HANDBOOK OF AIR CONDITIONING SYSTEM DESIGN OTHER McGR’AW-HILL .‘\MERIC.\N ~NSTITI:TE OF P H Y S I C S A~~IXIC.I\N Socrli~v 01: MIXIIANICAL Engineering Tables Metals Engineering-Design Metals Engineering-Processes Metals Properties AMERICAN SOCICTY OF TOOL AND HANDBOOKS OF INTEREST . American Institute of Physics Handbook ENGINLXRS . ASME Handbooks: MANUFACTURING ENGINEERS: Die Design Handbook Manufacturing Planning and Estimating Handbook Handbook of Fixture Design Tool Engineers Handbook ARCHITECTURAL RECORD . Time-Saver Standards BEEn<AN . Industrial Power Systems Handbook BRADY . Materials Handbook BURIKGTON . Handbook of Mathematical Tables and Formulas BURINCTON AND MAY . Handbook of Probability and Statistics with Tables CARROLL Industrial Instrument Servicing Handbook CONDON AND ODISHAW . Handbook of Physics CONSIDIN~ . Process Instruments and Controls Handbook CONSIDINE A N D R o s s Handbook of Applied Instrumentation CROCKER Piping Handbook D A V I S . Handbook of Applied Hydraulics DUDLEY Gear Handbook EMERICK . Heating Handbook FACTORY MUTUAL ENGINEERING DIVISION . Handbook of Industrial Loss Prevention FL&XE . Handbook of Engineering Mechanics HARRIS . Handbook of Noise Control HARRIS AND CREDE . Shock and Vibration Handbook HEYEL . The Foreman’s Handbook HUSKEY AND KORN . Computer Handbook JURAN . Quality Control Handbook KALLEN . Handbook of Instrumentation and Controls KING AND BRATER . Handbook of Hydraulics K N O W L T O N . Standard Handbook for Electrical Engineers KOELLE . Handbook of Astronautical Engineering KORN AND KORN . Mathematical Handbook for Scientists and Engineers LASSER . Business Management Handbook LAUGHNER AND HARGAN . Handbook of Fastening and Joining of Metal Parts LEGRAND . The New American Machinists’ Handbook MACHOL System Engineering Handbook MAGILL, HOLDEN, AND ACKLEY . Air Poilution Handbook MAXAS . National Plumbing Code Handbook MASTELL . Engineering Materials Handbook MARKS AND BAUXEISTER . Mechanical Engineers’ Handbook MAYNARD . Industrial Engineering Handbook MAYNARD . Top Management Handbook M O R R O W ’ Maintenance Engineering Handbook PERRY . Chemical Engineers’ Handbook PERRY . Engineering Manual ROSSNACEL Handbook of Rigging ROTHBART Mechanical Design and Systems Handbook SHAND Glass Engineering Handbook STASIAR Plant Engineering Handbook STREETER Handbook of Fluid Dynamics TOIJLOUKIAN Retrieval Guide to Thermophysical Properties Research Literature HANDBOOK OF AIR CONDITIONING SYSTEM DESIGN Carrier Air Conditioning Company . & I e M New York C GRAW-HILL Sara Francisco BOOK Toronto COMPANY London Sydney HANDBOOK apyright @ OF AIR 1965 by CONDITIONING McGraw-Hill, Inc. SYSTEM All Rights DESIGN Reserved. @ 1960, 1963, 1964, 1965 by Carrier Corporation. Printed in the United States of -4merica. This book, or parts thereof, may not be reproduced in any form without permission of the publishers. Lib7ary of Congress Catalog Card Number 65-17650. ISBN 07-010090-X 15 16 17 18 HDHO 8543210 . PREFACE The Handbook of Air Conditioning System Design is the first complete practical guide to the design of air conditioning systems. It embodies all the knowledge and experience gained over the past fifty years by the pioneer in the field, Carrier Air Conditioning Company. This handbook is tailored to the specific needs of the man responsible for the details of design, and, therefore, the foremost consideration has been the requirements of the consulting engineer. In fact, many of the concepts embody the up-to-date thinking of consulting engineers. If any one word best describes this work, it is the word “practical.” l l l l l l It is usable at all educational levels. I It provides practical data for professional designers who need optimum solutions on a day-to-day basis. It bridges the gap between air conditioning texts and manufacturers’ product catalogs. It provides proved system design techniques and assures quality of application with minimum service requirements. It provides guidance in simplified form. It provides a reference source employing the best techniques of indexing and format. This Handbook of Air Conditioning System Design is a companion piece to manufacturers’ product literature. Together the handbook and product literature make up a complete engineer’s manual. Those using this book for study will benefit from clear applicable examples presented in each of the engineering sections. In summary, this Handbook of Air Conditioning System Design is a quick reference for those actively engaged in designing’ air conditioning systems, a teaching work for those studying air conditioning system design, and a refresher for those engineers with wide experience in the field. * * * Grateful appreciation is hereby extended to those hundreds of Carrier engineers who generously contributed to the total body of knowledge herein, and to those consulting engineers, mechanica contractors, and architects who so willingly and enthusiastically contributed their experience to this project. Carrier Air Conditioning Company TENAGA EWBANK PRt$CE, LIBRARY CONTENTS p re f ace . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . P a r t 1 ..--..:LOAD--.-- ESTIMATING . .‘. I-l .:: . . . . . . . . . . . . . Building Survey and Load Estimate . . . . . . . . Design Conditions . . . . . . . . . . . . . . . . . . . . . . . Heat Storage, Diversity and Stratification . Solar Heat Gain thru Glass . . . . . . . . . . . . . . Heat and Water Vapor Flow thru Structures Infiltration and Ventilation . . . . . . . . . . . . . . Internal and System Heat Gain . . . . . . . . . . . Applied Psychrometrics .................. 1. 2. 3. 4. 5. 6. 7. 8. V . . . . . . . . . . . . . . . . . . . . . . . . l-l l-9 l-25 l-41 l-59 l-89 1-99 l-115 1 ,Part lib_ . 2.~ AIR DISTRIBUTION :. . . . . . . . . . . . . . . . . . . . 1. Air Handling Apparatus . . . . . . . . . . . . . . . . . . . . . . . . . 2. Air Duct Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3. Room Air Distribution . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-l . . . . . . 2-l 2-17 2-65 DESIGN .. :: . . . . . . . . . . . . . . . . . . . . . . . . . . ...... 3-1 Piping Design-General . . . . . . . . . . . . . . . . . . . . . . . . . Water Piping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Refrigerant Piping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Steam Piping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ...... ...... . 3-l 3-19 3-43 3-81 . . . : Part 4. REFRIGERANTS; BRINES; OIL.‘?: : :. :. . L.-.-.-..~ I 1. Refrigerants . . . . . . . . . . . . . 2. Brines . . . .‘... . . . . . . . . . . . . . . . . . . . . . . . 3. Refrigeration Oils . . . . . . . . ....... 4-l ....... ....... ....... 4-l 4-23 4-55 --Part :- -. 1. 2. 3. 4. 3 : PIPING .+ Part 5. WATER CONDITIONING .. 1. %ter Conditioning-General 2. Scale and Deposit Control . . . 3. Corrosion Control . . . . . . . 4. Slime and Algae Control . . . . 5. Water Conditioning Systems . 6. Definitions . . . . . . . . . . . . . . . . . . . . . . . . . . , . . . . . . . . . 5-1 . . . . . . . . . . . 5-l 5-l 1 5-19 5-27 5-31 5-47 ‘._ . . . . . . . . . . ’ I Part 6 . AIR HANDLING EQUIPMENT ........................... G-l ~ I. Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2. Air Coriclitioning Apparatus . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3. Unitary Equiprncnt . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4. :l(~rcssory Erlliiprncnt . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-I G-17 G--l 5 G-5 1 Part 7. REFRIGERATION EQUIPMENT . . . . . . . . . . . . . . . . . . . . . . . . 7-l 1. Reciprocating Refrigeration Mnrhine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2 . Centrifugal Rcfrigcrntion Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3. Absorption Refrigeration Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . -1. CombinaGon r\hsorption-Centrifugal System . . . . . . . . . . . . . . . . . . . . . . . . 5. ITent Rejection Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-1 7-21 7-33 7-47 7-55 Part 8. AUXILIARY EQUIPMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-l 1. 2. 3. 4. Centrifugal Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Motors and Motor Controls . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Boilers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Miscellaneous Drives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-l 8-17 8-5 1 8-61 Part 9. SYSTEMS AND APPLICATIONS . . . . . . . . . . . . . . . . . . . . . . . . . 9-l 1. Systems and Applications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-l Part 10. ALL-AIR SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1. 2. 3. 4. 5. 6. Convention& Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Constant Volume Induction System 1 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Multi-zone Unit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Dual-duct System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Variable Volume; Constant Temperature System . . . . . . . . . . . . . . . . . . . . . Dual Conduit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-I 10-l 10-9 10-17 lo-25 10-35 10-41 \ Part 11. AIR-WATER SYSTEMS .................................... 11-l 1. Induction Unit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2. Primary Air Fan-coil System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-l 1 l-23 Part 12. WATER AND DX SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12-1 1. Fan-coil Unit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2. DX Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12-1 12-11 Index .............................................................. I-l / k HANDBOOK OF AIR CONDITIONING SYSTEM DESIGN l-l Par-t 1 LOAD ESTIMATING CHAPTER 1. BUILDING SURVEY AND LOAD ESTIMATE The primary function of air conditioning is to maintain conditions that are (1) conducive to human comfort, or (2) required by a product, or process within a space. To perform this function, equipment of the proper capacity must be installed and controlled throughout the year. The equipment capacity is determined by the actual instantaneous peak load requirements; type of control is deterI xd by, the conditions to be maintained during ped~ and partial load. Generally, it is impossible to measure either the actual peak or the partial load in any given space; these loads must be estimated. It is for this purpose that the data contained in Part 1 has been compiled. Before the load can be estimated, it is imperative that a comprehensive suruey De made to assure accurate evaluation of the load components. If the building facilities and the actual instantaneous load within a given mass of the building are carefully studied, an economical equipment selection and system design can result, and smooth, trouble free performance is then possible. The heat gain or loss is the amount of heat instantaneously coming into or going out of the space. The actual load is defined as that amount of heat which is instantaneously added or removed by the enllipment. The instantaneous heat gain and the ’‘s. 11 load on the equipment will rarely be equal, because of the thermal inertia or storage effect of the building structures surrounding a conditioned space. Chapters 2, 4, 5, 6, and 7 contain the data from which the instantaneous heat gain or loss is estimated. Chapter 3 provides the data and procedure for applying storage factors to the appropriate heat gains to result in the actual load. Chapter 8 provides the bridge between the load estimate and the equipment selection. It furnishes the procedure for establishing the criteria to fulfill the conditions required by a given project. The basis of the data and its use, with examples, are included in each chapter with the tables and charts; also an explanation of how each of the heat gains and the loads manifest themselves. BUILDING SURVEY SPACE CHARACTERISTICS AND HEAT LOAD SOURCES An accurate survey of the load components of the space to be air conditioned is a basic requirement for a realistic estimate of cooling and heating loads. The completeness and accuracy of this survey is the very foundation of the estimate, and its importance can not be overemphasized. Mechanical and archi- tectural drawings, complete field sketches and, in some cases, photographs of important aspects are part of a good survey. The following physical aspects must be considered: 1. Orientation of building - Location of the ’ space to be air conditioned with respect to: a) Compass points -sun and wind effects. b) Nearby permanent structures - shading effects. c) Reflective surfaces - water, sand, parking lots, etc. 2. Use of space(s) - Office, hospital, department store, specialty shop, machine shop, factory, assembly plant, etc. 3. Physical dimensions of space(s) - Length, width, and height. 