicable codes and standards. 

4. Avoid using blast gates or other dampers. However, if 
blast gates are used for system adjustment, place each 
in a vertical section midway between the hood and the 
junction. To reduce tampering, provide a means of 
locking dampers in place after the adjustments have 
been made. (See Figure 5-24 for types.) 



5. Allow for vibration and expansion. If no other consid- 
erations make it inadvisable, provide a flexible con- 
nection between the duct and the fan. The fan housing 
and drive motor should be mounted on a common base 
of sufficient weight to dampen vibration or on a prop- 
erly designed vibration isolator. 

6. Exhaust fans handling explosive or flammable atmos- 
pheres require special construction (see Section 6.3,9). 

7. Do not allow hoods and duct to be added to an existing 
exhaust system unless specifically provided for in the 
original design or unless the system is modified. 

8. Locate fans and filtration equipment such that mainte- 
nance access is easy. Provide adequate lighting in 
penthouses and mechanical rooms. 

5.19.4 Codes: Where federal, state, or local laws conflict 
with the preceding, the more stringent requirement should be 
followed. Deviation from existing regulations may require 
approval. 

5.19.5 Other Types of Duct Materials: 

1. Avoid use of flexible ducts. Where required, use a 
noncollapsible type that is no longer than necessary. 
Refer to the manufacturer's data for friction and bend 
losses. 

2. Commercially available seamless tubing for small 
duct sizes (i.e., up to 6 in.) may be more economical 
on an installed cost basis than other types. 

3 . Plastic pipe may be the best choice for some applica- 
tions (e.g., corrosive conditions at low temperature; 
see Table 5-4.) For higher temperatures, consider fi- 
berglass or a coated duct. 

4. Friction losses for non-fabricated duct will probably 
be different than shown in Tables 5-5 and 5-6. For 
specific information, consult manufacturer's data. 

5.19.6 Testing: The exhaust system should be tested and 
evaluated (see Chapter 9). Openings for sampling should be 
provided in the discharge stack or duct to test for compliance 
with air pollution codes or ordinances. 

REFERENCES 

5.1 Loeffler, J.J.: Simplified Equations for HVAC Duct 
Friction Factors. ASHRAE Journal, pp. 76-79 (Janu- 
ary 1980). 

5.2 Guffey, S.E.: Air-Flow Redistribution in Exhaust 
Ventilation Systems Using Dampers and Static Pres- 
sure Ratios. Appl, Occup. Environ. Hyg. 
8(3): 168-177 (March 1993). 

5.3 The Kirk and Blum Manufacturing Co.: Woodwork- 
ing Plants, p. W-9. Kirk and Blum, Cincinnati, OH. 

5.4 Air Movement and Control Association, Inc.: AMCA 



Exhaust System Design Procedure 5-39 



Standard 210-74. AMCA, Arlington Heights, IL. 

5.5 Constance, J.A.: Estimating Air Friction in Triangular 
Ducts. Air Conditioning, Heating and Ventilating 
60(6)85-86 (June 1963). 

5.6 Wilson, D.J.: Contamination of Air Intakes from Roof 
Exhaust Vents. ASHRAE Transactions 82:1024-38 
ASHRAE, Atlanta, GA (1976). 

5.7 Wilson, D.J.: Flow Patterns Over Flat Roof Buildings 
and Application to Exhaust Stack Design. ASHRAE 
Transactions 85:284-95. ASHRAE, Atlanta, GA 

(1979). 

5.8 American Society of Heating, Refrigerating and Air- 
Conditioning Engineers: 1993 ASHRAE Handbook, 
Fundamentals Volume 14:1-14,18. ASHRAE, At- 
lanta, GA (1993). 

5.9 Clark, John: The Design and Location of Building 
Inlets and Outlets to Minimize Wind Effect and Build- 
ing Reentry. Amer. Indus. Hyg. Assoc. J. 26:262 

(1956). 

5.10 Lynch, J.R.: Computer Design of Industrial Exhaust 
Systems. Heating, Piping and Air Conditioning (Sep- 
tember 1968). 

5.11 Sheet Metal and Air Conditioning Contractors' Na- 
tional Assoc, Inc.: Round Industrial Duct Construc- 
tion Standards. SMACNA, Vienna, VA (1977). 

5.12 Sheet Metal and Air Conditioning Contractors' Na- 



tional Assoc, Inc.: Rectangular Industrial Duct Con- 
struction Standards. SMACNA , Vienna, VA (1980). 

5.13 American Welding Society: AWS D 1.1 -72. AWS, 
Miami, FL. 

5.14 Air Movement & Control Associations, Inc.: AMCA 
Publication 201. AMCA, Arlington Heights, IL. 

5.15 Wright, Jr., D.K.: A New Friction Chart for Round 
Ducts. ASHVE Transactions, Vol. 51, p. 303 (1945). 

5.16 Clarke, J.H.: Air Flow Around Buildings. Heating, 
Piping and Air Conditioning. 39(5): 145-154 (May 
1967). 

5.17 American Society of Heating, Refrigerating and Air- 
Conditioning Engineers: 1989 ASHRAE Handbook, 
Fundamentals Volume, p. 32.31. ASHRAE, Atlanta, 
GA(1989). 

5.18 Brandt, A.D.: Industrial Health Engineering. John 
Wiley & Sons, New York (1947). 

5.19 American Society of Heating, Refrigerating and Air- 
Conditioning Engineers: Heating, Ventilating, Air 
Conditioning Guide, 37th ed. ASHRAE, Atlanta, GA 
(1959). 

5.20 American Society of Heating, Refrigerating and Air 
Conditioning Engineers: 1993 ASHRAE Handbook, 
Fundamentals Volume, Chapter 5. ASHRAE Atlanta, 
GA(1993). 



5-40 



Industrial Ventilation 



R > .2D 






h e = 0.93 VP d 

PLAIN DUCT END 



Orifice 



G.49 VP, 



FLANGED DUCT END 



h e - 0.04 VP d 

BELLMOUTH ENTRY 





45* taper angle 



— C 



Z2— 



h e = 1.5 VP d 



SHARP-EDGED 
ORIFICE 



h e = 0.4 VP d (tapered take-off) 
h e = 0.65 VP d (no taper) 

STANDARD GRINDER HOOD 



TRAP OR SETTLING CHAMBER 



Fl 



TAPERED HOODS 

anged or unflanged; round, square or 

rectangular. is the major angie on 

rectangular hoods. 



y€X 






ENTRY LOSS (h d ) 


e 


ROUND 


RECTANGULAR 


15° 


0.15 VP 


0.25 VP 


30 u 


0.08 VP 


0.16 VP 


45 u 


0.06 VP 


0.15 VP 


60 u 


0.08 VP 


0.17 VP 


90 u 


0.15 VP 


0.25 VP 


120° 


0.26 VP 


0.35 VP 


150" 


0.40 VP 


0.48 VP 


180 u 


0.50 VP 


0.50 VP 



Face area (A ) at least 2 times the duct area. 



VP - Duct VP - VP d 

Note: 180* values represent 
round ducts butted into 
back of booth or hood 
without a rectangular to 
round transition. 



1.10 
1.00 
0.90 
0.80 
0.70 
0.60 
0.50 
0.40 
0.30 
0.20 
0.10 
0.00 



























































Rectangular & Square 








Transition to Round --> 


































i4> 


/ 






















\\ 






























— Conical 




'"■■-..„. 




" 




(Ref! 5-14) 



20 40 60 80 100 120140160180 
6, INCLUDED ANGLE IN DEGREES 



COMPOUND HOODS 

A compound hood, such as the 
slot/plenum shown to the right, 
would have 2 losses, one through 
the slot and the other through 
the transition into the duct. 

The slot entry loss coefficient, F s , 
would have a vaiue typically in the 
range of 1.00 to 1.78 (see Chapters 
3 and 10). 

The duct entry loss coefficient is given 
by the above data for tapered hoods. 




h e = F S VP S +F d VP d 



MISCELLANEOUS VALUES 



HOOD 




ENTRY LOSS 
COEFFICIENT f 


Abrasive blast chamber 






1.0 


Abrasive blast elevator 






2.3 


Abrasive separator 






2.3 


Elevators (enclosures) 






0.69 


Flanged pipe plus close 


elbow 




0.8 


Plain pipe plus close elb 


ow 




1.60 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HOOD ENTRY 
LOSS COEFFICIENTS 



DATE 



1-95 



FIGURE 



5-13 



Exhaust System Design Procedure 5-41 






Stamped 
(Smooth) 



b — p i e 



4- piece 



piece 



Mitered 





R/D 




0.5 


0775 


1.00 


1.50 


2.00 


2.50 


S lamped 


0.71 


0.30 


0.22 


0.15 


1 3 


1 ?■ 


b- piece 




r 0.46 


6.55^\ 0.24 


0.10 


0.17* 


4- piece 


_. 


0.50 


0.37 


0.27 


0.24 


0.2 ~0 


3- piece 


0.90 


0.54 


0.42 


0.34 


0.33 


0.33* 



2 x t r o p o I a t e d f r o rr < p u b ! i s h e d ci o t a 



OTHER ELBOW LOSS COEFFICIENTS 
Mitered, no vones ■ .2 

Mitered, turning vanes 0.6 

Flatback (R/D*= 2.5) 0.05 



(see higure 5-23) 



NOTE: Loss 'actors are assumed to be for elbows of ''zero length 
included to I he intersection of centerlines. 



iction losses should be 



ROUND ELBOW LOSS COEFFICIENTS 



(Ref. 50 3) 




w -■ 



r /d 


Aspect Rcbo. W/D 


0.25 


0.5 | 1.0 


2.0 


3.0 


4.0 


O.O(Mitred) 


1.50 


1.32 | 1.15 


1 04 


0.92 


0.86 


0.5 


36 i 1.21 ! 005 


0.21 


. 8 4 

" 0.20" 


0.79 

a 19 ' 


00 


0.45 


28 J 0/1 


1.5 


0.28 


0.18 j 0.13 


0.13 


0.12 


04 2 


2.0 


0.24 j 0.15 1 0.11 


0.11 


0.10 


0.10 : 


3.0 


0.24 | 045 | 0.11 


041 


i ■ 


0.10 



SQUARE & RECTANGUFAR ELBOW LOSS COEFFICIENTS 



AMERICAN CONFERENCE 
OF GOVERNMENTAL 

INDUSTRIAL HYGIEN1STS 



DUCT DESIGN DATA 
ELBOW LOSSES 



DATE 



1 95 



FIGURE 



5-14 



5-42 



Industrial Ventilation 



T"" 




f 5 



1 m o x . 



I 



Note: Branch en my loss assumed to occur 
in branch and is so calculated. 

Do not include an enlargement regain 
calculation for branch entry enlargements. 



Angle 
Degrees 



10 



40 



4- 



TO 



50 
"60~ 



30 



Loss Fraction o( Vi- : 
in Branch 



0.06 



0.09 



0.21 



0.28 
0.02 



I.0C 



BRANCH ENTRY LOSSES 




H, No. of 

Diameters 



D 



■IER CAP LOSSES 



Loss Fraction of VP 



0.75 D 
70 D 


0.18 
0.22 


- 


0.65 D 


0.50 




0.60 D 


0FM 




0.55 D 


0.56 


0.50 D 


0.75 


0.45 D 


1.0 





See Fia. 5-29 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIEN ISTS 



DUCT DESIGN DAT. 



DA 



1 - 95 



FIGURE 



5-15 



Exhaust System Design Procedure 5-43 



Table 5-5 Tabulated Friction Loss Factors 



Galvanized Sheet Metal Duct 



Diameter 




Friction Loss, No. VP 


per foot 






inches 


1000 fpm 


2000 fpm 


3000 fpm 


4000 fpm 


5000 fpm 


6000 fpm 


0.5 


1 .0086 


0.9549 


0.9248 


0.9040 


0.8882 


0.8755 


1 


0.4318 


0.4088 


0.3959 


0.3870 


0.3802 


0.3748 


1.5 


0.2629 


0.2489 


0.2410 


0.2356 


0.2315 


0.2282 


2 


0.1848 


0.1750 


0.1695 


0.1657 


0.1628 


0.1605 


2.5 


0.1407 


0.1332 


0.1290 


0.1261 


0.1239 


0.1221 


3 


0.1125 


0.1065 


0.1032 


0.1009 


0.0991 


0.0977 


3.5 


0.0932 


0.0882 


0.0854 


0.0835 


0.0821 


0.0809 


4 


0.0791 


0,0749 


0.0726 


0.0709 


0.0697 


0.0687 


4.5 


0.0685 


0.0649 


0.0628 


0.0614 


0.0603 


0.0595 


5 


0.0602 


0.0570 


0.0552 


0.0540 


0.0530 


0.0523 


5.5 


0.0536 


0.0507 


0.0491 


0.0480 


0.0472 


0.0465 


6 


0.0482 


0.0456 


0.0442 


0.0432 


0.0424 


0.0418 


7 


0.0399 


0.0378 


0.0366 


0.0358 


0.0351 


0.0346 


8 


0.0339 


0.0321 


0.0311 


0.0304 


0.0298 


0.0294 


9 


0.0293 


0.0278 


0.0269 


0.0263 


0.0258 


0.0255 


10 


0.0258 


0.0244 


0.0236 


0.0231 


0.0227 


0.0224 


11 


0.0229 


0.0217 


0.0210 


0.0206 


0.0202 


0.0199 


12 


0.0206 


0.0195 


0,0189 


0.0185 


0.0182 


0.0179 


13 


0.0187 


0.0177 


0.0171 


0.0168 


0.0165 


0.0162 


14 


0.0171 


0.0162 


0.0157 


0.0153 


0.0150 


0.0148 


15 


0.0157 


0.0149 


0.0144 


0.0141 


0.0138 


0.0136 


16 


0.0145 


0.0137 


0.0133 


0.0130 


0.0128 


0.0126 


17 


0.0135 


0.0127 


0.0123 


0.0121 


0.0119 


0.0117 


18 


0.0126 


0.0119 


0.0115 


0.0113 


0.0111 


0.0109 


19 


0.0118 


0.0111 


0.0108 


0.0105 


0.0103 


0.0102 


20 


0.0110 


0.0104 


0.0101 


0.0099 


0.0097 


0.0096 


21 


0.0104 


0.0098 


0.0095 


0.0093 


0.0092 


0.0090 


22 


0.0098 


0.0093 


0.0090 


0.0088 


0.0086 


0.0085 


23 


0.0093 


0.0088 


0.0085 


0.0083 


0.0082 


0.0081 


24 


0.0088 


0.0084 


0.0081 


0.0079 


0.0078 


0.0077 


25 


0.0084 


0.0080 


0.0077 


0.0075 


0.0074 


0.0073 


26 


0.0080 


0.0076 


0.0073 


0.0072 


0.0070 


0.0069 


27 


0.0076 


0.0072 


0.0070 


0.0069 


0.0067 


0.0066 


28 


0.0073 


0.0069 


0.0067 


0.0066 


0.0064 


0.0063 


29 


0.0070 


0.0066 


0.0064 


0.0063 


0.0062 


0.0061 


30 


0.0067 


0.0064 


0.0062 


0.0060 


0.0059 


0.0058 


31 


0.0065 


0.0061 


0.0059 


0.0058 


0.0057 


0.0056 


32 


0.0062 


0.0059 


0.0057 


0.0056 


0.0055 


0.0054 


H = 


wO.533 

0.0307 Q0612 













5-44 



Industrial Ventilation 



Table 5-5 Tabulated Friction Loss Factors (cont'd) 



Galvanized Sheet Metal Duct 



Diameter 




Friction Loss, No. VP 


per foot 






inches 


1000 fpm 


2000 fpm 


3000 fpm 


4000 fpm 


5000 fpm 


6000 fpm 


33 


0.0060 


0.0057 


0.0055 


0.0054 


0.0053 


0.0052 


34 


0.0058 


0.0055 


0.0053 


0.0052 


0.0051 


0.0050 


35 


0.0056 


0.0053 


0.0051 


0.0050 


0.0049 


0.0048 


36 


0.0054 


0.0051 


0.0049 


0.0048 


0.0047 


0.0047 


37 


0.0052 


0.0049 


0.0048 


0.0047 


0.0046 


0.0045 


38 


0.0050 


0.0048 


0.0046 


0.0045 


0.0044 


0.0044 


39 


0.0049 


0.0046 


0.0045 


0.0044 


0.0043 


0.0042 


40 


0.0047 


0.0045 


0.0043 


0.0042 


0.0042 


0.0041 


41 


0.0046 


0.0043 


0.0042 


0.0041 


0.0040 


0.0040 


42 


0.0045 


0.0042 


0.0041 


0.0040 


0.0039 


0.0039 


43 


0.0043 


0.0041 


0.0040 


0.0039 


0.0038 


0.0038 


44 


0.0042 


0.0040 


0.0039 


0.0038 


0.0037 


0.0036 


45 


0.0041 


0.0039 


0.0038 


0.0037 


0.0036 


0.0036 


46 


0.0040 


0.0038 


0.0037 


0.0036 


0.0035 


0.0035 


47 


0.0039 


0.0037 


0.0036 


0.0035 


0.0034 


0.0034 


48 


0.0038 


0.0036 


0.0035 


0.0034 


0.0033 


0.0033 


49 


0.0037 


0.0035 


0.0034 


0.0033 


0.0032 


0.0032 


50 


0.0036 


0.0034 


0.0033 


0.0032 


0.0032 


0.0031 


52 


0.0034 


0.0032 


0.0031 


0.0031 


0.0030 


0.0030 


54 


0.0033 


0.0031 


0.0030 


0.0029 


0.0029 


0.0028 


56 


0.0031 


0.0030 


0.0029 


0.0028 


0.0028 


0.0027 


58 


0.0030 


0.0028 


0.0027 


0.0027 


0.0026 


0.0026 


60 


0.0029 


0.0027 


0.0026 


0.0026 


0.0025 


0.0025 


62 


0.0028 


0.0026 


0.0025 


0.0025 


0.0024 


0.0024 


64 


0.0027 


0.0025 


0.0024 


0.0024 


0.0023 


0.0023 


66 


0.0026 


0.0024 


0.0023 


0.0023 


0.0023 


0.0022 


68 


0.0025 


0.0023 


0.0023 


0.0022 


0.0022 


0.0021 


70 


0.0024 


0.0023 


0.0022 


0.0021 


0.0021 


0.0021 


72 


0.0023 


0.0022 


0.0021 


0.0021 


0.0020 


0.0020 


74 


0.0022 


0.0021 


0.0020 


0.0020 


0.0020 


0.0019 


76 


0.0022 


0.0020 


0.0020 


0.0019 


0.0019 


0.0019 


78 


0.0021 


0.0020 


0.0019 


0.0019 


0.0018 


0.0018 


80 


0.0020 


0.0019 


0.0019 


0.0018 


0.0018 


0.0018 


82 


0.0020 


0.0019 


0.0018 


0.0018 


0.0017 


0.0017 


84 


0.0019 


0.0018 


0.0017 


0.0017 


0.0017 


0.0017 


86 


0.0019 


0.0018 


0.0017 


0.0017 


0.0016 


0.0016 


88 


0.0018 


0.0017 


0.0017 


0.0016 


0.0016 


0.0016 


90 


0.0018 


0.0017 


0.0016 


0.0016 


0.0015 


0.0015 


Hr = 


W0.533 

°0307 Q0612 













Exhaust System Design Procedure 5-45 



Table 5-6 Tabulated Friction Loss Factors 



Black Iron, Aluminum, Stainless Steei, PVC Ducts 



Diameter Friction Loss, No. VP per foot 

inches 1000 fpm 2000 fpm 3000 fpm 4000 fpm 



5000 fpm 6000 fpm 



0.5 


0.8757 


0.7963 


0.7533 


0.7242 


0.7024 


0.6851 


1 


0.3801 


0.3457 


0.3270 


0.3143 


0.3049 


0.2974 


1.5 


0.2333 


0.2121 


0.2007 


0.1929 


0.1871 


0.1825 


2 


0,1650 


0.1500 


0.1419 


0.1364 


0.1323 


0.1291 


2.5 


0.1261 


0.1147 


0.1085 


0.1043 


0.1012 


0.0987 


3 


0.1013 


0.0921 


0.0871 


0.0837 


0.0812 


0.0792 


3.5 


0.0841 


0.0765 


0.0724 


0.0696 


0.0675 


0.0658 


4 


0.0716 


0.0651 


0.0616 


0.0592 


0.0574 


0.0560 


4.5 


0.0621 


0.0565 


0.0535 


0.0514 


0.0499 


0.0486 


5 


0.0547 


0.0498 


0.0471 


0.0453 


0.0439 


0.0428 


5.5 


0.0488 


0.0444 


0.0420 


0.0404 


0.0392 


0.0382 


6 


0.0440 


0.0400 


0.0378 


0.0364 


0.0353 


0.0344 


7 


0.0365 


0.0332 


0.0314 


0.0302 


0.0293 


0.0286 


8 


0.0311 


0.0283 


0.0267 


0.0257 


0.0249 


0.0243 


9 


0.0270 


0.0245 


0.0232 


0.0223 


0.0216 


0.0211 


10 


0.0238 


0.0216 


0.0204 


0.0197 


0.0191 


0.0186 


11 


0.0212 


0.0193 


0.0182 


0.0175 


0.0170 


0.0166 


12 


0.0191 


0.0174 


0.0164 


0.0158 


0.0153 


0.0149 


13 


0.0173 


0.0158 


0.0149 


0.0143 


0.0139 


0.0136 


14 


0.0158 


0.0144 


0.0136 


0.0131 


0.0127 


0.0124 


15 


0.0146 


0.0133 


0.0125 


0.0121 


0.0117 


0.0114 


16 


0.0135 


0.0123 


0.0116 


0.0112 


0.0108 


0.0106 


17 


0.0125 


0.0114 


0.0108 


0.0104 


0.0101 


0.0098 


18 


0.0117 


0.0106 


0.0101 


0.0097 


0.0094 


0.0092 


19 


0,0110 


0.0100 


0.0094 


0.0091 


0.0088 


0.0086 


20 


0.0103 


0.0094 


0.0089 


0.0085 


0.0083 


0.0081 


21 


0.0097 


0.0088 


0.0084 


0.0080 


0.0078 


0.0076 


22 


0.0092 


0.0084 


0.0079 


0.0076 


0.0074 


0.0072 


23 


0.0087 


0.0079 


0.0075 


0.0072 


0.0070 


0.0068 


24 


0.0083 


0.0075 


0.0071 


0.0068 


0.0066 


0.0065 


25 


0.0079 


0.0072 


0.0068 


0.0065 


0.0063 


0.0062 


26 


0.0075 


0.0068 


0.0065 


0.0062 


0.0060 


0.0059 


27 


0.0072 


0.0065 


0.0062 


0.0059 


0.0058 


0.0056 


28 


0.0069 


0.0063 


0.0059 


0.0057 


0.0055 


0.0054 


29 


0.0066 


0.0060 


0.0057 


0.0055 


0.0053 


0.0052 


30 


0.0063 


0.0058 


0.0054 


0.0052 


0.0051 


0.0050 


31 


0.0061 


0.0055 


0.0052 


0.0050 


0.0049 


0.0048 


32 


0.0059 

yO.465 

00425 Q0602 


0.0053 


0.0050 


0.0048 


0.0047 


0.0046 


Hr = 













5-46 



Industrial Ventilation 



Table 5-6 Tabulated Friction Loss Factors (cont'd) 

Black Iron, Aluminum, Stainless Steel, PVC Ducts 



Diameter 




Friction Loss, No. VP 


per foot 






inches 


1000 fpm 


2000 fpm 


3000 fpm 


4000 fpm 


5000 fpm 


6000 fpm 


33 


0.0056 


0.0051 


0.0049 


0.0047 


0.0045 


0.0044 


34 


0.0054 


0.0050 


0.0047 


0,0045 


0.0044 


0.0043 


35 


0.0053 


0.0048 


0.0045 


0.0043 


0.0042 


0.0041 


36 


0.0051 


0.0046 


0.0044 


0.0042 


0.0041 


0.0040 


37 


0.0049 


0.0045 


0.0042 


0.0041 


0.0039 


0.0038 


38 


0.0048 


0.0043 


0.0041 


0.0039 


0.0038 


0.0037 


39 


0.0046 


0.0042 


0.0040 


0.0038 


0.0037 


0.0036 


40 


0.0045 


0.0041 


0.0039 


0.0037 


0.0036 


0.0035 


41 


0.0043 


0.0040 


0.0037 


0.0036 


0.0035 


0.0034 


42 


0.0042 


0.0038 


0.0036 


0.0035 


0.0034 


0.0033 


43 


0.0041 


0.0037 


0.0035 


0.0034 


0.0033 


0.0032 


44 


0.0040 


0.0036 


0.0034 


0.0033 


0.0032 


0.0031 


45 


0.0039 


0.0035 


0.0033 


0.0032 


0.0031 


0.0030 


46 


0.0038 


0.0034 


0.0033 


0.0031 


0.0030 


0.0030 


47 


0.0037 


0.0034 


0.0032 


0.0030 


0.0030 


0.0029 


48 


0.0036 


0.0033 


0.0031 


0.0030 


0.0029 


0.0028 


49 


0.0035 


0.0032 


0.0030 


0.0029 


0.0028 


0.0027 


50 


0.0034 


0.0031 


0.0029 


0.0028 


0.0027 


0.0027 


52 


0.0033 


0.0030 


0.0028 


0.0027 


0.0026 


0.0026 


54 


0.0031 


0.0028 


0.0027 


0.0026 


0.0025 


0.0024 


56 


0.0030 


0.0027 


0.0026 


0.0025 


0.0024 


0.0023 


58 


0.0029 


0.0026 


0.0025 


0.0024 


0.0023 


0.0022 


60 


0.0027 


0.0025 


0.0024 


0.0023 


0.0022 


0.0021 


62 


0.0026 


0.0024 


0.0023 


0.0022 


0.0021 


0.0021 


64 


0.0025 


0.0023 


0.0022 


0.0021 


0.0020 


0.0020 


66 


0.0024 


0.0022 


0.0021 


0.0020 


0.0020 


0.0019 


68 


0.0024 


0.0021 


0.0020 


0.0020 


0.0019 


0.0018 


70 


0.0023 


0.0021 


0.0020 


0.0019 


0.0018 


0.0018 


72 


0.0022 


0.0020 


0.0019 


0.0018 


0.0018 


0.0017 


74 


0.0021 


0.0019 


0.0018 


0.0018 


0.0017 


0.0017 


76 


0.0021 


0.0019 


0.0018 


0.0017 


0.0017 


0.0016 


78 


0.0020 


0.0018 


0.0017 


0.0017 


0.0016 


0.0016 


80 


0.0019 


0.0018 


0.0017 


0.0016 


0.0016 


0.0015 


82 


0.0019 


0.0017 


0.0016 


0.0016 


0.0015 


0.0015 


84 


0.0018 


0.0017 


0.0016 


0.0015 


0.0015 


0.0014 


86 


0.0018 


0.0016 


0.0015 


0.0015 


0.0014 


0.0014 


88 


0.0017 


0.0016 


0.0015 


0.0014 


0.0014 


0.0014 


90 


0.0017 


0.0015 


0.0015 


0.0014 


0.0014 


0.0013 


U = f 


x/0.465 
) 0495 













Q' 



I.602 



Exhaust System Design Procedure 5-47 



STATIC PRESSURE REGAINS FOR EXPANSIONS 

j~~ 4 D min 



Within due 




At end of duct 



Reqain (R), fraction of VP difference 


Toper anqie 
degrees 


1.25:1 


Ddnmete 
1.5:1 


r ratios 
1.75:1 


[ Vd, 

2:1 


2.5:1 


3 1/2 


0.92 


0.88 


0.84 


0.81 


0.75 


' 5 


0.88 


0.84 


0,80 


0.76 


0.68 


10 


0.85 


0.76 


0.70 


0.63 


•"0.53 


15 


0.83 


0.70 


0,62 


0.55 


0,43 


20 


0.81 


67 


67 


0.48 


0.45 


25 


0.80 


0.65 


0.55 


0.44 


(. 


30 


79 


0.63 


0.51 


0.41 


0.25 


Abrupt 00 


0.77 


0.62 


0.50 


0.40 


0.25 


Where: SP ? = SR + R(VP, -VP 2 ) 



Regain (R), fraction of inlet VP 


Toper length 

to inlet diam 

L/D 


1.2: 1 


Diarne 
1.3: 1 


ter raid 
1.4:1 


)S D ?/ 
" 1.5:1 


'3 

1.6:1 


1.7: 1 


1.0:1 


0.37 


0.39 


0.38 


0.55 


0.31 


2 7 


1.5:1 


0.39 


0.46 


73 A 7 


0.46 


n.4; 


0.41 


2.0:1 


0.42 


0.49 


0.52 


0.52 


0.51 


0.-19 


3.0:1 


44 


0.52 


0.57 


i)0> 

._.: ..:. _ 


: J 60 


0.59 
OOOl" ' 

~6'.72~ 


4.0: 1 


45 


. 5 


O.bO 0.6 5 


"a66 


5.0:1 


47 


0.56 


52 i 0.5:5 


7.5:1 


48 


58 


0.64 0.68 | 0.70 


Where: SP, - SP, - R(VP, ) * 



*When SP 2 -0 (atmosphere) SP, will be (■■-) 

The regain (R) will only be 70% of volue shown above when expansion follows a disturbance or 
elbow (including a fan) by less than 5 duct diameters. 



STATIC PRESSURE LOSSES FOR CONTRACTIONS 



©■ 



v> 







--CD 



Fapered contraction ; 

5P 2 =SP, - (VP 2 -VP, )-L(VP 2 -VP, ) 



Abrupt contraction 

SP 2 = SP,-(VP 2 VP,) K(VP 2 ) 



Taper ongSe 
degrees 


L(loss) 


5 


0.05 


10 


0.06 


15 


0.08 


20 


0.10 


25 


0.11 


30 


0.15 


45 


0.20 


80 


0.30 


over 60 


Abrupt contraction 



Ratio Va, 


K 


0.1 


0.48 


0.2 


0.46 


0.3 


0.42 


0.4 


O A 7 

(252 


0.4 


0.6 


0.26 


0.7 


0.20 



A^ duct, area, 



Note 
In c 

SP 



:alcula ting SP for expansion or contraction use algebraic signs: VP is ( + ), and usually 
is (H-) in discharge duct from fan, and SP is (-) in inlet duct, to fori. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DUCT DESIGN 



TA 



DAI 



1-95 



DIG U RE 



5-16 



5-48 



Industrial Ventilation 



TABLE 5-7A. Velocity Pressure to Velocity Conversion — Standard Air 



FROM: V = 4005 V VP 



V = Velocity, fpm 
VP = Velocity Pressure, "wg 



VP 


V 


VP 


V 


VP 


V 


VP 


V 


VP 


V 


VP 


V 


0.01 


401 


0.51 


2860 


1.01 


4025 


1.51 


4921 


2.01 


5678 


2.60 


6458 


0.02 


566 


0.52 


2888 


1.02 


4045 


1.52 


4938 


2.02 


5692 


2.70 


6581 


0.03 


694 


0.53 


2916 


1.03 


4065 


1.53 


4954 


2.03 


5706 


2.80 


6702 


0.04 


801 


0.54 


2943 


1.04 


4084 


1.54 


4970 


2.04 


5720 


2.90 


6820 


0.05 


896 


0.55 


2970 


1.05 


4104 


1.55 


4986 


2.05 


5734 


3.00 


6937 


0.06 


981 


0.56 


2997 


1.06 


4123 


1.56 


5002 


2.06 


5748 


3.10 


7052 


0.07 


1060 


0.57 


3024 


1.07 


4143 


1.57 


5018 


2.07 


5762 


3 20 


7164 


0.08 


1133 


0.58 


3050 


1.08 


4162 


1.58 


5034 


2.08 


5776 


3.30 


7275 


0.09 


1201 


0.59 


3076 


1.09 


4181 


1.59 


5050 


2.09 


5790 


3.40 


7385 


0.10 


1266 


0.60 


3102 


1.10 


4200 


1.60 


5066 


2.10 


5804 


3.50 


7493 


0.11 


1328 


0.61 


3128 


1.11 


4220 


1.61 


5082 


2.11 


5818 


3.60 


7599 


0.12 


1387 


0.62 


3154 


1.12 


4238 


1.62 


5098 


2.12 


5831 


3.70 


7704 


0.13 


1444 


0.63 


3179 


1.13 


4257 


1.63 


5113 


2.13 


5845 


3.80 


7807 


0.14 


1499 


0.64 


3204 


1.14 


4276 


1.64 


5129 


2.14 


5859 


3.90 


7909 


0.15 


1551 


0.65 


3229 


1.15 


4295 


1.65 


5145 


2.15 


5872 


4.00 


8010 


0.16 


1602 


0.66 


3254 


1.16 


4314 


1.66 


5160 


2.16 


5886 


4.10 


8110 


0.17 


1651 


0.67 


3278 


1.17 


4332 


1.67 


5176 


2.17 


5900 


4.20 


8208 


0.18 


1699 


0.68 


3303 


1.18 


4351 


1.68 


5191 


2.18 


5913 


4.30 


8305 


0.19 


1746 


0.69 


3327 


1.19 


4369 


1.69 


5206 


2.19 


5927 


4.40 


8401 


0.20 


1791 


0.70 


3351 


1.20 


4387 


1.70 


5222 


2.20 


5940 


4.50 


8496 


0.21 


1835 


0.71 


3375 


1.21 


4405 


1.71 


5237 


2.21 


5954 


4.60 


8590 


0.22 


1879 


0.72 


3398 


1.22 


4424 


1.72 


5253 


2.22 


5967 


4.70 


8683 


0.23 


1921 


0.73 


3422 


1.23 


4442 


1.73 


5268 


2.23 


5981 


4.80 


8775 


0.24 


1962 


0.74 


3445 


1.24 


4460 


1.74 


5283 


2.24 


5994 


4.90 


8865 


0.25 


2003 


0.75 


3468 


1.25 


4478 


1.75 


5298 


2.25 


6007 


5.00 


8955 


0.26 


2042 


0.76 


3491 


1.26 


4496 


1.76 


5313 


2.26 


6021 


5.50 


9393 


0.27 


2081 


0.77 


3514 


1.27 


4513 


1.77 


5328 


2.27 


6034 


6.00 


9810 


0.28 


2119 


0.78 


3537 


1.28 


4531 


1.78 


5343 


2.28 


6047 


6.50 


10211 


0.29 


2157 


0.79 


3560 


1.29 


4549 


1.79 


5358 


2.29 


6061 


7.00 


10596 


0.30 


2194 


0.80 


3582 


1.30 


4566 


1.80 


5373 


2.30 


6074 


7.50 


10968 


0.31 


2230 


0.81 


3604 


1.31 


4584 


1.81 


5388 


2.31 


6087 


8.00 


11328 


0.32 


2266 


0.82 


3627 


1.32 


4601 


1.82 


5403 


2.32 


6100 


8.50 


11676 


0.33 


2301 


0.83 


3649 


1.33 


4619 


1.83 


5418 


2.33 


6113 


9.00 


12015 


0.34 


2335 


0.84 


3671 


1.34 


4636 


1.84 


5433 


2.34 


6126 


9.50 


12344 


0.35 


2369 


0.85 


3692 


1.35 


4653 


1.85 


5447 


2.35 


6140 


10.00 


12665 


0.36 


2403 


0.86 


3714 


1.36 


4671 


1.86 


5462 


2.36 


6153 


10.50 


12978 


0.37 


2436 


0.87 


3736 


1.37 


4688 


1.87 


5477 


2.37 


6166 


11.00 


13283 


0.38 


2469 


0.88 


3757 


1.38 


4705 


1.88 


5491 


2.38 


6179 


11.50 


13582 


0.39 


2501 


0.89 


3778 


1.39 


4722 


1.89 


5506 


2.39 


6192 


12.00 


13874 


0.40 


2533 


0.90 


3799 


1.40 


4739 


1.90 


5521 


2.40 


6205 


12.50 


14160 


0.41 


2564 


0.91 


3821 


1.41 


4756 


1.91 


5535 


2.41 


6217 


13.00 


14440 


0.42 


2596 


0.92 


3841 


1.42 


4773 


1.92 


5549 


2.42 


6230 


13.50 


14715 


0.43 


2626 


0.93 


3862 


1.43 


4789 


1.93 


5564 


2.43 


6243 


14.00 


14985 


0.44 


2657 


0.94 


3883 


1.44 


4806 


1.94 


5578 


2.44 


6256 


14.50 


15251 


0.45 


2687 


0.95 


3904 


1.45 


4823 


1.95 


5593 


2.45 


6269 


15.00 


15511 


0.46 


2716 


0.96 


3924 


1.46 


4839 


1.96 


5607 


2.46 


6282 


15.50 


15768 


0.47 


2746 


0.97 


3944 


1.47 


4856 


1.97 


5621 


2.47 


6294 


16.00 


16020 


0.48 


2775 


0.98 


3965 


1.48 


4872 


1.98 


5636 


2.48 


6307 


16.50 


16268 


0.49 


2803 


0.99 


3985 


1.49 


4889 


1.99 


5650 


2.49 


6320 


17.00 


16513 


0.50 


2832 


1.00 


4005 


1.50 


4905 


2.00 


5664 


2.50 


6332 


17.50 


16754 



Exhaust System Design Procedure 5-49 



TABLE 5-7B. Velocity to Velocity Pressure Conversion — Standard Air 



FROSfi: V - 4005 V VP 



V = Velocity, Spm 
VP = Velocity Pressure, 



m 



V 


VP 


V 


VP 


V 


VP 


V 


VP 


V 


VP 


V 


VP 


400 


0.01 


2600 


0.42 


3850 


0.92 


4880 


1.48 


5690 


2.02 


6190 


2.39 


500 


0.02 


2625 


0.43 


3875 


0.94 


4900 


1.50 


5700 


2.03 


6200 


2.40 


600 


0.02 


2650 


0.44 


3900 


0.95 


4920 


1.51 


5710 


2.03 


6210 


2.40 


700 


0.03 


2675 


0.45 


3925 


0.96 


4940 


1.52 


5720 


2.04 


6220 


2.41 


800 


0.04 


2700 


0.45 


3950 


0.97 


4960 


1.53 


5730 


2.05 


6230 


2.42 


900 


0.05 


2725 


0.46 


3975 


0.99 


4980 


1.55 


5740 


2.05 


6240 


2.43 


1000 


0.06 


2750 


0.47 


4000 


1.00 


5000 


1.56 


5750 


2.06 


6250 


2.44 


1100 


0.08 


2775 


0.48 


4020 


1.01 


5020 


1.57 


5760 


2.07 


6260 


2.44 


1200 


0.09 


2800 


0.49 


4040 


1.02 


5040 


1.58 


5770 


2.08 


6270 


2.45 


1300 


0.11 


2825 


0.50 


4060 


1.03 


5060 


1.60 


5780 


2.08 


6280 


2.46 


1400 


0.12 


2850 


0.51 


4080 


1.04 


5080 


1.61 


5790 


2.09 


6290 


2.47 


1450 


0.13 


2875 


0.52 


4100 


1.05 


5100 


1.62 


5800 


2.10 


6300 


2.47 


1500 


0.14 


2900 


0.52 


4120 


1.06 


5120 


1.63 


5810 


2.10 


6310 


2.48 


1550 


0.15 


2925 


0.53 


4140 


1.07 


5140 


1.65 


5820 


2.11 


6320 


2.49 


1600 


0.16 


2950 


0.54 


4160 


1.08 


5160 


1.66 


5830 


2.12 


6330 


2.50 


1650 


0.17 


2975 


0.55 


4180 


1.09 


5180 


1.67 


5840 


2.13 


6340 


2.51 


1700 


0.18 


3000 


0.56 


4200 


1.10 


5200 


1.69 


5850 


2.13 


6350 


2.51 


1750 


0.19 


3025 


0.57 


4220 


1.11 


5220 


1.70 


5860 


2.14 


6360 


2.52 


1800 


0.20 


3050 


0.58 


4240 


1.12 


5240 


1.71 


5870 


2.15 


6370 


2.53 


1825 


0.21 


3075 


0.59 


4260 


1.13 


5260 


1.72 


5880 


2.16 


6380 


2.54 


1850 


0.21 


3100 


0.60 


4280 


1.14 


5280 


1.74 


5890 


2.16 


6390 


2.55 


1875 


0.22 


3125 


0.61 


4300 


1.15 


5300 


1.75 


5900 


2.17 


6400 


2.55 


1900 


0.23 


3150 


0.62 


4320 


1.16 


5320 


1.76 


5910 


2.18 


6410 


2.56 


1925 


0.23 


3175 


0.63 


4340 


1.17 


5340 


1.78 


5920 


2.18 


6420 


2.57 


1950 


0.24 


3200 


0.64 


4360 


1.19 


5360 


1.79 


5930 


2.19 


6430 


2.58 


1975 


0.24 


3225 


0.65 


4380 


1.20 


5380 


1.80 


5940 


2.20 


6440 


2.59 


2000 


0.25 


3250 


0.66 


4400 


1.21 


5400 


1.82 


5950 


2.21 


6450 


2.59 


2025 


0.26 


3275 


0.67 


4420 


1.22 


5420 


1.83 


5960 


2.21 


6460 


2.60 


2050 


0.26 


3300 


0.68 


4440 


1.23 


5440 


1.84 


5970 


2.22 


6470 


2.61 


2075 


0.27 


3325 


0.69 


4460 


1.24 


5460 


1.86 


5980 


2.23 


6480 


2.62 


2100 


0.27 


3350 


0.70 


4480 


1.25 


5480 


1.87 


5990 


2.24 


6490 


2.63 


2125 


0.28 


3375 


0.71 


4500 


1.26 


5500 


1.89 


6000 


2.24 


6500 


2.63 


2150 


0.29 


3400 


0.72 


4520 


1.27 


5510 


1.89 


6010 


2.25 


6550 


2.67 


2175 


0.29 


3425 


0.73 


4540 


1.29 


5520 


1.90 


6020 


2.26 


6600 


2.72 


2200 


0.30 


3450 


0.74 


4560 


1.30 


5530 


1.91 


6030 


2.27 


6650 


2.76 


2225 


0.31 


3475 


0.75 


4580 


1.31 


5540 


1.91 


6040 


2.27 


6700 


2.80 


2250 


0.32 


3500 


0.76 


4600 


1.32 


5550 


1.92 


6050 


2.28 


6750 


2.84 


2275 


0.32 


3525 


0.77 


4620 


1.33 


5560 


1.93 


6060 


2.29 


6800 


2.88 


2300 


0.33 


3550 


0.79 


4640 


1.34 


5570 


1.93 


6070 


2.30 


6900 


2.97 


2325 


0.34 


3575 


0.80 


4660 


1.35 


5580 


1.94 


6080 


2.30 


7000 


3.05 


2350 


0.34 


3600 


0.81 


4680 


1.37 


5590 


1.95 


6090 


2.31 


7100 


3.14 


2375 


0.35 


3625 


0.82 


4700 


1.38 


5600 


1.96 


6100 


2.32 


7200 


3.23 


2400 


0.36 


3650 


0.83 


4720 


1.39 


5610 


1.96 


6110 


2.33 


7300 


3.32 


2425 


0.37 


3675 


0.84 


4740 


1.40 


5620 


1.97 


6120 


2.34 


7400 


3.41 


2450 


0.37 


3700 


0.85 


4760 


1.41 


5630 


1.98 


6130 


2.34 


7500 


3.51 


2475 


0.38 


3725 


0.87 


4780 


1.42 


5640 


1.98 


6140 


2.35 


7600 


3.60 


2500 


0.39 


3750 


0.88 


48(H) 


1.44 


5650 


1.99 


6150 


2.36 


7700 


3.70 


2525 


0.40 


3775 


0.89 


4820 


1.45 


5660 


2.00 


6160 


2.37 


7800 


3.79 


2550 


0.41 


3800 


0.90 


4840 


1.46 


5670 


2.00 


6170 


2.37 


7900 


3.89 


2575 


0.41 


3825 


0.91 


4860 


1.47 


5680 


2.01 


6180 


2.38 


8000 


3.99 



5-50 



Industrial Ventilation 



TABLE 5-8. Area and Circumference of Circles 



Oiam. 


AREA 


CIRCUMFERENCE 


Oiam. 
in 


AREA 


ciRcuma 


FERENCE 


in 


Square 


Square 


Square 


Square 




Inches 


Inches 


Feet 


Inches 


Feet 


Snches 


Inches 


Feet 


Inches 


Feet 


1 


0.79 


0.0055 


3.14 


0.2618 


30 


706.9 


4.909 


94.2 


7.854 


1.5 


1.77 


0.0123 


4.71 


0.3927 


31 


754.8 


5.241 


97.4 


8.116 


2 


3.14 


0.0218 


6.28 


0.5236 


32 


804.2 


5.585 


100.5 


8.378 


2.5 


4.91 


0.0341 


7.85 


0.6545 


33 


855.3 


5.940 


103.7 


8.639 


3 


7.07 


0.0491 


9.42 


0.7854 


34 


907.9 


6.305 


106.8 


8.901 


3.5 


9.62 


0.0668 


11.00 


0.9163 


35 


962.1 


6.681 


110.0 


9.163 


4 


12.57 


0.0873 


12.57 


1.0472 


36 


1017.9 


7.069 


113.1 


9.425 


4.5 


15.90 


0.1104 


14.14 


1.1781 


37 


1075.2 


7.467 


116.2 


9.687 


5 


19.63 


0.1364 


15.71 


1.3090 


38 


1134.1 


7.876 


119.4 


9.948 


5.5 


23.76 


0.1650 


17,28 


1.4399 


39 


1194.6 


8.296 


122.5 


10.210 


6 


28.27 


0.1963 


18.85 


1.5708 


40 


1256.6 


8.727 


125.7 


10.472 


6.5 


33.18 


0.2304 


20.42 


1.7017 


41 


1320.3 


9.168 


128.8 


10.734 


7 


38.48 


0.2673 


21.99 


1.8326 


42 


1385.4 


9.621 


131.9 


10.996 


7.5 


44.18 


0.3068 


23.56 


1.9635 


43 


1452.2 


10.085 


135.1 


11.257 


8 


50.27 


0.3491 


25.13 


2.0944 


44 


1520.5 


10.559 


138.2 


11.519 


8.5 


56.75 


0.3941 


26.70 


2.2253 


45 


1590.4 


11.045 


141.4 


11.781 


9 


63.62 


0.4418 


28.27 


2.3562 


46 


1661.9 


11.541 


144.5 


12.043 


9.5 


70.88 


0.4922 


29.85 


2.4871 


47 


1734.9 


12.048 


147.7 


12.305 


10 


78.54 


0.5454 


31.42 


2.6180 


48 


1809.6 


12.566 


150.8 


12.566 


10.5 


86.59 


0.6013 


32.99 


2.7489 


49 


1885.7 


13.095 


153.9 


12.828 


11 


95.03 


0.6600 


34.56 


2.8798 


50 


1963.5 


13.635 


157.1 


13.090 


11.5 


103.87 


0.7213 


36.13 


3.0107 


52 


2123.7 


14.748 


163.4 


13.614 


12 


113.10 


0.7854 


37.70 


3.1416 


54 


2290.2 


15.904 


169.6 


14.137 


13 


132.73 


0.9218 


40.84 


3.4034 


56 


2463.0 


17.104 


175.9 


14.661 


14 


153.94 


1.0690 


43.98 


3.6652 


58 


2642.1 


18.348 


182.2 


15.184 


15 


176.71 


1.2272 


47.12 


3.9270 


60 


2827.4 


19.635 


188.5 


15.708 


16 


201.06 


1.3963 


50.27 


4.1888 


62 


3019.1 


20.966 


194.8 


16.232 


17 


226.98 


1.5763 


53.41 


4.4506 


64 


3217.0 


22.340 


201.1 


16.755 


18 


254.47 


1.7671 


56.55 


4.7124 


66 


3421.2 


23.758 


207.3 


17.279 


19 


283.53 


1.9689 


59.69 


4.9742 


68 


3631.7 


25.220 


213.6 


17.802 


20 


314.16 


2.1817 


62.83 


5.2360 


70 


3848.5 


26.725 


219.9 


18.326 


21 


346.36 


2.4053 


65.97 


5.4978 


72 


4071.5 


28.274 


226.2 


18.850 


22 


380.13 


2.6398 


69.12 


5.7596 


74 


4300.8 


29.867 


232.5 


19.373 


23 


415.48 


2.8852 


72.26 


6.0214 


76 


4536.5 


31.503 


238.8 


19.897 


24 


452.39 


3.1416 


75.40 


6.2832 


78 


4778.4 


33.183 


245.0 


20.420 


25 


490.87 


3.4088 


78.54 


6.5450 


80 


5026.5 


34.907 


251.3 


20.944 


26 


530.93 


3.6870 


81.68 


6.8068 


82 


5281.0 


36.674 


257.6 


21 .468 


27 


572.56 


3.9761 


84.82 


7.0686 


84 


5541.8 


38.485 


263.9 


21.991 


28 


615.75 


4.2761 


87.96 


7.3304 


86 


5808.8 


40.339 


270.2 


22.515 


29 


660.52 


4.5869 


91.11 


7.5922 


88 


6082.1 


42.237 


276.5 


23.038 



The usual sheet metal fabricator will have patterns for ducts in 0.5-inch steps through 5.5-inch diameter; 1 inch steps 6 inches through 20 inches 
and 2-inch steps 22 inches and larger diameters. 



Exhaust System Design Procedure 5-51 

TABLE 5-9. Circular Equivalents of Rectangular Duel Sizes 



a\' 


4.0 4.5 5.0 5.5 0.0 6.5 7.0 7.5 0.0 8.5 9.0 9.5 10.0 10.5 11.0 11.5 12.0 12.5 13.0 13.5 14.0 14.5 15.0 15.5 16.0 


3.0 


3.8 4.0 4.2 4.4 4.6 4.7 4.9 5.1 5.2 5.3 5.5 5.6 5.7 5.9 6.0 6.1 6.2 6.3 6.4 6.5 6.6 6.7 6.8 6.9 7.0 


3.5 


4.1 4.3 4.6 4.8 5.0 5.2 5.3 5.5 5.7 5.8 6.0 6.1 6.3 6.4 6.5 6.7 6.8 6.9 7.0 7.1 7.2 7.3 7.5 7.6 7.7 


4.0 


4.4 4.6 4.9 5.1 5.3 5.5 5.7 5.9 6.1 6.3 6.4 6.6 6.7 6.9 7.0 7.2 7.3 7.4 7.6 7.7 7.8 7.9 8.0 8.2 8.3 


4.5 


4.6 4.9 5.2 5.4 5.7 5.9 6.1 6.3 6.5 6.7 6.9 7.0 7.2 7.4 7.5 7.7 7.8 7.9 8.1 8.2 8.4 8.5 8.6 8.7 8.8 


5.0 


4.9 5.2 5.5 5.7 6.0 6.2 6.4 6.7 6.9 7.1 7.3 7.4 7.6 7.8 8.0 8.1 8.3 8.4 8.6 8.7 8.9 9.0 9.1 9.3 9.4 


5.5 


5.1 5.4 5.7 6.0 6.3 6.5 6.8 7.0 7.2 7.4 7.6 7.8 8.0 8.2 8.4 8.6 8.7 8.9 9.0 9.2 9.3 9.5 9.6 9.8 9.9 


.V 


6.0 7.0 8.0 9.0 10.0 11.0 12.0 13.0 14.0 15.0 16.0 17.0 18.0 19.0 20.0 22.0 24.0 26.0 28.0 30.0 32.0 34.0 36.0 38.0 40.0 


6.0 
7.0 
8.0 


6.6 

7.1 7.7 
7.6 8.2 8.7 


9.0 


8.0 8.7 9.3 9.8 


10.0 


8.4 9.1 9.8 10.4 10.9 


11.0 


8.8 9.5 10.2 10.9 11.5 12.0 


12.0 


9.1 9.9 10.7 11.3 12.0 12.6 13.1 


13.0 


9.5 10.3 11.1 11.8 12.4 13.1 13.7 14.2 


14.0 


9.8 10.7 11.5 12.2 12.9 13.5 14.2 14.7 15.3 


15.0 


10.1 11.0 11.8 12.6 13.3 14.0 14.6 15.3 15.8 16.4 


16.0 


10.4 11.3 12.2 13.0 13.7 14.4 15.1 15.7 16.4 16.9 17.5 


17.0 


10.7 11.6 12.5 13.4 14.1 14.9 15.6 16.2 16.8 17.4 18.0 18.6 


18.0 


11.0 11.9 12.9 13.7 14.5 15.3 16.0 16.7 17.3 17.9 18.5 19.1 19.7 


19.0 


11.2 12.2 13.2 14.1 14.9 15.7 16.4 17.1 17.8 18.4 19.0 19.6 20.2 20.8 


20.0 


11.5 12.5 13.5 14.4 15.2 16.0 16.8 17.5 18.2 18.9 19.5 20.1 20.7 21.3 21.9 


22.0 


12.0 13.0 14.1 15.0 15.9 16.8 17.6 18.3 19.1 19.8 20.4 21.1 21.7 22.3 22.9 24.0 


24.0 


12.4 13.5 14.6 15.6 16.5 17.4 18.3 19.1 19.9 20.6 21.3 22.0 22.7 23.3 23.9 25.1 26.2 


26.0 


12.8 14.0 15.1 16.2 17.1 18.1 19.0 19.8 20.6 21.4 22.1 22.9 23.5 24.2 24.9 26.1 27.3 28.4 


28.0 


13.2 14.5 15.6 16.7 17.7 18.7 19.6 20.5 21.3 22.1 22.9 23.7 24.4 25.1 25.8 27.1 28.3 29.5 30.6 


30.0 


13.6 14.9 16.1 17.2 18.3 19.3 20.2 21.1 22.0 22.9 23.7 24.4 25.2 25.9 26.6 28.0 29.3 30.5 31.7 32.8 


32.0 


14.0 15.3 16.5 17.7 18.8 19.8 20.8 21.8 22.7 23.5 24.4 25.2 26.0 26.7 27.5 28.9 30.2 31.5 32.7 33.9 35.0 


34.0 


14.4 15.7 17.0 18.2 19.3 20.4 21.4 22.4 23.3 24.2 25.1 25.9 26.7 27.5 28.3 29.7 31.1 32.4 33.7 34.9 36.1 37.2 


36.0 


14.7 16.1 17.4 18.6 19.8 20.9 21.9 22.9 23.9 24.8 25.7 26.6 27.4 28.2 29.0 30.5 32.0 33.3 34.6 35.9 37.1 38.2 39.4 


38.0 


15.0 16.5 17.8 19.0 20.2 21.4 22.4 23.5 24.5 25.4 26.4 27.2 28.1 28.9 29.8 31.3 32.8 34.2 35.6 36.8 38.1 39.3 40.4 41.5 


40.0 


15.3 16.8 18.2 19.5 20.7 21.8 22.9 24.0 25.0 26.0 27.0 27.9 28.8 29.6 30.5 32.1 33.6 35.1 36.4 37.8 39.0 40.3 41.5 42.6 43.7 


42.0 


15.6 17.1 18.5 19.9 21.1 22.3 23.4 24.5 25.6 26.6 27.6 28.5 29.4 30.3 31.2 32.8 34.4 35.9 37.3 38.7 40.0 41.3 42.5 43.7 44.8 


44.0 


15.9 17.5 18.9 20.3 21.5 22.7 23.9 25.0 26.1 27.1 28.1 29.1 30.0 30.9 31.8 33.5 35.1 36.7 38.1 39.5 40.9 42.2 43.5 44.7 45.8 


46.0 


16.2 17.8 19.3 20.6 21.9 23.2 24.4 25.5 26.6 27.7 28.7 29.7 30.6 31.6 32.5 34.2 35.9 37.4 38.9 40.4 41.8 43.1 44.4 45.7 46.9 


48.0 


16.5 18.1 19.6 21.0 22.3 23.6 24.8 26.0 27.1 28.2 29.2 30.2 31.2 32.2 33.1 34.9 36.6 38.2 39.7 41.2 42.6 44.0 45.3 46.6 47.9 


50.0 


16.8 18.4 19.9 21.4 22.7 24.0 25.2 26.4 27.6 28.7 29.8 30.8 31.8 32.8 33.7 35.5 37.2 38.9 40.5 42.0 43.5 44.9 46.2 47.5 48.8 


54.0 


17.3 19.0 20.6 22.0 23.5 24.8 26.1 27.3 28.5 29.7 30.8 31.8 32.9 33.9 34.9 36.8 38.6 40.3 41.9 43.5 45.1 46.5 48.0 49.3 50.7 


58.0 


17.8 19.5 21.2 22.7 24.2 25.5 26.9 28.2 29.4 30.6 31.7 32.8 33.9 35.0 36.0 38.0 39.8 41.6 43.3 45.0 46.6 48.1 49.6 51.0 52.4 


62.0 


18.3 20.1 21.7 23.3 24.8 26.3 27.6 28.9 30.2 31.5 32.6 33.8 34.9 36.0 37.1 39.1 41.0 42.9 44.7 46.4 48.0 49.6 51.2 52.7 54.1 


66.0 


18.8 20.6 22.3 23.9 25.5 26.9 28.4 29.7 31.0 32.3 33.5 34.7 35.9 37.0 38.1 40.2 42.2 44.1 46.0 47.7 49.4 51.1 52.7 54.2 55.7 


70.0 


19.2 21.1 22.8 24.5 26.1 27.6 29.1 30.4 31.8 33.1 34.4 35.6 36.8 37.9 39.1 41.2 43.3 45.3 47.2 49.0 50.8 52.5 54.1 55.7 57.3 


74.0 


19.6 21.5 23.3 25.1 26.7 28.2 29.7 31.2 32.5 33.9 35.2 36.4 37.7 38.8 40.0 42.2 44.4 46.4 48.4 50.3 52.1 53.8 55.5 57.2 58.8 


78.0 


20.0 22.0 23.8 25.6 27.3 28.8 30.4 31.8 33.3 34.6 36.0 37.2 38.5 39.7 40.9 43.2 45.4 47.5 49.5 51.4 53.3 55.1 56.9 58.6 60.2 


82.0 


20.4 22.4 24.3 26.1 27.8 29.4 31.0 32.5 33.9 35.4 36.7 38.0 39.3 40.6 41.8 44.1 46.4 48.5 50.6 52.6 54.5 56.4 58.2 59.9 61.6 


86.0 


20.8 22.9 24.8 26.6 28.3 30.0 31.6 33.1 34.6 36.1 37.4 38.8 40.1 41.4 42.6 45.0 47.3 49.6 51.7 53.7 55.7 57.6 59.4 61.2 63.0 


90.0 


21.2 23.3 25.2 27.1 28.9 30.6 32.2 33.8 35.3 36.7 38.2 39.5 40.9 42.2 43.5 45.9 48.3 50.5 52.7 54.8 56.8 58.8 60.7 62.5 64.3 



5-52 Industrial Ventilation 



TABLE 5-9. Circular Equivalents of Rectangular Duct Sizes (cont.) 



A B 



6.0 
7.0 
8.0 
9.0 
10.0 
11.0 
12.0 
13.0 
14.0 
15.0 
16.0 
17.0 
18.0 
19.0 
20.0 
22.0 
24.0 
26.0 
28.0 
30.0 
32.0 
34.0 
36.0 
38.0 
40.0 
42.0 
44.0 
46.0 
48.0 
50.0 
54.0 
58.0 
62.0 
66.0 
70.0 
74.0 
78.0 
82.0 
86.0 
90.0 



42.0 44.0 46.0 48.0 50.0 54.0 58.0 62.0 66.0 70.0 74.0 78.0 82.0 86.0 90.0 



D = 1 3 (Ax B)QS25 

^equiv ,0 ^ + B) - 25 

where: 

^equiv = equivalent round duct size for rectangular 
duct, in. 

A = one side of rectangular duct, in. 

B = adjacent side of rectangular duct, in. 



45.9 

47.0 48.1 

48.0 49.2 50.3 

49.1 50.2 51.4 52.5 
50.0 51.2 52.4 53.6 54.7 
52.0 53.2 54.4 55.6 56.8 59.0 
53.8 55.1 56.4 57.6 58.8 61.2 63.4 
55.5 56.9 58.2 59.5 60.8 63.2 65.5 67.8 

57.2 58.6 60.0 61.3 62.6 65.2 67.6 69.9 72.1 
58.8 60.3 61.7 63.1 64.4 67.1 69.6 72.0 74.3 76.5 

60.3 61.9 63.3 64.8 66.2 68.9 71.5 74.0 76.4 78.7 80.9 
61.8 63.4 64.9 66.4 67.9 70.6 73.3 75.9 78.4 80.7 83.0 85.3 
63.3 64.9 66.5 68.0 69.5 72.3 75.1 77.8 80.3 82.8 85.1 87.4 89.6 
64.7 66.3 67.9 69.5 71.0 74.0 76.8 79.6 82.2 84.7 87.1 89.5 91.8 94.0 
66.0 67.7 69.4 71.0 72.6 75.6 78.5 81.3 84.0 86.6 89.1 91.5 93.9 96.2 98.4 



Exhaust System Design Procedure 5-53 



TABLE 5-10. Air Density Correction Factor, df 




































ALTITUDE RELATIVE TO SEA LEVEL, 1 


t 














-5000 


-4000 


-3000 


-2000 


-1000 





1000 


2000 


3000 


4000 


5000 


6000 


7000 


8000 


9000 


10000 
















Barometric Pressure 
















"Hg 


35.74 


34.51 


33.31 


32.15 


31.02 


29.92 


28.86 


27.82 


26.82 


25.84 


24.89 


23.98 


23.09 


22.22 


21.39 


20.57 


"w 


486.74 469.97 453.67 437.84 422.45 407.50 392.98 378.89 365.21 : 


351.93 339.04 326.54 314.42 302.66 291.26 280.21 


Temp., 


































F 














Density Factor, dt 
















-40 


1.51 


1.46 


1.40 


1.36 


1.31 


1.26 


1.22 


1.17 


1.13 


1.09 


1.05 


1.01 


0.97 


0.94 


0.90 


0.87 





1.38 


1.33 


1.28 


1.24 


1.19 


1.15 


1.11 


1.07 


1.03 


1.00 


0.96 


0.92 


0.89 


0.86 


0.82 


0.79 


40 


1.27 


1.22 


1.18 


1.14 


1.10 


1.06 


1.02 


0.99 


0.95 


0.92 


0.88 


0.85 


0.82 


0.79 


0.76 


0.73 


70 


1.19 


1.15 


1.11 


1.07 


1.04 


1.00 


0.96 


0.93 


0.90 


0.86 


0.83 


0.80 


0.77 


0.74 


0.71 


0.69 


100 


1.13 


1.09 


1.05 


1.02 


0.98 


0.95 


0.91 


0.88 


0.85 


0.82 


0.79 


0.76 


0.73 


0.70 


0.68 


0.65 


150 


1.04 


1.00 


0.97 


0.93 


0.90 


0.87 


0.84 


0.81 


0.78 


0.75 


0.72 


0.70 


0.67 


0.65 


0.62 


0.60 


200 


0.96 


0.93 


0.89 


0.86 


0.83 


0.80 


0.77 


0.75 


0.72 


0.69 


0.67 


0.64 


0.62 


0.60 


0.57 


0.55 


250 


0.89 


0.86 


0.83 


0.80 


0.77 


0.75 


0.72 


0.69 


0.67 


0.64 


0.62 


0.60 


0.58 


0.55 


0.53 


0.51 


300 


0.83 


0.80 


0.78 


0.75 


0.72 


0.70 


0.67 


0.65 


0.62 


0.60 


0.58 


0.56 


0.54 


0.52 


0.50 


0.48 


350 


0.78 


0.75 


0.73 


0.70 


0.68 


0.65 


0.63 


0.61 


0.59 


0.57 


0.54 


0.52 


0.50 


0.49 


0.47 


0.45 


400 


0.74 


0.71 


0.69 


0.66 


0.64 


0.62 


0.59 


0.57 


0.55 


0.53 


0.51 


0.49 


0.48 


0.46 


0.44 


0.42 


450 


0.70 


0.67 


0.65 


0.63 


0.60 


0.58 


0.56 


0.54 


0.52 


0.50 


0.48 


0.47 


0.45 


0.43 


0.42 


0.40 


500 


0.66 


0.64 


0.61 


0.59 


0.57 


0.55 


0.53 


0.51 


0.49 


0.48 


0.46 


0.44 


0.43 


0.41 


0.39 


0.38 


550 


0.63 


0.61 


0.58 


0.56 


0.54 


0.52 


0.51 


0.49 


0.47 


0.45 


0.44 


0.42 


0.40 


0.39 


0.38 


0.36 


600 


0.60 


0.58 


0.56 


0.54 


0.52 


0.50 


0.48 


0.46 


0.45 


0.43 


0.42 


0.40 


0.39 


0.37 


0.36 


0.34 


700 


0.55 


0.53 


0.51 


0.49 


0.47 


0.46 


0.44 


0.42 


0.41 


0.39 


0.38 


0.37 


0.35 


0.34 


0.33 


0.31 


800 


0.50 


0.49 


0.47 


0.45 


0.44 


0.42 


0.41 


0.39 


0.38 


0.36 


0.35 


0.34 


0.32 


0.31 


0.30 


0.29 


900 


0.47 


0.45 


0.43 


0.42 


0.40 


0.39 


0.38 


0.36 


0.35 


0.34 


0.32 


0.31 


0.30 


0.29 


0.28 


0.27 


1000 


0.43 


0.42 


0.40 


0.39 


0.38 


0.36 


0.35 


0.34 


0.33 


0.31 


0.30 


0.29 


0.28 


0.27 


0.26 


0.25 



5-54 



Industrial Ventilation 



PSYCHROMETRIC CHART 

Barometric Pressure 29.92" Hg. 



170- 




45 50 



55 



60 65 70 75 80 85 

Dry Bulb Temperatures - °F 



90 



95 100 105 HO 115 



©American Air Filter Co. Inc., 1959 
Louisville, Ky. Form 1932 



FIGURE 5-17 



Total Heat Values - ASHAE Guide 
End Points - Ztmmsnnan & Lavlne 



Exhaust System Design Procedure 5-55 



HIGH U&PSBATUBgS 
Sarasaofric Pressure 29.92 la. Hg 



CMTHALPY OP REJECTED 
Oft ADDED WATER 



130 



ISO 170 180 190 gOO 210 220 230 240 



60 70 SO 



100 MO 



130 140 ISO 160 170 180 

DRY BULB TEMPERATURE F 




2)0 220 230 240 250 

COPYRIGHT 1944 CAfiRltR CORPORATION 



FIGURE 5-18 



Barometric Pressure 29,92 in Hg 



0.30 




500 



DRY BULB TEMPERATURE IN DEGREES F 

FIGURE 5-19. Psychrometric chart for humid air based on one pound dry weight (© 1951 American Air Filter Co., Inc., Louisville, KY) 






a 
a 
c 



< 



o 

a 




400 500 600 700 300 900 

DRY BULB TEMPERATURE -DEGREES F. 



tOOO 1100 1200 I3O0 1400 1500 



FIGURE 5-20 



-4 



5-58 Industrial Ventilation 



L / 



t~^. 



^ 




PREFERRED 



2 to 2.5 dia. I — f- 
center fine U--f 

radius (C.L.R.) 




XT 



.5 dia. 



ACCEPTABLE 
3LB0W RADIUS 



AVOID 



hibows should be 2 to 2.5 diameter cen tertine radius except 
where space does not oerrnit. See Fiq. 5-13 tor loss factor. 



-t D 



•V 



■i-|4- 




.V 



-v 



PREFERRED AVOID 

ASPECT RATIO (?) 

Elbows should have (^) and (jej equal to or greater than (1). 
See Fig. 5-13 for toss factor. 



Note: Avoid mitered elbows. If necessary, use only with clean 
air and provide turning vanes. Consult mfq. for turning 
vane loss factor. 



AMERIC AN C N FERE NC E 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



PRINCIPLES OF DUCT DESIGN 
ELBO WS 



DATE 2__g 5 



FIGURE 



5-21 



Exhaust System Design Procedure 5-59 






T 1Z- 



- RUBBER 
BELTING 



AIR FLOW 



SECTION 



SECTION \ 



ANGLE IRON -,-. 



REMOVABLE WEAR PLATE 
12 go. OR HEAVIER. 




FLANGE 



- ID MINI. 




\- REMOVABl E 
WEAR PLATE 



1.AT BACK ELBOW 



A" 



\ 
\ \ \ 
\ \ \ 
\ 
\ 
\ 




2D 



FLANGE 



— 3" MINIMUM 
CONCRETE 



4-X-f- 

v.. I >a 



CONCRETE REINFORCED El BOW 

NOTE: PROVIDE SOLID MOUNTING FOR CONCRETE REINFORCED ELBOWS. 



AMERIC AN CO NFERENC E 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENJSTS 



UVY DUTY ELBOWS 



DAI 



1 - 95 



-IGURi; 



5 22 



5-60 Industrial Ventilation 



Jr 




1L 



A/ 



;| ii i Of ! T ^ A f 



^ 



i>3- -i h 



LIDE 



C.I. HINGED DOOR 













o c o o 


O O o 




( 


I 


\ 


] 





\ 


' " / 




















SPLIT SLEEVE 
(Al SO LAN CONNECTION) 



PUL! OUT CAP 



AMERICAN CONFERENCE 
OF GOVERNMENTAL 

INDUSTRIAL HYGIENISTS 



t 




a 





CLE AND UT OPENINGS 



DATE 



1-95 



FIGURE 



5-23 



Exhaust System Design Procedure 5-61 




4> 

r. 

V 

r 


:> 

> 

A 



L DRILL AND RIVET OR BOLT 

AT FIXED POSITION, 





AMERICAN CONFERENCE 

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INDUSTRIAL HYGIENISTS 



BLAST GATES 



DATE 



1-95 



FIGURE 



5-24 



5-62 Industrial Ventilation 



DUCT ENLARGEMENTS 



^- See Fii 



r ig. 5-16 

PREFERRED 



AVOID 



DUCT CONTRACTIONS 



^ 



See Fig. 5-16 

PREFERRED 



AVOID 



SYMMETRICAL WYES 




30 
to 
60° 

PREFERRED 




60° 

PREFERRED 



AVOID 



AMERICAN CONFERENCE 

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INDUSTRIAL HYGIENISTS 



PRINCIPLES OF DUCT DESIGN 



DATE 1-95 



] FIGURE 5-25 



Exhaust System Design Procedure 5-63 



30 




/- 



PREFERRED 



ACCEPTABLE 



AVOID 



c 



— „.._iy 



PREFERRED 



ACCEPTABLE 
BRANCH ENTRY 



AVOID 



Br onches should enter ot gradual expansions and at on angle 
of 30° or less (preferred) to 45" if necessary. Expansion should 
be 15° maximum. See Fig. 5-15 for loss coefficients. 



Vm 



A 



Vm 



^- A_3= A r l- A 2 r ?.:■ 






'ERRED 



vrn Vm - Minimum transport velocity 
A ---■ Cross section area 



AVOID 



PROPER DUCT SIZE 
Size the duct to maintain the selected or higher 
transport velocity. 



AMERICAN CONFERENCE 

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INDUSTRIAL HYGIENISTS 



PRINCIPLES OF DUCT DESIGN 

BRANCH ENTRY 



DATE I __ 95 



FIGURE 



5-26 



5-64 Industrial Ventilation 



<- Straight inlet 
JL 



PREFERRED 



^V 




ACCEPTABLE 

Inlet elbow see note 



Note: 

See Chapter 6 for system ef f ect 
factors based on inlet and 
outlet duct arrangements. 



\ f~ Tapered inlet 



a 



PREFERRED 



-— 0\ ♦ 


-^^ 


[ ) i A 




^~"' / t 


^;:- 


K_-^ 




B — 


A 



r ~\ 




n 


c 


v J 




1 " 






\ - 


------ 




i . 


__ r 








i 





A™ twice wheel dia m in. 
B-- twice wheel dia min. 
n = wheel width min. 



soszt 



ACCEPTABLE 

Inlet elbow see note 




Use duct turn vanes to eliminate air 
spin or uneven loading of fan wheel. 



AMERICAN CONFERENCE 

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INDUSTRIAL HYC1ENISTS 



PRINCIPLES OF DUCT DESIGN 
FAN INLETS 



DATE 1-88 



1GURE 5- £7 



Exhaust System Design Procedure 5-65 



Wind unaffected by building 




V////////////////////////////////////////////////////////// ///////////X///7A 



Zone of 



recirculating flow 



A: Centerline flow patterns around a 
rectangular building 



Undisturbed flow 



1.5R 



Z1 Roof recirculation region 
Z2 High turbulence region 
23 Roof Wake boundry 



Wind 



Speed 




/////////////////////// ///////////////////////////////^ 



Building wake 

recirculation 

region 



B: Building Recirculation Cavities 



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INDUSTRIAL HYGIENISTS 



AIRFLOW AROUND 
BUILDINGS 



DATE 



5-94 figure 5-28 



5-66 



Industrial Ventilation 




) 

ue to 
turn and 
bouyancy 



Effective stack 
height 



A: Effective stack height 



From weather data: 

Days with max. wind vel. <9 m/s= 98% 

Avg wind vel.= 4.5 m/s 

* ^ . For Design: 

"^ N ) Assume 9 m/s, R = 1.5, 



Wind 



then stack veL 
be 13.5 m/s 




should 



Extensive Downwash into 
wake of stack 

B: Wake Downwash 



Trailing Vortices 



AMERICAN CONFERENCE 

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EFFECTIVE STACK HEIGHT 
AND WAKE DOWNWASH 



DATE 



4-94 "F IGURE 5 : -29 



Exhaust System Design Procedure 5-67 



D-rT 



T"n 



■bi- 



section A-A 




VERTICAL DISCHARGE 

NO LOSS 



1 1/2D + 1 



Bracket upper 

stack to 
discharge due 

(87/116 




Drain ■ 



OFFSET ELBOWS OFFSET STACK 

CALCULATE LOSSES DUE TO ELBOWS 



■ Drain 




(106) 



1. Rain protection characteristics of these caps are superior to a deflecting cap located 
0.75D from top of a stack. 

2. The length of upper stack is related to rain protection. Excessive additional distance may 
"blowout" of effluent at the gap between upper and lower sections. (86) 

WEATHER CAP 
Equal velocity contours 



STACKHEAD 

50 , ^^ ,12 



o 
o 



CD 
cn 

a 

o 

CO 



60 
75 



10 

8 

6 

4 

2 





CD 
CD 



O 

Q 



CO 

b o 



E Z 
6 




00 



100 



PREFERRED 
Deflects air upward 



12 10 8 6 4 2 
Diameters 



/H 3 \ 



iOT RECOMMENDED 



Deflects air downward 
AVOID 



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STACKHEAD DESIGNS 



DATE 



1-95 l FIGURE 5-30 



Chapter 6 

FANS 



6.1 INTRODUCTION 6-2 

6.2 BASIC DEFINITIONS 6-2 

6.2.1 Ejectors 6-2 

6.2.2 Axial Fans 6-2 

6.2.3 Centrifugal Fans 6-2 

6.2.4 Special Type Fans 6-2 

6.3 FAN SELECTION 6-6 

6.3.1 Consideration for Fan Selection 6-6 

6.3.2 Rating Tables 6-15 

6.3.3 Point of Operation 6-16 

6.3.4 Matching Fan Performance 

and System Requirements 6-17 

Figure 6-1 Air Ejectors 6-3 

Figure 6-2 Terminology for Axial and Tubular Centrifugal 

Fans 6-4 

Figure 6-3 Terminology for Centrifugal Fan Components . . 6-5 
Figure 6-4 Types of Fans: Impeller and Housing Designs/ 

Performance Characteristics and Applications . . 6-6 
Figure 6-5a Drive Arrangements for Centrifugal Fans . . . .6-11 
Figure 6-5b Drive Arrangements for Centrifugal Fans . . . .6-12 
Figure 6-5c Drive Arrangements for Axial Fans with 

or Without Evase and Inlet Box 6-13 

Figure 6-6 Estimated Belt Drive Loss 6-14 

Figure 6-7 Typical Fan Performance Curves 6-16 

Figure 6-8 System Requirement Curves 6-17 

Figure 6-9 Actual Versus Desired Point of Operation . . . .6-18 

Figure 6-10 Homologous Performance Curve 6-19 

Figure 6-11 In-Duct Heater 6-20 

Figure 6- 12 Fans: Parallel Operation 6-22 

Figure 6-13 Fans: Series Operation 6-23 

Figure 6-14 System Effect Factor 6-24 



6.3.5 Fan Laws 6-17 

6.3 .6 The Effect of Changing Rotation Rate or 

Gas Density 6-17 

6.3.7 Limitations on the Use of Fan Laws 6-19 

6.3.8 Fan Selection at Air Density Other Than 

Standard 6-19 

6.3.9 Explosive or Flammable Materials 6-21 

6.4 FAN INSTALLATION AND MAINTENANCE . . .6-21 

6.4.1 System Effect 6-21 

6.4.2 Inspection and Maintenance 6-25 

REFERENCES 6-25 



Figure 6-15 Fan Discharge Conditions 6-24 

Figure 6-16 Inlet Elbow 6-25 

Figure 6-17 Machinery Vibration Severity Chart 6-26 

Figure 6-18 System Effect Curves for Outlet Ducts — 

Centrifugal Fans 6-27 

Figure 6-19 System Effect Curves for Outlet Ducts — 

Axial Fans 6-28 

Figure 6-20 System Effect Curves for Outlet Elbows 

on Centrifugal Fans 6-29 

Figure 6-21 System Effect Curves for Various Mitered Elbows 

Without Turning Vanes 6-30 

Figure 6-22 System Effect Curves for Various Mitered 

Square Duct Elbows 6-31 

Figure 6-23 Non-Uniform Inlet Flows 6-32 

Figure 6-24 Non-Uniform Inlet Corrections 6-33 

Figure 6-25 System Effect Curves for Inlet Obstructions . . 6-34 

Figure 6-26 System Effect Curves 6-35 

Figure 6-27 System Effect Curves 6-36 



6-2 



Industrial Ventilation 



6.1 INTRODUCTION 

To move air in a ventilation or exhaust system, energy is 
required to overcome the system losses. This energy can be 
in the form of natural convection or buoyancy. Most systems, 
however, require some powered air moving device such as a 
fan or an ejector. 

This chapter will describe the various air moving devices 
that are used in industrial applications, provide guidelines for 
the selection of the air moving device for a given situation, 
and discuss the proper installation of the air moving device in 
the system to achieve desired performance. 

Selection of an air moving device can be a complex task, 
and the specifier is encouraged to take advantage of all 
available information from applicable trade associations as 
well as from individual manufacturers. 

6.2 BASIC DEFINITIONS 

Air moving devices can be divided into two basic classifi- 
cations: ejectors and fans. Ejectors have low operating effi- 
ciencies and are used only for special material handling 
applications. Fans are the primary air moving devices used in 
industrial applications. 

Fans can be divided into three basic groups: axial, centrifu- 
gal, and special types. As a general rule, axial fans are used 
for higher flow rates at lower resistances and centrifugal fans 
are used for lower flow rates at higher resistances. 

6.2.1 Ejectors: (see Figure 6-l)Are used sometimes when 
it is not desirable to have contaminated air pass directly 
through the air moving device. Ejectors are utilized for air 
streams containing corrosive, flammable, explosive, hot, or 
sticky materials that might damage a fan; present a dangerous 
operating situation; or quickly degrade fan performance. 
Ejectors also are used in pneumatic conveying systems. 

6.2.2 Axial Fans: There are three basic types of axial fans: 
propeller, tubeaxial, and vaneaxial (see Figures 6-2 and 6-3). 

Propeller Fans are used for moving air against low static 
pressures and are used commonly for general ventilation. Two 
types of blades are available: disc blade types when there is 
no duct present; narrow or propeller blade types for moving 
air against low resistances (less than l"wg). Performance is 
very sensitive to added resistance, and a small increase will 
cause a marked reduction in flow rate. 

Tubeaxial Fans (Duct Fans) contain narrow or propeller- 
type blades in a short, cylindrical housing normally without 
any type of straightening vanes. Tubeaxial fans will move air 
against moderate pressures (less than 2 M wg). 

Vaneaxial Fans have propeller configuration with a hub 
and airfoil blades mounted in cylindrical housings which 
normally incorporate straightening vanes on the discharge 
side of the impeller. Compared to other axial flow fans, 



vaneaxial fans are more efficient and generally will develop 
higher pressures (up to 8"wg). They are limited normally to 
clean air applications. 

6.2.3 Centrifugal Fans: (see Figures 6-4 and 6-5): These 
fans have three basic impeller designs: forward curved, radial, 
and backward inclined/backward curved. 

Forward curved (commonly called "squirrel cages") impel- 
lers have blades which curve toward the direction of rotation. 
These fans have low space requirements, low tip speeds, and 
are quiet in operation. They usually are used against low to 
moderate static pressures such as those encountered in 
heating and air conditioning work and replacement air 
systems. This type of fan is not recommended for dusts or 
particulates that would adhere to the short curved blades 
and cause unbalance. 

Radial Impellers have blades which are straight or radial 
from the hub. The housings are designed with their inlets and 
outlets sized to produce material conveying velocities. There 
is a variety of impeller types available ranging from "high 
efficiency, minimum material" to "heavy impact resis- 
tance" designs. The radial blade shape will resist material 
buildup. This fan design is used for most exhaust system 
applications when particulates will pass through the fan. 
These fans usually have medium tip speeds and are used 
for a variety of exhaust systems which handle either clean 
or dirty air. 

Backward Inclined/Backward Curved impeller blades are 
inclined opposite to the direction of fan rotation. This type 
usually has higher tip speeds and provides high fan efficiency 
and relatively low noise levels with "non-overloading" horse- 
power characteristics. In anon-overloading fan, the maximum 
horsepower occurs near the optimum operating point so any 
variation from that point due to a change in system resistance 
will result in a reduction in operating horsepower. The blade 
shape is conducive to material buildup so fans in this group 
should be limited as follows: 

• Single-Thickness Blade: Solid blades allow the unit to 
handle light dust loading or moisture. It should not be 
used with particulates that would build up on the 
underside of the blade surfaces. 

• Airfoil Blade: Airfoil blades offer higher efficiencies 
and lower noise characteristics. Hollow blades erode 
more quickly with material and can fill with liquid in 
high humidity applications. These should be limited to 
clean air service. 

6.2.4 Special Type Fans (see Figure 6-4): In-line Cen- 
trifugal Fans have backward inclined blades with special 
housings which permit a straight line duct installation. Pres- 
sure versus flow rate versus horsepower performance curves 
are similar to a scroll-type centrifugal fan of the same blade 
type. Space requirements are similar to vaneaxial fans. 



Fans 6-3 



"x: 



/ \ 



A 



INDUCED AIR 

"~Z~2^* — PRIMAR V AIR 



b_ 



A 

/ \ 

t \ \ 
i v / 






i 



i 



\ 




c 



"Q 



B 



PRIMARY AIR 



W 



EJECTOR FOR PNEUM/ 



KEYING 



D 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



AIR EJECT 



DATE 



rE 1-88 



FIGURE 



6-1 



6-4 Industrial Ventilation 



BELT TUBE 



.CASING 



INLET 



TUBULAR CENTRIFUGAL FAN4MRECT DRIVE 




OIFFUSER 



MOTOR 



CASING 



BEARING CASING 




TOBEAX1AL FAN-DIRECT DRIVE 

(IMPELLER DOWNSTREAM) 



BLADE 



HUB 



GUIDE VANE 



O 



IMPELLER 



VANEAXIAL FAN-BELT DRIVE 



Reprinted from AMCA Publication 201 -90, FANS AND SYSTEMS, by 
permission of the Air Movement and Control Association, lnc (61) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



TERMINOLOGY FOR AXIAL 
AND TUBULAR CENTRIFUGAL 

FANS 



DATE 5-92 



FIGURE 



6-2 



Fans 6-5 



HOUSING 



DIVERTER 



DISCHARGE 

OUTLET AREA 




FRAME 



INLET COLLAR 



Reprinted from AMCA Publication 201 -90, FANS AND SYSTEMS, by 
permission of the Air Movement and Control Association, lnc. ( 1) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



TERMINOLOGY FOE 

CENTRIFUGAL FAN 

COMPONENTS 



DATE 



5-92 I FIGURE 6-3 



6-6 



Industrial Ventilation 



Power Exhausters, Power Roof Ventilators are packaged 
units that can be either axial flow or centrifugal type. The 
centrifugal type does not use a scroll housing but discharges 
around the periphery of the ventilator to the atmosphere. 
These units can be obtained with either downward deflecting 
or upblast discharges. 

Fan and Dust Collector Combination: There are several 
designs in which fans and dust collectors are packaged in a 
unit. If use of such equipment is contemplated, the manufac- 
turer should be consulted for proper application and perform- 
ance characteristics. 

6.3 FAN SELECTION 

Fan selection involves not only finding a fan to match the 
required flow and pressure considerations but all aspects of 



an installation including the air stream characteristics, oper- 
ating temperature, drive arrangement, and mounting. Section 
62 discussed the various fan types and why they might be 
selected. This section offers guidelines to fan selection; how- 
ever, the exact performance and operating limitations of a 
particular fan should be obtained from the original equipment 
manufacturer. 

6.3.1 Considerations for Fan Seiection: 

CAPACITY 

Flow Rate (Q)\ Based on system requirements and ex- 
pressed as actual cubic feet per minute (acfm) at the fan inlet. 

Pressure Requirements: Based on system pressure require- 
ments which normally are expressed as Fan Static Pressure 
(FSP) or Fan Total Pressure (FTP) in inches of water gauge 



TYPE 



IMPELLER DESIGN 



HOUSING DESIGN 




Highest efficiency of all centrifugal fan 
designs. 9 to 16 blades of airfoil contour 
curved away from the direction of rotation. 
Air leaves the impeller at a velocity less 
than its tip speed and relatively deep 
blades provide for efficient expansion 
within the blade passages. For given 
duty, this will be the highest speed of the 
centrifugal fan designs. 




Scroll-type, usually designed to permit 
efficient conversion of velocity pressure 
to static pressure, thus permitting a high 
static efficiency; essential that clearance 
and alignment between wheel and inlet 
bell be very close in order to reach the 
maxiumum efficiency capability. Con- 
centric housings can also be used as in 
power roof ventilators, since there is effi- 
cient pressure conversion in the wheel. 



CO 






-J DC 

11 




Efficiency is only slightly less than that of 
airfoil fans. Backward-inclined or back- 
ward-curved blades are single thickness. 
9 to 16 blades curved or inclined away 
from the direction of rotation. Efficient for 
the same reasons given for the airfoil fan 
above. 




Utilizes the same housing configuration 
as the airfoil design. 



DC 




Simplest of all centrifugal fans and least 
efficient. Has high mechanical strength 
and the wheel is easily repaired. For a 
given point of rating, this fan requires 
medium speed. This classification 
includes radial blades (R) and modifi- 
ed radial blades (M), usually 6 to 10 in 
number. 




Scroll-type, usually the narrowest design 
of all centrifugal fan designs described 
here because of required high velocity 
discharge. Dimensional requirements of 
this housing are more critical than for air- 
foil and backward-inclined blades. 



CJ 



DC 

O 




Efficiency is less than airfoil and back- 
ward-curved bladed fans. Usually fab- 
ricated of lightweight and low cost con- 
struction. Has 24 to 64 shallow blades 
with both the heel and tip curved forward. 
Air leaves wheel at velocity greater than 
wheel. Tip speed and primary energy 
transferred to the air is by use of high 
velocity in the wheel. For given duty, 
wheel is the smallest of all centrifugal 
types and operates at lowest speed. 




Scroll is similar to other centrifugal-fan 
designs. The fit between the wheel and 
inlet is not as critical as on airfoil and 
backward-inclinded bladed fans. Uses 
large cut-off sheet in housing. 



FIGURE 6-4. Types of fans: impeiier and housing designs (see facing page) 



Fans 



6-7 



at standard conditions (0.075 lbm/ft 3 ). If the required pressure 
is known only at non standard conditions, a density correction 
(see Section 6.3.8) must be made. 

AIR STREAM 

Material handled through the fan: When the exhaust air 
contains a small amount of smoke or dust, a backward inclined 
centrifugal or axial fan should be selected. With light dust, 
fume or moisture, a backward inclined or radial centrifugal 
fan would be the preferred selection. If the particulate loading 
is high, or when material is handled, the normal selection 
would be a radial centrifugal fan. 

Explosive or Flammable Material: Use spark resistant con- 
struction (explosion-proof motor if the motor is in the air 



stream). Conform to the standards of the National Board of 
Fire Underwriters, the National Fire Protection Association 
and governmental regulations (see Section 6.3.9). 

Corrosive Applications'. May require a protective coating 
or special materials of construction (stainless, fiberglass, etc.) 

Elevated Air Stream Temperatures: Maximum operating 
temperature affects strength of materials and therefore must 
be known for selection of correct materials of construction, 
arrangement, and bearing types. 

PHYSICAL LIMITATIONS 

Fan size should be determined by performance require- 
ments. Inlet size and location, fan weight, and ease of main- 



PEBFORHflANCE CURVES 



PERFORMANCE CHARACTERISTICS' 



APPLICATIONS 







Highest efficiencies occur 50 to 65% of 
wide open volume. This is also the area of 
good pressure characteristics; the horse- 
power curve reaches a maximum near the 
peak efficiency area and becomes lower 
toward free delivery, a self-limiting power 
characteristic as shown. 



General heating, ventilating and air-con- 
ditioning systems. Used in large sizes for 
clean air industrial applications where 
power savings are significant. 



volume: flow rate 



ctIO 

& 8 

1 6 

UJ 

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in 9 
uj y - 

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8 10 



o 

LU 

o 



Operating characteristics of this fan are 
similar to the airfoil fan mentioned above. 
Peak efficiency for this fan is slightly lower 
than the airfoil fan. Normally unstable left of 
peak pressure. 



Same heating, ventilating, and air-con- 
ditioning applications as the airfoil fan. Also 
used in some industrial applications where 
the airfoil blade is not acceptable because 
of corrosive and/or erosion environment. 



VOLUME FLOW RATE 



;10 
- 8 
6 
4 
2 




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10 

8 

6 

4 

2 





Higher pressure characteristics than the 
above mentioned fans. Power rises con- 
tinually to free delivery. 



2 



4 



6 



8 10 



Used primarily for material-handing 
applications in industrial plants. Wheel can 
be of rugged construction and is simple to 
repair in the field. Wheel is sometimes 
coated with special material. This design 
also used for high-pressure industrial 
requirements. Not commonly found in 
HVAC applications. 



VOLUME FLOW RATE 



or 10 

Q - 8 

l n 
b 

LU 

% 4 

CO 

en o 

LU y - 
Q_ 



_ 




S^sp ^_^ 


/ ~ 


- uz/y^^ 


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2 4 6 8 10 
VOLUME FLOW RATE 



Pressure curve is less steep than that of 
backward-curved bladed fans. There is a 
dip in the pressure curve left of the peak 
pressure point and highest efficiency 
occurs to the right of peak pressure, 40 to 
50% of wide open volume. Fan should be 
rated to the right of peak pressure. Power 
curve rises continually toward free delivery 
and this must be taken into account when 
motor is selected. 



Used primarily in low-pressure heating, 
ventilating, and air-conditioning applica- 
tions such as domestic furnaces, central 
station units, and packaged air-con- 
ditioning equipment from room air-con- 
ditioning units to roof top units. 



Types of fans: Performance characteristics and applications. (*These performance curves reflect the general characteristics of various fans as commonly 
employed. They are not intended to provide complete selection criteria for application purposes, since other parameters, such as diameter and speed, are 
not defined.) 



6-8 



Industrial Ventilation 



TYPE 



IMPELLER DESIGN 



HOUSING DESIGN 



UJ 
Q_ 
O 
DC 




Efficiency is low. Impellers are usually of 
inexpensive construction and limited to 
low pressure applications. Impeller is of 2 
or more blades, usually of single thick- 
ness attached to relatively small hub. 
Energy transfer is primarily in form of 
velocity pressure. 



Simple circular ring, orifice plate, or ven- 
turi design. Design can substantially 
influence performance and optimum 
design is reasonably close to the blade 
tips and forms a smooth in let flow contour 
to the wheel. 



CO 

-J 
< 

3 



X 
< 
UJ 
CD 




Somewhat more efficient than propeller 
fan design and is capable of developing 
a more useful static pressure range. 
Number of blades usually from 4 to 8 and 
hub is usually less than 50% of fan tip 
diameter. Blades can be of airfoil or single 
thickness cross section. 




Cylindrical tube formed so that the run- 
ning clearance between the wheel tip 
and tube is close. This results in signifi- 
cant improvement over propeller fans. 



< 

ui 

z 




Good design of blades permits medium- 
to high-pressure capability at good effi- 
ciency. The most efficient fans of this type 
have airfoil blades. Blades are fixed or 
adjustable pitch types and hub is usually 
greater than 50% of fan tip diameter. 



[ 1 





Cylindrical tube closely fitted to the outer 
diameter of blade tips and fitted with a set 
of guide vanes. Upstream or downstream 
from the impeller, guide vanes convert the 
rotary energy imparted to the air and 
increase pressure and efficiency of fan. 



< 
O 

u. 
CC 

Z 

UJ 

o 




This fan usually has a wheel similar to the 
airfoil backward-inclined or backward- 
curved blade as described above. 
(However, this fan wheel type is of lower 
efficiency when used in fan of type.) 
Mixed flow impellers are sometimes 
used. 





1 1 




^ 


F 








\ i 





Cylindrical shell similar to a vaneaxial fan 
housing, except the outer diameter of the 
wheel does not run close to the housing. 
Air is discharged radially from the wheel 
and must change direction by 90 
degrees to flow through the guide vane 
section. 



•as. 






CO 

CC 

o 



© 
o 
ce 



© 
a. 



< 
CD 

u. 
CC 

z 
o 




Many models use airfoil or backward- 
inclined impeller designs. These have 
been modified from those mentioned 
above to produce a low-pressure, high- 
volume flow rate characteristic. In addi- 
tion, many special centrifugal impeller 
designs are used, including mixed-flow 
design. 



JELL 



Does not utilize a housing in a normal 
sense since the air is simply discharged 
from the impeller in a 360 degree pattern 
and usually does not include a configura- 
tion to recover the velocity pressure com- 
ponent. 



3 




A great variety of propeller designs are 
employed with the objective of high-vol- 
ume flow rate at low pressure. 




Essentially a propeller fan mounted in a 
supporting structure with a cover for 
weather protection and safety considera- 
tions. The air is discharged through the 
annular space around the bottom of the 
weather hood. 



FIGURE 6^ (continued). Types of fans: impeller and housing design 



Fans 



6-9 



PERFORMANCE CURVES 



PERFORMANCE CHARACTERISTICS* 



APPLICATIONS 




High flow rate but very low-pressure capa- 
bilities and maxiurnum efficiency is 
reached near free delivery. The discharge 
pattern of the air is circular in shape and the 
air stream swirls because of the action of 
the blades and the lack of straightening 
facilities. 



For low-pressure, high-volume air moving 
applications such as air circulation within a 
space or ventilation through a wall without 
attached duct work. Used for replacement 
air applications. 



2 4 6 8 10 
VOLUME FLOW RATE 




High flow-rate characteristics with medium- 
pressure capabilities. Performance curve 
includes a dip to the left of peak pressure 
which should be avoided. The discharge 
air pattern is circular and is rotating or whirl- 
ing because of the propeller rotation and 
lack of guide vanes. 



2 4 6 

VOLUME FLOW RATE 



Low- and medium-pressure ducted heat- 
ing, ventilating, and air-conditioning 
applications where air distribution on the 
downstream side is not critical. Also used in 
some industrial applications such as dry- 
ing ovens, paint spray booths, and fume 
exhaust systems. 




2 4 6 

VOLUME FLOW RATE 



High-pressure characteristics with medium 
volume flow rate capabilities. Performance 
curve includes a dip caused by aero- 
dynamic stall to the left of peak pressure, 
which should be avoided. Guide vanes cor- 
rect the circular motion imparted to the air 
by the wheel and improve pressure charac- 
teristics and efficiency of the fan. 



General heating, ventilating, and air-con- 
ditioning systems in low-, medium-, and 
high-pressure applications is of advantage 
where straight-through flow and compact 
instalation are required; air distribution on 
downstream side is good. Also used in 
industrial application similar to the tubeax- 
ial fan. Relatively more compact than com- 
parable centrifugal-type fans for same duty. 




Performance is similar to backward- curved 
fan, except lower capacity and pressure 
because of the 90 degree change in direc- 
tion of the air flow in the housing. The effi- 
ciency will be lower than the backward- 
curved fan. Some designs may have a dip 
in the curve similar to the axial-flow fan. 



Used primarily for low-pressure return air 
systems in heating, ventilating, and air-con- 
ditioning applications. Has straight- 
through flow configuration. 



2 4 6 8 10 
VOLUME FLOW RATE 



O 
a 



00 
CO 

UJ 

or 
a. 



- 








- 






= 


- 




^€ 


: 






PWR ^ 


~ 



2 4 6 8 
VOLUME FLOW RATE 



10 
8 
6 
4 

2 




Usually intended to operate without 
attached ductwork and therefore to operate 
against a very low-pressure head. It is usu- 
ally intended to have a rather high-volume 
flow rate characteristic. Only static pres- 
sure and static efficiency are shown for this 
type of product. 



10 



For low-pressure exhaust systems such as 
general factory, kitchen, warehouse, and 
commercial installations where the low- 
pressure rise limitation can be tolerated. 
Unit is low in first cost and low in operating 
cost and provides positive exhaust ventila- 
tion in the space which is a decided advan- 
tage over gravity-type exhaust units. The 
centrifugal unit is somewhat quieter than 
the axial unit desribed below. 







2 4 6 8 10 
VOLUME FLOW RATE 



Usually intended to operate without 
attached ductwork and therefore to operate 
against very low-pressure head. It is usually 
intended to have a high-volume flow rate 
characteristic. Only static pressure and 
static efficiency are shown for this type of 
product. 



For low-pressure exhaust systems such as 
general factory, kitchen, warehouse, and 
some commercial installations where the 
low-pressure rise limitations can be toler- 
ated. Unit is low in first cost and low in oper- 
ating cost and provides positive exhaust 
ventilation in the space which is a decided 
advantage over gravity-type exhaust units. 



Types of fans: performance characteristics and applications 



6-10 



Industrial Ventilation 



tenance also must be considered. The most efficient fan size 
may not fit the physical space available. 

DRIVE ARRANGEMENTS 

All fans must have some type of power source — usually 
an electric motor. On packaged fans, the motor is furnished 
and mounted by the manufacturer. On larger units, the motor 
is mounted separately and coupled directly to the fan or 
indirectly by a belt drive. A number of standard drive arrange- 
ments are shown in Figures 6-5 a, 6-5b, and 6-5c. 

Direct Drive offers a more compact assembly and assures 
constant fan speed. Fan speeds are limited to available motor 
speeds (except in the case of variable frequency controllers). 
Capacity is set during construction by variations in impeller 
geometry and motor speed. 

Belt Drive offers flexibility in that fan speed can be changed 
by altering the drive ratio. This may be important in some 
applications to provide for changes in system capacity or 
pressure requirements due to changes in process, hood design, 
equipment location or air cleaning equipment. V-belt drives 
must be maintained and have some power losses which can 
be estimated from the chart in Figure 6-6, 

NOISE 

Fan noise is generated by turbulence within the fan housing 
and will vary by fan type, flow rate, pressure, and fan effi- 
ciency. Because each design is different, noise ratings must 
be obtained from the fan manufacturer. Most fans produce a 
"white" noise which is amixture of all frequencies. In addition 
to white noise, radial blade fans also produce a pure tone at a 
frequency equal to the blade passage frequency (BPF): 



BPF-RPMxNxCF 



[6.1] 



where: 

BPF = blade passage frequency, Hz 

RPM = rotational rate, rpm 

N = number of blades 

CF = conversion coefficient, 1/60 

This tone can be very noticeable in some installations and 
should be considered in the system design. 

Because of its higher efficiency, the backward inclined 
impeller design is generally the quietest. However, for all fan 
types, non-uniform air flow at the fan inlet or outlet can 
increase the fan noise level. This is another problem related 
to "system effect" (see Section 6.4.1). 

Most fan manufacturers publish sound ratings for their 
products. There are a variety of ways to present the ratings. 
One popular way is to list sound power levels for eight ANSI 
standard octave bands. The sound power levels are typically 
in units called "decibels" (dB). The sound power level is a 
characteristic of a fan that varies with the fan speed and point 
of operation. 



For an installed fan, the surrounding environment affects 
the sound level that is measured or heard. Walls, floors, and 
other equipment reflect and absorb sound to varying degrees. 
The sound that reaches the listener will be different than the 
fan's rated sound power level. Typical sound measuring de- 
vices detect sound with a microphone and display sound 
pressure level in decibels. This sound pressure is an environ- 
ment-dependent measurement that changes with listener lo- 
cation and/or environment changes. 

While the decibel unit is used for sound power and sound 
pressure, the two measures are not interchangeable. For in- 
stance, 70 dB sound power is not 70 dB sound pressure. The 
decibel is not an absolute unit of measure. It is a ratio between 
a measured quantity and an agreed reference level. Both dB 
scales are logarithmic. The sound power is the log of the ratio 
of two power levels. The sound pressure is the log of the ratio 
of two pressure levels. The sound power scale uses a reference 
of 10~' 2 watts. The sound pressure scale uses a reference of 20 
xl0~ 6 N/M 2 . 

For an installed fan, the sound pressure levels are usually 
measured in dB using the "A" weighting scale. The A-weight- 
ing is used to measure environmental noise as it most closely 
reflects the human auditory response to noise of various 
frequencies. A sound level meter set on the "A" scale auto- 
matically integrates the noise of all frequencies to give a single 
dBA noise measurement. Expanded detail can be obtained by 
taking noise measurements with a meter capable of measuring 
the sound pressure level in each octave band. Such detail can 
help indicate the predominant source of a noise. 

The topic of sound is quite broad and there are many 
reference texts available to cover it. For a concise introduction, 
the ASHRAE Fundamentals Handbook (6 - 4) is a good starting 
point. 

SAFETY AND ACCESSORIES 

Safety Guards are required. Consider all danger points such 
as inlet, outlet, shaft, drive and clean out doors. Construction 
should comply with applicable governmental safety require- 
ments, and attachment must be secure. 

Accessories can help in the installation and in future main- 
tenance requirements. Examples might include drains, 
cleanout doors, split housings, and shaft seals. 

FLOW CONTROL 

There are various accessories that can be used to change 
fan performance. Such changes may be required on systems 
that vary throughout the day or for reduction in flow rate in 
anticipation of some future requirement. Dampers, variable 
pitch blades, and speed control are three common accessories 
used with fans. 

Dampers are installed directly on the fan inlet or outlet. 
Because they are in the air stream, dampers can build up with 



Fans 



6-11 



SW- Single Width 
SI -Single Inlet 



DW 

Dl - 



Double Width 
Double Inlet 



Arrangements 1. 3. 7 and 8 are also available with bearings mounted 
on pedestals or base set independent of the fan housing 




■ j=fr 



ARR. 1 SWSt For belt drive or di- 
rect connection Impeller overhung 
Two bearings on base 



T 





ARR. 2 SWSI For belt drive or di- 
rect connection. Impelleroverhung. 
Bearings in bracket supported by 
fan housing. 



ARR. 3 SWSI For belt drive or di- 
rect connection One bearing on 
each side and supported by fan 
housing. 



ARR. 3 DWDI For belt drive or di- 
rect connection One bearing on 
each side and supported by fan 
housing 






ARR. 4 SWSI For direct drive. Im- 
peller overhung on prime mover 
shaft. No bearings on fan. Prime 
mover base mounted or integrally 
directly connected. 



ARR. 7 SWSI For belt drive or di- 
rect connection. Arrangement 3 
plus base for prime mover 



ARR. 7 DWDI For belt drive or di- 
rect connection. Arrangement 3 
plus base for prime mover 



1 






f 

ML .- jjfrj 


'= 


^ ™ 1 





ARR. 8 SWSI For belt drive or di- 
rect connection. Arrangement 1 
plus extended base for prime 
mover. 



ARR. 9 SWSI For belt drive 
peller overhung, two bearings, > 
prime mover outside base 



3 



ARR. 10 SWSI For bell drive Im- 
peller overhung, two bearings, wtth 
pome mover inside base 



Reprinted from AMCA Publication 99-86, STANDARDS HAND- 
BOOK, by permission of the Air Movement and Control Association, 

Inc.* 6 - 1 ' 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DRIVE ARRANGEMENTS 
FOR CENTRIFUGAL 

FANS 



DATE 



5-92 | FIGURE 6~5a 



6-12 



industrial Ventilation 



sw- 

Sl ■ 



Single Width 
Single Inlet 



DW- 
Dl - 



Double Width 
Double Inlet 




d 

i i 
i i 



i i 




ARR. 1 SWSI WITH INLET BOX For ARR. 3 SWSI WITH INDEPENDENT ARR. 3 SWSI WITH INLET BOX AND 



belt drive or direct connection. I mpel- 
ler overhung, two bearings on base. 
Inlet box may be self-supporting. 



PEDESTAL For belt drive or direct, INDEPENDENT PEDESTALS For 



connection fan. Housing is self-sup- 
porting. One bearing on each side 
supported by independent pedestals. 



belt drive or direct connection fan. 
Housing is self-supporting. One 
bearing on each side supported by in- 
dependent pedestals with shaft ex- 
tending through inlet box. 




I I 




I l 



ARR. 3 DWDI WITH INDEPENDENT 
PEDESTAL For belt drive or direct 
connection fan. Housing is self-sup- 
porting. One bearing on each side 
supported by independent pedestals. 



ARR. 3 DWDI WITH INLET BOX AND 
INDEPENDENT PEDESTALS For 

belt drive or direct connection fan. 
Housing is self-supporting. One 
bearing on eachsidesupportedby in- 
dependent pedestals with shaft ex- 
tending through inlet box. 



ARR. 8 SWSI WITH INLET BOX For 

belt drive or direct connection. I mpel- 
ler overhung, two bearings on base 
plus extended base for prime mover. 
Inlet box may be self-supporting. 



Reprinted from AMCA Publication 99-86, STANDARDS HAND- 
BOOK, by permission of the Air Movement and Control Association, 

lnc.< 6 - 1 > 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DATE 



DRIVE ARRANGEMENTS 
FOR CENTRIFUGAL 

FANS 
5-92 I FIGURE 6-5b 



Fans 



6-13 




Lvase 



Iniet box and 
evase are 
optional on all 
arrangements. FT 



[m=| 



ARR, ] ARR. 1 TWO STAGE 

For belt drive or direct connection, impeller overhung. Two bearings 
located either upstream or downstream of impeller. 




ARR. 3 

F'or bel t drive or direct 
connection. Impeller between 
bearings that are on internal 
supports. Drive through inlet. 





ARR. 4 



ARR. 4 TWO STAGE 



For direct connection. Impeller 
overhung on motor shaft. No 
bearings on fen. Motor on 
internal supports. 




ARR 



For belt drive or direct connection. 
Arr. 3 plus common base for prime 
mover. 




i M 




z[ifi 


" ! 1 






1 



■ ARR. 8 (1 or 2 stage) 

For bel t drive or direc t 
connection. Arr. 1 plus 
common base for prime mover. 





ARR. 9 Motor on Casing ARR. 9 Motor on In 

For belt drive. Impeller overhung. Two bearings on iniemol 
Motor on casing or on integral base. Drive through belt fa 
NOTE: AN fan orientations may be horizontal or vertical. 

Reprinted from AMCA Publication 99-86 Standards Handbook, 

by permission of the Air Movement and Control Association Inc. (60 ) 



tegral Ba 
supports. 

iring. 



se 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DRIVE ARRANGEMENTS FOR 

AXIAL FANS WITH OR WITHOUT 

EVASE AND INLET BOX 



date n-96 



FIGURE 



6- 



5 c 



6-14 



Industrial Ventilation 



100 






80 
60 

* 40 

2 30 

I- 

1 20 

LU 

5 15 

o 

Q_ 

§ 10 
5 8 

5 6 

<n 4 
O 

UJ 3 

> 

Q 2 

1.5 
1 




































































































































































\ 










- R/ 


VNC 


5EO 


F D 


Rn 


• E LOSS 


>ES F 


: OP 


ST/! 


kHD 


AF 


ID BELT 


S 








V 




















i 


































' 


^ 












*^ 


-^ 






























^ 


"v 


** X 


r 












^ — 




























—<r 




















































-^_ 












































































































































0.3 0.4 0.6 0.8 1 2 3 4 6 8 10 20 30 40 60 80 100 200 300 400 600 

MOTOR POWER OUTPUT, hp 

HIGHER BELT SPEEDS TEND TO HAVE HIGHER LOSSES 
THAN LOWER BELT SPEEDS AT THE SAME HORSEPOWER 

*Drive losses are based on the conventional V-belt which has been the "work horse" of the drive industry 
for several decades. 

EXAMPLE 

® Motor power output, H mo , is determined to be 13.3 hp 

• The belts are the standard type and just warm to the touch immediately after shutdown 
® From chart, drive loss = 5.1% 

® Drive loss, H L = 0.051 x 13.3 

= 0.7 hp 

• Fan power input, H = 13.3 - 0.7 

= 12.6 hp. 

Reprinted from AMCA Publication 203-90, FIELD PERFORMANCE 
MEASUREMENT OF FAN SYSTEMS, by permission of the Air 
Movement and Control Association, Inc.* 6 ' 1 * 


AMERICAN CONFERENCE 
OF GOVERNMENTAL 

i TKTTM TCT^T A T TTVPTDMTOmO 


ESTIMATED BELT 
DRIVE LOSS 


1 liMJJUO 1 


IV 


LJAL 


j J 


LI 


lVj 


IJLj 


LNiv_ 


) 1 


J> 


date 5-92 I FIGURE 6-6 



Fans 



6-15 



material and may not be acceptable on material handling fans. 
Two types of dampers are available: 

• Outlet Dampers mount on the fan outlet to add 
resistance to the system when partially closed. These are 
available with both parallel and opposed blades. Selection 
depends on the degree of control required (opposed blade 
dampers will control the flow more evenly throughout the 
entire range from wide open to closed). 

• Inlet Dampers mount on the fan inlet to pre-spin air into 
the impeller. This reduces fan output and lowers 
operating horsepower. Because of the power savings, 
inlet dampers should be considered when the fan will 
operate for long periods at reduced capacities. 

Variable pitch blades are available with some axial- type 
fans. The fan impellers are designed to allow manual or 
automatic changes to the blade pitch. "Adjustable" impellers 
have a blade pitch that can be manually changed when the fan 
is not running. "Variable" impellers include devices to allow 
the blade pitch to be changed pneumatically or hydraulically 
while the fan is operating. 

A Variable Frequency Drive (VFD) may also be used to 
control flow. A VFD will control the fan speed, rather than 
varying the fan inlet flow conditions or the outlet area to 
change the fan's point of operation. This type of control varies 
both the flow rate and the fan static pressure. 

The VFD control unit is connected in-line between the 
electric power source and the fan motor. It is used to vary the 
voltage and frequency of the power input to the motor. The 
motor speed will vary linearly with the line frequency. Most 
VFD applications use a direct drive arrangement; however, 
belt drives are occasionally used. 

For a typical system with fixed physical characteristics, the 
attainable points of operation will fall on the system curve. 
For example, Figure 6- 1 shows points A 1 and A2 on a system 
curve. These two points of operation can be attained with a 
VFD by adjusting it for speeds of RPM] or RPM 2 . This will 
result in fan curve PQi or PQ 2 , respectively. 

VFDs do have disadvantages. They may have a low speed 
limitation. Most AC motors are designed to operate at their 
nameplate speeds. If a VFD is used to run a motor well below 
its nominal speed, the motor's efficiency will be reduced and 
losses will increase. This can increase motor heating and may 
cause damage. 

The VFD can cause harmonic distortion in the electrical 
input lines from the power source. This may affect other 
electrical equipment on the same power system. Such distor- 
tion can be reduced with the addition of isolation transformers 
or line inductors. 

To properly apply a VFD, the equipment supplier needs to 
know about its intended usage, about the building's power 
supply and about other electrical equipment in use. In general, 



for applications where the minimum system air flow is 80% 
or more of the maximum system air flow, the VFD's losses 
and higher initial cost may make use of the inlet damper a 
better choice for flow control. 

An advantage of the VFD or the Variable Pitch Blade over 
the dampers is often a dramatic power and noise reduction. 
However, these accessories usually require additional con- 
trolling equipment. An advantage of dampers is their rela- 
tively simple installation and use and their lower initial costs. 

6.3.2 Rating Tables: Fan size and operating RPM and 
Power usually are obtained from a rating table based on 
required air flow and pressure. Tables are based on FTP or FSP: 



Fan TP = (SP 0Ut)et + VP 0Ut)et ) - (SP jnlet + VP inlet ) 



FanSP = SP, 



outlet 



-SR, 



inlet 



-VR 



inlet 



[6.2] 



[6.3] 



Fan Rating Tables are based on requirements for air at stand- 
ard conditions (0.075 lbm/ft 3 ). If other than standard condi- 
tions exist, the actual pressure must be converted to standard 
conditions. See Section 6.3.8, "Selection at Air Densities 
Other Than Standard." 

The most common form of table is a "multi-rating table" 
(see Table 6-1) which shows a range of capacities for a 
particular fan size. For a given pressure, the highest mechani- 
cal efficiency usually will be in the middle third of the "CFM" 
column. Some manufacturers show the rating of maximum 
efficiency for each pressure by underscoring or similar indi- 
cator. In the absence of such a guide, the design engineer must 
calculate the efficiency from the efficiency equation 



Tl = 



QxFTP _Qx(FSP + VP outlet ) 



CFxPWR 



CFxPWR 



[6.4] 



where: 

r\ = Mechanical efficiency 
Q = Volumetric flow rate, cfm 
FTP = Fan total pressure, "wg 
FSP = Fan Static Pressure, "wg 
PWR = Power requirement, hp 
CF = Conversion Coefficient, 6362 

Even with a multi-rating table, it is usually necessary to 
interpolate in order to select fan RPM and BHP for the exact 
conditions desired. In many cases a double interpolation will 
be necessary. Straight line interpolations throughout the 
multi-rating table will introduce negligible errors. 

Certain types of fans may be offered in various Air Move- 
ment and Control Association^ 3) performance classes identi- 
fied as I through IV. A fan designated as meeting the 
requirements of a particular class must be physically capable 
of operating at any point within the performance limits for 
that class. Performance limits for each class are established in 
terms of outlet velocity and static pressure. Multi-rating tables 



6-16 



Industrial Ventilation 



TABLE 6-1. Example of Multi-Rating Table 

















^ 


r 






Inlet diameter: 13 


'O.D. 






Wheel diameter: 22%" 
















A 






Outlet 


area: 


.930 sq. ft, inside 






Wheel circumference: 5.92 ft. 






2" 


SP 


4"SP 


6"SP 


8"! 


5P 


10'SP 


12' 


SP 


14"SP 


16"SP 


18"SP 


20"SP 


22 "SP 


CFM 


ov 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM 


BHP 


RPM BHP 


RPM BHP 


RPM BHP 


930 
1 1 16 

1302 
1488 


1000 
1200 
1400 
1600 


843 
853 
866 
882 


0.57 
0.67 
0.77 
0.89 


1176 
1183 
1191 
1201 


1.21 
1.35 

1,51 
1.69 


1434 
1439 
1445 
1453 


1,93 
2.12 
2.33 
2.56 


1653 
1656 
1660 
1668 


2.75 
2.98 
3.22 
3.50 


1846 
1848 
1852 
1857 


3.64 
3.90 
4.20 
4,51 


2021 
2022 
2025 
2030 


4.59 
4.89 
5.23 
5.59 


2184 
2182 
2183 
2188 


5.62 
5.95 
6.31 

6.72 


2333 
2333 
2333 
2337 


6.68 
7.07 
7.47 
7.92 


2475 7.81 

2473 8.23 

2474 8.68 
2474 9,13 


2610 9.01 
2606 9.45 
2606 9.95 
2606 104 


2738 10.2 

2733 10.7 
2731 11.2 

2734 11.8 


1674 
1860 
2046 
2232 


1800 
2000 
2200 
2400 


899 
917 

937 
961 


1.01 
1.14 
1.29 
1.45 


1213 
1227 
1242 
1257 


1,88 
2,09 
2.32 
2.56 


1463 
1474 
1484 
1497 


2.81 
3.09 
3.37 
3.68 


1676 
1685 
1694 
1704 


3.81 
4.13 
4.48 
4.85 


1863 
1872 
1879 
1889 


4.86 

5.24 
5.63 
6.07 


2035 
2040 
2048 
2056 


5.98 
6.39 
6.84 
7.33 


2194 
2199 
2206 
2212 


7.16 
7.62 
8.13 
8.64 


2340 
2344 
2351 
2357 


8.38 
8.89 
9.43 
10.0 


2479 9.67 
2484 10.2 
2487 10.8 
2493 11.4 


2610 11.0 
2613 11.6 
2618 12.2 
2622 12.8 


2735 12.4 
2735 13.0 
2741 13.6 
2745 14.3 


2418 
2790 
3162 
3534 


2600 
3000 
3400 
3800 


984 
1038 
1099 
1164 


1.62 
2.02 
2.50 
3.07 


1275 
1313 
1358 
1407 


2.81 
3.36 
3.99 
4.69 


1513 
1543 
1580 
1620 


4.02 
4.73 
5.52 
6.37 


1717 

1744 
1775 
1812 


5.25 
6.11 
7.05 
8.09 


1900 
1924 
1952 
1984 


6.53 
7.52 
8.60 
9.79 


2065 
2088 
2115 
2144 


7,84 
8.96 
10.2 
11.5 


2222 
2241 
2265 
2290 


9.22 
10,4 
11.8 
13.3 


2364 
2383 
2405 
2428 


10.6 
12,0 
13,4 
15.0 


2501 12.1 
2517 13,5 
2538 15.1 
2562 16.8 


2631 13.6 
2644 15.1 
2665 16.8 
2684 18.6 


2750 15.1 
2766 16.7 
2783 18.5 
2803 20.5 


3906 
4278 
4650 
5022 
5394 


4200 
4600 
5000 
5400 
5800 


1232 
1306 
1380 
1457 
1535 


3.75 
4.56 

5.49 
6,56 
7,79 


1462 
1520 
1582 
1647 
1719 


5.48 
6.39 
7.41 
8.57 
9.93 


1665 
1717 
1770 
1827 
1885 


7.31 
8,38 
9,53 
10.8 
12.2 


1851 
1894 
1941 
1990 
2045 


9.19 
10.4 
11.7 
13.1 
14.7 


2018 
2058 
2100 
2146 
2194 


11.0 
12.4 
13.9 
15.5 
17.2 


2174 
2209 
2247 
2291 
2334 


12.9 
14.5 
16.1 
17.8 
19.7 


2320 
2355 
2390 
2428 
2469 


14.3 
16.5 
18.3 
20.2 
22.2 


2458 
2489 
2521 
2558 
2594 


16.8 
18.6 
20.5 
22.6 
24.7 


2587 18.7 
2614 20.6 
2645 22.7 
2681 25,0 
2717 27.3 


2708 20.6 
2736 22.7 
2766 25.0 
2798 27,3 
2830 29.8 


2825 22.5 
2852 24.8 
2883 27.3 



Performance shown is for fans with outlet ducts and with inlet ducts. BHP shown does not include belt drive losses. 



usually will be shaded to indicate the selection zones for 
various classes or will state the maximum operating RPM. 
This can be useful in selecting equipment, but class definition 
is only based on performance and will not indicate quality of 
construction. 

Capacity tables which attempt to show the ratings for a 
whole series of homologous fans on one sheet cannot be used 
accurately unless the desired rating happens to be listed on the 
chart. Interpolation is practically impossible since usually 
only one point of the fan curve for a given speed is defined in 
such a table. 



CO 

in 



cr. 




FLOW RATE (Q) 

FIGURE 6-7. Typical fan performance curve 



Today, most fan manufacturers have "electronic catalogs" 
available. These catalogs are computer programs which can 
be used to calculate the correct fan speed and horsepower 
based on input data such as desired flow rate and fan static 
pressure or fan total pressure. Some electronic catalogs in- 
clude estimates of the affects of various fan accessories such 
as dampers and inlet boxes. 

6.3.3 Point of Operation: Fans are usually selected for 
operation at some fixed condition or single "Point of Opera- 
tion." Both the fan and the system have variable performance 
characteristics which can be represented graphically as curves 
depicting an array of operating points. The actual "point of 
operation" will be the one single point at the intersection of 
the fan curve and the system curve. 

Fan Performance Curves: Certain fan performance vari- 
ables are usually related to volumetric flow rate in graphic 
form to represent a fan performance curve. Figure 6-7 is a 
typical representation where Pressure (P) and power require- 
ment (PWR) are plotted against flow rate (Q). Other variables 
also may be included and more detailed curves representing 
various fan designs are provided in Figure 6-4. Pressure can 
be either FSP or FTP. This depends on the manufacturer's 
method of rating. 

It should be noted that a fan performance curve is always 
specific to a fan of given size operating at a single rotation 
rate (RPM). Even with size and rotation rate fixed, it should 
be obvious that pressure and power requirements vary over a 
range of flow rates. 

System Requirement Curves: The duct system pressure also 
varies with volumetric flow rate. Figure 6-8 illustrates the 
variation of pressure (P) with flow rate (Q) for three different 
situations. The turbulent flow condition is representative of 



Fans 



6-17 



FLOW rate: (0) 

TURBULENT FLOW 
,2 



AP 



C0 Z 



FLOW RATE (Q) 

LAMINAR FLOW 
AP - CO 



FLOW RATE (Q) 

CONSTANT HEAD 
A P = C 

FIGURE 6-8. System requirement curves 

duct losses and is most common. In this case, the pressure loss 
varies as the square of the flow rate. The laminar flow condi- 
tion is representative of the flow through low velocity filter 
media. Some wet collector designs operate at or close to a 
constant loss situation. 

The overall system curve results from the combined effects 
of the individual components. 

6.3 A Matching Fan Performance and System 
Requirement: a desired point of operation results from the 
process of designing a duct system and selecting a fan. Con- 



sidering the system requirement or fan performance curves 
individually, this desired point of operation has no special 
status relative to any other point of operation on the individual 
curve. Figure 6-9 depicts the four general conditions which 
can result from the system design fan selection process. 

There are a number of reasons why the system design, fan 
selection, fabrication, and installation process can result in 
operation at some point other than design. When this occurs, 
it may become necessary to alter the system physically which 
will change the system requirement curve and/or cause a 
change in the fan performance curve. Because the fan per- 
formance curve is not only peculiar to a given fan but specific 
to a given rotation rate (RPM), a change of rotation rate can 
be relatively simple if a belt drive arrangement has been used. 
The "Fan Laws" are useful when changes of fan performance 
are required. 

6.3.5 Fan Laws: Fan laws relate the performance vari- 
ables for any homologous series of fans. A homologous series 
represents a range of sizes where all dimensional variables 
between sizes are proportional. The performance variables 
involved are fan size (SIZE), rotation rate (RPM), gas density 
(p), flow rate (Q), pressure (P), power requirement (PWR), 
and efficiency (r|). Pressure (P) may be represented by total 
pressure (TP), static pressure (SP), velocity pressure (VP), fan 
static pressure (FSP), or fan total pressure (FTP). 

At the same relative point of operation on any two perform- 
ance curves in this homologous series, the efficiencies will be 
equal. The fan laws are mathematical expressions of these 
facts and establish the inter-relationship of the other variables. 
They predict the effect of changing size, speed, or gas density 
on capacity, pressure, and power requirement as follows: 



Q, =Q. 



SIZE, 



SIZE, J ^RPM, 



RPM, 



P 2 =P 1 



SIZE 2 
I SIZE, 



Vrpm 2 n2 

^RPM, 



PWR 2 - PWR. 



5 'RPM, W 



SIZE, 



(^ SIZE, J I^RPM, 



£2 



[6.5] 



[6.6] 



[6.7] 



As these expressions involve ratios of the variables, any 
convenient units may be employed so long as they are consis- 
tent. Size may be represented by any linear dimension since 
all must be proportional in homologous series. However, 
impeller diameter is the most commonly used dimension. 

6.3.6 The Effect of Changing Rotation Rate or Gas 
Density: In practice, these principles are normally applied to 
determine the effect of changing only one variable. Most often 
the fan laws are applied to a given fan size and may be 
expressed in the simplified versions which follow: 

• For changes of rotation rate: 



6-18 



Industrial Ventilation 



DESIRED 




FLOW RATE (Q) 
A. FAN AND SYSTEM MATCHED 



°y / — DESIRED 




FLOW RATE (Q) 
8. WRONG FAN. 



if) 
(/) 

(_L.) 

cr 
a 




WRONG 



_.OW RATE (Q) 

;tem. 



ACTUAL 



Z) 

(/) 
co 

L.iJ 
CL 




FLOW RATE (Q) 
D. BOTH FAN AND SYSTEM WRONC 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



ACTUAL VERSUS DESIRED 
POINT OF OPERATION 



DATE 



1-88 



•IGURE 



6-9 



Fans 



6-19 



Flow varies directly with rotation rate; pressure var- 
ies as the square of the rotation rate; and power varies 
as the cube of the rotation rate: 



Q 2 -Q 1 



^RPM 2 
V RPM 1 



^1 



PWR 9 



^RPM 2 
V RPM 1 



= PWRJ 



(rpm, 



RPf^ 



[6.8] 

[6.9] 

[6.10] 



For changes of gas density: 

Flow is not affected by a change in density; pressure 
and power vary directly with density: 



Q 2 = Qi 



P 2 = P, 



Pi 



PWRp - PWR 1 



[6.11] 
[6.12] 

[6.13] 



6.3. 7 Limitations on the Use of Fan La ws: These e x- 
pressions are equations which rely on the fact that the per- 
formance curves are homologous and that the ratios are for 
the same relative points of rating on each curve. Care must 
be exercised to apply the laws between the same relative 
points of rating. 

Figure 6-10 contains a typical representation of two ho- 
mologous fan performance curves, PQ] and PQ 2 . These could 



1 oo r 




20 40 60 

FLOW RATE (0) 

FIGURE 6-10. Homologous performance curves 



80 



100 



be the performances resulting from two different rotation 
rates, RPMj and RPM 2 . Assuming a point of rating indicated 
as A] on PQ] there is only one location on PQ 2 with the same 
relative point of rating and that is at A 2 . The A| and A 2 points 
of rating are related by the expression 



^A2 ~ ^A1 



Q 



A 2 



Q 



Mj 



[6-14] 



The equation can be used to identify every other point that 
would have the same relative point of rating as Ai and A 2 . 
The line passing through M A 2 , A|" and the origin locates all 
conditions with the same relative points of rating. These lines 
are more often called "system lines" or "system curves." As 
discussed in Section 6.3.3, there are a number of exceptions 
to the condition where system pressure varies as the square of 
flow rate. These lines representing the same relative points of 
rating are "system lines" or "system curves" for turbulent flow 
conditions only. 

Where turbulent flow conditions apply, it must be under- 
stood that the system curves or lines of relative points of rating 
represent a system having fixed physical characteristics. For 
example the, "B 2 - B] M line defines another system which has 
lower resistance to flow than the "A 2 - A]" system. 

Special care must be exercised when applying the fan laws 
in the following cases: 

1. Where any component of the system does not follow 
the "pressure varies as the square of the flow rate" rule. 

2. Where the system has been physically altered or for 
any other reason operates on a different system line. 

6.3.8 Fan Selection at Air Density Other Than 
Standard: As discussed in Section 6.3.6, fan performance is 
affected by changes in gas density. Variations in density due 
to normal fluctuations of ambient pressure, temperature, and 
humidity are small and need not be considered. Where tem- 
perature, humidity, elevation, pressure, gas composition or a 
combination of two or more cause density to vary by more 
than 5% from the standard 0.075 Ibm/ft 3 , corrections should 
be employed. 

Rating tables and performance curves as published by fan 
manufacturers are based on standard air. Performance vari- 
ables are always related to conditions at the fan inlet. Fan 
characteristics are such that volumetric flow rate (Q) is unaf- 
fected but pressure (P) and power (PWR) vary directly with 
changes in gas density. Therefore, the selection process re- 
quires that rating tables are entered with actual volumetric 
flow rate but with a corrected or equivalent pressure. 

The equivalent pressure is that pressure corresponding to 
standard density and is determined from Equation 6.12 as 
follows: 



6-20 



Industrial Ventilation 



p =p 

1 d 'a 



0.075 



where: 

P e = Equivalent Pressure 

P a = Actual Pressure 

p a = Actual density, lbm/ft 3 

The pressures (P e and P a ) can be either Fan Static Pressure 
or Fan Total Pressure in order to conform with the manufac- 
turer's rating method. 

The fan selected in this manner is to be operated at the 
rotation rate indicated in the rating table and actual volumetric 
flow rate is that indicated by the table. However, the pressure 
developed is not that indicated in the table but is the actual 
value. Likewise, the power requirement is not that of the table 
as it also varies directly with density. The actual power 
requirement can be determined from Equation 6.13 as fol- 
lows: 



PWR a = PWR t 



0.075 



where: 

PWR a = Actual Power Requirement 
PWR t = Power Requirement in Rating Table 
p a = Actual Density, lbm/ft 3 

Fan selection at non-standard density requires knowledge of 
the actual volumetric flow rate at the fan inlet, the actual 
pressure requirement (either FSP or FTP, depending on the 
rating table used) and the density of the gas at the fan inlet. 
The determination of these variables requires that the system 
design procedure consider the effect of density as discussed 
in Chapter 5. 



EXAMPLE 

Consider the system illustrated in Figure 6-11 where the 
heater causes a change in volumetric flow rate and density. 
For simplicity, assume the heater has no resistance to flow 
and that the sum of friction losses will equal FSP. Using the 
Multi-Rating Table, Table 6-1, select the rotation rate and 
determine power requirements for the optional fan locations 
ahead of or behind the heater. 

Location 1 : Fan ahead of the heater (side "A" to "B" in Figure 

6-11). 

Step 1 . Determine actual FSP 
FSP = 1 "wg + 3 "wg 

= 4 "wg at 0.075 lbm/ft 3 . 

Step 2a. Density at fan inlet is standard. Therefore, 
enter rating table with actual volumetric flow 
rate at fan inlet, 1 000 acfm, and FSP of 4 "wg. 

b. Interpolation from Table 6-1 results in: 

RPM = 1182rpm 

PWR=1.32bhp 

Step 3. The fan should be operated at 1 182 rpm and 
actual power requirement will be 1.32 bhp. 

Location 2: Fan behind the heater (side "B" to "C" in Figure 
6-11). 

Step 1 . Determine actual FSP 

FSP = 1"wg + 3 "wg (as in explanation) 
= 4"wg at 0.0375 lbm/ft 3 

Step 2a. Density at fan inlet is not standard and a pres- 
sure correction must be made (using Equation 
6.12) to determine equivalent FSP. 



A 



1000 ACFM 
— 70 T — 



— 0.075 LBS/FT — — 

1 "wg FRICTION LOSS <§> 70 T 
(given) 



H 
E 

A 
T 
E 
R 



B 




2000 ACFM 

— 600 r - 



0.0375 LBS/FT 



3 : 'wg FRICTION LOSS @ 600 "F 
(given) 



FIGURE 6-11. In-duct heater 



Fans 



6-21 



FSP P = FSP, 



0.075^ 



0.075 ^ 



= 4"wgl "'"'" |=8"wg 
J \ 0.0375 J y 



Now, enter rating table with actual volumetric 
flow rate at fan inlet, 2000 acfm, and equiva- 
lent FSP, 8 M wg. 

b. Interpolation from Table 6-1 results in: 

RPM = 1692 rpm 

PWR = 4.39 bhp 

Step 3a. The fan should be operated at 1692 rpm, but 
actual power requirements will be affected by 
the density and can be determined by using 
Equation 6.13. 



PWR^PWR, 



0.075 



-4.39 



0.0375 ^ 
0.075 J 



= 2.2 bhp 

b. It should also be noted that a measurement of 
FSP will result in the value of 4 M wg (actual) 
and not the equivalent value of 8"wg. 

It will be noted that, regardless of location, the fan will 
handle the same mass flow rate. Also, the actual resistance to 
flow is not affected by fan location. It may appear then that 
there is an error responsible for the differing power require- 
ments of 1.32 bhp versus 2.2 bhp. In fact, the fan must work 
harder at the lower density to move the same mass flow rate. 
This additional work results in a higher temperature rise in the 
air from fan inlet to outlet. A fan located ahead of the heater 
will require less power and may be quieter due to the lower 
rotational speed. 

6.3.9 Explosive or Flammable Materials: When convey- 
ing explosive or flammable materials, it is important to rec- 
ognize the potential for ignition of the gas stream. This may 
be from airborne material striking the impeller or by the 
physical movement of the impeller into the fan casing. 
AMCA (6]) and other associations offer guidelines for both the 
manufacturer and the user on ways to minimize this danger. 
These involve more permanent attachment of the impeller to 
the shaft and bearings and the use of buffer plates or spark- 
resistant alloy construction. Because no single type of con- 
struction fits all applications, it is imperative that both the 
manufacturer and the user are aware of the dangers involved 
and agree on the type of construction and degree of protection 
that is being proposed. 

NOTE: or many years, aluminum alloy impellers have 
been specified to minimize sparking if the impeller were 
to contact other steel parts. This is still accepted, but 
tests by the U. S. Bureau qfMines (62) and others have 
demonstrated that impact of aluminum with rusty steel 



creates a "Thermite" reaction and thus possible ignition 
hazards. Special care must be taken when aluminum 
alloys are used in the presence of steel. 

6.4 FAN INSTALLATION AND MAINTENANCE 

Fan rating tests for flow rate, static pressure, and power 
requirements are conducted under ideal conditions which 
include uniform straight air flow at the fan inlet and outlet. 
However, if in practice duct connections to the fan cause 
non-uniform air flow, fan performance and operating effi- 
ciency will be affected. Location and installation of the fan 
must consider the location of these duct components to mini- 
mize losses. If adverse connections must be used, appropriate 
compensation must be made in the system calculations. Once 
the system is installed and operating, routine inspection and 
maintenance will be required if the system is to continue to 
operate at original design levels. 

6.4.1 System Effect: System effect is defined as the esti- 
mated loss in fan performance from this non-uniform air flow. 
Figure 6- 14 illustrates deficient fan system performance. The 
system pressure losses have been determined accurately and 
a suitable fan selected for operation at Point L However, no 
allowance has been made for the effect of the system connec- 
tions on fan performance. The point of intersection between 
the resulting fan performance curve and the actual system 
curve is Point 3. The resulting flow rate will, therefore, be 
deficient by the difference from 1 to 3. To compensate for this 
system effect, it will be necessary to add a "system effect 
coefficient" to the calculated system pressure. This will be 
equal to the pressure difference between Points 1 and 2 and 
will have to be added to the calculated system pressure losses. 
The fan then will be selected for this higher pressure (Point 
2) but will operate at Point I due to loss in performance from 
system effects. 

One commonly neglected system effect is a duct elbow at the 
fan inlet. For example, consider the fan shown in Figure 6-16. 

This fan has a four-piece 90° round duct elbow immediately 
in front of the inlet. There are no turning vanes inside the duct. 
The required flow rate is 5000 cfm and the system pressure 
losses are 8"wg at standard conditions (0.075 lb/ft3). Select- 
ing a fan without the system effect, using Table 6-1 , would 
result in a fan speed of 1987 rpm and power consumption of 
13.02 hp. 

With the elbow at the inlet, the air flow into the fan inlet 
will be degraded. Such a change in the air flow requires use 
of a system effect coefficient to select a fan that overcomes 
the degradation in performance. The system effect coefficient 
is used to determine a correction value, in inches water gauge, 
to be added to the system pressure losses. 

In this example, the duct diameter is 24 M with a turning 
radius of 48". This is a radius-to-diameter (r/d) ratio of 2.0. In 
Figure 6-21, Item C, we find the system effect curve to use is 



6-22 



Industrial Ventilation 




FLOW RATE 

rW0 IDENTICAL FANS 
RECOMMENDED 




FLOW RATE 

rW0 DIFFERENT FANS 
SATISFACTORY 



'-'!/ 



6> 




%% 



MOTES: 

1. TO ESTABLISH COMBINED FAN CURVE, IMF 

COMBINED AIR FLOW RATE, Q, iS THE SUM 
OF INDIVIDUAL FAN AIR "LOW RATES AT 
PO'NTS OF EQUAL PRESSURE 

2. TO ESTABLISH SYSTEM CURVE, INCLUDE 
LOSSES IN INDIVIDUAL FAN CONNECTIONS. 

3. SYSTEM CURVE MUST INTERSECT COMBINED 

FAN CURVE OR HIGHER PRESSURE FAN 
MAY HANDLE MORE AIR ALONE. 



FLOW RATE 

TWO DIFFERENT FANS 
UNSATISFACTORY 

WHEN SYSTEM CURVE DOES NOT CROSS COMBINED FAN 
CURVE, OR CROSSES PROJECTED COMBINED CURVE 
BEFORE FAN B, FAN B WILL HANDLE MORE AIR THAN 
FANS A AND B IN PARALLEL. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



FANS 
PARALLEL OPERA TION 



DATE W-96 



FIGURE 



6-12 



Fans 



6-23 




FLOW RATE 

TWO IDENTICAL FANS 

RECOMMENDED FOR BEST EFFICIENCY 




FLOW RATE 

TWO DIFFERENT FANS 
SATISFACTORY 




FLOW RATE 

TWO DIFFERENT FANS 
UNSATISFACTORY 



NOTES: 



1. TO ESTABLISH COMBINED FAN CURVE, THE 

COMBINED TOTAL PRESSURE IS THE SUM 
OF INDIVIDUAL FAN PRESSURES AT EQUAL 
AIR FLOW RATES, LESS THE PRESSURE LOSS IN 
THE FAN CONNECTIONS. 

2. AIR FLOW RATE THROUGH EACH FAN WILL BE 

THE SAME, SINCE AIR IS CONSIDERED 
INCOMPRESSIBLE. 

3. SYSTEM CURVE MUST INTERSECT 
COMBINED FAN CURVE OR LARGE FLOW RATE 
FAN MAY HANDLE MORE AIR ALONE. 



WHEN SYSTEM CURVE DOES NOT INTERSECT 
COMBINED FAN CURVE, OR CROSSES PROJECTED 
COMBINED CURVE BEFORE FAN B CURVE, FAN B 
WILL MOVE MORE AIR THAN FAN A AND B IN 
SERIES. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DATE 






SERIES OPERA TION 



5-96 



I 



figure: 



6-13 



6-24 



Industrial Ventilation 



r 






or 

Q_ 

~Z. 

CD 
00 
Ld 
Q 




SYSTEM EFFECT LOSS 
FACTOR @ DESIGN FLOW RATE 



FAN CATALOG 
CURVE 



ACTUAL PERFORMANCE 
OF FAN BECAUSE OF 
"SYSTEM EFFECT" 



DESIGN FLOW RATE 



FIGURE 6-14. System effect factor 

"R. " To find the system effect correction value in inches water 
gauge, we use the fan inlet velocity with Figure 6-27. Since 
the duct area is 3. 1 42 ft 2 , the velocity is 1592 fpm (5000 cfm 
- 3 . 1 42 ft 2 = 1 592 fpm). From Figure 6-27 we get a correction 
value of 0.1.9 "wg. This 0.19" value is added to the fan static 
pressure when selecting the fan from the multi-rating table. 
Select the fan for a static pressure of 8.19 "wg. Interpolating 
in Table 6-1, we find a selection for 5000 cfm and 8.19 "wg 
at 2002 rpm and 13.25 hp. This selection for a fan with an 
elbow at the inlet will result in operation at 5000 cfm and 8 
"wg drawing 13.25 hp. 



Note: The system effect coefficient compensates for the 
affect on the fan of an irregular air stream. This system 
effect coefficient is taken in addition to the friction loss 
used to calculate the system loss (Figure 5-13.) 

Figure 6- 1 5 illustrates typical discharge conditions and the 
losses which may be anticipated. The magnitude of the change 
in system performance caused by elbows and other obstruc- 
tions placed too close to a fan inlet or outlet can be estimated 
for the conditions shown on Figures 6-1 8 through 6-25. 



/ 






6- ?.Z 



JSS- SI:>L 

CURE 6 - - i 8 




NO LOSS 



i I 



l 1-^ 



s I 



EVAS!': 
CALCULATE FROM 
LiGU^L 5-- 16 AND 
SEE SECT! ON 5. lb 



FIGURE 6-15. Fan discharge conditions 



Fans 



6-25 




R 1220 mm 



1— 6 1 m m 



FIGURE 6-16. Inlet elbow 



Addition to system static pressure is given by reference to 
lettered curves in all but Figure 6-23. The additional static 
pressure, in "wg, is determined by obtaining the appropriate 
system effect coefficient from Figure 6-26 or 6-27 and mul- 
tiplying it by the fan inlet or discharge velocity pressure. 

A vortex or spin of the air stream entering the fan inlet may 
be created by non-uniform flow conditions as illustrated in 
Figure 6-24. These conditions may be caused by a poor inlet 
box, multiple elbows or entries near the inlet, or by other 
spin-producing conditions. Since the variations resulting in 
inlet spin are many, no System Effect Coefficients are tabu- 
lated. Where a vortex or inlet spin cannot be avoided or is 
discovered at an existing fan inlet, the use of turning vanes, 
splitter sheets, or egg crate straighteners will reduce the effect. 

6.4.2 Inspection and Maintenance: Material accumula- 
tion or abrasive wear on an impeller can cause a fan to "go 
out of balance." This unbalance will cause fan vibration. This 
may result in damage to or failure of the fan impeller, housing, 
bearings, or pedestal. Periodic cleaning and rebalancing of 
fans operating in air streams handling high material concen- 
trations is recommended. 

Regular observation of fan vibration levels can detect prob- 
lems before they develop to a damaging amplitude (see Figure 
6-17). Modern maintenance equipment permits the inspector 
to record vibration spectra. Review of changes in these spectra 
taken over time can indicate specific areas of developing 
problems with bearings, balance, belts or motors. Electronic 
or computerized vibration monitors are available to mount on 
fans used in critical operations. These devices can be set up 
with automatic alarm functions and/or to provide continuous 
information about a unit's vibration level. 

It is not uncommon, during fan installation or motor/starter 
maintenance, for the fan impeller rotation direction to be 



inadvertently reversed. Since fans do move a fraction of their 
rated capacity when running backward, incorrect rotation 
often goes unnoticed in spite of less effective performance of 
the exhaust system. 

Scheduled inspection of fans is recommended. Items 
checked should include: 

1 . Bearings for proper operating temperature (lubricate 
them on the manufacturer's recommended schedule). 

2. Excessive vibration of bearings or housing. 

3. Belt drives for proper tension and minimum wear. 

4. Correct coupling or belt alignment. 

5. Fan impeller for proper alignment and rotation. 

6. Impeller free from excess wear or material accumula- 
tion. 

7. Tight fan hold-down bolts. 

8. Tight fan impeller set screws or bushings. 

9. Proper installation of safety guards. 

Standard lockout/tagout procedures should be observed 
when servicing fan equipment or its associated duct. The 
electrical supply must be shut off and locked out at a discon- 
nect near the fan. When opening access doors or reaching into 
the fan inlet or outlet, the fan must be mechanically locked 
out by blocking the impeller from rotating. A warning tag 
should be used when blocking a fan. Do not open an access 
door while the fan is operating or coasting down. 

BE SURE to remove any inserted obstructions used to block 
impeller rotation when servicing is complete. 

REFERENCES 

6.1. Air Movement and Control Association, Inc.: AMCA 
Publication 201-90, Fans and Systems. AMCA, Ar- 
lington Heights, IL( 1990). 

6.2. Gibson, N.; Lloyd, F.C.; Perry, G.R.: Fire Hazards in 
Chemical Plants from Friction Sparks Involving the 
Thermite Reaction. Symposium Series No. 25. Insn. 
Chem. Engrs., London (1968). 

6.3. Air Movement and Control Association, Inc.: AMCA 
Publication 99-86, Standards Handbook. AMCA, Ar- 
lington Heights, IL (1986). 

6.4. American Society of Heating, Refrigeration, and Air- 
Conditioning Engineers, Inc.: 1993 ASHRAE Hand- 
book, Fundamentals Volume. ASHRAE, Atlanta, GA 
(1993). 



6-26 



Industrial Ventilation 



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AMERICAN CONFERENCE 
OF GOVERNMENTAL 

T IV T T\ T T O r P "O T a r t t\/p t m i\ T t o n~i o 


MACHINERY VIBRATION 
SEVERITY CHART 


1 IJNDUo 1 x\l 


a Li . 


Li i 


Ij 


IJ 


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DATE i ~0-9 6 I FIGURE 6-i7 



Fans 



6-27 



CENTRIFUGAL 
FAN 



CUTOFF 



BL^ST AREA 




DISCHARGE DUCT 




100% EFFECTIVE DUCT LENGTH 



TO CALCULATE 100% EFFECTIVE DUCT LENGTH, ASSUME A MINIMUM OF 2-1/2 DUCT DIAMETERS FOR 2500 
FPM OR LESS. ADD 1 DUCT DIAMETER FOR EACH ADDITIONAL 1000 FPM. 

EXAMPLE: 5000 FPM = 5 EQUIVALENT DUCT DIAMETERS. IF THE DUCT IS RECTANGULAR WITH SIDE 
DIMENSIONS a AND b, THE EQUIVALENT DUCT DIAMETER IS EQUAL TO (4ab/n) - 5 







12% 


25% 


50% 


100% 




No 


Effective 


Effective 


Effective 


Effective 




Duct 


Duct 


Duct 


Duct 


Duct 


Pressure 
Recovery 


0% 


50% 


80% 


90% 


100% 


Blast Area 
Outlet Area 


System Effect Curve 


0.4 


P 


R-S 


U 


W 


— 


0.5 


P 


R-S 


U 


W 


— 


0.6 


R-S 


S-T 


U-V 


W-X 


— 


0.7 


S 


U 


W-X 


— 


__ 


0.8 


T-U 


V-W 


X 


— 


— 


0.9 


V-W 


w-x 


— 


— 


— 


1.0 


— 


— 


— 


— 


— 



DETERMINE SEF BY USING FIGURE 6-26 OR 6-27 



Reprinted from AMCA Publication 201-90, FANS AND SYSTEMS, by 
permission of the Air Movement and Control Association, lnc. (s 1J 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SYSTEM EFFECT CURVES 
FOR OUTLET DUCTS- 
CENTRIFUGAL FANS 



date 5-92 



I 



FIGURE 



6-18 



6-28 Industrial Ventilation 




AXIAL FAN 




§ 



100% EFFECTIVE DUCT LENGTH 



& 



TO CALCULATE 100% EFFECTIVE DUCT LENGTH, ASSUME A MINIMUM OF 2-1/2 DUCT 
DIAMETERS FOR 2500 FPM OR LESS. ADD 1 DUCT DIAMETER FOR EACH ADDITIONAL 
1000 FPM. 

EXAMPLE: 5000 FPM = 5 EQUIVALENT DUCT DIAMETERS 





No 
Duct 


12% 

Effective 

Duct 


25% 

Effective 

Duct 


50% 

Effective 

Duct 


100% 
Effective 
Duct 


Tubeaxlal Fan 


— 


— 


— 


— 


— 


Vaneaxlal Fan 


U 


V 


W 


— 


— 



DETERMINE SEF BY USING FIGURE 6-26 OR 6-27 



Reprinted from AMCA Publication 201 -90, FANS AND SYSTEMS, by 
permission of the Air Movement and Control Association, Inc.* 6 1) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SYSTEM EFFECT CURVES 
FOR OUTLET DUCTS- 
AXIAL FANS 



DATE 



5-92 



FIGURE 



6-19 



Fans 



6-29 




POSITION A 



DETERMINE SEF BY USING FIGURES 6-26 AND 6-18 



For DWDi fans determine SEF using the curve for 
SWSI fans. Then apply the appropriate multiplier 
from the tabulation below 

MULTIPLIERS FOR DWDI FANS 
ELBOW POSITION A = AP X 1.00 
ELBOW POSITION B = AP X 1.25 
ELBOW POSITION C = AP X 1.00 
ELBOW POSITION D = AP X 0.85 



Reprinted from AMCA Publication 201 -90, FANS AND SYSTEMS, by 
permission of the Air Movement and Control Association, lnc. (61) 





Outlet 


No 


12% 


25% 


50% 


100% 


Blast Area 


Elbow 
Position 


Outlet 
Duct 


Effective 
Duct 


Effective 
Duct 


Effective 
Duct 


Effective 
Dud 


Outlet Area 




A 


N 





P-Q 


s 




0.4 


B 


M-N 


N 


O-P 


R-S 






C 


L-M 


M 


N 


Q 






D 


L-M 


M 


N 


Q 






A 


O-P 


P-Q 


R 


T 


0.5 


B 


N-O 


O-P 


Q 


S-T 






C 


M-N 


N 


O-P 


R-S 






D , 


M-N 


N 


O-P 


R-S 


cc 
















A 


Q 


Q-R 


S 


U 


§ 


0.6 


B 


P 


Q 


R 


T 




C 


N-O 


O 


Q 


S 


2 




D 


N-O 


O 


Q 


S 






A 


R-S 


S 


T 


V 


07 


B 


Q-R 


R-S 


S-T 


u-v 


u_ 




C 


P 


Q 


R-S 


T 






D 


P 


Q 


R-S 


T 


2 

1- 
co 
> 




A 


S 


S-T 


T-U 


W 


0.8 


B 


R-S 


S 


T 


V 


CO 




C 


Q-R 


R 


S 


U-V 


o 




D 


Q-R 


R 


S 


U-V 


2 




A 


T 


T-U 


u-v 


w 


0.9 


B 


s 


S-T 


T-U 


w 






C 


R 


S 


S-T 


V 






D 


R 


S 


S-T 


V 






A 


T 


T-U 


u-v 


w 


1.0 


B 


S-T 


T 


u 


w 






C 


R^S 


S 


T 


V 






D 


R-S 


S 


T 


V 





AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SYSTEM EFFECT CURVES 

FOR OUTLET ELBOWS 
ON CENTRIFUGAL FANS 



DATE 



5-92 



FIGURE 



6-20 



6-30 



Industrial Ventilation 



LENGTH - 
or DUCT 



r "\ 



Ly 















SYSTEM EFFECT FACTORS 


R / 


NO 2D 50 
DUCT DUCT DUCT 


~" 


N P R-S 



A. TWO-PIECE MiTEEED 90" ROUND SECTION ELBOW - NOT VANED 



LENGTH 
Oi r DUC1 



r~\ 



SYSTEM EFFECT FACTORS 



L_/ 



yj- 



% 


NO 


20 


5D 


DUCT 


DUCT 


DUCT 


0.5 





Q 


S 


0.75 


rj 


R-S 


I-U 


1.0 


R 


S--T 


u-v 


2.0 


R-S 


T 


u V 


3.0 


S 


T--U 


V 



B. THREE PIECE METERED 90' ROUND SECTION ELBOW -- NOT VANED 




SYSTEM EFFECT FACTORS 



R / 

7 D 


NO 


2D 


5D 


DUCT 


DUCT 


DUCT 


0.5 


P-Q 


R-S 


T 


0.75 


Q-R 


S 


U 


1.0 


R 


S--T 


U-V 


2.0 


R-S 


T 


u-v 


3.0 


S-T 


U 


v-w 



FOUR OR MORE PIECE METERED 90' ROUND SECTION ELBOW -- NOT VANED 



D = Diarnefer of the inlet collar. 

The inside area of the square duct (H X H) should be equal to the inside area of the fan inlet collar. 

+ The maximum permissible angle of any converging element of the transistion is 150 and for a diverging element 70 



Reprinted from AMCA Publication 201-76. FANS AND SYSTEMS, by 

permission of the Air movement and Control Association, Inc. (6.1) 



ERICAN CONFERENCE 
OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SYSTEM EFFECT CURVES FOR 
VARIOUS MITERED ELBOWS 
WITHOUT TURNING VANES 



DATE 



1- 



F1GURE 



6-21 



Fans 



6-31 



LENGTH 
OF DUCT 



r-\ t 




SYSTEM EFFECT FACTORS 



R / 


NO 


2D 


5D 


DUCT 


DUCT 


DUCT 


0.5 





Q 


S 


0.75 


p 


R 


S T 


1.0 


R 


S-T 


u-v 


2.0 


S 


T»U 


V 



A. SQUARE ELBOW WITH INLET TRANSITION -- NO TURNING VANES, 



LENGTH . 
OF DUCT 



i — v r 
l ^ 




SYSTEM EFFECT FACTORS 



X 


NO 


2D 


50 


DUCT 


DUCT 


DUCT 


0.5 


S 


T-U 


V 


1.0 


T 


u-v 


W 


2.0 


V 


v-w 


w-x 



B. SQUARE ELBOW WITH INLET TRANSITION — 3 LONG TURNING VANES. 



LENGTH 
OF DUCT 




SYSTEM EFFECT FACTORS 



R / 

7 D 


NO 


2D 


50 


DUCT 


DUCT 


DUCT 


0.5 


S 


T-U 


V 


1.0 


7 


u-v 


W 


2.0 


V 


v-w 


w X 



C. SQUARE ELEOW WITH INLET TRANSITION --■ SHORT TURNING VANES. 



D = 



vTF 



THE INSIDE AREA OF THE SQUARE DUCT (H X H) IS EQUAL TO THE INSIDE AREA CIRCUMSCRIBED BY THE FAN INLET COLLAR. 
THE MAXIMUM PERMISSIBLE ANGLE OF ANY CONVERGING ELEMENT OF THE TRANSITION IS 15", AND FOR A DIVERGING 
ELEMENT 7.5" 

Reprinted from AMCA Publication 201-76. FANS AND SYSTEMS, by 

permission of the Air movement and Control Association, Inc. (6.1) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SYSTEM EFFECT CURVES 

FOR OUTLET DUCTS - 

AXIAL FANS 



DATE 



5-92 



I 



FIGURE 



6-22 



6-32 



Industrial Ventilation 




A. NON-UNIFORM FLOW INTO A FAN INLET 
BY A 90* R OUND SECTION ELBOW - NO 
TURNING VANES. 



* Values shown are in modification of the 
original chart. 



THE REDUCTION IN FLOW RATE AND PRESSURE FOR 
THIS TYPE OF INLET CONDITION IS IMPOSSIBLE TO 
TABULATE. THE MANY POSSIBLE VARIATIONS IN 
WIDTH AND DEPTH OF THE DUCT INFLUENCE THE. 
REDUCTION IN PERFORMANCE TO VARYING DE- 
GREES AND THEREFORE THIS INLET SHOULD BE 
AVOIDED. FLOW RATE LOSSES AS HIGH AS 45% 
HAVE BEEN OBSERVED. EXISTING INSTALLATIONS 
CAN BE IMPROVED WITH GUIDE VANES OR THE 
CONVERSION TO SQUARE OR MITERED ELBOWS 
WITH GUIDE VANES. 



B. NON-UNIFORM FLOW INDUCED INTO FAN 
INLET BY A R ECT ANG ULAR INLET DUCT. 



Reprinted from AMCA Publication 201-76. FANS AND SYSTEMS, by 

permission of the Air movement and Control Association, Inc. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
I N D US TR I AL H YG I E N I S TS 



NON- UNIFORM INLET 
FLO 



DATE 



5-92 



FIGURE 



6 






Fans 6-33 



C. NON-UNIFORM FLOW INTO A FAN INlET BY AN INDUCED VORTEX, SPIN OR SWIRL. 



^y,^ 



TURNING 
VANES 




-V^K IMPELLER 

""""' ROTATION 



CORRECTED PRE- 
ROTATING SWIRL 




TURNING 
VANES 






I 



TURNING 
VANES 



IMPELLER 
ROTATION 



CORRECTED COUNTEF 
ROTATING SWIRL 



Reprinted from AMCA Publication 201-90 FANS AND SYSTEMS by 
permission of the Air Movement and Control Association Inc. (6.1) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



NON- UNIFORM INLET 
CORRECTIONS 



DATE 



5-92 



FIGURE 



6-24 



6-34 



Industrial Ventilation 




A. FREE INLET AREA PLANE 



INLET PLANE 
FAN WITH INLET COLLAR. 



■POINT OF TANGENT 
WITH FAN HOUSING SIDE 
AND INLET CONE RADIUS 





-INLET PLANE 
B. FREE INLET AREA PLANE — FAN WITHOUT INLET COLLAR. 



PERCENTAGE OF UNOBSTRUCTED 


SYSTEM EFFECT FACTORS 


INLET AREA 




100 


NO LOSS 


95 


0.26 


90 


0.40 


85 


0.53 


75 


0.8 


50 


1.6 


25 


2.0 



DETERMINE SEF BY CALCULATING INLET VELOCITY AND USING FIGURE 6-26 



Reprinted from AMCA Publication 210-90, FANS AND SYSTEMS, by 
permission of the Air and Control Association Inc. (6.1) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SYSTEM EFFECT CURVES 
FOR INLET OBSTRUCTIONS 



DATE 



5-92 



FIGURE 



6-25 



Fans 



6-35 



Loss Factor Equivalents 
for System Effect Curves® 



Curve 


' sys 


Curve 


•sys 


F 


16.0 


P 


1.98 


G 


14.3 


Q 


1.60 


H 


12.8 


R 


1.20 


1 


1 1.3 


S 


0.80 


J 


9.62 


T 


0.53 


K 


8.02 


U 


0.40 


L 


6.42 


V 


0.26 


M 


4.63 


W 


0.18 


N 


3.20 


X 


0.10 





2.51 







To use this table: 



1) Obtain the curve letter from Figures 
6-18 through 6-22 or Figure 6-25. 

2) For inlet system effects, multiply the 
equivalent loss coefficient from the above 
table by the fan inlet velocity pressure. 

3) For outlet system effects, multiply the 
equivalent loss coefficient from the above 
table by the fan outlet velocity pressure. 



*F sys values are in number of velocity pressures. For loss directly in "Wg, 
refer to Figure 6 — 27. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SYSTEM EFFECT CURVES 



DATE 



10-96 j FIGURE 6 : ~26 



6-36 



Industrial Ventilation 





5.0 




4.0 




3.0 




2.5 


o 


2,0 


^ 




tn 




UJ 




X 


1.5 


u 




f~ 




UJ 




01 




ZD 


1.0 


GO 

in 


0.9 


UJ 

01 


0.8 


Q_ 






0,7 




0.6 


Q< 




O 




O 


0.5 


< 




u_ 




1— 


0.4 


o 




LxJ 




U_ 




UJ 


0,3 


^ 




UJ 


0,25 


h- 




GO 




>~ 




GO 


0,2 



0.15 



0.1 



FGH1 J K L M NO 

















§ 


// 


// 


' / 


/, 














A 




// 


1 / 


// 


/ 














//// 


7/ 


/ 


/ / 


// 


/> 
















7/ 


/ 


// 


// 














777/ 


/ //7 / 




/ 


//, 


// 


' / 












1/ 

Wi 


7/ 




// 


i / 1 


/ 7 


/ 










/ 


' / , 


i 7 


/ J 


1 


i / 


7 


/ 








, ) 


1 


///// 


i 

/ 


7 7 


7 7 


/ 


7 > 


/ / 








, 


7 


7/7 




7 


7 / 


/ / 










/ 


'ik 


V. 


v// , 


I / 


/ / 




/ / 


/ t 








//, 


7 


/ 


/ / / 


/ / 








/ / 






7 




A 


/; 


/ / f 


/// 


7 / 












71 






/ 


/// 


// 










/ 




// 


A 




/ 


' /// 


7/ 


7 






/ / 


' 


y/ 


7 7 


/ 


/ 


/ 


//// 


7 / 
/ 


/: 


// 




1/ 


/ 


// 


7 




/ 


/ 

/ 

A 




/ / 


/! 