4. Ceiling height - Floor to Hoor height, Hoor to ceiling, clearance between suspended ceiling and beams. 5. Columns and beams - Size, depth, also knee braces. 6. Construction materials - Materials and thickness of walls, roof, ceiling, floors and partitions, and their relative position in the structure. 7. Surrounding conditions - Exterior color of walls and roof, shaded by adjacent building or sunlit. Attic spaces - unvented or vented, gravity or forced ventilation. Surrounding spaces conditioned or unconditioned - temperature of non-conditioned adjacent spaces, such as furnace or boiler room, and kitchens. Floor on ground, crawl space, basement. 8. Windows - Size and location, wood or metal l-2 I’ 9. 10. 11. 12. 13. 14. sash, single or double hung. Type of glass single or multipane. Type of shadi-ng device. Dimensions of reveals and overhangs. Doors - Location, type, size, and frequency of use. Stairways, elevators, and escalators- Location, temperature of space if open to unconditioned area. Horsepower of machinery, ventilated or not. People - Number, duration of occupancy, nature of activity, any special concentration. At times, it is required to estimate the number of people on the basis of square feet per person, or on average traffic. Lighting - Wattage at peak. Type - incandescent, fluorescent, recessed, exposed. If the lights are recessed, the type of air flow over the lights, exhaust, return or supply, should be anticipated. At times, it is required to estimate the wattage on a basis of watts per sq ft, due to lack of exact information. M o t o r s - Location, nameplate and brake horsepower, and usage. The latter is of great significance and should be carefully evaluated. The power input to electric motors is not necessarily equal to the rated horsepower divided by the motor efficiency. Frequently these motors may be operating under a continuous overload, or may be operating at less than rated capacity. It is always advisable to measure the power input wherever possible. This is especially important in estimates for industrial installations where the motor machine load is normally a major portion of the cooling load. APPl iances, business machines, electronic equipment - Location, rated wattage, steam or gas consumption, hooded or unhooded, exhaust air quantity installed or required, and usage. Greater accuracy may be obtained by measuring the power or gas input during times of peak loading. The regular service meters may often be used for this purpose, provided power or gas consumption not contributing to the room heat gain can be segregated. Avoid pyramiding the heat gains from various appliances and business machines. For example, a toaster or a waffle iron may not be used during the evening, or the fry kettle may not be used during morning, or not all business !ART 1. LOAD ESTIMATING machines in a given space may be used at the same time. Electronic equipment often requires individual air conditioning. The manufacturer’s recommendation for temperature and humidity variation must be followed, and these requirements are often quite stringent. 15. Ventilation - Cfm per person, cfm per sq Et, scheduled ventilation (agreement with purchaser), see Chapter 6. Excessive smoking or odors, code requirements. Exhaust fans-type, size, speed, cfm delivery. 16. Thermal storage - Includes system operating schedule (12, 16 or 24 hours per day) specifically during peak outdoor conditions, permissible temperature swing in space during a design day, rugs on floor, nature of surface materials enclosing the space (see Chapter 3). 17. Continuous or intermittent operation Whether system be required to operate every business day during cooling season, or only occasionally, such as churches and ballrooms. If intermittent operation, determine duration of time available for precooling or pulldown. LOCATION OF EQUIPMENT AND SERVICES , The building survey should also include information which enables the engineer to select equipment location, and plan the air and water distribution systems. The following is a guide to obtaining this information: 1. Available spaces - Location of all stairwells, elevator shafts, abandoned smokestacks, pipe shafts, dumbwaiter shafts, etc., and spaces for air handling apparatus, refrigeration machines, cooling towers, pumps, and services (also see Item 5). 2. Possible obstructions - Locations 0E all electrical conduits, piping lines, and other obstructions or interferences that may be in the way of the duct system. 3. Location of all jire walls and partitions Requiring fire dampers (also see Item Ih). 4. Location of outdoor air intakes - In reference to street, other buildings, wind direction, dirt, and short-circuiting of unwanted contaminants. 5. Power seruice - Location, capacity, current limitations, voltage, phases and cycle, 3 or 4 wire; how additional power (if required) may be brought in ,and where. 6. Water seroice - Location, size of lines, ca- . CHAP-I-El< I . ljacity, pressure, maximum temperature. 7. SLetlm .semice - Location, size, capacity, ternperature, pressure, type of return system. 8. I<efrigerntion, brine o r chilled waler (if f u r nished by customer)-Type of systen!, capacity, temperature, gpm, pressure. 9. Arrhitect7lral characteristics of spnce - For selection of outlets that will blend into the s p a c e design. 10. Existing air conveying equipment wnd ducts For possible reuse. 11. Drains - Location and capacity, sewage disposal. 12. Control facilities - Compressed air source and pressure, electrical. 13. Foundation and support - Requirements and facilities, strength of building. So&d and vibration control requirements _ I. Kelation of refrigeration and air handling apparatus location to critical areas. 15. Accessibility fo? moving equipment to the final location - Elevators, stairways, doors, accessibility from street. 16. Codes, local and national- Governing wiring, drainage, water supply, venting’of refrigeration, construction of refrigeration and air handling apparatus rooms, ductwork, fire dampers, and ventilation of buildings in general and apparatus rooms in particular. AIR CONDITIONING LOAD ESTIMATE The air conditioning load is estimated to provide the basis for selecting the conditioning equipment. It must take into account the heat coming into the sy-ace from outdoors on a design day, as well as the . being generated within the space, A design day is defined as: 1. A day on which the dry- and wet-bulb temperatures are peaking simultaneously (Chapter 2, “Design 1-3 l\LJII,DING SlJKVEY .\ND LO>\11 ESl‘IM.\I‘E Conditions”). 2. A day when there is little or no haze in the air to reduce the solar heat (Chapter 4, “Solar Heat Gain Thru Glass”). 3. All of the internal loads are normal (Chapter 7, “Internal and System Heat Gain”). The time of peak load can usually be established by inspection, although, in some cases, estimates must be made for several different times of the day. Actually, the situation of having all of the loads peaking at the same time will very rarely occur. To be realistic, various diversity factors must be applied 10 sonic of the lmtl compo11ents; rcl’er to cl/crfiter 3, “ljeat .~torage, ljiwrsity, and .\‘t1.nti/ir/ltion.” The infiltratiotl ;1nt1 ventilation air quantities arc estim;rtecl as dcscribetl in CilnpleY h. I;ig. I illustrates 211 air conditioning lmtl cstimatc form ;tntl is dcsigllctl to permit systematic load eVaI- uation. This form contains the references identified to the particular chapters of data and tables rcquirctl to estimate the various load components. OUTDOOR LOADS The loads from outdoors consist of: 1. Tile slln rays entering windows - Table 15, pages 44-49, and Table 16, pnge 52, provide data from which the solar lieat gain through glass is estimated. The solar heat gain is usually reduced by means of shading devices on the inside or outside of the windows; factors are contained in Table 16. In addition to this reduction, all or part of the window may be shaded by reveals, overhangs, and by adjacent buildings. Chart I, page 57, and Table 18, page 55, provide an , easy means of determining how much the window is shaded at a given time. A large portion of the solar heat gain is radiant and will be partially stored as described in Chapter 3. Tables 7 thru 11, pages 30-34, provide the storage factors to be applied to solar heat gains in order to arrive at the actual cooling load imposed on the air conditioning equipment. These storage factors are applied to peak solar heat gains obtained from Table 6, page 29, with overall factors from Table 16, page 52. 2. The sun rays striking the walls and roof - These, in conjunction with the high outdoor air temperature, cause heat to flow into the space. Tables 19 and 20, pages 62 and 63, provide equivalent temperature differences for sunlit and shaded walls and roofs. Tables 21, 22, 23, 24, 25, 27, and 28, pages 66-72, provide the transmission coefficients or rates of heat flow for avariety of roof andwall constructions. 3. The air temperature outside the conditioned space - A higher ambient temperature causes heat to flow thru the windows, partitions, and floors. Tables 25 and 26, pages 69 and 70, and Tables 29 and 30, pages 73 and 74, provide the transmission coefficients. The temperature differences used to estimate the heat flow thru these structures are contained in the notes after each table. l-4 l’.\Rl‘ I . LO,\11 I:S’I‘I,\I.\~I‘IN(~ HAP I TABLE REFERENCES AREA OR ITEM G 3 & 4 G L A S ~S S T O R A G OF OPERATION CONDITEE.-FEouroooR(OA, j Tau l - 3 1 6 . 1 7 ! 2 SOFTX, TE L & TRANS. So WALL S”FTX WALL So SQ P 6 2 I x: X TEL SaFlX! ROOF-SHADE” 20 4 , XI II XI ~63 PP 66.69 TBLS / 27.28 OUTDOOR AIR X i TBL 45 I -CIY,PLRSOH p97 i x p-Cr*iSa FT PEOPLE 90 FT , Yl I iTOLS 2 1 2 2 . 1 / A23,24 “R 251 ~ TBL1.9 FT ‘EMT,. ATION .5 ROOF FT X 1 FT X 1 / So F T X ’ JOOF-SUN 15 GAIN-WALLS WALL WALL 5 FOR ‘1: id roi 15 CORR I S o F T k’ -“-; : --’PP 44-49 ,<I PP 4 4 . 4 9 S o FT X’ GLASS --1 WITHOUT GLASS -1S T O R A G E SKYLIGHT ! SOLAR ,““SS 2 TBL” 9.lO”Rll 1 FT X, wL9-3++ PP 5 2 . 5 4 SQ E TABLE REFERENCES :STIMATE FACTOR / SOLAR GAIN-GLASS S O FT U T0~56&7 6 Wlill L A S S SUN G AIN OR TEMP DIFF. ! CluANTlrv iAF EF =tEF CFM SWNCIHG REYOLVlNc. 6 ALL GAIN-EXCEPT GLASS So PARTlTlON FT X Sa F -CEILI N G WALLS D O O R S - PEOPLE OPEN Doons-oooas IWTIL. ‘RITI”I( ‘i EXHAUST TBLS 4 6 . 4 7 : P 98 FAN FEET x TBL CRACK 7 1 . 7 2 NOTE X NOTFS. T So Fr X FLOOR Sa INFILTRATiON NOTE I’ FT ROOF 1 X ! X PP - CFM TBL 3 3 4 4 P 9 5 CFY/FT = - NQTF 1 AIR ‘X PP 73 EFFECTIVE sEFw*;“E;T ESHF 74 HEAT G-42 p g2 -!%ZE-CFMo 3 TBL 6 5 _ ADP E F F E C T I V__._... E Room SENS. EAT _ . ..~ HEFFECTIVE ROOM T O T A L HE A T C H A R T. FI G 3 3 P 1 1 6 = P 145. O R P S Y C H INDICATED ADP = SELECTED ADP = ----F DEHUMIDIFIED INTERNAL n APPARATUS APPARATUS DEWPOINT 1.08 X THRU P 76 p; TBL 29 :R 30 73.74 OUTDOOR T;pLS629Sj&6 PP 6%70X NO T E S X CFM 4 (i = ~ n VENTlLATlON CFM ,NFlLTRATl”N TRANS. = ~ AIR __ _-.-F QUANTITY P 121 (I--BF) X CT,,- F - Tm-F) =-F TpBppL;;4&3 PEOPLE X PEOPLE HP WATTS POWER oil K W x 3,4 A P P LI A N C E S . TBLS E TC . ADD,T,ONAL HE 50-52 PP 1 0 1 - 1 0 3 GAINS TBLS 5 4 . 5 7 A T So STORAGE X $;;,” kj;k;; 8 PP 107-109 X (TEMP ’ %iM 3UTLETI X TBL 5 3 P 10: x TBLS 12.14.49 35 ,R la!- _ LIGHTS sue SWING? T EMP . DIFF. EFFECTIVE 1.08 I F ACTOR ROOM _ - CFMc RISE SENS. HE A T = -FIRM--(IUTLET CFM n. SUPPLY FT X i TBL 1 4 P 38 1X (- T;$ RO %iY TOTAL HE A T TCYP ~$8, I ” 1.08 SAFETY SENS. F ~-1.08 X TOTAL SUE ROOM x O M AIR SENS H E X F QUANTITY A T DESIRED . pz.7 CFM: O,FF % P 113 R~JOM SE.NSlBLE H E A T W RESULTING ENT G LVG CONDITIONS AT APPARATUS O U T D O O R AIR NOTE 3 CFM E F F E C T I V E Y NOTE R O O M LATENT INFILTRATION PEOPLE PEOPLE STEAM P 121 BF ETC . ADDlTloNAL HE A T TBLS 5 0 . 