/ 


7 ' i 


/ 


7 


A 




7 


'/ 


/ 

7 


77 / 


// 


/ / 


/ 


/ 







p 

Q 



W 



6 7 8 9 10 



15 



20 25 30 



40 50 60 



AIR VELOCITY, FPM IN HUNDREDS 
(Air Density - 0.075 lbs/ft 3 ) 

*Enter the chart at the appropriate air velocity (on the abcissa) read up to the applicable 

curve, then across from the curve (to the ordinate) to find the SEP at standard 

air density. 

**Adapted for metric from AMCA Publication 201-90, FANS AND SYSTEMS, by permission 

of the Air Movement and Control Association, Inc. (6.1) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SYSTEM EFFECT CURVES 



DATE 6-92 



FIGURE 



6-27 



Chapter 7 

REPLACEMENT AND RECIRCULATED AIR 



7.1 INTRODUCTION 7-2 

7.2 REPLACEMENT AIR 7-2 

7.3 REPLACEMENT AIR DISTRIBUTION 7-4 

7.4 REPLACEMENT AIR FLOW RATE 7-5 

7.5 ROOM PRESSURE 7-5 

7.6 ENVIRONMENTAL CONTROL 7-5 

7.7 ENVIRONMENTAL CONTROL AIR FLOW RATE . 7-6 

7.8 AIR CHANGES 7-6 

7.9 AIR SUPPLY TEMPERATURES 7-6 

7.10 AIR SUPPLY VS. PLANT HEATING COSTS ....7-8 



Figure 7-1 
Figure 7-2 

Figure 7-3 

Figure 7-4 

Figure 7-5 
Figure 7-6 
Figure 7-7 
Figure 7-8 



Cold Zones Vs. Overheated Zones 7-2 

How Fan Performance Falls Off 

Under Negative Pressure 7-3 

Relationship Between Air Pressure and 

Amount of Force 7-4 

Throw Patterns and Distance From Different 

Register Adjustments 7-7 

Seasonal Air Ventilation 7-8 

Single Coil Steam Unit 7-9 

Steam Coil Piping 7-10 

Multiple Coil Steam Unit 7-11 



7.11 REPLACEMENT AIR HEATING EQUIPMENT. . . 7-9 

7.12 COST OF HEATING REPLACEMENT AIR 7-13 

7.13 AIR CONSERVATION 7-14 

7.13.1 Reduced Flow Rate 7-14 

7.13.2 Untempered Air Supply 7-16 

7.13.3 Energy Recovery 7-16 

7.13.4 Selection of Monitors 7-17 

7. 14 EVALUATION OF EMPLOYEE EXPOSURE 
LEVELS 7-1.8 

REFERENCES 7-20 

Figure 7-9 By-Pass Steam System 7-11 

Figure 7-10 Integral Face and By-Pass Coil 7-11 

Figure 7-11 Indirect Fired Unit 7-12 

Figure 7-12 Direct Fired Unit 7-12 

Figure 7-13 Direct Fired By-Pass Unit 7-13 

Figure 7-14 Recirculation Decision Logic 7-17 

Figure 7-15 Schematic Diagram of Recirculation 

Monitoring System 7-18 

Figure 7-16 Schematic of Recirculation From Air Cleaning 

Devices (Particulates) 7-19 



7-2 



Industrial Ventilation 



7.1 INTRODUCTION 

Chapters l through 6 describe the purpose, function, and 
design of industrial exhaust systems. As mentioned in Chapter 
I , Section 1 .2, supply systems are used for two basic purposes: 
to create a comfortable environment and to replace air ex- 
hausted from the building. It is important to note that while 
properly designed exhaust systems will remove toxic con- 
taminants, they should not be relied upon to draw outdoor air 
into the building. If the amount of replacement air supplied 
to the building is lower than the amount of air exhausted, the 
pressure in the building will be lower than atmospheric. This 
condition is called "negative pressure" and results in air 
entering the building in an uncontrolled manner through 
window sashes, doorways, and walls. In turn, this may lead 
to many undesirable results such as high velocity drafts, 
backdrafting, difficulty in opening doors, etc. 

To minimize these effects, design the mechanical supply 
systems to introduce sufficient outside air to avoid excessive 
negative or positive pressure conditions. A properly designed 
and installed air supply system can provide both replacement 
air and effective environmental control. Provided that impor- 
tant health and safety measures are taken, recirculation of the 
exhaust air may be an effective method that can substantially 
reduce heating and/or cooling costs. 

7.2 REPLACEMENT AIR 

Air will enter a building in an amount to equal the flow rate 
of exhaust air whether or not provision is made for this 
replacement. However, the actual exhaust flow rate will be 
less than the design value if the plant is under negative 
pressure. If the building perimeter is tightly sealed, thus 
blocking effective infiltration of outdoor air, a severe decrease 
of the exhaust flow rate will result. If, on the other hand, the 
building is relatively old with large sash areas, air infiltration 
may be quite pronounced and the exhaust system performance 
will decrease only slightly and other problems may occur. 



] EXHAUST 



rL 



o r~- 



COLD 

ZONE 



I 



© 



OVER HEATED 
ZONE 



COLD 
ZONE 



A 



< 

< 



Figure 7-1 . Under negative pressure conditions, workers in the cold zones 
turned up thermostats in an attempt to get heat. Because this did nothing 
to stop leakage of cold air, they remained cold while the center of plant was 
overheated. 



TABLE 7-1 . Negative Pressures and Corresponding Velocities 
Through Crack Openings (Calculated with air at room temperature, 
standard atmospheric pressure, C c = 0.6) 



Negative Pressure, "wg 



Velocity, fpm 



0.004 
0.008 
0.010 
0.014 
0.016 
0.018 
0.020 
0.025 
0.030 
0.040 
0.050 
0.060 
0.080 
0.100 
0.150 
0.200 
0,250 
0.300 
0.400 
0.500 
0.600 



150 

215 

240 

285 

300 

320 

340 

380 

415 

480 

540 

590 

680 

760 

930 

1080 

1200 

1310 

1520 

1700 

1860 



When the building is relatively open, the resultant in-plant 
environmental condition is often undesirable since the influx 
of cold outdoor air in the northern climates chills the perimeter 
of the building. Exposed workers are subjected to drafts, space 
temperatures are not uniform, and the building heating system 
is usually overtaxed (see Figure 7-1). Although the air may 
eventually be tempered to acceptable conditions by mixing as 
it moves to the building interior, this is an ineffective way of 
transferring heat to the air and usually results in fuel waste. 

Experience has shown that replacement air is necessary for 
the following reasons: 

1 . To insure that exhaust hoods operate properly. A lack 
of replacement air and the attendant negative pressure 
condition results in an increase in the static pressure 
the exhaust fans must overcome. This can cause a 
reduction in exhaust flow rate from all fans and is 
particularly serious with low-pressure fans such as 
wall fans and roof exhausters (see Figure 7-2). 

2. To eliminate high-velocity cross-drafts through win- 
dows and doors. Depending on the negative pressure 
created, cross drafts may be substantial (see Table 7- 1 ). 
Cross- drafts not only interfere with the proper opera- 



Replacement and Recirculated Air 7-3 



NEGATIVE PRESSURE IN BUILDING 




ORIGINAL SYSTEM 
PROPELLER FAN 



— LARGE FLOW LOSS 



NEGATIVE PRESSURE IN BUILDING 




ORIGINAL 
SYSTEM 

CENTRIFUGAL 
FAN 



SMALL FLOW LOSS 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HOW FAN PERFORMANCE FALLS 
OFF UNDER NEGATIVE PRESSURE 



DATE j _ QQ 



FIGURE 



7-2 



7-4 



Industrial Ventilation 



TABLE 7-2. Negative Pressures Which May Cause Unsatisfactory Conditions Within Buildings 



Negative Pressure "wg 



Adverse Conditions 



0.01 to 0.02 
0,01 to 0.05 

0.02 to 0.05 

0.03 to 0.10 

0.05 to 0.10 
0.10 to 0.25 



Worker Draft Complaints - High velocity drafts through doors and windows. 

Natural Draft Stacks Ineffective - Ventilation through roof exhaust ventilators, flow through 
stacks with natural draft greatly reduced. 

Carbon Monoxide Hazard - Back drafting will take place in hot water heaters, unit heaters, 
furnaces, and other combustion equipment not provided with induced draft. 

General Mechanical Ventilation Reduced - Air flows reduced in propeller fans and low 
pressure supply and exhaust systems. 

Doors Difficult to Open - Serious injury may result from non-checked, slamming doors. 

Local Exhaust Ventilation Impaired - Centrifugal fan exhaust flow reduced. 



6. 



tion of exhaust hoods, but also may disperse contami- 
nated air from one section of the building to another 
and can interfere with the proper operation of process 
equipment such as open top solvent degreasers. In the 
case of dusty operations, settled material may be dis- 
lodged from surfaces and result in recontamination of 
the workroom. 

To insure operation of natural draft stacks such as 
combustion flues. Moderate negative pressures can 
result in backdrafting of flues which may cause a 
dangerous health hazard from the release of combus- 
tion products, principally carbon monoxide, into the 
workroom. Back drafting may occur in natural draft 
stacks at negative pressures as low as 0.02 "wg (see 
Table 7-2). Secondary problems include difficulty in 
maintaining pilot lights in burners, poor operation of 
temperature controls, corrosion damage in stacks and 
heat exchangers due to condensation of water vapor in 
the flue gases. 

To eliminate cold drafts on workers. Drafts not only 
cause discomfort and reduce working efficiency but 
also may result in lower overall ambient temperatures. 

To eliminate differential pressure on doors. High dif- 
ferential pressures make doors difficult to open or shut 
and, in some instances, can cause personnel safety 
hazards when the doors move in an uncontrolled fash- 
ion (see Figure 7-3 and Table 7-2). 

To conserve fuel Without adequate replacement air, 
uncomfortable cold conditions near the building pe- 
rimeter frequently lead to the installation of more 
heating equipment in those areas in an attempt to 
correct the problem. These heaters take an excessive 
amount of time to warm the air and the over-heated air 
moving toward the building interior makes those areas 
uncomfortably warm (see Figure 7-1). This in turn 
leads to the installation of more exhaust fans to remove 
the excess heat, further aggravating the problem. Heat 
is wasted without curing the problem. The fuel con- 
sumption with a replacement air heating system usu- 



ally is lower than when attempts are made to achieve 
comfort without replacement air (see Section 7.10). 

7.3 REPLACEMENT AIR DISTRIBUTION 

Replacement air distribution is as critical as air volume 
(quantity) in industrial ventilation system design. Poor air distri- 
bution can destroy the control provided by a well-designed 
exhaust system. Non-turbulent air flow is particularly critical in 
indoor firing ranges, pharmaceutical plants, electronic compo- 
nents plants, some paint shops, and similar facilities. 

Designers often use the general heating, ventilating, and air 
conditioning system or plant air system to provide the supply 
(or replacement) air to replace air exhausted by local exhaust 
hoods. Unfortunately, plant air systems use high throw dif- 
fuses to mix the warmer/cooler air with the air already in the 
plant. Many times the throw distance is over 40 feet. This 
mixing effect causes turbulence near local exhaust systems. 
The local exhaust hoods must then be redesigned to draw in 
more air to control the contaminants and overcome the turbu- 
lence. This increases energy costs due to the need for larger 
fans and motors. Sometimes hoods cannot capture contami- 
nants and overcome turbulence. Hence, workers could still be 
overexposed even with a local exhaust system in place. There- 



1.00 
.75 
.50 

LU 

£* .25 

CO 

CO n 

uj O 

LK 

CL 

.25 



CK 
< 



.50 
.75 

1.00 



5.2 
5.9 
2.6 

1.3 



1.3 
- 2.6 
3.9 
5.2 



< 

O 

O 
CO 

cc 

LU 
(X 

CO 
Q 
Z 

O 
CL 



104 



or 78 

O 

O 

Q 52 

z: 

o 



26 




CT 
3 
CO 
CO 

LU 

cc 26 

Cl 

co S9 

? 

o 78 

CL 

104 









n 


b 


-8"x3' 
door 


-U 






20 sq 


ft 




O 






g 
G 



Figure 7-3. Relationship between air pressure and amount of force needed 
to open or close an average-sized door 



Replacement and Recirculated Air 



7-5 



fore, locate local exhaust hoods away from the turbulent 
effects of the plant air distribution system. 

If the supply air system does not sufficiently cool the 
employees, pedestal fans are often used. Pedestal fans also 
destroy contaminant control by causing turbulence near the 
local exhaust hood and should not be allowed. 

One method of providing non-turbulent air to the facility 
is to pass air through a supply air plenum built as part of the 
ceiling and/or through perforated duct. Cover the plenum face 
with perforated sheet metal. The ceiling plenum or duct runs 
should cover as large an area as possible to diffuse the air flow. 
A plenum wall providing cross-flow ventilation can be used 
only if the workers position themselves between the supply 
air system and the contaminant source and should not be used 
if the design velocity is over 100 fpm. 

Perforated drop-type ceilings work best in facilities with 
ceiling heights of less than 15 feet. Hoist tracks, lighting, and 
fire protection systems can be built into the ceiling. In some 
cases, fire protection will be required above and below the 
ceiling. Use perforated duct for ceilings over 15 feet. Perfo- 
rated duct manufacturers have computer programs to assist 
designers in determining duct sizes, shapes, and types as well 
as the location of pressure adjusting devices such as orifice 
plates and reducers. Air flow delivery in large bays may 
require supplemental air delivered at work stations to provide 
comfortable conditions for the workers. 

Feeding the air into the plenum is also critical. High veloc- 
ity flow into the plenum will cause the same turbulence prob- 
lems as the large d iff users commonly found in plants that 
attempt to throw the air to the floor. Consider feeding the 
plenum with perforated duct to diffuse the air inside the 
plenum. Another method of distributing air flow, from either 
a ceiling or wall-mounted plenum, is to design the plenum with 
two perforated plates, one fixed and one adjustable, located 
2-6 inches apart. Air flowing through slightly offset holes will 
encounter more resistance; thus air quantities passing through 
the low-flow areas will increase. The adjustable plates must be 
small enough to fine tune the air flow from the plenum. 

7.4 REPLACEMENT AIR FLOW RATE 

In most cases, replacement air flow rate should approxi- 
mate the total air flow rate of air removed from the building 
by exhaust ventilation systems, process systems and combus- 
tion processes. Determination of the actual flow rate of air 
removed usually requires an inventory of air exhaust locations 
and any necessary testing. When conducting the exhaust 
inventory, it is necessary not only to determine the quantity 
of air removed, but also the need for a particular piece of 
equipment. At the same time, reasonable projections should 
be made of the total plant exhaust requirements for the next 
one to two years, particularly if process changes or plant 
expansions are contemplated. In such cases it can be practical 
to purchase a replacement air unit slightly larger than immedi- 



ately necessary with the knowledge that the increased capac- 
ity will be required within a short time. The additional cost of 
a larger unit is relatively small and in most cases the fan drive 
can be regulated to supply only the desired quantity of air. 

Having established the minimum air supply quantity nec- 
essary for replacement air purposes, many plants have found 
that it is wise to provide additional supply air flow rate to 
overcome natural ventilation leakage and further minimize 
drafts at the perimeter of the building. 

7.5 ROOM PRESSURE 

While negative pressure can cause adverse conditions as 
described in Sections 7.1-7.3, there are situations where 
negative pressures are desired. An example is a room or area 
where a contaminant must be prevented from escaping into 
the surrounding area. It also may be desirable to maintain a 
room or area under positive pressure to maintain a clean 
environment. Either of these conditions can be achieved by 
setting and maintaining the proper exhaust/supply flow dif- 
ferential. Negative pressure can be achieved by setting the 
exhaust volumetric flow rate (Q) from the area to a level 
higher than the supply rate. A good performance standard for 
industrial processes is to set a negative pressure differential 
of 0.04 +/- 0.02 "wg. Conversely, positive pressure is 
achieved by setting the supply air flow rate higher than the 
exhaust rate. The proper flow differential will depend on the 
physical conditions of the area, but a general guide is to set a 
5% flow difference but no less than 50 cfrn. If the volume 
flows vary during either negatively or positively pressurized 
processes, it is easier to maintain the desired room pressure 
by adjusting the supply air. 

Some designers use transfer grilles and a pressure sensor 
in the room to maintain a desired room pressure. Air is allowed 
to seep from adjacent hallways, offices and other non-indus- 
trial areas. Do not use transfer grilles between areas where 
contaminant migration is possible. 

7.6 ENVIRONMENTAL CONTROL 

There are generally three types of industrial ventilation 
systems in most plants: 1) return air for the clean plant air; 2) 
a return air system where low level contaminants are diluted 
with fresh air (dilution ventilation); and 3) contaminant-laden 
air drawn through a local exhaust hood or ventilation system. 
In addition to toxic contaminants which are most effectively 
controlled by hoods, industrial processes may create an unde- 
sirable heat load in the work space. Modern automated ma- 
chining, conveying, and transferring equipment require 
considerable horsepower. Precision manufacturing and as- 
sembling demand increasingly higher light levels in the plant 
with correspondingly greater heat release. The resulting in- 
plant heat burden raises indoor temperatures, often beyond 
the limits of efficient and healthful working conditions and, 
in some cases, beyond the tolerance limits for the product. 



7-6 



Industrial Ventilation 



Environmental control of these factors can be accommo- 
dated through the careful use of the supply system. Industrial 
air conditioning may be required to maintain process specifi- 
cations and employee health . Many times the designer can use 
a setpoint higher than the 50-55 F used in conventional HVAC 
designs. ASHRAE gives basic criteria for industrial air con- 
ditioning in HVAC applications/ 7 9) (It must be noted that 
radiant heat cannot be controlled by ventilation and methods 
such as shielding, described in Chapter 2, are required.) 
Sensible and latent heat released by people and the process 
can be controlled to desired limits by proper use of ventilation. 

The HVAC industry uses automated building control and 
direct digital control (DDC) in many facilities. The technol- 
ogy can be applied to industrial ventilation with careful plan- 
ning. DDC uses computers and microprocessors tied to 
sensors and actuators to form a feedback and control system. 
DDC can be useful in industrial ventilation systems to control 
temperature, humidity, and relative room pressures. DDC 
systems can also track the system performance at hoods, fans, 
heating and cooling, and air pollution control equipment. 
DDC is especially useful in preventive maintenance. How- 
ever, DDC systems for industrial ventilation systems are 
complicated. Many are "one-of-a-kind" systems designed by 
a controls manufacturer and they require trained personnel to 
operate. 

Many industrial processes release minor amounts of "nui- 
sance" contaminants which, at low concentrations, have no 
known health effects but which are unpleasant or disagreeable 
to the workers or harmful to the product. The desire to provide 
a clean working environment for both the people and the 
product often dictates controlled air flow between rooms or 
entire departments. Evaluate the air streams returned into the 
facility to determine if the air pollution control devices (e.g., 
filters, cyclones) provide sufficient cleaning to prevent em- 
ployee exposure to "nuisance" contaminants. In addition, 
systems with known contaminants require controls listed in 
Section 7.12 and 7.13. The facility must employ trained 
mechanics and support a preventive maintenance program to 
sufficiently protect the workers. 

7.7 ENVIRONMENTAL CONTROL AIR FLOW RATE 

The design supply air flow rate depends on several factors 
including the health and comfort requirements. Sensible heat 
can be removed through simple air dilution (see Chapter 2 
under ventilation). 

"Nuisance" or undesirable contaminants can also be re- 
duced by dilution with outdoor air. The control of odors from 
people at various conditions of rest and work can be accom- 
plished with the outdoor air flow rate described in Chapter 2. 
However, these data apply mainly to offices, schools and 
similar types of environment and do not correspond well with 
the usual industrial or commercial establishment. Experience 
shows that when the air supply is properly distributed to the 



TABLE 7-3. Air Exchanges Vs. Room Sizes 




Room Size 


Air changes/ 
Room ft 3 minute 


Air changes/ 
hour 


40x40x12 high 
40 x 40 x 20 high 


19,200 11,650/19,200 = 0.61 
32,000 11,650/32,000-0.364 


36 
22 



working level (i.e., in the lower 8-10 ft of the space), outdoor 
air supply of 1-2 cfm/ft 2 of floor space will give good results. 
Specific quantities of outdoor air must be obtained from 
criteria developed by groups such as ASHRAE. 

7.8 AIR CHANGES 

"Number of air changes per minute or per hour" is the ratio 
of the ventilation rate (per minute or per hour) to the room 
volume. "Air changes per hour" or "air changes per minute" 
is a poor basis for ventilation criteria where environmental 
control of hazards, heat, and/or odors is required. The required 
ventilation depends on the problem, not on the size of the 
room in which it occurs. For example, let us assume a situation 
where 1 1 ,650 cfm would be required to control solvent vapors 
by dilution. The operation may be conducted in either of two 
rooms, but in either case, 1 1,650 cfm is the required ventila- 
tion. The "air changes," however, would be quite different for 
the two rooms. As can be seen in Table 7-3, for the same "air 
change" rate, a high ceiling space will require more ventilation 
than a low ceiling space of the same floor area. Thus, there is 
little relationship between "air changes" and the required 
contaminant control. 

The "air change" basis for ventilation does have some 
applicability for relatively standard situations such as office 
buildings and school rooms where a standard ventilation rate 
is reasonable. It is easily understood and reduces the engineer- 
ing effort required to establish a design criteria for ventilation. 
It is this ease of application, in fact, which often leads to lack 
of investigation of the real engineering parameters involved 
and correspondingly poor results. 

7.9 AIR SUPPLY TEMPERATURES 

Supply air temperature is controlled by the demand for 
heating and cooling. Factors to consider in maintaining a 
comfortable work environment for occupants are: setpoint 
temperature, humidity control, air distribution, and air flow 
rate. Where high internal heat loads are to be controlled, 
however, the temperature of the air supply can be appreciably 
below that of the space by reducing the amount of heat 
supplied to the air during the winter months and by deliber- 
ately cooling the air in the summer. When a large air flow rate 
is delivered at approximately space temperatures or somewhat 
below, the distribution of the air becomes vitally important in 
order to maintain satisfactory environmental conditions for 
the persons in the space. 

Maximum utilization of the supply air is achieved when the 



Replacement and Recirculated Air 



7-7 



air is distributed in the "living zone" of the space, below the 
8-10 foot level (see Figure 7-4). When delivered in this 
manner — where the majority of the people and processes are 
located — maximum ventilation results with minimum air 
handling. During the warm months of the year, large air flow 
in the working space at relatively high velocities is welcomed 
by the workers. During the winter months, however, care must 
be taken to insure that air velocities over the person, except 
when extremely high heat loads are involved, are kept within 
acceptable values (see Chapter 2, Table 2-5). To accomplish 
this, the air can be distributed uniformly in the space or where 
required for worker comfort. Heavy-duty, adjustable, direc- 
tional grilles and louvers have proven to be very successful in 
allowing individual workers to direct the air as needed. (71) 
Light gauge, stamped grilles intended for commercial use are 
not satisfactory. Suitable control must be provided to accom- 
modate seasonal and even daily requirements with a minim urn 
of supervision or maintenance attention. 

Chapter 2 describes the relative comfort that can be derived 
through adequate air flow control. Published tables of data by 
register and diffuser manufacturers indicate the amount of throw 
(projection) and spread that can be achieved with different 
designs at different flow rates (see Figure 7-4). Terminal veloci- 
ties at the throw distance can also be determined. 



Multiple point distribution is usually best since it provides 
uniformity of air delivery and minimizes the re-entrainment 
of contaminated air that occurs when large volumes are 
"dumped" at relatively high velocities. Depending on the size 
and shape of the space and the amount of air to be delivered, 
various distributional layouts are employed. Single point 
distribution can be used; however, it is usually necessary to 
redirect the large volume of air with a baffle or series of baffles 
in order to reduce the velocity close to the outlet and minimize 
re-entrainment In determining the number and types of reg- 
isters or outlet points, it also is necessary to consider the effect 
of terminal air supply velocity on the performance of local 
exhaust hoods. 

When large amounts of sensible heat are to be removed 
from the space during the winter months, it is most practical 
to plan for rapid mixing of the cooler air supply with the 
warmer air in the space. During the summer months, the best 
distribution usually involves minimum mixing so that the air 
supply will reach the worker at higher velocities and with a 
minimum of heat pickup. These results can be obtained by 
providing horizontal distribution of winter air over the 
worker's head, mixing before it reaches the work area and 
directing the air toward the worker through register adjust- 
ment for the summer months (see Figure 7-5). 



:i: 

o 



or: 

CO 



10' 



10 



0' 10' 20' 30' 40' 50' 60' 



_ r 


7 


T 








r / 


r J 


^ 


x - J 


/ / 




/ 


u 



"A" DEFLECTION 



-— CM 

— - CM 



10' 



0' 10' 20' 30' 40' 50' 



-^7""^ 


-7^ 


\7 ; 


V 


I i 


^ 


\../ 


-/_ 


/ 

u 


-- . ; 



"C" DEFLECTION 




DEFLECTION 




0' 10' 20' 







r^f 








{ 




L 


\ / 










^ 





"G" DEFLECTION 



PLAN VIEW 




UP PROJECTION HORIZONTAL PROJECTION 

SIDE VIEW 
FIGURE 7-4. Throw patterns and distance from different register adjustments (RER 7-2) 



DOWN PROJECTION 




7-8 



Industrial Ventilation 



Delivered air temperatures during the winter usually range 
from 65 F-68 F for work areas without much process heat or 
vigorous work requirement downward to 60 F or even 55 F 
where hard work or significant heat sources are involved. For 
summer operation, the temperature rise in indoor air can be 
estimated as described in Chapter 2. Evaporative cooling 
should be considered for summer operation. Although not as 
effective as mechanical refrigeration under all conditions, 
evaporative cooling significantly lowers the temperature of 



the outdoor air even in humid climates, improves the ability 
of the ventilation air to reduce heat stress, and costs much less 
to install and operate. 

7.10 AIR SUPPLY VS. PLANT HEATING COSTS 

Even if the supply air were drawn into the building simply 
by the action of the exhaust fans, during the winter months 
there will be an added burden on the plant heating system and 
fuel costs will rise. Experience has shown, however, that when 





10' APPROX. 



PULL CHAIN FOR 
FAST ADJUSTMENT 




SIDE WALL GRILLE 



WINTER - 



LOW AIR MOTION 
IN WORKING ZONE 



SUMMER 



HIGH AIR MOTION 
IN WORKING ZONE 






10' APPROX. 



PULL ROD FOR 
FAST ADJUSTMENT 




CEILING OUTLET 



WINTER - 



LOW AIR MOTION 
IN WORKING ZONE 



SUMMER 



HIGH AIR MOTION 
IN WORKING ZONE 



FIGURE 7-5. Seasonal air ventilation 



Replacement and Recirculated Air 7-9 



STEAM COIL 




'—FILTER SECTION 



FIGURE 7-6. Single coil steam unit 



the same flow rate of outdoor air is introduced through 
properly designed replacement air heaters, the overall fuel 
cost does not exceed previous levels and often is decreased. 
A partial explanation of this savings is more efficient heat 
transfer. The most important factor, however, is that a well- 
designed air supply system is not dependent on the plant space 
heating system; rather, the two systems operate in an inde- 
pendent fashion. The air supply system and the plant heating 
system can be understood best by considering the building as 
a whole. In order for an equilibrium to be established, the heat 
outflow from the building must balance the heat inflow. To 
obtain additional energy saving during downtime, design the 
supply system to provide sufficient heating to counter air enter- 
ing the building through infiltration and to prevent freezing. 

7.11 REPLACEMENT AIR HEATING EQUIPMENT 

Replacement air heaters are usually designed to supply 
100% outdoor air. The basic requirements for an air heater are 
that it be capable of continuous operation, constant delivered 
air flow rate, and constant preselected discharge temperature. 
The heater must meet these requirements under varying con- 
ditions of service and accommodate outdoor air temperatures 
which vary as much as 40 F daily. Standard design heating 
and ventilating units are usually selected for mixed air appli- 
cations, i.e., partial outdoor air and partial recirculated air; it 
is rare that their construction and operating capabilities will 
meet the requirements of industry. Such units are applicable 
in commercial buildings and institutional facilities where the 
requirements are less severe and where mixed air service is 
more common. 

Air heaters are usually categorized according to the source 
of heat: steam and hot water units, indirect-fired gas and oil 
units, and direct-fired natural gas and Liquified Petroleum 
Gas (LPG) units. Each basic type is capable of meeting the 
first two requirements — constant operation and constant 
delivered air flow rate. Variations occur within each type in 
relation to the third requirement, that of constant preselected 
discharge temperature. One exception to this rule is the direct- 
fired air heater where the inherent design provides a wide 
range of temperature control. Each type of air heater has 
specific advantages and limitations which must be understood 
by the designer in making a selection. 

Steam coil units were probably the earliest air heaters 



applied to general industry as well as commercial and institu- 
tional buildings (see Figure 7-6). When properly designed, 
selected, and installed, they are reliable and safe. They require 
a reliable source of clean steam at dependable pressure. For 
this reason they are applied most widely in large installations; 
smaller industrial plants often do not provide a boiler or steam 
capacity for operating a steam air heater. Principal disadvan- 
tages of steam units are potential damage from freezing or 
water hammer in the coils, the complexity of controls when 
close temperature limits must be maintained, high cost, and 
excessive piping. 

Freezing and water hammer are the result of poor selection 
and installation and can be minimized through careful appli- 
cation. The coil must be sized to provide desired heat output 
at the available steam pressure and flow. The coil preferably 
should be of the steam distributing type with vertical tubes. 
The traps and return piping must be sized for the maximum 
condensate flow at minimum steam pressure plus a safety 
factor. Atmospheric vents must be provided to minimize the 
danger of a vacuum in the coil which would hold up the 
condensate. Finally, the condensate must never be lifted by 
steam pressure. The majority of freeze-up and water hammer 
problems relate to the steam modulating type of unit which 
relies on throttling of the steam supply to achieve temperature 
control. When throttling occurs, a vacuum can be created in 
the coil and unless adequate venting is provided, condensate 
will not drain and can freeze rapidly under the influence of 
cold outdoor air. Most freeze-ups occur when outdoor air is 
in the range of 20-30 F and the steam control valve is partially 
closed, rather than when the outdoor air is a minimum tem- 
perature and full steam supply is on (see Figure 7-7). 

"Safety" controls are often used to detect imminent danger 
from freeze-up. A thermostat in the condensate line or an 
extended bulb thermostat on the downstream side of the coil 
can be connected into the control circuit to shut the unit down 
when the temperature falls below a safe point. As an alternate, 
the thermostat can call for full steam flow to the coil with 
shutdown if a safe temperature is not maintained. An obvious 
disadvantage is that the plant air supply is reduced; if the 
building should be subjected to an appreciable negative pres- 
sure, unit freeze-up still may occur due to cold air leakage 
through the fresh air dampers. 

The throttling range of a single coil unit can be extended 
by using two valves: one valve is usually sized for about 
two-thirds the capacity and the other valve one- third. Through 
suitable control arrangements both valves will provide 100% 
steam flow when fully opened and various combinations will 
provide a wide range of temperature control. Controls are 
complex in this type of unit and care must be taken to insure 
that pressure drop through the two valve circuits is essentially 
equal so as to provide expected steam flow. 

Multiple coil steam units (Figure 7-8) and bypass designs 
(Figure 7-9) are available to extend the temperature control 



7-10 



Industrial Ventilation 




6 



STEAM COIL 



5' 



r 

18" 



7 



STEAM SUPPLY 

PROVIDE STEAM FROM A CLEAN SOURCE 

MAINTAIN CONSTANT PRESSURE WITH REDUCING VALVES IE REQUIRED 

PROVIDE TRAPPED DRIPS FOR SUPPLY LINES 

SIZE SUPPLY PIPING FOR FULL LOAD AT AVAILABLE PRESSURE 
STRAINER 

1/32" DIAMETER MINIMUM PERFORATIONS 
DRIP TRAP 

INVERTED BUCKET TRAP PREFERRED 
CONTROL VALVE 

SIZE FOR MAXIMUM STEAM FLOW 

MAXIMUM PRESSURE DROP EQUAL TO 50% INLET STEAM PRESSURE 
VACUUM BREAKER 

1/2" CHECK VALVE TO ATMOSPHERE 
ALTERNATE VACUUM BREAKER 
STEAM COIL 

A. SIZE FOR DESIGN CAPACITY AT INLET STEAM PRESSURE (SUPPLY- VALVE DROP) 

B. VERTICAL COILS PREFFERED 

C. HORIZONTAL COILS MUST BE PITCHED 1/4" PER FOOT TOWARD DRAIN. 
6' MAXIMUM LENGTH RECOMMEMDED 

CONDENSATE TRAP 

A. INVERTED BUCKET PREFERRED 

B. SIZE TRAP FOR THREE TIMES MAXIMUM CONDENSATE LOAD AT PRESSURE 
DROP EQUAL TO 50% INLET PRESSURE 

C. INDIVIDUAL TRAP FOR EACH COIL 
CONDENSATE RETURN 

ATMOSPHERIC DRAIN ONLY 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



STEAM COIL PIPING 



DATE 



1 - 74 



FIGURE 



ry ry 



Replacement and Recirculated Air 7-11 



REHEAT COIL 

■PREHEAT COIL 




STEAM COIL. 



FILTER SECTION 



FIGURE 7-8. Multiple coil steam unit 



range and help minimize freeze-up. With multiple coil units, 
the first coil (preheat) is usually sized to raise the air tempera- 
ture from the design outdoor temperature to at least 40 F. The 
coil is controlled with an on-off valve which will be fully open 
whenever the outdoor temperature is below 40 F. The second 
(reheat) coil is designed to raise the air temperature from 40 
F to the desired discharge condition. Temperature control will 
be satisfactory for most outdoor conditions, but overheating 
can occur when the outdoor air temperature approaches 40 F 
(39 F + the rise through the preheat coil can give temperatures 
of 79-89 F entering the reheat). Refined temperature control 
can be accomplished by using a second preheat coil to split 
the preheat load. 

Bypass units incorporate dampers to direct the air flow. 
When maximum temperature rise is required, all air is directed 
through the coil. As the outdoor temperature rises, more and 
more air is diverted through the bypass section until finally 
all air is bypassed. Controls are relatively simple. The princi- 
pal disadvantage is that the bypass is not always sized for full 
air flow at the same pressure drop as through the coil, thus 
(depending on the damper position) the unit may deliver 



RACE DAMPER 




BY-PASS DAMPERS ■ 
FIGURE 7-9. By-pass steam system 



EiLTER SECTION 



differing air flow rates. Damper air flow characteristics are 
also a factor. An additional concern is that in some units the 
air coming through the bypass and entering the fan compart- 
ment may have a nonuniform flow and/or temperature char- 
acteristic which will affect the fan's ability to deliver air. 

Another type of bypass design, called integral face and 
bypass (Figure 7-10), features alternating sections of coil and 
bypass. This design promotes more uniform mixing of the air 
stream, minimizes any nonuniform flow effect and, through 
carefully engineered damper design, permits minimum tem- 
perature pickup even at full steam flow and full bypass. 

Hot water is an acceptable heating medium for air heaters. 
As with steam, there must be a dependable source of water at 
predetermined temperatures for accurate sizing of the coil. 
Hot water units are less susceptible to freezing than steam 
because of the forced convection which insures that the cooler 
water can be positively removed from the coil. Practical 
difficulties and pumping requirements thus far have limited 



-10 



G3 CS] 

D5J Co] 
Do] CS3 



Do] \m 

Do ] [Q] 

Do] Pol 

~D3ED 

D3 to] 

En na 

Do] Co] 




:l 



+ 50 



Do] Do] 

cs cs 
MM 

ETIsT 
ca ed 

ESHZ3 

-Gala 

G2 EG 
G2 EG 



G2 GG 
G] EG 
G3 EG 




FIGURE 7-10. Integral face and by-pass coil (7,4) 



7-12 



Industrial Ventilation 



COMBUSTION 
VENT FAN 



COMBUSTION 
CHAMBER 




FRESH AIR 
"LOUVERS 



RETURN AIR A 
LOUVERS 



FIGURE 7-11. Indirect-fired unit 



the application of hot water to relatively small systems: for a 
100 F air temperature rise and an allowable 100 F water 
temperature drop, 1 GPM of water will provide heat for only 
450 cfm of air. This range can be extended with high tempera- 
ture hot water systems. 

Hybrid systems using an intermediate heat exchange fluid, 
such as ethylene glycol, have also been installed by industries 
with critical air supply problems and a desire to eliminate all 
freeze-up dangers. A primary steam system provides the 
necessary heat to a converter which supplies a secondary 
closed loop of the selected heat exchange fluid. The added 
equipment cost is at least partially offset by the less complex 
control system. 

Indirect-fired gas and oil units (Figure 7-11) are widely 
applied in small industrial and commercial applications. Eco- 
nomics appear to favor their use up to approximately 10,000 
cfm; above this size the capital cost of direct- fired air heaters 
is lower. Indirect-fired heaters incorporate a heat exchanger, 
commonly stainless steel, which effectively separates the 
incoming air stream from the products of combustion of the 
fuel being burned. Positive venting of combustion products is 
usually accomplished with induced draft fans. These precau- 
tions are taken to minimize interior corrosion damage from 
condensation in the heat exchanger due to the chilling effect 
of the incoming cold air stream. The indirect-fired air heater 
permits the use of oil as a heat source and room air recircula- 
tion is permitted with this type of unit since the air stream is 
separated from the products of combustion. A third major 
advantage is that this type of unit is economical in the smaller 
sizes and is widely applied as a "package" unit in small 
installations such as commercial kitchens and laundries. 

Temperature control, "turn-down ratio," is limited to about 
3:1 or 5:1 due to burner design limitations and the necessity 
to maintain minimum temperatures in the heat exchanger and 
flues. Temperature control can be extended through the use 
of a bypass system similar to that described for single coil 
steam air heaters. Bypass units of this design offer the same 



advantages and disadvantages as the steam bypass units. 

Another type of indirect-fired unit incorporates a rotating 
heat exchanger. Temperature control can be as high as 20:1. 

Direct-fired heaters wherein the fuel, natural or LPG gas, 
is burned directly in the air stream and the products of com- 
bustion are released in the air supply have been commercially 
available for some years (Figure 7-12). These units are eco- 
nomical to operate since all of the net heating value of the fuel 
is available to raise the temperature of the air resulting in a 
net heating efficiency approaching 90%. Commercially avail- 
able burner designs provide turndown ratios from approxi- 
mately 25:1 to as high as 45:1 and permit excellent 
temperature control. In sizes above 10,000 cfm, the units are 
relatively inexpensive on a cost per cfm basis; below this 
capacity, the costs of the additional combustion and safety 
controls weigh heavily against this design. A further disad- 
vantage is that governmental codes often prohibit the recircu- 
lation of room air across the burner. Controls on these units 
are designed to provide a positive proof of air flow before the 
burner can ignite, a timed pre-ignition purge to insure that any 
leakage gases will be removed from the housing, and con- 
stantly supervised flame operation which includes both flame 
controls and high temperature limits. 

Concerns are often expressed with respect to potentially 
toxic concentrations of carbon monoxide, oxides of nitrogen, 
aldehydes, and other contaminants produced by combustion 
and released into the supply air stream. Practical field evalu- 



PROEILE 




E1LTER SECTION 



FIGURE7-12. Direct-fired unit 



Replacement and Recirculated Air 7-13 



ations and detailed studies show that with aproperly operated, 
adequately maintained unit, carbon monoxide concentrations 
will not be expected to exceed 5 ppm and that oxides of 
nitrogen and aldehydes are well within acceptable limits/ 7 3) 

A variation of this unit, known as a bypass design, has 
gained acceptance in larger plants where there is a desire to 
circulate large air flows at all times (see Figure 7-13). In this 
design, controls are arranged to reduce the flow of outdoor air 
across the burner and permit the entry of room air into the fan 
compartment. In this way, the fan air flow rate remains 
constant and circulation in the space is maintained. It is 
important to note that the bypass air does not cross the burner 
— 100% outdoor air only is allowed to pass through the 
combustion zone. Controls are arranged to regulate outdoor 
air flow and also to insure that burner profile velocity remains 
within the 1 imits specified by the burner manufacturer, usually 
in the range of 2,000 to 3,000 fpm. This is accomplished by 
providing a variable profile which changes area as the damper 
positions change. 

Inasmuch as there are advantages and disadvantages to both 
direct- fired and indirect-fired replacement air heaters, a care- 
ful consideration of characteristics of each heater should be 
made. A comparison of the heaters is given in Table 7-4. 

7.12 COST OF HEATING REPLACEMENT AIR 

As noted above, the cost of heating replacement air is 
TABLE 7-4. Comparison of Heater Advantages and Disadvantages 



TABLE PROFILE DAMPER 




Vv 



RECIRCULATED AIR x ^- ADJUSTABLE RECIRCULATING DAMPERS 
FIGURE 7-13. Direct fired by-pass unit 

probably the most significant annual cost of a ventilation 
system. Newer processes requiring cooling during the process 
must also be evaluated. Occupant comfort is more important 
than saving a few dollars in energy costs. Recent indoor air 
quality studies quantify diminished productivity when work- 
ers are uncomfortable. In addition to the equipment first cost, 
local building codes, and environmental regulations, designer 
experience in utility incentives and operating costs are in- 
volved in purchasing decisions. 

The American Society of Heating, Refrigerating, and 
Air Conditioning Engineers (ASHRAE), the U. S. De- 
partment of Energy, and others develop formulae and 
computer programs to determine the life-cycle costs of 
various equipment. The formulae and programs should 
not be used to determine the annual utility bill. Instead, 



Advantages 



Disadvantages 



Direct-fired Unvented: 

1. Good turndown ration -8:1 in small sizes; 25:1 in large sizes. Better 
control; lower operating costs. 

2. No vent stack, flue, or chimney necessary. Can be located inside 
walls of building. 

3. Higher efficiency (90%). Lower operating costs. (Efficiency based on 
available sensible heat.) 

4. Can heat air over a wide temperature range. 

5. First cost lower in large size units. 



1 . Products of combustion in heater air stream (some CO2, CO, oxides 
of nitrogen, and water vapor present). 

2. First cost higher in small size units. 

3. May be limited in application by governmental regulations. Consult 
local ordinances. 

4. Extreme care must be exercised to prevent minute quantities of 
chlorinated or other hydrocarbons from entering air intake, or toxic 
products may be produced in heated air. 

5. Can be used only with natural gas or LPG. 

6. Burner must be tested to assure low CO and oxides of nitrogen 
content in air stream. 



Indirect Exchanger: 

1 . No products of combustion; outdoor air only is discharged into 
building. 

2. Allowable in all types of applications and buildings if provided with 
proper safety controls. 

3. Small quantities of chlorinated hydrocarbons will not normally break 
down on exchanger to form toxic products in heated air. 

4. Can be used with oil, LPG, and natural gas as fuel. 

5. First cost lower in small size units. 

6. Can be used for recirculation as well as replacement. 



1 . First cost higher in large size units. 

2. Turn down ratio is limited — 3:1 usual, maximum 5:1. 

3. Flue or chimney required. Can be located only where flue or chimney 
is available. 

4. Low efficiency (80%). Higher operating cost. 

5. Can heat air over a limited range of temperatures. 

6. Heat exchanger subject to severe corrosion condition, Needs to be 
checked periodically for leaks after a period of use. 

7. Difficult to adapt to all combustion air from outdoors unless roof or 
outdoor mounted. 



7-14 



industrial Ventilation 



they are useful tools to compare the costs of various 
options in providing heating and cooling to an industrial 
ventilation system. 

The following two equations may be used to estimate replace- 
ment air heating costs on an hourly and yearly basis. They are 
based on average usage schedules and typical weather conditions 
rather than worst case conditions and maximum usage. 

Since there is an allowance for the efficiency of the replace- 
ment air unit, these equations will tend to give a low result if 
air is allowed to enter by infiltration only. They are also based 
on normal temperatures and moisture ratios and standard 
atmospheric pressure of 14.7 (101.4 kPa). Due to the heavy 
nature of the work in many industrial facilities, supply air may 
be cooler than for an office setting. Table 7-5 gives equation 
values (N) for supply air delivered at 70 F and 65 F. The 
humidity ratio (W) is assumed to be 0.01 pounds of water per 
pound of dry air. 



QN 
C-, = Hourly cost -0.001 — c 

q 



C 2 ^Yearly cost - 
where: 



0.154(Q)(dg)(T)(c) 

q 



[7.1] 



[7.2] 



Q = air flow rate, cfm 

N = required heat, BTU/hr/1000 cfm (Table 7-5 and 

Table 7-7) 
T = operating time, hours/week 
q = available heat per unit of fuel (Table 7-6) 
dg = annual degree days (Table 7-7) 
c = cost of fuel, $/unit 

EXAMPLE PROBLEM 1 

Find the hourly and yearly cost of tempering 10,000 cfm 
of replacement air to 70 F in St. Louis, Missouri, using oil at 
$1.35/gallon. 

Average winter temperature = 31 F 

u , , 0.001 QN 
Hourly cost = c 



= 0.001 x10 4 x^5^Lx$ 1.35 = $5.32 
106,500 



Yearly cost ^°- 154 )( 1 °)( 6023)(4Q) x$1.35 

y 106,500 

= $4,700 (assuming 40 hr/week 
of operation) 

The yearly cost is more representative because both the 
length and severity of the heating season are taken into 
account. 



TABLE 7-5. Required Heat for Outside Air Temperatures 


Avg. Outside Air 
Temperature, F 


N, Required Heat, 
BTU7hr/1,000cfm 

@70F 


N, Required Heat, 
BTU/hr/1,000cfm 

@65F 





77,000 


71,500 


5 


71,500 


66,000 


10 


66,000 


60,500 


15 


60,500 


55,000 


20 


55,000 


49,500 


25 


49,500 


44,000 


30 


44,000 


38,500 


35 


38,500 


33,000 


40 


33,000 


27,500 


45 


27,500 


22,000 


50 


22,000 


16,500 


55 


16,500 


11,000 


60 


11,000 


5,500 


65 


5,500 


— 



NOTE: Sensible Heat Equation used: q = 1.1 (cfm) delta t. Humidity ratio is 
assumed to be 0.01 pounds of moisture per pound of dry air, 



7.1 3 AIR CONSERVATION 

The supply and exhaust of air represent both a capital cost 
for equipment and an operating cost which is often sizable in 
northern climates. Concerned designers, recognizing these 
cost and energy conservation needs, are unanimous in their 
desire for reduced ventilation rates. 

There are four methods by which the cost of heating and 
cooling a large flow of outdoor air can be reduced: 1) reduc- 
tion in the total flow of air handled, 2) delivery of untempered 
outdoor air to the space, 3) recovery of energy from the 
exhaust air, and 4) recovery of warm, uncontaminated air 
from processes. The successful application of these engineer- 
ing methods without reduction in health hazard control and 
without impairing the inplant environment requires careful 
consideration. 

7.13.1 Reduced Flow Rate: A reduction of total air flow 

TABLE 7-6. Available Heat per Unit of Fuel 



Fuel 


Btu Per Unit 


Efficiency % 


Available Btu 
Per Unit 


Coal 


12,000 Btu/lb 
142,000 Btu/gal 


50 


6,000 


Oil 




75 


106,500 


Gas 








Heat 

Exchanger 
Direct Fired 


1,000 Btu/ft 3 


80 
90 


800 
900 



Replacement and Recirculated Air 



7-15 



TABLE 7-7. Heating Degree Day Normals and Average Winter Temperatures 



City 


Albany 


Boston 


Chicago 


Cleveland 


Detroit 


Minneapolis 


NY 


Phila- 
delphia 


Pitts- 
burgh 


St. Louis 


Wash., 
DC 


AvgTemp (F) 
Dec-Feb 


24 


22.4 


25 


28 


25.9 


16 


33.2 


33.3 


29 


32.2 


33.4 


Discharge Air 

Temp (F) 


Heating Degree Days 


80 


11782 


10409 


10613 


11343 


10959 


13176 


9284 


9652 


10797 


8943 


8422 


79 


11425 


10049 


10277 


10982 


10605 


12826 


8937 


9300 


10436 


8624 


8089 


78 


11062 


9690 


9940 


10621 


10256 


12478 


8596 


8954 


10076 


8310 


7764 


77 


10709 


9242 


9610 


10265 


9914 


12135 


8265 


8619 


9726 


8003 


7446 


76 


10356 


8994 


9283 


9915 


9581 


11797 


7938 


8285 


9379 


7702 


7139 


75 


10009 


8652 


8972 


9570 


9247 


11475 


7620 


7959 


9036 


7413 


6835 


74 


9669 


8317 


8656 


9229 


8920 


11142 


7308 


7641 


8702 


7121 


6538 


73 


9333 


7790 


8349 


8898 


8599 


10816 


7004 


7328 


8372 


6839 


6250 


72 


9007 


7668 


8046 


8567 


8291 


10496 


6706 


7028 


8050 


6560 


5974 


71 


8682 


7354 


7750 


8248 


7981 


10180 


6421 


6728 


7740 


6289 


5703 


70 


8364 


7046 


7468 


7928 


7678 


9870 


6146 


6438 


7429 


6023 


5438 


69 


8256 


6749 


7183 


7617 


7383 


9567 


5871 


6158 


7127 


5767 


5179 


68 


7750 


6458 


6905 


7313 


7100 


9269 


5606 


5886 


6833 


5523 


4929 


67 


7452 


6175 


6635 


7016 


6816 


8975 


5349 


5618 


6546 


5277 


4690 


66 


7162 


5903 


6373 


6722 


6543 


8687 


5101 


5360 


6272 


5053 


4455 


65 


6881 


5633 


6122 


6445 


6278 


8410 


4858 


5109 


5997 


4822 


4229 


64 


6607 


5370 


5875 


6165 


6020 


8131 


4621 


4864 


5734 


4595 


4014 


63 


6340 


5118 


5638 


5897 


5772 


7858 


4394 


4628 


5483 


4379 


3798 


62 


3081 


4873 


5399 


5636 


5533 


7590 


4176 


4397 


5234 


4168 


3588 


61 


5829 


4643 


5164 


5381 


5290 


7339 


3957 


4172 


5006 


3963 


3383 


60 


5586 


4399 


4936 


5140 


5054 


7086 


3747 


3952 


4769 


3761 


3182 



rate handled can be accomplished by conducting a careful 
inventory of all exhaust and supply systems in the plant. 
Determine which are necessary, which can be replaced with 
more efficient systems or hood designs, and which systems 
may have been rendered obsolete by changes. 

Numerous hood designs presented in Chapter 10 are in- 
tended specifically to provide for adequate contaminant cap- 
ture at reduced air flow rates. For instance, the use of 
horizontal sliding sash in the laboratory hood can provide a 
30% saving in exhaust air flow rate without impairing cap- 
ture velocity. The use of a tailored hood design, such as the 
evaporation hood shown in VS-35-40, provides good contami- 
nant capture with far lower exhaust flow rates than would be 
required for atypical laboratory bench hood. Low Volume-High 
Velocity hoods and systems such as those illustrated in VS- 



40-01 through VS-40-20 are used for many portable hand tool 
and fixed machining operations and can provide contaminant 
capture at far lower air handling requirements. 

Throughout industry there are many applications of win- 
dow exhaust fans and power roof exhausters to remove heat 
or nuisance contaminants which would be captured more 
readily at the source with lower air flow rates. Many roof 
exhausters, as noted earlier, have been installed initially to 
combat problems which were really caused by a lack of re- 
placement air. When air supply and balanced ventilation con- 
ditions are established, their use no longer may be necessary. 

Good design often can apply proven principles of local 
exhaust capture and control to reduce air flow rates with 
improved contaminant control. 



7-16 



Industrial Ventilation 



7.13.2 Untempered Air Supply: In many industries util- 
izing hot processes, cold outdoor air is supplied untempered 
or moderately tempered to dissipate sensible heat loads on the 
workers and to provide effective temperature relief for work- 
ers exposed to radiant heat loads. The air required for large 
compressors, as well as for cooling tunnels in foundries, also 
can come directly from outside the plant and thus eliminate a 
load that is otherwise replaced with tempered air. 

7.13.3 Energy Recovery: Energy recovery from exhaust 
air can be considered in two aspects: 1) the use of heat 
exchange equipment to extract heat from the air stream before 
it is exhausted to the outside and 2) the return (recirculation) 
of cleaned air from industrial exhaust systems. Heat ex- 
changer application to industrial exhaust systems has been 
limited primarily by the ratio of installed cost to annual return. 

Heat Exchangers — Air-to-air heat exchangers have been 
used to reduce energy consumption. This is achieved by 
transferring waste energy from the exhaust to replacement air 
streams of a building or process. The methods and equipment 
used will depend on the characteristics of the air streams. 
Major categories of equipment include heat wheels, fixed 
plate exchangers, heat pipes, and run-around coils. 

A heat wheel is a revolving cylinder filled with an air 
permeable media. As the exhaust air passes through the media, 
heat is transferred to the media. Since the media rotates, the 
warm media transfers heat to the cooler replacement air. 
Special care is required to ensure that this transfer does not 
cause a transfer of contaminants. 

A fixed plate exchanger consists of intertwined tunnels of 
exhaust and replacement air separated by plates (or sometimes 
a combination of plates and fins). The warm exhaust air heats 
the plates which in turn heat the cool replacement air on the 
other side of the plate. This exchanger uses no transfer media 
other than the plate forming wall of the unit. 

A heat pipe, or thermo siphon, uses a pipe manifold with 
one end in the warm exhaust air stream and the other in the 
cool replacement air stream. The pipe contains a fluid which 
boils in the warm exhaust air stream extracting heat and 
condenses in the cool replacement air stream releasing heat. 
Thus the heat pipe operates in a closed loop evaporation/con- 
densation cycle. 

A run-around coil exchanger uses a pair of finned-tube 
coils. A fluid circulates through the coils extracting heat from 
the warm exhaust air releasing heat to the cool replacement 
air. An advantage of the run-around coil is that the exhaust 
and supply duct systems can be separated by a significant 
distance which results in a reduced potential for re-entry; 
usually less duct in the systems and usually less roof area 
occupied by the units. 

Several factors are important in the selection of the appro- 
priate heat exchanger. A partial list is as follows: 

1. The nature of the exhaust stream. A corrosive or dust 



laden stream may need to be precleaned. 

2. The need to isolate the contaminated exhaust stream 
from the clean replacement air stream. 

3. The temperature of the exhaust stream. Unless the hot 
air stream is well above the desired delivery tempera- 
ture of the replacement air stream and the exhaust air 
stream is at elevated temperatures whenever heat is 
demanded by the replacement air stream, additional 
heating capacity will be required. 

4. Space requirements. Space requirements for some heat 
exchangers can be very extensive, especially when the 
additional duct runs are considered. 

5. The nature of the air stream. Many exhaust air streams 
are corrosive or dirty and special construction materi- 
als may be required. 

6. The need for a by-pass. During failure mode or sum- 
mer conditions, a by-pass will be required. 

Recirculation of Air from Industrial Exhaust Systems'. 
Where large amounts of air are exhausted from a room or 
building in order to remove particulates, gases, fumes, or 
vapors, an equivalent amount of fresh tempered replacement 
air must be supplied to the room . If the amount of replacement 
air is large, the cost of energy to condition the air can be very 
high. Recirculation of the exhaust air after thorough cleaning 
is one method that can reduce the amount of energy con- 
sumed. Acceptance of such recirculating systems will depend 
on the degree of health hazard associated with the particular 
contaminant being exhausted as well as other safety, technical 
and economic factors. A logic diagram listing the factors that 
must be evaluated is provided in Figure 7-14. (76) 

Essentially this diagram states that recirculation may be 
permitted if the following conditions are met: 

1. The chemical, physical, and toxicological charac- 
teristics of the chemical agents in the air stream to be 
recirculated must be identified and evaluated. Exhaust 
air containing chemical agents whose toxicity is un- 
known or for which there is no established safe expo- 
sure level should not be recirculated. 

2. All governmental regulations regarding recirculation 
must be reviewed to determine whether it is restricted 
or prohibited for the recirculation system under re- 
view. 

3. The effect of a recirculation system malfunction must 
be considered. Recirculation should not be attempted 
if a malfunction could result in exposure levels that 
would cause worker health problems. Substances 
which can cause permanent damage or significant 
physiological harm from a short overexposure shall 
not be recirculated. 

4. The availability of a suitable air cleaner must be deter- 
mined. An air cleaning device capable of providing an 



I. INITIAL DECISION 



IDENTIFY 

RECIRCULATION 

SYSTEM 



EVALUATE 
CHEMICAL AGENT 



CARCINOGEN OR 
low safe: EXPOSURE 
LEVEL 



NON-CARCINOGEN 
SAFE EXPOSURE 
LEVEL 



ASSESS RESULT 
OF FAILURE 



NON-ACCEPTABLE 



ACCEPTABLE 



CLEANER/MONITOR 
AVAILABILITY AND 
SUITABILITY 



I 



NOT AVAILABLE 
OR SUITABLE 



AVAILABLE AND 
SUITABLE 



I. DESIGN AND ASSESSMENT 









DESIGN AND 










ACCESS SYSTEM 










1 






1 




1 


^- 


WILL NOT MEET SAFE 
EXPOSURE LEVEL 




WILL MEET SAFE 
EXPOSURE LEVEL 


- 


hU ,_ 




1 


1 




INO 1 


ANALYZE COST 






1 








1 




COST HIGH 


COST SUITABLE 


1 


1 


fNOl 

III. SYSTEM EVALUATIC 


)N 


1 






CONSTRUCT AND 
INSTALL SYSTEM 






1 
















EVALUATE SYSTEM 
1 




1 




1 




WILL NOT MEET SAFE 
EXPOSURE LEVEL 




WILL MEET SAFE 
EXPOSURE LEVEL 




1 




1 






CORRECT DESIGN 




OPERATE SYSTEM 



FIGURE 7-14. Recirculation decision logic 

effluent air stream contaminant concentration suffi- 
ciently low to achieve acceptable workplace concen- 
trations must be available. 

5. The effects of minor contaminants should be reviewed. 
For example, welding fumes can be effectively re- 
moved from an air stream with a fabric filter; however 
if the welding process produces oxides of nitrogen, 
recirculation could cause a concentration of these 
gases to reach an unacceptable level. 

6. Recirculation systems must incorporate a monitoring 



Replacement and Recirculated Air 7-17 



system that provides an accurate warning or signal 
capable of initiating corrective action or process shut- 
down before harmful concentrations of the recircu- 
lated chemical agents build up in the workplace. 
Monitoring may be accomplished by a number of meth- 
ods and must be determined by the type and hazard of the 
substance. Examples include area monitoring for nui- 
sance type substances and secondary high efficiency 
filter pressure drop and on-line monitors for more haz- 
ardous materials. 

While all system components are important, special con- 
sideration should be given to the monitor. The prime requi- 
sites are that the monitor be capable of sensing a system 
malfunction or failure and of providing a signal which will 
initiate an appropriate sequence of actions to assure that 
overexposure does not occur. The sophistication of the moni- 
toring system can vary widely. The type of monitor selected 
will depend on various parameters (i.e., location, nature of 
contaminant — including shape and size — and degree of 
automation). 

7.13.4 Selection of Monitors: The safe operation of a re- 
circulating system depends on the selection of the best moni- 
tor for a given system. Reference 7.7 describes four basic 
components of a complete monitoring system which includes 
signal transfer, detector/transducer, signal conditioner, and 
information processor. Figure 7-15 shows a schematic dia- 
gram of the system incorporating these four components. It is 
quite likely that commercially available monitors may not 
contain all of the above four components and may have to be 
custom engineered to the need. 

In addition, the contaminant must be collected from the air 
stream either as an extracted sample or in toto. If a sample is 
taken, it must be representative of the average conditions of 
the air stream. At normal duct velocities, turbulence assures 
perfect mixing so gas and vapor samples should be repre- 
sentative. For aerosols, however, the particle size discrimina- 
tion produced by the probe may bias the estimated 
concentration unless isokinetic conditions are achieved. 

The choice of detection methods depends on the measur- 
able chemical and physical properties of the contaminants in 
the air stream. Quantifying the collected contaminants is 
generally much easier for particulate aerosols than for gases, 
vapors or liquid aerosols. 

Particulates: Where the hazardous contaminant constitutes 
a large fraction of the total dust weights, filter samples may 
allow adequate estimation of concentration. Better, if the 
primary collector (e.g., bag filters, cartridge filters) allows 
very low penetration rates, it may be economical to use high 
efficiency filters as secondary filters. If the primary filter fails, 
the secondary filter not only will experience an easily meas- 
ured increase in pressure drop, but will filter the penetrating 
dust as well — earning this design the sobriquet, "safety 



7-18 



Industrial Ventilation 




(nformation 1 
|processor| 



SIGNAL 
COND, 



RECIRCULATING 

EXHAUST 

SYSTEM 



ALARM 



INDICATOR/ 
RECORDER 



HUMAN 
INTERPRETATION 



CONTROLLER 



AUTO 
RESPONSE 



MANUAL 
RESPONSE 



FIGURE 7-15. Schematic diagram of recirculation monitoring system 

monitor" systems (see Figure 7-16). 

Non-p articulates: Continuously detecting and quantifying 
vapor and gas samples reliably and accurately is a complex 
subject beyond the scope of this manual. 

Air Sampling Instruments for Evaluation of Atmospheric 
Contaminants, published by ACGIH, (7 8) describes and evalu- 
ates different air monitoring devices. The monitor in a recir- 
culating system must be capable of reliably monitoring 
continuously and unattended for an extended period of time. 
It must also be able to quickly and accurately sense a change 
in system performance and provide an appropriate warning if 
a preselected safety level is reached. In order to function 
properly, monitors must be extremely reliable and properly 
maintained. Monitors should be designed so that potential 
malfunctions are limited in number and can be detected easily 
by following recommended procedures. Required mainte- 
nance should be simple, infrequent, and of short duration. 

7.14 EVALUATION OF EMPLOYEE EXPOSURE LEVELS 

Under equilibrium conditions, the following equations may 
be used to determine the concentration of a contaminant 
permitted in the recirculation return air stream: 



C R = 



(!-ti)(C e 



-K R C M ) 



1-[(K r )(1-ti)] 



[7.3] 



where: 



C R = air cleaner discharge concentration after 
recirculation, mg/m 3 
t] = fractional air cleaner efficiency 

C E = local exhaust duct concentration before recircu- 
lation, mg/m 3 

K R = coefficient which represents a fraction of the 
recirculated exhaust stream that is composed 



of the recirculation return air (range to 1 .0) 
C M = replacement air concentration, mg/m 3 



CB=^-(C G -C M )(1-f) + (C -C M )f 



+ K B C R +(1-K B )(C M ) 



[7.4] 



where: 

C B = 8-hr TWA worker breathing zone concentration 

after recirculation, mg/m 3 
Q B = total ventilation air flow before recirculation 
Q A = total ventilation air flow after recirculation 
C G = general room concentration before recircula- 
tion, mg/m 3 

f = coefficient which represents the fraction of time 
the worker spends at the work station 

C = 8-hr TWA breathing zone concentration at work 
station before recirculation 

K B = fraction of worker's breathing zone air that is 
composed of recirculation return air (range to 
1.0) 

The coefficients K R , K B and f are dependent on the work 
station and the worker's position in relation to the source of 
the recirculation return air and the worker's position in rela- 
tion to the exhaust hood. The value of K R can range from to 
1 .0 where indicates no recirculation return air entering the 
hood and 1.0 indicates 100% recirculation air entering the 
hood. Similarly, the value of K B can range from to 1 .0 where 
indicates there is no recirculation return air in the breathing 
zone and 1.0 indicates that the breathing zone air is 100% 
recirculated return air. The coefficient "f ' varies from where 
the worker does not spend any time at the work station where 
the air is being recirculated to 1 .0 where the worker spends 



Replacement and Recirculated Air 7-19 




AUTOMATIC ACTIVATION . 
OF BY- PASS ON 
HIGH PRESSURE 



EXHAUST OUTLET 
DAMPER MOTOR 



DIRECT RECIRCULATED AIR 
AWAY FROM WORKERS 

ULTRA HIGH EFFICIENCY 
FILTER 



PRESSURE SWITCH 



ALARM 



MANOMETER 



MANOMETER ACROSS 
FABRIC COLLECTOR 



FIGURE 7-16. Schematic of recirculation from air cleaning devices (particulates) 



100% time at the work station. 

In many cases it will be difficult to attempt quantification 
of the values required for solution of these equations for an 
operation not yet in existence. Estimates based on various 
published and other available data for the same or similar 
operations may be useful. The final system must be tested to 
demonstrate that it meets design specifications. 

An example of use of Equations 7.3 and 7.4 and the effect 
of the various parameters is as follows: 

Consider a system with 10,000 cfm total ventilation before 
recirculation (Q B ) consisting of 5,000 cfm of general exhaust 
and 5,000 cfm of local exhaust. The local exhaust is recircu- 
lated resulting in 1 0,000 cfm after recirculation air flow consist- 
ing of 5,000 cfm recirculated and 5,000 cfm fresh air flow. 

Assume poor placement of the recirculation return (K R and 
K B = 1) and that the worker spends all his time at the work 
station (f = 1); the air cleaner efficiency (rj) = 0.95; exhaust 
duct concentration (C E ) - 500 ppm; general room concentra- 
tion (C G ) = 20 ppm; replacement air concentration (C M ) = 5 
ppm; work station (breathing zone) concentration before re- 
circulation (C ) "=35 ppm; and a contaminant TLV of 50 ppm. 

Equation 7.3 gives recirculation air return concentration: 

(1-0.95) (500-1x5) 9R . m 

R = = 2b.1 ppm 

1- [(1)(1-0.95)] 

Equation 7.4 gives the worker breathing zone concentra- 



tion: 



Q B 



C B =7f- (C G -C M ) (1-f) + (C -C M )f + K B C R 



(1-K B )(C M ) 



1 10 4 * 



10* 



(20-5) (1-1) + (35-5) (1) + (26.1) 



+ (1-1) (5) 

= 56.1 ppm 

Obviously, 56.1 ppm exceeds the TLV of 50 ppm and there- 
fore is unacceptable. 

In order to achieve lower concentrations (C B ), the system 
configuration must be redesigned so that only 50% of the 
recirculation return air reaches the work station. Thus, K R and 
K B are reduced to 0.5 . Substituting these new data in Equation 
7.4, the breathing zone concentration calculates as 45.3 ppm. 
This is lower than the TLV of 50 ppm and therefore accept- 
able. 

Several potential problems may exist in the design of 
recirculated air systems. Factors to be considered are: 

1. Recirculating systems should, whenever practicable, 
be designed to bypass to the outdoors, rather than to 



7-20 



Industrial Ventilation 



recirculate, when weather conditions permit. If a sys- 
tem is intended to conserve heat in winter months and 
if adequate window and door openings permit suffi- 
cient replacement air when open, the system can dis- 
charge outdoors in warm weather. In other situations 
where the work space is conditioned or where me- 
chanically supplied replacement air is required at all 
times, such continuous bypass operation would not be 
attractive. 

2. Wet collectors also act as humidifiers. Recirculation 
of humid air from such equipment can cause uncom- 
fortably high humidity and require auxiliary ventila- 
tion or some means must be used to prevent excess 
humidity. 

3. The exit concentration of typical collectors can vary 
with time. Design data and testing programs should 
consider all operational time periods. 

4. The layout and design of the recirculation duct should 
provide adequate mixing with other supply air and 
avoid uncomfortable drafts on workers or air currents 
which would upset the capture velocity of local ex- 
haust hoods. 

5. A secondary air cleaning system, as described in the 
example on particulate recirculation, is preferable to a 
monitoring device because it is usually more reliable 
and requires a less sophisticated degree of mainte- 
nance. 

6. Odors or nuisance value of contaminants should be 
considered as well as the official TLV values. In some 
areas, adequately cleaned recirculated air, provided by 
a system with safeguards, may be of better quality than 
the ambient outdoor air available for replacement air 
supply. 

7. Routine testing, maintenance procedures, and records 
should be developed for recirculating systems. 

8. Periodic testing of the workroom air should be pro- 
vided. 

9. An appropriate sign shall be displayed in a prominent 
place reading as follows: 



CAUTION 

AIR CONTAINING HAZARDOUS SUBSTANCES 
IS BEING CLEANED TO A SAFE LEVEL IN THIS 
EQUIPMENT AND RETURNED TO THE BUILD- 
ING. SIGNALS OR ALARMS INDICATE MAL- 
FUNCTIONS AND MUST RECEIVE IMMEDIATE 
ATTENTION: STOP RECIRCULATION, DIS- 
CHARGE THE AIR OUTSIDE, OR STOP THE 
PROCESS IMMEDIATELY. 

REFERENCES: 

7.1. American Industrial Hygiene Association: Heating 
and Cooling Man and Industry. AIHA, Akron, OH 
(1969). 

7.2. Hart and Cooley Manufacturing Co.: Bulletin E-6. 
Holland, MI. 

7.3. Hama, G.: How Safe Are Direct-Fired Makeup Units? 
Air Engineering, p. 22 (September 1962). 

7.4. National Fire Protection Association, 1 Batterymarch 
Park, P. O. Box 9101, Quincy, MA 02269-9101. 

7.5. American Society of Heating, Refrigeration and Air 
Conditioning Engineers: Heating Ventilating and Air 
Conditioning Guide. ASHRAE, Atlanta, GA (1963). 

7.6. Hughes, R.T.;Amendola, A.A.: Recirculating Exhaust 
Air: Guides, Design Parameters and Mathematical 
Modeling. Plant Engineering (March 18, 1982). 

7.7. National Institute for Occupational Safety and Health: 
The Recirculation of Industrial Exhaust Air — Sym- 
posium Proceedings. Pub. No. 78-1 41 Department of 
Health, Education and Welfare (NIOSH), Cincinnati, 
OH (1978). 

7.8. American Conference of Governmental Industrial Hy- 
gienists: Air Sampling Instruments for Evaluation of 
Atmospheric Contaminants, 8th Edition. ACGIH, 
Cincinnati, OH (1995). 

7.9. American Society of Heating, Ventilating and Air 
Conditioning Engineers: HVAC Application. 
ASHRAE, Atlanta, GA (1995). 



Chapter 8 

VENTILATION ASPECTS OF INDOOR AIR QUALITY 



8.1 

8.2 

8.3 



8.4 



INTRODUCTION 8-2 

DILUTION VENTILATION FOR INDOOR AIR 

QUALITY 8-2 

HVAC COMPONENTS AND SYSTEM TYPES . . 8-2 

8.3.1 Components 8-2 

8.3.2 Types of Systems 8-3 

HVAC COMPONENTS, FUNCTIONS AND 
MALFUNCTIONS 8-4 

8.4.1 Outdoor Air 8-4 

8.4.2 Dampers 8-5 

8.4.3 Air Cleaning 8-5 



8.4.4 Heating/Cooling Coils 8-6 

8.4.5 Fans 8-6 

8.4.6 Humidifier/Dehumidifier 8-6 

8.4.7 Supply Air Distribution 8-6 

8.4.8 Supply Air Diffuser 8-7 

8.4.9 Return Air Grilles 8-8 

8.4.10 Return Air 8-8 

8.4.11 Fan Coil Units 8-8 

8.5 HVAC COMPONENT SURVEY OUTLINE 8-8 

REFERENCES 8-10 



Fig 


8-1 


Fig 


8-2 


Fig 


8-3 


Fig- 


8-4 


Fig. 


8-5 


Fig. 


8-6 


Fig. 


8-7 


Fig- 


8-8 


Fig. 


8-9 



Single Duct Constant Volume With Reheat . . 8-11 
Single Duct Constant Volume With Bypass . . 8-12 
Single Duct Variable Air Volume With Reheat 8-13 
Single Duct Variable Air Volume With Induction 8-14 
Single Duct Variable Air Volume with Fan 

Powered Devices and Reheat 8-15 

Dual Duct Constant Air Volume 8-16 

Single Duct Constant Air Volume Multi-Zone . 8-17 

Air Coil Unit 8-18 

Zone Heat Pump System 8-19 



Fig. 


8-10 


Fig. 


8-11 


Fig. 


8-12 


Fig. 


8-13 


Fig. 


8-14 


Fig. 


8-15 


Fig. 


8-16 


Fig. 


8-17 


Fig. 


8-18 



Filter Efficiency Vs. Particle Size 8-20 

Dampers: Parallel and Opposed Blade 8-21 

Typical HVAC Air Filters 8-22 

Heating/Cooling Coils 8-23 

Terminology for Centrifugal Fan Components . 8-24 

HVAC Humidifier/Dehumidifier 8-25 

HVAC System Self-Contained and Equipment 

Room 8-26 

Air Supply Diffusers and Return Air Grilles . . 8-27 

Typical Partitioned Office Air Pattern 8-28 



8-2 



Industrial Ventilation 



8.1 INTRODUCTION 

There are two ventilation aspects which are major causes 
of the complaints noted in the vast majority of reported 
problems from all parts of this and other countries. They are 
complaints of unsatisfactory indoor air quality (which may be 
due to the lack of sufficient outdoor air for dilution of "nor- 
mal" indoor airborne contaminants) and the failure to deliver 
supply air properly to the occupied zones. 

Indoor air quality is defined as the overall quality of the 
indoor air and includes biological, chemical, and comfort 
factors. This chapter is designed to familiarize the reader with 
heating, ventilation, and air conditioning (HVAC) systems 
used in office and similar spaces. The individual components 
of a typical HVAC system are defined, and the operation of 
the more common types of HVAC systems found are dis- 
cussed. 

8.2 DILUTION VENTILATION FOR INDOOR AIR QUALITY 

The oil shortage and the resulting energy crisis of the late 
1960s and early 1970s is considered by some as the most 
significant cause of the current indoor air quality concern. In 
the past, when energy costs were relatively low, the design of 
heating, ventilation and air conditioning (HVAC) systems for 
buildings included the infiltration of outside air through 
doors, windows, and other sources. Also, up to 25% (8 - n out- 
door air was supplied by the system, in addition to the 
infiltration, for general ventilation purposes. The outdoor air 
had the effect of diluting the "normal" indoor contaminants 
to a very low level of concentration, which had little effect on 
the occupants. 

Since the energy crisis resulted in major increases in energy 
costs, an extensive effort was made to reduce the infiltration 
of outdoor air by constructing the building as airtight as 
possible. Outdoor air supplied by the HVAC system was 
reduced to a minimum and in some instances eliminated 
entirely. Airborne contaminants found in indoor environ- 
ments were present in extremely small quantities and had not 
been a health problem in the past due to the dilution effect of 
the outdoor air. New concepts of office design that utilize 
fabric partitions, particle board furniture, increased use of 
carpets, office copy machines, etc., have increased the poten- 
tial for indoor contaminants. As buildings became more en- 
ergy efficient, there was an increase in complaints of 
stuffiness, drowsiness, tiredness, eye irritation, throat irrita- 
tion, and stale air. 

Existing health standards are not usually violated by the 
low-level concentrations, and the only current legal require- 
ment for outdoor air is found in the building codes. The 
Uniform Building Code (8 2) is the most widely accepted stand- 
ard for providing outdoor air. Section 605 states that 5 cfm of 
outdoor air per occupant shall be mechanically supplied to all 
parts of the building during occupancy. Carbon dioxide con- 
centrations from occupant respiration within a space are often 



used as an indicator of the quantity of outdoor air being 
supplied to that space. When the indoor air concentration 
reaches approximately 800-1000 ppm (excluding external 
combustion sources), complaints may escalate. As the carbon 
dioxide levels increase, the number of complaints will in- 
crease more rapidly. 

In 1989, the American Society of Heating, Refrigeration 
and Air Conditioning Engineers (ASHRAE) developed and 
adopted ASHRAE Standard 62-1989, "Ventilation for Ac- 
ceptable Indoor Air Quality. M < 8 3) The standard recognized the 
health problems resulting from the changes in construction 
and HVAC methods. It is based on occupancy of spaces and 
provides the outdoor air requirement for that space. Require- 
ments for outdoor air for offices based on an occupancy of 
seven people per 1000 square feet is currently 20 cfm per 
person. This is based on a total occupancy, including tran- 
sients, and is in addition to the usual HVAC requirements. 
The standard is expected to satisfy the requirements for 80% 
or more of the occupants. 

Provision for delivery of the outdoor air for dilution of the 
normal indoor airborne contaminants in the occupied space is 
a major factor of indoor air quality considerations. It is obvi- 
ous that if the outdoor air included as part of the total supply 
air is not delivered to the occupied zone, the potential for 
unsatisfactory indoor air quality increases. Another important 
factor in the delivery of the air to the occupied zone is the 
location of the supply and return air grilles to avoid short-cir- 
cuiting. Ideally, the supply air diffusers and the air grilles are 
so located that a uniform flow of air through the space occurs 
to avoid both stagnant air and drafts. 

Temperature and humidity can play a role in how people 
perceive indoor environment. ASHRAE Standard 55-1 992 (8 4) 
provides guidance in design and maintenance of indoor ther- 
mal environments. ASHRAE recommends temperature 
ranges of 67 to 76 F in winter (heating season) and 72 to 81 
F in summer (cooling season). However, complaints may 
increase when temperatures rise above 74 F. Similarly, it is 
preferable to keep relative humidities above 20-30% during 
the heating season and below 60% during the cooling season. 
ASHRAE also suggests limits on air movement. The average 
air movement in an occupied space should not exceed 30 fpm 
in winter or 50 fpm in summer. 

8.3 HVAC COMPONENTS AND SYSTEM TYPES 

When considering the ventilation aspects of HVAC sys- 
tems, the type of system and its components should be re- 
viewed for potential sources or causes of complaints regarding 
indoor air quality. Detailed descriptions of the systems and 
components can be found in the Systems and Equipment 
volume of the ASHRAE Handbooks 5 > 

8.3.1 Components: The components that make up HVAC 
systems generally include the following: 



Ventilation Aspects of Indoor Air Quality 8-3 



A. HVAC System: HVAC system refers to the equipment 
and distribution system used for heating, ventilating, 
cooling, humidifying, dehumidifying, and cleansing 
air for a building or building zone for the purpose of 
comfort, safety, and health of the occupants. 

B. Dampers: Dampers are devices of various types used 
to vary the volume of air passing through an outlet, 
inlet, or duct. 

C . Outdoor A ir (Fresh A ir; Replacement A ir; Compen- 
sating Air): Outdoor air used to replace all or part of 
the air in a building or building space. 

D. Return Air: Return air is air that has been in the 
building for a period of time and is returned to the 
HVAC system. Varying percentages of return air are 
exhausted outdoors with the remaining air (recircu- 
lated air) mixed with outdoor air for conditioning and 
distribution. 

E. Mixing Plenum: A mixing plenum is a chamber 
within an HVAC system where outdoor air is mixed 
with returned air. The mixed air, after cleaning and 
conditioning, comprises the supply air for the building. 

F. Air Cleaners: Air cleaners are devices designed to 
remove atmospheric airborne impurities such as dusts, 
gases, vapors, fumes, and smoke. (Air cleaners include 
air washers, air filters, electrostatic precipitators and 
charcoal filters.) 

G. Heating Coils: Heating coils are heat transfer devices 
which utilize hot water, steam, or electricity to heat the 
supply air. 

H. Cooling Coils: Cooling coils are heat transfer devices 
which utilize chilled water or a refrigerant to cool the 
supply air. 

I. Condensate Pan (Drip Tray; Defrost Pan): A vessel 
or tray under the cooling coil to receive water extracted 
from the supply air by condensation from the cooling 
coil. 

J. Humidifier/Dehumidifier: Humidifier/dehumidifiers 
are devices to add/remove moisture to/from the supply 
air. 

K. Fans (Supply and Return): Fans are devices for mov- 
ing ventilation air through the HVAC system. 

L. Supply Air: Supply air is conditioned ventilation air 
delivered to zones within a building. 

M. Control Zone: Control zone is a space or group of 
spaces within a building served by an HVAC system. 
Depending on the space requirements, the control zone 
may be designated as core or interior zone and/or 
perimeter zone. 



N. Occupied Zone: The occupied zone is the region 
within an occupied space between 3 and 72 inches 
above the floor. 

O. Control Box (Variable Air Volume, Bypass, Dual 
Duct): Control boxes are devices to which the supply 
air may be delivered by the HVAC system prior to 
delivery to the supply diffuser. These boxes may in- 
clude means of controlling supply air temperature and 
volume to the diffuser or multiple diffusers within a 
HVAC zone. 

P. Supply Air Diffusers: Supply air diffusers are devices 
whose function is to deliver the supply air to the 
occupied zone and to provide a desired distribution 
pattern. The diffusers may be circular, square, rectan- 
gular, linear slots, louvered, fixed, adjustable, or a 
combination. 

Q. Return Air Grilles: Return air grilles may be louvered 
or perforated coverings for openings located in the 
sidewall, ceiling or floor of a zone through which the 
return air enters. The return air grilles may be directly 
connected to an open return air plenum or to a ducted 
return air system. 

R. Return Air Plenum: A return air plenum is the space 
usually located above the ceiling where the return air 
is collected from a zone prior to entering the return air 
system. 

S. Economizer: An economizer is a control system which 
reduces the heating and cooling load through the use 
of outdoor air for free cooling when the total heat of 
the return air exceeds the total heat of the outdoor air. 

8.3.2 Types of Systems: There are different types of 
HVAC systems: singie-duct systems, dual-duct systems, 
multi-zone systems, and special systems. These systems may 
be considered basic and subject to variations that are neces- 
sary to meet specific requirements. The following descrip- 
tions of the basic systems are intended as a guide and the 
referenced ASHRAE Handbook should be reviewed for sys- 
tem details and variations. 

Single-Duct Systems may be either a constant or a variable 
air volume system. The constant volume system maintains 
constant air flow with the temperature of the supply air 
controlled in response to the space load. See Figures 8-1 and 
8-2. The system may be a single zone, a zoned reheat, multi- 
ple-zone modification or a by-pass variation using a by-pass 
box in lieu of reheat constant volume primary system with a 
variable air volume secondary system. A variable air volume 
(VAV) system controls the temperature within a zone by 
varying the supply air volume. See Figures 8-3, 8-4, and 8-5. 
This type of system may include reheat at the terminals, 
induction unit, fan-powered distribution box, dual conduit, 
and variable diffusers. 



8-4 



Industrial Ventilation 



Dual-Duct Systems condition all the air in a central appa- 
ratus and distribute it to the conditioned zones through two 
parallel mains, one carrying cold air and the other warm air. 
The system may be a constant volume type single fan and with 
or without reheat capability. See Figure 8-6. Also, the system 
may be VAV which mixes the cold and warm air in various 
volume combinations depending on the zone load. In both 
system types, the cold and warm air is delivered to a dual duct 
box which mixes the air prior to delivery to the supply air 
diffuser. 

Multizone Systems supply several zones from a centrally 
located HVAC unit. Supply air for the different zones consists 
of mixed cold and warm air through zone dampers in the 
central HVAC unit in response to zone thermostat control. 
From there, the supply air is distributed through the building 
by single zone ducts which, in turn, supply the airto the zone 
diffusers. See Figure 8-7. 

Fan Coil Units are usually located along the outdoor wall 
of a building for heating and cooling the perimeter up to 15 
feet from the outdoor wall. These units may have a through- 
wail duct for outdoor air and can be totally self contained or 
have the heating and cooling media supplied from a central 
mechanical room. See Figure 8-8. Controls for temperature 
and operation will vary although control of the outdoor air is 
usually at the unit and beside the nearest occupant. 

Zone Heat Pumps are packaged HVAC units that may 
provide the heating and cooling for individual zones within a 
building. These units vary in how the heating and cooling 
media is provided, but the function is generally constant (see 
Figure 8-9). Also, these units may be located within the 
individual zone above the ceiling in the return air space or 
remotely such as on the building roof. The supply air is 
delivered to the entire zone through a duct distribution and 
diffuser system. Return air for a pump located in the building 
enters the return air plenum above the ceiling due to zone 
pressure and migrates to the unit for reconditioning. For the 
remote unit, the return air is ducted from the ceiling plenum 
or from return air grilles to the unit. Outdoor air for interior 
units may be provided by a separate system and delivered to 
the return air plenum above the ceiling. Some building codes 
require that the outdoor air be directly supplied to the interior 
units. For the remote unit located on the roof, the outdoor air 
may be provided by the unit on the return air side through a 
damper that usually is set manually. 

8.4 HVAC COMPONENTS, FUNCTIONS, AND 
MALFUNCTIONS 

8.4.1 Outdoor Air: The outdoor air requirement for a 
space or an entire building must satisfy the need for acceptable 
indoor air quality and the need to replace air removed from 
the space or building by process or other exhaust. For indoor 
environment, ASHRAE Standard 62-1989, "Ventilation for 
Acceptable Indoor Air Quality," (83) is the accepted design 



criteria. Replacement air, however, will depend on factors 
such as total exhaust volume and pressure differential require- 
ments of the space or building plus the evaluation of potential 
airborne contaminants that may be generated inside or outside 
the building. For example, in the "open concept" type of office 
layout where partitions approximately five feet high enclose 
office spaces, the supply air has a tendency to ventilate only 
the space between the partitions and the ceiling. Very little, if 
any, of the supply air enters the actual occupied space directly 
to provide the necessary dilution. This allows contaminants 
in the occupied space to increase in concentration resulting in 
the potential for unsatisfactory air quality complaints. 

ASHRAE Standard 62-1989 recommends the measure- 
ment and documentation of the outdoor air intake volumetric 
flow rate on all configurations of HVAC systems. The pri- 
mary purpose of this requirement is to control the level of 
carbon dioxide, human odors, and the normal airborne con- 
taminants generated within the space. The published ventila- 
tion rates are based on occupancy or space usage and on an 
assumed occupant density. If the occupant density increases 
or the space usage increases, a degradation of the indoor air 
quality will occur which, in turn, will require an increase in 
outdoor air. 

It is possible to estimate the percentage of outdoor air by 
equation using the return, outdoor, and mixed air tempera- 
tures. The percentage of outdoor air also can be determined 
by equation using the carbon dioxide concentrations in the 
same air flow areas. Using these results, the estimated volu- 
metric flow rate of the outdoor air can be determined. The 
equations are as follows: 

Temperature Method: 



%outdoorair = ^A_jMA x100 



W ~ *OA 



where: 



t RA = temperature, return air 
t MA = temperature, mixed air 
t 0A = temperature, outdoor air 

Carbon Dioxide Method: 

o/ *j ppm/RA-ppm/MA 

% outdoor air = — — x 1 00 

ppm/RA-ppm/OA 

where: 

ppm/RA = CO2 concentration, return air 
ppm/MA = CO2 concentration, mixed air 
ppm/OA = CO2 concentration, outdoor air 

Given the legal implications, the direct measurement and 
documentation of the outdoor air volumetric flow rate is 
recommended. 



Ventilation Aspects of Indoor Air Quality 8-5 



TABLE 8-1 (8e) Relationships among extent of complaints regarding 
indoor air quality, C0 2 levels, and outdoor air ventilation rates 

Outdoor Air 
Ventilation Rate/Person 



COMMENTS 



C0 2 (ppm) CFM 



Us 



16.5 



10 



Occasional complaints, 600 35 

particularly if the air 
temperature rises 

Complaints are more 800 21 

prevalent 

Insufficient replacement air, 1000 15 

complaints more general 



The concentration of carbon dioxide within a space may 
provide a good indication of the outdoor air being delivered 
to the space. A study conducted by the Ontario Inter-Minis- 
terial Committee on Indoor Air Quality reported on the rela- 
tionship between levels of complaints, carbon dioxide 
concentrations, and the outdoor air ventilation rates. (86 > The 
results are indicated in Table 8-1 . 

Location of outdoor air intake may be found on the building 
roof, side wall, at ground level, or possibly at all three loca- 
tions for very large building complexes. Figure 5-28, "Air 
Flow Around Buildings," clearly illustrates the potential for 
airborne contaminants to enter the building through any of the 
outdoor air intakes. Sources of potential airborne contami- 
nants from the building and from sources remote or adjacent 
to the building should be thoroughly investigated. Assistance 
with this investigation should be requested of industrial hy- 
giene and environmental organizations who have responsi- 
bilities for the building. The location of the outdoor air intakes 
will be affected by the atmospheric air flow over the building 
as will the location and height of any exhaust stacks. Criteria 
for the atmospheric air flow characteristics and stack heights 
may be found in Figures 5-28, 5-29, and 5-30 of this manual 
or in the Fundamentals volume of the 1993 ASHRAE Hand- 
book, Chapter 14, "Airflow Around Buildings." (87) 

Roof intakes generally are located within a few feet of the 
roof surface. Standing water on the roof from weather condi- 
tions, HVAC equipment drains, or other sources present the 
potential for biological growth and entry into the intake. 
Unless discharged vertically above the recirculation region 
(see Figure 5-28), building exhaust systems from restrooms, 
processes within the building, and restaurant kitchens have 
significant potential for re-entry. 

Building sidewall intakes have the potential for entry of 
airborne contaminants from street level automotive traffic, 
shipping and receiving docks, and adjacent buildings. In the 
building wake region, see Figure 5-28, the potential for re-en- 
try increases for the outdoor air intakes and open windows or 
doors since the pressure in the recirculation region is lower 



than the surrounding area. Airborne debris such as leaves, 
paper, and atmospheric dirt may tend to collect on the intake 
bird screens which can reduce the intake area and may reduce 
the flow rate into the intakes. 

Ground level outdoor air intakes are possibly the least 
desirable location of the three described. This location offers 
the potential for air quality problems caused by standing 
water, automotive emissions, and as a collection point for dirt 
and debris. The security of a building can be compromised 
through the ground level intakes by the deliberate addition of 
foreign materials. 

8.4.2 Dampers: A typical office building HVAC system 
will include outdoor and return air dampers. Air flow through 
these dampers will vary over a wide range depending on the 
damper opening settings and the space or building require- 
ments. Indoor air quality problems often result if the outdoor 
air damper is not designed or adjusted to allow introduction 
of sufficient outdoor air for the current use of the building. 
Outdoor air requirements for acceptable indoor air quality 
indicate that the actual volumetric flow rate through the 
damper sections be monitored. When the outdoor air and 
return air dampers are combined in an HVAC system, there 
may be an imbalance in volumetric flow rates. 

It is customary to report the volumetric air flow rate through 
dampers in terms of damper opening. However, damper open- 
ing is not linearly proportionate to volumetric flow rate. 
Another misconception regarding dampers is that a "closed" 
damper will leak approximately 10%. Closed dampers may 
not leak at all. 

Dampers are mechanical devices (either parallel or opposed 
blade, see Figure 8-1 1) that require routine maintenance and 
periodic settings checks to assure proper air flow passage. 
Actuators, connecting arms and damper bearings are compo- 
nents that can affect the air flow if not properly connected or 
adjusted. This is an area where potential problems affecting 
the indoor air quality are not uncommon. In older buildings, 
the practice of disconnecting the outdoor air dampers to 
conserve energy is fairly common. This practice has been 
found by current surveys in some older buildings and also in 
newer buildings which have been occupied for an extended 
period. The result, of course, is a significant potential source 
of unsatisfactory indoor air quality complaints. 

8.4.3 Air Cleaning: The requirements for air cleaning vary 
according to the space or building requirements. There are, 
however, some basic factors that should be included in the 
design of the HVAC filter section. It is considered essential 
that all ventilation supply air, including outdoor and recircu- 
lated air, pass through a prefilter and a high efficiency final 
filter. Depending on requirements, filters such as charcoal, 
potassium permanganate, HEPA, and others may be speci- 
fied. See Figure 8-12 for various types. 



8-6 



Industrial Ventilation 



For example, paper dust is one of the contributors to 
unsatisfactory indoor air quality. The paper dust in itself is an 
irritant to the eyes and respiratory system. Also, many papers 
are chemically treated which tends to compound the irritant 
effect. The dust generated enters the return air section of the 
HVAC system, and may be reintroduced to the space being 
served by the system. Studies (8 - 8 ~ 8l0) have indicated that a 
significant percentage of the paper dust will be removed by high 
efficiency type filters which, as stated, should be included in the 
HVAC filter section. Figure 8-10(8.5) shows approximate effi- 
ciency versus particle size for typical air filters. 

In some older HVAC equipment, also in perimeter fan coil 
units and self-contained heat pumps, low efficiency filters are 
noted. It may be possible to replace these filters with medium 
efficiency filters of the same dimensions. The medium efficiency 
pleated filter has more than twice the filtering area, thereby 
increasing the interception of the airborne particulates without a 
significant increase in the static pressure requirements. 

The air cleaning or filter section of the HVAC system 
requires routine maintenance for replacing dirty filters, or in 
some instances, cleaning a reusable type of air cleaner. Rou- 
tine maintenance would also bring attention to damaged filters 
or filter frames and uneven air flow (by the dirt pattern on the 
face of the filters.) Even though a regular maintenance pro- 
gram may be in force, the filter sections, both pre- and final 
filters, present a potential source for indoor air quality com- 
plaints. 

Some HVAC systems utilize self-contained heat pumps to 
control conditions in specific zones. These heat pumps can be 
located above the ceiling in the return air plenum near the zone 
being served. This location results in a difficult situation in 
terms of providing service for the unit. The zonal heat pump 
usually has a low efficiency filter which can be completely 
blocked or missing due to the difficulty of servicing. 

8.4.4 Heating/Cooling Coils: Heating and cooling coils 
must be free of damage, especially the heat transfer fins. 
Irregularities in the fins will result in unequal heat transfer and 
will provide an area for dirt and other materials to accumulate. 
Air cleaning sections are not 100% perfect in removing the 
airborne contaminants regardless of efficiency ratings. Con- 
venient access to the coil section for inspection, cleaning and 
maintenance is essential to the proper functioning of the coil. 
See Figure 8-13. 

Cooling coils require some additional considerations. The 
supply air will pass through at a relatively low velocity, and 
the heat transfer will condense moisture on the coil fins. This 
moisture will drain to the condensate pan below the coil. 
Provision must be made to properly discharge the condensate. 
Since the air cleaning section is not perfect, some airborne 
contaminants will reach the cooling coils. The moisture accu- 
mulating on the coil fins will collect a significant percentage 
of these contaminants, which may adhere to the fins or drain 



to the pan below with the condensate. Accumulations of these 
contaminants create a source of molds, spores, bacteria, etc., 
that may enter the supply air stream. The condensate pan drain 
may allow condensate to accumulate at or near the outdoor 
air intake and can re-enter the HVAC system. It is essential 
that the cooling coil condensate pan be properly drained. The 
pan drain must be directed away from any outdoor air intake. 
The coil and pan must be inspected and cleaned on a regular 
basis. Microorganisms may proliferate if this is not done. 

8.4.5 Fans: HVAC systems vary in size and complexity 
over a wide range as do the fans as the system prime mover 
of volumetric flow for both supply, outdoor and return air. 
The fans may be the axial or centrifugal type with inlet vanes, 
outlet dampers, variable speed, direct or belt drive (see Figure 
8-14). Also, the fan or fans may be inside the housing of a 
self-contained HVAC unit or a separate component in a 
mechanical room or penthouse. See Figure 8-16 for a typical 
layout of the self-contained and mechanical room system. 

Since the fan is the prime mover of the HVAC system, a 
preventive maintenance program usually will reveal any po- 
tential malfunctions before they occur. Failures of the fan are 
usually noted immediately and corrected by maintenance. 
There is, however, a maintenance procedure (lubrication of 
moving components) that may be a source of odor complaints 
by the building occupants. Over lubrication, which is not an 
uncommon practice, may place a small quantity of the lubri- 
cant in the air flow into the fan. This may cause the blades to 
become coated and the lubricant odor to be carried into 
occupied areas. 

8.4.6 Humidifiers/Dehumidifiers: The incorporation of 
humidifiers/dehumidifiers is dependent on the space require- 
ments of the building. Humidifiers add moisture to the supply 
air by direct water or by steam spray. Dehumidifiers remove 
moisture from the supply air by a desicc ant-type filter or by 
cooling coils. Of the two processes, humidification is more 
widely used in HVAC systems (see Figure 8-15). The equip- 
ment used for humidification has a reputation for requiring a 
high level of maintenance for proper operation. For this 
reason, it is fairly common to find that the humidifier has been 
shut off — especially in office buildings. 

Both humidifying and dehumidifying are associated with 
water and dampness. This association presents the potential 
for the growth of molds, spores, bacteria, etc., that may enter 
the supply air flow. Proper drainage of any water or moisture 
generated by either process is essential 

8.4. 7 Supply Air Distribution: The air supply distribution 
system should be through sheet metal, steel and aluminum, or 
some type of non-fibrous duct material. There is increasing 
concern that fibrous materials such as fiberglass board ducts 
may produce fibers that may be potentially harmful. The use 
of interior duct insulation should also be avoided to eliminate 
the possibility of fibers entering the supply air stream. Supply 



Ventilation Aspects of Indoor Air Quality 8-7 



air duct systems should be designed in accordance with ac- 
cepted standards as detailed in current publications such as 
the ASHRAE Handbook series, standards of the Sheet Metal 
and Air Conditioning Contractors National Association 
(SMACNA), National Fire Protection Association, and other 
applicable criteria sources. 

The physical condition of the air supply duct system is 
important in the overall evaluation. Duct systems are usually 
located above the ceiling in the return air space together with 
utility lines, sprinkler lines, computer cables, etc. This space 
is relatively small, and when repairs, rearrangements, instal- 
lations, etc., occur, damage to the duct system may not be 
noticed but may affect the air distribution. 

It is common practice to connect the supply air duct to 
mixing boxes and/or diffusers with flexible duct. The fric- 
tional resistance can be up to five times that of sheet metal, 
and the manufacturer's data should be reviewed. Bends and 
turns using flexible duct will compound the losses and have 
a tendency to reduce the cross-section, which may in turn 
reduce the air volume. Improper hangers and supports also 
have the same tendency and results. The use of flexible ducts 
should be limited to minimum lengths, properly supported 
and securely fastened at each end. 

8.4.8 Supply Air Diffuser: The function of the supply air 
diffuser is to deliver and distribute the supply air throughout 
the occupied zone. Diffusers are available in a wide variety 
of types, shapes, and sizes — all of which will provide the 
proper air volume according to the supplier. See Figure 8-17 
for illustrations of various types of diffusers. The suppliers or 
manufacturers usually rate the diffusers in terms of supply air 
volume, static pressure drop, and the "throw" or pattern of the 
air delivery. Also, the published data will include illustrations 
of the air flow pattern created by the diffuser in a totally empty 
space and rely on the "coanda effect" for mixing the air in the 
space. However, when the space is occupied by people, 
equipment, file cabinets, cubicle partitions, library shelves, 
etc., the supply air pattern changes dramatically. This results 
in less supply air to the occupied zone. This ventilation aspect 
can be easily recognized through the use of a simple hand-held 
smoke test. 

Some of the more common problems associated with dif- 
fusers are as follows: 

1 . Variable air volume HVAC systems with fixed-supply 
air diffusers vary the air flow rate depending on tem- 
perature demand. Even if the minimum outdoor air is 
provided at all times, the reduced flow rate through the 
diffusers will reduce the throw and flow pattern. This 
may result in some areas within the occupied zone 
receiving little or no supply air. There are diffusers that 
automatically adjust for reduced air flow rates to main- 
tain a constant throw and flow pattern utilizing the 
"coanda effect." Reports from the field vary over a wide 



range. Some reports state that technical maintenance 
is relatively high to assure proper functioning. Others 
report that the pressure required increases as the slot 
area decreases which decreases the air flow rate. 

2. Variable air volume systems may include constant air 
flow to the diffusers through terminal boxes serving 
specific zones. These boxes contain an air supply fan 
with sensors and controls that draw air from the return 
air system or the specific zone based on the flow rate 
from the main system. Even though this will assure a 
constant air flow rate to the diffuser, it also presents 
some potential problems by localized recirculation 
within the specific zone. 

3. Supply air diffusers with fixed blades, diffusers cov- 
ered by a perforated plate, linear fixed diffusers, and 
fluorescent light troffers direct the supply air across 
the ceiling depending on the coanda effect for delivery 
to the occupied zone. In the "open concept" office 
layout with 5-ft high partitions enclosing office spaces, 
the supply air from the diffusers described will have a 
tendency to provide continuous supply only to the 
space between the partitions and the ceiling. See Fig- 
ure 8-18. Very little if any of the supply air enters the 
occupied zone directly which may result in com- 
plaints. This particular ventilation aspect may occur 
even though the system is providing 100% outdoor air 
and is often referred to as "short circuiting." The flow 
pattern above the partitions can usually be observed by 
using the simple smoke test. Another test that is done 
which may give a more qualitative result is the meas- 
urement of the carbon dioxide concentrations in the 
occupied zone and in the space between the partitions 
and ceiling. A concentration in the occupied zone that 
is significantly higher than the concentration above the 
partitions indicates that possibly up to 75% of the 
supply is above the partitions. 

4. Supply air diffusers with adjustable blades are avail- 
able in the multi-directional ceiling type, linear dif- 
fusers with adjustable T-bars, sidewall supply grilles, 
and other types. The adjustable feature does offer a 
means of better directing the supply air to the occupied 
zone. However, locations of the diffusers and the 
adjustment of the blades is critical to the distribution 
of the supply air. Improper adjustment may result in 
complaints by the occupants of excessive drafts. 

Location and type of supply air diffusers should be such 
that a continuous flow of air through the occupied space will 
occur at all times. Avoid situations that result in localized 
recirculation or short circuiting to the return air system. In 
general, the air flow pattern through a space by the supply air 
should receive critical attention and can be characterized in 
terms of ventilation efficiency. Two efficiencies should be 
considered: system efficiency and ventilation efficiency. Sys- 



8-8 



Industrial Ventilation 



tern efficiency is defined as the ratio of the actual volumetric 
flow rate to a specific space to the design volumetric flow rate 
for that specific space. Ventilation efficiency is defined as the 
ratio of the actual volumetric flow rate to a specific occupied 
zone to the design volumetric flow rate for that specific 
occupied space. Location and type of supply air diffusers are 
critical in the development of good ventilation efficiency. 
Design criteria in ASHRAE Standard 62-1989 will assist the 
design engineer in this effort. 

8.4.9 Return Air Grilles: The return grilles have the func- 
tion of receiving or exhausting air from a space through the 
return air system. Also, it is the function of the return air grilles 
to enhance the flow of the supply air through the space. The 
size and number of return grilles must be such that 100% of 
the supply air can be returned to the return air system. Loca- 
tion of the return air grilles influences the air flow pattern 
through the space and proper location will minimize localized 
recirculation zones. 

There is little design data available on the placement of the 
return air grilles but the location should be considered as 
important as the location of the supply air diffusers. The 
short-circuiting of the supply air directly to the ceiling return 
grilles may result in less than 50% of the supply air reaching 
the occupied zone. Development of an air flow pattern 
through an occupied zone from the supply diffusers to the 
return air grilles is a primary consideration. 

8.4.10 Return Air: The return air system may be either an 
open plenum type or a ducted system, both of which are 
typically located above the ceiling. In the return system, a 
static pressure balance between return air points must be part 
of the system design. It is obvious that the open ceiling plenum 
cannot be balanced by design which accounts for difficulty in 
providing a balanced supply air volume. For ducted return, 
the approach is similar to an industrial exhaust system. The 
static pressure in each run must be balanced by design at their 
junction which also accounts for difficulty in providing a 
balanced supply air volume. 

Pressure differentials at any junction are limited to 20% 
which is the maximum correction possible by damper. For 
differentials over this limit, redesign is necessary. Flexible 
duct is used at times to connect return air grilles to the ducted 
return. Since the negative pressure will tend to collapse the 
flexible duct, this practice should be avoided. 

8.4.11 Fan Coil Unit: The fan coil units used for HVAC 
are commonly located around the perimeter of a building and 
serve up to 15 feet from the outdoor wall. See Figure 8-8 for 
an illustration of a typical fan coil unit. These units may be 
totally self-contained with automatic controls; may include a 



through-wall duct for outdoor air; may have remote heating 
and cooling media or may be controlled manually at the unit. 
Since the fan coil units are rather compact, the filters are 
relatively small and in the low efficiency range. This will tend 
to increase the maintenance requirements since the return air 
is at the floor level — a potential significant source for dirt 
and possibly other contaminants. A provision for outdoor air 
may be a feature of the fan coil unit especially for units used 
to provide the HVAC for the building perimeter. The outdoor 
air intake is normally screened and may, over time, become 
blocked by dirt and debris from the outside atmosphere. Also, 
the intake may be located on an outside ledge of the building, 
depending on the building design, which may be a roosting 
area for birds. The outdoor air intake presents a significant 
source for contaminants and a difficult location to maintain. 

8.5 HVAC COMPONENT SURVEY OUTLINE 



The responsibility for monitoring the indoor air quality 
within a building may be assigned to an office individual, the 
building maintenance department, an outside environmental 
firm, or an HVAC maintenance contractor. In order to meet 
this responsibility, the assignee should conduct periodic walk- 
through surveys of the HVAC system and its components. The 
assignee should have a procedure or outline of the system compo- 
nents in order to conduct the survey. Basic information required to 
develop the procedure would including the following: 

1. The mechanical plans and specifications for the 
HVAC system to be surveyed including modifications 
or rearrangements, which are essential to conducting 
the survey. 

2. A detailed description of the type of HVAC system, 
its features and functions, especially for those who are 
not thoroughly acquainted with the system. 

3 . The current test and balance reports which can provide 
information on air distribution and design vs. perform- 
ance data. These reports may also indicate a specific 
component problem such as outdoor air requirements. 

4. Reports of complaints regarding the indoor air quality 
(which should include the nature and location). These 
reports are essential to conducting this survey. They 
may indicate a component problem such as a discon- 
nected diffuser and the lack of air movement in an 
occupied zone. 

In addition to the basic information, the walk-through 
survey includes observation or inspection of each of the 
HVAC system components for potential malfunction. The 
procedure or survey outline of the components together with 
specific notes follows. 



Ventilation Aspects of Indoor Air Quality 8-9 



1 . Outdoor Air (see Figure 5-28): 

A. Intake location and physical condition 



WALK-THROUGH SURVEY OUTLINE 

6. Humidifier/Dehumidifier (see Figure 8-15): 

A. Type and general condition 



B. Building exhaust stacks and vent pipes adjacent to 
intake 

C. Cooling tower; type and location 

D. Building entryways, doors, and windows as poten- 
tial entries for airborne contaminants 

E. Areas adjacent to the building as potential sources: 
shipping/receiving docks, parking lots, high traffic 
roads, adjacent buildings and operations, etc. 

2. Dampers (see Figure 8-11): 

A. Outdoor air; type and physical condition 

B. Return air; type and physical condition 

C. Face and bypass; type and physical condition 

D. Exhaust/pressure relief; type and physical condition 

3. Air Cleaning (see Figure 8-12): 

A. Type and general condition 

B. Prefilter; type, efficiency, and condition 

C. Final filter; type, efficiency, and condition 

4. Heating/Cooling Coils (see Figure 8-13): 

A. Pre-heat; type and condition 

B. Cooling; type and condition 

C. Condensate pan and drain 

D. Re-heat; type and condition 

5. Fans/Blowers (see Figure 8-14): 

A. Supply air; type and condition 

B. Return air; type and condition 

C. Exhaust/pressure relief; type and condition 



B. Condensate pan and drain 

7. Supply Air Distribution: 

A. Duct system; type and general condition 

B. Control box; type and condition 

C. Control box function 

D. Control box/d iff user connection; type and condition 

8. Supply Air Diffusers (see Figure 8-17): 

A. Type and general condition 

B. Characteristics of area served 

C. Number of diffusers this area 

D. Occupied zone air flow pattern; smoke test results 

E. Obstructions to flow pattern 

9. Return Air Grilles (see Figure 8-17): 

A. Type, location, and general condition 

B. Air flow pattern, supply to return; smoke test 

C. Obstructions to flow pattern 

10. Return Air System: 

A. Open plenum; general condition, location of return 
air opening; return air fan/duct 

B. Ducted return; location and general condition 

C. Balancing dampers; type and condition 

11. Miscellaneous Potential Contaminant Sources: 

12. General Comments and Notes: 



8-10 



Industrial Ventilation 



REFERENCES 

8.1 The Trane Company: Trane Air Conditioning Manual, 
(February, 1961). 

8.2 International Conference of Building Officials: Light, 
Ventilation and Sanitation, Section 605, Uniform 
Building Code, (1988). 

8.3 American Society of Heating, Refrigeration and Air 
Conditioning Engineers: Ventilation for Acceptable 
Indoor Air Quality, ASHRAE Standard 62-1989. 
ASHRAE, Atlanta, GA (1989). 

8.4 American Society of Heating, Refrigeration and Air 
Conditioning Engineers, Thermal Environmental 
Conditions for Human Occupancy, ASHRAE Stand- 
ard 55-1990. ASHRAE, Atlanta, GA (1992). 

8.5 American Society of Heating, Refrigeration and Air 
Conditioning Engineers: ASHRAE Handbook, 



HVAC Systems and Equipment. ASHRAE, Atltanta, 
GA(1992). 

8.6 Rajhans, G.S.: Findings of the Ontario Inter-Ministe- 
rial Committee on Indoor Air Quality. In: Proceedings 
of the ASHRAE/SOEN Conference, IAQ *89, pp. 195- 
223. ASHRAE, Atlanta, GA (1990). 

8.7 American Society of Heating, Refrigeration and Air 
Conditioning Engineers: ASHRAE Handbook, Fun- 
damentals volume. ASHRAE, Atlanta, GA (1989). 

8.8 Bauer, E.J.; et al.: Use of Particle Counts for Filter 
Evaluation. ASHRAE Journal . ASHRAE, Atlanta, 
GA (October 1973). 

8.9 Duffy, G: Filter Upgrades. Engineered Systems 
(July/August 1993). 

UO Ottney, T.C.: Particle Management for HVAC Sys- 
tems. ASHRAE Journal, p. 23. ASHRAE, Atlanta, GA 
(July 1993). 



Ventilation Aspects of Indoor Air Quality 8-11 



Reheat coil 



\ 



Humidifier- 



Cooling coi 
Preheat coil 



.V X 



Outside ai 



-y- 



f>l 



y 



Filters 



Supply air 
fan 



Condensate pan 



Supply air' 
diff users 
zone // 1 



/ \ 



Zone 
thermostat 



Reheat coil 



\ 

Supply air 
diffusers 
-~y0 zone //2 



Exhaust aii- 



Zone ^ 
thermos tai 



Return air 



Return air 
fan 



y = Damper 



NOTE; See text regarding outside air 
requirements and distribution. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SINGLE DUCT CONSTANT 
VOLUME WITH REHEAT 



DATE 



4-94 



FIGURE 



8-1 



8-12 Industrial Ventilation 



Bypass box — 



-^Bypass to return air 



Supply air 
diffusers 
zone // 1 



Humidifier - 



Outside air 



Cooling coil 
Preheat coil ~x 



* \ 



1 



I 

Supply air 
v / fan 



^ i 



l 



/ ' X 



-© 



Zone 
thermostat 



o 
ex. 



Exhaust air 

c 1 — y- - 



v Filters 



-Condensate pan 

Bypass box 



j 



Bypass to return air 



Supply air 
diffusers 
II) zone ■// 2 



Zone -- 

thermostat 



Return air 



Return air 
fan 



^ — Dampers 



NOTE: See text regarding outside air 
requirements and distribution. 



::rican conference 
of governmental 
ndustrial hygienists 



SINGLE DUCT CONSTANT 
VOLUME WITH BYPASS 



DATE 



4-94 \ 



figure: 



8-2 



Ventilation Aspects of Indoor Air Quality 8-13 



Reheat coil — \ 
VAV box-—-. \ 



S u p p I y a i r 
V V dif f users 

\- ^, zone //■ 1 / 



Humidifier- 



Outside air 




^ 



> 



| Supply air 
^ :> VAV fan 



> 
> 



i I 









Zone 
thermostat 



— Reheat coil 



Filters 



\ 



X 



"1 

Condensate pan J 

/ 
VAV box / 



J 






o 



l^ Supply air 

\ diffusers 

X "^:D zone //1 

Zone ~ ' 

thermostat 



E x h a u s t air 



(X-^y. 



Return air 



Return air 
VAV fan 



^ - Dampers 



NOTE: See text regarding outside air 
requirements and distribution. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL H YGIENISTS 



SINGLE DUCT VARIABLE; 
AIR VOLUME WITH REHEAT 



DA' 



'e 4 - 94 



FIGURE 



8-3 



8-14 Industrial Ventilation 



induced air — s 



Cooling coil 
Preheat coil - x 



V 



Outside air 



It 



VAV induction box- 



Humidifier 



V: 



Exhaust air 



Fitters 



Supply air 
Z> VAV fan 



Condensate pan 

VAV induction box- 



Supply air —y 
diffusers / 
Z o n e // 1 / 



/'\ 



-a) 



Zone 
thermostat 



I n d u c e d a i r 



— -P 



\ 



Supply air 
diffusers 
LP Zone H-2 



Zone 
thermostat 



Return air 



Return air 
VAV fan 



$ - Dompers 



NOTE: See text regarding outside air 
requirements and distribution. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRI AL H YG1 ENISTS 



SINGLE DUCT VARIABLE 
AIR VOLUME 'WITH INDUCTION 



DATE 



4-94 I' 



1GURE 



8-4 



Ventilation Aspects of Indoor Air Quality 8-15 



Return air 

Fan powered 
VAV box 






Cooling coif 
Preheat coil 



Humidifier 



Outside air 



y 



o 



Exhaust air 



?-^m/- 



\ 



\ 



Supply air 
\ y VAV fan 



Supply air -y 
diffusers / 

zone // 1 / 

- — -& <> 



/ " \ 



-® 



Zone 
thermostat 



Return air 



Filters 



Condensate pan 



S 



X " x 



Fan powered — 
VAV box 



Zone ^ 
thermostat 



\ 

Supply air 
diffusers 
I) Zone //2 



Return air 



Return air 
VAV fan 



$ = Dampers 



NOTE: See text regarding outside air 
requirements and distribution. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SINGLE DUCT VARIABLE AIR 

VOLUME WITH FAN POWERED 

DEVICES AND REHEAT 



DATE 



4^94 



FIGURE 



8-5 



8-16 Industrial Ventilation 



-Cooling coil 



Cooling coil 
Preheat coil -v 



Condensate pan 



Humidifier 



Outside air 

c 3 — — y- 



j 




Supply air 
~y f a n 



-Or* 



-f-fc 



Dual duct 
\ mixing box 



/■ ^ii- * 

1 '-Reheat coil 7 XX 

(j)--^ Supply air 

_ dif fusers 

,, Zone , , zone #1 

thermostat ;/ 



g®3 



Cooling c o i ! 



Filters 



Condensate pan 



! T~ Dual duct 

I \ mixing box 

1 fc=; - 



i 



-Reheat coil 



/ 



/ \ 



Supply air 

- 7 dif fusers 

Zone 

thermostat zone # 2 



Exhaust air 



— ./- 



f \_ 

Return air 
fan 



Return air 



^ = Dampers 



NOTE: See text regarding outside air 
requirements and distribution. 



A M E R I C A N C N F E R E N C E 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DUAL DUCT CONSTANT 
AIR VOLUME 



DATE ^ _ p4 



FIGURE 



8-6 



Ventilation Aspects of Indoor Air Quality 8-17 



Face & by-pass dampers — 
zone control clampers \ 



Heating coif 
Humidifier 



Preheat coil - 



Outside air 




o 

CD 

or 



Exhaust air 



<^—y- 



Cooling coil 
Condensate pan 



Suppiy air 
diffusers 
jS zone //2 

Zone 
thermostat 



Return air 



Return air 
fan 



^ - Dampers 



NOTE: See text regarding outside air 
requirements and distribution. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SINGLE DUCT CONSTANT 
AIR VOLUME MULTI-ZONE 



DATE 



4 — Q4 figure 8-7 



8-18 Industrial Ventilation 



Outside 
wall 



Outside Air 
(Optional) 



Room supply air 




Supply fan 




Condensate 

pan 




-^— L_* 




Filter 



Recirculated air 



Return 
air 



^ — Damp 



ers 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



AIR COIL UNIT 



DATE 



4-94 



FIGURE 



8-8 



Ventilation Aspects of indoor Air Quality 8-19 



Cooling coil 
Preheat coil 



Outside air 



-/- 



Exhaust air 



Heat pump zone #1 
(heat or cool) 

Filter 



Humidifier 



> 
) 




1 Filters V 



Supply air 
fan 

Condensate pan 




Supply air 
dlffusers 
zone #1 



n f ^ 



Zone 
thermostat 

Supply air to 
plenum is ducted 
directly to or in 
near vicinity of 
heat pump 



Filter 



Heat pump zone #2 
(heat or cool) 




Supply air 
dlffusers 
zone #2 




/} f \ 



Zone 
thermostat 




Return air 



Return air 
fan 



$ - Damp 



ers 



NOTE: See text regarding outside air 
requirements and distribution. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



ZONE HEAT 
PUMP SYSTEM 



DATE 4 _ g/f_ 



FIGURE 



8-9 



8-20 



Industrial Ventilation 



100 



80 



&^ 



- 60 



>- 

o 

z: 

UJ 

O 



40 



20 



































(A) 




































































//© 














®L/ 








© 






























(® / 




























/ ^ / 




® 90-95% EFFICIENCY 
© 80-85% EFFICIENCY 
© 30-45% EFFICIENCY 
© 30-45% EFFICIENCY 
© 25-30% EFFICIENCY 
© 70-80% ARRESTANCE 





































































0.3 0.4 0.5 0.6 0.8 



4 5 6 



10 



PARTICLE SIZE, MICROMETERS 



Approximate Efficiency Versus Particle Size 
for Typical Air Filters (See notes 1 & 2) 



NOTE: 1. Compiled and averaged from manufacturer data 
Efficiency and arrestance per ASHRAE Standard 
52 - 76 Test Methods. 
2. Caution: Curves are approximations only for general guidance. 
Values from them must not be used to specify air filters, 
since a generally recognized test standard does not exist. 



From: ASHRAE Equipment 



[ERICAN CONFERENCE 
OF GOVERNMENTAL 
NDUSTRIAL HYGIENISTS 



FIL TEE EFFICIENC Y 
VS PARTICLE SIZE 



DATE 8~96 



FIGURE 8—10 



Ventilation Aspects of Indoor Air Quality 8-21 




PARALLE L BLA DE 



o 
I 



A-WIDE 




OPPO SED BLA DE 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DAMPERS: 
PARALLEL AND OPPOSED BLADE 



DATE P~97 



I 



FIGURE 



8-11 



8-22 Industrial Ventilation 




High Efficiency 
Disposable 
Cell Fi-ter 



Super Inception or 
Absolute Filter 








Washable Metal Filter 






MM- 

^ >-( >-<{ )-< W""\ > 




I hrowaway 
i'ype Filter- 



Disposable Medio Filter 




Disposable Media 
Filter 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



TYPICAL HVAC AIR FILTERS 



DATE 



2-97 



FIGURE 



8-12 



Ventilation Aspects of Indoor Air Quality 8-23 












.:•-•:;> 









: i^ggm 

?l si 



m 



m- 



Tfffl'Pli 1 !! 



!.i 



■ 



A M h : E I C A N C N F' E R E N C E 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENIS' 



//£/! TING/COOLING COILS 



DATE 



02-97 



FIGURE 



5-15 



8-24 Industrial Ventilation 



HOUSING 



DIVERTER 



DISCHARGE 

OUTLET AREA 




FRAME 



INLET COLLAR 



Reprinted from AMCA Publication 201 -90, FANS AND SYSTEMS, by 
permission of the Air Movement and Control Association, lnc. c 1} 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DATE 5-92 



TERMINOLOGY FOR 

CENTRIFUGAL FAN 

COMPONENTS 



FIGURE 



8-14 



Ventilation Aspects of Indoor Air Quality 8-25 



J WET AIR IN 



COOLING COILS 



r~o 



CONDITIONED. AIR OUT 






DRAIN 



v — CONDENSATE PAN 



REFRIGERATION DEHUMIDIFICATION 



\j ^IEL£J=ow_ 



SNOAM 

TKT" 



Jf 



f 

DRAIN 

:am mumidrter 



w w aaa i B a M Mfl ni^^ 



JKWM»M«il«« 



P^^^«M^^M8Mim^ 



I ! 

V WET AIR IN 

v„ _ __ 



V" 



-y— '" " 

ELIMINATORS 



ATOMIZER ^ 



ATOMIZING HUMIDIFIER Wi'TH 
FILTER ELIMINATOR 



OPTIONAL 



.-.v.v.v.v.v.v.v.v.v.-. ( TTZZZZTZZZZZ. 

/ A 

/ 

ELIMINATORS ■■ / 




ir 



PNEUMATIC ATOMIZING HUMIDlFlEi 



AMERIC AN CONFERENCE 

OE GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HVAC 
HUMIDIFIER \ DEHUM1DIFIER 



DAT! 



02-97 



FIGURE 



8-15 



8-26 



Industrial Ventilation 



SUPPLY 
AIR 




SUPPLY 
AIR 



DUAL 
DUCT 



HVAC SELF CONTAINED 



PENTHOUSE ROOE 




OUTDOOR 
AIR 



SUPPLY AIR 
SINGLE DUCT 



HVAC CENTRAL EQUIPMENT ROOM 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HVAC SYSTEM 

SELF CONTAINED 

AND EQUIPME NT ROOM 

date 8-96 F 



FIGURE 



8-16 



Ventilation Aspects of Indoor Air Quality 8-27 









/ 



uM=Hl// 



I 



SQUARE 




PERFORATED 




0'fr W)) 



yjy 



LINEAR 



ROUND 



fcHBfciBt: 



... H.JL. LJ J ... L|U 



Iflffl 



* ! 



mSf 




SIDE WALL 



AIR SUPPLY DIFFUSERS 



ADJUSTABLE BLADE 



SIDE WALL 













— i L 














- 


„|_L 


4- 


~*ji~' 


- 








rT~ 


















i 


















— J— 1— i— 
















i ■ ! 


- 














r"!"!™ 5 
















; i 1 














i_Xj_ 














Ttt 1 






















Cl 


:i 


, 


N 


G 











RETURN AIR GRILLES 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



AIR SUPPLY DIFFUSERS 8c 

RETURN APR GRILLES 



DATE 8-96 



FIGURE 



8-17 



8-28 Industrial Ventilation 




Q i n 



SDE WALL SUPPLY 






A/S 




CD 



R/£ 



j v: 



R/A 




R/A 



C E 



NG SUPPi 



GRILLES 



AMERICAN CONFERENCE 

OF GOVERNMENTAL, 
NDUSTRIAL HYGIENiSTS 



T YPICA I, PA R Tl 'TIG NED 



OFFICE AIR P ATT EI 



?N 



DATE 



8-96 



•IGliRE 



8-18 



Chapter 9 

TESTING OF VENTILATION SYSTEMS 



9.1 INTRODUCTION 9-2 

9.1.1 Initial Test 9-2 

9.1.2 Periodic Test 9-2 

9.2 MEASUREMENTS OF VOLUMETRIC 

FLOW RATE 9-2 

9.2. 1 . Air Velocity Instruments 9-3 

9.3 CALIBRATION OF AIR MEASURING 

INSTRUMENTS 9-6 

9.3.1 Design of a Calibrating Wind Tunnel 9-6 

9.3.2 Use of Calibrating Wind Tunnel 9-7 

9.4 PRESSURE MEASUREMENT 9-11 

9.4.1 Static Pressure 9-11 

Figure 9-1 Rotating Vane Anemometer 9-3 

Figure 9-2 Swinging Vane Anemometer 9-4 

Figure 9-3 Anemometer Applications 9-5 

Figure 9-4 Thermal Anemometer 9-6 

Figure 9-5 Calibration Wind Tunnel 9-8 

Figure 9-6 Calibration 9-9 

Figure 9-7 Static Tap Connections 9-1 1 

Figure 9-8 Pi tot Tube Measurement 9-12 

Figure 9-9 Standard 'Pitot Tube 9-14 



9.4.2 Hood Static Pressure 9-11 

9.4.3 Hood Static Pressure Interpretation 9-13 

9.4.4 Velocity Pressure 9-14 

9.5 PITOT TRAVERSE METHOD 9-15 

9.6 CORRECTIONS FOR NON-STANDARD 

CONDITIONS 9-21 

9.6.1 Example Traverse Calculations 9-22 

9.7 CHECK-OUT PROCEDURES 9-23 

9.7.1 Difficulties Encountered in Field 

Measurement 9-27 

REFERENCES 9-27 



Figure 9-10A 1 0-Point Pitot Traverse in a Circular Duct ... 9-15 

Figure 9-10B 6-Point Pitot Traverse in a Circular Duct .... 9-15 

Figure 9-11 Velocity Pressure Distributions 9-16 

Figure 9-12 Pitot Traverse Points in a Rectangular Duct' . . 9-17 

Figure 9-13 U-Tube Manometer 9-20 

Figure 9-14 Inclined Manometer 9-20 

Figure 9-15 Aneroid Gauge 9-21 

Figure 9-16 Sample System 9-25 

Figure 9-17 Survey Form 9-26 



9-2 



Industrial Ventilation 



9.1 INTRODUCTION 

Every ventilation system should be tested at the time of 
initial installation to verify the volumetric flow rate(s) and to 
obtain other information which can be compared with the 
original design data. Testing is necessary to verify the setting 
of blast gates, fire dampers, and other air flow control devices 
which may be a part of the system. Initial testing will provide 
a baseline for periodic maintenance checks and isolation of 
system failures should a malfunction occur. Many govern- 
mental codes require initial and periodic testing of exhaust 
systems for certain types of processes. Exhaust system test 
data are also useful as a basis for design of future installations 
where satisfactory air contaminant control is currently being 
achieved. 

The tests described in this text pertain to ventilation sys- 
tems only. Environmental tests should be conducted prior to 
and after installation to verify system performance. In these 
cases, the services of a qualified industrial hygienist may be 
required. 

9.1.1 Initial Test: The Pitot tube and manometer are the 
standard for initial field testing of equipment when used as 
described. However, other instruments may be used. As noted 
later, all instruments must be calibrated. Identify on the test 
sheet the instruments and procedures used. A sample survey 
form is located at the end of the chapter (Figure 9-17). The 
following steps outline the recommended procedure and the 
minimum data necessary for a thorough initial ventilation test. 

1 . Review the system specifications and drawings to de- 
termine the relative location and sizes of ducts, fittings, 
and associated system components. Where possible, 
pertinent prints should be carried to the test site. 

2. Inspect the system to determine that its installation is 
in accordance with the specifications and drawings. 
Check such items as fan rotation, belt slippage, and 
damper settings. 

3 . Include a drawing of the system as installed. Select and 
identify test locations. 

4. Measure the volumetric flow rate, fan static pressure, 
fan speed, motor speed, motor amperes, and the tem- 
perature of the air in the system. Also determine pres- 
sure drops across all components such as coils, fittings, 
and air cleaning equipment. 

Fan speed may be measured directly at the end of the 
fan shaft using a revolution counter and a watch. A 
tachometer or stroboscopic measuring device also may 
be used. 

The operating amperage is obtained with an ammeter. 
The readings taken on each lead on the three-phase 
current should be averaged and compared to the rating 
on the motor name plate at the operating voltage to 
determine if the motor is operating within its rated 



range. 

For some tests, moisture content of the air in the system 
and/or ambient barometric pressure should be obtained 
also. See example calculations in Section 9.6. 

5. Record the test data and design specifications on the 
data sheet. Calculate test results following the format 
on the data sheets. 

6. Compare the test data with design specifications. De- 
termine if alterations or adjustments of the system are 
necessary to meet specifications, codes, or standards. 

7. If alterations or adjustments are made, retest the sys- 
tem, and record the final test data. On the drawing, note 
any physical changes that were made in the system. 

9.1 .2 Periodic Test: The performance of a system should 
be checked periodically. If there have been no alterations to 
the system, this can be done by static pressure measurements 
and close visual inspection. Measurements also can be made 
continuously by means of an operating console or other 
remote readout system. 

The following is the recommended procedure with sugges- 
tions as to types of measurements needed to perform the 
periodic tests: 

1 . Refer to the initial data sheet for test locations. 

2. Inspect the system for physical damage (broken, cor- 
roded, collapsed duct, etc.) and proper operation of com- 
ponents (fan, damper, air cleaner, controls, burner, etc.). 

3. Measure static pressure at the same locations used in 
the initial test. 

4. Compare measured static pressures with initial test. 
From these comparisons, determine if the system is 
performing at initial levels. 

5. Make and record any correction required. 

6. Recheck the system to verify performance if correc- 
tions have been made. 

Whenever alterations have been made to the system, a new 
initial test is necessary following the procedures outlined 
under Section 9.1.1, "Initial Test." 

9.2 MEASUREMENTS OF VOLUMETRIC FLOW RATE 

The most important measurement in testing of systems is 
the measurement of the volumetric flow rate in cfm. This 
should be done before balancing the system is attempted. The 
commonly used instruments are of the velocity measuring 
type rather than quantity meters. Therefore, it is necessary to 
obtain not only the average air velocity through an opening 
or duct, but also the net cross-sectional area at the point of 
measurement. The volumetric flow rate then can be deter- 
mined from the equation: 



Q=VA 



[9.1] 



Testing of Ventilation Systems 9-3 



TABLE 9-1 . Characteristics of Flow Instruments 



Instrument 


Range, fpm 


Hole Size 
(for ducts) 


Range, 
Temp* 


Dust, Fume 
Difficulty 


Calibration 
Requirements 


Ruggedness 


General 
Usefulness 
and Comments 


PITOT TUBES with inclined manometer 














Standard 


600 -up 


3/8" 


Wide 


Some 


None 


Good 


Good except at 
low velocities 


Small Size 


600 - up 


3/16" 


Wide 


Yes 


Once 


Good 


Good except at 
low velocities 


Double 


500 - up 


3/4" 


Wide 


Small 


Once 


Good 


Dirty air stream 


SWINGING VANE ANEMOMETERS 
















25-10,000 


1/2-1" 


Medium 


Some 


Frequent 


Fair 


Good 


ROTATING VANE ANEMOMETERS 














Mechanical 


30-10,000 


Not for duct use 


Narrow 


Yes 


Frequent 


Poor 


Special; limited 
use 


Electronic 


25-200 
25 - 500 
25-2000 
25 - 5000 


Not for duct use 


Narrow 


Yes 


Frequent 


Poor 


Special; can 
record; direct 
reading 



Temperature range: Narrow, 20-150 F; Medium, 20-300 F; Wide, 0-800 F 



where: 

Q = volumetric flow rate, (cfm) 

V = average linear velocity, feet per minute (fpm) 

A = cross-sectional area of duct or hood at the meas- 
urement location, ft 3 

9.2.1 Air Velocity Instruments: The volumetric flow rate 
of an exhaust system can be determined by the use of various 
types of field instruments which measure air velocity directly. 
Typically, these instruments are used at exhaust and discharge 
openings or, depending on size and accessibility, inside a duct. 
The field technique is based on measuring air velocities at a 
number of points in a plane and averaging the results. The 
average velocity is used in Equation 9.1 to determine the 
volumetric flow rate. Due to the difficulty of measuring the 
area of an irregularly shaped cross-section and the rapid 
change in velocity as air approaches an exhaust opening, meas- 
urements obtained should be considered an approximation of the 
true air flow. All instruments should be handled and used in strict 
compliance with the recommendations and directions of the 
manufacturers. Table 9-1 lists some characteristics of typical air 
velocity instruments designed for field use. 

Rotating Vane Anemometer (Figure 9-1): This instrument 
is accurate and can be used to determine air flow through large 
supply and exhaust openings. Where possible, the cross-sec- 
tional area of the instrument should not exceed 5.0% of the 
cross-sectional area of the duct or hood opening. The standard 
instrument consists of a propeller or revolving vane connected 
through a gear train to a set of recording dials that read the 
linear feet of air passing in a measured length of time. It is 
made in various sizes; 3", 4", and 6" are the most common. It 



gives average flow for the time of the test (usually one 
minute). The instrument requires frequent calibration and the 
use of a calibration card or curve to determine actual velocity. 
The instrument may be used for either pressure or suction 
measurements using the correction coefficients listed by the 
manufacturer. The standard instrument has a useful range of 
200-3000 fpm; specially built models will read lower velocities. 

Direct-recording and direct-reading rotating vane ane- 
mometers are available. These instruments record and meter 




FIGURE 9-1. Rotating vane anemometer 



9-4 



Industrial Ventilation 



electrical pulses developed by a capacitance or inductive 
transducer. The impulses are fed to the indicator unit where 
they are integrated to operate a conventional meter dial. 
Readings as low as 25 fpm can be measured and recorded. 

The standard 4" rotating vane anemometer is unsuited for 
measurement in ducts less than 20". in diameter as it has too 
large a finite area and its equivalent cross-sectional area is 
difficult to compute. The conventional meter is not a direct- 
reading velocity meter and must be timed. It is fragile and care 
must be used in dusty or corrosive atmospheres. Newer units 
of 1" diameter which can be used in ducts as small as 5 M in 
diameter are available. 

Swinging Vane Anemometer (Figure 9-2): This instrument 
is extensively used in field measurements because of its 
portability, wide-scale range, and instantaneous reading fea- 
tures. Where accurate readings are desired, the correction 
coefficients in Table 9-2 should be applied. The instrument 
has wide application and, by a variety of fittings, can be used 
to check static pressures and a wide range of linear velocities. 
The minimum velocity is 50 fpm unless specially adapted for a 
lower range. The instrument is fairly rugged and accuracy is 
suitable for most field checks. Uses of the swinging vane ane- 
mometer and its various fittings are illustrated in Figure 9-3. 

Before using, check the meter for zero setting by holding 
it horizontal and covering both ports so that no air can flow 
through. If the pointer does not come to rest at zero, an 
adjustment must be made to correct the starting point. The 
meter should be used in an upright position and, when using 
fittings, it must be held out of the air stream so that the air 
flows freely into the opening. The length and inside diameter 
of the connecting tubing will affect the calibration of the 
meter. When replacement is required, use only connecting 
tubing of the same length and inside diameter as that origi- 
nally supplied with the meter. 



TABLE 9-2. Correction Factors for the Swinging Vane and Thermal 
Anemometers 




Grille Openings 



Correction Factor C F 
(percent) 



Pressure 
More than 4 in wide and up to 600 in area, 
free opening 70% or more of gross area, no 
directional vanes. Use free-open area. 

Suction 
Square punched grille (use free-open area) 
Bar grille (use gross area) 
Strip grille (use gross area) 
Free open, no grille 



93 



78 

73 

No correction 



Where temperatures of an air stream vary more than 30 F 
from the standard temperature of 70 F and/or if the altitude is 
greater than 1,000 feet, it is advisable to make a correction for 
temperature and pressure. Corrections for change in density 
from variations in altitude and temperature can be made by 
using the actual gas density (p) shown in Equation 9.9 in the 
following equation: 



V =V 

v c w r 



where: 



0.075 



[9.2] 



V c = corrected velocity, fpm 
V r = velocity reading of instrument, fpm 
p = actual gas density, lbm/ft 3 

Use at Supply Openings: On large (at least 3 ft 2 ) supply 
openings where the instrument itself will not block the open- 
ing seriously and where the velocities are low, the instrument 
itself may be held in the air stream with the air impinging 
directly in the left port. When the opening is smaller than 3 
ft 2 and/or where the velocities are above the "No Jet" scale, 
appropriate fittings must be used. 

Because the velocity and static gradient in front of an 
exhaust opening is steep, the finned opening of the fitting must 
be held flush with the exhaust opening. If the opening is 
covered by a grille, hold the fin directly against the grille and 
use the correction coefficients listed in Table 9-2 when com- 
puting exhaust volumes. 



Q - C F VA 

where: 



[9.3] 



FIGURE 9-2. Swinging vane anemometer 



C F = correction coefficient in percent of scale read- 
ing. 

While it can be used to measure air velocities, static pres- 
sure, and total pressure in ducts, it has several disadvantages. 
Used in place of a Pitot tube for velocity or total pressure 
measurements, it necessitates a much larger hole in the duct, 
often difficult and impractical to provide. When the velocities 



Testing of Ventilation Systems 9-5 



G) 






<% 





v v 



DIFFUSER 





^ 





C^ 




DIFFUSER 



BOXBOARD CONE 
(COMMERCIALLY AVAILABLE) 



AREA SIZED FOR 
MAXIMUM VELOCITY 
OF 400 FPM 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



ANEMOMETER 
APPLICATIONS 



DATE 1-88 



FIGURE 



9-3 



9-6 



Industrial Ventilation 



are high, there may be no appreciable errors at the high end 
of the scale and the instrument tends to read low on the 
discharge side of the fan and high on the inlet side. 

The presence of dust, moisture, or corrosive material in the 
atmosphere presents a problem since the air passes through 
the instrument. In those instruments calibrated for use with a 
filter (the filter must always be used), the fi Iter itself is a source 
of error because as the filter becomes plugged, its resistance 
increases and thus alters the amount of air passed to the 
swinging vane. The instrument requires periodic calibration 
and adjustment. 

Thermal Anemometer (Figure 9-4): This type of instrument 
employs the principle that the amount of heat removed by an 
air stream passing a heated object is related to the velocity of 
the air stream. Since heat transfer to the air is a function of the 
number of molecules of air moving by a fixed monitoring 
point, the sensing element can be calibrated as a mass flow 
meter as well as a velocity recorder. Commercial instruments 
use a probe which consists of two integral sensors: a velocity 
sensor and a temperature sensor. The velocity sensor operates 
at a constant temperature — typically about 15 F above 
ambient conditions. Heating energy is supplied and controlled 
electrically by a battery-powered amplifier in the electronics 
circuit. The electrical current required to maintain the probe 
temperature in conjunction with the temperature sensor will 
provide an electrical signal which is proportional to the air 
velocity and is displayed on either a digital or analog meter. 
Additional features often include time integration of fluctuat- 
ing readings and air temperature at the probe. Displays are 
available in either English or S.I. units. 

The velocity sensor should be used with care in normal field 
use and is insensitive to mild particulate contamination. The 
probe can be used directly to measure air velocity in open 
spaces at air exhaust and supply air openings. Attachments 




are available to measure velocity pressure. Due to the small 
diameter of the probe, measurements can be made directly 
inside of ducts using the measurement techniques described 
later for Pitot traverses. When used at supply or exhaust 
openings covered by grilles, the correction coefficient listed 
in Table 9.2 should be used. 

Battery charging and maintenance is extremely important 
and the battery voltage must be checked prior to instrument 
use. The correction coefficients for this instrument are the 
same as a swinging vane anemometer (see Table 9.2). Instru- 
ments of this type require both initial and periodic calibration. 

Smoke Tubes: Low velocity measurements may be made 
by timing the travel of smoke clouds through a known dis- 
tance. Smoke trail observations are limited to velocities less 
than 150 fpm since high air velocities diffuse the smoke too 
rapidly. Commercially available, smoke tubes and candles are 
useful in the observation of flow patterns surrounding exhaust 
or supply openings. They also can be used for checking air 
movement and direction in plant space. 

The visible plume is corrosive and should be used with care 
near sensitive processes or food preparation. Smoke candles 
are incendiary and thus cannot be used in flammable atmos- 
pheres. They should not be hand-held. Alternative methods of 
observing air flow patterns include the use of soap bubbles, water 
vapor cooled by dry ice (C02), and heated vegetable oil. 

Tracer Gas: The principle of dilution sometimes is used to 
determine rate of air flow. A tracer gas is metered continu- 
ously into one or more intake ports (hood or duct openings) 
along with the entering air stream. After thorough mixing and 
system equilibrium has been established, air samples are 
collected at some point downstream — usually at or near the 
effluent point — and the concentration of the tracer gas in the 
exit stream is determined. The rate of air flow is readily 
calculated from the degree of dilution noted in the exit and 
feed gas concentrations (rate of air flow equals rate of feed 
divided by tracer gas concentrations)/ 9 5) 

The tracer gas usually is selected on the basis of the 
following: 1) ease of collection and analysis, 2) not present 
naturally in the process being studied, 3) not absorbed chemi- 
cally or physically in the duct system, 4) non-reactive with 
other constituents of the gas stream, and 5) non-toxic, non- 
explosive, and non-odorous. Some frequently used tracer 
gases are sulfur hexafluoride and carbon dioxide. 

9.3 CALIBRATION OF AIR MEASURING INSTRUMENTS 

Direct-reading meters need regular calibration because 
they can be easily impaired by shock (dropping, jarring), dust, 
high temperatures, and corrosive atmospheres. Meters should 
be calibrated regularly and must be calibrated if they will not 
adjust to zero properly or if they have been subjected to rough 
handling and adverse atmospheres. 



FIGURE <M. Thermal Anemometer 



9.3.1 Design of a Calibrating Wind Tunnel: A typical 



Testing of Ventilation Systems 9-7 



calibrating wind tunnel for testing air flow meters must have 
the following components: 

1 . A satisfactory test section. This is the section where 
the sensing probe or instrument is placed; it must be 
uniform in air flow both across the air stream and in 
line with the air flow. A section with a pronounced 
vena contracta and turbulence will not give satisfac- 
tory results. 

2. A satisfactory means of precisely metering the airflow. 
The meter on this system must be accurate and with 
large enough scale graduations so that the volumetric 
flow rate is indicated within ± 1%. For convenience 
and time saving, a fixed single reading meter such as 
a venturi meter or orifice meter is preferable to a 
multi-point traverse type instrument such as a Pitot 
tube. 

3. A means of regulating and effecting air flow through 
the tunnel. For usual calibrations of instruments used 
on heating, ventilating, and industrial exhaust systems, 
test velocities from approximately 50 to 8000 fpm are 
needed. Air flow regulation must be such that there is 
no disturbance in the test section. The regulating de- 
vice must be easily and precisely set to the desired 
velocities. The fan must have sufficient capacity to 
develop the maximum velocity in the test section 
against the static pressure of the entire system. 

To provide a satisfactory uniform flow in the test section, 
a bell-shaped stream! ine entry is necessary (Figure 9-5). There 
are various designs for this entry. One type is the elliptical 
approach in which curvature is similar to a one-quarter section 
of an ellipse in which the semi-major axis of the ellipse is 
equal to the duct diameter to which the entry is placed and the 
semi-minor axis is two-thirds of the semi-major axis. This 
type of entry can be made on a spinning lathe. 

Actually, any type of smooth curved, bell-shaped entry 
which directs the air into the duct over a 1 80° angle should be 
satisfactory. A readily available entry is atubaorSousaphone 
bell. This bell entry should be connected to a 5.5" diameter 
smooth, seamless plastic tube. Ridges, small burrs, or obstruc- 
tions should be filed so a smooth connection between horn 
and tube results. 

For calibrating larger instruments such as the lower veloc- 
ity swinging vane anemometer (Alnor velometer) and the 
rotating vane anemometer, a large rectangular test section of 
transparent plastic at least 2.5 ft 2 in cross-sectional area can 
be constructed with curved air foil inlets as shown in Figure 
9-6. A fine mesh screen placed deep in the enclosure will assist 
in providing a uniform air flow in the test section. 

A sharp-edged orifice, venturi meter, or a flow nozzle can 
be used as a metering device. Of these, the sharp-edged orifice 
has more resistance to flow but is more easily constructed, 
and it can be designed to be readily interchangeable for several 



orifice sizes. The orifice can be mounted between two flanged 
sections sealed with gaskets as shown in Figure 9-5. Each 
orifice should be calibrated using a standard Pitot tube and 
manometer prior to use. For velocity measurements below 
2,000 fpm, a micromanometer should be used. (96) 

Table 9-3 lists calculations for three sizes or orifices: 
1.400", 2.625" and 4.900" diameters. When the orifices are 
placed in a 7" diameter duct and made to the precise dimen- 
sions given, no calibration is needed and the tabulated data in 
the Table will give volumetric flow rates within ± 5% over 
the range of values shown for standard air density. 

A centrifugal fan with sufficient capacity to exhaust 1 , 100 
cfm at 10 "wg static pressure is needed for a wind tunnel with 
a 5.5" diameter test section using an orifice meter. Radial and 
backwardly inclined blade centrifugal fans are available with 
the required characteristics. The air flow can be changed with 
an adjustable damper at the discharge, a variable speed motor, 
or an adjustable drive on the fan. 

The air flow for a sharp-edged orifice with pipe taps located 
1" on either side of the orifice can be computed from the 
following equation for 2"- 14" diameter ducts: 



Q = 6KD 2 



[9.4] 



where: 

Q = volumetric flow rate, cfm 

K= coefficient of air flow 

D = orifice diameter, inches 

h = pressure drop across orifice, "wg 

p = density, Ib/m/ft 3 

The coefficient, K, is affected by the Reynolds number — 
a dimensionless value expressing flow conditions in a duct. 
The following equation gives a simplified method of calcu- 
lating Reynolds number for standard air: 

R-8.4DV [9.5] 

where: 

R= Reynolds number, dimensionless 
V = velocity of air through orifice, fpm 

The coefficient, K, can be selected from Table 9-4. (93) 

9.3.2 Use of Calibrating Wind Tunnel: Air velocity meas- 
uring instruments must be calibrated in the manner in which 
they are to be used in the field. Swinging vane and rotating 
vane anemometers are placed in the appropriate test section 
on a suitable support and the air velocity varied through the 
operating range of interest. Heated thermocouple instruments 
are calibrated in the same manner. Special Pitot tubes and duct 
probes of direct-reading instruments are placed through a 
suitable port in the circular duct section and the air velocity 
varied through the operating range of interest. Heated thermo- 



9-8 Industrial Ventilation 



STREAMLINE INLET 



3 HP MOTOR WITH VARIABLE DRIVE 
500 TO 3670 RPM 

ALTERNATE DAMPER -— \ 



ORIFICE - SEE DETAIL 



X 



32" ■• 

5 1/2" DIAM 



1 1 ..^-i-^- 



70" 



"1 



7 



Jc 



FLANGE 



\ 



7" DIAM 



PLASTIC TUBE 



* 

a 



STRAIGHTENERS 



TEST SECTION 

FOR HIGH VELOCITY METERS 
WITH SMALL TEST PROBES 
IN TEST AIR STREAM. 



M 



MANOMETER - 6 ; ' INCLINE 
15" VERTICAL 



FAN 



CALIBRATION WIND TUNNE 



PIPE TAPS — \— i 



r- 



- 20" 

SCREEN — , 



35" 




l 





7" DIAM 



FLANGE 



T 

Do 



SHARP EDGE ORIFICE 
1/8" STEEL PLATE 




GASKET 



V - BRACKET ON ROD 
TRANSPARENT PLASTIC 



ORIFICE DETAIL, 



TEST SECTION 

FOR LOW VELOCITY METERS WITH 
LARGE AREA IN TEST AIR STREAM. 



::rican conference 

OF GOVERNMENTAL 
STRIAL HYGIENISTS 



CALIBRATION WIND TUNNEL 



DATE _/-£# 



FIGURE 



9-5 



Testing of Ventilation Systems 9-9 



15 OR MORE 



r 



12" OR LESS 



iSrlh 



hF 7 

L - SUPPORT ROD 
RING STAND 



% 



TEST SECTION 
VELOMETER EXHAUST JET IN TEST SECTION 



^ 



(o) ^ 



Vm 15" OR MORE* 
IF STAND IS 
USED 



12" OR LESS -■ 



SHEET RUBBER SEAL 



l S 






TEST SECTION 



HEATED THERMOCOUPLE PROBE IN TEST SECTION 

SCREEN 




BRACKET 



TEST SECTION 
LARGE AIR METER IN TEST SECTION 



KEEP TEST SECTION ENTRANCE 
CLEAR OF OBSTRUCTIONS AND 
FREE OF DRAFTS 



AMERICAN CONFERENCE 

OF .GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



CALIBRATION 



DATE 



1-88 



FIGURE 



9-6 



9-10 



Industrial Ventilation 



TABLE 9-3. Orifice Flow Rate (scfm) Versus Pressure Differential (in. of water) 



AP 




ORIFICE SIZE 




AP 




ORIFICE SIZE 




AP 




ORIFICE SIZE 




"WC 


1.4" 


2.625 


4.90" 


"WC 


1.4" 


2.625" 


4.90" 


"WC 


1.4" 


2.625" 


4.90" 


0.02 






57.1 


1.22 


28.7 


101.4 


410.3 


4.10 


52.3 


185.3 


746 


0.04 




18.7 


78.8 


1.24 


28.9 


102.3 


413.6 


4.20 


52.9 


187.5 


755 


0.06 




22.8 


95.3 


1.26 


29.2 


103.1 


416.9 


4.30 


53.5 


189.7 


763 


0.08 




26.2 


109.2 


1.28 


29.4 


103.9 


420.1 


4.40 


54.1 


191.9 


772 


0.10 




29.3 


121.5 


1.30 


29.6 


104.7 


423.4 


4.50 


54.7 


194.0 


781 


0.12 




32.1 


132.6 


1.32 


29.8 


105.5 


426.5 


4.60 


55.3 


196.2 


789 


0.14 




34.6 


142.8 


1.34 


30.1 


106.3 


429.7 


4.70 


55.9 


198.3 


797 


0.16 




37.0 


152.3 


1.36 


30.3 


107.1 


432.9 


4.80 


56.5 


200.4 


806 


0.18 




39.2 


161.2 


1.38 


30.5 


107.9 


436.0 


4.90 


57.1 


202.4 


814 


0.20 




41.3 


169.6 


1.40 


30.7 


108.6 


439.1 


5.00 


57.6 


204.4 


822 


0.22 




43.3 


177.6 


1.42 


30.9 


109.4 


442.2 


5.10 


58.2 


206.5 


830 


0.24 




45.2 


185.2 


1.44 


31.2 


110.2 


445.2 


5.20 


58.8 


208.5 


838 


0.26 




47.0 


192.6 


1.46 


31.4 


110.9 


448.3 


5.30 


59.3 


210.4 


846 


0.28 




48.8 


199.6 


1.48 


31.6 


111.7 


451.3 


5.40 


59.9 


212.4 


854 


0.30 




50.5 


206.5 


1.50 


31.8 


112.4 


454.3 


5.50 


60.4 


214.3 


862 


0.32 




52.1 


213.0 


1.52 


32.0 


113.2 


457.2 


5.60 


61.0 


216.3 


869 


0.34 




53.7 


219.4 


1.54 


32.2 


113.9 


460.2 


5.70 


61.5 


218.2 


877 


0.36 




55.3 


225.6 


1.56 


32.4 


114.6 


463.1 


5.80 


62.0 


220.0 


884 


0.38 




56.8 


231.6 


1.58 


32.6 


115.4 


466.0 


5.90 


62.6 


221.9 


892 


0.40 




58.3 


237.5 


1.60 


32.8 


116.1 


468.9 


6.00 


63.1 


223.8 


899 


0.42 




59.7 


243.2 


1.62 


33.0 


116.8 


471.8 


6.10 


63.6 


225.6 


907 


0.44 




61.1 


248.8 


1.64 


33.2 


117.5 


474.7 


6.20 


64.1 


227.4 


914 


0.46 




62.4 


254.3 


1.66 


33.4 


118.2 


477.5 


6.30 


64.6 


229.2 


921 


0.48 




63.8 


259.6 


1.68 


33.6 


118.9 


480.3 


6.40 


65.1 


231.0 


928 


0.50 


18.5 


65.1 


264.9 


1.70 


33.8 


119.6 


483.1 


6.50 


65.6 


232.8 


935 


0.52 


18.8 


66.4 


270.0 


1.72 


34.0 


120.3 


485.9 


6.60 


66.1 


234.6 


942 


0.54 


19.2 


67.6 


275.0 


1.74 


34.2 


121.0 


488.7 


6.70 


66.6 


236.3 


949 


0.56 


19.5 


68.9 


280.0 


1.76 


34.4 


121.7 


491.5 


6.80 


67.1 


238.1 


956 


0.58 


19.9 


70.1 


284.8 


1.78 


34.6 


122.4 


494.2 


6.90 


67.6 


239.8 


963 


0.60 


20.2 


71.3 


289.6 


1.80 


34.8 


123.1 


496.9 


7.00 


68.1 


241.5 


970 


0.62 


20.6 


72.4 


294.3 


1.82 


35.0 


123.8 


499.7 


7.10 


68.5 


243.2 


977 


0.64 


20.9 


73.6 


298.9 


1.84 


35.2 


124.4 


502.4 


7.20 


69.0 


244.9 


984 


0.66 


21.2 


74.7 


303.4 


1.86 


35.4 


125.1 


505.0 


7.30 


69.5 


246.5 


990 


0.68 


21.5 


75.8 


307.9 


1.88 


35.5 


125.8 


507.7 


7.40 


69.9 


248.2 


997 


0.70 


21.8 


76.9 


312.3 


1.90 


35.7 


126.4 


510.4 


7.50 


70.4 


249.9 


1003 


0.72 


22.1 


78.0 


316.7 


1.92 


35.9 


127.1 


513.0 


7.60 


70.9 


251.5 


1010 


0.74 


22.4 


79.1 


320.9 


1.94 


36.1 


127.8 


515.6 


7.70 


71.3 


253.1 


1017 


0.76 


22.7 


80.2 


325.2 


1.96 


36.3 


128.4 


518.2 


7.80 


71.8 


254.7 


1023 


0.78 


23.0 


81.2 


329.3 


1.98 


36.5 


129.1 


520.8 


7.90 


72.2 


256.4 


1029 


0.80 


23.3 


82.2 


333.5 


2.00 


36.6 


129.7 


523.4 


8.00 


72.7 


257.9 


1036 


0.82 


23.6 


83.2 


337.5 


2.10 


37.5 


132.9 


536.2 


8.10 


73.1 


259.5 


1042 


0.84 


23.9 


84.2 


341.6 


2.20 


38.4 


136.0 


548.6 


8.20 


73.6 


261.1 


1048 


0.86 


24.1 


85.2 


345.5 


2.30 


39.3 


139.0 


560.8 


8.30 


74.0 


262.7 


1055 


0.88 


24.4 


86.2 


349.4 


2.40 


40.1 


142.0 


572.6 


8.40 


74.5 


264.2 


1061 


0.90 


24.7 


87.2 


353.3 


2.50 


40.9 


144.9 


584.3 


8.50 


74.9 


265.8 


1067 


0.92 


25.0 


88.1 


357.2 


2.60 


41.7 


147.8 


595.7 


8.60 


75.3 


267.3 


1073 


0.94 


25.2 


89.1 


361.0 


2.70 


42.5 


150.6 


606.9 


8.70 


75.7 


268.8 


1079 


0.96 


25.5 


90.0 


364.7 


2.80 


43.3 


153.3 


617.9 


8.80 


76.2 


270.4 


1085 


0.98 


25.8 


91.0 


368.4 


2.90 


44.0 


156.0 


628.6 


8.90 


76.6 


271.9 


1091 


1.00 


26.0 


91.9 


372.1 


3.00 


44.8 


158.7 


639.2 


9.00 


77.0 


273.4 


1097 


1.02 


26.3 


92.8 


375.7 


3.10 


45.5 


161.3 


649.6 


9.10 


77.4 


274.9 


1103 


1.04 


26.5 


93.7 


379.3 


3.20 


46.2 


163.8 


659.9 


9.20 


77.9 


276.4 


1109 


1.06 


26.8 


94.6 


382.9 


3.30 


46.9 


166.4 


670.0 


9.30 


78.3 


277.8 


1115 


1.08 


27.0 


95.5 


386.4 


3.40 


47.6 


16B.8 


679.9 


9.40 


78.7 


279.3 


1121 


1.10 


27.3 


96.3 


390.0 


3.50 


48.3 


171.3 


689.7 


9.50 


79.1 


280.8 


1127 


1.12 


27.5 


97.2 


393.4 


3.60 


49.0 


173.7 


699.3 


9.60 


79.5 


282.2 


1132 


1.14 


27.8 


98.1 


396.9 


3.70 


49.7 


170.1 


708.8 


9.70 


79.9 


283.6 


1138 


1.16 


28.0 


98.9 


400.3 


3.80 


50.3 


178.4 


718.2 


9.80 


80.3 


285.1 


1144 


1.18 


28.2 


99.8 


403.7 


3.90 


51.0 


180.7 


727.5 


9.90 


80.7 


286.5 


1150 


1.20 


28.5 


100.6 


407.0 


4.00 


51.6 


183.0 


736.6 


10.00 


81.1 


287.9 


1155 



Testing of Ventilation Systems 



9-1! 




DUCT WALL - DRILL ALL HOLES 1/16" D. OR LESS. MAINTAIN 
INNER SURFACE OF DUCT SMOOTH AND FLUSH. 



ONE HOLE RUBBER STOPPER, CONNECTING TUBE AND 
RUBBER HOSE. 



PERMANENT TAP -1/8" TEE WITH 1/8" PIPE PLUG, T-HANDLE AND 
CLEANING WIRE. EQUIP WITH 1/8" PLUG WHEN NOT IN USE. USE 
1/8" COUPLING IF USED FOR PERMANENT CONNECTION TO U-TUBE. 



1/8" PETCOCK AND COUPLING WITH 3/16" COPPER TUBE 
SOLDERED TO PETCOCK. USE CAREFULLY WHEN HIGH 
PRESSURES ARE MEASURED TO PREVENT GAUGE LIQUID 
SURGE ON OPENING COCK. 



FIGURE 9-7. Static tap connections 

couple instruments are calibrated in the same manner. Special 
Pitot tubes and duct probes of direct-reading instruments are 
placed through a suitable port in the circular duct section and 
calibrated throughout the operating range (Figure 9-6). (99) 

9.4 PRESSURE MEASUREMENT 

At any point in an exhaust system, three air pressures exist 
which can be compared to the atmospheric pressure immedi- 
ately surrounding the system. Typically, these pressures are 
measured in inches water gauge ("wg) and are related to each 
other as follows: 

TP-SP + VP [9.6] 

where: 

TP = total pressure, M wg 

SP = static pressure, "wg 

VP = velocity pressure, "wg 

Static pressure is that pressure which tends to burst or 
collapse a duct and is positive when the pressure is above 
atmospheric and negative when below atmospheric. Velocity 
pressure is the pressure resulting from the movement of air 
and is always positive. Total pressure is the algebraic sum of 
the static pressure and velocity pressure and can be either 
positive or negative (see Figure 9-8). 

9.4.1 Static Pressure is measured by a pressure measur- 
ing device, usually a simple U-tube manometer filled with oil, 
water, or other appropriate liquid and graduated in inches 
water gauge or similar reading pressure gauge. A vertical 
manometer is suitable for most static pressure measurements. 
The use of an inclined manometer will give increased accu- 
racy and permits reading of lower values. For field measure- 
ment, one leg of the manometer is open to the atmosphere and 



the other leg is connected with tubing held flush and tight 
against a small opening in the side of the pipe. Additional 
information concerning manometers and their construction 
can be found in References 9. 1 and 9.2. 

The location of the static pressure opening is usually not 
too important in obtaining a correct measurement except that 
one should avoid pressure measurement at the heel of an 
elbow or other location where static pressure may be incorrect 
because the direction of the velocity component is not parallel 
with the duct wall. It is usually advisable to drill 2-4 pressure 
holes at uniform distances around the duct in order to obtain 
an average and to detect any discrepancy in value. 

The static pressure opening should be flush with the inner 
surface of the pipe wall and there should be no burrs or 
projections on the inner surface. The hole should be drilled, 
not punched. A l/l6"-l/8" hole is usually satisfactory since 
the size is not too important except for some types of instru- 
ments where air actually flows through the device (see Figure 
9-7). The recommendations of the manufacturer concerning 
the size of the static pressure opening should be followed. A 
second method less likely to involve error is to use the static 
pressure element of a Pitot tube as shown in Figure 9-8. In 
use, the instrument must be pointed upstream and parallel to 
the duct for accurate measurement. 