5 2 GAINS TRANS. SUPPLY PP x TBLS 1 4 . 4 8 : X t,tOO LEAKAGE NOTE 3 LOSS E N S I B L E : NOTES NOTE 3 CORR B X TBL 40 P Q SUB T LATENT % ____~ GRAB CFM x N CFM XN”TE R O O M T,,,-TLDB- C H A R T : TEWB- F. T LWS-F 1. p;My~~-B”~~ (“B) T E M P E R A T U R E Form O T A L 2. ~;&MOISTURE CO 3. NORMALLY . US E &CFM VENTILATION” EYER. WHEN ,NF,LTRAT,“N 1s T o B E DETERM,NE .-CFM OUTDOOR AIR.” AIR O T E EZO. 4. W H E N I N F I L T R A T I O N IS N O T 1” BE O F F S E T . A N ” “ C F M VENTILATION IS L E S S T H A N “ C F M INFILTRATION.” T H E N T H E E X C E S S INFILTRATI” ,s Acco”NT~oFOR HERE LATENT HEAT H E A T n HEAT 1 F X cl--~121BF) X Z”a,~s Without HEAT N T E N T P 1 2 1 BF x a,68 x T O T A L I.08 X (1 --P 12lEF) Y 0 . 6 8 TEL 6 0 S U B TO T A L FL, p ‘Q -+ KY:,,“, pdj, G R A N D Carrier Masthead = TBLS E L O W P 110 ROOM RETURN C H A R T 3 RETURN DUFT P 1 1 0 DUCT P 112 Oh + HEAT GA,” % + LEAK. GAlN With CT,,-F-TT,,-F) NOTES X ROOM CFM X NOTE 2 OUTDOOR LATENT: X (T em---F - Tm--F) = PP 38.100 % E F F E C T I V E S PSYCH. GII/LI1 x 0 . 6 8 2 X NOTE ZGI,!-8 EFFECTIVE 8 T,,-F-t&+’ T,,-F+?BF X FRO” NOTE 101.103 P 113 DUCT I R W DIFFERENCE FROM TOP “F ESTIMAT FROM T “F TBL 58 P 1 0 9 S0.f~ F ACTOR O”TD”“R A EDB 1.08 LDB X --__. SAFETY X H E A T LB/HI x 1050 P 107 APPLIANCES. VAPOR X HEAT CFM Noi~ 4 1F S E N S I B L E T O T A L H E A T Cwrier M”sthe”dForm n E5024. FIG . 1 - AIR CONDITIONING LOAD ESTIMATE . (GR,LB) DI F F E R E N C E O P ESTIMAT FOR “ C F M OUTDOOR A I R . ” H o w OFFSET. R E F E R T o PA G E 9 2 T CHAPTER I. I1UILDINC; SlJKVliY .\NI> LO,\D 4. The clir 71~ipr)~ ~MS~~W - i\ Iiighcr wpoi pressure surrounding contlitionctl space ca~scs water v a p o r t o liow tilt-u the l)uilding niaterials. This load is signilicant only in low dcw1)oint al~l~lications. Tile data required t o estinlate this load is containccl in Table JO, p(/ge Sf. In comlort applications, this load is neglected. 5 . The wield blorui,lg trgtrilrst ~1 side of tlte I~rliltl- irjg-1Vintl c;iuscs the outdoor air that is higlicr in tempcraturc ant1 moisture content to infiltrate thru the cracks around the doors and windows, resulting in locali~ctl scnsiblc ant1 latent heat gains. All or part of this infiltration may be ofFset by air being introduced thru the apparatus for ventilation purposes. Cl/npler h contains the estimating data. Ozltdoor niy usunlly wquired 1-s ESI‘IM;\TE /(II. ueiltilntiotl pz+oses - Outdoor air is usually necessary to flush out the space and keep the odor level down. This ventilation air imposes a cooling and dehumidifying load on the apparatus bec a u s e t h e h e a t and/or moisture must be removed. i\Iost air conditioning equipment permits some outdoor air to bypass the cooling surface (see Chapter 8). This bypassed outdoor air becomes a load within the conditioned space, similar to infiltration; instead of coming thru a crack around the window, it enters the room thru the supply air duct. The amount of bypassed outdoor air depends on the type of equipment used as outlined in Chnptey 8. Table $5, page 97, provides the data from which the ventilation requirements for most comfort applications can be estimated. ’ “ie foregoing is that portion of the load on the a; .anditioning equipment that originates outside the space and is common to all applications. INTERNAL LOADS Chapter 7 contains the data required to estimate the heat gain from most items that generate heat within the conditioned space. The internal load, or heat generated within the space, depends on the character of the application. Proper diversity and usage factor should be applied to all internal loads. AS with the solar heat gain, some of the internal gains consist of radiant heat which is partially stored (as described in Chnpter 3), thus reducing the load to be impressed on the air conditioning equipment. Generally, internal heat gains consist of some or all of the following items: 1. People - The human body thru metabolism c VI generates heat within itself ant1 rcleascs it by radiation, convection, and evaporation I’rom the su~l;~cc, mtl by coI1vection arid evaporation in the respiratory tract. The amount of heat generated 2nd rcleasetl depends on surrountling tcniperaturc and on the activity level ol’ the person, as listed in Ttrble fS, f-“1ge 100. 2. Liglct.9 - Illi~niinaiits convert electrical power into light ant1 licat (refer to Cllclptel- 7). Some of the heat is radiant and is partially stored (see C/upter 3). 3 . Appliuuces - Rcstaiirants, hospitals, laboratories, and some specialty shops (beauty shops) have electrical, gas, or steam appliances which release heat into the space. Tables 50 th?-u 52, pnges 101-103, list the recommendecl heat gain values for most appliances when not hooded. If a positive exhaust hood is used with the appliances, the heat gain is reduced. 4. Electric calcfilnting machines - Kefer to manufacturer’s data to evaluate the heat gain from electric calculating machines. Normally, not all of the machines would be in use simultaneously, and, therefore, a usage or diversity factor should be applied to the full load heat gain. The machines may also be hooded, or partially cooled internally, to reduce the load on the air conditioning system. 5. Electric motors - Electric motors are a signifi- cant load in industrial applications and should be thoroughly analyzed with respect to operating time and capacity before estimating the load (see Item 13 under “Space Characteristics and Heat Load Sozlrces”). It is frequently possible to actually measure this load in existing applications, and should be so done where possible. Table 53, page 105, provides data for estimating the heat gain from electric motors. 6. Hot pipes and tanks - Steam or hot water pipes running thru the air conditioned space, or hot water tanks in the space, add heat. In many industrial applications, tanks are open to the air, causing water to evaporate into the space. Tables 54 thou 58, pages 107-109 provide data for estimating the heat gain from these sources. 7. Miscellaneous sowces - There may be other sources of heat and moisture gain within a space, such as escaping steam (industrial cleaning devices, pressing machines, etc.), absorption oE water by hygroscopic materials (paper, textiles, etc.); see Chapter 7. In addition to the heat gains from the indoor and outdoor sources, the air conditioning equipment and duct system gain or lose hcnt.. The fans and pumps requirctl to distribute the air or water thru the system add heat; heat is also added to supply and return air ducts running thru warmer or hot spaces; cold air may leak out of the supply duct and hot air may leak into the return duct. The procedure for estimating the heat gains from these sources in percentage of room sensible load, room latent load, and grand total heat load is contained in Clmrt 3, ;b~ge I IO, and Tables 59 rind 60, pa.ges 111-113. HEATING LOAD ESTIMATE The heating load evaluation is the foundation for selecting the heating equipment. Normally, the ating load is estimated for the winter design temperatures (Chnpter 2) usually occurring at night; therefore, no credit is taken for the heat given off by internal sources (people, lights, etc.). This estimate must take into account the heat loss thru the building structure surrounding the spaces and the heat required to offset the outdoor air which may infiltrate and/or may be required for ventilation. Chnpter 5 contains the transmission coefficients and procedures for determining heat loss. Chapter 6 contains the data for estimating the infiltration air quantities. Fig. 2 illustrates a heating estimate form for calculating the heat loss in a building structure. Another factor that may be considered in the evaluation of the heating load is temperature swing. Capacity requirements may be reduced when the temperature within the space is allowed to drop a few degrees during periods of design load. This, of -ourse, applies to continuous operation only. Table pnge 20, provides recommended inside design conditions for various applications, and Table 13, page 37, contains the data for estimating the possible capacity reduction when operating in this manner. The practice of drastically lowering the temperature to 50 F db or 55 F db when the building is unoccupied precludes the selection of equipment based on such capacity reduction. Although this type of operation may be effective in realizing fuel economy, additional equipment capacity is required for pickup. In fact, it may be desirable to provide the additional capacity, even if continuous operation is contcmplatctl, because of pickup required after forcctl shutdown. It is, thcrcl’ore, evident that the use of storage in reducing the heating load for the purpose of equipment selection should be applied with care. HIGH ALTITUDE LOAD CALCULATIONS Since air conditioning load calculations are based on pounds of air necessary to handle a load, a decrease in density means an increase in cfm required to satisfy the given sensible load. The weiglit of air required to meet the latent load is decreased because of the higher latent load capacity of the air at higher altitudes (greater gr per lb per degree tlitference in dewpoint temperature). For the same dry-bulb and percent relative humidity, the wctbulb temperature decreases (except at saturation) as the elevation above sea level increases. The following adjustments are required for high altitude load calculations (see Chapter 8, Table 66, page 148): Design room air moisture content must be adjusted to the required elevation. . Standard load estimating methods and forms are used for load calculations, except that the factors affecting the calculations of volume and sensible and latent heat of air must be multiplied by the relative density at the particular elevation. Because of the increased moisture content of the air, the effective sensible heat factor must be corrected. EQUIPMENT SELECTION After the load is evaluated, the equipment must be selected with capacity sufficient to offset this load. The air supplied to the space must be of the proper conditions to satisfy both the sensible and latent loads estimated. Chapter S, “Applied PsychrometTics ,” provides procedures and examples for determining the criteria From which the air conditioning equipment is selected (air quantity, apparatus dewpoint, etc.). . (:II.\I”TTR I. I:1711.1>1NC; SllRVI<Y ,\ND I>O.\D HEATING CONDITIONS l-7 ESTIM,\'I‘I~ TEMPERATURE OF AIR ENTERING UNIT TOTAL TRANSMISSION LOSS FORM El0 F1c.2 - HEATING LOAD ESTIMATE l-9 CHAPTER 2. DESIGN CONDITIONS This chapter presents the data from which the outdoor design conditions are established for various localities and inside design conditions for various applications. The design conditions established determine the heat content of the air, both outdoor and inside. They directly affect the load on the air conditioning equipment by influencing the transmission of heat across the exterior structure and the difference in heat content between the outdoor and inside air. For further details, refer to OUTDOOR DESIGN CONDITIONS - SUMMER AND WINTER The outdoor design conditions listed in Table 1 are the industry accepted design conditions as published in AR1 Std. 530-56 and the 1958 ASHAE :. The conditions, as listed, permit a choice of G ouLdoor dry-bulb and wet-bulb temperatures for different types of applications as outlined below. B.NORMAL DESIGN CONDITIONS - SUMMER Normal design conditions are recommended for use with comfort and industrial cooling applications where it is occasionally permissible to exceed the design room conditions. These outdoor design conditions are the simultaneously occurring dry-bulb and wet-bulb temperatures and moisture content, which can be expected to be exceeded a few times a year for short periods. The dry-bulb is exceeded more frequently than the wet-bulb temperature, and usually when the wet-bulb is lower than design. When cooling and dehumidification (dehydration) are performed separately with these types of applications, use the normal design dry-bulb tem- perature for selecting the sensible cooling apparatus; use a moisture content corresponding to the normal design wet-bulb temperature and 80% rh for selecting the dehumidifier (dehydrator). Daily range is the average difference between the high and low dry-bulb temperatures for a 24-hr period on a design day. This range varies with local climate conditions. A, MAXIMUM DESIGN CONDITIONS-SUMMER Maximum summer design conditions are recommended for laboratories and industrial applications where exceeding the room design conditions for even short periods of time can be detrimental to a product or process. The maximum design dry-bulb and wet-bulb temperatures are simultaneous peaks (not individual peaks). The moisture content is an individual peak, and is listed only for use in the selection of separate cooling and dehumidifying systems for closely controlled spaces. Each of these conditions can be expected to be exceeded no more than 3 hours in a normal summer. NORMAL DESIGN CONDITIONS - WINTER Normal winter design conditions are recommended for use with all comfort and industrial heating applications. The outdoor dry-bulb temperature can be expected to go below the listed temperatures a few times a year, normally during the early morning hours. The annual degree days listed are the sum of all the days in the year on which the daily mean temperature falls below 65 F db, times the number of degrees between 65 F db and the daily mean temperature. ’ l-10 PART I. LO/\D ES?