9.4.2 Hood Static Pressure: The hood static pressure 
method of estimating air flow into an exhaust hood or duct is 
based on the principle of the orifice; i.e., the inlet opening 
simulating an orifice. This method is quick, simple, and 
practical. It is a fairly accurate estimation of the volumetric 
air flow in branch exhaust ducts if the static pressure or suction 
measurement can be made at a point one to three duct diame- 
ters of straight duct downstream from the throat of the exhaust 



9-12 Industrial Ventilation 



TOTAL PRESSURE = STATIC PRESSURE + VELOCITY PRESSURE 





TOTAL PRESSURE 
BELOW ATMOSPHERE 



STATIC PRESSURE 
BELOW ATMOSPHERE 



J 



VELOCITY PRESSURE 
ABOVE ATMOSPHERE 




AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRf AL HYGIENISTS 



Testing of Ventilation Systems 9-13 



TABLE 9-4. Values of K in Equation 9.10 for Different Orifice Diameters to Duct-Diameter Ratios (d/D) and Different Reynolds Numbers* 

Reynolds Number in Thousands 



d/D 


25 


50 


100 


230 


500 


1000 


10,000 


0.100 


0.605 


0.601 


0.598 


0.597 


0.596 


0.595 


0.595 


0.200 


0.607 


0.603 


0.600 


0.599 


0.598 


0.597 


0.597 


0.300 


0.611 


0.606 


0.603 


0.603 


0.601 


0.600 


0.600 


0.400 


0.621 


0.615 


0.611 


0,610 


0.609 


0.608 


0.608 


0.450 


0.631 


0.624 


0.619 


0.617 


0.615 


0.615 


0.615 


0.500 


0.644 


0.634 


0.628 


0.626 


0.624 


0.623 


0.623 


0.550 


0.663 


0.649 


0.641 


0.637 


0.635 


0.634 


0.634 


0.600 


0.686 


0.668 


0.658 


0.653 


0.650 


0.649 


0.649 


0.650 


0.717 


0.695 


0.680 


0.674 


0.670 


0.668 


0.667 


0700 


0.755 


0.723 


0.707 


0.699 


0.694 


0.692 


0.691 


0.750 


0.826 


0.773 


0.747 


0.734 


0.726 


0.723 


0.721 



*For duct diameters of 2" to 14" inclusive. 



inlet and if an accurate analysis of the hood entry loss can be 
made. 

This technique involves the measuring of hood static pres- 
sure by means of a U-tube manometer at one or more holes 
(preferably four, spaced 90° apart), one duct diameter down- 
stream from the throat for all hoods having tapers, and three 
duct diameters from the throat for flanged or plain duct ends. 
The holes should be drilled 1/1 6"— 1/8" in diameter or less; the 
holes should not be punched as inwardly projecting jagged 
edges of metal will disturb the air stream. The U-tube ma- 
nometer is connected to each hole in turn by means of a 
thick-walled soft rubber tube and the difference in the height 
of the water columns is read in inches. 

If an elbow intervenes between the hood and the suction- 
measurement location, the pressure loss caused by the elbow 
should be subtracted from the reading to indicate the suction 
produced by the hood and throat alone (see Chapter 5, Figure 

5-14). 

The values for hood entry loss coefficient (h e ) for various 
hood shapes are listed in Chapter 5, Figure 5-13. When the 
hood static pressure (SP h ) is known, the volumetric flow rate 
can be determined by the following equation: 



Q-4005AC e JSP h 



[9.8] 



Q-1096A 



SP h 



(1 + h e )p 



[9.7] 



where: 

Q = volumetric flow rate, cfm 
A = area of duct attached to hood, ft 2 
SP h = U-tube average manometer reading, "wg 
h e - hood entry loss coefficient 
p = actual gas density, Ib/m/ft 3 

For standard air, Equation 9.7 becomes: 



As noted above, the coefficient of entry, C e , can also be 
used in conjunction with SP h to determine Q. To facilitate 
repeated measurements, it is sometimes convenient to post the 
value of C e directly on the hood near the point where SP h is 
measured. 

9.4.3 Hood Static Pressure Interpretation: If the hood 
static pressure is known while a system is functioning prop- 
erly, its continued effectiveness can be assured so long as the 
original value is not changed. Any change from the original 
measurement can only indicate a change in velocity in the 
branch and, consequently, a change in volumetric flow 
through the hood. This relationship will be true unless: 1) a 
hood design change has affected the entrance loss; 2) there 
are obstructions or accumulations in the hood or branch ahead 
of the point where the hood static pressure reading was taken; 
or 3) the system has been altered or added to. Depending on 
the location of the obstruction in the duct system, restrictions 
of the cross-sectional area will reduce the air flow although 
hood suction may increase or decrease. 

Pressure readings vary as the square of the velocity or 
volumetric flow rate. To illustrate, an indicated reduction in 
static pressure readings of 30% would reflect a volumetric 
flow rate (or velocity) decrease of 6%. 

A marked reduction in hood static pressure often can be 
traced to one or more of the following conditions: 

1. Reduced performance of the exhaust fan caused by 
reduced shaft speed due to belt slippage, wear, or 
accumulation on rotor or casing that would obstruct air 
flow. 

2. Reduced performance caused by defects in the exhaust 
piping such as an accumu lation in branch or main ducts 



9-14 



Industrial Ventilation 



<'K 



iji 

4, 

I K -1 



5 IN.= 16 D 



2.500 IN.: 



D 



0.250 IN. — , 



-0.125 IN. DIA. 






IMI 



v/////M///////;jmm zm 



, k\N k\\\\\\\v\^^ 




0.312 IN.= 1 D 



0.937 IN RADIUS 



90' ± 1' 




8 HOLES - 0.04 IN. DIA. 
EQUALLY SPACED 
FREE FROM BURRS 



0.156 IN. RAD. 



NOSE SHALL BE FREE 
FROM NICKS AND BURRS. 



..INNER TUBING - APPROX. 
0.125 IN O.D. x 21 B&S GUAGE 



SECTION A-A 

NOTE :OTHER SIZES OF PITOT TUBES WHEN REQUIRED, MAY BE BUILT USING THE 
SAME GEOMETRIC PROPORTIONS WITH THE EXCEPTION THAT THE STATIC 
ORIFICES ON SIZES LARGER THAN STANDARD MAY NOT EXCEED 0.04 IN. IN 
DIAMETER. THE MINIMUM PITOT TUBE STEM DIAMETER RECOGNIZED 
UNDER THIS CODE SHALL BE 0.10 IN. IN NO CASE SHALL THE STEM 
DIAMETER EXCEED 1/30 OF THE TEST DUCT DIAMETER. 



"j-CV STATIC 



PRESSURE 



- OUTER TUBING 

0.312 IN O.D. x APPROX. 18 B&S GUAGE 



TOTAL PRESSURE 



FIGURE 9-9. Standard Pitot tubes 



due to insufficient conveying velocities, condensation 
of oil or water vapors on duct walls, adhesive charac- 
teristics of material exhausted, or leakage losses 
caused by loose clean-out doors, broken joints, holes 
worn in duct (most frequently in elbows), poor con- 
nection to exhauster inlet, accumulations in ducts or 
on fan blades. 

3 . Reduced air flow rate also can be charged to additional 
exhaust duct openings added to the system (sometimes 
systems are designed for future connections and more 
air than required is handled by present branches until 
future connections are made) or change of setting of 
blast gates in branch lines. Blast gates adjust the air 
distribution between the various branches. Tampering 
with the blast gates can seriously affect such distribu- 
tion and therefore they should be locked in place 
immediately after the system has been installed and its 
effectiveness verified. Fan volume control dampers 
also should be checked. 

4. Reduced volumetric flow may be caused by increased 
pressure loss through the dust collector due to lack of 
maintenance, improper operation, wear, etc. These 



effects will vary with the collector design. Refer to 
operation and maintenance instructions furnished with 
the collector or consult the equipment manufacturer. 

9.4.4 Velocity Pressure: For measuring velocity pressure 
to determine air velocity, a standard Pitot tube may be used. 
A large volume of research and many applications have been 
devoted to the subject of flow measurements by this instru- 
ment, which was developed by Henry Pitot in 1734 while a 
student in Paris, France. A standard Pitot tube (see Figure 9-9) 
needs no calibration if carefully made and the accuracy of 
velocity pressure readings obtained are considered to be ac- 
curate at velocities above 600 fpm (see Table 9-1). For more 
details concerning specifications and application of the Pitot 
tube, see the "Standard Test Code" published by the American 
Society of Heating, Refrigerating and Air Conditioning Engi- 
neers and the Air Moving and Conditioning Association/ 9 - 1, 9 - 4) 

The device consists of two concentric tubes — one meas- 
ures the total or impact pressure existing in the air stream; the 
other measures the static pressure only. When the annular 
space and the center tube are connected across a manometer, 
the difference between the total pressure and the static pres- 
sure is indicated on the manometer. This difference is the 



Testing of Ventilation Systems 9-15 



QQQQ Q 

CNOO^t CN ^r 

OO-CN K) 

oooo o 



J, 



LO XT ^tOO^t 

00 [\ LO^[\ 

co r-. oo cd en 

o 6 ooo 



f 

i a q 


Q 


Q 


a a 


| r- -t 


^ 


en 


UD K) 


I m m 


O 


cn 


-t -=* 


i en oo 


|\ 


CN 


*- o 



o o 




o 



o o 



10 POINT PIT0T TRAVERSE 
IN A CIRCULAR DUCT. 

(GREATER THAN 6" DIAM.) 
10 OR 20 LOCATIONS IN CENTERS 
OF EQUAL ANGULAR AREA. 



FIGURE 9-1 OA. 10-point Pilot traverse in a circular duct 

velocity pressure. 

The velocity pressure can be used to compute the velocity 
of the air stream if the density of the air is known. The 
following equation can be used: 



V-1096 |— 



where: 

VP = velocity pressure, "wg 
p = actual gas density, lb/m/ft 3 

Where air is at standard conditions (p 
Equation 9.9 becomes: 



[9.9] 



V - 4005VVP 



0.075 lb/m/ft 3 ), 



[9.10] 



For example, if the temperature of the air stream varies more 
than 30 F from standard air (70 F and 29.92 "Hg) or the altitude 
of the site is more than 1,000 feet above or below sea level or 
the moisture content of the air is 0.02 lb/lb of dry air or greater, 
the actual gas density (p) must be used. 

Velocity pressure versus velocity tables for standard air can 
be found in Chapter 5 (Tables 5-7A and 5-7B). These tables 
can be used for air at densities other than standard conditions 
by determining an equivalent velocity pressure. 



VP. 



VP 
df 



6 POINT PIT0T TRAVERSE 
IN A CIRCULAR DUCT. 

(6" DIAM. OR LESS) 

6 OR 12 LOCATIONS IN CENTERS 

OF EQUAL ANGULAR AREA. 



[9.11] 



FIGURE 9-1 OB. 6-point Pitot traverse in a circular duct 

where: 

VP e = equivalent velocity pressure, M wg 

VP m = measured velocity pressure, "wg 

df = density factor coefficient 

The equivalent VP then can be used in the velocity pressure 
versus velocity table selected to give the actual velocity at 
duct conditions. 

A number of techniques can be used to determine the 
volumetric flow rate at hood openings and at other points in 
an exhaust system using the fluid flow principles previously 
described. The method selected will depend on the degree of 
accuracy required, time available for testing, and the type of 
test data required. It is extremely important that measurements 
taken at the time of the tests include all necessary information 
to determine the gas density to permit the calculation of the 
actual velocity and volumetric flow rate. 

9.5 PITOT TRAVERSE METHOD 

Because the air flow in the cross-section of a duct is not 
uniform, it is necessary to obtain an average by measuring VP 
at points in a number of equal areas in the cross-section. The 
usual method is to make two traverses across the diameter of 
the duct at right angles to each other. Readings are taken at 
the center of annular rings of equal area (see Figures 9-10A 
and 9-1 OB). Whenever possible, the traverse should be made 
7 14 duct diameters or more downstream from any major air 



9-16 Industrial Ventilation 



VP MAX . 
10 



^ 



VP MAX. 



A: FULLY DEVELOPED VP DISTRIBUTION 



VP MAX. 



VP MAX. 




GOOD VP DISTRIBUTION. (ALSO SATISFACTORY FOR 
FLOW INTO FAN INLETS. BUT MAY BE UNSATIS- 
FACTORY FOR FLOW INTO INLET BOXES - MAY 
PRODUCE SWIRL IN BOXES.) 



VP MAX 
10 



VP MAX. 




VP MAX. — 




C: SATISFACTORY VP DISTRIBUTION - MORE THAN 
75% OF VP READINGS GREATER THAN VP MAX . 

10 



D: DO NOT USE! UNSATISFACTORY VP DISTRIBUTION - 
LESS THAN 75% OF VP READINGS GREATER THAN 
VP MAX . 
10 



VP MAX 



VP MAX. 




VP MAX. 

10 — *- 


— VP MAX.*- 




f 


"—"■"-% 


i r 




y 


35% 






1 


\ ' 


* \ 


) 


K-^ 


35% 






/ 


\y 


j 


r*-— ^^ 



E: DO NOT USE! UNSATISFACTORY VP DISTRIBUTION - 
LESS THAN 75% OF VP READINGS GREATER THAN 
VP MAX , 
10 



F: DO NOT USE! UNSATISFACTORY VP DISTRIBUTION - 
LESS THAN 75% OF VP READINGS GREATER THAN | 
VP MAX . 
10 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



VELOCITY PRESSURE 
DISTRIBUTIONS 



date n-90 



FIGURE 



9-11 



Testing of Ventilation Systems 9-17 



a © ; g 

O $ O Q 

o © o © 

©GO© 



PITOT TRAVERSE POINTS IN A RECTANGULAR 
DUCT. CENTERS OF 1 6 TO 64 EQUAL AREAS. 
LOCATIONS NOT MORE THAN 6" APART. 



FIGURE 9-12. Pitot traverse points in a rectangular duct 

disturbance such as an elbow, hood, branch entry, etc. Where 
measurements are made closer to disturbances, the results 
must be considered subject to some doubt and checked against 
a second location. If agreement within 10% of the two tra- 
verses is obtained, reasonable accuracy can be assumed and 
the average of the two readings used. Where the variation 
exceeds 10%, a third location should be selected and two air 
flows in the best agreement averaged and used. The use of a 
single centerline reading for obtaining average velocity is a 
very coarse approximation and is NOT recommended. 

The reason for the uncertainty and variation in measure- 
ments is the non-uniformity of air flow after a disturbance. 
Figure 9-1 1 shows some air flow patterns that could develop 
after a disturbance and the resulting difficulties in obtaining 
good reliable measures are evident. 

For round ducts 6" and smaller, at least 6 traverse points 
should be used. For round ducts larger than 6" diameter, at 
least 10 traverse points should be employed. The number of 
traverse locations on each diameter and the number of tra- 

TABLE 9-5. Distance from Wall of Round Pipe to Point Reading 
(nearest 1/8 inch) for 6-Point Traverse 



Duct 
DIA 


Ri 

.043 DIA 


R 2 
.146 DIA 


*3 

.296 DIA 


R 4 
.704 DIA 


Rs 

.854 DIA 


R 6 

.957 DIA 


3 


1/8 


1/2 


7/8 


21/8 


21/2 


2 7/8 


31/2 


1/8 


1/2 


1 


21/2 


3 


3 3/8 


4 


1/8 


5/8 


1 1/8 


2 7/8 


3 3/8 


3 7/8 


41/2 


1/4 


5/8 


13/8 


31/8 


3 7/8 


41/4 


5 


1/4 


3/4 


1 1/2 


31/2 


41/4 


4 3/4 


51/2 


1/4 


3/4 


15/8 


3 7/8 


4 3/4 


51/4 


6 


1/4 


7/8 


13/4 


41/4 


51/8 


5 3/4 



verse diameters required are determined by the need for 
accuracy and the symmetry of the measured values. Where 
uniform velocity pressure profiles exist, a single traverse 
along one diameter may be adequate. Where the values are 
moderately skewed, the use of two diameters is recom- 
mended. For greater accuracy, a third diameter should be 
used. Six, ten, and twenty point traverse points for various 
duct diameters are given in Tables 9-5, 9-6, and 9-7. To 
minimize errors, a Pitot tube smaller than the standard 5/16" 
O.D. should be used in ducts less than 12" in diameter. 

For square or rectangular ducts, the procedure is to divide 
the cross-section into a number of equal rectangular areas and 
measure the velocity pressure at the center of each. The 
number of readings should not be less than 16. However, 
enough readings should be made so the greatest distance 
between centers is approximately 6" (see Figure 9-12). 

The following data are essential and more detailed data may 
betaken if desired: 

• The area of the duct at the traverse location. 

e Velocity pressure at each point in the traverse. 

• Temperature of the air stream at the time and location 
of the traverse. 

The velocit y pressure r eadings obtained are converted to 
velocities and the velocities, not the velocity pressures, are 
averaged. Where more convenient, the square root of each of 
the velocity pressures may be averaged and this value then 
converted to velocity (average). The measured air flow is then 
the average velocity multiplied by the cross-sectional area of 
the duct (Q = VA). Where conditions are not standard, see 
"Corrections for Non- Standard Conditions." 

The Pitot tube cannot be used for measuring velocities less 
than 600 fpm in the field. It is susceptible to plugging in air 
streams with heavy dust and/or moisture loadings. A vibration 
free mounting is necessary if using a liquid manometer. See 
Reference 9. 10 for special instrumentation which can be used 
to measure low velocities. 

Modified Pitot Tubes: Modified Pitot tubes have been 
made in an effort to reduce plugging difficulties encountered 
in heavy dust streams or to increase manometer differentials 
enabling the measurement of lower velocities in the field. 
These are referred to as M S"-type (Staubscheide) tubes. They 
usually take the form of two relatively large impact openings, 
one facing upstream and the other facing downstream. Such 
tubes are useful when thick-walled ducts, such as boiler 
stacks, make it difficult or impossible to insert a conventional 
Pitot tube through any reasonably sized port opening. They 
require only initial calibration for all conditions. 

Measurements made with an "S"- type Pitot tube cannot be 
used directly. The tube first must be calibrated against a 
standard Pitot tube and the velocity pressure measured cor- 
rected to the actual velocity pressure. 



9-18 



Industrial Ventilation 



TABLE 9- 


6. Distance from Wall of Round Pipe to Point of Reading (nearest 


1/8 inch) for 10-Point Traverse 






DUCT DIA 


R, 
0.026DIA 


R 2 
0.082DIA 


R 3 

0.146DIA 


R 4 
0.226DIA 


R 5 

0.342DIA 


R 6 

0.658DIA 


R7 
0774DIA 


R 8 

0.854DIA 


R 9 

0.918DIA 


R 10 

0.974DIA 


4 


1/8 


3/8 


5/8 


7/8 


13/8 


2 5/8 


31/8 


3 3/8 


3 5/8 


3 7/8 


41/2 


1/8 


3/8 


5/8 


1 


1 1/2 


3 


31/2 


3 7/8 


41/8 


4 3/8 


5 


1/8 


3/8 


3/4 


11/8 


13/4 


31/4 


3 7/8 


41/4 


4 5/8 


4 7/8 


51/2 


1/8 


1/2 


3/4 


1 1/4 


17/8 


3 5/8 


41/4 


4 3/4 


5 


5 3/8 


6 


1/8 


1/2 


7/8 


13/8 


2 


4 


4 5/8 


51/8 


51/2 


5 7/8 


7 


1/8 


5/8 


1 


15/8 


2 3/8 


4 5/8 


5 3/8 


6 


6 3/8 


6 7/8 


8 


1/4 


5/8 


1 1/8 


13/4 


2 3/4 


51/4 


61/4 


6 7/8 


7 3/8 


7 3/4 


9 


1/4 


3/4 


1 1/4 


2 


31/8 


5 7/8 


7 


7 3/4 


81/4 


8 3/4 


10 


1/4 


7/8 


1 1/2 


21/4 


3 3/8 


6 5/8 


7 3/4 


81/2 


91/8 


9 3/4 


11 


1/4 


7/8 


15/8 


21/2 


3 3/4 


71/4 


81/2 


9 3/8 


10 1/8 


10 3/4 


12 


3/8 


1 


13/4 


2 3/4 


41/8 


7 7/8 


91/4 


101/4 


11 


115/8 


13 


3/8 


1 


17/8 


2 7/8 


41/2 


81/2 


101/8 


11 1/8 


12 


12 5/8 


14 


3/8 


1 1/8 


2 


31/8 


4 3/4 


91/4 


10 7/8 


12 


12 7/8 


13 5/8 


15 


3/8 


11/4 


21/4 


3 3/8 


51/8 


9 7/8 


115/8 


12 3/4 


13 3/4 


14 5/8 


16 


3/8 


1 1/4 


2 3/8 


3 5/8 


51/2 


101/2 


12 3/8 


13 5/8 


14 3/4 


15 5/8 


17 


1/2 


13/8 


21/2 


3 7/8 


5 3/4 


111/4 


131/8 


141/2 


15 5/8 


161/2 


18 


1/2 


1 1/2 


2 5/8 


41/8 


61/8 


117/8 


13 7/8 


15 3/8 


161/2 


171/2 


19 


1/2 


11/2 


2 3/4 


41/4 


61/2 


121/2 


14 3/4 


161/4 


17 1/2 


181/2 


20 


1/2 


15/8 


2 7/8 


41/2 


6 7/8 


131/8 


151/2 


171/8 


18 3/8 


191/2 


22 


5/8 


13/4 


31/4 


5 


71/2 


141/2 


17 


18 3/4 


201/4 


213/8 


24 


5/8 


2 


31/2 


51/2 


81/4 


15 3/4 


181/2 


201/2 


22 


23 3/8 


26 


5/8 


21/8 


3 3/4 


5 7/8 


8 7/8 


171/8 


201/8 


221/4 


23 7/8 


25 3/8 


28 


3/4 


21/4 


41/8 


6 3/8 


9 5/8 


18 3/8 


215/8 


23 7/8 


25 3/4 


271/4 


30 


3/4 


21/2 


4 3/8 


6 3/4 


101/4 


19 3/4 


231/4 


25 5/8 


271/2 


291/4 


32 


7/8 


2 5/8 


4 5/8 


71/4 


11 


21 


24 3/4 


27 3/8 


29 3/8 


311/8 


34 


7/8 


2 3/4 


5 


7 3/4 


115/8 


22 3/8 


261/4 


29 


311/4 


331/8 


36 


1 


3 


51/4 


81/8 


12 3/8 


23 5/8 


27 7/8 


30 3/4 


33 


35 


38 


1 


31/8 


51/2 


85/8 


13 


25 


29 3/8 


321/2 


34 7/8 


37 


40 


1 


31/4 


5 7/8 


9 


13 5/8 


26 3/8 


31 


341/8 


36 3/4 


39 


42 


1 1/8 


3 3/8 


61/8 


91/2 


14 3/8 


27 5/8 


321/2 


35 7/8 


38 5/8 


40 7/8 


44 


1 1/8 


3 5/8 


6 3/8 


10 


15 


29 


34 


37 5/8 


40 3/8 


42 7/8 


46 


11/4 


3 3/4 


6 3/4 


103/8 


15 3/4 


301/4 


35 5/8 


391/4 


421/4 


44 3/4 


48 


1 1/4 


4 


7 


10 7/8 


16 3/8 


315/8 


371/8 


41 


44 


46 3/4 



Testing of Ventilation Systems 



9-19 



TABLE 9-7. Distance from Wall of Round Pipe to Point of Reading (nearest 1/8 inch) for 20 Point Traverse 



Ouct 
dio 


R, 

0.013 D 


R 2 

0.0390 


#3 
0.0670 


R 4 

0.097 D 


R s 

a 1290 


Re 

0.1650 


Ry 

0.204 D 


Rs 

02500 


R s 

0.3060 


Rio 
0.3880 


R„ 

0.6120 


Rf2 
0.6940 


Rl3 
07500 


R /4 
0.7960 


R/5 
0.8350 


R/6 
0.87ID 


Rl7 
0.9030 


Rf8 
0.9330 


R/9 
096/0 


Reo 
0.9870 


40 


1/2 
24 1/2 


1 1/2 
27 3/4 


2 5/8 

30 


3 7/8 
31 7/8 


5 1/8 
33 3/8 


6 5/8 
34 7/8 


8 1/8 
36 1/8 


10 

37 3/8 


12 1/4 
38 1/2 


15 1/2 
39 1/2 


42 


1/2 
25 3/4 


1 5/8 
29 1/8 


2 7/8 
31 1/2 


4 1/8 
33 3/8 


5 3/8 
35 1/8 


6 7/8 
36 5/8 


8 5/8 
37 7/8 


10 1/2 
39 1/8 


12 7/8 
40 3/8 


16 1/4 
41 1/2 


44 


1/2 
26 7/8 


1 3/4 
30 1/2 


3 
33 


4 1/4 
35 


5 5/8 
36 3/4 


7 1/4 
38 3/8 


9 
39 3/4 


11 
41 


13 1/2 
42 1/4 


17 1/8 
43 1/2 


U6 


5/8 
28 1/8 


1 3/4 
31 7/8 


3 1/8 
34 1/2 


4 1/2 
36 5/8 


6 
38 3/8 


7 5/8 
40 


9 3/8 
41 1/2 


11 1/2 
42 7/8 


14 1/8 
44 1/4 


17 7/8 
45 3/8 


48 


5/8 
29 3/8 


1 7/8 
33 1/4 


3 1/4 
36 


4 5/8 
38 1/4 


6 1/4 
40 1/8 


7 7/8 

41 3/4 


9 3/4 
43 3/8 


12 

44 3/4 


14 3/4 
46 1/8 


18 5/8 
47 3/8 


50 


5/8 
30 5/8 


2 
34 5/8 


3 3/8 
37 1/2 


4 7/8 
39 3/4 


6 1/2 
41 3/4 


8 1/4 
43 1/2 


10 1/4 
45 1/8 


12 1/2 
46 5/8 


15 3/8 

48 


19 3/8 
49 3/8 


52 


5/8 
31 7/8 


2 
36 1/8 


3 1/2 
39 


5 
41 3/8 


6 3/4 
43 1/2 


8 1/2 
45 1/4 


10 5/8 
47 


13 

48 1/2 


15 7/8 

50 


20 1/8 
51 3/8 


54 


5/8 
33 


2 1/8 
37 1/2 


3 5/8 
40 1/2 


5 1/4 
43 


7 
45 1/8 


8 7/8 
47 


11 

48 3/4 


13 1/2 
50 3/8 


16 1/2 
51 7/8 


21 

53 3/8 


56 


3/4 
34 1/4 


2 1/8 
38 7/8 


3 3/4 

42 


5 3/8 

44 5/8 


7 1/4 
46 3/4 


9 1/4 
48 3/4 


11 3/8 

50 5/8 


14 

52 1/4 


17 1/8 
53 7/8 


21 3/4 
55 1/4 


58 


3/4 
35 1/2 


2 1/4 

40 1/4 


3 7/8 
43 1/2 


5 5/8 
46 1/8 


7 1/2 
48 1/2 


9 1/2 
50 1/2 


11 7/8 
52 3/8 


14 1/2 
54 1/8 


17 3/4 
55 3/4 


22 1/2 
57 1/4 


60 


3/4 
36 3/4 


2 3/8 
41 5/8 


4 
45 


5 7/8 
47 3/4 


7 3/4 
50 1/8 


9 7/8 
52 1/4 


12 1/4 

54 1/8 


15 
56 


18 3/8 
57 5/8 


23 1/4 
59 1/4 


62 


3/4 
37 7/8 


2 3/8 
43 


4 1/8 
46 1/2 


6 
49 3/8 


8 
51 3/4 


10 1/4 

54 


12 5/8 
56 


15 1/2 
57 7/8 


19 

59 5/8 


24 1/8 
61 1/4 


64 


3/4 
39 1/8 


2 1/2 
44 3/8 


4 1/4 
48 


6 1/4 
50 7/8 


8 1/4 
53 1/2 


10 1/2 

55 3/4 


13 1/8 

57 3/4 


16 

59 3/4 


19 5/8 
61 1/2 


24 7/8 
63 1/4 


66 


7/8 
40 3/8 


2 5/8 
45 3/4 


4 3/8 
49 1/2 


6 3/8 
52 1/2 


8 1/2 
55 1/8 


10 7/8 
57 1/2 


13 1/2 
59 5/8 


16 1/2 
61 5/8 


20 1/4 
63 3/8 


25 5/8 
65 1/8 


68 


7/8 
41 5/8 


2 5/8 
47 1/8 


4 1/2 
51 


6 5/8 
54 1/8 


8 3/4 
56 3/4 


11 1/4 
59 1/4 


13 7/8 
61 3/8 


17 

63 1/2 


20 7/8 
65 3/8 


26 3/8 
67 1/8 


70 


7/8 
42 7/8 


2 3/4 
48 1/2 


4 3/4 
52 1/2 


6 3/4 
55 3/4 


9 
58 1/2 


11 1/2 
61 


14 1/4 
63 1/4 


17 1/2 
65 1/4 


21 1/2 
67 1/4 


27 1/8 
69 1/8 


72 


7/8 
44 


2 3/4 
50 


4 7/8 
54 


7 
57 1/4 


9 1/4 

60 1 '* 


11 7/8 

62 3/4 


14 3/4 
65 


18 

67 1/8 


22 

69 1/4 


28 

71 1/8 


74 


7/8 
45 1/4 


2 7/8 
51 3/8 


5 
55 1/2 


7 1/8 
58 7/8 


9 1/2 
61 7/8 


12 1/8 
64 1/2 


15 1/8 
66 7/8 


18 1/2 
69 


22 5/8 
71 1/8 


28 3/4 
73 1/8 


76 


1 
46 1/2 


3 
52 3/4 


5 1/8 
57 


7 3/8 
60 1/2 


9 7/8 
63 1/2 


12 1/2 
66 1/8 


15 1/2 
68 5/8 


19 

70 7/8 


23 1/4 
73 


29 1/2 
75 


78 


1 
47 3/4 


3 
54 1/8 


5 1/4 
58 1/2 


7 1/2 
62 1/8 


10 1/8 
65 1/8 


12 7/8 
67 7/8 


15 7/8 
70 1/2 


19 1/2 
72 3/4 


23 7/8 
75 


30 1/4 
77 


80 


1 

49 


3 1/8 
55 1/2 


5 3/8 

60 


7 3/4 

63 5/8 


10 3/8 
66 7/8 


13 1/8 
69 5/8 


16 3/8 
72 1/4 


20 

74 5/8 


24 1/2 
76 7/8 


31 
79 



9-20 



Industrial Ventilation 






2-0=2" 

FIGURE 9-13. U-tube manometer 



/ + 1 =2 



6-4=2" 



Other modified forms of the Pitot tube are the air foil 
pitometer, the Pilot venturi, and the air speed nozzle, to name 
afew.< 9 - 7 « 9 - 8 > 

Pressure Sensors: Pressure sensors can be used in conjunc- 
tion with the pitot tube to measure pressures existing within 
ventilation systems. These devices are described below. 

U-Tube Manometer. The vertical U-tube (see Figure 9-13) 
is the simplest type of pressure gauge. Usually calibrated in 
inches water gauge, it is used with various fluid media such 
as alcohol, mercury, oil, water, kerosene and special manome- 
ter fluids. The U-tube may be used for either portable or 
stationary applications. Available commercial units offer a 
wide latitude in range, number of columns, and styles. Tubes 
are usually of all-plastic construction to minimize breakage. 
One leg may be replaced by a reservoir or well (well-type 
manometer) with the advantage of easier manometer reading. 



Inclined Manometer (Figure 9-14): Increased sensitivity 
and scale magnification is realized by tilting one leg of the 
U-tube to form an inclined manometer or draft gauge. The 
inclined manometer gives increased accuracy and permits 
lower readings. In commercial versions, only one tube of the 
small bore is used and the other leg is replaced by a reservoir. 
The accuracy of the gauge is dependent on the slope of the 
tubes. Consequently, the base of the gauge must be leveled 
carefully and the mounting must be firm enough to permit 
accurate leveling. The better draft gauges are equipped with 
a built-in level, leveling adjustment and, in addition, a means 
of adjusting the scale to zero. Some models include over-pres- 
sure safety traps to prevent loss of fluid in the event of pressure 
surges beyond the manometer range. 

A modification of the inclined manometer is the inclined- 
vertical gauge in which the indicator leg is bent or shaped to 
give both a vertical and inclined portion — the advantage is 




FIGURE 9-14. Inclined manometer 



Testing of Ventilation Systems 9-21 




FIGURE 9-1 5. Aneroid gauge 

smaller physical size for a given range while retaining the 
refined measurement afforded by the inclined manometer. As 
in the U-tube and inclined gauges, the commercial units 
available offer a wide choice in range, number of columns, 
and calibration units. 

Aneroid Gauges: This type of gauge is used as a field 
instrument in ventilation studies for measuring static, veloc- 
ity, or total pressure with a Pitot tube or for single tube static 
pressure measurements. A number of manufacturers offer 
gauges suitable for the measurement of the low pressures 
encountered in ventilation studies. Perhaps the best known of 
this type is the Magnehelic™ gauge (Figure 9-15). The prin- 
cipal advantages of this gauge can be listed as follows: easy 
to read, greater response than manometer types; very portable 
— small physical size and weight; absence of fluid means less 
maintenance; and mounting and use in any position is possible 
without loss of accuracy. Principal disadvantages are that the 
gauge is subject to mechanical failure, requires periodic cali- 
bration checks, and occasional recalibration. 

Electronic Aneroid Gauges: Commercial instruments are 
now available which will measure and record static pressure 
as well as integrate velocity pressure directly to velocity using 
the pressure sensing principles of an aneroid gauge. This type 
of instrument can be connected directly to a standard Pitot 
tube and used in the same manner as a U-tube manometer. 
The instruments are light in weight, easily hand-held, and can 
be equipped with an electronic digital display or print recorder 
with measurement data in either English or S.I. units. Because 
they are battery powered, periodic servicing is required as is 
calibration. 



9.6 CORRECTIONS FOR NON-STANDARD CONDITIONS 

Air velocities sometimes are measured at conditions sig- 
nificantly different from standard. If these conditions are 
ignored, serious errors can be introduced in the determination 
of the actual duct velocity and the volumetric flow rate(s) in 
the system. Elevation, pressure, temperature, and moisture 
content all affect the density of the air stream. The actual 
density present in the system must be used in either Equation 
9.2 or 9.9 to determine the actual velocity. 

Correction for changes in elevation, duct pressure, and 
temperature can be made independently of each other with 
reasonable accuracy. The individual correction coefficients 
are multiplied together to determine the change from standard 
air density. The actual air density becomes: 



p = 0.075 df and 
df = CF e CF p CF t 



[9.12] 
[9.13] 



where: 



CF e = correction for elevations outside the range of 

±1000 ft 
CF p = correction for local duct pressures greater than 

± 20 "wg 
CF t = correction for temperatures outside the range of 

40 to 100 F 

One exception to this general rule is when elevations sig- 
nificantly different from sea level are coupled with high 
moisture content. Where this occurs, a psychrometric chart 
based upon the barometric pressure existing at the elevation 
of concern should be used. See Chapter 5 for an explanation 
of the determination of density using a psychrometric chart 
when moisture content and temperature are significantly dif- 
ferent from standard. 

The correction coefficient for elevation, CF e , can be given by 



CF e -[l-(6.73x10^)(z)] 5258 



[9.14] 



where: 

z = elevation, ft. 

The correction coefficient for local duct pressure, CF p , can 
be given by 



CF n 



407 + SP 
407 



[9.15] 



where: 

SP = static pressure, "wg. (Note that the algebraic 
sign of SP is important.) 

The correction coefficient for temperature, CF t , can be 
given by 



CF t = 



530 
t + 460 



[9.16] 



9-22 



Industrial Ventilation 



where: 

t - dry-bulb temperature, F (Note: Algebraic sign 
oft must be used) 

Density factors (df) for various altitudes, barometric pres- 
sures, and temperature conditions are shown on Table 9-8. 

Example 1: A velocity pressure reading of 1.0 "wg was 
taken with a Pitot tube in a duct where the dry-bulb tempera- 
ture is 300 F, the moisture content is negligible and the static 
pressure is -23.5 "wg. The system is installed at an elevation 
of 5000 feet. What would the density and actual velocity be 
at that point? 

As the moisture content is unimportant, Equations 9.12 and 
9.13 can be used directly to determine the density. 

The individual correction coefficients can be found from 
Equations 9.14 through 9.16 as 



CF e = [1 - (6.73 x 1 0" 6 )(5000)J 



5.258 



0.84 



CF n 



407-23.5 
407 



CF,=- 



530 



0.94 



0.70 



300 + 460 
Then the density at this condition would be 

p = (0.075)(0.84)(0.94)(0.70) - 0.0415 Ibm/ft 3 
and the velocity from Equation 9.9 would be 



V = 1096 



I, 



10 



0415 



= 5380 fpm 



Note that an error of 26% would result if standard density had 
been assumed. 

Example 2; A swinging vane anemometer is used to 
determine the velocity in a duct at sea level where the dry bulb 
temperature is 250 F, the SP = -10 "wg and moisture is 
negligible. What is the actual duct velocity if the anemometer 
reading is 3150? 

The temperature correction coefficient is 0.75 from Equa- 
tion 9.16 and the density would be 

p = (0.75)(0.075) = 0.0563 Ibm / ft 3 
Therefore, the actual velocity in the duct would be 



V„ 



,Vdf 



V, 



corrected 



= 31 50 J— 
\0.i 



0.075 



0563 



(3150)(1.15)=3636fpm 



9.6.1 Example Traverse Calculations: Measurement of 
air velocity at non-standard conditions requires calculation of 
the true air velocity, accounting for difference in air density 
due to air temperature, humidity, and barometric pressure. 



The following calculations illustrate the method of calculation 
and the effect of varying air density. 

1 . Standard Conditions: 

Air Temp. .= 79 F; Wet-Bulb Temp. = 50 F 
Barometer = Std. (29.92 n Hg); 24" Duct Diameter 



Pitot Traverse #1 



Pitot Traverse #2 
(1 to Traverse #1) 



Traverse Pt. VP, 



M 



V* 



Traverse Pt. 



VPm Vs* 



1 


0.22 


1879 


1 


0.23 


1921 


2 


0.28 


2119 


2 


0.27 


2081 


3 


0.32 


2260 


3 


0.33 


2301 


4 


0.33 


2301 


4 


0.34 


2335 


5 


0.34 


2335 


5 


0.34 


2335 


6 


0.35 


2369 


6 


0.35 


2369 


7 


0.33 


2301 


7 


0.34 


2335 


8 


0.32 


2230 


8 


0.32 


2260 


9 


0.30 


2193 


9 


0.32 


2230 


10 


0.24 


1962 
21949 


10 


0.25 


2003 
22170 


Calculated from Equation 9.S 


) or Chapter ! 


5, Table 5-7 






Average 


Velocity, V e 


_ 2194S 


+ 22170 


44119 





20 



20 



= 2205.9 = 2206 fpm 

Q s = VA = 2206x3.142 = 6931.2 - 6931 scfm 

2. Elevated Temperature: 

Air Temp. - 150 F; Wet-Bulb Temp. - 80 F 
Barometer = Std.; 24" Outside Diameter Duct 



Pitot Traverse #1 



Pitot Traverse #2 
(1 to Traverse #1) 



Traverse Pt. 


VP M 


v s * 


Traverse Pt. 


VP M 


V s * 


1 


0.22 


2015 


1 


0.23 


2060 


2 


0.28 


2275 


2 


0.27 


2235 


3 


0.32 


2430 


3 


0.33 


2465 


4 


0.33 


2470 


4 


0.34 


2505 


5 


0.34 


2505 


5 


0.34 


2505 


6 


0.35 


2540 


6 


0.35 


2540 


7 


0.33 


2470 


7 


0.34 


2505 


8 


0.32 


2395 


8 


0.32 


2430 


9 


0.30 


2355 


9 


0.32 


2395 


10 


0.24 


2105 
23560 


10 


0.25 


2150 
23790 



•Calculated from Equation 9.9 or Chapter 5, Table 5-7 



Testing of Ventilation Systems 9-23 



To determine the air velocity at standard conditions (V s ) 
for each VP M , the density (p) can be calculated using Equa- 
tions 9.12 and 9.16: 

p - 0.075 x (530 -610) = 0.065lbm / ft 3 

Using Equation 9.11, each VP M is multiplied by the ratio 
0.075 ■+ 0.065 and the resulting V s values averaged. 



Average Velocity, V s = 



23560 + 23790 47350 



20 



20 

= 2368 fpm 
Q s = VA = 2368 x 3.142 = 7440.3 = 7440 scfm 

Short Method: 

Find: "standard velocity" average from measured VP S = 2206 
fpm (From # 1 ) 

VP for 2206 fpm = 0.30 (Equation 9.10); at 150 F, 
density = 0.075 - 0.87 = 0.065 lb/m/ft 3 

VP S = VPm x (0.075 - 0.065) = 0.346 = 0.35 "wg 

V s = 2370 fpm 

3 . Elevated Temperature and Moisture: 

Air Temp - 150 F; Wet-Bulb Temp. = 140 F 
Barometer = Std.; 24" Outside Diameter Duct 



Pitot Traverse #1 



Pilot Traverse #2 
(1 to Traverse #1) 



Traverse Pt. VP, 



M 



v* 



Traverse Pt. 



VP M 



9 
10 



0.22 

028 
0.32 
0.33 
0.34 
0.35 
0.33 
0.32 
0.30 
0.24 



2100 
2370 
2530 
2570 
2610 
2645 
2570 
2490 
2450 
2190 
24525 



1 
2 
3 
4 
5 
6 
7 
8 
9 
10 



023 
0.27 
0.33 
0.34 
0.34 
0.35 
0.34 
0.32 
0.32 
0.25 



2145 
2325 
2570 
2610 
2610 
2645 
2610 
2530 
2490 
2235 
24770 



^Calculated from Equation 9.9 or Chapter 5, Table 5-7 

To determine the air velocity at standard conditions (V s ) 
for each VP M , the air density (p) can be calculated using the 
psychrometric charts found in Chapter 5. 

p = 0.075 x 0.80 (density coefficient -mixture) 
= 0.06 lb/m/ft 3 



Using Equation 9.1 1, each VP M is multiplied by the ratio 
0.075 ^ 0.06 and the resulting V s values averaged. 



Average velocity, V s : 



24525 + 24770 



20 



49295 



-2460 



20 

= 2465 fpm 
(V s may be found also by the Short Method found in #2.) 



Q 



actual 



= V C A = 2465x3.142 



= 7745cfmofairandwatermixture 
Weight of mixture = Q s x 0.075 x d 

= 7745 x 0.075 x 0.80 = 465 lb 

From psychrometric charts, weight of water in mixture 
0.15 lb. H 2 0/lb dry air. 

Weight of dry air 

(weight of mixture) 
(weight of dry air + moisture per Ibm dry air) 

465 
1.15 

404lbm/min 



Q = 



404 



Stddensity 0.075 
Alternate Method: 



= 5387 scfm 



From Chapter 5, humid volume = 19.3 ft 3 of mixture/lb dry 
air (Interpolate). 

O 774*1 

Weight of dry air = -?*- = I±^± = 403 lb 
y y 19.3 19.3 

^ 403lbm _ _ _ 

CL = = 5373 scfm 

s 0.075 

4. High or Low Altitudes: 

Qs ~ V s x A where V s can be obtained from Equations 
9.2 and 9.9 in conjunction with Table 9.8 or Equations 
9.12 and 9.13. 

9.7 CHECK-OUT PROCEDURES 

The following procedure may be used on systems (see 
Figure 9- 1 6) that were designed to balance without the aid of 
blast gates. It is intended as an initial verification of the design 
computations and contractors* construction in new systems, 
but it may be used also for existing systems when design 
calculations are available or can be recomputed. It does not 
detect poor choices of design criteria such as low conveying 
or capture velocities, and consequently, will not reveal inade- 



9-24 



Industrial Ventilation 



TABLE 9-8. Air Density Correction Factor, df 















ALTITUDE RELATIVE TO SEA LEVEL, ft 














-5000 


-4000 


-3000 


-2000 


-1000 





1000 


2000 


3000 


4000 


5000 


6000 


7000 


8000 


9000 


10000 
















Barometric Pressure 
















"Hg 


35.74 


34.51 


33.31 


32.15 


31.02 


29.92 


28.86 


27.82 


26.82 


25.84 


24.89 


23.98 


23.09 


22.22 


21.39 


20.57 


"w 


486.74 469.97 453.67 437.84 422.45 407.50 392.98 378.89 365.21 : 


J51.93 339.04 326.54 314.42 302.66 291.26 280.21 


Temp., 


































F 














Density Factor, df 
















-40 


1.51 


1.46 


1.40 


1.36 


1.31 


1.26 


1.22 


1.17 


1.13 


1.09 


1.05 


1.01 


0.97 


0.94 


0.90 


0.87 





1.38 


1.33 


1.28 


1.24 


1.19 


1.15 


1.11 


1.07 


1.03 


1.00 


0.96 


0.92 


0.89 


0.86 


0.82 


0.79 


40 


1.27 


1.22 


1.18 


1.14 


1.10 


1.06 


1.02 


0.99 


0.95 


0.92 


0.88 


0.85 


0.82 


0.79 


0.76 


0.73 


70 


1.19 


1.15 


1.11 


1.07 


1.04 


1.00 


0.96 


0.93 


0.90 


0.86 


0.83 


0.80 


0.77 


0.74 


0.71 


0.69 


100 


1.13 


1.09 


1.05 


1.02 


0.98 


0.95 


0.91 


0.88 


0.85 


0.82 


0.79 


0.76 


0.73 


0.70 


0.68 


0.65 


150 


1.04 


1.00 


0.97 


0.93 


0.90 


0.87 


0.84 


0.81 


0.78 


0.75 


0.72 


0.70 


0.67 


0.65 


0.62 


0.60 


200 


0.96 


0.93 


0.89 


0.86 


0.83 


0.80 


0.77 


0.75 


0.72 


0.69 


0.67 


0.64 


0.62 


0.60 


0.57 


0.55 


250 


0.89 


0.86 


0.83 


0.80 


0.77 


0.75 


0.72 


0.69 


0.67 


0.64 


0.62 


0.60 


0.58 


0.55 


0.53 


0.51 


300 


0.83 


0.80 


0.78 


0.75 


0.72 


0.70 


0.67 


0.65 


0.62 


0.60 


0.58 


0.56 


0.54 


0.52 


0.50 


0.48 


350 


0.78 


0.75 


0.73 


0.70 


0.68 


0.65 


0.63 


0.61 


0.59 


0.57 


0.54 


0.52 


0.50 


0.49 


0.47 


0.45 


400 


0.74 


0.71 


0.69 


0.66 


0.64 


0.62 


0.59 


0.57 


0.55 


0.53 


0.51 


0.49 


0.48 


0.46 


0.44 


0.42 


450 


0.70 


0.67 


0.65 


0.63 


0.60 


0.58 


0.56 


0.54 


0.52 


0.50 


0.48 


0.47 


0.45 


0.43 


0.42 


0.40 


500 


0.66 


0.64 


0.61 


0.59 


0.57 


0.55 


0.53 


0.51 


0.49 


0.48 


0.46 


0.44 


0.43 


0.41 


0.39 


0.38 


550 


0.63 


0.61 


0.58 


0.56 


0.54 


0.52 


0.51 


0.49 


0.47 


0.45 


0.44 


0.42 


0.40 


0.39 


0.38 


0.36 


600 


0.60 


0.58 


0.56 


0.54 


0.52 


0.50 


0.48 


0.46 


0.45 


0.43 


0.42 


0.40 


0.39 


0.37 


0.36 


0.34 


700 


0.55 


0.53 


0.51 


0.49 


0.47 


0.46 


0.44 


0.42 


0.41 


0.39 


0.38 


0.37 


0.35 


0.34 


0.33 


0.31 


800 


0.50 


0.49 


0.47 


0.45 


0.44 


0.42 


0.41 


0.39 


0.38 


0.36 


0.35 


0.34 


0.32 


0.31 


0.30 


0.29 


900 


0.47 


0.45 


0.43 


0.42 


0.40 


0.39 


0.38 


0.36 


0.35 


0.34 


0.32 


0.31 


0.30 


0.29 


0.28 


0.27 


1000 


0.43 


0.42 


0.40 


0.39 


0.38 


0.36 


0.35 


0.34 


0.33 


0.31 


0.30 


0.29 


0.28 


0.27 


0.26 


0.25 



quate control due to this type of error. Agreement with design 
within ± 10% is considered acceptable. 

1. Determine volumetric flow in duct with a pitot tr- 
averse. If volumetric flow matches design, go to Step 
4; otherwise, continue with la. 

a. Check fan size against plan; 

b. Check fan speed and direction of rotation against 
design; 

c. Check fan inlet and outlet configuration againstplan. 

2. If a discrepancy is found and corrected, return to Step 
1 . If not, measure fan inlet and outlet static pressures 
and compute the fan static pressure. Using fan table, 
check flow, fan static pressure and fan speed (RPM). 
If agreement is acceptable although at some other 
operating point than specified, fan is satisfactory, and 
trouble is elsewhere in the system. Go to Step 3. 

3. If fan inlet static pressure is greater (more negative) 
than calculated in the design, proceed to Step 4. If fan 
outlet static pressure is greater (more positive) than 
design, proceed to Step 8. 

4. Measure hood static pressure on each hood and check 
against design. If correct, go to Step 10; otherwise, 
continue with Step 4a. 

a. Check size and design of hoods and slots against 
plan; 



b. Examine each hood for obstructions. 

5. After all hood construction errors and obstructions 
have been corrected, if hood static pressures are cor- 
rect, return to Step 1; if too low, proceed to Step 6. 

6. Measure static pressure at various junctions in ducts 
and compare with design calculations. If too high at a 
junction, proceed upstream until static pressures are 
too low and isolate the trouble. In an area where the 
loss exceeds design: 

a. Check angle of entries to junctions against plan; 

b. Check radii of elbows against plan; 

c. Check duct diameters against plan; 

d. Check duct for obstructions. 

If the static pressure is too low, proceed downstream 
and locate the trouble. 

7. After correcting all construction details which deviate 
from specifications, return to Step 1. 

8. Measure pressure differential across air cleaning de- 
vice and check against manufacturer's data. If loss is 
excessive, make necessary corrections and return to 
Step 1 . If loss is less than anticipated, proceed to Step 
8a. 

a. Check ducts, elbows, and entries as in Step 6a and 
6d. 



Testing of Ventilation Systems 9-25 




POINT 



MEASUREMENT 



LOCATION OF MEASUREMENT 



MEASUREMENT USE 



HOOD STATIC 
PRESSURE 



DISTANCE FROM HOOD - 

3 PIPE 0'S-FLANGED OR PLAIN HOOD 

1 0-TAPERED HOOD 



1. ESTIMATE FLOW:Q = 4005CeA^/SP h 

2. CHECK POINT FOR HOOD AND 
SYSTEM PROFORMANCE. 



VELOCITY AND 
STATIC PRESSURE 



BRANCH AND MAINS-PREFERABLY 7.5 
0'S STRAIGHT RUN DOWNSTREAM FROM 
NEAREST AIR DISTURBANCE 
( EL, ENTRY, ETC..) 



1. TRANSPORT VELOCITY 

2. EXHAUST VOLUME: Q=VA 

3. SP AS SYSTEM CHECK POINT 



CENTERLINE VP 



SMALL DUCTS LOCATION AS ABOVE. 
CENTERLINE VELOCITY READING ONLY. 



ROUND DUCT ONLY. USE ON SMALL 
DUCTS WHERE TRANSVERSE 
IMPRACTICAL OR WHERE APPROXIMATE 
VOLUME WANTED. 



STATIC, VELOCITY 
AND TOTAL 
PRESSURES 



INLET AND OUTLET OF FAN-ANY TWO 

OF THREE READINGS AT EACH 

LOCATION 



1. FAN STATIC AND TOTAL PRESSURES 



FSP= SP 
TP= SP - 



SP, - VP; 

SPi + VP 



VP: 



3. 



MOTOR SIZE OR GFM ESTIMATE 

CFM_ x JP__ _ 

BHP^ "^^g~~ M "f~ r'FAN 

SP AS SYSTEM CHECK POINT 



STATIC PRESSURE 



INLET AND OUTLET OF COLLECTOR 
DIFFERENTIAL PRESSURE 



1. COMPARE PRESSURE DROP WITH 
NORMAL OPERATING RANGE 

2. CHECKPOINTS FOR MAINTAINENCE. 
READINGS ABOVE OR BELOW 
NORMAL INDICATE PLUGGING, WEAR 
OR DAMAGE TO COLLECTOR 
ELEMENTS, NEED OF CLEANING 



IN ADDITION TO THE ABOVE, FACE VELOCITY (HOOD FACE) AND CAPTURE VELOCITY (POINT OF 
CONTAMINANT DISPERSION ) MEASUREMENTS ARE USUALLY MADE TO DEFINE HOOD PERFORMANCE. 
OBSERVATION OF AIR FLOWS SURROUNDING EXHAUST OPENINGS MAY BE VISUALLY AUGMENTED BY 
USE OF SMOKE GENERATORS, TRAILS, AND STREAMERS. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SAMPLE SYSTEM 



DATE 



1-88 



FIGURE 



9-16 



9-26 



Industrial Ventilation 



PLANT 

OPERATION EXHAUSTED. 



.DEPT.. 



. DATE . 

_ BY- 



LINE SKETCH SHOWING POINTS OF MEASUREMENT 



DATE SYSTEM INSTALLED 



HOOD AND TRANSPORT 


VELOCITY 








POINT 


DUCT 


VP 
IN. H 2 


SP 

IN. H 2 


FPM 

(Tbi. 9-1) 


CFM 
Q =VA 


REMARKS 


D 


AREA 

(Tbi. 5-5) 



















































































































































PITOT TRAVERSE 

PITOT READINGS- SEE TABLES 9-1 TO 9-4 



FAN 
TYPE 
SIZE 



POINTS 


VP 


VEL 


VP 


VEL. 


VP 


VEL 


1 














2 














3 














4 














5 














6 














7 














8 














9 














10 














TOTAL VEL. 












AVERAGE VEL. 








CFM 









POINT 


DIA. 


SP 


VP 


TP 


CFM 


INLET 












OUTLET 













FAN SP_ 
MOTOR 



(SEE SECTION 6) 



NAMF 


SI7F 


HP F 


I W 


COLLECTOR 
TYPF A- SI7F 



POINT 


DIA. 


SP 


ASP 


INLET 








OUTLET 









NOTES 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SURVEY FORM 



date l-Qd 



FIGURE g^2"/ 



Testing of Ventilation Systems 9-27 



b. Check system discharge type and dimensions 
against plans. 

9. If errors are found, correct, and return to Step 1. If no 
errors can be detected, recheck design against plan, 
recalculate, and return to Step 1 with new expected 
design parameters. 

10. Measure control velocities at all hoods where possible. 
If control is inadequate, redesign, or modify hood. 

1 1 . The above process should be repeated until all defects 
are corrected and hood static pressures and control 
velocities are in reasonable agreement with design. 
The actual hood static pressures should then be re- 
corded for use in periodic system checks. A file should 
be prepared containing the following documents: 

• System plan 

• Design calculations 

• Fan rating table 

• Hood static pressures after field measurement 

• Maintenance schedule 

• Periodic hood static pressure measurement log 

• Periodic maintenance log 

9.7.1 Difficulties Encountered in Field Measurement: 

The general procedures and instrumentation for the measure- 
ment of air flow have been previously discussed. However, 
special problems connected with air flow necessitate a some- 
what more detailed discussion. 

Some of these special problems are as follows: 

1. Measurement of air flow in highly contaminated air 
which may contain corrosive gases, dusts, fumes, 
mists, or products of combustion. 

2. Measurement of air flow at high temperatures. 

3. Measurement of air flow in high concentrations of 
water vapor and mist. 

4. Measurement of air flow where the vel ocity is very low 
(see Reference 9.ll). 

5. Measurement of air flow in locations of turbulence and 
non-uniform air flow; e.g., discharge of cupolas, loca- 
tions near bends, enlargements or discharges from 
exhaust fans. 

6. Measurement of air flow in connection with isokinetic 
sampling when the velocity is constantly changing. 

Selection of Instruments: The selection of the proper 
instrument will depend on the range of air flow to which the 
instrument is sensitive; its vulnerability to high temperatures, 
corrosive gases, and contaminated atmospheres; its portabil- 
ity and ruggedness and the size of the measuring probe relative 



to the available sampling port. A brief summary of the char- 
acteristics of a few of the instruments which can be used is 
given in Table 9.1. 

In many cases, conditions for air flow measurement are so 
severe that it is difficult to select an instrument. Generally 
speaking, the Pitot tube is the most serviceable instrument; it 
has no moving parts, is rugged, and will stand high tempera- 
tures and corrosive atmospheres when it is made of stainless 
steel. It is subject to plugging, however, when it is used in a 
dusty atmosphere. It cannot be used for measurement of low 
velocities. A special design of Pitot tube can be used for dusty 
atmospheres. In many cases, it is difficult to set up an inclined 
manometer in the field because many readings are made from 
ladders, scaffolds and difficult places. This greatly limits the 
lower range of the Pitot tube. A mechanical gauge can be used 
in place of a manometer. A mechanical gauge is estimated to 
be accurate to 0.02 "wg with proper calibration. 

For lower velocities, the swinging-vane anemometer pre- 
viously described can be used if conditions are not too severe. 
The instrument can be purchased with a special dust filter 
which allows its use in light dust loadings. It can be used in 
temperatures up to 100 F if the jet is exposed to the high 
temperature gases only for a very short period of time (30 
seconds or less). It cannot be used in corrosive gases. If the 
very low velocity jet is used, a hole over 1 " in diameter must 
be cut into the duct or stack. 

For very low velocities, anemometers utilizing the heated 
thermocouple principle can be used under special conditions. In 
most cases, these anemometers cannot be used in temperatures 
above 300 F. Contact the manufacturer to determine to what 
degree the thermocouple probe will withstand corrosive gases. 

In sampling work where a match of velocities in the sam- 
pling nozzle and air stream under changing velocities is 
required, the null method is sometimes used. This method 
uses two static tubes or inverted impact tubes, one located 
within the sampling nozzle and the other in the air stream. 
Each is connected to a leg of the manometer; the sampling 
rate is adjusted until the manometer reading is zero. 

REFERENCES: 

9.1 American Society of Heating, Refrigerating and Air 
Conditioning Engineers: 1 985 Fundamentals Volume. 
ASHRAE, Atlanta, GA (1985). 

9.2 Air Moving and Control Association, Inc.: Field Per- 
formance Measurements, Publication 203-8 1 . AMCA, 
Arlington Heights, IL (1981). 

9.3 Brandt, A.D.: Industrial Health Engineering, John 

Wiley and Sons, New York (1947). 

9.4 Air Movement and Control Association, Inc.: Test 
Code for Air Moving Devices, AMCA Standard 210- 
S6. AMCA, Arlington Heights, IL (1986). 

9.5 Farant, J.P; McKinnon, D.L.; McKenna, T.A.: Tracer 



9-28 



Industrial Ventilation 



Gases as a Ventilation Tool: Methods and Instrumen- 
tation. In: Ventilation '85 — Proceedings of the First 
International Symposium on Ventilation for Contami- 
nant Control, pp. 263-274. Elsevier Press, Amsterdam, 
The Netherlands (1986). 

9.6 American Society of Mechanical Engineers: Power 
Test Codes, Chapter 4, Flow Measurement. ASME, 
Lubbock, TX (1959). 

9.7 First, M.D.; Silverman, L.: Airfoil Pitometer, Indus- 
trial and Engineering Chemistry, 42:301-308 (Febru- 
ary 1950). 

9.8 American Society of Mechanical Engineers: Fluid 



Meters — Their Theory and Applications. ASME, 
Lubbock, TX (1959). 

9.9 Hama, G.: A Calibrating Wind Tunnel for Air Meas- 
uring Instruments. Air Engr. 41:18-20 (December, 
1967). 

9.10 Hama, G: Calibration of Alnor Velometers, American 
Industrial Hygiene Association Journal (December 
1958). 

9. 1 1 Hama, G.; Curley, L.S.: Instrumentation for the Meas- 
urement of Low Velocities with a Pitot Tube. Air 
Engineering (July 1967) and American Industrial Hy- 
giene Association Journal (May-June 1967). 



Chapter 10 

SPECIFIC OPERATIONS 



The following illustrations of hoods for specific operations 
are intended as guides for design purposes and apply to usual 
or typical operations. In most cases, they are taken from 
designs used in actual installations of successful local exhaust 
ventilation systems. All conditions of operation cannot be 
categorized, and because of special conditions (i.e., cross- 
drafts, motion, differences in temperature, or use of other 
means of contaminant suppression), modifications may be in 



order. 

Unless it is specifically stated, the design data are not to be 
applied indiscriminately to materials of high toxicity, e.g., 
beryllium and radioactive materials. Thus the designer may 
require higher or lower air flow rates or other modifications 
because of the peculiarities of the process in order to ade- 
quately control the air contaminant. 



Group 


Operation 


Print No. 


Old No. 


Page 


10.10 Cleanrooms 








10-6 




Cleanroom Ducted Module 


VS-10-01 




10-8 




Cleanroom Pressurized Plenum 


VS-10-02 




10-9 




Cleanroom Return Air Arrangements 


VS-10-03 




10-10 


10.15 Filling Operations 








10-11 




Barrel Filling 


VS-15-01 


< VS-303 


10-12 




Bag Filling 


VS-15-02 


VS-301 


10-13 




Bag Tube Packer 


VS-15-03 


VS-302 


10-14 




Weigh Hood Assembly — Dry Material 


VS-15-10 




10-15 




Weigh Hood Details — Dry Material 


VS-15-11 




10-16 




Toxic Material Bag Opening 


VS-15-20 


VS-1001 


10-17 




Shaft Seal Enclosure 


VS-15-21 


VS-210 


10-18 




Sampling Box 


VS-15-30 


VS-211 


10-19 


10.20 Foundry Operations 








10-20 




Foundry Shakeout — Enclosing 


VS-20-01 


VS-110 
VS-112 


10-21 




Foundry Shakeout — Side Draft 


VS-20-02 


VS-110 
VS-111 


10-22 




Foundry Shakeout 


VS-20-03 


VS-112 


10-23 




Shell Core Making 


VS-20-10 


VS-114 


10-24 




Core Making Machine — Small Rollover Type 


VS-20-11 


VS-115 


10-25 


10.25 Gas Treatment 








10-26 




Fumigation Booth 


VS-25-01 


VS-921 


10-27 




Fumigation Booth Notes 


VS-25-02 


VS-921 .1 


10-28 




Ethylene Oxide Sterilizers 


VS-25^10 




10-29 




Ethylene Oxide Sterilizer Notes 


VS-25-11 




10-30 




Ethylene Oxide Sterilizer Hood Details 


VS-25-12 




10-31 



10-2 



Industrial Ventilation 



Group 


Operation 


Print No. 


Old No. 


Page 


10.30 Kitchen Equipment 








10-32 




Dishwasher Ventilation 


VS-30-01 


VS-912 


10-33 




Kitchen Range Hoods 


VS-30-10 


VS-910 


10-34 




Kitchen Range Hood 


VS-30-11 


VS-911 


10-35 




Charcoal Broiler & Barbeque Pit Ventilation 


VS-30-12 


VS-913 


10-36 


10.35 Laboratory Ventilation 








10-37 




Typical Laboratory Hood 


VS-35-01 


VS-203 


10-40 




General Use Laboratory Hood Notes 


VS-35-02 


VS-205 


1041 




Perchloric Acid Hood Notes 


VS-35-03 


VS-205.1 


10-42 




Work Practices for Laboratory Hoods 


VS-35-04 


VS-205.2 


10-43 




Biological Safety Cabinet — Class II, Type A 


VS-35-10 




10-44 




Biological Safety Cabinet — Class II, Type B 


VS-35-11 




10-45 




Dry Box or Glove Hood for High Toxicity & 
Radioactive Materials 


VS-35-20 


VS-202 


10-46 




Horizontal Laminar Flow Clean Bench (Product 
Protection Only) 


VS-35-30 


VS-918.2 


10-47 




Vertical Laminar Flow Clean Bench (Product 
Protection Only) 


VS-35-31 


VS-918.1 


10-48 




Specialized Laboratory Hood Designs 


VS-35-40 


VS-206 


10-49 




Oven Exhaust 


VS-35-41 




10-50 


10.40 Low Volume-High Velocity Exhaust 
Systems 








10-51 




Extractor Head for Cone Wheels and Mounted 
Points 


VS-40-01 


VS-801 


10-52 




Hood for Cup Type Surface Grinder and Wire 
Brushes 


VS-40-02 


VS-802 


10-53 




Pneumatic Chisel Sleeve 


VS-40-03 


VS-803 


10-54 




Extractor Head for Small Radial Grinders 


VS-40-04 


VS-804 


10-55 




Extractor Hood for Disc Sander 


VS-40-05 


VS-805 


10-56 




Extractor Tool for Vibratory Sander 


VS-40-06 


VS-806 


10-57 




Typical System Low Volume-High Velocity 


VS-40-20 


VS-807 


10-58 


10.45 Machining 








10-59 




Metal Cutting Bandsaw 


VS-45-01 


VS-418 


10-60 




High Toxicity Materials Milling Machine Hood 


VS-45-02 


VS-209 


10-61 




Metal Shears High Toxicity Materials 


VS-45-03 


VS-208 


10-62 




Cold Heading Machine Ventilation 


VS-45-04 


VS-919 


10-63 




Lathe Hood 


VS-45-05 


VS-207 


10-64 


10,50 Material Transport 








10-65 




Bucket Elevator Ventilation 


VS-50-01 


VS-305 


10-66 



Specific Operations 10-3 



Group 


Operation 


Print No. 


Old No. 


Page 




Bin & Hopper Ventilation 


VS-50-10 


VS-304 


10-67 




Conveyor Belt Ventilation 


VS-50-20 


VS-306 


10-68 




Toxic Material Belt Conveying Head Pulley 


VS-50-21 


VS-1002 


10-69 




Toxic Material Conveyor Belt Loading 


VS-50-22 




10-70 




Rail Loading 


VS-50-30 




10-71 




Truck Loading 


VS-50-31 




10-72 


10.55 Metal Melting Furnaces 








10-73 




Melting Furnace Crucible, Non-Tilt 


VS-55-01 


VS-103 


10-74 




Melting Furnace, Tilting 


VS-55-02 


VS-106 


10-75 




Melting Furnace — Electric, Top Electrode 


VS-55-03 


VS-105 


10-76 




Melting Furnace — Electric, Rocking 


VS-55-04 


VS-104 


10-77 




Melting Pot & Furnace 


VS-55-05 


VS-906 


10-78 




Crucible Melting Furnace — High Toxicity Material 


VS-55-06 


VS-201 


10-79 




Induction Melting Furnace— Tilting 


VS-55-07 




10-80 




Pouring Station 


VS-55-10 


VS-109 


10-81 




Fixed Position Die Casting Hood 


VS-55-20 


VS-904 


10-82 




Mobile Hood, Die Casting 


VS-55-21 


VS-905 


10-83 


10.60 Mixing 








10-84 




Mixer and Muller Hood 


VS-60-01 


VS-107 


10-85 




Air Cooled Mixer and Muller 


VS-60-02 


VS-108 


10-86 




Banbury Mixer 


VS-60-10 


VS-901 


10-87 




Rubber Calendar Rolls 


VS-60-11 


VS-902 


10-88 




Roller Mill Ventilation 


VS-60-12 


VS-902.1 


10-89 


10.65 Movable Exhaust Hoods 








10-90 




Moveable Exhaust Hoods 


VS-65-01 




10-91 




Granite Cutting and Finishing 


VS-65-02 


VS-909 


10-92 




Hawley Trav-L-Vent Perspective Layout 


VS-65-03 




10-93 


1070 Open Surface Tanks 








10-94 




Open Surface Tanks 


VS-70-01 


VS-503 


10-104 




Open Surface Tanks 


VS-70-02 


VS-503.1 


10-105 




Push-Pull Hood Design Data for Widths Up to 10' 


VS-70-10 


VS-504 


10-106 




Push-Pufl Hood Design Data 


VS-70-11 


VS-504.1 


10-107 




Push Nozzle Manifold Pressure 


VS-70-12 


VS-504.2 


10-108 




Solvent Degreasing Tanks 


VS-70-20 


VS-501 


10-109 




Solvent Vapor Degreasing 


VS-70-21 


VS-501.1 


10-110 


10.75 Painting Operations 








10-111 




Large Paint Booth 


VS-75-01 


VS-603 


10-112 



10-4 



Industrial Ventilation 



Group 


Operation 


Print No. 


Old No. 


Page 




Small Paint Booth 


VS-75-02 


VS-604 


10-113 




Trailer Interior Spray Painting 


VS-75-03 


VS-605 


10-114 




Large Drive-Through Spray Paint Booth 


VS-75-04 


VS-606 


10-115 




Paint Booth Vehicle Spray 


VS-75-05 


VS-601 


10-116 




Dip Tank 


VS-75-06 


VS-502 


10-117 




Drying Oven Ventilation 


VS-75-20 


VS-602 


10-118 




Paint Mix Storage Room 


VS-75-30 




10-119 


10,80 Mechanical Surface Cleaning and 
Finishing 








10-120 




Abrasive Blasting Room 


VS-80-01 


VS-101 


10-121 




Abrasive Blasting Cabinet 


VS-80-02 


VS-101.1 


10-122 




Tumbling Mills 


VS-80-03 


VS-113 


10-123 




Grinding Wheel Hood — Surface Speeds Above 
6500 sfpm 


VS-80-10 


VS-4111 


10-124 




Grinding Wheel Hood — Surface Speeds Below 
6500 sfpm 


VS-80-11 


VS-411 


10-125 




Surface Grinder 


VS-80-12 


VS-417 


10-126 




Core Grinder 


VS-80-13 


VS-102 


10-127 




Vertical Spindle Disc Grinder 


VS-80-14 


VS-410 


10-128 




Horizontal Double-Spindle Disc Grinder 


VS-80-15 


VS-408 


10-129 




Swing Grinder 


VS-80-16 


VS-414 


10-130 




Abrasive Cut-Off Saw 


VS-80-17 


VS-401 


10-131 




Hand Grinding Bench 


VS-80-18 


VS-412 


10-132 




Portable Chipping and Grinding Table 


VS-80-19 


VS-413 


10-133 




Manual Buffing and Polishing 


VS-80-30 


VS-406 


10-134 




Buffing Lathe 


VS-80-31 


VS-407 


10-135 




Backstand idler Polishing Machine 


VS-80-32 


VS-402 


10-136 




Straight Line Automatic Buffing 


VS-80-33 


VS-405 


10-137 




Circular Automatic Buffing 


VS-80-34 


VS-404 


10-138 




Metal Polishing Belt 


VS-80-35 


VS-403 


10-139 


10.85 Vehicle Ventilation 








10-140 




Service Garage Ventilation — Overhead 


VS-85-01 


VS-907 


10-141 




Service Garage Ventilation — Underfloor 


VS-85-02 


VS-908 


10-142 




Exhaust System Requirements for Typical Diesel 
Engines Under Load 


VS-85-03 


VS-908.2 


10-143 




Ventilated Booth for .Radiator Repair Soldering 


VS-85-10 




10-144 


10.90 Welding and Cutting 








10-145 




Welding Ventilation Bench Hood 


VS-90-01 


VS-416 


10-146 




Welding Ventilation — Movable Exhaust Hoods 


VS-90-02 


VS-416.1 


10-147 



Specific Operations 10-5 



Group 


Operation 


Print No. 


Old No. 


Page 




Production Line Welding Booth 


VS-90-03 




10-148 




Torch Cutting Ventilation 


VS-90-10 


VS-916 


10-149 




Robotic Application 


VS-90-20 




10-150 




Metal Spraying 


VS-90-30 


VS-415 


10-151 


10.95 Woodworking 








10-152 




Band Saw 


VS-95-01 


VS-706 


10-153 




Floor Table Saw 


VS-95-02 




10-154 




Radial Arm Saw 


VS-95-03 


VS-709 


10-155 




Swing Saw 


VS-95-04 


VS-707 


10-156 




Table Saw Guard Exhaust 


VS-95-05 




10-157 




Single Drum Sander 


VS-95-10 


VS-705 


10-158 




Multiple Drum Sander 


VS-95-11 


VS-704 


10-159 




Disc Sanders 


VS-95-12 


VS-703 


10-160 




Optional Jet Stripper for Disk Sander 


VS-95-12a 




10-161 




Horizontal Belt Sanders 


VS-95-13 


VS-702 


10-162 




Horizontal Belt Sander, Push-Pull System 


VS-95-14 


VS-702.1 


10-163 




Jointers 


VS-95-20 


VS-701 


10-164 




Exhaust Plenum Retrofit for Orbital Hand Sander 


VS-95-30 




10-165 




Auxiliary Exhaust Retrofit for Air Powered Orbital 
Hand Sander 


VS-95-31 




10-166 


10.99 Miscellaneous Operations 








10-168 




Screens 


VS-99-01 


VS-307 


10-169 




Table Slot 


VS-99-02 


VS-505 


10-170 




Canopy Hood 


VS-99-03 


VS-903 


10-171 




Indoor Pistol and Small Bore Rifle Range 
Ventilation 


VS-99-04 


VS-914 


10-172 




Fiuidized Beds 


VS-99-05 


VS-915 


10-173 




Outboard Motor Test 


VS-99-06 


VS-920 


10-174 




Mortuary Table 


VS-99-07 




10-175 




Furniture Stripping Tank 


VS-99-08 




10-176 



10-6 



Industrial Ventilation 



10.10 CLEANROOMS 

U.S. Federal Standard 209E< 10101) establishes standard 
classes of air cleanliness for airborne particulate levels in 
cleanrooms and clean zones. This standard is issued by the 
General Services Administration of the United States. While 
nominally a publication for use by federal agencies, FED- 
STD-209E has been adopted by American industry. It pre- 
scribes methods for class verification and monitoring of air 
cleanliness. It also addresses certain other factors that affect 
control of airborne contaminants. 

FED-STD-209E does not address the physical, chemical, 
radiological, or viable nature of airborne contaminants. It also 
does not address the occupational health concerns of employ- 
ees working in cleanroom environments. 

A cleanroom class is the statistically allowable number of 
particles, greater than or equal to 0.5 micrometers in size, per 
cubic foot of air. Cleanroom classes are shown in Table 
10.10.1. 

In order to meet the class limits, a high efficiency particu- 
late air (HEPA) or ultra low penetration air (ULPA) filter is 
required. A HEPA filter is a disposable, extended-media, 
dry-type filter in a rigid frame with a minimum particle 
collecting efficiency of 99.97% for 0.3 micrometer, thermally 
generated dioctylphthlate (DOP), or specified alternate, aero- 
sol particles at a maximum clean resistance of 1.0 "wg when 
tested at rated air flow capacity. An ULPA filter is a dispos- 
able, extended-media, dry-type filter in a rigid frame with a 
minimum particle collecting efficiency of 99.999% for par- 
ticulate diameters between 0.1 and 0.2 micrometers in size. 

Military specifications^ 102) and publications^ 103) by the 
Institute of Environmental Sciences (IES) define HEPA and 
ULPA filter construction. Filters having an efficiency even 
higher than an ULPA filter are available from some compa- 
nies specializing in cleanrooms and air filtration. 

TABLE 10.10.1. Class limits in particles per cubic foot of size 
equal to or greater than particle sizes shown.* 



Room 




Measured Particle Size (Micrometers) 




Class 


0.1 


0.2 


0.3 


0.5 


5.0 


1 


35 


7.5 


3 


1 


NA 


10 


350 


75 


30 


10 


NA 


100 


NA** 


750 


300 


100 


NA 


1000 


NA 


NA 


NA 


1000 


7 


10,000 


NA 


NA 


NA 


10,000 


70 


100,000 


NA 


NA 


NA 


100,000 


700 



*The class limit particle concentrations shown are defined for class purposes only 
and do not necessarily represent the size distribution to be found in any particular 
situation. 

**Not applicable. 



The primary design considerations for cleanrooms are the 
supply air flow rate, the air flow patterns within the clean- 
room, the method for recirculating the air from the cleanroom 
and the filter efficiency. 

Air is supplied to the cleanroom by an air handling system 
containing the components needed for heating, cooling, and 
humidity control. Noise is readily transmitted to the clean- 
room so very slow fan speeds, vibration isolation and noise 
control devices are important design considerations. The air 
circulation system will also contain two or three stages of 
prefiltration. This allows the final filters in the cleanroom 
ceiling to remain in place for very long periods of time. A final 
filter life of ten years or more is typical for Class 100 and 
better cleanrooms. 

Air from the supply system enters the cleanroom through 
either a ducted module or a pressurized plenum. VS-10-01 
shows the ducted module arrangement. Ducted modules con- 
taining HEPA or ULPA filters are connected to the main air 
supply duct by flexible branch ducts. The modules usually 
contain an internal baffle for balancing the air exhaust which 
must be at a uniform velocity across the face of the filter. The 
ducted modules are mounted in a T-bar grid and sealed with 
either solid gaskets or a liquid gel sealant. The ducted mod- 
ules, because of long filter life, usually are considered to be 
throwaway items; however, some arrangements do permit the 
replacement of filters from within the cleanroom. A ducted 
module system offers maximum flexibility for cleanroom 
modification. 

VS- 10-02 shows a pressurized plenum arrangement. A 
heavy-duty grid system is suspended from the ceiling with 
suspension rods and the HEPA or ULPA filters sealed in the 
grid with liquid gel or solid gaskets. The entire plenum is 
pressurized by the air supply system to allow a uniform flow 
of air through the filters to the cleanroom below. A pressur- 
ized plenum system will usually cost less than a ducted 
module system for large cleanrooms. 

VS- 10-03 shows raised floor and low sidewall arrange- 
ments. Air is returned through a utility chase to the cleanroom 
supply air system. To provide better particulate control, the 
raised floor arrangement is preferred. The low sidewall return 
should not be used for vertical downflow cleanrooms more 
than 14 feet wide in order not to disrupt laminar flow at the 
work area. 

1ES-RP-CC-006-84< 10I04 > contains testing methods for 
characterizing the performance of cleanrooms. It defines 
terms having special meaning and describes test procedures 
to assure proper cleanroom operation. Uniform air flow is 
defined as unidirectional with all velocity readings within 
20% of the average velocity of the work area. The air velocity 
at the work area is generally about 1 00 fpm; however, design 
conditions may require velocities of 10 fpm or lower. 



Specific Operations 



10-7 



REFERENCES 

10.10.1 U.S. General Services Adminstration: FED-STD- 
209E, Federal Standard, Cleanroom and Work Sta- 
tion Requirements, Controlled Environment. 
Federal Supply Service, General Services Admini- 
stration, Washington, DC (June 15, 1988). 

10.10.2 U.S. Army: MIL-F-51068(EA), Specification Fil- 
ters, Particulate, High-Efficiency, Fire Resistant, 



Biological Use, General Specifications For. Com- 
mander, U.S. Army Armament Research and Devel- 
opment Command, Aberdeen Proving Ground, MD 
(October 4, 1982). 

10.10.3 Institute of Environmental Sciences: 1ES-RP-CC- 
001.3, HEPA and ULPA Filters. IES, Mount Pros- 
pect, IL. 

10.10.4 Institute of Environmental Sciences: IES-RP-CC- 
006, Testing Cleanrooms. IES, Mount Prospect, IL. 



10-8 



Industrial Ventilation 




Air is supplied to the ducted modules from the air distribution duct 
through flexible branch ducts, which are secured at the bottom ends 
by clamps. A damper (not shown) on the inside of the collar allows 
balancing of the air flowing from the module to the cleanroom. 




Ducted modules are mounted in 2' x 4' grids and sealed with gaskets 
or a liquid gel sealant. Tear drop or recessed lighting (shown) 
provides illumination. 



I AMERICAN CONFERENCE 
OF GOVERNMENTAL 
NDUSTRIAL HYGIENISTS 



CLEANROOM 
DUCTED MODULE 



date 12-90 1 FIGURE VS-10-01 



Specific Operations 10-9 



Distribution plate- 



H.E.P.A. Filter 
(shown removed) 



Plenum space 




Light fixture 



Suspended frame/ 
system — 



— Vertical airflow 
through room 



Supply air enters a pressurized plenum and strikes a distribution plate. Each 
2' x 4' opening in the support structure contains a HEPA filter. Filters are 
sealed around the perimeter with gaskets or a fluid ic sealant. The framing 
structure is supported from the plenum ceiling by suspension rods. Tear drop 
lighting is shown. 



AMERICAN CONFER] 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



CLEAN ROOM 
PRESS URIZED PLENUM 



DATI 



12-90 



FIGURE 



VS-10- 02 



10-10 



Industrial Ventilation 



Raised floor 

Utility compartment — 



-T?=& 




1 



"— Air flow 



Raised floor with depressed slab. Air is returned through a utility 
compartment to the air supply system. 