‘IM,\?‘ING TABLE l-OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER STATE AND CIT Y ALABAMA Anniston Birmingham Mobile Montgomery ARIZONA’ MAXIMUM DESIGN COND.-SUMMER July at 3:00 PM DI r y I Bulb (F) DryBulb (F) 95 95 95 95 90 105 65 76 81 94 26 30 TlJCSOll 105 100 110 72 70 78 - 77 85 93 30 95 95 76 78 - 104.5 117.5 I6 I6 25 ARKANSAS Fort Smith little Rock CALIFORNIA Bakersfield El Centro Eureka Fresll0 Laguno Beach Long Beach Los Angeles Oakland Montague Pasadena Red Bluff Sacramento San Bernadine San Diego Son Francisco San Jose Williams 105 110 90 105 70 78 65 74 54 94 52 76 90 90 85 70 70 65 78 78 60 9 5 100 100 70 70 72 70 62 73 I8 I05 85 85 72 68 65 65 75 60 IO 17 91 70 76.5 Durango 95 95 Fort Collins Grand Junction Pueblo 95 95 CONNECTICUT Bridgeport Hartford New Haven Waterbury 64 65 65 65 - 75 95.9 82 70 103.0 94 94 68 88 74 102 68 86.2 I4 16 14 117.5 Apalachicola Jacksonville Key West Miami 95 95 98 91 80 78 78 79 I31 117.5 I 12.5 I31 Pensacola Tampa Tallahorsee 95 95 78 117.5 117.5 80 - 78.4 24 25 78 7.7 SW 5.4 E 6,894 1,108 5.0 w 5.2 NW 2,376 4,853 I46 1036 6.7 N 7.0 E 6.0 NW 30 25 4758 2403 7.0 N 8.0 NW 35 30 1391 6.0 SW 74.4 30 2680 35 35 1596 3137 25 2823 - 94 95 0 0 0 - 1 5 82 7.0 w 2.0 w mtiuda deg) 8.3 E 8.3 NW 448 324 7.3 5.4 NW 499 43 132 287 6.4 NE 84 155.6 99 82 150.5 92 81 150.5 34 34 31 32 32 35 33 10 47 261 I7 34 34 34 38 2.635, 42 34 40 39 7.2 SE 305 I16 6.3 NW 7.5 N 26 I7 34 33 38 100 86 37 39 7.0 s 7.5 s 5.22 I 6,558 40 37 5613 5558 6.0 SE 4.4 NW 7.9 NW 4,587 4,770 41 39 38 6113 5880 7.0 s 7.0 s 8.7 NW 9.4 N 9 58 23 41 42 41 42 N W 134 40 7.8 NW 72 39 IO.0 S W 99 733 694 IO 293 5.0 w 0 COLUMBIA Washington 7242 1441 3226 3009 99.3 62 63 95 OF 110 89.4 117.5 -10 25 8.0 N 9.9 N 7.5 NW 25 68 78 1 83 145.5 - 99 95 2806 261 I 1566 207 I 25 -10 30 2.5 99 102 99 DIST. 126.9 60 70 75 75 75 Wilmington 78 155.6 110 95 93 95 DELAWARE 2 ilavathan 4bove Se0 Level (ft) I 40 - COLORADO Denver r 103 103 DATA Avg. Velocity and DryAnnual Prevailing Direction ~(pr/lb o f I B u l b I D e g r e e I Summer Winter I c i,; a i r ) 1 IFI 1 Doyr 8 30 WIND ontentt Bulb i VI 90 II3 1 lolrture -Dry- 19 I9 12 I5 Flagstaff Phoenix Winslow Yuma ‘_ Vetiulb IF) - I NORMAL DESIGN COND. WINTER 0 5.0 s FLORIDA 78 95 ‘Correspondr to dry-bulb and wet-l 3 temperatures listed, and is corrected for altitude of city. tcorrespondr to peek dewpoint temperature, corrected for altitude. 25 25 45 35 20 30 25 - 1252 II85 59 I85 1281 571 1463 - 5 . 0 SW, 8.0 SW 9.0 SE 7.0 SE 8.4 9.0 NE 10.6 NE 10.1 E 23 I8 23 II 30 30 25 26 6.0 NE 10.9 N 8.6 NE N 408 25 31 68 28 30 (:I I.\I”I’I~:I< 11. I>I<,SI(;N 1-11 (:ONI)I’I‘IONS TABLE 1 -OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (Contd) NORMAL DESIGN COND.-SUMMER July at 3:00 PM STATE AND CITY MAXIMUM DESIGN COND.-SUMMER July at 3:00 PM NORMAL DESIGN COND. WINTER WIND Elevation Above Sea Level Winter i vtt) DATA GEORGIA Atlanta Augusta Brunswick Il.7 NW 6.5 NW 975 195 NW 9.5 NW 408 42 Columbus Macon Savannah 6.7 IDAHO Boise Lewiston Pocotello fwin Falls ‘- ‘NOIS 3iro Lhicogo Danville Moline Peoria Springfield INDIANA Evansville Fort Wayne Indianapolis 95 95 95 1 I ii 54.5 44 65 65 65 / ;i 9698 96 31 28 28 61 / ‘it” 1 I9 76 76 77 71 92.6 100 1 I04 / 80 1 140.6 -10 5678 5 - 5 -IO 5109 674 1 --I-- 5.0 N W 9.1 SE 4.1 E 8.9 SE / -I: / :;:: / 10.0 763 4,468 9.8 12.0 SW NW / NE 8.3 s I I.9 NW I9 20 18 102 100 99 82 150.5 0 -IO - 1 0 4410 7.0 S W 6232 8.0 S W 5458 9.0 S W - 5 117.5 I23 117.5 I8 I8 - 5 -I5 -I5 - 2 0 102 6252 6375 6820 78 78 I25 132 20 21 ‘opeko ,Vichita 100 100 78 75 109.5 98 I9 21 95 78 1 17.5 22 99 I I I LOUISIANA Alexandria New Orleans Shreveport I 5425 Ill -IO -IO - 1 5 110 -IO -IO 5075 106' 79 i I 126.9 38 41 40 42 40 42 42 43 I I 00 I 637 1,111 I 4417 4792 I 7.0 S W 1201 i 39 38 13.3 SW 9.8 SW I 42 41 43 43 39 38 39 5069 4644 46 43 42 41 41 40 6.0 S W 10.0 s 11.0 s 44 594 602 603 8.2 SW 11.5 N W 95 95 33 33 32 37 42 40 + Fort Dodge Keokuk Sioux City Waterloo KANSAS Concordia Dodge City Salino 34 34 31 319 594 124 IOWA Cedar Rapids Davenport Des Moines Dubuque KENTUCKY Lexington Louisville 2,705 W I03 106 103 117.5 99 104.5 South Bend Terre Haute 109 Latitude (deeI N 8.6 N 8.8 SE 989 459 38 38 89 9 197 32 30 33 45 45 44 Belfast Eortport Millinocket Presque Isle Portland Rumford I 90 I 70 I 781131 I / I 5: I - 2 0 I *Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city. tcorresponds to peak dewpoint temperature, corrected for altitude. 8445 I 7.0 s 12.6 W / NW 10.4 NW 100 44 45 46 L 47 1 47 44 44 ’ l-12 I’AKT I . LO:\11 IiS’I‘lhl,\l‘lNG TABLE l-OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.) l- STATE AND CITY MAXIMUM DESIGN COND.-SUMMER July at 3:00 PM DryBulb (F) MARYLAND Baltimore Cambridge Cumberland _ _ .-Frederick Frostburg Salirbury WetBulb (F) 95 78 95 75 DryBulb (F) 117.5 DryBulb (F) 18 MICHIGAN Alpeno Big Rapids Detroit Escano ba 75 75 95 75 95 75 -I 9> 9.5 95 Ludington Marquette Saginaw Soult Ste Marie 1 ‘3 -10 0 -10 96 -10 -15 0 z+- 93 93 i 75 95 1 75 17 17 19 101 20 99 20 + L j ‘11 93 73 95 75 ~__ St. Cloud St. Paul / I 79 4487 90 99 96 103 20 20 -10 -10 96 I Summer / Winter 6.0 SW 1 8.2 NW NW I I 19 17 -2.5 -25 -20 102 t 79 131.1 + MISSOURI Columbia Kansas City Kirksville St. Louis St. Joseph Springfield 79 78 100 100 78 76 95 / 70 18 98 20 104 82 L__- / 66 ‘55.6 -I 71 56 Gc 49 20 97 1 77.4 c I temperature, corrected for altitude. 199 625 42 42 42 42 615 45 I 8278 6560 8777 t WI 6702 c 111.0 SW NW 10.0 SW 12.0 SW /9.5 NW 7149 t ‘1.9 w , 10.6 NW / 8.9 SE 1,128 7975 9.0 s 7213 7.7 SE 6.3 N 8.3 8.9 SW 10.3 NW SW 1 9.0s 8.0 S 111.8s 9.3 NW ‘0.9 SE j 12.4 W 84’6 t i 7930 8032 7591 7604 47 * 5070 4962 -25 -20 -20 -30 ;a w 9723 7966 4596 5596 4569 j , 9.8 SW 9307 c 6’9 8.0 w /12.’ Lzw 7458 8745 -10 -IO *Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city. peak dewpoint 9.0 SW 5.0 SW 4.0 SW 6.0 SW 0 0 5 0 40 40 40 W 2330 2069 2 2 3 2 39 39 39 -__.-~ 6743 15 10 10 - 14 . - 70 Latitude kg) c 0 -IO -10 108 66 95 / 109 ‘08 78 -- I 96 c 90 83 -25 -20 Above Sea Level (ft) 5936 c -20 103 95 95 tion Avg. Velocity and Prevailing Direction W NW -10 -I5 -IO -I5 -10 -10 - 5 -IO 98 t MlSSlSSlPPl Jackson teridian ucksburg 135.9 c / 93 73 95 75 MINNESOTA Alexandria Duluth Minneapolis 102 102 0 - 5 -IO 0 99 75 75 I EleVa- DATA L 0 5 -/‘02 17 _- Flint Grand Rapids Kalamazoo Lansing tCorresponds to AntWa Degree Days WIND L-.--. 104 .ew Bedford /lymouth Springfield Worcester Helena Kalispell Miles City Misroula B u l b (gr/lb o f (F) dry air) r - 5 - 5 10 MASSACHUSETTS Amherst Boston Fall River -__ Fitchburg Lowell Nantucket MONTANA Billings Butte Great Falls Hovre Moisture Net- contentt NORMAL XSIGN C O N D . WINTER 5.6 S E 3’6 410 226 32 32 32 ( : I l.\l”1’l~:I~ 1-13 2. I)lL‘5l(;N C:ONI)I~I‘IONS TABLE 1 -OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.) NORMAL DESIGN COND.-SUMMER July at 3:00 PM STATE AVG. DAILY RANGE MAXIMUM DESIGN CON&-SUMMER July at 3:00 PM NORMAL DESIGN COND. WINTER WIND DATA ElevaLatitud kg) NEBRASKA Grand Island 41 41 42 41 41 43 North Platte Omaha Valentine York NEVADA tar Vegas RetlO Tonopah ‘%memucca h HAMPSHIRE Berlin Concord 115 95 75 65 40 41 95 65 40 90 73 i 102 66 66.9 20 - 5 5 - 1 5 562 I 5812 6357 7.0 S W 7.0 S W s 6.0 W 9.9 SE 8.1 NE 1,882 4,493 5.42 1 4,293 I I I 95 j 14 I I I I / I 17’ 36 40 30 42 45 43 43 43 43 I 5 Camden East Oronae + Newark Cit; Jersey Paterson Sandy Hook Trenton 95 95 95 75 75 95 99 99 78 a 14 14 117.5 99 95 14 82 145.5 0 81 140.6 0 0 5015 13.0 S W 15.8 N W 10.0 S W 5500 96 13.0 S W 13.0 S W 9.0 S W NW 17.1 N W 16.1 10.9 N W 125 30 173 10 10 56 39 41 40 41 41 41 41 41 40 35 32 36 77 _ lens F a l l s Ithaca Jamestown Lake Placid Roche&r Schenectady Syracuse I I I I 126.9 - 5 - 2 5 - 1 0 - 1 5 1 + I 1 I 95 93 93 / I I 75 75 75 I 102 I 18 I 95 6925 8305 12.0 S W 1 17.1 w 8.0 10.5 NW W 604 458 43 42 43 43 43 43 42 42 - 44 41 45 i 43 43 43 43 43 44 36 35 37 36 34 ‘Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city. tcorresponds to peak dewpoint temperature, corrected for altitude. i TABLE 1 -OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.) NORMAL DESIGN COND.-SUMMER MAXIMUM DESIGN COND.-SUMMER AVG. DAILY NORMAL DESIGN COND. STATE AND CITY NORTH DAKOTA Birmarck Devils Lake c Fargo Grand Forks Williston OHIO Akron Cincinnati Cleveland Columbus Dayton Lima Sandusky Toledo Youngstown OKLAHOMA Ardmore Bartlesville Oklahoma City Tulsa 95 95 95 75 78 75 99 117.5 99 95 95 76 78 104.5 123 19 22 19 23 23 95 95 95 75 75 75 99 99 99 19 19 106 101 81 79 - 5 0 0 145.5 135.9 95 99 99 4990 6144 7.0 SW 11.0 s 8.5 SW 14.7 SW 104 553 651 41 39 42 -10 0 - 5 5506 5412 9.0 SW 8.0 SW 11.6 SW 11.1 SW 724900 40 40 41 0 -IO 6095 6269 10.0 SW II.0 12.1 SW 608 589 1,186 I 101 101 77 77 108 101.5 21 I 104 106 42 42 41 i 79 t OREGON Baker Eugene Medford Pendleton Portland Roseburg Wamic 1 / I -5 -15 / 7197 I ( I 1 I 366 / 44 PENNSYLVANIA Altoono Bethlehem Erie Harrisburg New Castle Oil City Philadelphia Pittsburgh Reading Scranton Warren ~ Williamsport 95 95 95 75 78 75 99 117.5 105 18 14 14 97 98 95 95 75 75 99 99 14 95 RHODE ISLAND Block Island Pawtucket Providence 95 93 93 75 75 75 99 102 102 SOUTH CAROLINA Charleston Columbia Greenville 95 95 95 78 75 76 117.5 99 104.5 17 17 17 98 SOUTH DAKOTA Huron Rapid City Sioux Falls 95 95 95 75 70 75 106 05 99 19 22 20 I06 103 / *Corresponds to dry-bulb ond wet-bulb temperatures listed, tcorrespondr to peak dewpoint 79 126.9 0 0 0 - 5 -15 - 5 4739 5430 10.0 SW 9.OSW 5232 6218 6.0 SW 15 IO 10 1066 2488 3059 42 40 40 11.0 NW 11.6 W 26 1,248 9.0 7.6 SW NW NW 311 746 525 40 41 41 42 10.5 SW 8.0 SW 8.4 9 401 902 33 34 35 14 14 a2 155.6 76 71 and is corrected for altitude of city. temperature, corrected for altitude. 10.0 SW 7.0 N E , ,- _ ----_-.., _. l-15 (:f I \l”l~I~:l< 2. I)LSI~;N (:ONl)I-I‘IoNS TABLE l-OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.) STATE AND CITY TENNESSEE Chattanooga Johnson CiLy -Knoxville Memphis Nashville TEXAS Abilene Amarillo Austin Brownsville Corpus Chrirti Dallas Del Rio El Po’so Fort Worth Galveston Houston Palestine Port Arthur San Antonio UTAH Modem Logan Ogden Salt take City VERMONT Bennington Burlington Rutland DryBulb (F) Net. Bulb WI 95 76 95 95 95 75 78 78 100 100 100 -.- 95 95 100 100 100 100 95 95 100 95 100 l-I-onirteunrat* MAXIMUM DESIGN COND.-SUMMER July at 3100 PM gr/lb of dry air) DryBulb VI DryBulb (F) 104.5 I8 98 I7 103.5 100 117.5 18 103 117.5 17 98 ---i-- 74 72 78 80 80 78 78 69 78 80 80 78 79 78 124 109.5 95.. 