Floor 



Utility compartment 



/ / 








Air flow 



Low sidewall grille return through a utility compartment to the 

air supply system. Room width is limited to 14 feet if laminar 

air flow is to be achieved. The distance from the top of the 
grille to the floor should not exceed 18 inches. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



-yivr 



CLE AN ROOM RETbn 
AIR ARRANGEMENTS 



DATE 12-90 



ncur 



RE VS-10-03 



Specific Operations 



1.0-11 



10.15 FILLING OPERATIONS 

Filling operations have special considerations that should 
be addressed when designing hoods. An enclosed space is not 
empty but rather is filled with air. When material enters the 
space, it forces the air out which in turn can carry some of the 
material with it. Also, additional air can be entrained by the 
material stream entering the enclosed space. This effect is a 
function of the size of the particles and the distance the 
material must fall. These two effects must be considered when 
designing hoods for material handling situations. 

If there are any openings in the walls of the container which 
is being filled, some "splashing" of the material can occur. 
This can lead to loss of material through cracks and openings 
in the receiving vessel. The design of the ventilation system 
should take this effect into account. 

The proper choice of exhaust flow rate is critical. If too little 
air is exhausted, the air displaced by the falling material may 
exceed the exhaust rate and the contaminant may not be 
adequately controlled. If too much air is exhausted, excess 
material could be entrained into the exhaust air stream. As this 
material often is the product, excess product loss could occur. 

VS- 15-01 illustrates four different ways of controlling 
barrel or drum filling operations. VS- 15-02 illustrates bag 
filling and weighing. VS- 15-03 depicts a bag tube packer. 
VS-15-10 and VS-15-1 1 depict a weighing hood where dry 
materials are removed from a bulk pack and weighed into 
smaller bags. 00 15 3) These smaller bags are then packed into a 



container. Bags containing toxic materials can be opened 
within an enclosing hood such as shown on VS-15-20. 00 154) 

VS-15-30 shows how to extract a toxic liquid from a 
process line or vessel for analysis and VS-15-21 shows a 
possibility of controlling leaks around rotating shafts that 
enter containers. 00 155) 

REFERENCES 

10.15.1 Hama, G.M.: Ventilation Control of Dust from 
Bagging Operations. Heating and Ventilating, p. 91 
(April, 1948). 

10.15.2 Cooper, T.C.: Control Technology for a Dry Chemi- 
cal Bagging and Filling Operations. Monsanto Ag- 
ricultural Products Co., Cincinnati, OH (1983). 

10.15.3 Gressel, M.G.; Fischback, T.J.: Workstation Design 
Improvements for the Reduction of Dust Exposures 
During Weighing of Chemical Powders. Applied 
Industrial Hygiene 4:227-233 (1989). 

10.15.4 Goldfield, J.; Brandt, F.E.: Dust Control Techniques 
in the Asbestos Industry. A paper presented at the 
American Industrial Hygiene Conference, Miami 
Beach, FL (May 12-17, 1974). 

10.15.5 Langner, R.R.: How to Control Carcinogens in 
Chemical Production. Occupational Health and 

Safety (March-April, 1977). 



10-12 Industrial Ventilation 



Close clearance 




1 slot 



W L - 



4 min. 



Q = 1 00 cfm/ft barrel top (minimum) 
Minimum (duct velocity = 3500 fpm 
h e = 1.78 VP S + 0.25 VP d 



Feed spout 
4" min. dia 



Flex duct 




Exhaust duct 

- 45° 



Q = 50 cfm x drum diam. (ft) 
Minimum duct velocity = 3500 fpm 



h £ 



0.25 VP H 




Q = 1 50 cfm/ft of open face area 
Minimum duct velocity = 3500 fpm 
h e = 0.25 VP d (45 u taper) 




Q = 300-400 cfm 

Minimum duct velocity = 3500 fpm 



h £ 



0.25 VP H 



Note 1: Air displaced by material feed rate may require higher exhaust flow rates. 
Note 2: Excessive air flow can cause loss of product. 
Reference: 10.15.1 



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BARREL FILLING 



DATE 1 ^Q 1 [FIGURE YS-15~01 



Specific Operations 10-13 



500 fpm maximum 




— Hood attached to bin 



Principal dust source 



Scale support 



Q = 400-500 cfm - non-toxic dust 
1000-1500 cfm - toxic dust 

Minimum duct velocity = 3500 fpm 
h e = 0.25 VP d 



Note: Care must be taken such that too much air 
is not used, as valuable product will be 
pulled into the exhaust system. 



Reference: 10.15.2 



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BAG FILLING 



DATE 



l-gi [figure; YS-15-Q2 



10-14 



Industrial Ventilation 




— C 



Q ~ 500 cfm per filling tube 
= 500 cfm at Feed hopper 
~ 950 cfm at Spill hopper 

Minimum duct velocity = 3500 fpm 

h e = 0.25 VP d for take-off at A and C 
1.0 VP d for take-off at B 



Reference: 10.15.2 



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BAG TUBE PACKER 



DATE 



1-91 



| FIGURE VS-15-03 



Specific Operations 10-15 



Booth 




Scale 



Bag Container 



Air Shower 



Dry Material 
Container Hood 



Dry Material Container 



NOTE: See VS-15-11 for design details 



Reference 1 0. 1 5.3 



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WEIGH HOOD ASSEMBLY 
DRY MATERIAL 



DATE 



1-qI [figure ys-15^10 



10-16 Industrial Ventilation 



BOOTH 




Q c 



50 cfm/ft of face open area. 



L and W to fit operation 

Minimum duct velocity = 3500 fpm 

h e = 1.78 VP S + 0.25 VP d 

Dry material container hood is extension of 
booth slot 



Configure to fit equipment. 



AIR SHOWER 




Qc 



I 00 Ucfm 



L s = 3 feet. (Can be longer if required to fit 
workstation but do not exceed 1/2 booth 
length; L) 



0.25" pegboard or equivalent, 20 percent 
maximum open area. 



DRY MATERIAL CONTAINER HOOD 

Hood is extension of booth slot. An additional 
takeoff(s) may be used if required for hood air flow 
distribution. 






Airflow and hood slot design per VS-15-01 

12" to 24" Diameter 
24" Maximum 



Reference 1 0. 1 5.3 



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WEIGH HOOD DETAILS 
DRY MATERIAL 



DATE 



2-91 1 



FIGURE 



VS-15-lt 



Specific Operations 10-17 



To exhaust 



To exhaust 

Compactor io take 
fiber bags — — —^ 



Open grille work sheif 
(under' hood) 




Light inside hood 



6" Rubber Curtain 



Hopper connected to 
screw feed, chute, etc. 



Q = minimum 250 cfm/ft 2 of open area 
Minimum duct velocity = 3500 fpm 
h e = 0.25 VP d 



Reference: 1 0. 1 5.' 



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TOXIC MATERIAL 
BAG OPENING 



DATE 



1 - 91 



[FIGURE VS~~15~ 20 



10-18 



Industrial Ventilation 



Air inlet 



Impeller 
shaft 




Optional slinger 

Used to prevent the process fluid 

from creeping along the shaft. 



Q = 500 cfm/ft of open area 
(typically 10-40 cfm) 
Note: Sufficient air must be provided to 

dilute flammable gases and/or vapors 
to below 25% of LEL. See Chapter 2. 
Duct velocity ~ 2000 f pm 



h e = 1.78VP S 



0.25VP, 



Note: Similar hood is appropiate for unions 



Reference: 10.15.5 



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SHAFT SEAL ENCLOSURE 



DATE 



1-91 I FIGURE VS- 15^21 



Specific Operations 10-19 



Ram type 
sampling valve 




Process line or vessel 



Slots or perforated plate. 





Door 

Swing out or vertical sliding 
Interlock desirable to prevent 
sample extraction unless door 
is closed. 



Q = 125 cfm/ft of open area (door area) minimum 
Duct velocity = 2000 fpm 
h e = 1 .78 VP s + 0.50 VP d 



NOTE: Sufficient air must be provided when door closed 
to dilute flammable gases and or vapors to 25 % 
of LEL. See Chapter 2 



Reference: 1 0. 1 5.5 



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SAMPLING BOX 



DATE 



X- 91 | figure ys- 15^30 



10-20 



Industrial Ventilation 



10.20 FOUNDRY OPERATIONS 

Foundry operations include many operations common to 
other industries. Some of these operations are covered in the 
following subsections of this chapter: 

10.45 Machining 

10.50 Material Transport 

10.55 Metal Melting 

10.60 Mixing 

10.80 Surface Cleaning 

10.90 Welding and Cutting 

This subsection addresses operations that are more unique 
to the foundry industry: casting shakeout and core making. 

10.20.1 Casting Shakeout: Foundry shakeout ventilation 
rates depend on the type of enclosure and the temperature of 
the sand and castings. The enclosing shakeout hood (VS-20- 
01) requires the smallest air flow rate. The side draft shakeout 
hood (VS-20-02) requires additional air flow rates but pro- 
vides improved access for casting and sand delivery and for 
casting removal. The downdraft shakeout (VS-20-03) is the 
least effective in controlling contaminant and requires the 



highest ventilation rates. It is not recommended for hot cast- 
ings. The shakeout hopper below the shakeout table requires 
additional exhaust ventilation equivalent to 10% of the 
shakeout hood exhaust rate. 

Particular attention should be paid to the conveyor remov- 
ing sand from the shakeout. This conveyor requires hoods and 
ventilation as described in Section 10.50. 

Rotary tumble mills used for shakeout should be treated as 
an enclosing hood with a minimum inward velocity of 150 
fpm through any opening. 

10.20.2 Core Making: Core making machines require 
ventilation to control reactive vapors and gases such as amines 
and isocynates that are used in the core making process. A 
minimum capture velocity of 75 fpm is required. However, a 
ventilation rate as high as 250 cfm/ft 2 of opening may be 
necessary for adequate control of contaminant emissions. 
When cores are cured in ovens, adequate ventilation control 
of the oven is required. 

REFERENCES 

10.20.1 American Foundrymen's Society, Inc.: Foundry 
Ventilation Manual. AFS, Des Plaines, IL (1985). 



Specific Operations 10-21 



ENCLOSING HOOD 

Provides best control with least flow rate 
Minimum duct velocity = 4000 fpm 
h e = 0.25 VP d 



Working openings, 
keep as small as 
possible. 



Molds in 
here 



Mold 
conveyor 




Shakeout 




Castings 
out here 



Shakeout exhaust, minimum 


* 


Type of hood 


Hot castings 


Cool castings 


Enclosing ** VS-20-01 


2 
200 cfm/ft opening 

At least 200 cfm/ft 2 
grate area 


200 cfm/ft opening 

At least 150 cfm/ft 2 
grate area 


Two sides and 1/3 top 
area enclosed** VS-20-02 


2 

300 cfm/ft grate area 


2 
275 cfm/ft grate area 


Side hood (as shown or 
equivalent) ** VS-20-02 


400-500 cfm/ft 2 grate 
area 


350-400 cfm/ft grate 
area 


Double side hood ** VS-20-02 


n 

400 cfm/ft grate area 


300 cfm/ft grate area 



*Choose higher values when 

(1) Castings are quite hot 

(2) Sand to metal ratio is 

(3) Cross — crafts are high 

** Shakeout hoppers require an additional 10% exhaust. 



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FOUNDRY SHAKEOUT 
ENCLOSING 



DATE 



10-90 \™»™ VS-20-01 



10-22 Industrial Ventilation 




Moveable panels to secure 
desired distribution. 



H^L 



Channel iron guard 



Optional top 
take-off. 



0.2 W— - 



Baffle to edge 
__y_of grate 




gidly 
aced 



Minimum practical clearance — 
Velocity through openings 2000 fpm 

SIDE-DRAFT HOOD 

Minimum duct velocity = 4000 fpm. 
h p = 1.78 VP q + 0.25 VP H 



Blank wall in this position is 
almost as good as double hood. 




Minimize 
clearance 




Rigidly braced 



DOUBLE SIDE-DRAFT 

Proportions same as single side-draft hood except for overhang. 

Minimum duct velocity = 4000 fpm 
Slots sized for 2000 fpm 
h e = 1.78 VP S + 0.25 VP d 



See VS — 20 — 01 for exhaust rates 



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FOUNDRY SHAKE0UT 
SIDE DRAFT 



DATE 



10-90 FIGURE VS-20-02 



Specific Operations 10-23 



If feeder enclosure is over 10 feet long, provide exhaust 
at hopper. See VS-50-10 and VS-99-01 



Grate 





Minimum area = 
4 x duct area 

END VIEW 



—-Enclose pan feeder or belt completely. 
Exhaust at transfer to elevator. 

SIDE VIEW 



HOPPER EXHAUST DETAIL 



ELEVATOR 
See VS-50-01 



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FOUNDRY SHARED UT 



DATE 



11-90 



figure VS-20-03 



10-24 



Industrial Ventilation 




Canopy hood 



~Mil|ll|ll|l!|!i|ll[ 
^t±J#^4l l+i 4 l-M LL| XX [U j- 

rr|[T|Ti|i!iiniii|ii|ii|ii|iii 
^u-i-UMJ||ii|ii[i[|[i|Si|ii| 



-To suit — — -- 

operation 




Slotted 
side draft 
hood. 




Use side baffle on canopy hood 



Canopy hood: 



Q = 250 cfm/ft canopy - single unit 
2 
150 cfm/ft canopy - double unit 

h e = 0.25 VP d 



Note: Slotted side draft hoods required to remove 
smoke as hot cores emerge from machine. 
Minimum capture velocity = 150 fpm 



Side draft hood: 



Q = 150(10X + A) where A equals hood area 
h e = 1.78 VP S + 0.25 VP d 



Note: Conveyor or cooling area require ventilation for 

large cores. Scrap conveyor or tote boxes may also 
require additional ventilation- 
Minimum duct velocity = 3500 fpm 



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SHELL CORE MAKING 



DATE 



10-90 1™™* VS-20-10 



Specific Operations 10-25 




Top view of take-off 
connection 



Roll -over handle — < 



\ 



Hood. Closed on ends, 
top and sides. 




Seaf around shaft 



SIDE VIEW 



Q = 200 cfm/ft of open face area 
Minimum duct velocity = 3500 fpm 
h e = 0.25 VP d 

Note: Elbow and rotating connection losses 
not included. 



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CORE MAKING 
SMALL ROLL-0 



CH 

t TYPE 



DATE 



10-90 



FIGURE VS-20-11 



10-26 



Industrial Ventilation 



10.25 GAS TREATMENT 

The handling of gas cylinders for industrial operations 
requires special attention. In addition to the potential safety 
problems associated with transportation and use of com- 
pressed gas cylinders, the gas inside the cylinders can escape 
through leaky valves and fittings. During connection and 
disconnection of the gas lines, due to the operating pressures, 
gas can be released. 

This section of VS-prints illustrates uses of toxic gases 



during fumigation (VS-25-01 and -02) and during ethylene 
oxide sterilization (VS-25-10, -1 1 and -12). 

REFERENCES 

10.25.1 Mortimer, V.D.; Kercher, S.L.; O'Brien, D.M.: Ef- 
fective Controls for Ethylene Oxide — A Case 
Study. Applied Industrial Hygiene 1(1): 15-20 
(1986). 

10.25.2 Hama, G.M.: Ventilation for Fumigation Booths. 
Air Engineering (December 1964). 



Specific Operations 10-27 



Note 7 



Air 

inlet door 




- Note 3 



Note 2 



Ventilation Rates 



Allow 60 minute purge time 

Ventilation rate must be 20 air changes per hour or greater. 

Design must provide: 

(a) 500 fpm velocity or greater through air inlet door 
when large access door is closed 

and 

(b) at least 100 fpm through all openings when large 
access door is open. 



Reference: 10.25.2 



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FUMIGATION BOOTH 



DATE 



1-91 | FIGURE VS-25-01 



10-28 



Industrial Ventilation 



NOTES: 

1. Provide an air inlet with automatic damper closure; damper must be interlocked with 
fan circuit to open only when fan is turned on. Size opening for a minimum velocity of 
500 fpm. Air inlet must be located so purge air sweeps entire booth. 

2. Loading door must be opened only when booth has been completely purged. Provide 
gaskets, screw clamps, and brackets for applying uniform pressure for a gas-tight fit. 

3. Provide ventilated cabinet for gas cylinders in use and being stored. Fan must be on 
continuously and exhaust approximately 500 cfm to produce a negative pressure in 
the cabinet when the doors are closed. 

4. Provide nozzle openings for introducing fumigant gas. A circulating cabinet fan should 
also be provided for obtaining good mixture of fumigant gas. 

5. Mechanical fan damper must be provided that closes tightly when fan is shut off 
during fumigation and opens when fan is turned on. Damper controls should be 
interlocked with fan controls. 

6. Fan for ventilating fumigation booth must be sized to dilute air to safe limit in required 
time. Use vertical, outside, discharge stack away from windows, doors, and air 
intakes. 

7. Fumigant gas cylinder cabinet fan must run continuously. 

8. Control switches for fan and lights and an air flow switch-actuated pilot light are 
recommended. 

9. Red warning light to indicate booth is under fumigation as a protection against 
careless entry is recommended. 

10. To facilitate penetration of fumigant gas and subsequent airing out, mattresses should 
be loaded with separators to allow free air space around each mattress. 

11. Fumigants with no odor-warning properties should be used together with an odor- 
indicating chemical. 

12. Where toxic fumigants are used, a leak test should be made on the booth. The booth 
first should be tested by lighting several large smoke candles in it with doors and 
dampers closed. Leaks can be noted by the presence of smoke at the point of escape. 
Where highly diffusible toxic gases are used, an additional test should be made with 
the booth under charge, at doors and dampers, with a sensitive detecting meter or 
sampling device. 



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FUMIGATION BOOTH NOTES 



DATE 



1-91 



figure VS-25-02 



Specific Operations 10-29 



Q c (Note 1) Qo(Note 2) 



Q A (Note 5) 



Q R (Note 6) 



Slot or Con op y Hood 
(see VS-25-11) 




Aerator 



Ethylene oxide 
cylinders 



^ And Siphon 

Air Gap 



v — Slern \zer 

(see VS-25 11) 



S e e not e 6 



TYPE OF EQUIPMENT 


EXHAUST FLOW RATE 
(Q R ) CFM 


Electrically heated 

gas sterilizers (< 10 ft ' 5 ) 


o.io ! 


Steam heated gas 
or steam sterilizers 


0.25 


A e r a I o r a n d i n s t r u m e n i 

washer units 


0.15 



See VS-25 -11 
See VS-25- 12 



lor notes, 
for notes 



Reference: 10.25. 



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ETHYLENE OXIDE STERILIZERS 



DATE 



; - 9i 



™ URE VS 25 10 



10-30 



Industrial Ventilation 



NOTES: 

1. The ethylene oxide (EtO) supply cylinders should be placed in a ventilated cabinet or a 
partially enclosed hood with an exhaust rate, Q c , of at least 100 cfm/ft 2 of open area. 

2. The anti-syphon air gap in the sterilizer evacuation drain line should be enclosed and 
ventilated. The enclosure should have one or two openings to allow air, G D , to enter 
and to prevent liquid, which might back up from the drain, from reaching the sterilizer 
evacuation line. In lieu of a greater value specified by the sterilizer/vacuum pump 
manufacturer, Q D should be approximately 50 cfm and the openings sized to maintain 
approximately a 600 fpm face velocity. 

3. The overpressure relief valve should be vented to carry EtO out of the building if it 
should ever open. With a sealed line connecting the valve with the ventilation duct, 
there will be no ventilation volume, Q v , except when the valve opens. Consult the 
sterilizer manufacturer for the proper size of this line; too much resistance could 
interfere with proper venting of the chamber. 

4. A hood should be placed above the sterilizer door to remove EtO rising from the 
chamber when the sterilizer door is "cracked" open a few inches for approximately 15 
minutes before the sterilized items are removed from the chamber. See VS-25-12 for a 
discussion of the exhaust volume, Q s , requirements. 

5. If an aerator is installed, its door should be hinged, and it should be placed beside the 
sterilizer so that the doors of the gas sterilizer and aerator open away from each other 
to facilitate transferring the sterilized items. Consult the manufacturer for the required 
air flow, Q^. 

6. The room behind the wall enclosing the steritizer(s) and other equipment should be 
exhausted adequately to handle the air driven to the ceiling by the thermal gradients 
caused by the heated equipment. The ideal arrangement would be to have a properly 
sized vent above each piece of heated equipment. The total Or should be the sum of 
the values for each piece of heated equipment (see VS-25-10) plus 100 cfm/ft 2 of open 
area for transfer vents placed in the upper portion of the room. However, federal 
hospital standards specify that, for a recess room containing a gas sterilizer, the 
volume exhausted in one hour should be at least ten times the room volume. Transfer 
vents placed in the lower portion of the room will help the influx of air to supply the 
thermal air currents and would not add to the total exhaust requirement. 

7. All air that could contain EtO should be exhausted through a ventilation system which 
does not have vents in any other rooms. The discharge of the fan on the roof should be 
located so that the exhausted air will not re-enter the building or expose people 
outside the building. This ventilation system should have a flow sensor/alarm to warn if 
it is not functioning properly. If there is the possibility of lint in the exhausted air, use 
a differential pressure sensor or some other type that will not be clogged or stuck open 
by the accumulation of lint. 



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ETHYLENE OXIDE STERILIZER 

NOTES 



DATE 




figure ys-25 



Specific Operations 10-31 



H slot ]_ 




slot 



H slot * 3 " 



L slot - X + °- 66H slot 



D = 



slot face to be at 45"-90° 
angle with plane of enclosure 



Q =75 cfm/ft slot length 

slot 



Y = 2" for airing out sterilizer 
chamber. 



hood I 




Q, = 100 L, ,W 

hood hood 



H, = 12 to 24 

hood 



W = 



nood 



hood 



L, ,= X + 0.66H 
hood 



hood 



Y - 2" for airing out sterilizer 
chamber. 



CANOPY HOOD 



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ETHYLENE OXIDE STERILIZER 
HOOD DETAILS 



DATE 



02-91 | FIGURE VS J j 12 



10-32 



Industrial Ventilation 



10.30 KITCHEN EQUIPMENT 

The purpose of an exhaust system for kitchen equipment is 
to control heat, humidity and grease vapor released into the 
space by cooking or dishwashing equipment. A secondary 
consideration is the control of combustion products associated 
with the heat source which may be vented separately or 
through the hood itself. 

National Fire Protection Association (NFPA) Standard 
9500.30. i) describes grease filter construction as well as hood 
construction necessary to maintain hood integrity in the event 
of a fire. Welded seam construction is preferred and some- 
times required by public health authorities to assure cleanabil- 
ity and ease of maintenance. The National Sanitation 
Foundation Standard No. 4 (i0302) also lists hood construction 
requirements for cleanability and integrity in the cooking and 
food zones within the hood. In all cases, the local health 
authorities having jurisdiction should be consulted for con- 
struction requirements prior to hood fabrication. 

Fire is a primary concern with all cooking equipment. Each 
hood will require some type of fire suppression consistent 



with local fire code requirements. The system selected must 
not compromise sanitation or endanger workers due to loca- 
tion or system activation. Hood or duct penetrations by fire 
suppression piping, etc., must be sealed to prevent short 
circuiting of air or loss of fire arrestance. 

For high temperatures situations such as exposed flames or 
charcoal, the grease filters must be sufficiently removed from 
the heat source to prevent ignition. Fan selection may require 
use of high temperature fan components and consideration of 
the effect of change in air density. 

REFERENCES 

10.30.1 National Fire Protection Association: Standard for 
the Installation of Equipment for the Removal of 
Smoke and Grease-Laden Vapors from Commercial 
Cooking Equipment, (Standard 96). NFPA, Quincy, 
MA (1987). 

10.30.2 National Sanitation Foundation: Commercial Cook- 
ing and Hot Food Storage Equipment, Standard No. 
4. NSF, Ann Arbor, MI (1986). 



Specific Operations 1 0-33 




Pitch duct 
toward hood 



1/2 H 




CANOPY HOODS 



Q = 250 cfm/ft of door area- each end 
Minimum duct velocity = 1000 - 3000 fpm 
h e - 0.25 VP d 



Dishwasher 



6" min 



2" slot 
around hood 

6" 




1/2 H 



SLOT HOODS 



Q = 150 cfm/ft 2 of door area (150WH) 

each end 
Minimum duct velocity = 1000 - 3000 fpm 
h e = 1.00 VP S + 0.25 VP d 



Dishwasher 




EXHAUSTED VESTIBULES 



Q = 150 cfm/ft of entrance and exit area 
Minimum duct velocity = 1000 - 3000 fpm 
h e = 0.50 VP d 



Note: If direct exhaust connections are provided from dishwasher body, cap these 
connections and use external hoods. 



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DISHWASHER VENTILATION 



date W-90 FIGURE VS-30-01 



10-34 



Industrial Ventilation 



Ducts 6' on center (max.) 
for large hoods pH 



Grease filter 






"1" 
- H- 



- H- 



- f j 

_ | 1 

J I 



H = 4' max. ^_. 

„_J 



Cooking equipment 




6' min. 
■ overhang on 
three sides 



Removable-^ 
drip pan 



HOOD AGAINST WALL 
Q = 80 cfm/ft 2 of hood area (80 WL) 

Not less than 50 cfm/ft 2 of face area (50 PH) 
P = perimeter of hood = 2W + L 

Duct velocity = 1000 - 4000 fpm, to suit conditions 
h e = (filter resistance + 0.1") + 0.50 VP d ( straight take off) 
h e = (filter resistance + 0.1") + 0.25 VP d ( tapered take off) 



6 ' maximum 



I 




" r 


I 


1 " 


~ r 


T~ 


~1 " 


~-~r 


.... | 


I — 


- H - - 


--H-- 


_ + _ 


— 1__ 


— i — 


- + - 


1„_ 


--H- 


1 


I 


_ J _ _ 


-_L_ 


_x_ 


__J_- 


_ - L _ 


„X_ 


__J__ 


_._L„ 





X 



H = 4' max. 

_J 



Island cooking area 



Grease filters- 



45°- 60° 



6 min. 
- overhang on 

all sides 




Drip pan 



ISLAND TYPE HOOD 



P — perimeter of hood 
= 2W + 2L 
Q = 125 cfm/ft' of hood area (125 WL) 

Not less than 50 cfm/ft 2 of face area (50 PH) 

Minimum duct velocity - 1000 — 4000 fpm, to suit conditions 

h e - (filter resistance + 0.1") + 0.50 VP d (straight take off) 

h e = (filter resistance + 0.1") + 0.25 VP d (tapered take off) 

Note: See VS — 30— 11 for information about filters and fans for range hoods. 



AMERIC AN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



KITCHEN RANGE HOODS 



DA' 



r E 1Q-90 



figure 



E VS-30-10 



Specific Operations 10-35 




Ducts 6' on center maximum Plenum 



. 



1/^Z^mzL 



Filters 



20" minimum 



*:< 



Cooking equipment 



Face or ends can be 
opened for filter 
removal — — — — — 



1 ' max 
set— back 



Filter mounting height 
See note 4 below 



Closed ends 
desirable 




3' max. 



LOW SIDE WALL HOOD 

= 200 cfm/lineal ft of cooking surface (200L) 
Minimum duct velocity - 1000 - 4000 fpm, to suit conditions 
h e = (filter resistance + 0.1") + 0.50 VP d ( straight fake off) 
h e - (filter resistance + 0.1") + 0.25 VP d (tapered take off) 



NOTES FOR KITCHEN HOODS 

Filters: 1. Select practical filter size. 

2. Determine number of filters required from manufacturer's data. 
(Usually: 2 cfm maximum exhaust for each square inch of filter area.) 

3. install at 45° -60° to horizontal. Never horizontal. 

4. Filter mounting height (Reference 10.30.1) 

a. No exposed cooking flame 18'' minimum to lowest edge of filter. 

b. Charcoal and similar fires 4' minimum to lowest edge of filter. 

5. Shield filters from direct radiant heat. 

6. Provide removable grease drip pan. 

7. Clean pan and filters regularly. 

Fan: 1. Use upblast discharge fan. Downblast is not recommended. 

2. Select fan for design and SP resistance of filters and duct. 

3. Adjust fan specification for expected exhaust air temperature. 



AMERICAN CONFERENCE 
OF GOVERNMENTAL 



'CHEN RANGE HOOD 



INDUSTRIAL HYGIENISTS 




™urf: VS 30 11 



^■v-r^z^^o^ 



10-36 



Industrial Ventilation 




— Grease filters 
Metal sides 



Lower edge of filters 

at least 4'-6" above fire 



CHARCOAL BROILER 

Q = 100 LH 

Minimum duct velocity = 1000 — 3000 fpm 



K 



= (filter resistance + 0.1") + 0.50 VP (straight take off) 
= (filter resistance + 0.1") + 0.25 VP (tapered take off) 



Glass to be 
pyrex or high 
temperature type - 




— Grease filters 
Metal sides 



Lower edge of filters at 
east 3' — 6" above fire 



BARBEQUE PITS Notes: 1. If hood is more than 12 feet 

long use multiple takeoffs 6 
feet on center. 

2. See VS-30-11 for information 

Q - 100 WH (maximum open door area, ft 2 ) ° bou * fI ^rs and fans for 

Minimum duct velocity - 1000 - 3000 fpm 

h e - (filter resistance + 0.1") + 0.50 VP (straight take off) 

h e = (filter resistance + 0.1") + 0.25 VP (tapered take off) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



CHARCOAL BROILER AND 
BARBEQUE PIT VENTILATION 



DATE 



10-90 | F1GURE VS-30-12 



Specific Operations 10-37 



10.35 LABORATORY VENTILATION 

The primary method of contaminant control within the 
laboratory is exhaust ventilation and, in particular, laboratory 
hoods. This section presents information on laboratory hoods 
but expands to other types of ventilation control such as 
biological safety cabinets, clean benches, and other local 
exhaust systems found in the laboratory. 

10.35.1 Laboratory Hoods: In most cases, laboratory 
hoods will be purchased from manufacturers specializing in 
the design and construction of laboratory hoods. VS-35-01 
shows atypical laboratory hood design. VS-35-02 describes 
general-use laboratory hoods and VS-35-03 describes per- 
chloric acid hoods. VS-35-04 describes work practices for 
laboratory hoods. 

Several features are essential to the proper performance of 
the hood. The most important aspect of the hood is the 
aerodynamic entry characteristics. For the hood to adequately 
control contaminants, the entry must be smooth. This usually 
is achieved with an airfoil sill at the leading edge of the 
workbench. Often, beveled jambs at the side wall entry will 
improve the air flow. 

In many cases, good performance correlates with uniform 
face velocity. To achieve a uniform face velocity, many hood 
manufacturers provide adjustable slots in the plenum at the 
back of the hood. Although the adjustment will allow for 
unusual conditions such as large hot plates for sample diges- 
tions, inappropriate adjustment of the slots can have a detri- 
mental effect on hood performance/ 1035 -° 

Supply Air Distribution: For typical operation of a labora- 
tory hood, the worker stands at the face of the hood and 
manipulates the apparatus in the hood. The indraft at the hood 
face creates eddy currents around the worker's body which 
can drag contaminants in the hood along the worker's body 
and up to the breathing zone. The higher the face velocity, the 
greater the eddy currents. For this reason, higher face veloci- 
ties do not result in greater protection as might be supposed. 

Room air currents have a large effect on the performance 
of the hood. Thus, the design of the room air supply distribu- 
tion system is as important in securing good hood perform- 
ance as the face velocity of the hood. American Society of 
Heating, Refrigeration and Air Conditioning (ASHRAE) re- 
search project RP-70 results, reported by Caplan and Knut- 
son, (1035 2) conclude in part: 

1. Lower breathing zone concentrations can be attained 
at 50 cfm/ft 2 face velocities with good air supply 
distribution than at 150 cfrn/ft 2 with poor air distribu- 
tion. With a good air supply system and tracer gas 
released at 8 liters per minute inside the hood, breath- 
ing zone concentrations can be kept below 0. 1 ppm and 
usually below 0.05 ppm. 

2. The terminal throw velocity of supply air jets should 



be no more than one-half the hood face velocity; such 
terminal throw velocities are far less than conventional 
practice. 

3. Perforated ceiling panels provide a better supply sys- 
tem than grilles or ceiling diffusers in that the system 
design criteria are simpler and easier to apply, and 
precise adjustment of the fixtures is not required. 

For the reasons described, an increased hood face velocity 
may be self-defeating because the increased air volume han- 
dled through the room makes the low-velocity distribution of 
supply air more difficult. 

Selection of Hood Face Velocity: The interaction of supply 
air distribution and hood face velocity makes any blanket 
specification of hood face velocity inappropriate. Higher 
hood face velocities will be wasteful of energy and may 
provide no better or even poorer worker protection. The 
ANSI/ASHRAE Hood Performance Test< l0353 > may be used 
as a specification. The specified performance should be re- 
quired of both the hood manufacturer and the designer of the 
room air supply system. 

The specification takes the form: 

AUyyy, Alyyy, or AMyyy 

where: 

AU identifies an "as used" test 

Al identifies an "as installed" test 

AM identifies an "as manufactured" test 

yyy = control level, ppm, at the breathing zone of 
the worker. 

Any well-designed airfoil hood, properly balanced, can 
achieve < 0.10 ppm control level when the supply air distri- 
bution is good. Therefore, it would seem appropriate that the 
"AM" requirements would be < 0. 10 ppm. The "AU" require- 
ment involves the design of the room supply system and the 
toxicity of the materials handled in the hood. The "AU" 
specification would be tailored to suit the needs of the labo- 
ratory room location. 

For projected new buildings, it is frequently necessary to 
estimate the cost of air conditioning early — before the 
detailed design and equipment specifications are available. 
Forthat early estimating, the guidelines listed in Table 10.35.1 
can be used. 

10.35.2 Biological Safety Cabinets: Biological safety 
cabinets (BSCs) are classified as Class 1; Class IT, Types A, 
B1,B2 and B3; and Class III. 

Class I BSC provides personnel and environmental protec- 
tion but does not protect the product. The front panel can be 
open, allowing room air to enter the cabinet, sweep the inner 
surfaces and exhaust out the duct. A front closure panel with 
glove ports may be installed. If gloves are installed, air is 



10-38 



Industrial Ventilation 



TABLE 10.35.1. Laboratory Hood Ventilation Rates 



Condition 



cfm/ft 2 
Open Hood Face 



1 . Ceiling panels properly located with average panel face velocity < 40 fpm. (10 ' 35 ^ Horizontal sliding sash hoods. No 60 
equipment in hood closer than 12 inches to face of hood. Hoods located away from doors and traffic ways.* 

2. Same as 1 above; some traffic past hoods. No equipment in hoods closer than 6 inches to face of hood. Hoods located 80 
away from doors and traffic ways.* 

3. Ceiling panels properly located with average pane! face velocity < 60 fpm < 10 - 35 - 2 ) or ceiling diffusers properly located; 80 
no diffuser immediately in front of hoods; quadrant facing hood blocked; terminal throw velocity < 60 fpm. No 

equipment inhood closer than 6 inches to face of hood. Hoods located away from doors or traffic ways.* 

4. Same as 3 above; some traffic past hood. No equipment in hood closer than 6 inches to face of hood. 100 

5. Wall grilles are possible but not recommended for advance planning of new facilities 

*Hoods near doors are acceptable if 1) there is a second safe egress from the room; 2) traffic past hood is low; and 3) door is normally closed. 



drawn through a secondary opening equipped with a roughing 
filter. A laboratory hood, as shown in VS-35-20, could be 
considered a Class I BSC if the exhausted air is passed through 
HEP A filters prior to release to the atmosphere. 

Class II BSCs provide personnel, product, and environ- 
mental protection. Class II cabinets differ in the proportion of 
air recirculated within the cabinet; velocity of air flow to the 
work surface; where the exhausted air is discharged; and 
whether the contaminated air plenum is under positive pres- 
sure. A Type A cabinet (VS-35-10) may discharge the ex- 
hausted air, after HE PA filtration, directly into the room. Type 
A cabinets which discharge into the work area are not recom- 
mended for use with gases or vapors. A primary application 
is for sterile packaging. Care is required while decontaminat- 
ing the cabinet. 

Type B hoods (VS-35-1 1) discharge the exhaust but may 
recirculate within the cabinet. Type Bl cabinets recirculate 
about 30% of the air within the BSC and typically exhaust the 
remainder outside the laboratory (i.e., exhaust air is not dis- 
charged back into the room). The contaminated plenum is 
under negative pressure. Type B2 cabinets are referred to as 
"total exhaust" cabinets as the contaminated air is exhausted 
to the atmosphere after HEPA filtration without recirculation 
in the cabinet or return to the laboratory room air. Type B3 
BSCs have HEPA filtered downflow air that is a portion of 
the mixed downflow and inflow air from a common exhaust 
plenum. 

Class I'll BSCs (VS-35-20) provide the highest level of pro- 
tection to personnel and the environment. The cabinet is totally 
enclosed with operations conducted through attached gloves. See 
"National Sanitation Foundation Standard No. 49" (I0354) for 
descriptions and requirements of the various classes of BSCs. 

10.35.3 Clean Benches: Clean benches can be divided 
into laminar flow and exhausted clean benches. 

Laminar flow clean benches provide product protection 
only. In a laminar flow clean bench, room air is HEPA-fil- 
tered, directed across the work area and discharged back to 



the room. Air may be directed horizontally as depicted in 
VS-35-30 or vertically as in VS-35-3 1 . Neither of these hoods 
provide worker protection. Workers using the Horizontal 
Laminar Flow Clean Bench are exposed to the product as the 
air sweeps across the product into the worker's face. Workers' s 
arms or other objects protruding into the Vertical Laminar 
Flow Clean Bench opening may cause contaminated air to 
spill into the room. Personal protective equipment or general 
ventilation should be provided as needed. 

Other types of clean benches incorporate the same general 
principles of biological safety cabinets and utilize HEPA 
filtered laminar flow within the hood to provide product 
protection and exhaust sufficient air to ensure flow into the 
hood at the face to provide operator protection. 

10.35.4 Laboratory Equipment: Some laboratory equip- 
ment such as evaporation hoods (VS-35-40), discharge from 
instruments such as 1CP or A A and some ovens (VS-35-41) 
require local exhaust ventilation to adequately control con- 
taminant releases. Often, specially designed ventilation spe- 
cific to the operation provides better control than using a 
laboratory hood to control these releases. 

REFERENCES 

10.35. 1 Knutson, G.W.: Effect of Slot Position on Labora- 
tory Fume Hood Performance. Heating, Piping and 
Air Conditioning (February, 1984). 

1 0.35.2 Caplan, K.J.; Knutson, G. W.: Influence of Room Air 
Supply on Laboratory Hoods. American Industrial 
Hygiene Association Journal 43(I0):738— 746 
(1982). 

10.35.3 American Society of Heating, Refrigerating and Air 
Conditioning Engineers: ANSI/ASHRAE Standard 
110-1995, Method of Testing the Performance of 
Laboratory Fume Hoods, ASHRAE, Atlanta, GA 
(1995). 

10.35.4 National Sanitation Foundation: Standard 49, Class 
II (Laminar Flow) Biohazard Cabinetry. NSF, Ann 



Specific Operations 10-39 



Arbor, MI (1987). 

10.35.5 U. S. Air Force: Technical Order 00-25-203: Stand- 
ards and Guidelines for Design and Operation of 
Clean Rooms and Clean Work Stations. Office of 
Technical Services, Department of Commerce, 



Washington, DC (July 1963). 

10.35.6 Harris, W.P.; Christofano, E.E.; Lippman, M: Com- 
bination Hot Plate and Hood for Multiple Beaker 
Evaporation. American Industrial Hygiene Associa- 
tion Journal 22(4) (August 1961). 



10-40 



Industrial Ventilation 



Optional room air 
by-pass does not 
open until sash is 
closed 25-30 % 

Air toil jarnb 

Moveable sash 

can have horizontal 

sliding panels 

Recessed bottom 
Airtoil sill 




Exhaust duct 

Adjustable top slot 



Sash closes by -pass 
when raised 



Fixed center slot 
Rear baffle 

Adjustable bottom slot 



VERTICAL SASH AIRFOIL HOOD 



For safety shield, at least 
one sosh 16" max. width 




Q = 80-100 cfm/ft full open area 
depending on quality of supply 
air distribution and uniformity 
of face velocity 

h e = 0.5 VP d 

Duct velocity = 1000-2000 fpm to 
suit conditions 



Airfoil 
sill 



Design specifications: 

General use laboratory hoods-See VS-35-02 
Perchloric acid -See VS-35 — 03 

"Auxiliary Air" or "Compensating" hoods 
furnish some replacement air at hood face, 
design varies with vendor. 

Work practices - See VS-35-04 



HORIZONTAL SASH 
AIRFOIL HOOD 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



TYPICAL LABORATORY HOOD 



DATE 



02-91 1 F1GURE VS-35- 01 



Specific Operations 10-41 



GENERAL USE LABORATORY HOODS 

A. Provide uniform exhaust air distribution in hood. Adjust baffles and air flow for < 10 percent 
variation in point-to-point face velocity with sash in maximum open position. 

B. Locate hood away from heavy traffic aisles and doorways. Hoods near doors are acceptable 
if: 1 ) there is a second safe means of egress from room, 2) traffic past hood is low, and 3) door 
is normally closed. 

C. Use corrosion-resistant materials suitable for expected use. 

D. Provide air cleaning on exhaust air if necessary and adequate stack height to minimize re-entry 
of contaminants or to comply with air pollution regulations. 

E. Avoid sharp corners at jambs and sill. Tapered or round hood inlets are desirable. An airfoil 
shroud at sill is important. 

F. Provide filters for radioactive materials in greater than "exempt" quantities. 

G. By-pass opening in hood is desirable to avoid excessive indraft under partially closed sash 
condition. Opening to be baffled to prevent splash from eruption in hood as shown in VS-35-01 . 

H. Provide tempered or conditioned replacement air to laboratory. Replacement air volume to be 
selected for desired air balance with adjoining spaces. 

I. In order to reduce air flow volumes, local exhaust hood should be considered instead of 
laboratory bench hoods for fixed setups. 

J. For air conservation, use horizontal sliding sash with airfoil sill. 

K. All bench hoods should have a recessed work surface and airfoil sill. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



GENERAL USE LABORATORY 
HOODS 



DATE 



02- 91 



I 



FIGURE 



VS-35-02 



10-42 



Industrial Ventilation 



PERCHLORIC ACID HOODS 

Perchloric acid is extremely dangerous because it is a very strong oxidizer. When the acid reacts 
with organic material, an explosive reaction product may be formed. 

1 . Do not use perchloric acid in a hood designed for other purposes. Identify perchloric acid hoods 
with large warning signs. 

2. Provide exhaust ventilation and room supply air with minimal challenge to the hood. 

3. Utilize local exhaust ventilation within the hood to minimize condensation of vapors inside the 
hood. 

4. Locate all utility controls outside the hood. 

5. Materials of construction for this type of hood and duct must be nonreactive, acid resistant, and 
relatively impervious. AVOID ORGANIC MATERIALS unless known to be safe. Stainless steel 
type 31 6 with welded joints is preferred. Unplasticized polyvinyl chloride or an inorganic ceramic 
coating, such as porcelain, is acceptable. 

6. Ease of cleanliness is paramount. Use stainless steel with accessible rounded corners and 
all-welded construction. 

7. The work surface should be water tight with a minimum of 0.5-inch dished front and sides and 
an integral trough at the rear to collect the washdown water. 

8. Design washdown facilities into the hood and duct. Use daily or more often to thoroughly clean 
perchloric acid from the exhaust system surfaces. 

9. Each perchloric acid hood should have an individual exhaust system. Slope horizontal runs to 
drain. Avoid sharp turns. 

10. Construct the hood and duct to allow easy visual inspection. 

11. Where required, use a high-efficiency (greater than 80%) wet collector constructed for per- 
chloric acid service. Locate as close to the hood as possible to minimize the accumulation of 
perchloric acid in the exhaust duct. 

12. Use only an acid-resistant metallic fan protected by an inorganic coating or an air injector. 

13. Lubricate the fan with a fluorocarbon-type grease. 

14. Locate the fan outside the building. 

15. The exhaust discharge must terminate out-of-doors, preferably using a vertical discharge cap 
that extends well above the roof eddy zone. See Figure 5.30. 



AMERIC AN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



PERCHLORIC ACID 
HOOD NOTES 



DATE Q2-91 



FIGURE 



VS- 



35- 



03 



Specific Operations 10-43 



WORK PRACTICES FOR LABORATORY HOODS 

No large, open-face hood with a low face velocity can provide complete safety for a worker standing 
at the face against all events that may occur in the hood. The hood may not adequately protect the 
worker from volatile or otherwise airborne contaminants with a TLV in the low part-per-billion range. 
For more ordinary exposures, a properly designed hood in a properly ventilated room can provide 
adequate protection. However, certain work practices are necessary for the hood to perform 
capably. The following work practices are generally required; more stringent practices may be 
necessary in some circumstances. 

1. Conduct all operations that may generate air contaminants at or above the appropriate TLV 
inside a hood. 

2. Keep all apparatus at least 6 inches back from the face of the hood. A stripe on the bench 
surface is a good reminder. 

3. Do not put your head in the hood when contaminants are being generated. 

4. Do not use the hood as a waste disposal mechanism except for very small quantities of volatile 
materials. 

5. Do not store chemicals or apparatus in the hood. Store hazardous chemicals in an approved 
safety cabinet. 

6. Keep the hood sash closed as much as possible. 

7. Keep the slots in the hood baffle free of obstruction by apparatus or containers. 

8. Minimize foot traffic past the face of the hood. 

9. Keep laboratory doors closed (exception: some laboratory designs require lab doors to be 
open). 

10. Do not remove hood sash or panels except when necessary for apparatus set-up; replace sash 
or panels before operating. 

11. Do not place electrical receptacles or other spark sources inside the hood when flammable 
liquids or gases are present. No permanent electrical receptacles are permitted in the hood. 

1 2. Use an appropriate barricade if there is a chance of explosion or eruption. 

13. Provide adequate maintenance for the hood exhaust system and the building supply system. 
Use static pressure gauges on the hood throat, across any filters in the exhaust system, or 
other appropriate indicators to ensure that exhaust flow is appropriate. 

14. If hood sash is supposed to be partially closed for the operation, the hood should be so labeled 
and the appropriate closure point clearly indicated. 



AMERICAN CONFERENCE WORK PRACTICES FOR 

OF GOVERNMENTAL I LABORATORY HOODS 

INDUSTRIAL HYGIENISTS 



DATE 



02-91 r IGURE VS-35-04 



10-44 



Industrial Ventilation 



HEPA exhaust 
filter 



Balancing 
damper 

HEPA supply 
filter 



Vertical laminar 

Air flow 70-100 fpm 



Positive 

pressure 

plenum 




Blower 



tr 



mi 





Expanded metal 
exhaust cover 



Exhaust filter 
access 



Supply filter 
access 



View screen 



Work zone 
with solid 
work surface 



Work opening 

Air flow 80-100 fpm 



Negative pressure plenum 



CLASS II TYPE A 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



BIOLOGICAL SAFETY CABINET 
CLASS II TYPE A 



DATE 



02-91 1 FIGURE VS-~35~T0 



Specific Operations 



10-45 



Vertical 
I a m i n a r 
air fiow 
70-100 
f pm 




Exhaust 
filter 

HEPA 
supply filter 



Diffuser 

Air flow 
80-100 
f pm 



Vertical 

laminar 
air flow 
70-100 
f pm 




/\ 



~yt 



" t t t t T T 

iRarMcle free 
/workV area 




HEPA 

supply 

filter 



Diffuser 

Air flow 
80-100 
f pm 



RECIRCULATING AIR HOOD * 



100% EXHAUST AIR HOOD 



Recirculating Air Hoods are not recommended for use with gases or vapors. 

Note: See "National Sanitation Foundation Standard 49" (10.35.4) for 
requirements and definitions of classes. 

For product protection only, see VS-35-30 and VS-35-31. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



BIOLOGICAL SAFETY CABINET 
CLASS II, TYPE B 



DATE 



02-91 



FIGURE 



VS-35-11 



10-46 



Industrial Ventilation 



To HEPA filter 



I Air // 
I : ock// 




^vr 



door area and 0.25" 



SP 



Q - 50 cfm/ft 2 of open 

on a closed system. 

h e = 0.50 VP d 

Duct velocity = 2000-4000 fpm 

Pilfers : 1 . inlet air filters in doors. 

2. Roughing filter at exhaust connection to hood. 

3. HEPA filter 
AN facilities totally enclosed in hood. Exterior controls may 
Arm length rubber gloves are sealed to glove port rings. 

Strippable plastic on interior and air cieaner on exhaust outlet may be 
used to facilitate decontamination of the system. 

Pilfer units may be installed in the doors to allow the air flow necessary 
for burners etc. 

For filters, see Chapter 4. 



be advisable. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIEN1STS 



DRY BOX OR GLOVE HOOD 
FOR HIGH TOXICITY & RADIOACTIVE 
A TERIALS 



DATE 02-91 



FIGURE 



■35-20 



Specific Operations 10-47 



Blower 



HEPA fillers 



Lights 




Sides, top and rear 
flush with filter outer 
edges 



Air 
intake 



Prefilter 




Face velocity ~ 90 fpm + 20 fpm 



Note : Total power input must be considered 
as part of air conditioning load. 

This hood does not provide protection 
for the operator. 



Reference 10.35.5 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HORIZONTAL LAMINAR FLOW 

CLEAN BENCH 
(PRODUCT PROTECTION ONLY) 



DATE 



UU Vi |ra"RE VS-35-30 



10-48 Industrial Ventilation 



Blower 



Roughing filler 
if needed 



Sliding glass door 




HEPA filter 



Grating work area 



Blower located outside 
cabinet 





— Damper 



Exhaust 

"Treat' 1 to suit contaminant 



Vertical velocity - 90 fpm with average minimum 

uniformity ± 20 fpm 

Duct velocity = 2000 - 4000 fpm to suit conditions 

Clean station for control of air particles 



Notes: Supply and exhaust should be maintained equal by 
flow meter control techniques. 

This hood does not provide protection for the operator. 

Do not use with toxic material. 



AMERICAN CONFERENCE 

OF' GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



VERTICAL LAMINAR FLOW 
CLEAN BENCH 
(PRODUCT PROTECTION ONLY) 



DATE 92-91 



figure VS-35-31 



Specific Operations 10-49 



Shelf 



Strip heaters built -i 

into bench \ ] '■, 



Size plenum for 
1000 fpm down flow 







Size holes or slots 
for 2000 fpm 



©©o©eo©©o©©0©©© 

^_ Height to suit glassware 




EVAPORATION BENCH 

Q= 20 cfm/foot of hood or 50 HL 
Minimum duct velocity ~ 2000 fpm 
h e = 1.78VP S + 0.25VP d 



Strip heaters built — . 

into shelves \ 



^ 



JJ 





ggSSSg^TgZ 



7 r7r77T77 yTr 



rr ^ T . rrr . rT ^^ r T7 




Slots: size for 2000 fpm 



Height to suit glassware 



~T 



o 





o 


o 




o 


o 





o 









EVAPORATION HOOD 

Q = 20 cfm/foot of shelf or 50 HL for each shelf 
Minimum duct velocity = 2000 fpm 
h e = 1.78VP S + 0.25VP d 



Reference 10.35.6 



AMERIC AN CO NFERENC E 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SPECIALIZED LABOR A TOR Y 
HOOD DESIGNS 



DATE 



02-91 



FIGURE 



VS-35-40 



10-50 Industrial Ventilation 



W 




Oven (dashed line) 



To p View 



T 

12" 

i. 
r 




Maximum 2" gap — 



Skirts on sides 
and back desirable 



Oven Door 



Oven 



Q = 200 - 400 cfm 

Minimum duct velocity 1000 - 3000 fpm 

h = 0.25 VP, 
e a 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



O VEN EXHA LIST 



DAT 



E 02-91 1 FIGURE VS-35-41 



Specific Operations 10-51 



10.40 LOW VOLUME-HIGH VELOCITY EXHAUST 
SYSTEMS 

The low volume-high velocity (LVH.V) exhaust system is 
a unique application of exhaust which uses small volumes of 
air at high velocities to control dust from portable hand tools 
and machining operations. Control is achieved by exhausting 
the air directly at the point of dust generation using close-fit- 
ting, custom-made hoods. Capture velocities are relatively 
high but the exhaust volume is low due to the small distance 
required. For flexibility, small diameter, light-weight plastic 
hoses are used with portable tools. This results in very high 
duct velocities but allows the application of local exhaust 
ventilation to portable tools which otherwise would require 
larger flow rates and large duct sizes when controlled by 
conventional exhaust methods. The resulting additional bene- 
fit is the reduction of replacement air requirements. 

This technique has found a variety of applications although 
its use is not common. Rock drilling dust has been controlled 
by using hollow core drill steel with suitable exhaust holes in 
the drill bits. Air is exhausted either by a multi-stage turbine 
of the size generally used in industrial vacuum cleaners or, in 
the case of one manufacturer/ 1040 ■■> by the exhaust air from 
the pneumatic tool which operates a Venturi to withdraw air 
from the drill. Some applications use flexible connections to 
a central vacuum system to aid in the control of graphite dust 
at conventional machining operations. One- to two-inch di- 
ameter flexible hose was used with simple exhaust hoods 
mounted directly at the cutting tool. In a similar application 
for the machining of beryllium/ 10 402) a central vacuum system 
utilizing 1.5-inch I.D. flexible hoses was employed. The 
exhaust hoods were made of lucite or transparent material and 
were tailor-made to surround the cutting tools and much of 
the work. Exhaust flow rates vary from 120-150 cfm with 
inlet velocities of 1 1,000-14,000 fpm. In another applica- 
tion/ 10403 ) a portable orbital sanding machine has been fitted 
with a small exhaust duct surrounding the edge of the plate. 
A fitting has been provided to connect this to the flexible hose 
of a standard domestic vacuum cleaner. 

VS-40-01 to VS-40-06 illustrate a custom-made line of 
exhaust hoods available. 00 404) The required air flow rates 
range from 60 cfm for pneumatic chisels to 380 cfm for swing 
grinders. Due to the high entering velocities involved, static 
pressures are in the range of 7" to 14" of mercury (95 to 290 
M wg). This high pressure is necessary to create the high 
capture velocities at the source to control the dust. However, 
there are disadvantages associated with high velocities: 1) 
small metal parts can be sucked into the hood; 2) coolants may 



be disturbed; and 3) very high noise levels may be produced. 

10.40.1 Design—Calculations: With the exception of the 
proprietary system mentioned which can be purchased as a 
"package," the design calculations for these systems are 
largely empirical and little performance data are available for 
the user. In normal ventilation practice, air is considered to be 
incompressible since static pressures vary only slightly from 
atmospheric pressure. However, in LVHV systems the ex- 
treme pressures required introduce problems of air density, 
compressibility, and viscosity which are not easily solved. 
Also, pressure drop data for small diameter pipe, especially 
flexible tubing, is not commonly available. For practical 
purposes, the turbine exhauster should be selected for the 
maximum simultaneous exhaust flow rate required. Resis- 
tance in the pipe should be kept as low as possible; flexible 
tubing of less than 1- to 1.5-inch diameter should be limited 
to 10 feet or less. In most applications, this is not a severe 
problem. 

The main consideration in piping for such systems is to 
provide smooth internal configuration so as to reduce pressure 
loss at the high velocities involved and to minimize abrasion. 
Ordinary pipe with threaded fittings is to be avoided because 
the lip of the pipe or male fitting, being of smaller diameter 
than the female thread, presents a discontinuity which in- 
creases pressure loss and may be a point of rapid abrasion. 

For dust exhaust systems, a good dust col lector and pri mary 
separator should be mounted ahead of the exhauster to mini- 
mize erosion of the precision blades and subsequent loss in 
performance. Final balance of the system can be achieved by 
varying the length and diameters of the small flexible hoses. 

It must be emphasized that although data are empirical, 
LVHV systems require the same careful design as the more 
conventional ones. Abrupt changes of direction, expansions, 
and contractions must be avoided, and care must always be 
taken to minimize pressure losses. 

REFERENCES 

10.40.1 Thor Power Tool Company, Aurora, IL. 

10.40.2 Chamberlin, Richard I.: The Control of Beryllium 
Machining Operations. American Medical Associa- 
tion. Archives of Industrial Health, Vol. 19, No. 2 
(February 1959). 

10.40.3 Master Power, Inc., Westminster, MD. 

10.40.4 Hoffman Air and Filtration Div., Clarkson Indus- 
tries, Inc., New York. 



10-52 Industrial Ventilation 



Cone wheel used for 
internal grinding on 
castings and dies 




Q = 25-60 cfm/inch diam. 
Branch static pressure = 7 to 1 4" Hg 
Slot velocity = 24,000 to 39,000 fpm 
Flexible hose = 1" to 1 1/2" ID 
Extension hose = up to 8 ft long* 



Grinding wheel size< 



— 1" to 3" diam 
1 " to 4" long 



Peripheral speed - 6,000 to 10,000 linear fpm 

*Hose lengths may be extended up to 
a maximum of 50 ft using larger sizes 
between the too! hose and the tubing system. 



Reference 1 0,40.4 



:.;rican conference 
of governmental 
n d us tri a l h yg ie nis is 



EXTRACTOR HEAD FOR CONE 
WHEELS AND MOUNTED POINTS 



DATE 10-90 



'I CURE 



VS- 40-01 



Specific Operations 10-53 





Adapter plate 
to tit grinder 



Hood adjustable for wheel wear 
j I Hood 



t: 



- •lzxzzzz.zzz/zzxzz.x;'. zzzzzz.z.zr - 



Minimum clearance 



Q = 25-60 cfm/inch diam. or width 

Branch static pressure = 7 to 1 4 " Hg 

Slot velocity = 30,000 to 39,000 fpm 

Flexible hose = 1" to 2" ID 

Extension hose — Up to 8 ft long* 

Peripheral speed = 6,000 to 12,000 linear fpm 

*Hose lengths may be extended up to a 
maximum of 50 ft by using larger sizes 
between the tool hose and the tubing 
system. 



Reference 1 0.40.4 




Hood fitted to grinder 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HOOD FOR CUP TYPE 

SURFACE G RENDER 

AND WERE BRUSHES 



DATE 



10-90 



FIGURE 



VS-40-02 



10-54 



Industrial Ventilation 



Dust is extracted through slots built info 
the rubber sleeve; slots should be on both 
sides of the cutting edge of the chisel 




Q = 25 — 60 cfm/inch diarn. 
Branch static pressure = 7 to 14 'Hg 
Slot velocity = 24,000 to 39,000 fpm 
Flexible hose = f to 1 1/2" ID 
Extension hose ~ Up to 8 ft long* 
Chisel sizes = 13/16" octagona 
7/8" octagonal 
7/8" hexagonal 
*Hose lengths may be extended 
a maximum of 50 ft by using 
sizes between the tool and the tubing 
system. 



Slots 



up to 
larger 



Reference 1 0.40.4 



ERICAN CONFERENCE 
OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



PNEUMATIC CHISEL SLEEVE 



DATE W^90 |FIGURE VS-40~03 



Specific Operations 10-55 




O 



-Ov Adjustable 



X 



\ 



\ 






These extractor heads have been specif icaily designed 
for work done inside casting or in awkward places 
when radial wheels of small diameter are most suitable. 
The heads are narrower than the grinding wheels and can 
precede the wheel when a groove is being ground. 



^r ' 




Peripheral dust captured 
\ p- Fine dust controlled 
,_ — Heavy particles 



Q ~ 70-150 cfm/inch dlam. 
Branch static pressure = 7 to 1 4" Hg 
Slot velocity = 25,000 to 39,000 fpm 
Flexible hose = 1" to 1 1/2" ID 
Extension hose = up to 8 ft long* 



Grinding wheel sizes - 2" diam. xl/2" wide 
fo 8" diam. x 2" wide 



Peripheral speed = 6,000 to 15,000 linear fpm 



*Hose lengths may be extended 
up to a maximum of 50 ft by 
using larger sizes between the 
tool hose and the tubing system. 



Reference 10.40.4 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



EXTRACTOR HEAD FOR SMALL 
RADIAL GRINDERS 



DATE 



10-90 l FICURE VS-40 04 



10-56 Industrial Ventilation 




7 » 14 "Hg 



~ 10-30 cfm/inch diam. 
Branch static pressure = 7 to 14 "Hg 
Slot velocity = 24,000 to 39,000 fprn 
Flexible hose = 1" to 2" ID 
Extension hose ~ Up to 8 tt long* 



Sanding disc size — 2 to 9 diam. 

Peripheral speed - 4,500 - 14,000 linear fpm 

*Hose lengths may be extended up to a 
maximum of 50 ft by using 
larger sizes between the tool 
hose and the tubing system. 




Reference 1 0.40.4 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



EXTRACTOR HOOD FOR 
DISC SANDER 



DATE W-90 



FI 



cure VS-40-05 



Specific Operations 10-57 





7 - 14 Hg 



This design is suitable for sanders 
running up to 20,000 cycles per minute 




Q = 5 to 1 5 cfm/inch of perimeter 
Branch static pressure = 7 to 14" Hg 
Slot velocity = 15,000 to 39,000 fpm 
Flexible hose = 1 1/4" to 2" ID 
Extension hose = Up to 8 ft long* 



*Hose lengths may be extended up to a 
maximum of 50 ft by using larger sizes 
between the tool hose and the tubing 
system. 



Reference 10.40.4 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



EXTRACTOR TOOL FOR 
VIBRATORY SANDER 

| F1GURE ys- 40-06 



DATE 10-90 



10-58 



Industrial Ventilation 



System Notes 

Bell and socket, smooth-flow type 
tubing and fittings should be used 
throughout the system. 

When system is used for vacuum 
cleaning of abrasive materials, Schedule 
No. 40 pipe and cast iron drainage fittings, 
or heavier, should be used in place of 
tubing. 



To Atmosphere 







I.D. Plastic 






Hose Size 




cf m 


(inches) 


Disk sanders 3-9 inch diam. 


60-175 


1-1.5 


Vibratory pad sander - 4"x9" 


100 


1.25 


Router, 1/8"-f 


80-100 


1-1.25 


Belt sander 3" -4000 fpm 


70 


1 


Pneumatic chisel 


60 


1 


Radial wheel grinder 


70 


1 


Surface die grinder, 1/4" 


60 


1 


Cone wheel grinder 


90 


1.25 


Cup stone grinder, 4" 


100 


1.25 


Cup type brush, 6" 


150 


1.5 


Radial wire brush, 6" 


90 


1.25 


Hand wire brush 3" x 7" 


60 


1 


Rip out knife 


175 


1.5 


Rip out cast cutter 


150 


1.5 


Saber saw 


120 


1.5 


Swing frame grinder 2" x 1 8" 


380 


2.5 


Saw abrasive 3" 


100 


1.25 



2 1/8" 2 1/2" 3" 



2 \/l 



2 1/8" 



2 1/8" 



7" -72,000 rpm Chipping 6" -10,000 rpm 6"x1\ 10,000 rpm ( ; 

Disc sander hammer Cup stone grinder Wheel grinder U 




> separator 



2 1/8" 



Typi ca l Lay ou' 

Reference 1 0.40.4 



Swing frame 
grinder 




> 2 1/2" 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



TYPICAL SYSTEM- 
LOW VOLUME- HIGH VELOCITY 



DATE W-90 



FIG 



;ur E ys-40-20 



Specific Operations 1 0-59 



10.45 MACHINING 

The primary purpose of m etal cutting machines is to finish rough 
parts formed by other processes to specific dimensions. Finishing 
and shaping may be accompl ished by a variety of cutting tools such 
as saws, broaches, and chisel-shaped tool bits held in fixtures with 
fixed or movable drives. Cutting is accomplished by creating 
relative motion between the tool edge and the material blank. Chips 
of varying sizes are produced; chip size depends on the material 
being cut, feed rate of the tool, and relative speed or feed between 
the tool and the metal being shaped. 

Non-traditional methods of metal cutting and shaping in- 
clude electrochemical, electrodischarge, wire electrodis- 
charge, and laser beam machining. With the exception of the 
laser beam, each of the processes utilizes a circulating oil or 
water-base dielectric to facilitate molecular erosion as well as 
to remove process heat and particulate debris. The rate of 
metal removal is controlled carefully by regulating the flow 
of electrical current between the shaped anode or wire and the 
workpiece. The laser beam is used in a dry environment and 
metal cutting is accomplished by vaporizing the workpiece 
along the cutting edge with a focused beam of high energy 
light. The process is flexible and a variety of metallic and 
non-metallic materials can be shaped by this means. (I045 i} 

It is estimated that up to 97% of the work involved in 
conventional metal cutting results in heat. The rate of heat 
removal must be controlled carefully in order to protect both 
the cutting tool and the metallurgy of the work being cut. 
Where convection or radiant cooling is insufficient, a cutting 
fluid can be used to reduce friction, carry away generated heat, 
and, more commonly, flush away metal chips produced by the 
cutting process. Cutting fluids include straight-chained and 
synthetic mineral oils as well as soluble oil emulsions in 
water. A variety of water soluble lubri-coolants (1-5% mix- 
ture of lubricants, emulsifiers, rust inhibitors, and other 
chemicals in water) are used commonly, particularly for high 
speed metal working machines. In some applications the 
lubri-coolant mixture is applied as a mist by using a small 
volume of liquid in a high velocity air stream. In the more 
usual situation, liquid is applied by flooding the tool in the 
cutting zone to flush away cutting debris. The latter type 
system requires a low pressure pump with valves; filters; 
settling chamber to separate the fluid from the chips; and a 
reservoir which permits recirculation. Where liquids cannot 
be used, low temperature nitrogen or carbon dioxide gas can 
be used as a cooling media for both the tool and the cutting 
surface as well as a means of dispersing particulate debris. 

The hazards created by skin exposure to the water lubri-cool- 
ant mixtures, particulates and oil mist/vapor produced in the 
transfer of heat is best handled with engineering controls — pri- 
marily ventilation. An additional health concern is the fact that 
soluble oil emulsions provide a breeding ground for bacteria and, 
therefore, it is common practice to add biocides to prevent odor 
generation and decomposition of the oil mixture. Biocides and 



other additives may be primary skin irritants or cause hyper- 
sensitive dermatitis. It is for these reasons that mist, vapors, 
and particulates must be controlled adequately/ 1045 2) 

Mist and vapors from machining operations can be control- 
led by a combination of machine enclosure and local exhaust 
ventilation. Exhaust hoods and enclosures should be designed 
so the machine can be serviced easily and the operation 
observed when required. Hood sides should act as splash 
guards since an indraft of air will not stop liquid directly 
thrown from rotating parts. AH components should be robust 
and rigidly supported. To facilitate maintenance, service and 
tool adjustment, portions of the hood enclosure which are not 
permanently fixed should be designed for easy removal. 
Thought should be given also to the use of sliding, hinged, or 
bellows-connected panels in locations where frequent access 
is required. All windowed openings must be shatter-proof 
with appropriate internal lighting. All non-fixed panels should 
be designed with overlapping, drip-proof edges. The use of 
gaskets or seals on abutting panels is not recommended. 
Ventilation rates vary; however, a minimum of 100 fpm 
indraft usually is required to prevent vapor and mist from 
exiting the enclosure. A typical machine enclosure will re- 
quire a volumetric flow rate of from 400-500 cfm minute. 
Additional air may be required to control heat generated 
within the enclosure as well as to maintain adequate vision. 
Where coolant flumes are used for chip transport, additional 
exhaust ventilation is required to control air entrapment. 
Baffles above the liquid level are beneficial and flumes should 
be enclosed to the extent possible. 

Local ventilation control is the preferred method — particu- 
larly in machine environments which are temperature control- 
led with refrigerated air conditioning systems. In more open 
workrooms, the use of dilution ventilation may be adequate 
to control air contaminants. For further information on dilu- 
tion ventilation, see Chapter 2 of this manual. 

REFERENCES 

10.45.1 Rain, Carl: Non-traditional Methods Advance Ma- 
chining Industry. High Technology (November/De- 
cember 1957). 

1 0.45.2 O'Brien, Dennis; Frede, John C: Guidelines for the 
Control of Exposure to Metal Working Fluids. Na- 
tional Institute for Occupational Safety and Health 
(February 1978). 

10.45.3 Schulte, H.F.; Hyatt, E.C.; Smith, Jr., F.S.: Exhaust 
Ventilation for Machine Tools Used on Materials of 
High Toxicity. American Medical Association Ar- 
chives of Industrial Hygiene and Occupational 
Medicine, Vol. 5, No. 21 (January 1952). 

10.45.4 Mitchell, R.N.; Hyatt, E.G.: Beryllium— Hazard 
Evaluation and Control Covering a Five-Year Study. 
American Industrial Hygiene Quarterly, Vol. 18, 
No. 3 (September 1957). 



10-60 Industrial Ventilation 



^. 



r ~~ 'a 



o 



*> 



>: 



Booth sized - 
to suit 
work 




Q at booth = 225 cfm/ft open area 
Q at bottom = 350 cfrr, 

Minimum duct velocity - 4000 f pm 

h e = 1 -75 VPd , at point A 



RICAN CONFERENCE 
OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



METAL CUTTING- BANDS. AW 



date 10-90 F1GURE VS-45-01 



Specific Operations 10-61 




«&b«^^ 






^— Transparent cover 
normally closed 



? 




Q-300 cfm/ft 2 of open area 
Minimum duct velocity = 3500 fpm 

he = 0.35 VPd 



Reference 10.45.4 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HIGH TOXICITY MATERIALS 
MILLING MACHINE HOOD 



DATE 



11-90 | FIGURE VS-45-02 



10-62 



Industrial Ventilation 







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o 




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2 


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x 



Reference 10.45.3 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



METAL SHEA RS 
HIGH TOXICITY MATERIALS 



DATE 



11-90 



FIGURE 



VS-45-03 



Specific Operations 10-63 




END VIEW 



Hinged die cover 




^--~ Parts discharge 



SIDE VIEW 



Cold header : 

2 

- 750 cfm/ft of die opening 

Minimum duct velocity = 3500 fpm 

h = 1 .0 VP + 0.25 VP. 
e s a 

Parts discharge and container : 

Q = 100 cfm/ft of hood length 
Minimum duct velocity = 3500 fpm 
h„ = 0.25 VP, 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



COLD HEADING MACHINE 
VENTILA TION 



DATE 



11-90 8 FIGURE VS-45-04 



10-64 



Industrial Ventilation 




Reference 10.45.4 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



LATHE HOOD 



DATE 



11-90 | FIGURE VS- 45=05 



Specific Operations 1 0-65 



10.50 MATERIAL TRANSPORT 

Ventilation of material transport systems generally requires 
the use of an exhausted enclosure because of the motion and 
quantity of material involved. If the enclosure were perfectly 
air tight there would be no need for exhaust. However, there 
usually are cracks and other leak points in addition to the 
openings necessary for personnel and material access. 

For enclosures where there is little motion and low material 
quantity, the exhaust rate is the product of the total openings 
in square feet and some velocity between 50 and 200 fpm. 
However, in some cases the inward flow of material and 
entrained air can overwhelm the exhaust flow rate calculated 
on the basis of enclosure openings. In such cases the material 
flow rate, the dustiness of the material, and the height of fall 
in transferring from one surface to another must be considered 
in the system design/ 1050 K I0502) Other design factors include: 

1. The rate of air induction into the space. 

2. The location of cracks or other openings in relation to 
the "splash" or agitation of material during transfer. 

3. The need to avoid excessive product withdrawal. 

4. Adequate air flow for dilution of interior concentra- 
tions for visibility or safety from explosions. 

10.50.1 Bucket Elevators: Air motion caused by the 
bucket moving within the elevator is not significant. The 
motion of buckets in one direction is offset by the opposite 
flow. Consequently, an exhaust rate of 100 cfm/ft 2 of elevator 
cross-section is adequate for most elevator applications (see 
VS-50-01 for details). Additional ventilation is required as 
materials enter and leave the elevator (see VS-50-10, VS-50- 
20 and VS-50-2 1 ). Handling hot material often causes signifi- 
cant thermal buoyancy which requires increased exhaust 
ventilation to overcome this challenge. 

10.50.2 Conveyors: Dust from the operation of belt con- 
veyors originates mainly at the tail pulley where material is 
received and at the head pulley where material is discharged. 
The exhaust requirement at the head pulley is generally small 
because air is induced downward and away from this transfer 
point. An exhaust rate of 1 50-200 cfm/ft 2 of opening often is 
adequate. 

At the tail pulley, the exhaust requirements are determined 
by the amount of air induced by the delivery chute. An exhaust 
of 350 cfm/ft 2 of belt width often is adequate where the 



material does not fall more than 3 feet. The exhaust point 
should be located at least twice the belt width away from the 
point where the material hits the belt. Where the material falls 
more than 3 feet, additional exhaust is required (see VS-50-20 
for details). Note that very dry or dusty material may require 
flowrates 1.5 to 2.0 times these values. 

Belt conveyors should be covered and exhausted at 30 foot 
intervals at arateof350 cfm/ft of belt width. Vibrating feeders 
should be exhausted at a rate of 500 cfm/ft of feeder width. 
Rubber or canvas flexible seals should be provided from the 
feeder sides and end to the hopper sides and end. 

The conveying of toxic material requires additional care in 
enclosure design to ensure that no air leaks out and that 
sufficient access is available for inspection and cleanout. The 
head pulley should be equipped with a scraper or brush (see 
VS-50-21). 

10.50.3 Bin and Hopper Ventilation: For the mechanical 
loading of bins and hoppers, the exhaust rates previously 
listed for belt conveyors are appropriate. An exhaust rate of 
150 cfm/ft 2 of hopper is adequate for manual loading opera- 
tions. The enclosure should cover as much of the hopper 
opening as possible. 

10.50.4 Loading and Unloading: For loading and un- 
loading operations, a ventilation rate of 1 50-200 cfm/ft 2 of 
enclosure opening is adequate provided the enclosure is large 
enough to accommodate the "splash" effect. The entrance to 
enclosure for truck dumps should be covered with flaps to 
minimize ventilation requirements. Rotary or bottom car 
dumps generally are exhausted at the rate of 50—100 cfm/ft 2 
of hopper area. 

REFERENCES 

10.50.1 DallaValle, J.'M: Exhaust Hoods. Industrial Press. 
New York (1946). 

10.50.2 Hemeon, W.C.L.: Plant and Process Ventilation. 
Industrial Press, New York (1963). 

10.50.3 Rajhans, G.S.; Bragg, G.M.: Engineering Aspects of 
Asbestos Dust Control. Ann Arbor Science Publica- 
tions, Inc. Ann Arbor, Ml (1978). 

10.50.4 National Grain and Feed Association: Dust Control 
for Grain Elevators. NGFA, Washington, DC 
(1981). 



10-66 Industrial Ventilation 



Alternate exhaust point 



Preferred exhaust 
point 



Additional ventilation for 
hopper, bin, or screen 
see Vs-50-10 & 
VS-99-01 



For casing only 

Q =100 cfm/ft casing cross section 
Minimum duct velocity — 3500 fpm 

h e = 1 .0 VPd or calculate fro 
individual losses 



Take-off at top for hot materials, 
at top and bottom if elevator is over 
30 ft high, otherwise optional. 




CONVEYOR BELT DISCHARGE VENTILATION 



BELT SPEED 


FL0WRATE 


Less than 200 fpm 


350 cfm/ft of belt width. Not less than 150 cfm/ft of opening. 


Over 200 fpm 


500 cfm/ft of belt width. Not less than 200 cfm/ff 2 of opening. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



BUCKET ELEVATOR 
VENTILATION 



DATE 



1-91 l F1GURE VS- 50-01 



Specific Operations 10-67 



Enclosed loading point- 




Minimum duct velocity = 3500 fpm 
Q — 200 cfm/ft of all open are 
h e - 0.25 VP d 



MECHANICAL LOADING 



Belt speed 
Less than 200 fpm 

Over 200 fpm 



Flowrate 
350 cfm/ft of belt width. 
Not less than 150 cfm/ft of opening. 

500 cfm/ft of belt width. 

Not less than 200 cfm/ft of opening. 



rBo-oth to accommodate barrel, bag, etc. 




Grate bars 



Hopper / 




Booth to cover as much 
of hopper as possible 



Minimum duct velocity — 3500 fpm 
Q = 150 cfm/ft 2 face 
h e = 0.25 VP d 

MANUAL LOADING 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



BIN & HOPPER VENTILATION 



DATE 



1 _ 91 jfigure VS _ 50 ^ 10 



10-68 



Industrial Ventilation 




45 



24" min 



1 . Conveyor transfer less 
than 3' talk For greater 
fall, provide additional 
exhaust at lower belt. 
See 3 below. 
h a = 0.25 VP, 



belt width 



Close face to 
bottom of belt 



As close as 
practical 



\(Se 



Elevator 

exhaust 

e VS-50 



,A 



Tote box 



vw^- 



■01) 




2. Conveyor to elevator with 
magnetic separator. 
h e = 0.25 VP d 

DESIGN DATA 
Transfer points: 

Enclose to provide 150 - 200 fpm indraft 
at all openings. (Underground mining 
tunnel ventilation will interfere with 
conveyor exhaust systems.) 



2" clearance for load' 
on belt 




DETAIL OF BELT OPENING 



Chute to belt transfer and conveyor 
transfer, greater than 3' fall. 
Use additional exhaust at (V) 
for dusty material as follows; 
Belt width 12" -36", 0^700 cfm 
Belt width above 36", 0-1000 cfm 

h = 0.25 VP 

e d 



Note: Dry, very dusty materials 
require exhaust flowrates 
2,0 times stated values. 



may 
1.5 



to 



= 350 cfm/ft belt width for belt 

speeds under 200 fpm. (minimum) 
= 500 cfm/ff belt width for belt 
speeds over 200 fpm and for 
magnetic separators, (minimum) 

Minimum duct velocity = 3500 fpm 

h e = 0.25 VP d 

Conveyor belts: 

Cover belt between transfer points 
Exhaust at transfer points 
Exhaust additional 350 cfm/ft. of belt 
width at 30' intervals. Use 45 
tapered connections. 



AMERIC AN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



CONVEYOR BELT VENTILATION 



DATE 



l-Ql ■ JFIGURE VS- 50^20 



Specific Operations 10-69 



Settling box 



Internal skirt board 



Troughing belt — 




, — Totally enclosed conveyor, 

/ leakage tactor depends on 
type ot construction 



Cfeanout and inspection 
doors 



Scraper conveyor 



Return belt scraper 
or brush 



Q - 250 cfm/ft of open area 
Minimum duct velocity = 3500 fpm 
h e = 0.4 VP d 



Reference: 1 0.50.3 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



TOXIC MATERIAL BELT 
CONVEYING HEAD PULLEY 



DATE 1-Ql 



FIGURE VS~ 50-21 



10-70 Industrial Ventilation 



l~— Square setiling box 
/ (Designed for 150-200 fpm velocity) 



Exhaust duct 



Internal dust 
shield 



Flexible skirt 
board 




Rock box 



Completely 

enclosed 

conveyor 



Axial loading chute 
T roughing belt 
Clean-out and inspection 



Q = 250 cfm/ft of open area 
Minimum duct velocity = 3500 fpm 
h P = 0.4 VP H 



Reference: 10.50.3 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
IND USTEI AL H YGIENISTS 



TOXIC MATERIAL 
CONVEYOR BELT IDA 



DATE 



7- <)l l llf:im: I;-' no :■::■: 



Specific Operations 10-71 




10-72 Industrial Ventilation 



Telescoping 
grain spout 




Air conduit 
must be flex 
hose or swivel 
oint 



Air velocity 
500 fpm 



Outer sleeve 

must be telescoping 

or collapsible 



-Dead box 



Must maintain 
1 2" clearance 
to be most 
effective 



Grain pile 



Reference 10.50.4 




AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



TRUCK LOADING 



DATE 



2-91 



figure ys- 50-31 



Specific Operations 10-73 



10.55 METAL MELTING FURNACES 

This set of VS-prints describes hood designs for a variety 
of metal melting furnaces, including electric induction, carb- 
on arc, convention, and crucible, which use natural gas or 
electric resistance elements as the heat source. Exhaust ven- 
tilation usually is required to control specific oxides associ- 
ated with the metal being melted or contaminants carried in 
the scrap charge. In some cases, a single hood will suffice for 
charging, melting, and pouring. In other cases, a separate 
hood, remote from the primary melter, may be required for 
charging because of the nature of the charge or the large open 
area necessary at this phase of the operation. This is true 
particularly of electric arc furnaces which are completely 
open for charging. After charging, the port exhaust can be 
used to achieve control during the remainder of the melting 
and pouring cycles. 

All metal melting will produce a slag which must be 
removed prior to pouring. This activity may produce a signifi- 
cant release of oxides and may require a separate exhaust 
system for oxide and/or dross control Where metal purifica- 
tion is performed directly within the furnace or melting vessel, 



such as the addition of oxygen or chlorine, additional exhaust 
may be required to contain the rapidly generated plume. 

All systems must be designed to include the increase in air 
temperature under operating conditions to insure an adequate 
air flow into the hood. The air temperature rise is usually 
relatively low except where metal innoculants or oxidizers are 
added to the molten charge. During this phase of metal 
melting, a significant temperature rise will occur and it is 
customary to provide a large hood in which gas expansion can 
take place. 



REFERENCES 

10.55.1 American Air Filter Co.: 
(January 1946). 



Rotoclone Dust Control 



10.55.2 Kane, J.M.: Foundry Ventilation. The Foundry (Feb- 
ruary and March 1 946). 

10.55.3 Kane, J.M.: The Application of Local Exhaust Ven- 
tilation to Electric Melting Furnaces. In: Transac- 
tions of the American Foundrymen's Association, 
Vol. 52, p. 1351 (1945). 



10-74 Industrial Ventilation 




Crane track 



Close end with panel -^ 



Sliding panels 
on rollers 



-Track for panels - 




/ 2 
Q ~ 200 cfm/ft of opening including doors, 

plus products of combustion correcfed 

for temperature. Row of crucibles 

Minimum duct velocity — 3500 fpm 



h e = 0.5 VP d 



Note: Same principle of sliding or swinging 
doors is applied to individual furnace 
enclosures. 



Exhaust stack 




T 

i 



"Fireproof drop panel from roof. 



Canopy to clear crane; or provide slot for crane bridge; 
or separate cranes inside and outside; or use 
manual crucible removal. 



Q — 200 cfm/ft of total opening, minimum, plus 
products of combustion corrected for temperature. 



4ERICAN CONFERENCE 
OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



MELTING FURNACE 
CRUCIBLE, NO N- TILT 



DATE 



02-91 ™ure VS-55-01 



Specific Operations 



10-75 




-Door guides- 



Counterweighfed or 
spring-loaded sliding 
doors front and back 
if necessary — \ 



^ Door to extend beiow 

top of furnace if 
possible. 




Solid side 
panels 



Q - 200 LW; but not less than 
200 cfrn/ft of all openings 
with doors open. Correct for products 
of combustion and temperature. 

Minimum duct velocity ~ 3500 fpm 

h e = 0.25 VP d 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



(ELTING FURNACE 
TILTING 



DATE 



02- 91 



FIGURE 



ro p: p: 

u ~~ O O ' 



02 



10-76 



Industrial Ventilation 



Flanged by-pass connection; no exhaust 
during furnace tilting and pouring 



<T 



Hood is fastened 
to furnace roof 
and swings with 
roof — \ 



Furnace body 



~\ 



Exhaust transition 



— Electrodes 




Slag door hood 



Counterweight 



Hood over pouring 
spout 



Close Capture: 

For Q, SP and operating temperature, consult manufacturer 
Approximate exhaust rate = 2500 scfm/ton of charge 

Alternate designs: 

1 . Some exhaust designs utilize direct furnace roof tap. For details 
consult manufacturer. 

2. Canopy hoods require large exhaust and are not recommended 
Canopy hoods can be used as secondary hoods to capture 
fugitive emissions. 



References 10.55.1, 10.55.2, 10.55.3 



ERICAN CONFERENCE 
OF GOVERNMENTAL 
NDUSTRIAL HYGIENISTS 



MELTING- FURNACE 
ELECTRIC, TOP ELECTRODE 



date Q2-91 Ifigure vs-55-03 



Specific Operations 10-77 




Hingec 
door 




Q — 400 cfm/ft of opening 
Minimum duct velocity = 3500 fpm 
h = 1.78 VR. + 0.25 VP H 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



MELTING- FURNACE! 
E LEG ERIC ROCKING 



DATE 



02-91 



figure: 



55-04 



10-78 



Industrial Ventilation 



Q = 100-200 cfm/fi of 
opening plus volume of 
products of combustion 

Duct velocity = 3500 fpm 



h e = 0.50 VP d 



* Correct for temperature 
See Chapter 3 
for additional 
information. 



Pot - ^ 



Furnace 




NOTE: Separate flue required if com- 
bustion gases are not vented 
through the hoods. 



— Work openings. 
I Keep as small as practical 
Doors advisable, 



Dross poi 



Door for dross pan 
removal. 



Dross chute, min angle - 60 



STATIONARY FURNACE OR MELTING POT 



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MELTING POT 8c FURNACE 



DATE 



02-91 | FIGURE VS- 55-05 



Specific Operations 10-79 



Flange is necessary 
180° slot if possible 




Flange 



\ Size for 1/2 

y slot velocity 



Q= 175 cfm/ft of furnace top with curved slot 

and flanges. 
Slot velocity- 2000 fpm 
Minimum duel velocity= 3500 fpm 
Entry losses^ 1.78 VP S + 0.25 VP d 



AMERIC AN C NFERENC E 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



CRUCIBLE MELTING FURNACE 
HIGH TOXICITY MATERIAL 



DATE 



02 91 | F1GURE VS 55 06 



10-80 Industrial Ventilation 



Electrical cables 

and cooling r Pouring channel 




Rotating joint 



FRONT VIEW 



Slot 




Charging door 




SIDE VIEW 



Rotating joint 
TOP VIEW 



Q = 350 cfm/ft. open area. Correct for 
temperature and combustion products. 

Entry loss = 1 .78 VP S + 0.5 VP d 

Slot velocity = 2000 fpm 

Minimum duct velocity = 3500 fpm 




ISOMETRIC VIEW 



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INDUCTION MELTING FURNACE 
-TILTING 



DATE 02-91 1 FIGURE VS-55- 07 



Specific Operations 10-81 




Use top baffle when 
operations permit 



Wide flange 



— - 1.6 W m I n . -*- 



7-Z^ C? r TV~7~7~? r 7^rrZ u. 



Mold 



— Conveyor 
SMALL MOLDS 




Ciose clearance 



Q ~ 200 (10 X -f A) where A equals hood area. 

Minimum duct velocity = 3500 fpm 

h e = 1.78 VP S + 0.25 VP d 

Use slots for uniform distribution, size slots 

for 2000 fpm 



PARTIAL SIDE ENCLOSURE 




Use slots for distribution 
Slot velocity = 2000 fpm 



T - 3' 



Note; 

For large molds and ladles 

provide large - draft hood 

similar to shakeout. 

Q = 400 cfm/ft 2 working area. 



Q = 200 - 300 cfm/ft of hood length. 



AMERICAN CONFERENCE 
OF GOVERNMENTAL 

INDUSTRIAL HYGIENISTS 



POURING STATION 



DATE 



02-91 | F1GURE VS-55-10 



10-82 



Industrial Ventilation 



Flange type fitting for 
easy removal of hood 
(if necessary) -••- - . 



._J 



Hinged baffle for preventing short 
circuiting of air. 




W 



h~ 3/4 D 



Note: Place hood as close to machine as 
possible, if more than 4 inches 
from back of machine, hinged side 
baffles should be used. 



Note: Products of combustion 
require separate flue or 
may be vented into hood. 



Q = 300WH 

h e = 0.25 VP d 

Minimum duct velocity = 2000 fpm. 



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OF GOVERNMENTAL, 
INDUSTRIAL HYGIENISTS 



FIXED POSITION 



7 CASTING HOOD 



DATE 02-91 



figure VS-55- 



20 



Specific Operations 10-83 



Duct and mobile hood 
match here 




Crane beam 



Die hoist 



Hood travels on die 
hoist crane 



Q = 300WL 

Minimum duct velocity = 2000 fpm 

h e ^ 0.25 VP d 



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INDUSTRIAL HYGIEN1STS 



MOBILE HOOD 
DIE CASTING 



DATE 



02-91 



FIGUR 



E VS-55 21 



10-84 



Industrial Ventilation 



10.60 MIXING 

Mixing operations combine a large variety of materials, 
usually without significant chemical reactions. This section 
includes categories of mixing operations. 

10.60.1 Mixing and Mulling: Mixers and mullers require 
exhaust ventilation to provide a minimum velocity of 1 50 fpm 
through all openings. Additional ventilation may be required 
when flammable solvents are used. The dilution ventilation 
rates should maintain concentrations within the muller below 
25% of the Lower Explosive Limit (LEL). Some codes or 
standards may require ventilation rates which ensure the 
concentration of flammable vapor is maintained below 20% 
of the LEL. 



chemical reactions or may be caused by the elevated tempera- 
ture of the mixed materials. Particulate emissions can occur 
during additions as well as during mechanical blending. 

The roller mill ventilation shown in VS-60-12 encloses the 
roller mill to the maximum extent possible except for a front 
opening of sufficient height and width to permit operator 
access and material entry /removal. An air curtain directed 
upward towards the top of the enclosure provides a barrier to 
contaminant escape but still permits operator access and 
material entry/removal. The design values provided are criti- 
cal for proper operation. 

REFERENCES 



10.60.2 Roll Mixing: Themachinery shown in this subsec- 
tion is used to mix and blend quantities of viscid materials, 
such as rubber and plastic, with additives that are dry powers 
or liquids. Emissions of gases and vapors may evolve due to 



10.60.1 Hampl,V.; Johnston, O.E.; Murdock, D.M.: Appli- 
cation of an Air Curtain Exhaust System at a Milling 
Process. American Industrial Hygiene Association 
Journal, 49(4): 167-175 (1988). 



Specific Operations 



10-85 



Skip hoist hood 
Q = 250 LW 



Skip — - 




To prevent condensation, insulation, 

strip heaters or dilution fitting may 
be necessary. — — 



Enclosing 
hood 



Skip hoist hood 

h P =1.78 VPo -I- 0.25 VP r 



d 



<? 




-*^ 


- Slots 

Opening for — ^ 
skip loading \ 




/ 




! '• 


A 




I I 










I I 








| I 






L — 


i i 




Skip 


I I 

I I 
i I 










* i i 



H-JJ I 



/ ; v / 



\ I ' T 
S I 




150 cfm/ft through all openings but not less than: 



Mutler diarn. feet 


Exhaust, cfm 


4 


750 


6 


900 


7 


1050 


8 


1200 


10 


1575 



Minimum duct velocity = 4000 fpm 
h e = 0.25 VP d 

Notes: 1. Other types of mixers: enclose as much as possible and provide 150 cfm/ft 
oi remaining openings 

2. When flammable solvents are used in mixer, calculate minimum exhaust rate 
for dilution to 25% of the LEL. See Chapter 2 

3. For air cooled mullers see VS — 60 — 02 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



MIXER AND MULLER HOOD 



DATE 



11-90 1 RGURE VS-60 01 



10-86 



Industrial Ventilation 



Loading 
hopper 




To prevent condensation 
insuiation, strip heaters 
or dilution fitting may 
be necessary. 



Tight enclosure — 



Batch hopper 

Side hood or — 
booth 



Bond hoppe r— 
■Enclosing hood 




Low — velocity 
duct used with 
cooling type 
m u I ! e r . 



Cooling fan 
blow — through 
arrangement 



Minimum exhaust rate, cfm 


Muiler type 


Location 

| No cooling 


Blow-though 
cooling 


Draw — though 
cooling 


Batch hopper Note 1 
Bond hopper 600 
Muller: Note 2 
4' diameter 750 
6' diameter 900 
7' diameter 1050 
8' diameter 1200 
10' diameter 1575 


600 

600 

Note 3 


Note 1 

600 
Note 3 



Minimum duct velocity = 4500 fpm 
h e = 0.25 VP 
Notes: 

1 . Batch hopper requires separate exhaust with blow-through cooling. With other fan 
arrangement (muller under suction), separate exhaust may not be required. (If skip 
hoist is used, see VS-60 — 01) 

2. Maintain 150 fpm velocity through ail openings in muller hood. Exhaust flow rates 
shown are the minimum for control. 

3. Cooling mullers do not require additional exhaust if maintained in dust tight 
condition. Blow-through fan must be off during loading. If muller is not dust tight, 
exhaust as in note 2 plus cooling air flow rate. 

4. When flammable solvents are used in mixer, calculate minimum exhaust flow rate 
for dilution to 25% of the LEL. See Chapter 2. 



AMERIC AN CONFERENCE 

OF GOVERNMENTAL 
I N D US TRI A 1, H YG IE N IS TS 



AIR COOLED 
MIXER AND MULLER 



DATE 



- 90 



F,GURE VS- 60 02 



Specific Operations 10-87 



Charging hood 



To suit method of 
feed to mixer. 



Platform 



„_V„._ 




6 D i a m . duct 



-800 cf rrr 



Trunnion exhaust 



= 200-300 cf m/ft open face area. 

500 cfm/ft of belt width if belt feeder used. 
Duct velocity - 3500 fpm minimum. 
h e = 0.25 VP d at hood 



1.0 VF^ at irunnion 



Consult mixer manufacturer for specific recommendations. 



eric an conference 
of governmental 
:ndustri al h ygienists 



BANBURY MIXER 



DATE 



11-90 l F1GURE VS-60 10 



10-88 Industrial Ventilation 



"^<:> 




Flange — 



v 



Side baffles 
desirable \ 



-^r 



"■o-.. 











' 1 












( 

V 


r 


X 




' J 




/ 


— --. 












-..]..,. 




( 




I 




\ 
\ 



■ Safety 
brake - 






_l 






o 


Rolls 


( 












[ 






/ 




i 

I 


r " / " 

i / 

V"7 


i 



Better location of 
brake bar 



= 125 cfm/ft open face area 
Minimum duct velocity = 2000 fpm 
h e ^ 0.25 VP d 



NOTE: Both sides may be ooen 




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Specific Operations 10-89 



i 





Operator opening ^y 




See note 4 ^ 

_ _. i 










/ 




Supply y/ 
manifold —^ 





r 



SIDE VIEW 



i 1 

-^-■■"T 



L d 



\ 0.25' Holes (diam.) 

0.75" Apart (center to center) 



FRONT VIEW 

d > 0.145 -\/~C ft (Manifold diameter) 

Q s = 29.6 L cfm (Supply flow) 

Q E = 320 yx" L cfm (Exhaust flow) 

SP S ~ Manifold pressure = +1.2 "wg 



Notes: 1 . All dimensions in feet, except as otherwise noted. 
2. For access openings other than operator opening, 
increase Q E by 100 cfm/fi of opening area. 



3. If operator opening is required on both sides of mil 
total E will be sum 

4. X not to exceed 6 ft. 



total G E will be sum of Q E for both sides. 



Reference 10.60.1 



AMERICAN CONFERENCE 

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ROLLER MILL VENTILA TION 



DATE 



11-90 | FIGURE VS 60-12 



10-90 



Industrial Ventilation 



10.65 MOVABLE EXHAUST HOODS 

Movable exhaust hoods provide control for moving con- 
taminant sources. In general, movable hoods are associated 
with flexible exhaust ducts; traveling exhaust hoods, swivel, 
slip, or telescoping joints in duct sections; or systems which 
separate the hood from the duct for access to the process. 

Flexible exhaust duct is possibly the most common way of 
providing a movable exhaust hood. A section of flexible duct 
connects to a relatively small exhaust hood. The duct section 
and hood may be supported by a counter-weighted or spring- 
loaded, hinged arm that allows the positioning of the exhaust 
hood near the source of contaminant generation. This type of 
device is known as "snorkel," "elephant trunk," or "flex-arm" 
exhaust. Illustrations of this type of exhaust in Chapter 10 
include Welding Exhaust (VS-65-0 1) and Granite Cutting and 
Finishing (VS-65-02). Flexible exhaust duct use is also illus- 
trated in Barrel Filling (VS-15-01), Metal Spraying (VS-90- 
30) and Service Garage Ventilation (VS-85-01 and 
VS-85-02) and low volume/high velocity systems (Section 
10.40). Frictional resistance can be very high in the flexible 
duct section as can the negative pressure or suction. Materials 
used in the construction of the duct may be metal or non-metal 
and the losses vary over a wide range depending on the type 
and use. The application data provided by the manufacturer 
must be included in the design development of the system. 
When used, flexible duct should be non-collapsible with 
minimal length to reduce undesirable bends which will result 
in excessive static pressure losses. 

Traveling exhaust hoods may be used for a variety of 
operations where the contaminant source moves from one 
point to another. This type is more suited to heavy-duty 
requirements than the flexible exhaust duct. Examples of 



these operations include flame and plasma cutting, foundry 
pouring, heavy abrasive cutting, and similar operations. An 
illustration of a traveling exhaust hood is the Hawley Trav-L- 

Vent(VS-65-03).< 10651 > 

Telescoping or slip joints are duct sections that overlap, 
slide, or rotate to allow a section of one exhaust duct to slide 
into or rotate around another section of duct. This arrangement 
allows an exhaust hood to be moved from one position to 
another without disconnecting the duct or hood. Illustrations 
of the slip joint include the Core Grinder (VS-80-13) and 
Granite Cutting and Finishing (VS-65-02). This joint or duct 
section may include a swivel feature to permit rotation of the 
hood away from the process equipment being exhausted and 
may be used in the horizontal or vertical plane. Guide rails 
may be required for the horizontal application while pulleys 
and counter-weights may be required for vertical applications. 

Separating exhaust duct sections is another method of 
providing a movable exhaust hood. This concept requires that 
the exhaust duct separate at or near the exhaust hood when 
the hood is moved for access to the process equipment or the 
process equipment moves. Illustrations of this method include 
hoods for Top Electrode Melting Furnace (VS-55-03), Core 
Making Machine— Small Roll-Over Type (VS-20-11), and 
Mobile Hood, Die Casting (VS-55-21). Alignment of the 
exhaust duct when the hood is in place is critical, and any 
opening at this point should be included in the total exhaust 
calculations. Also, during the separation period, little or no 
exhaust control will be available at the contaminant source. 

REFERENCES 

10.65.1 Vulcan Engineering Co.: Hawley Trav-L-Vent 
Equipment Specifications and Layout. Helena, AL. 



Specific Operations 10-91 



Overhead 
support - 



To exhaust 
system 



Swivel 




P 



1 Cleanaut 



%%-p 

Galvanized hood — ' 
FLEXIBLE EXHAUST CONNECTIONS 






Swivel base and angle 
iron duct support arm 



Support 





Telescoping flex 
duct support 



Swivel joints 




j 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
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MOVEABLE EXHAUST HOODS 



- 



DATE 



02- 91 



figure VS-65- 01 



10-92 Industrial Ventilation 




r- 4" I.D. flex, duct (rubber) or metal 
/ duct with telescope joints 



View A-A 



- Ball joint 



§Sfiz-i 




3" x 8" opening with 
3" metal flange 



PNEUMATIC HAND TOOLS 

Q = 400 cfm minimum, tool 10"max distance from hood 
Minimum duct velocity ~ 4000 fpm 



Flexible duct to 
branch duct 
Attach to machine 

/™ Chisel 



Stone 



Abrasive blasting to be done in a room 
or cabinet; 500 fpm at all openings. See 
"Abrasive Blasting", VS-80-01 



V 




SURFACE MACHINE HOODS 



Ho od 
Baby surfacer 
Medium surfacer 



cfm Branc h dia m . 

400 4" 

600 5" 



h e = 1.0 VP d (at point A) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
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DATE 02-91 



GRANITE CUTTING 

AND FINISHING 

T CURE VS-65-02 



Specific Operations 10-93 




Exhaust 
hood 



Typical Design Ranges 

Q = 2500 - 10,000 cfrn 
V = 3500 - 4000 fprn 
h e = 6"wg approx. total 
SP h = calculate separately 



Note: This is a patented system with many variations- 
Consult manufacturer for applications. 



See Reference 10.65.1 



AMERICAN CONFERENCE 

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INDUSTRIAL HYGIENISTS 



HA WLE Y TRA V- L - VENT' 
PERSPECTIVE LAYOUT 



DATE 



02-91 P' GURE VS-65-03 



10-94 



Industrial Ventilation 



10.70 OPEN SURFACE TANKS 

Ventilation rates for plating, cleaning, and other open sur- 
face tank operations will depend on a number of parameters 
which include materials, tank configuration and location, and 
type of ventilation system. This section describes four 
hood/ventilation types: enclosing and canopy hoods, lateral 
exhaust, and push-pull. 

Enclosing hoods usually consist of a lateral hood with one 
end panel (two sides open) or panels at both tank ends (one 
side open). This hood configuration can provide increased 
efficiency by reducing the effects of cross-drafts and by 
directing more of the hood air flow over the tank open surface. 

Canopy hoods may be open on four sides (free standing) or 
on three sides (such as against a wall). Control is achieved by 
air flow into the hood. It is, however, difficult in many cases 
to achieve sufficient control velocity without excessive air 
flow rates. Canopy hoods should not be used with highly toxic 
materials, in locations where high cross-drafts are unavoid- 
able, or where the worker must bend over the tank. 

Lateral exhaust consists of a slot hood which controls 
emissions by pulling air across the tank. A single slot may be 
used on one side of the tank where the tank width is 36 inches 
or less. For widths greater than 36 inches and where the 
process configuration will allow, two slot hoods on opposite 
sides of the tank or a slot hood along the tank centerline may 
be used. A single slot may be used up to a tank width of 48 
inches but only if the material hazard class is low and if 
cross-drafts are not present.(see Section 1 0.70.1) 

The air flow required will be that necessary to achieve a 
minimum control velocity determined by the hazard class of 
the material used for operation and the particular tank/venti- 
1 at ion system configuration. The procedure for determining 
the class and minimum control velocity for the three preced- 
ing hood types is provided in Tables 10.70.1-10.70.7 and the 
accompanying text. Exhaust flow for a canopy hood is deter- 
mined from VS-99-03 and for a booth hood from Figure 3-1 1 
where W is the total opening width. The exhaust flow for a 
lateral hood is determined from Table 10.70.4. 

Air and/or mechanical agitation of the tank solution may 
be used as an aid to the plating or cleaning process. Mechani- 
cal agitation creates a rolling motion and usually will not 
affect tank emissions. However, air agitation creates a boil- 
ing-like condition and may significantly increase tank emis- 
sions, thus creating need for increased exhaust flow to provide 
effective control. 

Push-pull ventilation consists of a push jet located on one 
side of a tank with a lateral exhaust hood on the other 
gjdgOo.70.1-10.70.3) Tank emissions are controlled by the jet 
formed over the tank surface. The jet captures the emissions 
and carries them into the hood. As the jet velocity, at all 
locations across the tank, is higher than the maximum control 
velocities specified for canopy, enclosing, or lateral exhaust 



hoods (Table 10.70.3), the push-pull exhaust flow is deter- 
mined on the basis of that necessary to capture the jet flow 
and is independent of the hazard classification. Push-pull 
design criteria are provided in VS-70-10, -1 1 and -12. 

10.70.1 Tank Design Considerations: 

1 . Duct velocity = any desired velocity (see Chapter 3). 

2. Entry loss = 1.78 slot VP plus duct entry loss for slot 
hoods. For canopy or enclosure hoods, entry loss = 
duct entry loss. 

3. Maximum slot hood plenum velocity = 1/2 slot veloc- 
ity (see Chapter 3). 

4. Slot velocity = 2000 fpm unless distribution is pro- 
vided by well-designed, tapered takeoff. 

5. Provide ample area at the small end of the plenum. 

6. If L = 6' or greater, multiple takeoffs are desirable. If 
L = 10' or greater, multiple takeoffs are necessary. 

7. Tank width (W) means the effective width over which 
the hood must pull air to operate (e.g., where the hood 
face is set back from the edge of the tank, this setback 
must be added in measuring tank width). 

If W = 20", slot on one side is suitable. 

If W = 20 - 36 M , slots on both sides are desirable. 

IfW = 36-48", use slots on both sides or along tank 
centerline or use push-pull. A single slot along one side 
should not be used unless all other conditions are 
optimum. 

If W = 48" or greater, local exhaust usually is not 
practical. Consider using push-pull. 

Enclosure can be used for any width tank if process 
will permit. 

It is not practicable to ventilate across the long dimen- 
sion of a tank whose ratio W/L exceeds 2.0. It is 
undesirable to do so when W/L exceeds 1 .0 

8. Liquid level should be 6" to 8" below top of tank with 
parts immersed. 

9. Lateral hood types A, C, D and E (VS-70-01 and -02) 
are preferred. Plenum acts as baffle to room air cur- 
rents. 

10. Provide removable covers on tank if possible. 

1 L Provide duct with cleanouts, drains and corrosion-re- 
sistant coating if necessary. Use flexible connection at 
fan inlet. 

12. Install baffles to reduce cross-drafts. A baffle is a 
vertical plate the same length as the tank and with the 
top of the plate as high as the tank is wide. If the exhaust 
hood is on the side of the tank against a building wall 
or close to it, it is perfectly baffled. 



Specific Operations 



10-95 



13. Replacement air to the tank area must be supplied 
evenly and directed toward the tank from above or in 
front of the tank so that cross-drafts do not occur. 

Flow Rate Calculation for Good Conditions: (No cross- 
crafts, adequate and well-distributed replacement air.) 

1. Establish process class by determining hazard poten- 
tial from Tables 1 0.70. 1 and 10.70.2; information from 
Threshold Limit Values, Solvent Flash Point, Solvent 
Drying Time Tables in Appendices A and B and 
Tables 10.70.5-10.70.8. 

2. Process class can also be established directly from 
Tables 10.70.5-10.70.8 if process parameters are 
known. 

3. From Table 10.70.3, choose minimum control velocity 
according to hazard potential; evolution rate (process 
class); and hood design (see Table 10.70.5 for typical 
processes). 

4. From Table 10.70.4, select the cfm/ft 2 for tank dimen- 
sions and tank location. 

5. Multiply tank area by value obtained from Table 
10.70.4 to calculate required air volume. 

EXAMPLE 

Given: Chrome Plating Tank 6' x 2.5' 

High production decorative chrome 
Free standing in room 
No cross-drafts 

a. Tank Hood. See VS-70-01. Use hood "A" long 6' 
side. Hood acts as baffle 

b. Component — Chromic Acid (Chromium, metal; 
water-soluble CrVI) 

Hazard potential: A (from Table 10.70.1; from Ap- 
pendix A: TLV = 0,05 mg/m 3 ; from Appendix A: 
Flash point = Negligible) 

Rate of Evolution: 1 from Table 10.70.2; from Table 
10.70.6: Gassing rate = high) 
Class: A-l 

Control Velocity = 150 fpm (from Table 10.70.3) 
Minimum Exhaust Rate = 225 cfm/ft 2 (from Table 
10.70.4; Baffled tank, W/L - 0.42) 
Minimum Exhaust Flow Rate = 22 5x15 = 33 75 cfm 

c. Hood Design 

Design slot velocity = 2000 fpm 

Slot area = Q/V = 3375 cfm/2000 fpm = 1 .69 ft 2 

Slot Width = A/L = 1 .69 ft 2 /6 ft = 0.281* = 3.375" 

Example (continued) 

Plenum depth = (2)(slot width) = (2)(3.375) = 6.75" 



TABLE 10.70.1. Determination of Hazard Potential 

HYGIENIC STANDARDS 

Hazard Gas and Vapor Mist Flash Point 

Potential (see Appendix A) (see Appendix A) (see Appendix B) 



A 
B 
C 
D 



0-10 ppm 0-0.1 mg/m 3 

11-100 ppm 0.11-1.0 mg/m 3 

101-500 ppm 1.1-10 mg/m 3 

Over 500 ppm Over 10 mg/m 3 



Under 100 F 
100-200 F 
Over 200 F 



TABLE 10.70.2. Determination of Rate of Gas, Vapor, 
or Mist Evolution 



Liquid 
Rate Temperature (F) 


Degrees 

Below Boiling 

Point (F) 


Relative Evaporation* 

(Time for 100% 

Evaporation) 


Gassing** 


1 Over 200 


0-20 


Fast (0-3 hours) 


High 


2 150-200 


21-50 


Medium (3-12 hours) 


Medium 


3 94-149 


51-100 


Slow (12-50 hours) 


Low 


4 Under 94 


Over 100 


Nil (Over 50 hours) 


Nil 



*Dry Time Relation (see Appendix B). Below 5 — Fast; 5-15 — Medium; 
15-75 — Slow; 70-over — Nil. 

**Rate of gassing depends on rate of chemical or electrochemical action and 
therefore depends on the material treated and the solution used in the tank and 
tends to increase with 1) amount of work in the tank at any one time, 2) strength of 
the solution in the tank, 3) temperature of the solution in the tank, and 4) current 
density applied to the work in electrochemical tanks. 



Duct area = Q/V = 3375 cfm/2500 fpm = 1 .35 ft 2 
Use 16" duct, area =1.396 ft 2 
Final duct velocity - Q/A - 3375/1.396 - 2420 fpm 
Hood SP = Entry loss + Acceleration 

= 1 .78 VP S + 0.25 VP d + 1.0 VP d (see 

Chapter 3) 
= (1.78 x 0.25") + (0.25 x 0.37") + 0.37" 
= 0.45 + 0.09 + 0.37 
HoodSP= 0.91" 

REFERENCES 

10.70.1 Huebener, D.J.; Hughes, R.T.: Development of 
Push-Pull Ventilation. American Industrial Hygiene 
Assoc. Journal, Vol. 46:262-267 (1985). 

10.70.2 Hughes, R.T.: Design Criteria for Plating Tank 
Push-Pull Ventilation. In: Ventilation '85: Proceed- 
ings of the First International Symposium on Venti- 
lation for Contaminant Control. Elsie ver Press, 
Amsterdam, the Netherlands (1986). 

10.70.3 Sciola, V.: Private Communication, Hamilton 
Standard. 



10-96 



Industrial Ventilation 



TABLE 10.70.3. Minimum Control Velocity (FPM) for Undisturbed Locations 



Class (see Tables Enclosing Hood Lateral Exhaust 

10.70.1 & 10.70.2) One Open Side Two Open Sides (see VA-70-01 & 70-02) (Note 1) 



Canopy Hoods 

(see Figure 3-8 & VS-99-03) 

Three Open Sides Four Open Sides 



A-1 and A-2 (Note 2) 

A-3(Note2), B-1,B-2, and C-1 

B-3,C-2,andD-1(Note3) 

A-4 (Note 2) C-3, and 
D-2(Note3) 

B-4, C-4, D-3(Note3),andD-4- 



100 


150 


75 


100 


65 


90 


50 


75 



150 


Do not use 


Do not use 


100 


125 


175 


75 


100 


150 


50 


75 


125 



- Adequate General Room Ventilation Required (see Chapter 2). 



Notes: 1 . Use aspect ratio to determine air volume; see Table 10.70.4 for computation. 

2. Do not use canopy hood for Hazard Potential A processes. 

3. Where complete control of hot water is desired, design as next highest class. 



TABLE 1 0.70.4. Minimum Rate, cfm/ft 2 of Tank Area for Lateral Exhaust 







cfm/ft 2 to maintain required minimum control velocities at following 










tank width (\N\ .. 

— ratios 

tank length {L J 






Required Minimum 












Control Velocity, fpm 












(from Table 10.70.3) 


0.0-0.09 


0.1-0.24 


0.25-0.49 


0.5-0.99 


1.0-2.0 (Note 2) 


Hood against wall or flanged (see Note 1 below and Section 10.70.1. Note 12). See VS-70-01 A and VS-70-02DandE. 


50 


50 


60 


75 


90 


100 


75 


75 


90 


110 


130 


150 


100 


100 


125 


150 


175 


200 


150 


150 


190 


225 


[250] Note 3 


[250] Note 3 


Hood on free standing tank (see 


Note 1). See VS-70-01 B and VS-70-02 F. 








50 


75 


90 


100 


110 


125 


75 


110 


130 


150 


170 


190 


100 


150 


175 


200 


225 


250 


150 


225 


[250] Note 3 


[250] Note 3 


[250] Note 3 


[250] Note 3 



Notes: 1 . Use W/2 as tank width in computing W/L ratio for hood along centerline or two parallel sides of tank. See VS-70-01 B and C and VS-70-02 F. 

2. See Section 10.70.1, Notes 6 and 7. 

3. While bracketed values may not produce 150 fpm control velocity at ail aspect ratios, the 250 cfm/ft 2 is considered adequate for control. 



Specific Operations 10-97 



TABLE 10.70.5. Typical Processes Minimum Control Velocity (fpm) for Undisturbed Locations 



Operation 


Contaminant 


Hazard 


Contaminant 
Evolution 


Lateral Exhaust 

Control Velocity 

(SeeVS-70-01& 

VS-70-02) 


Collector 
Recommended 


Anodizing Aluminum 


Chromic-Sulfuric Acids 


A 


1 


150 


X 


Aluminum Bright Dip 


Nitric + Sulfuric Acids 
Nitric + Phosphoric Acids 


A 
A 


1 
1 


150 
150 


X 
X 


Plating — Chromium 
Copper Strike 


Chromic Acid 
Cyanide Mist 


A 
C 


1 
2 


150 
75 


X 
X 


Metal Cleaning (Boiling) 


Alkaline Mist 


C 


1 


100 


X 


Hot Water (if vent desired) 
Not Boiling 
Boiling 


Water Vapor 


D 
D 


2 
1 


50* 
75* 




Stripping — Copper 
Nickel 


Alkaline-Cyanide Mists 
Nitrogen Oxide Gases 


C 
A 


2 
1 


75 
150 


X 
X 


Pickling — Steel 


Hydrochloric Acid 
Sulfuric Acid 


A 

B 


2 
1 


150 
100 


X 
X 


Salt Solution 
Bonderizing & Parkerizing) 
Not Boiling 
Boiling 


Water Vapor 
Water Vapor 


D 
D 
D 


2 
2 

1 


50* 
50* 
75* 




Salt Baths (Molten) 


Alkaline Mist 


C 


1 


100 


X 


*Where complete control of water \ 


/apor is desired, design as next highest class 











00 



TABLE 10.70.6. Airborne Contaminants Released by Metallic Surfaced Treatment, Etching, Pickling, Acid Dipping and Metal Cleaning Operations 







Component of Bath Which May be 


Physical and Chemical Nature of 




Usual Temp. 


e 


Process Type 


Notes 


Released to Atmosphere (13) 


Major Atmospheric Contaminant 


Class (12) 


Range F 


2. 


Surface Treatment Anodizing Aluminum 




Chromic-Sulfuric Acids 


Chromic Acid Mist 


A-1 


95 


< 


Anodizing Aluminum 




Sulfuric Acid 


Sulfuric Acid Mist 


B-1 


60-80 


3 


Black Magic 




Cone. Sol. Alkaline Oxidizing Agents 


Alkaline Mist, Steam 


C-1 


260-350 


M 


Bonderizing 


1 


Boiling Water 


Steam 


D-2,1 (14, 15) 


140-212 


S3 


Chemical Coloring 




None 


None 


0-4 


70-90 


©" 

3 


Descaling 


2 


Nitric-Sulfuric, Hydrofluoric Acids 


Acid Mist, Hydrogen Fluoride Gas, Steam 


B-2,1 (15) 


70-150 




Ebonoi 




Cone. Sol. Alkaline Oxidizing Agents 


Alkaline Mist, Steam 


C-1 


260-350 




Galvanic-Anodize 


3 


Ammonium Hydroxide 


Ammonia Gas, Steam 


B-3 


140 




Hard-Coating Aluminum 




Chromic-Sulfuric Acids 


Chromic Acid Mist 


A-1 


120-180 




Hard-Coating Aluminum 




Sulfuric Acid 


Sulfuric Acid Mist 


B-1 


120-180 




Jetal 




Cone. Sol. Alkaline Oxidizing Agents 


Alkaline Mist, Steam 


C-1 


260-350 




Magcote 


4 


Sodium Hydroxide 


Alkaline Mist, Steam 


C-3,2 (15) 


105-212 




Magnesium Pre-Dye Dip 




Ammonium Hydroxide-Ammonium Acetate 


Ammonia Gas, Steam 


B-3 


90-180 




Parkerizing 


1 


Boiling Water 


Steam 


D-2,1 (14,15) 


140-212 




Zincete Immersion 


5 


None 


None 


D-4 


70-90 




Etching Aluminum 




Sodium Hydroxide-Soda Ash-Trtsodium 
Phosphate 


Alkaline Mist, Steam 


C-1 


160-180 




Copper 


6 


Hydrochloric Acid 


Hydrogen Chloride Gas 


A-2 


70-90 




Copper 


7 


None 


None 


D-4 


70 




Pickling Aluminum 




Nitric Acid 


Nitrogen Oxide Gases 


A-2 


70-90 




Aluminum 




Chromic, Sulfuric Acids 


Acid Mists 


A-3 


140 




Aluminum 




Sodium Hydroxide 


Alkaline Mist 


C-1 


140 




Cast Iron 




Hydrofluoric-Nitric Acids 


Hydrogen Fluoride-Nitrogen Oxide Gases 


A-2,1 (15) 


70-90 




Copper 




Sulfuric Acid 


Acid Mist, Steam 


B-3,2 (15) 


125-175 




Copper 


8 


None 


None 


D-4 


70-175 




Duralumin 




Sodium Flouride, Sulfuric Acid 


Hydrogen Fluoride Gas, Acid Mist 


A-3 


70 




Inconel 




Nitric, Hydrofluoric Acids 


Nitrogen Oxide, HF Gases, Steam 


A-1 


150-165 




Inconel 




Sulfuric Acid 


Sulfuric Acid Mist, Steam 


B-2 


160-180 




Iron and Steel 




Hydrochloric Acid 


Hydrogen Chloride Gas 


A-2 


70 




Iron and Steel 




Sulfuric Acid 


Sulfuric Acid Mist, Steam 


B-1 


70-175 




Magnesium 




Chromic-Sulfuric, Nitric Acids 


Nitrogen Oxide Gases, Acid Mist, Steam 


A-2 


70-160 




Monel and Nickel 




Hydrochloric Acid 


Hydrogen Chloride Gas, Steam 


A-2 


180 




Monel and Nickel 




Sulfuric Acid 


Sulfuric Acid Mist, Steam 


B-1 


160-190 




Nickel Silver 




Sulfuric Acid 


Acid Mist, Steam 


B-3,2 (15) 


70-140 




Silver 




Sodium Cyanide 


Cyanide Mist, Steam 


C-3 


70-210 




Stainless Steel 


9 


Nitric, Hydrofluoric Acids 


Nitrogen Oxide, Hydrogen Fluoride Gases 


A-2 


125-180 




Stainless Steel 


9,10 


Hydrochloric Acid 


Hydrogen Chloride Gas 


A-2 


130-140 




Stainless Steel 


9,10 


Sulfuric Acid 


Sulfuric Acid Mist, Steam 


B-1 


180 




Stainless Steel Immunization 




Nitric Acid 


Nitrogen Oxide Gases 


A-2 


70-120 




Stainless Steel Passivation 




Nitric Acid 


Nitrogen Oxide Gases 


A-2 


70-120 





TABLE 10.70.6- Airborne Contaminants Released by Metallic Surfaced Treatment, Etching, Pickling, Acid Dipping and Metal Cleaning Operations (con't) 







Component of Bath Which May be 


Physical and Chemical Nature of 




Usual Temp. 


Process 


Type Notes Released to Atmosphere (13) 


Major Atmospheric Contaminant 


Class (12) 


Range F 


Acid Dipping 


Aluminum Bright Dip 


Phosphoric, Nitric Acids 


Nitrogen Oxide Gases 


A-1 


200 




Aluminum Bright Dip 


Nitric, Sulfuric Acids 


Nitrogen Oxide Gases, Acid Mist 


A-2,1 (15) 


70-90 




Cadmium Bright Dip 


None 


None 


D-4 


70 




Copper Bright Dip 


Nitric, Sulfuric Acids 


Nitrogen Oxide Gases, Acid Mist 


A-2,1 (15) 


70-90 




Copper Semi-Bright Dip 


Sulfuric Acid 


Acid Mist 


B-2 


70 




Copper Alloys Bright Dip 


Nitric, Sulfuric Acids 


Nitrogen Oxide Gases, Acid Mist 


A-2,1 (15) 


70-90 




Copper Matte Dip 


Nitric, Sulfuric Acids 


Nitrogen Oxide Gases, Acid Mist 


A-2,1 (15) 


70-90 




Magnesium Dip 


Chromic Acid 


Acid Mist, Steam 


A-2 


190-212 




Magnesium Dip 


Nitric, Sulfuric Acids 


Nitrogen Oxide Gases, Acid Mist 


A-2,1 (15) 


70-90 




Monel Dip 


Nitric, Sulfuric Acids 


Nitrogen Oxide Gases, Acid Mist 


A-2,1 (15) 


70-90 




Nickel and Nickel Alloys Dip 


Nitric, Sulfuric Acids 


Nitrogen Oxide Gases, Acid Mist 


A-2,1 (15) 


70-90 




Silver Dip 


Nitric Acid 


Nitrogen Oxide Gases 


A-1 


70-90 




Silver Dip 


Sulfuric Acid 


Sulfuric Acid Mist 


B-2 


70-90 




Zinc and Zinc Alloys Dip 


Chromic, Hydrochloric Acids 


Hydrogen Chloride Gas (If HCI 
attacks Zn) 


A-4,3 (15) 


70-90 


Metal Cleaning 


Alkaline Cleaning 11 


Alkaline Sodium Salts 


Alkaline Mist, Steam 


C-2,1 (15) 


160-210 




Degreasing 


Trichloroethylene-Perchloroethylene 


Trichloroethylene-Perchloroethylene 

Vapors 

Petroleum-Coal Tar Vapors 


B (16) 


188-250 




Emulsion Cleaning 


Petroleum-Coal Tar Solvents 


B-3,2 (15) 


70-140 










(17) 


70-140 




Emulsion Cleaning 


Chlorinated Hydrocarbons 


Chlorinated Hydrocarbon Vapors 


(17) 


70-140 



Notes: 1 Also Aluminum Seal, Magnesium Seal, Magnesium Dye Set, Dyeing Anodized 

Magnesium, Magnesium Alkaline Dichromate Soak, Coloring Anodized Aluminum. 

2 Stainless Steel Before Eiectropoiishing. 

3 On Magnesium. 

4 Also Manodyz, Dow-12. 

5 On Aluminum. 

6 Dull Finish. 

7 Ferric Chloride Bath. 

8 Sodium Dichromate, Sulfuric Acid Bath and Ferrous Sulfate, Sulfuric Acid Bath. 

9 Scale Removal, 
10 Scale Loosening. 



11 Soak and Electrocleaning. 

12 Class as described in Table 10.70.3 based on hazard potential (Table 10.70.1) 

and rate of evolution (Table 10.70.2) for usual operating conditions. Higher temperatures, agitation 
or other conditions may result in a higher rate of evolution. 

13 Hydrogen gas also released by many of these operations. 

14 Rate where essentially complete control of steam is required, Otherwise, adequate dilution 
ventilation may be sufficient. 

15 The higher rate is associated with the higher value in the temperature range. 

16 For vapor degreasers, rate is determined by operating procedure. See VS-70-20. 

17 Class of operation is determined by nature of the hydrocarbon. Refer to Appendix A. 



c/) 



O 

-13 

63 

O* 
3 






© 

© 



TABLE 10.70.7. Airborne Contaminants Released by Electropofishing, Electroplating and Eiectroless Plating Operations 









Component of Bath Which May be 


Physical and Chemical Nature of 




Usual Temp. 


Process 


Type 


Nol 


es Released to Atmosphere (19) 


Major Atmospheric Contaminant 


Class (18) 


Range F 


Electropofishing 


Aluminum 




Sulfuric, Hydrofluoric Acids 


Acid Mist, Hydrogen Flouride Gas, Steam 


A-2 


140-200 




Brass, Bronze 




Phosphoric Acid 


Acid Mist 


B-3 


68 




Copper 




Phosphoric Acid 


Acid Mist 


B-3 


68 




Iron 




Sulfuric, Hydrochloric, Perchforic Acids 


Acid Mist, Hydrogen Chloride Gas, Steam 


A-2 


68-175 




Monel 




Sulfuric Acid 


Acid Mist, Steam 


B-2 


86-160 




Nickel 




Sulfuric Acid 


Acid Mist, Steam 


B-2 


86-160 




Stainless Steei 




Sulfuric, Hydrofluoric, Chromic Acids 


Acid Mist, Hydrogen Flouride Gas, Steam 


A-2,1 (20) 


70-300 




Steel 




Sulfuric, Hydrochloric, Perchloric Acids 


Acid Mist, Hydrogen Chloride Gas, Steam 


A-2 


68-175 



Strike Solutions 


Copper 

Silver 

Wood's 


Nickel 




Cyanide Salts 
Cyanide Salts 
Nickel Chloride, Hydrochloric Acid 


Cyanide Mist 
Cyanide Mists 
Hydrogen Chloride Gas, 


Chloride Mist 


C-2 
C-2 
A-2 


70-90 
70-90 
70-90 


Eiectroless Plating 


Copper 
Nickel 




2 


Formaldehyde 
Ammonium Hydroxide 


Formaldehyde Gas 
Ammonia Gas 




A-1 
B-1 


75 
190 


Electroplating Alkaline 


Platium 

Tin 

Zinc 




3 


Ammonium Phosphate, Ammonia Gas 

Sodium Stannate 

None 


Ammonia Gas 

Tin Salt Mist, Steam 

None 




B-2 
C-3 
D-4 


158-203 
140-170 
170-180 



Electroplating Fiuoborate Cadmium 




Fiuoborate Salts 


Fiuoborate Mist, Steam 


C-3,2 (20) 


70-170 


Copper 




Copper Fiuoborate 


Fiuoborate Mist, Steam 


C-3,2 (20) 


70-170 


Indium 




Fiuoborate Salts 


Fiuoborate Mist, Steam 


C-3,2 (20) 


70-170 


Lead 




Lead Fluoborate-Fluoboric Acid 


Fiuoborate Mist, Hydrogen Fluoride Gas 


A-3 


70-90 


Lead-Tin Alloy 




Lead Fluoborate-Fluoboric Acid 


Fiuoborate Mist 


C-3,2 (20) 


70-100 


Nickel 




Nickel Fiuoborate 


Fiuoborate Mist 


C-3,2 (20) 


100-170 


Tin 




Stannous Fiuoborate, Fluoboric Acid 


Fiuoborate Mist 


C-3,2 (20) 


70-100 


Zinc 




Fiuoborate Salts 


Fiuoborate Mist, Steam 


C-3,2 (20) 


70-170 


Electroplating Cyanide Brass, Bronze 


4,5 


Cyanide Salts, Ammonium Hydroxide 


Cyanide Mist, Ammonia Gas 


B-4,3 (20) 


60-100 


Bright Zinc 


5 


Cyanide Salts, Sodium Hydroxide 


Cyanide, Akaline Mists 


C-3 


70-120 


Cadmium 


5 


None 


None 


D-4 


70-100 


Copper 


5,6 


None 


None 


D-4 


70-160 


Copper 


5,7 


Cyanide Salts, Sodium Hydroxide 


Cyanide, Alkaline Mists, Steam 


C-2 


110-160 


Indium 


5 


Cyanide Salts, Sodium Hydroxide 


Cyanide, Alkaline Mists 


C-3 


70-120 


Silver 


5 


None 


None 


D-4 


72-120 


Tin-Zinc Alloy 


5 


Cyanide Salts, Potassium Hydroxide 


Cyanide, Alkaline Mists, Steam 


C-3,2 (20) 


120-140 


White Alloy 


5,8 


Cyanide Salts, Sodium Stannate 


Cyanide, Alkaline Mists 


C-3 


120-150 


Zinc 


5,9 


Cyanide Salts, Sodium Hydroxide 


Cyanide, Alkaline Mists 


C-3,2 (7) 


70-120 









TABLE 10.70.7. Airborne Contaminants Released by Electropolishing, Electroplating and Electroless Plating Operations (con't) 



Process 



Type 



Component of Bath Which May be 
Notes Released to Atmosphere (19) 



Physical and Chemical Nature of Usual Temp. 

Major Atmospheric Contaminant Class (18) Range F 



Electroplating Acid 



Chromium 




Chromic Acid 




Chromic Acid Mists 


Copper 


10 


Copper Sulfate, Sulfuric Acid 


Sulfuric Acid Mist 


Indium 


12 


None 




None 


Indium 


13,14 


Sulfamic Acid, Sulfamate Salts 


Sulfamate Mist 


Iron 




Chloride Salts, Hydrochloric Acid 


Hydrochloric Acid Mist, Steam 


Iron 


12 


None 




None 


Nickel 


3 


Ammonium Fluoride, 


Hydrofluoric Acid 


Hydrofluoric Acid Mist 


Nickel and Black 


12,15 


None 




None 


Nickel 










Nickel 


9,12 


Nickel Sulfate 




Nickel Sulfate Mist 


Nickel 


13,14 


Nickel Sulfamate 




Sulfamate Mist 


Palladium 


15 


None 




None 


Rhodium 


12,17 


None 




None 


Tin 




Tin Halide 




Halide Mist 


Tin 


12 


None 




None 


Zinc 




Zinc Chloride 




Zinc Chloride Mist 


Zinc 


12 


None 




None 



A-1 

B-4,3 (20,21) 

D-4 

C-3 

A-2 

D-4 

A-3 

C-4 (22) 

B-2 
C-3 
D-4 
D-4 
C-2 
D-4 
B-3 
D-4 



90-140 
75-120 
70-120 
70-90 

190-210 
70-120 

102 
70-150 

70-90 

75-160 

70-120 

70-120 

70-90 

70-120 

75-120 

70-120 



Notes: 1 Arsine may be produced due to the presence of arsenic in the metal or polishing bath. 

2 Alkaline Bath. 

3 On Magnesium, 

4 Also Copper-Cadmium Bronze. 

5 HCN gas may be evolved due to the acidic action of CO2 in the air at the surface of the bath. 

6 Conventional Cyanide Bath. 

7 Except Conventional Cyanide Bath. 

8 Albaloy, Spekwhite, Bonwhite (Alloys of Copper, Tin, Zinc). 

9 Using Insoluble Anodes. 

10 Over 90 R 

11 Mild Organic Acid Bath. 

12 Sulfate Bath. 

13 Sulfamate Bath. 



14 AirAgitated. 

15 Chloride Bath. 

16 Nitrite Bath. 

17 Phosphate Bath. 

18 Class as described in Table 10.70.3 based on hazard potential (Table 
10.70.1) and rate of evolution (Table 10.70.2) for usual operating conditions. 
Higher temperatures, agitation, high current density or other conditions may result 
in a higher rate of evolution. 

19 Hydrogen gas also released by many of these operations. 

20 The higher rate is associated with the higher value in the temperature range. 

21 Baths operated at a temperature of over 140 F with a current density of over 45 
amps/fr and with air agitation will have a higher rate of evolution. 

22 Local exhaust ventilation may be desired to control steam and water vapor. 



o 

fS 

s 

rs 

o 

ft 



© 
1 

© 



o 



TABLE 10.70.3. Airborne Contaminants Released by Stripping Operations 



Coating to be Stripped 


Base Metal 
(Footnote) 


Component of Batch Which May be 
Released to Atmosphere (f) 


Physical and Chemical Nature of Major 
Atmospheric Contaminant 


Class (e) 


Usual Temp. 
Range F 


Anodized Coatings 


1,7 




Chromic Acid 


Acid Mist, Steam 


A-2 




120-200 


Black Oxide Coatings 


14 




Hydrochloric Acid 


Hydrogen Chloride Gas 


A-3,2 


(g) 


70-125 


Brass and Bronze 


8,14 


(a) 


Sodium Hydroxide, Sodium Cyanide 


Alkaline, Cyanide Mists 


C-3,2 


(g) 


70-90 


Cadmium 


8,14 

2,4,14 


(a) 


Sodium Hydroxide, Sodium Cyanide 
Hydrochloric Acid 


Alkaline, Cyanide Mists 

Acid Mist, Hydrogen Chloride Gas 


C-3,2 
A-3,2 


(g) 
(g) 


70-90 
70-90 


Chromium 


7,8,14 

2,4,8,14 

2,4,8,18 


(a) 
(a) 


Sodium Hydroxide 
Hydrochloric Acid 
Sulfuric Acid 


Alkaline Mist, Steam 
Hydrogen Chloride Gas 
Acid Mist 


C-3 
A-2 
B-2 




70-150 
70-125 
70-90 



Copper 



8,14 




Sodium Hydroxide, Sodium Cyanide 


Alkaline, Cyanide Mists 


7,12,14 


(b) 


None 


None 


14 


(a) 


Alkaline Cyanide 


Cyanide Mist 


1 




Nitric Acid 


Nitrogen Oxide Gases 


18 


(a) 


Sodium Hydroxide-Sodium Sulfide 


Alkaline Mist, Steam 



C-3,2 


(g) 


70-90 


D-4 




70-90 


C-3,2 


(g) 


70-160 


A-1 




70-120 


C-2 




185-195 



3 

a 



< 



© 

3 



Gold 



4,5,6,8,9,14 (a) Sodium Hydroxide, Sodium Cyanide 

4,5,18 (a) Sulfuric Acid 



Alkaline, Cyanide Mists 
Acid Mist 



C-3,2 
B-3,2 



(g) 
(g) 



70-90 
70-100 



Lead 



13 (c) Acetic Acid, Hydrogen Peroxide 

14 (a),(c) Sodium Hydroxide 



Oxygen Mist 
Alakaline Mist, Steam 



D-3 
C-3,2 



(g) 



70-90 
70-140 



Nickel 



Silver 



2,4 Sulfuric, Nitric Acids 

2,4 (a) Hydrochloric Acid 

2,4,14 (a) Sulfuric Acid 

7 Hydrofluoric Acid 

14 Fuming Nitric Acid 



Nitrogen Oxide Gases 
Hydrogen Chloride Gas 
Acid Mist 

Hydrogen Fluoride Gas 
Nitrogen Oxide Gases 





1,18,19 


(a),(d) 
(a) 


Hot Water 
Sulfuric Acid 


Steam 

Acid Mist, Steam 


D-2 
B-3,2 


(h) 

(g) 


200 
70-150 


Phosphate Coatings 


15 
16 




Chromic Acid 
Ammonium Hydroxide 


Acid Mist, Steam 
Ammonia Gas 


A-3 
B-3,2 


(g) 


165 
70-% 


Rhodium 


10 




Sulfuric, Hydrochloric Acids 


Acid Mist, Hydrogen Chloride Gas 


A-3,2 


(g) 


70-100 



1 




2,11 




8,14 


(a) 


17 


(a) 


2,3,4 


(a\ 


2,14 


W 


14 


(a) 



Nitric Acid 

Sulfuric, Nitric Acids 

Sodium Hydroxide, Sodium Cyanide 

Sodium Cyanide 



Nitrogen Oxide Gases 
Nitrogen Oxide Gases, Steam 
Alkaline, Cyanide Mists 
Cyanide Mist 



A-1 




70-90 


A-1 




180 


C-3 




70-90 


C-3 




70-90 


B-4,3 


(g) 


70-90 


C-3 




70-90 


A-3,2 


(g) 


70-90 


C-2 




7(^200 



Tin 



Ferric Chloride, Copper Sulfate Acetic Acid 
Sodium Hydroxide 
Hydrochloric Acid 
Sodium Hydroxide 



Acid Mist 
Alkaline Mist 
Hydrogen Chloride Gas 
Alkaline Mist, Steam 



TABLE 10.70.8. Airborne Contaminants Released by Stripping Operations (con't) 








Base Metal Component of Batch Which May be 
Coating to be Stripped (Footnote) Released to Atmosphere (f) 


Physical and Chemical Nature of Major 
Atmospheric Contaminant 


Class (e) 


Usual Temp. 
Range F 


Zinc 1 Nitric Acid 

8,14 Sodium Hydroxide, Sodium 


Nitrogen Oxide Gases 
Alkaline, Cyanide Mists 


A-1 
C-3 


70-90 
70-90 



8. 


Nickel 


9. 


Nickel Alloys 


10. 


Nickel Plated Brass 


11. 


Nickel Silver 


12. 


Non-Ferrous Metals 


13. 


Silver 



Base Metal: 1. Aluminum 

2. Brass 

3. Bronze 

4. Copper 

5. Copper Alloys 

6. Ferrous Metals 

7. Magnesium 

Notes: (a) Electrolytic Process. 

(b) Refers only to steel (14) when Chromic, Sulfuric Acids Bath is used. 

(c) Also Lead Alloys. 

(d) Sodium Nitrate Bath. 

(e) Class as described in Table 10.70. 3 based on hazard potential (Tablel 0.70.1) and rate of evolution 
(Table 10.70.2) for usual operating conditions. Higher temperatures, agitation or other conditions 
may result in a higher rate of evolution. 



14. 


Steel 


15. 


Steel (Manganese Type Coatings) 


16. 


Steel (Zinc Type Coatings) 


17. 


White Metal 


18. 


Zinc 


19. 


Zinc Base Die Castings 



(f) Hydrogen gas also released by some of these operations. 

(g) The higher rate is associated with the higher value in the temperature range. 

(h) Rate where essentially complete control of steam is required. Otherwise, adequate dilution 
ventilation may be sufficient. 



in 

o 






o 



10-104 Industrial Ventilation 



1 2"min 




Plenum acts as flange- 
Slot sized for 2000 fpm 
S 




A. UPWARD PLENUM 



Partial covers advisable if — 
possible - on any type tank 



-| |— 2S 



J 




! 




. . i 


/ I 


[___\ 


L 


Tankj 


:::=::::: 



















1 2 min 




B. DOWNWARD PLENUM 



A — 



T - 




L 
T 



2S 



A i 

C. CENTRAL SLOT 




Section A — A 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



OPEN SURFACE TANKS 



DATE 12-90 | FIGURE VS-70-01 



Specific Operations 10-105 




A 

off vapors L r J 



Work gives 

after removal from 

tank. 




To suit work 



Section 



D. PICKLING TANK 




Slot 



Extend over tank as far 

as possible 



~xr} 




12" min 



E. LATERAL 




Max. plenum velocity ~ 1/2 slot velocity 
p 12" min 



Slot velocity 2000 fpm 
Sloped plenum desirable 




Inside radius desirable if space permits 
— Slot 



— — 2 S min 

A 



F. END TAKE-OFF 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



OPEN SURFACE TANKS 



DATE 



12-90 ™ure VS-70-02 



p 



10-106 



Industrial Ventilation 



*n 


^/ 




o 




_____ Hood Opening _ 






®- 


lu 


ru 1 n 




Vj 




o» 



— Hood 



Sea 




<TOGiO 



H 



r© 



OOOOOOOOO 



r 



© 




Liquid Temperature (t) 



Push nozzle manifold (T) -Circular, rectangular or square. Manifold cross-sectional 

area should be at least 2.5 times the total nozzle flow area. 

Push nozzle angle @ - 0' to 20' down. 

Nozzle openings @ - 1/8" to 1/4" slot or 5/32" to 1/4" dia. holes with 3 to 8 

dja. spacing. Outer holes or slot ends (4) must be 1/2" to 
1" inside tank inner edges. 

Exhaust opening (s) - Size to achieve 2000 fpm slot velocity. Outer edges of 

opening(7) must extend to edge of tank including flanges. 

Liquid surface (jb) — Tank freeboard must not exceed 8 M with parts removed. 
Push nozzle supply Q. = 243 VX 

where Q, = push nozzle supply, cfm/ft manifold length 

A, = total nozzle opening per foot of manifold length 

Total push supply Q s = Q, x L cfm 

2 

Exhaust flow Q^ - 75 cfm/ft tank surface area for t < 1 50 F 

Q = (0.4 T -h 15) cfm/ft 2 tank surface area for t > 1 50 F. 

Tank surface area = L (length of tank) x W (width of tank) 

Design Procedure: Select nozzle opening within above limits and calculate push supply 
and exhaust air flow. See VS-70-11 and VS-70-12. 

Reference 10.70.1, 10.70.2, & 10.70.3 



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PUSH-PULL HOOD DESIGN 
DATA FOR WIDTHS UP TO 10' 



DATE 



4-94 



FIGURE VS ^ 70 _ 1Q 



Specific Operations 10-107 



In push-pull ventilation, a nozzle pushes a jet of air across the vessel surface into an exhaust 
hood. Effectiveness of a push jet is a function of its momentum which can be related to the 
prod uct of the nozzle supply air flow (Qj) and the nozzle exit velocity (Vj). For a jet used for plating 
tanks or other open surface vessels, a push supply flow can be determined from: 

where: Qj = push nozzle supply, cfm per foot of push nozzle plenum length 

Aj = nozzle exit area, ft2/perfoot of push nozzle plenum length 

Using this approach, a push nozzle design is first selected and the nozzle exit area (Aj) 
determined. 

The push nozzle manifold may be round, rectangular or square in cross-section. The push nozzle 
may be a 1/8" to 1/4" horizontal slot or 5/32" to 1/4" diameter drilled holes on 3 to 8 diameter 
spacing. 

It is important that the air flow from the nozzle be evenly distributed along the length of the supply 
plenum. To achieve this, the total nozzle exit area should not exceed 40% of the plenum 
cross-sectional area. Multiple supply plenum inlets should be used where practical. 

The push nozzle manifold should be located as near the vessel edge as possible to minimize 
the height above the liquid surface. The manifold should be adjustable to optimize the push jet 
angle. The manifold axis can be angled down a maximum of 20o to permit the jet to clear 
obstructions and to maintain the jet at the vessel surface. It is essential any opening between 
the manifold and tank be sealed. 

An exhaust flow of 75 cfm/ft2 of vessel surface area should be used for tank liquid temperatures 
(t) of 1 50 F or lower. For tank liquid temperatures greater than 1 50 F use an exhaust flow of (0.4t 
+ 15) cfm/ft2. These flow rates are independent of the "class" used in determining exhaust flow 
for side draft hoods. "Control velocity" is achieved by the push jet blowing over the tank and will 
be considerably higher than that which can be achieved by a side draft hood. The purpose of 
the exhaust hood is to capture and remove the jet — not to provide capture velocity. A flanged 
hood design is to be used wherever practical. The exhaust hood should be located at the vessel 
edge so as not to leave a gap between the hood and the vessel. 

Design and location of an open surface vessel encompasses a number of variables. In some 
cases vessel shape, room location, cross-drafts, etc., may create conditions requiring adjustment 
of the push and/or pull flow rates in order to achieve effective control. Cross-draft velocities over 
75 fpm, very wide vessels (eight feet or more), or very large or flat surface parts may require 
increased push and/or pull flows. To account for the effects of these variables, a flow adjustment 
of ±20% should be designed into the push and +20% into the pull flow system. Wherever practical, 
construction and evaluation of a pilot system is recommended. Once designed and installed, 
push-pull systems can be initially evaluated by use of a visual tracer technique and appropriate 
flow adjustments can be made as required. 

The exhaust hood opening should be sized to assure even flow distribution across the opening. 
This can be achieved by sizing the slot for 2000 fpm slot velocity. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



VSIGN DATA. 
PUSH-PULL HOOD 



DATE 



4-94 



FIGURE 



70 ././ 



10-108 Industrial Ventilation 



en 



OT 



2.00 



CO 
CO 

£ LOO 

a 0.80 
< 

^ 0.60 

Q 
_) 
O 

u. 

;E 

< 0.40 



M 
M 
O 



</> 0.20 

CL 




0.002 



0.020 



0.004 0.006 0.010 

NOZZLE FLOW AREA (Aj ) » ft 2 /ft manifold length 



PUSH NOZZLE SUPPLY = 243 /Aj cfm/ft of length for nozzles with 1/8 to 1/4 

inch wide slots or 5/32 to 1 /4 inch diameter holes on 
3 to 8 diameter spacing. 

For holes Aj (ft 2 /ft)= 0.065 x hole diameter (in)/hote spacing (no. of diameters (in).) 



For slot Aj(ft 2 /ft)= S ' Qt W ^ th ° n) (See VS-70-10) 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DATE 



PUSH NOZZLE 
MANIFOLD PRESSURE 

4-94 | FIGURE VS-70-12 



Specific Operations 1 0-1 09 



No sloi near 
take — off 




Inside radius 
desirable 



Maximum plenum 
velocity = 500 fpm 



Cover when not in use 



n 



-Ir 



^r 



J \r 




Section A— A 



Q = 50 LW 
Slot velocity 



2000 fpm 
0.25 VP d 
Duct velocity ~ 2000 minimum 



h e = 1.78 VP S + 0.25 VP d 



Also provide: 1 . Separate flue for combustion products 

2. For cleaning operation, appropriate respiratory protection is necessary. 

3. For pit units, the pit should be mechanically ventilated. 

4. For further safeguards, see VS — 501.1 

NOTE: Provide down draft grille for p^rts that cannot be 
removed dry; Q = 50 cfm/ft grille area. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SOLVENT DEGREASING TANKS 



date 12-90 | FIGURE V3-70-20 



10-110 



Industrial Ventilation 



Solvent vapor degreasing refers to boiling liquid cleaning systems utilizing trichloroethylene, 
perchlorethylene, methylene chloride, freons® or other halogenated hydrocarbons. Cleaning action 
is accomplished by the condensation of the solvent vapors in contact with the work surface 
producing a continuous liquid rinsing action. Cleaning ceases when the temperature of the work 
reaches the temperature of the surrounding solvent vapors. Since halogenated hydrocarbons are 
somewhat similar in their physical, chemical and toxic characteristics, the following safeguards 
should be provided to prevent the creation of a health or life hazard: 

1 . Vapor degreasing tanks should be equipped with a condenser or vapor level thermostat to keep 
the vapor level below the top edge of the tank by a distance equal to one-half the tank width or 
36 inches, whichever is shorter. 

2. Where water type condensers are used, inlet water temperature should not exceed 80 F and 
the outlet temperature should not exceed 1 1 F. For some solvents, lower water temperatures 
may be required. 

3. Degreasers should be equipped with a boiling liquid thermostat to regulate the rate of vapor 
generation, and with a safety control at an appropriate height above the vapor line to prevent 
the escape of solvent in case of a malfunction. 

4. Tanks or machines of more than 4 square feet of vapor area should be equipped with suitable 
gasketed cleanout or sludge doors, located near the bottom, to facilitate cleaning. 

5. Work should be placed in and removed slowly from the degreaser, at a rate no greater than 1 1 
fpm, to prevent sudden disturbances of the vapor level. 

6. CARE MUST BE TAKEN TO PREVENT DIRECT SOLVENT CARRYOUT DUE TO THE 
SHAPE OF THE PART. 

7. Maximum rated workloads as determined by the rate of heat transfer (surface area and specific 
heat) should not be exceeded. 

8. Special precautions should be taken where natural gas or other open flames are used to heat 
the solvent to prevent vapors* from entering the combustion air supply. 

9. Heating elements should be designed and maintained so that their surface temperature will not 
cause the solvent or mixture to breakdown* or produce excessive vapors. 

10. Degreasers should be located in such a manner that vapors* will not reach or be drawn into 
atmospheres used for gas or electric arc welding, high temperature heat treating, combustion 
air or open electric motors. 

11. Whenever spray or other mechanical means are used to disperse solvent liquids, sufficient 
enclosure or baffling should be provided to prevent direct release of airborne vapor above the 
top of the tank. 

12. An emergency quick-drenching facility should be located in near proximity to the degreaser for 
use in the event of accidental eye contact with the degreasing liquid. 

*Electric arcs, open flames and hot surfaces will thermally decompose halogenated hydrocarbons to toxic and corrosive 
substances (such as hydrochloric and/or hydrofluoric acid). Under some circumstances, phosgene may be formed. 



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SOLVENT VAPOR DEGREASING 



DATE l~9l 



FIG 



3URE VS- 70-21 



Specific Operations 



10-111 



10.75 PAINTING OPERATIONS 

Application of industrial paints and coatings usually is 
accomplished by one of three techniques: air-atom ization, 
electrostatics, or airless methods. Potential health hazards 
exist from exposure to solid and liquid aerosols as well as to 
solvent vapors. In addition to the airborne exposures, hazards 
include the use of flammable and combustible liquids and the 
accumulation of flammable paint residues. Fire safety and 
proper electrical wiring are important concerns in most paint 
applications/ 10 - 75 - 1 * 10 - 75 ^ 

Control of airborne pollutants by ventilation may be ac- 
complished through the use of spray booths such as shown in 
VS-75-01 and VS-75-04. The typical booth is a partial enclo- 
sure of sheet metal construction with openings for conveying 
the work piece into and out of the booth. Several factors are 
important in the performance of these booths. Booth depth is 
critical; spray rebound may escape from shallow booths and 
increase exposures. The size of the booth is governed princi- 
pally by the size of the object being coated. Sufficient space 
must be provided to permit air flow on all sides of the object, 
to provide room to work, and to enable the air to enter the 
booth in a smooth, controlled manner without excessive wrap- 
around. In some cases, downdraft booths may be employed 
when large objects are painted. 

In air- atom ization applications, the most common spray 
technique, it is important to use the minimum air pressure 
needed to accomplish the task. Excess air pressure results in 
increased dispersion of the paint and overspray as well as poor 
work quality. 

Airless application results in aerosols with fewer particles 
in the respirable range. One study (l ° 75 3) suggests that approxi- 
mately 20% of the particles in air-atom ization applications are 
less than 12 microns while airless methods produce aerosols 
with only 2% less than this value. The larger aerosols pro- 
duced by the airless technique will deposit more efficiently 
on the work piece, due to impaction, than the smaller particles 
produced by the compressed air method. 

Electrostatic applications result in more efficient deposi- 
tion of paint aerosols due to electrostatic forces. As a result, 
ventilation air flow requirements for control of electrostatic 
applications tend to be lower than for compressed air methods. 

Many spray booths are equipped with disposable particu- 



late filters which become loaded over time and result in 
increased pressure loss. This loss eventually can reduce air 
flow to unacceptable levels and, hence, system performance 
must be monitored. Water wash systems are available for 
cleaning particulate matter from the exhausted air but do little 
for solvent vapors. Fan selection is an important component 
of a spray booth installation. Often the fan is an integral part 
of the system when purchased and may be installed in a 
different configuration than originally designed. This can 
result in reduced air flow, particularly if additional system 
resistance is encountered in the actual installation. 0075 4) 

Work practices remain an important aspect of controlling 
exposure to paint aerosols and solvent vapors. The worker 
should not stand downstream of the object being sprayed. A 
turntable can help to facilitate easy access to all sides of the 
object without the worker having to move. Extension arms on 
spray guns should be employed for hard-to-reach cavities. 
Proper location of the booth with respect to replacement air 
and obstructions is essential. Locating booths in corners or 
near disruptive air currents can defeat the protection of these 
hoods. Poor location of the booth may result in turbulent air 
flow which may reduce the protection provided by the booth. 

Respiratory protection may be required in applications 
using toxic materials. This includes heavy metal pigments and 
organics such as isocyanates in urethane paints and amines in 
epoxy paints. 

REFERENCES 

10.75.1 National Fire Protection Association: Flammable 
and Combustible Liquids Code. No. 30. NFPA. Bos- 
ton, MA (1990). 

10.75.2 National Fire Protection Association: National Elec- 
tric Code. No. 70. NFPA. Boston, MA, (1990). 

10.75.3 National Institute for Occupational Safety and 
Health: An Evaluation of Engineering Control Tech- 
nology for Spray Painting. DHHS (NIOSH) Pub. 
No. 81-121; NTIS Pub. No. PB-82- 1 62-264. Na- 
tional Technical Information Service, Springfield, 
VA(1981). 

10.75.4 Burgess, W.A.; Ellenbecker, M.J.; Treitman, R.D.: 
Ventilation for Control of the Work Environment. 
John Wiley and Sons, NY (1989). 



10-112 



Industrial Ventilation 




.r 


--^■-^ 


D 






"L 








E 










C 












v// 




v//, 




//// 
Y//, 



^O 




Y/W// 


VA 


Y 


m 


V//, 



1 . Split baffle or filters 
B = 0.75 D 
Baffle area = 0.75 WH 
For filter area, see note 2 



2 . Angular baffle 
E = D + 6" 
Baffle area = 0.40 WH 
For filter area, see note 2 



Air spray paint design data. 

Any combination of duct connections and baffles may be used. Large, deep booths 
do not require baffles. Consult manufacturers for water-curtain designs. Use explosion 
proof fixtures and a non-sparking fan. Electrostatic spray booth requires automatic 
high-voltage disconnects for conveyor failure, fan failure or grounding. 



Walk-in booth 
W — work size + 6' 
H ~ work size + 3' (minimum — 7') 
C = work size + 6' 

Q = 1 00 cfm/ft booth cross section 

"Y- for 



Operator outside booth 



W 
H 
C 
Q 



May be 75 cfm/ft^ lor very 

large, deep booth. Operator may 

require a NIOSH cerified respirator. 

h e = 1.78 VP S + 0.50 VP^ (baffles) 

h e = Dirty filter resistance + 0.50 VPd (filters) 

Duct velocity = 2000 fpm 



work size + 2' 

work size -I- 2' 

0.75 x larger front dimension 

100 - 150 cfm/ft 2 of open 

area, including conveyor 

ope n i n g s . 



Airless spray paint design 
Q = 60 cfm/ft 2 booth cross 
section, walk — in booth 

Q = 60 - 100 cfm/ft 2 of 

total open area, operator 
outside of booth 



Notes : 1. Baffle arrangements shown are for 
air distribution only. 

2. Paint arresting filters usually selected 
for 100 - 500 fpm, consult manu- 
facturer for specific details. 

3. For construction an safety, consult 
NFPA - See Reference 10.75.1. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



LARGE PAINT BOOTH 



DATE 



1 - 91 



FIGURE 



VS-75-01 



Specific Operations 10-113 





_B L 


C 

















cr^ 


D 






A 




B 

-d — — &»~ 




C 


— g^ 



















n 


J2 


c 
r ^ 









1_ A 



45" m i n 





Solid baffle 
B = 0.75 D 
Baffle area 



2. Angular baffle 3. Splif baffle or fillers 

B = D + 6" B = D -I- 6" 

0.60 WH Baffle area = 0.60 WH Baffle area = 0.75 WH 



Air spray design data 

Any combination of branch ducts and baffles may be used. 

W = work size + 1 2" 

H = work size + 1 2'' 

C — 0.75 VV or H, whichever is larger 

Q = 200 cfm/ft 2 (200 WH) - for face area up to 4 ft 2 

~ 150 cfm/ft 2 ™ for face area over 4 ft 2 
h e = 1.78 VP S + 0.25 VP d (baffles) 

= dirty filter resistance + 0.25 VP^ (fiiters) 
Duct velocity = 2000 fpm 

Airless spray paint design data 

Q = 125 cfm/ft 2 (125 WH) - for face area up to 4 ft 2 
= 100 cfm/ft 2 ~~ for face area over 4 ft 2 

Notes: 1 . Baffle arrangements shown are for air distribution only. 

2. Paint arresting filters usually selected for 100 — 500 fpm, consult 
manufacturer for specific details. 

3. Por construction and safety, consult NPPA (Reference 10.75.1). 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SMALL PAINT BOOTH 



DATE 



1 91 



FIGURE 



YS- 75 - 02 



10-114 



Industrial Ventilation 




Use window or / 

opening at opposite / V 
end of inlet. / ^ 



v 



Q — 50 cfm/ft 2 of cross-sectional trailer area 

h e = 0.25 VP d 

Minimum duel" velocity ~ 2000 fpm 



Notes; 

1. Paint arresting filters usually selected for 100-500 fpm, 
consult manufacturer for specific details. 

2. For construction and safety, consult NFPA, Reference 10,75.1. 

3. Operator must wear an appropriate, NI0SH certified respirator. 



A ME R'I'C AN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIEN] 



S'l 



TRAILER INTERIOR 
SPRA Y PAINTING 



DATE 



l-Ql [FIGURE VS-75-03 



Specific Operations 



10-115 



Door stop in floor 




Paint arresting filters 
in door 



W 



Air filters in door desirable 

PLAN VIEW 




Door stop in floor 



Use vertical discharge - 




ELEVATION 



00 cfm/ft 2 of cross sectional arec 



Latch to close doors 
tightly 



(When W'x H is greater than 150 ft 2 , Q = 50 cfm/ft 2 ) 
h e ~ 0.50 VPcj plus resistance of each filter bank when dirty 
Minimum duct velocity ~ 2000 fpm 



Notes: 

1 . Exhaust fan interlock with make — up air supply and compressed 
air to spray gun is desirable. 

2. Paint arresting filters usually selected for 100-500 fpm. 
Consult manufacturer for specific details. 

3. Por construction and safety, consult NPPA, Reference 10.75.1. 

4. Por airless spray painting use 

Q = 60 cfm/ft of cross section area. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



LARGE DRIVE THROUGH 

SPRAY PAINT BOOT! 



DATE 



1-91 



I 



'{CURE 



VS-75 -04 



10-116 



industrial Ventilation 




Air filters in doors desirable 



To fan and discharge (fan should- - 

have inspection door) 

PLAN VIEW 

Q ~ 1 00 cfm/ft 2 of cross sectional area 

(when W x H is greater than 150 ft 2 , Q - 50 cfm/ft 2 ) 
h e - 0.50 VP c j plus resistance of each filter bank when dirty 
Minimum duct velocity ~ 2000 fpm 

Paint arresting filters to be sized for 100-500 cfm/ft 2 
of filter. Consult manufacturer for specific details. 



D D D 

□ a d 




Alternate exhaust duct 



ELEVATION 
Typical filter installation 

Note: For airless spray painting use 

Q ~ 60 cfm/ft 2 of cross-sectional area 



For construction and safety 

consult NFPA code (Reference 10.75.1). 



Paint arresting filters 



duct diameter 4- 6" 



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OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



PAINT BOOTH 
VEHICLE SPRAY 



DATE 



1-91 



figure ys-75-05 



Specific Operations 1.0-117 



Locale takeoffs 15 feet on center 
Q = 50 cfm/fi 2 drain board area, 

but not less than 100 fpm indraft 

through openings 
h e = 0.25 VP e 
Minimum duct velocity ~ 2000 fpm 




/ 
For best results enclose 
dra inboard as a tunnel 



To suit 
work 



1000 fpm maximum 
plenum velocity 

Q = 125 cfm/ft of tank and drainboard area 

Slot velocity = 2000 fpm 

h e == 1.78 VP S + 0.25 VP d 

Minimum duct velocity = 2000 fpm 

For air drying in a room or enclosure, see Chapter 2 for 
dilution ventilation required. 

For construction and safety, consult NFPA codes, Reference 10.75.1 



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OF GOVERNMENTAL 
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DIP TANK 



DATE 



1 - 91 



FIG 



jure VS-75 06 



10-118 Industrial Ventilation 




Siot for conveyor 

Extend as low as possible 
to clear work 




Side baffles 
are desirable 



SLOT TYPE 

Q = 100 cfm/ft 2 door plus 1/2 
Products of combustion 

Minimum duct velocity = 2000 fpm 
h e = 1.0 VP S + 0.25 VP d 

Size plenum for 
1000 fpm maximum 



Siot on three sides with 
V s = 2000 fpm. 

Locate on inside or outside of door. 



CANOPY TYPE 

Q = 200 cfm/ft 2 



of hood face 
plus 1/2 products of combustion 

= 0.25 VPd 



h, 



Duct velocity = 2000 fpm 



Notes: 

1 . Eor dryers, include rate of 
water vapor liberated. 

2. Eor flammable solvent drying 
refer to Chapter 2, ''General 
Industrial Ventilation". 

3. Hoods at each end of oven. Reduce 
size of doors as much as possible. 
Separate vent must be added for 
products of combustion. 