65 95 65 90 90 73 73 WetBulb (F) NORMAL DESIGN COND. WINTER Vloisture htentt :gr/lb of dry air) Avg. Velocity and Annual tI Degree Days Summer I 6.0 S W 79 83 7.7 NW ElSVotlon Above Sea Level (ft) / 689 6.0 S W 7.0 S W 8.0 w Lotitude (deeI 35 - 36 36 35 36 - 93 101 96 I 105 80 101 72 100 81 I9 102 83 66 25 97 61 25 102 2573 4196 1679 628 965 2367 1501 2532 2355 II74 1315 2068 1532 1435 9.0 s 11.0 s 9.0 13.0 8.0 10.0 9.0 10.0 9.0 8.0 32 35 31 26 28 33 29 32 33 29 30 32 30 21 SE SE s SE E s s S 7.8 SE 68 - 91 8.0 s 11.6 S 4,446 4,222 38 42 41 41 308 VIRGINIA Cape Henry Lynchburg Norfolk Richmond Roancke vVASHINGTON North Head Seattle Spokane Tacoma Tatoosh Island Walla Wallo Wenatchee Yakima WEST VIRGINIA Bluefield Charleston Elkinr Huntington Martinsburg Porkersburg Wheeling 95 95 95 95 95 78 75 78 78 76 99 95 90 85 85 93 85 65 65 65 64 86 106 95 90 95 65 65 65 37 37 37 38 38 11.0 s 6.0 SW 5367 70 68 7.0 N 7.0 SW I05 W 102 76 98 *Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city. tcorrerponds to peak dewpoint temperature, corrected for altitude. 16.1 9.8 SE 6.2 S W 8.0 18.9 5.4 s 4928 4.0 SE 48 48 47 48 46 48 47 603 37 38 39 38 39 39 40 , TABLE l-OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.) STATE AND CITY NORMAL DESIGN COND.-SUMMER July at 3:00 PM Moisture Wet- content* Bulb (gr/lb of (F) dry air) DryBulb (F) WISCONSIN Ashland Eau Claire Green Bay L o Crosre Madison Milwaukee WYOMING Carper Cheyenne Lander Sheridan AVG. DAILY RANGE 95 95 95 95 75 75 75 75 99 99 103.5 99 95 95 65 65 68.5 66 DryBulb (F) MAXIMUM DESIGN COND.-SUMMER July at 3:00 PM Moisture vi.,- contentt Bulb (gr/lb of (F) dry air) DryBulb IF) 14 / 99 -r 17 100 18 96 , 14 99 28 28 79 131.1 161.2 83 102 NORMAL DESIGN COND. WINTER DryBulb (F) WIND DATA Avg. Velocity and Annual Prevailing Direction Degree Days Summer Winter -20 - 2 0 - 2 0 - 2 5 -I5 - 1 5 7931 742 I 7405 7079 8.0 6.0 8.0 9.0 -20 -15 -18 - 3 0 7536 8243 7239 9.0 s 5.0 S W 5.0 N W S S SW SW 10.5 9.3 10.1 12.1 Elevation Above Sea Level (ft) SW NW SW s I NW w 885 589 673 938 619 SW 13.3 N W 3.9 4.9 N W 5,321 6,139 5,448 3,773 PROVINCE AND CITY Lethbridge MCMUrKly Medicine Hat BRITISH COLUMBIA Estevon Point Fort Nelson Penticton Prince George Prince Rupert VClllCOtJVC3~ I NEW BRUNSWICK Campbellton Fredericton Moncton Saint John NEWFOUNDLAND Corner Brook Gander Goose Bay Saint Johns 43 42 44 45 90 90 - I 66 71 68 77 - 90 65 80 67 90 71 90 75 I 9520 9 3 10320 9 2 ’ 8650 2 5 8650 2 3 3 3 4 3 i 78 - 83.5 107 / I 9.7 8.9 10.1 7.6 7.9 15.0 9.1 9.0 3 4 3 2 2 2 9 9 9500 6910 5230 5410 10930 16810 10630 -11 - 6 8830 -8 -3 8700 8380 -I - 3 - 2 6 1 9210 9440 12140 8780 NORTHWEST TERRITORIES Aklovik Fort Norman Frobirher Resolute Yellowknife *Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city. tcorrerponds to peak dewpoint temperature, corrected for altitude. 3,540 2,219 2,190 3,018 1,216 2,365 51 54 55 50 57 50 7.2 8.0 7.7 12.3 20 1,230 1,121 2,218 170 22 228 49 59 50 54 54 49 48 14.7 6.4 12.0 1,200 115 894 786 50 59 54 50 9.2 42 164 48 46 13.8 14.9 248 119 46 45 17.2 10.3 19.3 40 482 144 463 , 49 49 53 48 30 300 68 56 682 69 65 9.9 3.7 l7 - 3 8 - 6 - 3 2 8 11 15 Victoria MANITOBA Brandon Churchill The Pas Winnipeg 42 45 45 44 43 43 . CANADA ALBERTA Calgary Edmonton Grand Prairie Latitude be) 11.5 7.9 9.2 . 62 1-17 (:fl.\l”l’l-I< II. I)ICSI(iN (:ONI)I’I’IONS TABLE 1 -OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.) NORMAL DESIGN COND.-SUMMER July at 3:00 PM CANADA PROVINCE AND CITY COND.-SUMMER DryBulb VI Direction f------I ----I NORMAL DESIGN COND. WINTER DryBulb VI WIND DATA Avg. Velocity and ‘reVallill Annual Degree Day5 Summer Winter Elevation Above SW Level (ft) Latitude idee) NOVA SCOTIA 90 Halifax Sydney Yarmouth ONTARIO Fort William Hamilton Kapuskodng Kingston Kitchener &don North Boy OttOWCl Peterborough Souix Lookout Sudbury Timminr Toronto Windsor Soult Ste. Marie PRINCE ISLAND - 75 107 7570 8220 7520 Ia +--I 90 75 107 93 75 102 .93 75 102 - 2 4 0 - 3 0 -11 - 3 -I - 2 0 -IS -11 - 3 3 - 1 7 - 2 6 0 3 ,’ ,.-’ 9.6 13.1 13.5 83 197 136 45 46 44 T 8.4 9.6 10.0 644 303 752 11.9 11.3 912 1,210 48 43 49 44 43 43 46 45 44 50 47 - 7810 7380 8830 9.6 8.9 7020 8.1 8380 a.7 =---I+ ,48 43 42 47 + EDWARD Charlottetown ,:’ /’ 10350 6890 11790 9.9 6.6 - 3 H QUEBEC Arvido Knob Lake Mont Joli Montreal Port Harrison Quebec City Seven Islands Sherbrooke Three Rivers 90 90 75 107 90 75 107 -~__ 1 - 1 9 - 4 0 -II - 9 - 3 9 - 1 2 - 2 0 - 1 2 - 1 3 10440 -I 8130 9.9 9070 9.0 74 8.2 375 46 55 48 46 58 47 50 45 46 8610 t SASKATCHEWAN Prince Albert Regina Soskatoon Swift Current 11.3 90 90 90 90 71 70 70 92.5 81 -41 - 3 4 - 3 7 - 3 3 11430 10770 10960 9660 YUKON TERRITORY DaV4?.0n Whitehorse *Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city. tcorrespondr to peak dewpaint temperature, corrected for altitude. 12.4 10.7 4.9 12.1 9.7 14.6 15040 1 8.7 1,414 1,884 1,645 2,677 53 50 52 50 1,062 2,289 64 61 . 1-18 I’.‘IlR-I‘ CORRECTIONS TO OUTDOOR DESIGN CONDITIONS FOR TIME OF DAY AND TIME OF YEAR The normal design conditions for summer, listed in Table 1, are applicable to the month of July at about 3:00 P.M. Frequently, the design conditions at other times of the day and other months of the year must be known. I. LOAD ESTIhI.\?‘ING Solution: Normal design conditions for New York in July at 3:00 p.m. are 95 F db, 75 F wb (Table I). Daily range in New York City is 14 F db. Yearly range in New York City = 95 - 0 = 95 F db. Correction for time of day (12 noon) from Table 2: I Dry-bulb = -5 F Wet-bulb = -1 F Correction for time of year (October) from Table 3: Table 2 lists the approximate corrections on the dry-bulb and wet-bulb temperatures from 8 a.m. to 12 p.m. based on the average daily raqge. The drybulb corrections are based on analysis of weather data, and the wet-bulb corrections assume a relat i v e l y c o n s t a n t dewpoint t h r o u g h o u t t h e 24-hr period. Dry4,ulb = -16 F Wet-bulb = - 8 F Design conditions at 12 noon in October (approximate) : Dry-bulb = 95 - 5 - 16 = 74 F Wet-bulb = 75 - 1 - 8 = 6 6 F INSIDE COMFORT DESIGN CONDITIONS SUMMER Ta6le 3, lists the approximate corrections of the dry-bulb and wet-bulb temperatures from March to November, based on the yearly range in dry-bulb I 3erature (summer normal design dry-bulb minus w,,lter nprmal design dry-bulb temperature). These corrections are based on analysis of weather data and are applicable only to the cooling load estimate. The inside design conditions listed in Table 4 are recommended for types of applications listed. These conditions are based on experience gathered from many applications, substantiated by ASHAE tests. The optimum or deluxe conditions are chosen where costs are not of prime importance and for comfort applications in localities having summer outdoor design dry-bulb temperatures of 90 F or less. Since all of the loads (sun, lights, people, outdoor air, etc.) do not peak simultaneously for any prolonged periods, it may be uneconomical to de’sign for the optimum conditions. Example J - Corrections to Design Conditions Given: A comfort application in New York City. Find: The approximate dry-bull, and wet-bulb temperatures at 12:00 noon in October. TABLE 2-CORRECTIONS IN OUTDOOR DESIGN TEMPERATURES FOR TIME OF DAY (For Cooling Load Estimptes) DAILY YGE OF ’ .-MPERATURE* (F) 10 15 my i../ 25 30 35 40 45 / . _ /,;UN .,, DRYOR WETBULB B Dry-Bulb W&-Bulb Dry-Bulb Wet&lb - 9 - 2 -12 - 3 Dry-Bulb Wet-Bulb Dry-Bulbs Wet-Bulb Dry-Bulb Wet-Bulb -14 - 4 -I6 - 4 -18 - 5 -10 - 3 -IO - 3 -12 - 3 - 5 - 1 - 5 -1 - 6 - 1 Dry-Bulb Wet-Bulb Dry-Bulb Wet-Bulb Dry-Bulb Wet-Bulb. -21 - 6 -24 - 7 -26 - 7 -14 - 4 -16 - 4 -17 - 5 - 7 - 2 -8 - 2 - AM PM 10 / - 12 7 2 9 2 TIME -. I -5- 1 -5, ->1 \ -a - 2 ,....+:-I 0 - 1 0 - 1 0 - 1 0 - 1 0 - 1 0 - 1 0 - 2 0 I 3 -A 0 0 0 0’ 00 0 0 0 9 d 1) 0 0 0 0 4 I I -1 0 - 1 0 -1 0 6 I - 2 - 1 - 2 -1 -3-1 -5 - 1 -6 - 1 -7 -2 -1 0 -I 0 - 3 1 4 1 -8 - 2 -1Q - 3 - a - 2 -10 - 3 -11 - 3 -13 - 3 -15 - 4 - 1 0 -1 0 - 2 -I - 6 - 1 - 7 - 2 -8 - 2 -12 - 3 -14 - 4 -16 - 4 -18 - 5 -21 - 6 -24 - 0 * T h e d a i l y range of dry-bulb temperature is the difference between the highest and lowest dry-bulb temperature during o d e s i g n day. (See Table I for the value of daily range for a particular city). Equation: Outdoor design temperature at ony time = Outdoor design temperature from 10 8 I I 24-hour Table I + Correction from above table. - 12 I - 9 - 2 -14 - 4 -16 - 4 -18 - 5 -21 - 6 -24 - 7 -20 - 9 -31 -10 period on a typical (:fI.\I”I’b:I~ 2. l-19 I)l3I(;N (:oKl)I~l‘I<)NS TABLE 3-CORRECTIONS IN OUTDOOR DESIGN CONDITIONS FOR TIME OF YEAR Cooling (For, YEARLY RANGE OF TEMPERATURE(Fj* 120 TIME March I 90 / OF YEAR April Drv-Bulb - 3 0 ---IS 95 Estimates) DRY- OR WET-BULB _ii 100 Load ;;j ~ ‘-;i zj+; __~~.-I_ - 4 20 -I1 10 -5 - 2 _I- Wet-Bulb - 3 0 -15 Dry-Bulb Wet-Bulb Dry-Bulb Wet-Bulb - 2 9 - - 1 4 - 2 9 -14 -19 -10 -19-p -IO -10 -5 - Dry-Bulb Wet-Bulb -29 -14 -19 -10 -IO -5 - 2 9 -14 -19 -10 -9 -5 a5 Dry-Bulb Wet-Bulb 80 Dry-Bulb Wet-Bulb 75 Dry-Bulb Wet-Bulb 70 Dry-Bulb Wet-Bulb 65 Dry-Bulb Wet-Bulb 60 Dry-Bulb Wet-Bulb 55 Dry-Bulb Wet-Bulb 50 Dry-Bulb Wet-Bulb -- -20 -IO -11 -5 - 4 - 2 -IO -5 3 2 3 2 0 0 0 0 e~wIqge~ -I_. - 0 I -6 0 -3 0 -b 0 -3 $ ;i -I7 - B -331--16 -17 - 2 9 -14 2; - a - 2 7 -14 - 2 7 -I4 - a 0 0 0 0 1; 0 0 0 0 -6 -3 -6 -3 - 1 6 -3 - 2 0 0 0 0 -16 - B -26 -14 - 3 - 2 0 0 0 0 -5 -3 - 1 6 -25 - 1 4 j - a ,, ‘Yearly range of temperature is the difference between the wmmer and winter normal desigq dry-bulb temperatures (Table 1). Equation: Outdoor design temperature = Outdoor design temperature from Table 1 + Correct& from above table. The commercial inside design conditions are recommended for general comfort air conditioning ap plications. Since a majority of people are comforta!2’ t 75 F or 76 F db and around 45y0 to 50% rh, tht --Ltermostat is set to these temperatures, and these conditions are maintained under partial loads. As the peak loading occurs (outdoor peak dry-bulb and wet-bulb temperatures, 100% sun, all people and lights, etc.), the temperature in the space rises to thi design point, usually 78 F db. If the temperature in the conditioned space is forced to rise, heat will be stored in the building mass. Refer to Chapter 3, “Heat Storage, Diversity and Stmtification,” for a more complete discussion of heat storage. With summer cooling, the temperature swing used in the calculation of storage is the difference between the design temperature and the normal thermostat setting. The range of summer inside design conditions is provided to allow for the most economical selection of equipment. Applications of inherently high sen- sible heat factor (relatively small latent load) usually result in the most economical equipment selection if the higher dry-bulb temperatures and lower relative humidities are used. Applications with low sensible heat factors (high latent load) usually result in more economical equipnient selection if the lower drybulb temperatures and higher relative humidities are used. INSIDE COMFORT DESIGN CONDITIONSWINTER For winter season operation, the inside design conditions listed in Table 4 are recommended for general heating applications. With heating, the temperature swing (variation) is below the comfort condition at the time of peak heating load (no people, lights, or solar gain, and with the minimum outdoor temperature). Heat stored in the building structure during partial load (day) operation reduces the required equipment capacity for peak load operation in the same manner as it does with cooling. f ’ i PAR-l- I l-20 . LO,\D E!xIhI,\‘-rIKG TABLE 44tECOMMENDEDINSIDE DESIGN CONDITIONS*-SUMMER AND WINTER WINTER SUMMER With Commercial Prodice TYPE OF APPLICATION Humidification Without Humidification GENERAL COMFORT Apt., House, Hotel, Office Hospital, School, etc. RETAIL SHOPS (Short term occupancy) Bank, Barber or Beauty Shop, Dept. Store, Supermarket, etc. LOW SENSIBLE HEAT FACTOR APPLICATIONS (High Latent load) Auditorium, Church, Bar, Restaurant, Kitchen, etc. FACTORY COMFORT Assembly Areas, Machinina Rooms. etc. I I I I I , * he room design dry-bulb temperature should be reduced when hot radiant panels ore adjacent to the occupant and increased when cold Panels are ~di,,en+ to compensate for the increase or decrease in radiant heat exchange from the body. A hot IX Cold Panel m’JY be unshaded glass or 9l“ss block wjidows (hot in summer, cold in winter) and thin partitions with hot or cold spaces adjacent. An unheated slab floor on the ground or walls below the ground level are cold panels during the winter and frequently during the summf~ ~Iso. Hot tanks, fUrnoceS Or machines are hot Panels. TTemperoture swing ir above the thermostat setting at peak summer load conditionsJTemperature swing is below the thermostat setting at peak winter load conditions (no lights, People Or solar heat gain). **Winter humidification in retail clothing shops is recommended to maintain the quality texture of goods- INSIDE INDUSTRIAL DESIGN CONDITIONS Table 5 lists typical temperatures and relative humidities used in preparing, processing, and manufacturing various products, and for storing both raw and finished goods. These conditions are only typical of what has been used, and may vary with applications. They may also vary as changes occur in processes, products, and knowledge of the effect o f temperature and humidity. In all cases, the temperature and humidity conditions and the permissible limits of variations on these conditions should be established by common agreement with the customer. Some of the conditions listed have no effect on the product or process other than to increase the efficiency oE the employee by maintaining comfort conditions. This normally improves workmanship and uniformity, thus reducing rejects and production cost. In some cases, it may be advisable to compromise between the required conditions and comfort conditions to maintain high quality commensurate with low production cost. Generally, specific inside design conditions are required in industrial applications for one or more of the following reasons: 1. A constant temperature level is required for close tolerance measuring, gaging, machining, or grinding operations, to prevent expansion and contraction of the machine parts, machined products and measuring devices. Normally, a constant temperature is more im-; portant than the temperature level. A constant: relative humidity is secondary in nature but’, should not go over 457, to minimize formation of heavier surface moisture film. Non-hygroscopic materials such as metals, glass, plastics, etc., have a property of capturing water molecules within the microscopic surface crevices, forming an invisible, non-continuous surface film. The density of this film increases when relative humidity increases. Hence, this film must, in many instances, be held below a critical point at which metals may etch, or the electric resistance of insulating materials is significantly decreased. 2. Where highly polished surfaces are manufac-tured or stored, a constant relative humidity and temperature is maintained, to minimize increase in surface moisture film. The temperature and humidity should be at, or a little (:F1.\l”l’lCK 2. I)ESI(;N (:ONI)I’I’IONS below, the comfort conditions to minimize perspiration of the operator. Constant temperature ant1 humidity may also be r~quirctl in machine rooms to prevent etching or corrosion of the parts of the machines. With applications of this type, if the conditions are not maintained 24 hours a day, the starting of air conditioning after any prolonged shutdown shoultl bc tlonc carefully: (1) During the summer, the moisture accumulation in the space should be reduced before the temperature is reduced; (2) During the winter, the moisture should not be introduced before the materials have a chance to warm up if they are cooled during shutdown pcriotls. 3. Control of relative humidity is required to maintain the strength, pliability, and regain of hydroscopic materials, sucll as textiles and pAper. The humidity must also be controlled in some applications to reduce the effect of static electricity. Development of static electric charges is minimized at relative humidities of 55% or higher. 4. The temperature and relative humidity control are required to regulate the rate of chemical or biochemical reactions, such as drying of l-21 varnishes or sugar coatings, preparation of synthetic fibers or chemical compounds, fermentation of yeast, etc. Generally, high temperatures with low humidities increase drying rates; high temperatures increase the rate of chemical reaction, and high temperatures and relative humidities increase such processes as yeast fermentations. 5. Laboratories require precise control of both temperature and relative humidity or either. Roth testing and quality control laboratories are frequently designed to maintain the ASTM Standard Conditions’ of 73.4 F db and 50% rh. 6. With some industrial applications where the load is excessive and the machines or materials do not benefit from controlled conditions, it may be advisable to apply spot cooling for the relief of the workers. Generally, the conditions to be maintained by this means will be above normal comfort. *Published in ASTM pamphlet dated 9-29-48. These conditions have also been approved by the Technical Committee on Standard Temperature and Relative Humidity Conditions’ of the FSB (Federal Specifications Board) with one variation: FSB permits 24%/,, whereas ASTM requires *2710 permissable humidity tolerance. l-24 I’,\RT I. LOAD ESTIMATING TABLE S-TYPICAL INSIDE DESIGN dONDlTlONS-INDUSTRIAL (Listed conditions INDUSTRY are only PROCESS ABRASIVE rtanufacture BAKERY )ough M i x e r :ermen~ing ‘roof Box bread Cooler Zold Room &cake-up Rm. Zake Mixing lrackerr & Biscuih Nrapping jtorogeDried typical; r Ingred. design 75-80 45-50 75-80 75-82 92-96 70-80 40.45 78-82 95-105 60-65 60-65 40-50 70-75 80-85 80-85 65-70 50 60-65 70 55-65 45-70 55-60 80 35 Water 32-35 - Wax Paper 70-80 40-50 Sugar are established 50-65 CERAMICS 30-32 Grain 80 . 32-34 Lager 32-35 75 Ale 40-45 75 40-45 75 55 75 32-35 75 80-85 60-65 75-80 40-50 50-55 55-60 Ale Racking CANDYCHOCOLATE Cellar Candy Centers Hand Dipping Rm. Enrobing Rm. EnrobingLoading End Enrober Stringing Tunnel Packing Pan Specialty Rm. General Storage CANDY-HARD CHEWING GUM Mfg. Mixing 8 Cooling TlIllIlel Packing storage D r y i n g - J e l l i e s , Gum C o l d Rm.Marshmallow Mfg. Rolling Stripping Breaking Wrapping 80 90 70 40-45 65 70-75 65-/O 75-80 75-80 55 65-75 65-75 120-150 - - 50 13 , 40-50 1DP - 40 30-40 40-45 DP - 55 40-45 45-50 15 75-80 45-50 77 68 72 74 74 33 63 53 47 58 PROCESS efroctory \olding Rm. lay storage lecol 8 Decorating DRYBULB (Fl 110-150 80 60-80 ___ 75-80 ackoging 75-80 Ifa. 65-70 DISTILLING w a g e Grain ELECTRICAL PRODUCTS lectronic & X - r a y Coils & Trans. Winding ube Arrem. lectrical Inst. M f g . 8. t a b . ‘hermortot Assem. & Calib. humidistat Assem. 6. &lib. 3o.w Tol. Assem. Aeter Asrem. Test iwi~chgear- -_ 60 --__ Liquid Yeast 32-34 M-f g .~~~~~___ 60-75 ~- 45-60 Aging 65-72 50-60 Fuse 8. Cut-Out Assem. CCID. W i n d i n a Paper storage Conductor Wrapping .ightning Arrestor Circuit Brkr. Assem. & Test ?ectifiersProcess Selenium & Copper Oxid< Plates Fermenting CellarLager requirements) COSMETICS 55-60 ~60 ___75 Liauid Yeasl customer CEREAL storageHops by INDUSTRY 80-85 Shortening conditions DRYBULB (F) 30-45 -. 70-75 Fresh Ingred. BREWERY final FURS Drying Shock Treatmeni Storage GLASS Zutting Vinyl Lam. Rm. LEATHER D r y i n g - 72 68 70 76 76 72 74-76 50-55 40-45 60-63 73 50 73 -7 3 75 68 50 50 65-70 20-40 76 30-60 74 30-40 110 18-20 40-50 55-65 Con ‘leg. Tanned Chrome Tanned storage LENSESOPTICAL Fusing Grinding MATCHES Mfg. D rying storage MUNITIONS MetalPercussion Elementr- 80 72-74 70-75 Drying Parts D&a Points Black Powder Drying Condition & Load Powder Type Fuse toad Tracer Pellets , Comfort I 50 ” 50 40 (;1-1,\1”1‘1:11 2. l)I,.sI(;s (:oIUI)l’I’IoNS 1-23 TABLE 5-TYPICAL INSIDE CONDITIONS-INDUSTRIAL (Contd) (Listed conditions INDUSTRY PHARMACEU. TICAL PHOTO MAT E R I A L only Powder design DRYBULB (F) . 70-80 30-35 conditions are TEXTILES 15-35 -75-80~~. 80 3570-80 1 40 80 35 - PRECISION MACHINING Spectrographic Anal. sear Matching 8. Assem. jtorageGasket Cement 8 Glue Aachinings Gogin& A s s e m . Adjusting Precision Ports Honing REFRIGERATION EQUIPMENT Valve Mfg. Compressor Assem. Refrigerator Asrem. resting RUBBER DIPPED GOODS Mfg. Cementing Surgical Articles storage Before Mfg. .ob. (ASTM std.) :otton O p e n i n g 8. P i c k i n g Cording D r a w i n g 6. R o v i n g requirements) DRYBULB (F) PROCESS Cotton, cont. Ring Spinning Conventional - -__ - 80-85 80-85 80-85 60-70 78-80 78-80 75 60-65 70-85 65-70 75 55-65 55-60 C___-.___ arding, Spinning 75.80 60 W eaving 80 80 __--.-. -_. ___...WOOlWlS ~-Pickers 80-85 60 ~.._._ _ __----Carding 80-85 65-70 ~__- -_ __.Spinning 80-85 50-60 ~-.._ --~__ _____ Dressino 75-80 60 -. Weaving~___ __tight Goods 80-85 55-70 ~__-_ - - - -.-. Heavy 80-85 60-65 -__~---_ Drawing 75 50-60 -__.._ ___ I W o~-rrteds - _-.. C a r d i n g , C o m b i n g I, B Giliing 80-85 60-70 ___storage 70-85 75-80 ~~--.. Drawing 80-85 50-70 ___Cap Spinning 80-85 50-55 __- ~S p o o l i n g , Winding I 75-80 --__ - 55-60 Weavino 80 50-60 A---Finishing 75-80 60 -__ silk .-. ~. _ Prep. 8. Dressing -. 80 60-65 Weavinq 8. Spin&a 80 65-70 Throwina 80 60 __-.- _ ?ayon -__ Spinning 80-90 50-60 -__ Throwing 80 55-60 .__--_ Weovina --___Regenerated 80 50-60 --__ Acetate 8 0 55-60 --.-_ Spun Rayon 80 80 ~Picking 75-80 50-60 Carding, Roving, Drawing 80-90 50-60 __Knitting - -~ --. Viscose or Cuprammonium 80-85 65 __~. Synthetic Fiber Prep. & Weaving Viscose 80 60 ___--_ _ _ _ CdO”eS.2 80 70 25-30 ~__ 45-65 60 15-25 Comfort _1 35-45 M u l t i c o l o r Litho. Pressroom Stockroom Sheet & W e b P r i n t . Storage, Foldino, etc. customer Cloth __Room -.-Combing Linen 70 30-50 80 40 78-80 5-10 80 35 78 40-50 75 35-40 Comfort 80 35 70-80 20-30 Comfort Comfort 90 90 (cont.) by Long Draft Frame Spinning S p o o l i n g , W o r p i n 19 _Weaving Film _) Hfg.Thermo S e t t i n g Compounds Cellophane established INDUSTRY Drying C u t t i n g & Packing storoge~_~-_ Film Base, Film Paper, Coated Paper Safety Film lot Press-Resin :old Press TEXTILES final Storage Before Mfg. After Mfa. Milling Rm. Tablet Compressing Tablet Coating Effervescent~-__Tablet 8. Powder Hypodermic Tablet Colloids Cough Syrup Glandular Prod. Ampule Mfg. Gelatin Capsule Capsule storage Microanalysis Biological Mfg. Liver Extract Serums Animal Rm. PLYWOOD PRINTING typical; PROCESS rlitrote PLASTIC are Comfort Comfort Nylon TOBACCO C i g a r 84 Cigarette Mfg. Softening S t e m m i n g 8 Strippin 9 storage 8. Prep. Conditioning Qcking 8. S h i p p i n g 80 50-60 70-75 90 75-85 78 75 75 55-65 85.88 75 70 75 60 l-25 CHAPTER 3. HEAT STORAGE, DIVERSITY AND S’JXATIFTCAT~ON 7%~ i~or111;~1 load cstini;tling proccdurc has been to evaluate the instanlancous heat gain to a space and to assume tllat the equilmlcnt will I;cmovc the hcnt at this rate. Generally, it was found tllat the ec~~~ipwmt sclccted on this basis was oversized and therefore cnpablc of maintaining much lower room conditions than the original design. Extensive annlysis, research and testing have shown that the reasons for this arc: 1. Storage of heat in the building structure. 9 Non-simultaneous occurrcncc of the peak of the individual loads (diversity). 