4. Eor construction and safety, consult 
NEPA code (Reference 10.75.1). 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGTENISTS 



WING OVEN VENTILATION 



DATE 



1-91 



figure VS-75-20 



Specific Operations 10-119 



Paint mix 

pot (typ.) V"V 



Q(exh) = 10 to 12 air changes/hour 

Q(supply) = 1 to 12 air changes/hour 

Balance room slightly negative 

Q(exh) = 1 .05 Q(supply) 

Entry loss = 1.78 VP S + 0.05 VP D 

Minimum duct velocity - 2000 fpm 

Stack velocity - 3500 fpm 



Floor line 



For construction and safety consult NFPA Code 
Reference 10.75.1 and 10.75.2 



f 



^Z<-^,_^' 



A \ 



Supply 




(or 



AMERICAN CONFERENCE 



OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



PAINT MIX 
STORAGE ROOM 



DATE 



91 



FICURE VS- 75 30 



10-120 



Industrial Ventilation 



10.80 MECHANICAL SURFACE CLEANING AND 
FINISHING 

Mechanical surface cleaning is generally used to clean a 
surface in preparation for painting, welding, or other opera- 
tions. The surfaces may be coated with paint, rust, or oxida- 
tion; plated with other metals; or covered with molding sand, 
inorganic, organic, or biological matter. Mechanical cleaning 
may be accomplished by abrasive blasting, wire wheels, sand 
paper/sanding belts, grinding wheels, or use of abrasive chips 
in tumbling mills. The capture velocity needed to entrain large 
particles is often very high and the collection hood must be 
positioned so the materials are directed toward the hood. A 
minimum duct transport velocity of 3500 fpm is needed but 
4000 to 5000 fpm is recommended. A hood that encloses as 
much of the operation as practical is desired. The toxicity of 
the material removed must be considered when cleaning 
mechanically. Complete enclosures may be used or the 
worker may need to wear a respirator in addition to using local 
exhaust ventilation. 

For many grinding, buffing and polishing operations, regu- 
lations from the Occupational Safety and Health Administra- 
tion (OSHA) (l0 - 80l) and National Fire Protection Association 
(NFPA)< l0802 > may apply. 

10.80.1 Abrasive Cleaning: VS-80-0l,-02,and-03 show 
suggested designs for abrasive blasting and tumbling mills. A 
supplied air respirator must be used in abrasive blasting 
rooms. 

10.80.2 Grinding: Mechanical surface finishing uses or- 
ganic bonded wheels, cones, saws or other shapes rotating at 
a high rate of speed to smooth a surface; reduce an object or 
part in size; or perform other operations. As the object is being 
surfaced or finished, metallic particles are removed and leave 
the object at a high speed. In addition, the abrasive wheel is 
reduced in size and generates particles that must be controlled. 
Frequently, grinding is accomplished using fluids to keep the 
parts cool. This cooling fluid will be emitted as an aerosol or 
mist and needs to be controlled and provisions must be made 
in the duct to drain off the liquids that accumulate. 

The hood used to capture the particles should enclose the 
operation as much as possible and be positioned to take 
advantage of the velocity and direction of the particles as they 
are generated. Design specifications for grinding and surfac- 
ing operations are shown in VS-80-10 through VS-80-19. 



10.80.3 Buffing and Polishing: The same principles ap- 
ply for buffing and polishing as for grinding and surfacing. 
The buffing wheel or belt should be enclosed as much as 
practical and positioned to take advantage of the centrifugal 
force of the particles as they leave the wheel or belt. The 
minimum duct velocity for the generated particles is 3500 fpm 
and 4500 fpm if the material is wet or sticky. Since many 
varieties of metals and alloys are buffed and polished, it is 
extremely important not to mix ferrous and non-ferrous met- 
als in the same exhaust systems (see NFPA codes)/ 10 80 - 2) 
VS-80-30 through VS-80-35 show suggested designs for 
buffing and polishing. 

REFERENCES 

10.80.1 U. S. Department of Labor, Occupational Safety and 
Health Administration: 29 CFR.1910. US- 
DOL/OSHA, Washington, DC (1970). 

10.80.2 National Fire Protection Association: National Fire 
Codes — in particularNFPA-65 (Processing and Fin- 
ishing of Aluminum); NFPA-68 (Guide for Explo- 
sion Prevention Systems); NFPA-77 (Practice on 
Static Electricity); NFPA-91 (Installation of Blow- 
ers and Exhaust Systems for Dust, Stack and Vapor 
Removal or Conveying); NFPA-480 (Storage, Han- 
dling and Processing of Magnesium); NFPA-481 
(Production, Processing, Handling and Storage of 
Titanium); NFPA-482 (Production, Processing, 
Handling and Storage of Zirconium); and NFPA- 
561 (Manufacture of Aluminum and Magnesium 
Powder). NFPA, Quincy, MA. 

10.80.3 National Institute for Occupational Safety and 
Health: Recommended Industrial Ventilation 
Guidelines. DHEW (NI.OSH) Pub. No. 76-162; 
NTIS Pub. No. PB-266-227. National Technical In- 
formation Service, Springfield, VA (1975). 

10.80.4 American Foundrymen's Society, Inc.: Foundry 
Ventilation Manual. AFS, Des Plaines, IL (1985). 

10.80.5 National Institute for Occupational Safety and 
Health: Ventilation Requirements for Grinding, 
Buffing and Polishing Operations. DHEW (NIOSH) 
Pub. No. 75-107; NTIS Pub. No. PB-277-332. Na- 
tional Technical Information Service, Springfield, 
VA(1975). 



Specific Operations 10-121 



3500 f pm, minimum 



>-- 



To dust collector 




Floor grille 



iniiijiiiixii 



SECTION THROUGH TYPICAL R00I 



Q - 60-100 cfm/ft 2 of floor for downdrafi with typical choice 80 cfm/ft^ 
Q = 1 00 cfm/fi^ of wall for crossdraft- 

Lower control velocities may be used depending on toxicity of the contaminant, 
object and blasting media and the size of the blasting room. 

Notes: 1 . The above ventilation is for operator visibility and to control escape 
of contaminants info adjacent work areas. 

2. Operator in an abrasive blasting room is required to wear appropriate 
NIOSH certified respiratory protection. 

3. For rotary tables use 200 cfm/ft^ of total opening (taken without 
curtains). 

4. For blasting cabinets see VS — 80 — 02. 



Reference 10.80.3 and 10.80.4 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
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ABRASIVE BLASTING 
ROOM 



DATE 



02- 91 



FIGURE VS 80 01 



10-122 Industrial Ventilation 



Air intake vents baffled — * 



Screened viewing 
window ■• — 



Rubber gloves 
attached to 
cabinet 




To Exhaust System 



Door with 
dust tight 
gasket 



Return to blasting hopper for reuse 



Notes: 20 air changes per minute 

At least 500 fpm inward velocity at al! openings 

Minimum duct velocity = 4000 fpm 

If cabinet has self-contained dust collector, 

consult manufacturer for losses. 

h a = 1.0 VR, 



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ABRASIVE BLASTING- CABINET 



date Q2-91 I F1GURE VS- 80 02 



Specific Operations 10-123 



- End bell 



Hinged access door- 



r°i 



Holes in disc 



Open 



^-- Trap 




SECTION THROUGH HOLLOW 
TRUNNION TUMBLER 

Minimum duct velocity = 5000 fpm 
Entry loss (h e ) depends on design and 
typically ranges from 3 to 9 " wg. 



Air slot velocity : - 
400 fpm. minimum 



STAVE MILL 
(END SECTION) 

Minimum duct velocity - 3500 fpm 
h e = 0.25 VP d 



EXHAUST RATES 



Square mill 
side diam in. 


Round 
i.D. in 


mill 
inches 


Exhaus 


t cfm** 


Trunnion 


Stave 




Up to 


24 incl. 


430 


800 


Up to 24 inci. 


24 


- 30 


680 


900 


25 - 30 


31 


- 36 


980 


980 


31 - 36 


37 


- 42 


1330 


1330 


37 - 42 


43 


- 48 


1750 


1750 


43 - 48 


49 


- 54 


2200 


2200 


49 - 54 


55 


- 60 


2730 


2730 


55 - 60 


61 


- 66 


3300 


3300 


61 - 66 


67 


- 72 


3920 


3920 


67 - 72 




4600 


4600 



* Low-loss designs have large air inlet openings in end bell. 

Holes in end discs are sized for velocities of 1250 — 1800 fpm. 

** For lengths over 72'', increase exhaust rate proportionately 



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TUMBLING 



LLS 



DATE 



02-91 | FIGURE VS-80-03 



10-124 



Industrial Ventilation 



Adjustable tongue (keep adjusted to not 
more than 1 /4" from wheel ) 



clearance 



2 




H2 



7" 



r i" 

L _L s p a c e 
4 H 



SPECIAL TOOL REST 
Reference 10.10 



Chip trap if desired 



EXHAUST FLOW RATES, cfm 



Wheel diam. 
inches 


Wheel 

width 

inches 


Good 
enclosure* 


Poor 
enclosure 


Up to 5 


1 


220 


390 


5 to 10 


1.5 


390 


610 


10 to 14 


2 


500 


740 


14 to 16 


2 


610 


880 


16 to 20 


3 


740 


1000 


20 to 24 


4 


880 


1200 


24 to 30 


5 


1200 


1600 


30 to 36 


6 


1600 


2000 



*Special hood and tool rest as shown, no more than 25% ot the wheel exposed. 
Minimum duct velocity = 4000 fpm 

h = 0.65 VP d for straight take-off 
= 0.40 VP d for tapered take-off 



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GRINDING WHEEL HOOD 
SURFACE SPEEDS 
ABOVE 6500 sfpm 



DATE 02- 91 



figure VS-80-10 



Specific Operations 



10-125 



L 



\ / 
V 



7 



I / \ 

k: \J 

I 



Adjustable tongue (keep adjusted to not 
more than 1/4" from wheel) 

m 

1" clearance 
J 



1 to H 




Chip trap 
if desired 



EXHAUST FLOW RATES, cfr 



Wheel diam 
inches 


Wheel 
width 
inches 


Good 
enclosure* 


Poor 
enclosure 


Up to 5 


1 


220 


220 


5 to 10 


1 .5 


220 


300 


10 to 14 


2 


300 


500 


14 to 16 


2 


390 


610 


16 to 20 


3 


500 


740 


20 to 24 


4 


610 


880 


24 to 30 


5 


880 


1200 


30 to 36 


6 


1200 


1600 



* No more than 25% of wheel exposed. 
Minimum duct velocity = 4000 fpm 

h e = 0.65 VP d for straight takeoff 
0.40 VP d for tapered takeoff 



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GRINDING WHEEL HOOD 
SURFACE SPEEDS 
BELOW 6500 sfprn 



DAT 



02- 91 



FIGURE 



VS-80-11 



10-126 industrial Ventilation 



Grinding wheel 



Example for: 
x = 4" 
A = 3" x 4.5 ? 



Vs 
1000 


Q , c f m 
"52 


I 2000 


100 


3000 


1 60 


4000 


210 


5000 


260 


6000 


310 


7000 


360 


8000 


420 


9000 


470 


10000 


520 



45" Taper wheel 




Metal strip 



Q = 0.043 V s (10X + A) 

Minimum duct velocity = 3500 fpm 

h e = 0.25 VP d 

X = distance from hood face to center of wheel, ft 

9 

A = hood face area, ft 

V s = Wheel Speed, surface feet per min. (SFM) 

= 7T (D/12) R 

D ~ diameter in inches 

R = rpm of qrindinq wheel 

Reference 10.80.5 



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SURFACE GRINDER 



DA 



TE 02-91 



figure VS-80-12 



Specific Operations 10-127 




Slip and swivel join! 



Access door — \ / 



Supports 



Cut to suit around frame 



Slot 



L- Disc _| 
Diameter 



Use canvas or rubber baffles 
to surround disc as far as 
possible 



6^ 



Disc diameter 


Duct diameter 


cfm 


Up to 20" 


6" 


900 


20" to 30" 


' 8" 


1600 


30" to 53" 


12" 


3500 


53" to 72" 


16" 


6300 



Minimum duct velocity = 4000 fpm 
Minimum slot velocity = 2000 fpm 

h e = 1.0 VP s + 0.40 VP d 

(2) h + elbow losses + joint Josses 



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CORE GRINDER 



DATE 



OR 01 J™ 1TP -c ??? 80 13 



10-128 



Industrial Ventilation 




Ring attached to hood at 
convenient locations 




Adjustable to 
clear grinder 



Angle of slots to be in relation to rotation 



EXHAUST FLOW RATE, cfm 



Disc diameter 


1/2 or more of disc covered 


Disc not covered 


inches 


No.* 


Exhaust flow rate, cfm 


No.* 


Exhaust flow rate, 


cfm 


Up to 20 


1 


500 


2 


780 


20 to 30 


2 


780 


2 


1500 


30 to 53 


2 


1800 


4 


3500 


53 to 72 


2 


3100 


5 


6000 





* Number of exhaust outlets around periphery of hood or equal distribution 
provided by other means. 



Minimum slot velocity = 2000 fpm 
Minimum duct velocity = 4000 fpm 
h e = 1.0 VP S + 0.5 VP d 



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VERTICAL SPINDLE DISC 
GRINDER 



DATE 02-91 I F1GURE VS-80-14 



Specific Operations 10-129 



? 
A 



Close clearance 






n::::::: 



:o> 



---pp 



Work 



LX^ 



r Work 




Endless belt conveyor or 
any other method. 



Disc diameter 
inches 


Exhaust flow rate 
cfm 


Up to 19 


610 


19 to 25 


880 


25 to 30 


1200 


30 to 53 


2000 


53 to 72 


6300 



Section A — A 



Note: If the disc is tightly enclosed by machine 
housing, then exhaust from the housing is 
acceptable. 



Minimum duct velocity ~ 4000 fpm 
h e - 0.65 VP d straight take-off 
= 0.45 VP d tapered take-off 



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HORIZONTAL DO UBLE- SPINDLE 
DISC GRINDER 



DATE 



02-91 



FIGURE 



: VS-80-15 



10-130 



Industrial Ventilation 



Branch takeoff at top or back. Central location 
or multiple branches if several booths are used. 



Additional adjoining 
booths if needed. 



45" slope 




Booth encloses grinder 
frame and suspension. 



Grinder to operate in or 
close to face opening 



Keep width as small 
as practical. 



For- a large opening, 4 ' to 6' wide 
Q = 150 cfm/ft 2 of opening 



For a small opening, T to 2'-6" with 
grinder in front 



Q 



200 cfm/ft of opening 



Minimum duct velocity = 3500 fpm 
h e = 0.25 VP d 



NOTE: Small local exhaust hoods mounted behind 

grinder wheel may trap the stream of sparks, 
but are usually not effective in control of 
airborne dust. 



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SWING GRINDER 



DATE Q2-91 



? IGURI 



E VS- 80-16 



Specific Operations 10-131 



Reduce open area 
with baffles 



/ Top takeoff (optional) 



Booth width to suit 
regular work 



Hinged side doors may be 
opened for longer pieces — 

Saw operates at face 

of booth — 



Close In area under 
table - - 




Hinged c I e a n o u t 
door 



= 250 cfm/ft of open face area 
Minimum duct velocity = 4000 fpm 
h e = 0.50 VP d (no taper) 

0.25 VP d (with 45" taper) 



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ABRASIVE CUT- OFT 
S.A W 



DATE 



02-91 



F'IGUR 



E VS-80-17 



10-132 



Industrial Ventilation 



Back and side shields highly 
desirable, enclose sides and 
top to make booth if practical 



Bench top — — 




Cleanout doors 
or drawers. 



Tapered takeoff necessary 
for distribution. 



45° min 



END VIEW 



Q = 150-250 cfm/ft of bench area. 

Minimum duct velocity = 3500 fpm 

h e = 0.25 VP d 

If slots are used for distribution 

h e = 1.78 VP S + 0.25 VP d 

Notes: 1. If grinding in a booth, use 100 fpm face velocity. 

2. For downdraft grilles in floor: Q=100 cfm/ft of working area. 

3. Provide equal distribution. 

4. Provide for cleanout . 



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HAND GRINDING BENCH 



DATE 02-91 f™F YS- 80-18 



Specific Operations 10-133 




Opening \o be 
sized to handie 
3/4 of total air 
at 1000 fpm 



Opening to be sized to handle 
1/4 of total air at 200 fpm 



X> 



\ 



\ 



\ 2/3 of duct 
width 




Baffle 
plate 



Q = 150 cfm/ft opening 
Minimum duct velocity - 3500 fpm 
h e = 0.25 VPd 



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PORTABLE CHIPPING AND 

GRINDING TABLE 



DATE 



02-91 | F1GURE VS-80-19 



10-134 



Industrial Ventilation 



Adjustable iongue - keep 
as close to wheel as 
possible. 



h- 0.25 D at least. 




Direction of rotation. 



1 .5" for hard wheel. 
3" for soft whee 




— - 0.75 D if 1 

possible 



Trap with cleanout 
when desirable. 



Minimum duct velocity : 3500 fpm, 

4500 fpm if material is wet or sticky 
h e = 0.65 VP d for straight take-off. 
h e = 0.40 VP d for tapered take-off. 



Wheel diam. 
inches 


Wheel width 
inches 


Exhaust flow rate 
cf m 


Exhaust flow rate 
cf m 






Good enclosure * 


Poor enclosure 


Up to 9 


2 


300 


400 


over 9 to 1 6 


3 


500 


610 


over 16 to 19 


4 


610 


740 


over 19 to 24 


5 


740 
1040 


1200 


over 24 to 30 


6 


1 500 


over 30 to 36 


6 


1200 


2000 



* not more than 25% of the wheel is exposed 

Note : Consult applicable NFPA codes Reference 10.80.2 

Caution : Do not mix ferrous and non — ferrous metals 
in same exhaust system. 



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MANUAL 

BUFFING AND POLISHING 



DATE 



02- 91 B FIG, - ,RE !.••' (U)--i(- 



Specific Operations 10-135 



-5 



Adjustable tongue - keep 
minimum clearance. 



Keep side clearance 
to a minimum. 



When main duct Is overhead, 
locate preferred take-off 
location as shown. 



Adjustable hopper — 



Cleanout door 




Alternate 
take-off. 



Wheel diam, 
inches 


Wheel width 
inches 


Exhaust flow rate 
cfm 


Up to 9 


2 


400 


9 to 16 


3 


610 


16 to 19 


4 


740 


19 to 24 


5 


1200 


24 to 30 


6 


1500 


30 to 36 


6 


1900 



Note: For wider wheels than listed, increase cfm with width 
Minimum duct velocity = 3500 fpm 

4500 fpm if material is wet or sticky 
h e - 0.40 VP d 

Notes; 1. Consult applicable NFPA codes. See Reference 10.80.2 

2. For titanium, aluminum, and magnesium, eliminate hopper 
use 5000 fpm through hood cross-section. 

3. Caution ; Do not mix ferrous and non-ferrous metals in same exhaust system. 



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BUFFING LATHE 



DATE 



02- 91 



FIGURE VS-80 31 



10-136 



Industrial Ventilation 



?z 


T 



Side^ opening should be minimal 
1/4" maximum desired 



0.75 D 



Adjustable tongue; not more than 1/4' 
from belt 



Hinged side panel for 

maintenance 



Clean out door 




Belt width 
inches 


Exhaust flow role 
cf m 


Exhaust flow rate 
cfm 


Good enclosure * 


Enclosure 


1 1/2 


220 


300 


2 


390 


610 


3 


500 


740 


4 


610 


880 


5 


880 


1200 


6 


1200 


1570 



* Hood as shown; no more than 25% of wheel 
exposed. 

h e = 0.40 VP d 

Minimum duct velocity - 3500 fpm, 4500 fpm if wet or sticky. 

Notes: 

1. Consult applicable NFPA codes, 10.80.2 

2. For titanium, aluminum and magnesium eliminate hopper and use 
5000 fpm through hood cross section. 

3. Caution: do not mix ferrous and non-ferrous 
metals in same exhaust system. 



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HACKSTAND IDLER 
POLISHING MACHINE 



DATE 



02-91 l FIGURE VS^ 80^32 



Specific Operations 10-137 



Use one branch duct 
/~ for each wheel 



Hinged access doors for maintenance, 
normally closed 




Slow speed belt conveyor 



Q = 500 cfm/wheel, minimum 

Not less than 250 cfm/ft total open area 

Minimum duct velocity = 3500 fpm, 4500 fpm if 

material is wet or sticky 



h f 



1 .78 VR, + 0.25 VP H 



Note: 

1. Consult applicable NFPA standards. Reference 10,80.2 

2. Caution: Do not mix ferrous and non-ferrous 
metals in same exhaust system. 

3. Wheel adjustments on outside of enclosure. 

4. For highly toxic material, enclose the return strand 
of the belt conveyor . 



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INDUSTRIAL HYGIENISTS 



STRAIGHT LINE AUTOMATIC 
BUFFING 



DATE 



0S~ 91 



figure VS-80-33 



10-138 



Industrial Ventilation 




Access door 



Q = 500 cfm/wheel, minimum 

Not less than 250 cfm/ft tota! open area 
Minimum duct velocity ~ 3500 fpm, 4500 fpm if 

material is wet or sticky 
h e ^ 1.78 VP S + 0.25 VP d 

On small, 2 or 3 spindle machines, one take-off may be used. 
Multiple take-offs desirable. 

Note: 

1. Consult applicable NFPA standards, Reference 1.0.80.2 

2. Caution: Do not mix ferrous and non-ferrous 
metals in same exhaust system. 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
.INDUSTRIAL H YGIENISTS 



CIRCULAR AUTOMATIC BUFFING 



DATE 



02-91 



figure VS-80-34 



Specific Operations 10-139 



Flange 



Sliding tongue 






! /h w "T 







1 — Bell tension 
Side hinged 



"" " " " 

Beit width, inches 


Exhaust flow rate, cfm 


Up to 3 


220 


3 lo 5 


300 


5 to 7 


390 


7 to 9 


500 


9 to 1 1 


610 


11 io 13 


740 



Minimum duct velocity - 3500 fpm. 4500 fpm if 
material is wet or sticky 
h = 0.65 VPdfor straight take-off 
0.45 VPdfor tapered take-off 

Notes: 
1 . Consult applicable NFPA codes Reference 1 0.80.2 
2. Caution : Do not mix ferrous and non-ferrous 
metals in same exhaust system 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
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METAL POLISHING- BELT 



DATE 



02-91 



FIGURE 



7S-80- 35 



.10-140 



Industrial Ventilation 



10.85 VEHICLE VENTILATION 

The objective of providing ventilation for vehicles in an 
environment is to keep a worker's exposure to toxic exhaust 
fumes and gases below the TLV, both the TWA and STEL, 
or other appropriate standards. This can be achieved either by 
dilution or local exhaust ventilation. 

It is difficult to establish dilution ventilation requirements 
accurately for the operating vehicles in a plant. For an existing 
facility, the designer has the opportunity to measure the 
emission in the field. Standard techniques can be used to 
measure gas flow rates, composition and temperatures, and 
contaminant levels. Using the equations in Chapter 2 and the 
measurements, the dilution rates can be calculated. However, 
it is not always possible to accurately determine the contami- 
nant generation rate because generation is not uniform. More- 
over, no such data are available to the designer for new 
vehicles. 

The use of dilution ventilation is usually considered only 
after rejection of the source capture concept. Common reasons 
for rejecting source capture (local exhaust) are operating 
interference problems or layout constraints. For lift trucks or 
cars in motion or idling outside of stalls, local exhaust is not 
feasible. Hence the only method for control of health hazards 
is dilution ventilation. 

Over the years, some empirical rates have been developed 
which have been applied successfully to achieve contaminant 
control. The recommended dilution rates based on average 
operating conditions are: 

5,000 cfm/propane fueled lift truck 

8,000 cfm/gasoline fueled lift truck 

5,000 cfm/operating automobile 

1 0,000 cfm (or more)/operating truck 

100 cfm/horsepower for diesel fueled vehicle 

The above dilution rates for lift trucks apply under the 
following conditions^ 1085 !) 

1 . A regular maintenance program incorporating final 
engine tuning through carbon monoxide analysis of 
exhaust gas must be provided. CO concentration of 
gases should be limited to 1% for propane fueled 
trucks; 2% for gasoline fueled trucks. 

2. The periods of lift truck engine operation do not ex- 
ceed 50% of the working day (total engine operation 
of lift truck equal to or less than 4 hours in an 8-hour 
shift). 

3. A reasonably good distribution of air flow must be 
provided. 



4. The volume of space must amount to 1 50,000 rWlift 
truck or more. 

5. The lift truck is powered by an engine of less than 60 
HP. 

Where actual operating conditions vary from the above, the 
ventilation rate should be increased. On the other hand, me- 
chanical ventilation may not be required in large buildings 
where lift truck operation is intermittent and where natural 
infiltration based on a maximum of one air change/hour for 
the net building volume exceeds the recommended dilution 
ventilation rate. 

The alternative to dilution ventilation is to capture the 
contaminant at the source by installing local exhaust ventila- 
tion. For stationary vehicles in service garages, effective 
systems are shown in VS-85-01 (overhead) and VS-85-02 
(under floor). The systems should be connected directly to the 
vehicle exhaust and should terminate outdoors above the roof. 
The design procedure outlined in Chapter 5 must be followed. 
For friction loss data of flexible ducts, manufacturers should 
be contacted. As with all flexible systems, the length of 
flexible duct must be minimized, and non-collapsible duct 
should be used. Unnecessary and/or sharp bends should be 
avoided. Exhaust requirements for automobiles are shown in 
VS-85-02 and for diesel engines in VS-85-03. 

The requirements for parking or storage garages should be 
based on short-term exposure of drivers to exhaust emissions 
when entering or departing. A continuous supply of 500 cfm 
fresh air/parking space should be adequate. Additional venti- 
lation may be required if there are long periods of engine 
idling (winter warm-ups, loading, etc.) or if the general traffic 
pattern is such that clusters of vehicles arrive or depart. 

Attendant booths of parking garages should be pressurized 
with a supply of fresh air from uncontaminated sources. 

For indoor loading docks, continuous supply of 2 cfm/ft 2 
of dock area should be adequate where truck motors are shut 
off except when entering or leaving the dock. 

REFERENCES 

10.85.1 Hama, G.M.; Butler, Jr., K.E.: Ventilation Require- 
ments for Lift Truck Operation. Heating, Piping and 
Air Conditioning (January, 1970). 

10.85.2 Goldfield, J.; Sheehy, J.W.; Gunter, B.J.; Daniels, 
W.J.: An Affordable Ventilation Control for Radia- 
tor Repair Shops. Ventilation '91: 3rd International 
Symposium on Ventilation for Contaminant Con- 
trol. American Conference of Governmental Indus- 
trial Hygienists, Cincinnati, OH (1993). 



Specific Operations 10-141 



Roof 



Stackhead - See Fig. 5-31 




Main duct: 

Plenum design best - size for 2000 fpm maximum 

or design as in chapter 5. 



At least 5' 



- 7 _ ? -r—y 7 — z — 7 / / / / ,~~~~ 7 ~- ~~ T ~~~ 7 z z _ z _^ = ^2_^ T 7 



Hose can be 
counterweighted 




zz::5 



\i 



Fan 



1 0'— 1 2' from 
floor 



All joints soldered 



Flexible joint 



# % 



For dual tailpipes 
use one hose with "Y" 
or use two outlets per 
stall. 



Floor 



Vehicle horsepower (hp) 


cf m/vehicle 


Flexible duct 
diam. 


Branch 
connection 


Up to 200 


100 


3" 


4" 


Over 200 


200 


4" 


4" 


Diesel trucks 




See VS-85-03 





On dynamometer test rolls 

Automobiles and light duty trucks = 350 cfm 

Heavy duty trucks = 1200 cfm minimum 

For friction loss of flexible duct; consult manufacturers' data. 

See VS-85-02 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SERVICE GARAGE VENTILATION 
VERHEAD 



DATE 



12-90 l FIGURE VS~-85 ::: 01 



10-142 



Industrial Ventilation 



Flex duct to 
tailpipe 



^ Double or single floor plates suitable. 

x - Self -closing floor plates desirable. 



Note : In ventilating a garage use either the overhead or under floor system. 

Exhaust to be discharged above roof. i 

To fan and - --.^ I 
discharge above ^ 
roof 




30" -4- 5" 
Along ceiling of fioor below, or in trench. 
If in trench, drain tile with cemented 
joints is suitable. Must be sloped and 
drained for flushing. 



Size main for 
2000 fpm or 
less. 



Sump or dry 
well 



UNDER FLOOR SYSTEM 
EXHAUST REQUIREMENTS* 



Type 


cfm per vehicle 


Flex duct ID ( m i n ) 


Automobiles and trucks up to 200 hp 


100 


3" 


Automobiles and trucks over 200 hp 


200 


4" ** 


Diesel 


See 


VS-85-03 



On dynamometer test rolls 

Automobiles and light duty trucks = 350 cfm 

Heavy duty trucks = 1 200 cfm minimum 

3" diarn. permissible for short runs with proper fan. 

For friction loss of flexible duct; consult manufacturers' data. 




Use adapters 
on dual exhausts 
and special 
tailpipes. 




AMERICAN CONFERENCE 

OF GOVERNMENTAL 
NDUSTR] AL H YGIENISTS 



SERVICE GARAGE 
VENTILA TION UNDERFLOOR 



DATE 



12-90 pGME VS-- 85^02 



Specific Operations 10-143 



2200 






2000 
1800 _ 
1600 

1400 

E 

r> 1200 

| 1000 _ 

X 
LU 

800 _ 
600 „ 
400 
200 

. 






i 

1 


[ 
























{, 


I 
























-I 


J j 


























y 




























' 




i 
























NA — Normally Aspirated 
TC - Turbo Charged 








r 












/ 


































































































/ 


























z 






_ ^' 


~/ 1 


















z 




_ ^ 


















^y 


Z 


/ | 






— 


— 


c 














Z 




> 






| 














/ 


/ 


/ 






i 




































y 


/ 




















_...-- 












A 


5L 


(UM 


----- 












...._---- 




■-■""'"^ 
















■-~^~ 
























10 

Exhaust, c 
For specif 
EPA engine 


00 2000 3000 4000 
Engine speed, rpm 

fm = acfm + 20% excess 

c design information request manufacturers 13 

a bench test 


mode 


AMERICAN CONFERENCE 
OF GOVERNMENTAL 

1 T 'M "H T T C 1 H" 1 1 A T O VP T T? a ~ T o n" 1 o 


EXHAUST SYSTEM REQUIREMENTS 

FOR TYPICAL DIESEL ENGINES 

UNDER LOAD 


1 IJNJJ Uo I K1.AL hi Y 


criJ 


_M\ 


lo 


] o 


i ■*■!■;■: /.;_; (j() |r-.. ;lil :: >■ .#•: (/ \j 



10-144 Industrial Ventilation 




Mln. 45' 











L Lif1 
-*~T — w — 












Water tank 

















i 



Q 



Water tank 



Lift r 



Elevation View 




Enclosure 
opening 

-*— W —m 
Plan View 



Note 1 



Q - 190 cfm/ft z 

opening (L x W) 
Minimum duct velocity = 3000 f pm 
h e ^ 0.25 VP d 
L ~ Height of ventilated opening in mm. 
W ~ Width of ventilated opening in mm. 



Notes: \ _ Enclosure can be made of metal or fire resistant curtain 
material 



Reference: 10,85.3 



Edge of 
tank 



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OF GOVERNMENTAL 
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VENTILATED BOOTH FOR 
RADIATOR REPAIR SOLDERING 



DATE 



4-94 l F1GURE VS~8{f~W 



Specific Operations 



10-145 



10.90 WELDING AND CUTTING 

The purpose of welding and cutting ventilation is to control 
gases, fumes, and particulate generated during the welding 
and cutting operations. 

10.90.1 Hazards: The generation rate of fumes and gases 
varies with the composition of the base metal, fluxes, and 
fillers, and with the rate and depth of welding. Exposure to 
the welder varies with the generation rate, duration and fre- 
quency of operations, work practices (particularly distance of 
the plume from the breathing zone), and the effectiveness of 
ventilation. 

Contaminants from welding may include: 

1. Fume from the base metals and filler or electrode 
metals. 

2. Fume from coatings (e.g., zinc oxide from galvanized 
surfaces, thoriafrom T.I.G. welding, and fluorides and 
N0 2 from electrode coatings). 

3. Ozone due to ionization of oxygen by the ultraviolet 
light from arc welding. 

4. Carbon monoxide from ultraviolet effects on carbon 
dioxide in shield gas. 

5. Shield gases such as carbon dioxide, helium and argon. 

6. Fluoride gases and other thermal decomposition prod- 
ucts of fluxes and electrode coatings. 

7. Flammable gases such as acetylene. 

There are welding tasks that present enhanced hazards such 
as welding on materials containing or contaminated with 
heavy metals or welding in the presence of flammable vapors 
or halogenated hydrocarbons. If such welding is required, 
extraordinary precautions must be taken on a case-by-case 
basis. Even in the absence of such hazard materials, any 



welding operation in a confined space is potentially lethal and 
requires continuous and copious dilution ventilation. 

10.90.2 General Recommendations: 

1. Choose hood designs in the following descending 
order of effectiveness: enclosing hoods; vacuum noz- 
zles; fixed slot/plenum hood on a worktable or rectan- 
gular hood fixed above a worktable; moveable hood 
above a worktable; moveable hood hanging freely or 
overhead canopy; dilution ventilation. 

2. Integrate planning for ventilation systems with plan- 
ning for materials handling. 

3. Place welding curtains or other barriers to block cross- 
drafts . 

4. Install turntables, work rests, and other aids to improve 
utilization of the hoods. 

5. Avoid recirculating filtered air from welding hoods 
back into occupied spaces unless the welding is low 
hazard and produces low quantities of gaseous con- 
taminants. 

6. Face velocity for enclosing hoods should be 1 00-130 
fpm with the higher values used for poor conditions 
such as high cross-draft velocities. 

7. Capture velocity for non-enclosing hoods should be 
100-170 fpm with the higher values used for poor 
conditions such as high cross-draft velocities and with 
higher hazard levels. 

Enclosing hoods are by far the most effective in control! ing 
welding contaminants; however, they restrict access and force 
reconsideration of material and product handling. Capturing 
hoods are less effective than enclosures but for low hazard 
conditions can be adequate if properly used. 



10-146 



Industrial Ventilation 



r 45" taper angle 

Slots-size for 2000 fpm 



Baffles are 
desirable 




Maximum plenum velocity 
1 /2 slot velocity 

Q = 350 cfm/ft of hood length 

Hood length - required working space 

W = 24" maximum, if W>24" see chapter 3 

Minimum duct velocity = 2000 fpm 

h e = 1.78 VP S + 0.25 VP d 

General ventilation, where local exhaust can not be used: 



Rod, diam. 


Cf m/welder 


5/32 


1000 
1500 


3/16 


1/4 


3500 


3/8 


4500 



or 



A. 


For open areas, where welding fume can 

rise away from the breathing zone: 

cfm required = 800 x lb/hour rod used 


B. 


For enclosed areas or positions where fume 
does not readily escape breathing zone: 
cfm required = 1 600 x lb/hour rod used 



For toxic materials higher airflows ore necessary and operator 
may require respiratory protection equipment. 

Other types of hoods 

Local exhaust: See VS-90-02 

Booth: For design see VS — 90 — 30 

Q = 100 cfm/ft 2 of face opening 

MIG welding may require precise air flow control 



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I N D U S T R I A L H YG I E N I S T S 



WELDING VENTJLA TI0N 
BENCH HOOD 



DATE 



l-Ql fFicuRE Y2-90-01 



Specific Operations 10-147 



Overhead 
support 



To exhaust 



Swive 




Q 



O 



Galvanized hood 
FLEXIBLE EXHAUST CONNECTIONS 




Cleanout 




PLAIN DUCT 



CONE HOOD 



FLANGED HOOD 



RATE OF EXHAUST 


X, inches 


Plain duct 
cfm 


Flange or 
cone, cfm 


Up to 6 


335 


250 


6-9 


755 


560 


9-12 


1335 


1000 



GENERAL VENTILATION, where local exhaust cannot 

be used 



Rod, diam. 


cfm/welder 


5/32 


1000 


3/16 


1500 


1/4 


3500 


3/8 


4500 



Face velocity = 1 500 f pm 
Minimum duct velocity = 3000 fpm 
Plain duct entry loss = 0.93 VPd 
Flange or cone entry loss = 0.25 VPd 

Notes: 

1 . Locate work as close as 
possible to hood. 

2. Hoods perform best when located 
to the side of the work. 

3. Ventilation rates may be 
inadequate for toxic materials. 

4. Velocities above 100-200 fpm 
may disturb shield gas. 



OR 



A. For open areas, where welding fume can rise 

away from the breathing zone: 

cfm required = 800 x lb/hour rod used 

B. For enclosed areas or positions where fume 

does not readily escape breathing zone: 
cfm required = 1600 x ib/hour rod used 



For toxic materials higher airflows are necessary 
and operator should use respiratory protection 
equipment. 

Other types of hoods 
Bench, see VS-90-01 
Booth, for design see VS — 90 — 30 
Q = 100 cfm/ft^ of face opening 



AMERICAN CONFERENCE 

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INDUSTRIAL HYGIENISTS 



WELDING VENTILATION 
MOVABLE EXHAUST HOODS 



DATE 



l-Ql [FIGURE y S _ 9 Q_Q2 



10-148 



Industrial Ventilation 



Flow line 



Q = 1000 to 1200 cfm/lineor ft of booth 

H e = filter loss + 0.5 VP d 

Minim urn duet velocity — 3500 fpm 






Window 




Grilles 1 2" to 50" high 
D e s i a n for 5 5 t d m 



x:> 



E x h a u 
stack 



Duct 



i \ 



ilk 



/, 



' V V V v v\ 

vv v! V vv' 

V V V ' 

v ^lv V < 



\ 



JE 



X 






IERICAN CONFERENCE 
OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



DATE 



PRODUCTION LINE 
WELDING BOOTH 

2-Qi jfigure vs-90-03 



Specific Operations 



10-149 



Slots sized for 
2000 fprn 



6 foot center 
to center 
maximum 




Cleanout 
doors 



45 taper 
angle 



Enclose base 
of bench 



Q - 1 50 cfm/ft of gross bench area 
Minimum duct velocity — 4000 fpm 



h e = 1.75 



VP. 



+ 0.25 VP d 



AM 



i;r 



[CAN CONFERENCE 



OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



TORCH CUTTING VENTILATION 



DATE 



|FIGURE VS-90- 10 



10-150 Industrial Ventilation 



r- Flexible weld 
\ hose 

rrc ( ( ( ( 




10" 



1/2" -J 



Floor line 



Slot velocity = 2000 fpm 
Hood located within 4" 
from source to maintain 
200 fpm capture velocity 



Rigid pipe 



Flexible 
weld hose 




Slot nozzle 

see detail above 



h e = 1.78 Vl|+ 0.25 VP d 

Minimum duct velocity = 3500 fpm 



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ROBOTIC APPLICATION 



DATE 



02-91 1 FIGURE VS-96^20 



Specific Operations 10-151 



no 



Slot velocity = 2000 fpm- 



Face open- 




45° taper angle 



grille top work bench 



Clean-out doors 
METALLIZING BOOTH 
Non-toxic: Q = 1 50 cfm/ft 2 face area _Toxic: Provide appropriate NIOSH 



certified respirator 

Q = 200 cfm/ft 2 face area 



Minimum duct velocity = 3500 fpm 

h e = 1 .78 VP S + 0.25 VP d 

Small lathe, etc., may be mounted in booth 



l^a, — i2' m i n .— »-| 



Gun (on fool post) 



n. 



Y-- : \ 

\ J 



l~ Flex duct to allow 
/ movement full length 
of work 



_ — T 




: 



l i \ l> 



a 



tr 



Hood extends as low as possible to 
^ clear lathe rail. Hood may be 
LOCAL HOOD connected to move with tool rest- 
Note: Local hood may not be satisfactory for spraying toxic metals. 
= 200 cfm/ft^ face openings 
Minimum duct velocity ~ 3500 fpm 
h e = 0.25 VPd 



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METAL SPRAYING 



DAT! 



l-gi j figure ys-90-30 



10-152 



Industrial Ventilation 



10.95 WOODWORKING 

Woodworking equipment generates large amounts of wood 
dust by abrasive or cutting action. It is important to provide 
good ventilation for all equipment as the broad particle size 
distribution of wood dust creates the potential for health and 
housekeeping problems as well as fire hazards. Excessive 
amounts of dust, if allowed to accumulate inside equipment 
and in shop areas, can create fire or explosion hazards. An 
additional consideration should be the toxicity of the wood 
species used. 

In many instances, woodworking equipment, such as saws 
and sanders, generates air flow patterns which make dust 
control difficult. Exhaust hoods should enclose the operation 
as much as possible. Where the equipment tends to eject wood 
dust (e.g., at sanding belt pulleys) the exhaust hood should be 
placed in the ejection path. 

Enclosures must incorporate cleanout doors to prevent dust 
build-up. Duct velocities should be maintained at a minimum 
of 3500 fpm to prevent settling and subsequent clogging of 
the duct. 

Exhaust flow rates will vary with equipment type and size. 
Design data are provided for a number of operations in 



VS-10-95-1 through 10-95-20 and in Table 10.95.1. Addi- 
tional information for hand-held sanders using Low Volume- 
High Velocity (LV-HV) can be found in sub-section 10.40. 
Where information for a specific operation is not provided, 
data for similar listed operations can be used. 

REFERENCES 

10.95.1 Hampl, V.; Johnston, O.: Control of Wood Dust 
from Horizontal Belt Sanding. American Industrial 
Hygiene Association Journal 46(10)567-577 
(1985). 

10.95.2 Hampl, V.; Johnston, O.: Control of Wood Dust 
from Disc Sanders. Applied Occupational Hygiene 
6(1 1):938-944 (November 1991). 

10.95.3 Topmiller, J.L.; Watkins, D.A.; Schulman, S.A.; 
Murdock, D.J.: Controlling Wood Dust from Orbital 
Hand Sanders. Applied Occupational Hygiene 
11(9): 1131-1 138 (September 1996). 

10.95.4 Hampl, V.; Topmiller, J.L.; Watkins, D.S.; Mur- 
dock, D.J.: Control of Wood Dust from Rotational 
Hand-Held Sanders. Applied Occupational Envi- 
ronmental Hygiene 7(4):263-270 (April 1992). 



Specific Operations 10-153 



Blade 




Blade 

- Hood slotted to 
enclose blade 

—Hinged door 
for cleanout 



Slotted wood 
block 

Top hood 




TOP HOOD DETAIL 



1 — Entire base enclosed 
on all sides 



Blade width, 
inches 


Exhaust flow rate, cfm 


Bottom 


Top 


Total 


Up to 2 

2 to 3 

3 to 4 

4 to 6 
6 yo 8 


350 
350 
550 
550 
550 


350 
550 
800 
1100 
1400 


700 
900 
1350 
1650 
1950 



Minimum duct velocity = 3500 fpm 
h e = 1.75 VP d (Point ® in duct riser) 



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BAND SAW 



DATE 12-90 



FIGURE 



VS-95-01 



10-154 Industrial Ventilation 



Saw blade 



i& 



1/2" minimum, size for 
2000 fpm 



45" taper angle 



\ 

Cleanout door and drawer 



Saw blade diameter, 
inches 


Exhaust flow rate, 
cfm 


Up to 16 


350 


16 to 24 


440 


over 24 


550 


Saw with dado blade 


550 



Minimum duct velocity = 4000 fpm 
h e = 1.78 VP S + 0.25 VP d 



Floor 



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OF GOVERN 



CONFERENCE 



TAL 



:nd 



:al hygienists 



FLOOR TABLE SA W 



date 12-90 FIGURE VS -95-02 



Specific Operations 10-155 



500 cfm 



430 cfm 



Blast gate 



1 3/4" inside diam. 
flexible hose 




Hood 4 1/2" wide 



Minimum duct velocity — 4000 fpm 
h e ~ 3.5 VPd (point A in duct riser) 

For booth enclosure, see VS — 80-17 



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RADIAL ARM SAW 



DATE 



12-90 | FIGURE VS-95-03 



10-156 Industrial Ventilation 




Type hood where table 
is cut through 




Type hood where table 
is not cut through 




Front view 
of hood 



j r 

/ 

I 

tZ [__ 



Front view 
of hood 



Saw diameter, 
inches 


Exh- flow rate 
cf m 


Up to 20 inch 


350 


over 20 


440 



Minimum duct velocity = 4000 fpm 
h e = 1.78 VP S + 0.25 VP d 



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SWING SAW 



date 12-90 1 FICURE VS-95-04 



Specific Operations 10-157 



1/4" 



Saw Blade 



SEE VS-95-02 



Q G = 100 CFM 



Saw Table 




Side View 



Guard Frame 




2" 



Table 




End View 



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TABLE SAW GUARD EXHAUST 



date 4-94 |figure ys-95-05 



10-158 Industrial Ventilation 



Slot 



Single drum 



Table 







Drum surface 
square inches 


Exhaust flow rate 
cf m 


Up to 200 
(and less than 10" diam.) 


350 


200 to 400 


550 


400 to 700 


790 


700 to 1400 


1 100 


1400 to 2400 


1400 



Minimum duct velocity = 3500 fpm 
Entry loss depends on hood design. 
h e = 1.78 VP S + 0.25 VPd as illustrated 



AMERIC AN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



SINGLE DRUM SANDER 



DATE 12-90 



FIGURE VS-95-10 



Specific Operations 10-159 



C I e a n o u t 




Drum covers necessary. Hinge 
or otherwise provide for 
mainfenance. 



Exhaust fl 


dw rates 


Drum length 


Total exhaust for 
machine cfm/drum * 


Up to 31" 


550 


31" to 49" 


790 


49" to 67" 


1 100 


over 67" 


1400 


Brush roils 


350 cfm at brush 



Note: Provide one more take off than 
the number of drums. 

Minimum duct velocity = 3500 fpm 
h e = 0.25 VP d 



AMERICAN CONFERENCE 

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MULTIPLE DRUM SANDER 



DATE 



12^90 



FIGURE 



VS-95-11 



10-160 



Industrial Ventilation 



Duct-B 




Disc diameter, 
inches 


Totai exhaust flow rate 
cfm 


Applies to 
duct 


Up to 12 


350 


A 


12 to 18 


440 


A 


18 to 26 


550 


A 


26 to 32 


700 * 


A-B 


32 to 38 


900 * 


A-B 


38 to 48 


1300 ** 


A-B-C 



* Two bottom branches. 
** One top and two bottom branches. 

Minimum duct velocity — 3500 fpm 
Entry loss depends on hood design. 
h e = 1.0 VP S + 0.25 VPd as illustrated 



Duct-A 



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DISC SANDERS 



date 12-90 |™ure VS-95-12 



Specific Operations 



10-161 



Work 
Table 



0.625" diam. 



Jet Stripper 

location, maun ■ 
to work table 




0.625 



0.1 25 ' diam. to 

pressure gauge 

pressure line optional 



Jet Stripper 
(see note 1) 



NOTES: 

1. Jet stripper may be added as shown, ft is used in 

addition to the ventilation specified in VS— 95-12. 

2. Q e is same as shown in VS— 95-12. 

3. Clearance between stripper nozzle and disk is 0.25". 

4. Nozzle opening diameter is 0.035". 

5. Total air flow (Q s ) to stripper = 1.8 L. (L in inches) 

6. Stripper inlet pressure = 15 psi. 

The Jet Stripper is patented (#5,099,616). Use of the devise (not for sale) is 
permissible. To obtain information regarding license for commercial production 
of the devise contact NIOSH, 1 -800-35NIOSH. 
Reference 10.95.2 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



OPTIONAL JET STRIPPER 
FOR DISK SANDER 



DATE 



11/96 |™e vs-95-12a 



10-162 



Industrial Ventilation 



Head 




Y7 7777;77ZZ7 ZZZA 




Tail 
Table can raise and lower. 




Head 



HORIZONTAL BELT SANDERS 




Belt width, 


Exhaust flow rate ChM 


inches 


Head end 


Tail end 


Total 


up to 6 


440 


350 


790 


6 to 9 


550 


350 


900 


9 to 14 


800 


440 


1200 


over 1 4 


1 100 


550 


1700 



Minimum duct velocity = 3500 fpm 
h e = 0.40 VP d 



AME.RIC AN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



HORIZONTAL BELT SANDERS 



DATE 12-90 



FIGURE 



VS- 95-13 



Specific Operations 10-163 



Main Hood: 

W = belt width, in. 

^E ~ ^ c ^ m P er ' nc ' 1 °^ kelt w 'dth 
Auxiliary Hood: 

I_a - auxiliary hood length to extend 3" 
Into end of sanding area. 

Q^ = 135 cfm per ft of hood length 



Belt 



Note 1 



r jet 




ood 



Workrest- 



Note 2- 



Auxiliary hood 



Notes: 

1 . Tips of stripper jets must be positioned 0.25" from 
belt and inside hood face. 

2. Keep clearances between hood, belt and workrest 
to a minimum. 



L 



-2.5" 



Auxiliary hood 



Orifice 



wm^A 




"E 0.125" 
L 0.063" 



W//// //W \^ 



0.043" 



Tube and Orifice Detail 
Reference 10.95.1 



(CkD 


r- 4 tubes per 


inch 




J y- Orifice 




^H w 




^~~x 




ripper Jet 




belt v 


/idth, in. 





w 

Q s = 0.5 cfm per tube, 

inlet pressure 10 — 12 psig 



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HORIZONTAL BELT SANDER 
PUSH-PULL SYSTEM 



DATE 



12-90 l FIGURE VS- 95-14 



10-164 Industrial Ventilation 



Rear table 



Front table 



jL 



Fence 



Velocity at this - 
space, 2000 fprn 
minimum 



j o r 



7TW 




1/2" minimum clearance 



Clean out or dead-end cap. 



121 



Knife length, 
inches 


Exhaust flowrate 
cf m 


Up to 6 


350 


6 to 12 


440 


12 to 20 


550 


over 20 


800 



Minimum duct velocity = 4000 fpm 
h = 1.0 VP C + 0.25 VP H 



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JOINTERS 



date 12-90 1 F1GURE VS- 95-20 



Specific Operations 10-165 



Sander Air Supply 




■Pienum exhaust 
copper fitting 1" elbow 
or equivalent. 



— — 0.375" 



0.75 




Note 1 



Holes to match 

pad mounting holes (Note 4) 



1. Plenum machined from metal of plastic. 

2. Plenum can be used with air or electrically powered sander. 

3. Qe = 50 cfm at approximately 25" wg. 

4. Mounting holes located to match sander pad mounting holes. 

The Exhaust Plenum Retrofit has a patent pending. Use of the device (not for 
sale) is permissible. To obtain information regarding the license for commercial 
production of the devise contact NIOSH, 1 -800-35NI0SH 

Reference: 10.95.3 



AMERICAN CONFERENCE 

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EXHAUST PLENUM RETROFIT 
FOR ORBITAL HAND SANDER 

[figure vs-95-30 



DATE 12-96 



10-166 Industrial Ventilation 



Note 1 




-Modified Sander Pad 



Note 1 



slot depth = 1/2 d 



slot radius — 1/2 d 




Existing 
Pad Holes 



NOTES: 

1. Additional exhaust plenum to 
fit sander. 

wg 



2. 
3. 



Q e = 50 cfm at 25" 



Sander Air Supply 

^b^ Normal Sander Aspirator Exhaust 
^- Additional Exhaust, Q e 



Qp 



Qp 



- 120° 




Modified Sander Pad. 



Q e is in addition to normal sander 
aspirator exhaust. 
4. Additional exhaust may be supplied 
by standard shop vacuum cleaner. 

The Auxiliary Exhaust Modification Configuration is patented (#5,105,858). 

Use of this device (not for sale) is permissible. To obtain information regarding 

license for commercial production of the device contact NIOSH, 1 -800 — 35NI0SH. 



Reference: 10.95.4 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



AUXILIARY EXHAUST RETROFIT 

FOR AIR POWERED RANDOM 

ORBITAL HAND SANDER 



DATE 



12/97 f^uRE VS- 95-31 



Specific Operations 1 0-1 67 



TABLE 10,95,1 ■ j^isceflganeous Woodworking Machinery not Seven an VS Prints 

The following list of recommended exhaust volumes is for average-sized woodworking machines and is based on many years of experience. It must be noted that 
some modern, high-speed, or extra-large machines will produce such a large volume of waste that greater exhaust volumes must be used. Similarly, some small 
machines of the home workshop or bench type may use less exhaust air than listed. 



Self-feed Table Rip Saw 

Exhaust Flow Rate, cfm 



Saw Diameter, inches 


Bottom 




Top 


Total 


Up to 16 inclusive 

Over 16 

Self-feed, not on table 


440 
550 
800 




350 
350 
550 


790 

900 

1350 




Gang Rip Saws 

Exhaust Flow Rate, cfm 




Saw Diameter, inches 


Bottom 




Top 


Total 



Up to 24, inclusive 
Over 24 to 36, incl. 
Over 36 to 48, incl. 
Over 48 



550 

800 

1100 

1400 



350 
440 
550 
550 



900 
1240 
1650 
2060 



Belt Width, inches 



Vertical Belt Sanders 

(rear belt and both pulleys enclosed) 

and 
Top Run Horizontal Belt Sanders 

Exhaust Flow Rate, cfm 



Up to 6, incl. 
Over 6 to 9, incl. 
Over 9 to 14, incl. 
Over 14 



440 

550 

800 

1100 



Swing Arm Sander: 440 cfm 



Single Planers or Surfacers 

Exhaust Flow Rate, cfm 



Up to 20" knives 
Over 20" to 26" knives 
Over 26" to 32" knives 
Over 32" to 38" knives 
Over 38" knives 



785 

1100 
1400 
1765 
2200 



Double Planers or Surfacers 

Exhaust Flow Rate, cfm 



Up to 20" knives 
Over 20" to 26" knives 
Over 26" to 32" knives 
Over 32" to 38" knives 
Over 38" knives 



Bottom 

550 

785 
1100 
1400 
1400 



Top 

785 
1100 
1400 
1800 
2200 



Total 

1355 
1885 
2500 
3200 
3600 



Scolders, Matchers, & Sizers 

Exhaust Flow Rate, cfm 



Size, inches 



Bottom 



Top 



Right 



Left 



Up to 7, incl. 
Over 7 to 12, incl. 
Over 12 to 18, incl. 
Over 18 to 24, incl. 
Over 24 



440 

550 

800 

1100 

1400 



550 

800 

1100 

1400 

1770 



350 
440 
550 
800 
1100 



350 
440 
550 
800 
1100 



Exhaust Flow Rate, cfm 



Sash stickers 

Woodshapers 

Tenoner 

Automatic Bathe 

Forming lathe 

Chain mortise 

Dowel machine 

Panel raiser 

Dove-tail and lock corner 

Pulley pockets 

Pulley stile 

Glue jointer 

Gainer 

Router 

Hogs 

Up to 12" wide 

Over 12" wide 
Floorsweep 

6" to 8" diameter 



550 

440 to 1400 

Same as moulder 

800 to 5000 

350 to 1400 

350 

350 to 800 

550 

550 to 800 

550 

550 

800 

350 to 1400 

350 to 800 

1400 
3100 

800 to 1400 



10-168 



Industrial Ventilation 



10.99 MISCELLANEOUS OPERATIONS 

In the previous sections of the chapter, hood ventilation 
sketches were grouped together because they provided venti- 
lation concepts for similar operations, used the same ventila- 
tion approach, or were applicable within the same industry. 
However, not all hood ventilation sketches are so easily 
categorized. 

This section provides a location for those hood ventilation 
sketches that do not fit in other sections. Some have a unique 
application such as VS-99-04 for the Pistol Range. Others 
have such broad application that they could appear in many 
of sections (e.g., the canopy hood in VS-99-03). In other 
cases, this section will be used for new ventilation sketches 
for a particular application or industry. Such sketches will 
reside in this section until other hood ventilation sketches are 
developed and a new section formed. Finally, this section will 
be used for tabular presentation of specific design parameters 
for a variety of operations which could not be adequately 
described in previous sections. 

REFERENCES 

1 0.99. 1 Pennsylvania Department of Labor and Industry: 
Abrasive Wheel Manufacture. Safe Practices Bulle- 
tin No. 13. 

10.99.2 Hartzell Propeller Fan Company, Bulletin 1001 . 

10.99.3 Goldfield, J.; Brandt, F.E.: Dust Control Techniques 
in the Asbestos Industry. A paper presented at the 
American Industrial Hygiene Conference, Miami 
Beach, FL (May 12-17, 1974). 

10.99.4 Hama, CM.: Ventilation Control of Dust from Bag- 
ging Operations. Heating and Ventilating (April 
1948). 

10.99.5 Hutcheson, J.R.M.: Environmental Control in the 
Asbestos Industry of Quebec. C 1 MM Bulletin 
64(712):83-89 (August 1971). 

10.99.6 Private Communications. Occupational Health Pro- 
tection Branch, Ontario Ministry of Labour, Ontario, 

Canada (October, 1976). 

10.99.7 Hama, G.; Frederick, W.; Monteith, H.: Air Flow 
Requirements for Underground Parking Garages. 
American Industrial Hygiene Association Journal, 

Vol. 22, No. 6 (December 1961). 

10.99.8 Kane, J. M.: Design of Exhaust Systems. Heating and 
Ventilating 42:68 (November 1945). 

10.99.9 Oddie, W.M.: Pottery Dusts: Their Collection and 



Removal. Pottery Gazette 53:1280 (1928). 

10.99.10 B.F. Sturtevant Company: What We Make. Catalog 
No. 500. 

10.99.11 Brandt, A.D.: Industrial Health Engineering, John 
Wiley and Sons, New York (1 947). 

10.99.12 Kane, J.M.: Foundry Ventilation. The Foundry (Feb- 
ruary and March 1946). 

10.99.13 Kane, J.M. Foundry Ventilation. University of 
Michigan In-Service Training Course (October 
1945). 

10.99.14 Fen, O.E.: The Collection and Control of Dust and 
Fumes from Magnesium Alloy Processing. Peters- 
Dalton, Inc. (January 1945). 

10.99.15 Postman, B.F.: Practical Application of Industrial 
Exhaust Ventilation for the Control of Occupational 
Exposures. American Journal of Public Health, 
30:149(1940). 

10.99 16 New York Department of Labor: Rules Relating to 
the Control of Silica Dust in Stone Crushing Opera- 
tions. Industrial Code Rule No. 34 (July 1942). 

10.99.17 DallaValle, J.M.: Exhaust Hoods. Industrial Press, 
New York (1946). 

10.99.18 Hatch, T.; et al: Control of the Silicosis Hazard in 
Hard Rock Industries. II. An Investigation of the 
Kelley Dust Trap for Use with Pneumatic Rock 
Drills of the Jackhammer Type. Journal of Industrial 

Hygiene 14:69 (February 1932). 

10.99.19 Hay, Capt., P.S.: Modified Design of Hay Dust 
Trap. Journal of Industrial Hygiene 12:18 (January 
1930). 

10.99.20 Riley, E.C.; et al: How to Design Exhaust Hoods for 
Quartz-Fusing Operations. Heating and Ventilating, 

37:23 (April 1940). 

10.99.21 Riley, E.C.; DallaValle, J.M.: A Study of Quartz- 
Fusing Operations with Reference to Measurement 
and Control of Silica Fumes. Public Health Reports 

54:532 (1939). 

10.99.22 Yaglou, C.P.: Ventilation of Wire Impregnating 
Tanks Using Chlorinated Hydrocarbons. Journal of 
Industrial Hygiene and Toxicology 20:401 (June 
1938). 

10.99.23 American Air Filter Co., Inc.: Usual Exhaust Re- 
quirements (for) Grain Elevators, Feed and Flour 
Mills. Louisville, KY (April 1956). 



Specific Operations 1 0-1 69 



Feed 



45° min. slope 

/ -- Flexible connection if desired 

Top take-off preferred 

-— Complete enclosure 
Screen 

— Oversize 




FLAT DECK SCREEN 



l2, 



Q =200 cfm/ft through hood openings, but not less than 

50 cfm/ff screen area. No increase for multiple decks 
Minimum duct velocity = 3500 fp rn 
h e = 0.50 VP d 



Complete — ^ 

enclosure 

Screen - 



45 min. slope 




CYLINDRICAL SCREEN 



Feed 



Q = 1 00 cfm/ft circular cross section of 
screen; at least 400 cfm/ft of 
enclosure opening 

Minimum duct velocity - 3500 fpm 

h Q = 0.50 VP, 



Oversize 



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SCREENS 



DAT! 



12-90 



-IGURE 



VS-99-01 



10-170 



Industrial Ventilation 



£ 



Slot velocity 2000 fpm 



45 




= 50-100 cfm/ft of table top. 

Minimum duct velocity — 2000 fpm 

h e = 1.78 VP S + 0.25 VP d 

Note: See "Open Surface Tanks'', VS-70-01 and VS-70-02 

for other suitable slot types. Air flow rate may 

be calculated on dilution basis if data is available. 

Maximum plenum velocity = 1/2 slot velocity. 

Large plenum essential for good distribution. 



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TABLE SLOT 



DATE 12^90 



figure ys-99-02 



Specific Operations 



10-171 




/ 

45 Minimum 



H I — 0.4H 



T 

H 



TANK 

OR 

PROCESS 



Not to be used where material is toxic and worker must bend 

over tank or process. 

Side curtains are necessary when cross — drafts are present. 



Q = 1.4PHV 


for open type canopy. 

P = perimeter of tank, feet. 

V = 50-500 fpm. See Chapter 3 


Q = (W + L)HV 


for two sides adjacent enclosed. 
W & L are open sides of hood. 
V = 50-500 fpm. See Chapter 3 


Q = WHV 
or 
LHV 


for three sides enclosed, (booth) 
V - 50-500 fpm. See Chapter 3 



h e = 0.25 VP d 

Duct velocity = 1000-3000 fpm 



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CANOPY HOOD 



DATE 



12-90 



FIGURE 



VS- 99- 0c 



10-172 



Industrial Ventilation 



If baffle is used, it must be located at 
the extreme down range end of the booth 




]- 1 0' ceiling height 



Baffles below bench should 
be avoided. 

If baffles are present use a 
bar positioned to provide a 
6 inch space between shooter 
and bench. 



Double 0.25" pegboard wall will provide better distribution 
than single. Wall must be constructed with spacers to 
assure free air circulation within wall. 



Q(minimum) = 50 HW, but not less than 20 cfm/ff" 
of room cross-sectiona! area 



Notes: 

Replacement air distribution: 

Uniform air distribution necessary. 
Perforated rear wall or ceiling plenum system 
preferred. Grilles and diffuse rs are not 
recommended. 

NIOSH certified dust respirator for lead is necessary 

during clean — up and lead removal from bullet trap. 

Acoustical material on walls, ceiling and thick fabric 
on bench top are recommended 



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INDOOR PISTOL AND 
SMALL BORE RIFLE 
RANGE VENTILATION 



DATE 12-90 



figure ys^99-04 



Specific Operations 10-173 



45 minimum slope ~| 




Slots 



Side baffles desirable 



6" freeboard in tank 



Use at least two slots. One at 
bottom of hood. 



Q = 150 cfm/ft"of bed (150LW) 
Slot velocity = 2000 fpm 
he= 1.78 VP S + 0.25 VP d 
Minimum duct velocity = 3500 fpm 
W not to exceed 36" 



Free board must be maintained to prevent material carryout. 



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FLUID IZ ED BEDS 



DATE 



12-90 FIGURE: VS-99-05 



10-174 



Industrial Ventilation 




Overflow pipe 



Q = 200 L x W 

h e = 1.78 slot VP S + 0.25 duct VPd 

Duct velocity = 2500 - 3000 fpm 



Sic 



size for 1500 - 2000 fpm 



AMERICAN CONFERENCE 

OF GOVERNMENTAL 
INDUSTRIAL HYGIENISTS 



OUTBOARD MOTOR TEST 



DATE 12-90 



figure ys-99-06 



Specific Operations 10-175 




Cross-section of 
table and hood 



Q total = Q, + Q 2 = 725 cfm 
Slot Velocity = 2000 fpm 
Duct Velocity = 2000 fpm 
h e = 1.78 VP + 0.5 VP d 



Note 1: Position of exhaust take-offs may vary due to table configuration 
place to assure same airflow distribution into slot. 



Reference 10.99.24 



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INDUSTRIAL HYGIENISTS 



MORTUARY TABLE 



DATE 



12-90 l F1GURE VS-99-07 



10-176 



Industrial Ventilation 



-> Q 



Slot hood 



Perforated plate 



>" 

9 



Slide damper -^ 
(full width of hood) 



Reference 10.99.25 



:j 



rjvA 



Q s _> 



^ xi ; L 



End View 



Y, 



-1 



Baf-'le 
Note 




45" 




^o ft °o l o y 




Slot hood 



Note 1 

Perforated piaie 

10% open area 

1/2" hole 



Tank drain 



Q s " 45 cfm/ft ~ of tank area 

Q, = 65 cfm/ft" of tank area 

Plenum velocity < 1/2 slot velocity 
Slot velocity = 2000 fpm 
h e = 1.78 VP d +,25 VP S 



Note 1 : Build perforated plate in 
sections for easy cleaning 
Allow freeboard height at least 9'' 

Note 2: Baffle to extend to ends of tank 

6" (minimum) above top of workpiece 

Note 3: Adjust slot to achieve 
Qg/CL flow distribution 



AMERICAN CONFERENCE 

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FURNITURE STRIPPING 
TANK 



DATE 



4-94 



figure VS-99-08 



Specific Operations 1 0-1 77 



TABLE 10.99.1. Grain Elevators, Feed Bfltlls, Flour frills (10 "- 23) 

The following data are offered as guides. Air flow rates can vary considerably depending on degree of enclosure, flow rate of material, and dustiness of the grain. 
Minimum duct velocity - 3500 fpm. Ventilation control is desirable for these operations to minimize the explosive characteristics of grain dusts and to preserve 
plant housekeeping standards. 



Operation 


Hood Design 


Air Volume 


Bag Loading 


VS-15-02, VS-15-03 
Booth VS-1 5,01 


As shown 

100 cfm/ft 2 open face area 


Belt Discharge 


To belt — VS-50-20 
To bin — VS-50-10 
To elevator — VS-50-1 , VS-50-20 


350 cfm/ft of belt width up to 200 fpm belt speed 
500 cfm/ft of belt width over 200 fpm belt speed 
Increase 1/3 if material drop is over 10 ft 


Bins 


Direct exhaust. Use taper. 


500 cfm/bin 


Bucket Elevator 


VS-50-01 


100 cfm/ft 2 cross-section 


Cleaning Machines 


Consult manufacturer 




Distributors 


Enclose discharge 200 fpm in-draft through 
enclosure openings. 


No. of Spouts 


Diameter of Spouts 


0-6 

6-12 

12-24 


Exhaust 
cfm 


6" 7" 8" 9" 


550 675 950 1250 
950 1250 1500 1900 
1500 1900 2250 2750 


Feed Grinders 


Consult manufacturer 




Floor Dump 


Booth 


200 cfm/ft 2 open face area 


Floor Sweep 




950 cfm in 4" x 8" opening 


Garner Bin 


Direct exhaust. Use taper. 


dm = 1.25 x bushels/min 


Mixers 


Ventilated cover 


Mixer Capacity 


Exhaust, cfm 


Up to 0.5 ton 
0.5 to 1.5 tons 

Over 1 ,5 tons 


300 
675 
950 


Percentage Feeders 


Enclosed conveyor 


200 cfm at each feeder 


Purifiers 


Enclosure 


30 - 40 cfm/ft 2 screen area 


Roll Stands 


Enclosure 


60 cfm/lineal ft 


Scales 


Enclosure 


Scale Capacity, Bushels 


Exhaust, cfm 


Up to 5 
6 to 10 
Over 11 


250 
400 
600 


Scale Hopper 


Direct exhaust. Use taper. 


cfm = 1.25 x bushels/min 


Screw Conveyor 


Direct exhaust. Use taper. 


200 cfm — ducts on 30 ft centers 


Sifters 


Enclosure 


200 cfm/compartment 


Track Sink 


Direct exhaust from hopper. Use taper. 


100 cfm/ft 2 grate area 


Tripper Car 


Belt discharge. VS-50-01, VS-50-10, 
VS-50-20. Spout ends — tapered connec- 
tion. Spillage — exhaust under head pulley. 


See "Belt Discharge" above. 200 cfm/ft 2 spout cross-section 90 cfm/ft 
belt width 



10-178 



Industrial Ventilation 



TABLE 10.99.2. i©ous Specific Operation Standards 



Operation or Industry 


Ventilation 


Minimum Design 
Duct Velocity (fpm) 


Reference No. and 
Remarks 


Type of Hood 


Air Flow or Capture Velocity 


Abrasive Wheel Mfg. 
Grading screen 
Barrels 

Grinding wheel dressing 


Enclosure — booth 
Close canopy 

Enclosure — booth 


50 fpm at face 
400 fpm at face 

400 fpm at face 


4000 
4000 

3000 


10.99.1 

Bbls. receive dust from 
cyclone 


Aluminum Furnaces 


Enclosure 


150-200 fpm through opening 


2000 


10.99.2 


Asbestos 
Bagging 
Carding 
Crushing 
Drilling of panels 

containing asbestos 
Dumping 
Grinding of brake shoes 

Hot press for brake shoes 
Mixing 

Preform Press 
Screening 

Spool winding 
Spinning and twisting 

Weaving 


Enclosure — booth 
Enclosure 
Enclosure 
Moveable hood 

Booth 
Enclosure 

Enclosure 
Booth 
Enclosure 
Enclosure 

Local Hoods 
Partial 

Canopy with baffles 


250 fpm through all openings 
1600 cfm/card 

150 fpm through all openings 
400 fpm capture velocity 

250 fpm face velocity 

400 fpm minimum capture at 

the tool rest 
250 fpm through all openings 
250 fpm face velocity 
250 fpm through all openings 
200 fpm through all openings 

but not less than 25 cfm/ft 2 

screen areas 
50 cfm/spool 
50 cfm/spool 

50 fpm through openings 


3500 
3500 
3500 
4500 

3500 
3500 

3500 
3500 
3500 
3000 

3500 
3500 

3500 


10,99.3, 10.99.4 

10.99.5 
10.99.6 

10.99.6 
10.99.6 

10.99.6 
10.99.6 
10.99.6 
10.99.3 

Hinged front panels and skirt, 
wet twisting preferred 
Wet weaving preferred 


Auto Parking Garage 


2-Level 


500 cfm/parking space 




10.99.7 


Ceramic 
Dry pan 
Dry Press 

Aerographing 
Spraying (lead glaze) 


Enclosure 
Local at die 
Local at die 
At supply bin 
Booth 
Booth 


200 fpm through all openings 

500 cfm 

500 cfm 

500 cfm 

100 fpm (face) 

400 fpm (face) 


3500 
3500 
3500 
3500 

2000 


10.99.8, 10.99.9, 10.99.10 
Automatic feed 
Manual feed 
Manual feed 


Coating Pans 
(pharmaceutical) 


Air flow into opening of pan 


100-150 fpm through 
opening 


3000 


10.99.8, 10.99.11 

If heated air supplied to pan, 

add volume of heated air to 

exhaust 


Cooling Tunnels (foundry) 


Enclosure 


75-100 cfm per running foot 
of enclosure 


— 


10.99.12, 10.99.13 


Core Knockout (manual) 


Large side-draft or semi- 
booth — exhaust near floor 


200-250 cfm/ft 2 dust 
producing working area 


3500 


10.99.12, 10.99.13, 
10.99.14 


Core Sanding (on lathe) 


Downdraft under work 


100 fpm at source 


3500 


10.99.15 


Crushers and Grinders 


Enclosure 


200 fpm through openings 


3500 


10.99.16 


Drilling (rocks) 


Special trap (see references) 


60 cfm — vertical 
(downward) work 
200 cfm — horizontal work 




10.99.17, 10.99.18, 

10.99.19 

May vary with size and speed 

of drill 


Forge (hand) 


Booth 


200 fpm at face 


1500 


10.99.2 


Outboard Motor Test Tank 


Side draft 


200 cfm/ft 2 of tank opening 


— 




Packaging Machines 


Booth 
Downdraft 
Complete enclosure 


50 -100 fpm at face 
95 -150 fpm down 
100 -400 fpm opening 


3000 

to 

4000 




Paper Machine 


Canopy 


200 - 300 fpm at face 


1500 


10.99.20, 10.99.21 



Specific Operations 10-179 



TABLE 10.99-2. IfflsceBlaneous Specific Operation Standards (eon't) 



Operation or Industry 


Ventilation 


Minimum Design 
Duct Velocity (fpm) 


Reference No. and 
Remarks 


Type of Hood 


Air Flow or Capture Velocity 


Quartz Fusing 


Booth on bench 


150-200 fpm at face 


— 


10.99.20, 10.99.21 


Rotary Blasting Table 


Enclosure 


500 fpm through all openings 
when in operation 


3500 




Silver Soldering 


Free hanging 


100 fpm at source 


2000 




Steam Kettles 


Canopy 


150 fpm at face 


2000 




Varnish Kettles 


Canopy 


200-250 fpm at face 


1500 


10.99.2, 10.99.10 


Wire Impregnating 


Covered tanks 


200 cfm/ft 2 of opening 




10.99.22 

Chlorinated naphthalenes & 

diphenyls 



BIBLIOGRAPHY 



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11-1 



11-2 



Industrial Ventilation 



34. Postman, B. F., Practical Application of Industrial Ex- 
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52. National Board of Fire Underwriters, Standard for Class 
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55. Schulte, H. F., Hyatt, E. C, and Smith, Jr., F. S., Exhaust 
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56. Mitchell, R. N., and Hyatt, E. C, Beryllium—Hazard 
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57. Hemeon, W. C. L., Plant and Process Ventilation, Indus- 
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58. Manufacturing Chemists' Association, Technical Data 

on Plastics, February, 1957. 

59. First, M. W., and Silverman, L., Airfoil Pitometer, In- 
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66. National Fire Protection Association, Ventilation of 
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February, 1940. 



Contents 



11-3 



70. Langley, M. Y, Harris, Jr., R.. L, Lee, D. H. K, Calcu- 
lation of Complex Radiant Heat Load from Surrounding 
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71. Thor Power Tool Company, Aurora, Illinois. 

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73. The Black and Decker Tool Company, Townson, MD. 

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75. Alexander, J. M, Croley, Jr., J. J., and Messick, R. R, 
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76. Trickier, C. J., Engineering Letter E-4R, New York 
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78. New York State Department of Labor, Division of In- 
dustrial Hygiene. 

79. Air Conditioning, Heating and Ventilating, Vol. 60, No. 

3, March, 1963. 

80. Hama, George, M.S, Frederick, W, Sc.D., and Mon- 
teith, H, M.S, Air Flow Requirements for Underground 
Parking Garages, American Industrial Hygiene Associa- 
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81. Feiner, B., and Kingsley, I., Ventilation of Industrial 
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cember, 1956, pp. 82-89. 

82. U. S. Air Force Technical Order 00-25-203, Standards 
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July, 1963. 

83. Federal Standard No. 209B, Clean Room and Work 
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85. Constance, J. A., Estimating Air Friction in Triangular 
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86. McKarns, J.S., Confer, R. G., and Brief, R. S, Estimat- 
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102. Airflow Developments Ltd., Lancaster Rd., High Wy- 
combe, Bucks, England. 

103. Heating and Cooling for Man and Industry, American 
Industrial Hygiene Association, 1969. 



11-4 



Industrial Ventilation 



104. ASHRAE Guide & Data Book, American Society of 
Heating, Refrigeration and Air Conditioning Engineers, 
1961, p 243. 

105. F. W. Dwyer Company, Michigan City, IN. 

106. HPAC Data Sheet, How to Design Drain Type Stacks, 
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107. Air Movement and Control Association, Inc., 30 W. 
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108. Adapted from U. S. Dept, of Labor, Occupational Safety 
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1 1 0. Hama, George M., and Butler, Jr., Kerrel E, Ventilation 
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1 12. Hama, George, and Bonkowski, K. J., Ventilation Re- 
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1 22. Hatch, T.F, Economy in the Design of Exhaust Systems. 

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128. Ventilation and Air Contracting Contractors Association 
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133. Leith, David, Gibson, Dwight D, and First, Melvin W, 
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135. Rajhans, G. S, and Bragg, G. M, Engineering Aspects 
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136. Caplan, K. J, and Knutson, G.W, Laboratory Fume 
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137. Caplan, K.J, and Knutson, G.W, Laboratory Fume 
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138. Sheet Metal and Air Conditioning Contractors' National 
Assn., Inc., Round Industrial Duct Construction Stand- 



Contents 



11-5 



ards, 1977, 8224 Old Courthouse Rd, Tysons Corner, 
Vienna, VA 22180. 

139. Sheet Metal and Air Conditioning Contractors' National 
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143. Hughes, R. T., Unpublished data. 

144. American Society of Heating, Refrigerating and Air 
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145. Air Moving and Control Association, Inc., AMCA Pub- 
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146. American Society of Mechanical Engineers, Fluid Me- 
ters - Their Theory and Applications, 1959. 

147. Farant, J. P., McKinnon, D. L., and McKenna, T. A., 
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149. U. S. Dept. of Health, Education and Welfare, PHS, 
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150. U. S. Air Force, AFOSH Standard 16L2. 

151. U. S. Dept. of Health and Human Services, PHS, CDC, 
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Revised Criteria, 1986. 

152. American Welding Society, (AWS D 1.1 -72), P. O. Box 
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153. Gibson, N., Lloyd, F. C, and Perry, G. R., Fire Hazards 
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Thermite Reaction, Symposium Series No. 25, Insn. 
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154. Hughes, R.T., and Amendola, A. A., Recirculating Ex- 
haust Air: Guides, Design Parameters and Mathematical 
Modeling, Plant Engineering, March 18, 1982. 



155. U. S. Dept. of Health, Education and Welfare (NIOSH), 
The Recirculation of Industrial Exhaust Air - Sympo- 
sium Proceedings, Pub. No. 78-141, 1978. 

156. American Conference of Governmental Industrial Hy- 
gienists, Air Sampling Instruments for Evaluation of 
Atmospheric Contaminants, 6th Ed., Chapters U and V, 
Cincinnati, OH, 1983. 

1 57. Baturin, V.V., Fundamentals Industrial Ventilation, Per- 
gamon Press, NY, 1 972. 

158. U. S. Public Health Service, Air Pollution Engineering 
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159. U. S. Public Health Service, Air Pollution Engineering 

Manual, Publication No. 999-AP-40, 1973. 

160. HampI, V., and Johnson, O.E.,: Control of Wood Dust 
from Horizontal Belt Sanding, American Industrial Hy- 
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161. HampI, V., Johnston, O.E., and Murdock, D.M.: Appli- 
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162. Air Movement and Control Association, Inc., AMCA 
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sity Dr., Arlington Heights, IL 60004. 

163. FED-STD-209D, Federal Standard, Cleanroom and 
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164. MIL-F-51477(EA), Specification Filters, Particulate, 
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165. IES-RP-CC-002-86, HEPA Filters, Institute of Environ- 
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166. IES-RP-CC-006-84, Testing Clean Rooms, Institute of 
Environmental Sciences, 940 East Northwest Highway, 
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167. Cooper, T. C, Control Technology for a Dry Chemical 
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168. Gressel, M. G., Fischback, T.J., Workstation Design 
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trial Hygiene, 4:227-233, 1989. 

169. American Foundrymen's Society, Inc., Foundry Venti- 
lation Manual, Des Plaines, IL, 1985. 



11-6 



Industrial Ventilation 



170. Mortimer, V. D, Kerder, S. L. and O'Brien, D. M, 
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171. Hama, G. M, Ventilation for Fumigation Booths, Air 
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172. National Fire Protection Association Standard, NFPA 
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174. Knutson, G. W, Effect of Slot Position on Laboratory 
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175. Caplan, K. J. and Knutson, G. W., Influence of Room 
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176. American Society of Heating, Refrigerating and Air 
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177. National Sanitation Foundation, Standard 49, Class II 
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178. U. S. Air Force Technical Order 00-25-203, Standards 
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Contents 



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APPENDICES 



Appendix A Threshold Limit Values for Chemical Substances in the Work 

Environment with Intended Changes for 1997 12-3 

Appendix B Physical Constants/Conversion Factors 12-21 



12-1 



Appendix A 



12-3 



APPENDIX A 



1997 

Threshold Limit Values 

for Chemical Substances 

in the 

Work Environment 

Adopted by ACGIH 
with Intended Changes 



INTRODUCTION TO THE 
CHEMICAL SUBSTANCES 

Threshold Limit Values (TLVs) refer to a