3. Stratification of heat, in some cases. This chapter contains the data and procedures for determining the load the equipment is actually picking LIP at any one time (actual cooling load), taking into account the above factors. Application of these data to the appropriate individual heat gains results in the actual cdbling load. The actual cooling load 1s generally considerably below the peak total instantaneous heat gain, thus requiring smaller equipment to perform a specific job. In addition, the air quantities and/or water quantities are reduced, resulting in a smaller overall system. Also, as brought out in the tables, if the equipment is operated somewhat longer during the peak load periods, and/or the temperature in the space is allowed to rise a few degrees at the peak periods during cooling operation, a further action in required capacity results. The smaller system operating for longer periods at times of peak load will produce a lower first cost to the customer with commensurate lower demand charges and lower operating costs. It is a well-known fact that equipment sized to more nearly meet the requirements results in a more efficient, better operating system. Also, if a smaller system is selected, and is based on extended periods of operation at the peak load, it results in a more economical and efficient system at a partially loaded condition. Since, in most cases, the equipment installed to perform a specific function is smaller, there is less margin for error. This requires morexexacting engineering including air distribution design and system balancing. With multi-story, multi-room application, it is usu;~lly tlcsirable to provide some flcsibility in the air sitlc or room load to ;~llow I’or individual room control, load pickup, etc. Ccncrnlly, it is recommended that the full retl~lction from storage and diversity be taken on the overall rcfrigcrntion or I~iiiltling load, with some degree of conservatism on the air side or room loads. This degree should be determined by the cnginccr from project requiremcnts and customer desires. A system so designed, full reduction on refrigeration load and less than full reduction on air side or room load, meets all of the flcsibility requirements, except at time of peak load. In addition, such a system has a low owning and operating cost. STORAGE OF HEAT IN BUILDING STRUCTURES The instantaneous heat gain in a typical comfort application consists of sun, lights, people, transmission thru walls, roof and glass, infiltration and ventilation air and, in some cases, machinery, appliances, electric calculating machines, etc. A large portion of this instantaneous heat gain is radiant heat which does not’ become an instantaneous load on the equipment, because it must strike a solid surface and be absorbed by this surface before becoming a load on the equipment. The breakdown on the various instantaneous heat gains into radiant heat and convected heat is approximately as follows: HEAT GAIN SOURCE Solar, without inside blinds Solar, with inside blinds Fluorescent Lights Incandescent Lights People* Transmission+ Infiltration and Ventilation Machinery or AppIianceQ RADIAN? * C O N V E C T I V E HEAT H E A T 100% 58% 50% 80% 40%, 6’3% - 20.80% 42% 50% 20% 20% 40% 100YO 80s200(, *The remaining 40% is dissipated as latent load. +Transmission load is considered to he 100yO convective load. This load is normally a relatively small part of the total load, and for simplicity is considered to I)e the instantaneous load on the equipment. $Thc load from machinery or appliances varies, depending upon the temperature of the surface. The higher the surface temperature, the greater the radiant heat load. CONSTANT SPACE TEMPERATURE AND EQUIPMENT OPERATING PERIODS ,,\s tlic r;ttli:\nt heat from S O L I I ‘ C ~ S s h o w n i n t h e above table strikes ; I solid surface (walls, floor, cciling, etc..), it is absorbed, r a i s i n g the tcmperaturc at the surface of the n~aterinl above that insitlc the material and the air adjacent to the surt’ace. This temperature ditfcrcncc causes heat flow into the material by conduction and into the air by convection. The heat conducted away from the surface is stored, and the heat convected lrom the surface becomes an instantaneous cooling load. The portion of radiant heat being stored depends on the ratio of the resistance to heat flow into the material and the resistance to heat flow into the air film. With most construction materials, the resistance to heat flow into the material is much lower than the air resistance; therefore, most of the radiant heat will oe stored. However, as this process of absorbing radiant heat continues, the material becomes warmer and less capable oE storing more heat. is the actual cooling load that results in an avcragc construction al)plication with the space tempcrature held constant. The reduction in the peak heat gain is approximately zlO(;h and the peak load l a g s the peak heat gain by approximately 1 hour. The cross-hatched areas (I’is. 3) represent the Heat Stored and the Stored Heat Removed from the construction. Since all ol the heat coming into a space must be rcmovctl, these two areas are equal. FIG. 3 - ACTUAL COOLING LOAD, SOLAR HEAT G AIN , W EST E X P O S U R E, A VERAGE C O N S T R U C T I O N The relatively constant light load results in a large portion being stored just after the lights are turned on, with a decreasing amount being stored the longer the lights are on, as illustrated in I;ig. f. The upper and lower curves represent the instantaneous heat gain and actual cooling load from /Z~orescent lights with a constant space temperature. The cross-hatched areas are the Heat Stored and the , Stored Heat Removed from the construction. The dotted line indicates the actual cooling load for the first clay if the lights are on longer than the period shown. Figs. 3 and -f illustrate the relationship between the instantaneous heat gain and the actual cooling load in average construction spaces. With light construction, less heat is stored at the peak (less storage capacity available), and with heavy construction, more heat is stored at the peak (more storage capacity available), as shown in Fig. 5. This aspect affects the extent of zoning required in the design of a system for a given building; the lighter the building construction, the more attention should be given to zoning. The upper curve OF I;ig. 5 is the instantaneous solar heat gain while the three lower curves are the actual cooling load for light, medium and heavy constrz~tion respectively, with a constant temperature in the space. One more item that significantly affects the storage of heat is the operating period of the air conditioning equipment. All of the curves shown in FIG. 4 - ACTUAL COOLING LOAD FROM FLUORESCENT LIGHTS, AVERAGE CONSTRUCTION FIG. 5 - ACTUAL COOLING LOAD, SOLAR HEAT G AIN , LIGHT, MEDIUM AND HEAVY CONSTRUCTION The highly varying and relatively sharp peak of the instantaneous solar heat gain results in a large part of it being stored at the time oE peak solar heat gain, as illustrated in Fig. 3. The upper curve in Fig. 3 is typical of the solar j~at gnin for a west exposure, and the lower curve . ! (:I1.\l’~l’I~,l< 3. IIL\‘I’ s’I’oli.\c;~. I~i\‘~.IIsI’I‘\~. .\KT) ~1‘1I.\~1’11~1~:.\ 1-27 I’IOS I;i,q.s. 3, J, ~rrlrl 5 illrlstratc the actual cooling lo;~cl [or 2.1.horrr operxr.ion. 11’ the ccluipiicnt i s s h u t tlotvn alter 16 li011rs ol oper;ition, some of the storctl heat rcln;titls in tlic building construction. This heat moist 1x2 rcniovctl @cat in miist ccl~al hcxt out) a t i d w i l l ;ippear ;ls ;I ~~iilltlow~i load when tlic cquilxnci~t is turned on tlic next day, ;IS illustrated ill Fig. 6. :\cltiing the pidlclown loacl to the cooling load for that day results in the actual cooling load lor lci-hour ope~lion, as illi~stratcd in Pig. 7. The upper curve represents the instantaneous heat gain and the lower curve the ~C~UCLI cooling loall for that day with a constant temperature maintained within the space during the operating period of the equipment. The dotted line represents the additional cooling load from the heat left in the buildconstruction. The temperature in the spacex.,..s during the shutdown period lrom the nighttime transmission load and the stored heat, and is brought back to the control point during the pulldown period. Shorter periods ol operation increase the pulldown load because more stored heat is left in the building construction when the equipment is shut off; I;ig. 8 illustrates the pulldown load for X 2 - h o u r TIME IHR) , L FIG, 8 - PULLDOWN LOAD, SOLAR HEAT G AIN , W I -ST EXPOSIJRE, l.Z-HOLIR O PERATION operation. Adding this pulldown load to the cooling load for that day results in the actual cooling load for 12-hour operation, as illustrated in Fig. 9. The upper and lower solid curves are the instantaneous heat gain and the actual cooling load in average construction space with a constant temperature maintained during the operating period. The cross-hatched areas again represent the Heat Stored and the Stored Heat Kemoved from the jtruction. The light loud (fluorescent) is shown in Fig. 10 for 12- and 16-hour operation with a constant space temperature (assuming lo-hour operation ol’ lights). 16 ACTUAL COOLING LOAD v TIME (HA) 12 FIG. 9 - ACTUAL COOLING LOAD, SOLAR HEAT G AIN , W EST EXPOSURE , 12-130~~ O PERATION I N S T A N T A N E O U S H E A T GAP4 FIG. 6 - PULLDOWN LOAD, SOLAR HEAT G AIN , W EST EXPOSURE , 16-HOIJR O PERATION FIG. 10 - ACTUAL COOLING, LOMI FROM FLUORESCENT LIGHTS, 12- AND I6-HOIIR O PERATION Basis of Tables 7 thru 12 Storage Load Factors, Fintl: .\. The actual cooling load from the solar heat gain in July at 4 p.m., 40” North latitude with the air conditioning equipment operating 24 hours during the pc;t!i loarl pcriotls and a constant temperature maintained within the room. Solar and Light Heat Gain 12., 16., and 24-hour Operation, Constant Space Temperature 'l‘l~csc t;tl)lcs tl~~vclopctl These arc c a l c u l ; t t c t l , using Cr01n ;I s e r i e s ol t e s t s i n tests were contluctctl i n a actual I)roceclurc 0. The I)uiltlings. office I)uildings, su- ;~ntl residences throughout this country. l>cr”‘;lrkcts, J‘hc Iil;tgnitutlc or the stor:tge effect is determined largely by the thermal capacity or heat holding capacity of the niatcrials surrounding the space. The thermal capacity of a material is the weight times the hpccific heat of the material. Since the specific heat of most construction material is approximately 0.20 I%tu/ (lb) (F), the thermal capacity is directly proportional to the weight of the material. Therefore, the data in the tables is based on weight of the tcrials surrouncling the space, per cooling load at 8 p.m. for the same conditions. Solution: The weight per sq ft of floor arca of this room (values 01,. tained from Chn/7ter 5) is: Orltsitlc wall = 120 x ;; ; -<&yj x ,‘) x 1% ,l,/sq ft (Table 31, page 64) = 25.2 Il)/sq ft floor area “0X8X3 x 22 Il,/sq ft 20 x 20 (Table 26, page 70) = 13.2 lh/sq ft floor area Partitions = l/s x Floor = l/z x $zyyqf x 59 lb/q ft i (Table 29, page 73) = 29.5 Ib/sq ft floor area 20 x 20 square Eoot of 110or arca. Solar and Light Heat Gain 12-, 16-, and 24.hour Operation, Constant Space Temperature Tables 7 t1~~z~ II are used to determine the actual cooling load from the solar heat gain with a constant temperature maintained within the space for different types of construction and periods of operation. Iliith both the 12- and l&hour factors, the starting time is assumed to be 6 a.m. suntime (7 a.m. Daylight Saving Time). The weight per sq ft of types of construction are listed in Tables 21 thru 33, pages 66-76. The actual cooling load is determined by multiplying the storage load factor from these tables for any or all times by the peak solar heat gain for ” e particular exposure, month and latitude desired. :\ i’able 6 is a compilation of the peak solar heat gains for each exposure, month and latitude. These values are extracted from Table 15, page 44. The peak solar heat gain is also to be multiplied by either or both the applicable over-all factor for shading devices (Table 16, page 52) and the corrections list& under Table 6. Reduction in solar heat gain from the shading of the window by reveaIs and/or overhang should also be utilized. Example J - 20 x 20 = I/z x “0 x 59 Il,/sq ft (Table 29,page 73) = 29.5 Il)/sq ft floor area Ceiling Use of Tables 7 thrti 12 Storage Load Factors, Actual Cooling Load, Solar Heat Gain Given? ti ti A 20 ft X 20 ft X 8 ft outside office room with &inch sand aggregate concrete floor, with a floor tile finish, 2l/$inch solid sand plaster partitions, no suspended ceiling, and a 12-inch common brick outside wall with i/s-inch sand aggregate plaster finish on inside surface. A 16 ft X 5 ft steel sash window with a white Venetian blind is in the outside wall and the wall faces west. . NOTE: One-half of the partition, floor and ceiling thickness is used, assuming that the spaces above and below are conditioned and are utilizing the other halves for storage of heat. . Total weight per sq ft of floor area = 25.2 + 13.2 + 29.5 + 29.5 = 97.4 lb/sq ft. The overall factor for the window with the white Venetian blind is 0.56 (Table i6, page 52) and the correction for steel sash = 1 / .85. .I. Storage factor, 4 p.m. = 0.6G (Thble 7) The peak solar heat gain for a west exposure in July at 40” North latitude = 164 Btu/(hr)(sq ft), (Table 6). Actual cooling load = 5 x 16 x 164 x .56 x-j& X 0.66 = 5700 Btu/hr ( ) II. Storage factor, 8 p.m. = .20 (Table 7) Actual cooling load = . ( 5 x 16 x 164 x .56x-$- 1 x .20 = 1730 Btu/hr Table I-3 is used to determine the actual cooling load from the heat gain from lights. These data may also be used to determine the actual cooling load from: People - except in densely populated areas such as auditoriums, theaters, etc. The radiant heat exchange from the body is reduced in situations like this because there is relatively less surface available for the body to radiate to. Some appliances and machines that operate periodically, with hot exterior surfaces such as ovens, dryers, hot tanks, etc. NOTE: For Items 1 and 2 above, use values listed for fluorescent exposed lights. ! (;H:\I”I‘ER 3. I-IlC.\‘I‘ S’I‘OKAGE. I>IVI:KSI~l‘Y, 1-X) AI .\Nl> ~l‘li.\‘l‘ll~l~:.\‘l’lON = 51’31) Iltu/l1~. ns 111~ pcoplc nrrivc ;tt R ;l.~n.). TABLE 6-PEAK SOLAR HEAT GAIN THRU ORDINARY GLASS* Btu/(hr)(sq it) NORTH LAT. MONTH JUne July B May Aug & April Sept 8 March Ott B Feb Nov 8 Jon Dee ?O June July K May Aug & April Sept 8 March Ott S Feb Nov B Jon Dee !,"+QOO -J&l”* July 8 May Aug & April Sept 8 March Ott & Feb Nov & Jon Dee 20" /-“ \ ,30" ) / ‘l...-.-A’ June July 8. May Aug 8 April Sept & March Ott & Feb Nov & Jon D.X JWle July 8 May - Aug B April Sept 8 March Ott & Feb Nov 8 Jon Dee 5 0 ” JWle July B May Aug 8 April ‘Sept 8 March Ott 8 Feb Nov B Jon DeC Nt - 59 48 25 10 10 10 10 40 30 13 10 10 9 9 '26 19 11 IO 9 8 8 20 EXPOSURE NORTH LATITUDE NE E SE 156 153 141 147 152 42 52 79 118 141 153 156 14 14 I4 ‘4. 34 a2 156 55 14 14 14 55 “a’ 79 52 42 163 163, 163 152 147 12 148 130 103 155 158 66 155 143 137 37 28 154 ,138 118 a7 52 26 18 163 164 94 127 149 106 14 14 163 73 85 113 140 26 65 147 128 121 160 164 lb7 111 141 149 161 90 100 129 152 21 30 160 163 165 164 165 11 9 a 7 158 135 116 105 6 17 15 11 9 7 5 5 133 127 102 58 35 12 10 162 lb 126 117 94 58 29 9 7 ‘64 162 149 122 100 lb3 162 162 111 125 146 lb2 163 Steel Sash or No Sash X1/.85 or 1.17 120 63 ‘05 145 159 163 54 69 102 140 162 156 a6 148 166 165 lb4 163 135 143 157 93 158 138 105 64 47 163 157 127 -116 106 ‘38 158 167 153 141 E EXPOfURE Solar Gain Correction 28 73 161 163 lb 14 11 8 5 4 3 5 - 66 67 Haze -15%(Max) 42 52 79 118 141 153 66 94 127 149 101 163 73 a5 113 140 lb0 164 167 NW ioriz MONTH 147 156 226 Dee Nov 8 Jon Ott 8 Fab Sept 8 March Aug 8 April July 8 May June I53 148 130 103 G7 37 28 250 247 230 210 202 Dee Nov 8 Jan Ott 8 Feb Sept & March Aug 8 April July 8 May June 154 138 118 a7 52 26 18 250 251 247 233 208 i,ao 170 DeC Nov 8 Jan Ott 8 F e b Sept 8 March Aug 8 April July 8 May J u n e DeC Nov 8 Jon Ott 8 Feb Sept 8 Mfarch Aug 8 April July & May June 30" 12 250 246 235 212 179 145 131 133 127 102 58 35 12 10 237 233 214 183 129 103 85 Dee Nov a?. Jan Ott 8 Feb Sept 8. March Aug & April July 8, May June 40" Dee Nov & Jan Ott & Feb Sept 8 March Aug & April July & May June 50" 152 163 lb7 163 152 147 155 158 163 lb4 155 143 137 160 163 165 163 147 128 121 I53 141 118 79 52 42 66 90 100 129 152 163 161 164 165 158 135 ye :131, 108 90 39 162 162 116 lb 111 125 146 162 163 156 ,J32, 148 135 143 157 163 157 127 SOUTH LAT. W 105 (64 1-6-2~ 149 122 100 a6 164 lb3 158 138 105 64 116 47 N W W 233 245 250 245 233 226 126 117 94 58 29 9 7 JO” 20" tlorir - SOUTH LATITUDE Altitude +0.7% per 1000 ft Dewpoint Above 67 F -7y0 per 10 F Dewpoint Below 67 F +770 p e r 1 0 F South Lot Dee or Jon +7% *Abstracted from Table 15, page 43. ~Solar heat gain on North exposure (in North latitudes) or on South exposure (in South latitudes) consists primarily of diffuse radiation which is esrentially constant throughout the day. The solar heat gain values for this exposure ore the average for the 12 hr period (6 mm. to 6 P.?). The storage factors in Tables 7 thru 11 (1ssutne that the solar heat gain on the North (or South) exposure is constant. I ~ l’.\R’l I. I.O,\I) I~S’l’lhf.\-l’lN(~ l-00 TABLE 7-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS W&i INTERNAL SHADE* 24 Hour Operation, Constant Space Temperaturet EXPOSURE (North 1.1) WEIGHTS (lb per sq fi of floor area) SUN TIME AM 6 7 8 9 PM 10 11 12 1 2 3 4 5 6 7 A M l 8 9 10 11 12 1 2 3 4 EXPOSURE (South Lat) 5 Northeast 150 a over 100 30 .47 .58 54 .42 .27 .2, .20 .,P .I8 .I7 .16'.14 .I2 1.09 .08 .07 .06 '.06 .05 .05 .04 .04 .04 .03 . 4 8 . 6 0 . 5 7 . 4 6 . 3 0 . 2 4 2 0 . , 9 . , 7 ;. I 6 1. I 5 '. 1 3 . l l I . 0 8 . 0 7 . 0 6 . 0 5 . 0 5 . 0 4 . 0 4 . 0 3 . 0 3 . 0 2 . 0 2 55 .76 .73 .58 .36 .24 .,9 .17 .I5 .I3 .I2 .I1 .07 .04 .02 .02 .Ol .Ol 0 0 0 0 0 0 Southeast East 1 5 0 8 DYB, 100 30 .39 .56 .62 .59 .49 .33 .23 2, .20'.18 ,I7 .I5 .I2 .I0 .09 .08 .08 .07 .06‘ .05 .05 .05 .04 .04 . 4 0 . 5 8 . 6 5 . 6 3 . 5 2 . 3 5 . 2 4 . 2 2 I . 2 0 . 1 8 I. 1 6 . I 4 /. I 2 . 0 9 . 0 8 . 0 7 . 0 6 . 0 5 . 0 5 . 0 4 . 0 4 . 0 3 . 0 3 . 0 2 0 0 0 0 0 0 .46 .70 .80 .79 .64 .42 .25 .,9 .,6'.14 1.11 .09 .07 .04 .02 .02 .Ol .Oi East 150 8 eve, 100 Northeast 30 .04 .28 .47 59 .64 .62 33 .4, 27 .24 .21 .I9 .16,.14 .,2'.1, .I0 .09 .08 .07 .06 .06 .05 .05 .03 .28 .47 .61 .67 .65 .57 .44 .29 .24 .21 .I8 .I5 .I2 .lO .09 .08 .07 .06 .05 .05 .04 .04 .03 0 0 0 0 0 0 0 .30 .57 .75 .84 .8, .69 30 .30 .20 .17 .I3 .09 .05 .04 .03 .02 .Ol 150 8 OYW 100 30 . 0 6 . 0 6 . 2 3 . 3 8 5 , . 6 0 . 6 6 . 6 7 . 6 4 ' . 5 9 , . 4 2. 2 4 . 2 2 . I 9 . I 7 . I 5 . I 3 . 1 2 . I 1 . I 0 . 0 9 . 0 8 . 0 7 . 0 7 .04 .04 .22 .38 52 .63 .70 .71 .69 .59 .45 .26 .22 .I8 .I6 .13 .12 .lO .09 .08 .07 .06 .06 .05 .I0 .21 .43 .63 .77 .86 .88 .82 .56 .50 .24 .I6 .I1 .08 .05 .04 .02 .02 .Oi .Ol 0 0 0 0 North 150 a over 100 .08 .08 .09 .,O .,, .24 .39 .53 .63 .66 .61 .47 .23 .I9 .18 .16 .I4 .I3 .I1 .lO .09 .08 .08 .07 .07 .08 .08 .08 .I0 .24 .40 .55 .66 .70 .64 .50 .26 20 .I7.I5 .13 .I1 .I0 .09 .08 .07 .06 .05 Southeast South Southwest 30 West Northwest .86 .79 .60 .26 .I7 .12 .08 .05 .04 ,.03 .OZ .Ol .Ol 0 0 150 8 ever 100 30 .08 .09 .09 .lO .lO .I0 .I0 .I8 .36 .52 .63 1.65 .55 .22 .I9 .I7 .I5 .14 .I2 .I1 .lO .09 .08 .07 .07 .08 .08 .09 .09 .09 .09 .I8 .36 .54 .6_6 .68 .60 .25 .20 .17 .I5 .13 .11 .lO .08 .07 .06 .05 .03 .04 .06 .07 .08 .08 .08 .,9 .42 .65 .81 .85 .74 .30 .I9 .I3 .OP .06 .05 .03 ,02 .02 .Ol 0 150 8 over 100 .08 .09 .lO .,O .lO .,O .,O .,O .I6 .33 .49 .61 .60 .I9 .I7 .I5 .13 .I2 .lO .09 .08 .08 .07 .06 .07 .08 .09 .09 .I0 .I0 .I0 .lO .16 .34 .52 .65 .64 23 .I8 .I5 .12 .11 .09 .08 .07 I.06 .06 .05 30 North and Shade .03 .04 .06 .07 .09 .23 .47 .67 .81 150 a over 100 30 .03 .05 .07 .08 .09 .09 .I0 .I0 .17 .39 .63 .80 .79 .28 .I8 .I2 .09 .06 .04 .03 .02 .02 .Ol . Northwest West Southwest 0 .08 .37 .67 .7, .74 .76 .79 3, .83 .84 .86 .87 .88 .29 .26 .23 ,20 .I9 .17 15 .lA .I2 .ll .I0 .06 .31 .67 .72 .76 ,79 .81 .83 .85 .87 .88 .90 .91 .30 .26 .22 .19 .16 .I5 .I3 .12 .lO .09 .08 0 . 2 5 . 7 4 . 8 3 . 8 8 . 9 1 . 9 4 t.96 .96 .98 .98 .99 .99 .26 . 1 7 , . 1 2 . 0 8 . 0 5 . 0 4 . 0 3 , . 0 2 . O l , . O l . O l South and Bhade Equation: Cooling Load, Btu/hr = [Peak solar heat gain, Btu/(hr) (sq ft), (Table 6)] x [Window oreo, sq ft] X [Shade factor, ,Hoze factor, etc., (Chapter 4)] X [Storage factor, (above Table at desired time)] *Internal shading device is any type of shade located on the inside of the glass. tThese factors apply when maintaining a CONSTANT TEMPERATURE in the space during the operating period. Where the allowed to swing, additional storage will result during peak load periods. Refer to Table 13 for applicable storage factors. $Weight temperature p e r sq ff o f floor(Weight of Outside Walls, lb) + Room on Bldg Exterior (One or more outside walls) = Room in Bldg Interior (No outside walls) = Ceiling, lb) ‘% (Weight of Partitions, Floor and Ceiling, lb) -.__ Floor Area in Room, sq ft (Weight of Outside Walls, lb) Basement Room (Floor on ground) = _____-__ Entire Building or Zone = % (Weight of Partitions, Floorznd Floor Area in Room, sq ft + (Weight of Floor, lb) +___% (Weight of Partitions and Floor Area in Room, sq ft (Weight of Outside Wall, Partitions, Floors, Ceilings, Structural Members ond Supports, lb1 _____-. -Air Conditioned Floor Area, sq ft With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug. Weights per sq ft of common types of construction are contained in Tables 21 fhru 33, pcrges 66 thru 76. Ceiling,~lb) __-__ is (:I-1.\l”l’I~:lI 3. III,:.\ I ‘ S’I 01<.\c;1:. I)IvI~:lisl’l’\r’, 1-o 1 .\NI) s’l~I~.\‘I‘II~I~:,\‘l’loi\[ TABLE 8-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS WITH BARE GLASS OR WITH EXTERNAL SHADE: 24 EXPOSURE ( N o r t h Lot) 1 WEIGHT$ lib oer %a It if fi00r Ad Hour Operation, Constant Space Temperaturet SUN TIME AM 6 7, 8 9 11 12 AM PM I 10 1 2 3 4 6 5 7 9 8 10 1t 12 1 2 3 4 5 Northeast 150 a oY*r 100 30 .I7 27 .33 .33 21 29 .27 ,125 .23 .22 20 .,P .I7 .I5 .I4 .I2 .I1 .I0 .09 .08 .07 .07 .06 .06 .I9 .31 28 .39 .36 .34 27 .24 .22 .2l .I9 .17 .I6 .I4 .I2 .I0 .07 .08 .07 .06 .05 .05 .04 .03 .31 56 .65 .6l .46 .33 .26 .2l .I8 .I6 .I4 .I2 .09 .06 .04 1.03 .OZ .Ol .Ol .Ol 0 0 0 0 East 150 B ov*r 100 30 .I6 .26 .34 .39 .40 .38 .34 .30 .28 .26 .23 .22 .20 .I6 .29 .40 .46 .46 .42 .36 .3l .28 .25 .23 .20 .I8 .27 50 .67 .73 .68 .53 .38 .27 .22 .I8 .I5 .I2 .09 150 a OYBl 100 .08 .I4 .22 .3l .38 .43 .44 .43 I.39 .35 .32 .29'.26 .05 .I2 .23 .35 .44 .49 .51 .47 .4l .36 .3, .27 .24 Southeast 30 South thwert, 150 & *Ye* 100 '.I1 .I0 .lO .I0 .I0 .14 .21 .29 .36 .43 .47 .46‘.40 .34 .30 .27 .24'.7.2 .20 .I8 .16 .14 .I3 .12 .09 .09 .08 .09 .09 .14 .22 .3l .42 30 .53 .5l .44 .35 .29 .26 .22 .I9 .17 .I5 .I3 .I2 .I1 .09 30 Northwest North / .02 .03 .05 .06 .08 .I2 .34 53 .68 .78 .78 .68 .46 .29 .20 .I4 .09 .07 .05 .03 .02 .02 .Ol .I2 .I1 .I1 .I0 .I0 .lO .I0 .13 .I9 .27 .36 .42 .44 .38 .33 .29 .26 23 .21 .18 .I6 .I5 .I3 .,2 .09 .09 .09 .09 .09 .09 .I0 .I2 .I9 .30 .40 .48 .51 .42 .35 .30 .25 .22 .19 .I6 .I4 .I3 .I1 .09 .02 .03 .05 .06 .07 .07 .08 .I4 .29 .49 1.67,.76 I.75 53 .33 .22 I.15 .I1 30 .02 .04 .05 .07 .08 .09 .I0 .I0 .13 I.27 .48 .65 .73 .49 .3l .21 .I6 .I0 .07 .05 .04 .03 .OZ .Ol Equation: Cooling toad, 30 1. I 6 1. 2 3 ! . 3 3 /. 4 l 1. 4 7 /. 5 2 /. 5 7 / . 6 l /. 6 6 /. 6 9 /’7 2 i. 7 4 ’ . 5 9 i. 5 2 i i 1. 4 6 . 4 2 1. 3 7 1. 3 4 /. 3 l /. 2 7 1. 2 5 . 2 3 / . 2 1 / . I 7 1 .I1 53 .44 .5l .57 .62 .66 .70 .74 .76 .79 .80 I.60 Sl .44 .37 .32 .29 .27 .23 .21 .18 .I6 .I3 0 .48 I.66 I.76 .82 .87 .9l I.93 .95 .97 .9e .98 I.52 .34 .24 .I6 I.11 I.07 1.05 .04 .02 .02 .Ol .Ol Btu/hr = [Peak solar heat gain, Btu/(hr) (sq ft), (Table Northwest West .08 .05 .04 1.03 .02 .O, .lO .I0 .I0 .,O .I0 .I0 .I0 .I0 .I2 .I7 .25 .34 .39 .34 .29 .26 .23 .20 .I8 .16 .14 .I3 .I2 .I0 .08 .09 .09 .09 .09 .09 .09 .09 .,, .I9 I.29 .40 .46 .40 .32 .26 .22 .I9 .16 1.14 .I3 .I1 .I0 .08 150180~Ver North .O, 150 8 OVW 100 and Shade East Northeast .I0 .I0 .13 .20 .28 .3S .42 .48 .5l .5l ,413 .42 .37 .33 .29 .26 .23 .2l .I9 .I7 .I5 .I4 .I3 .12 .07 .06 .I2 .20 .30 .39 .48 .S4 .58 .57 .53 .45 .37 .3l .27 .23 .20 .I8 .I6 .I4 .I2 .I1 .I0 .08 0 0 .I2 .29 .48 .64 .75 .82 .8l .75 .6l .42 ,213 .I9 .I3 .09 .06 .04 .03 .02 .Ol .Ol 0 0 150 8 OYW 100 Southeast .I8 .40 .59 .72 .77 .72 .60 .44 .32 .23 .I8 .14 150 a over 100 30 30 West 0 .I3 .I2 .lO .09 .08 .OB .07 .06 .I1 .09 .08 .08 .06 .06 .05 .04 .02 .Ol .Ol .Ol .Ol 0 0 .Ol EXPOSURE (South L-t) Southwest I South and Shade 611 X [Window area, sq ft] x [Shade factor, Haze factor, etc., (Chapter 4)] X [Storage factor, (above Table at desired time)] $Bare glass-Any window with no inside shading device. Windows with shading devices on the outside or shaded by external projections are considered bare glass. tThese factors apply when maintaining a CONSTANT TEMPERATURE in the space during the operating period. Where the allowed to swing, additional storage will result during peak load periods. Refer to Table 13 for applicable storage factors. $Weight temperature per sq ft of floor- ,om on Bldg Exterior (One or more outside walls) Room in Bldg interior (No outside walls) = (Weight of Outside Walls, lb) + % (Weight of Partitions, Floor and Ceiling, lb) = --.-Floor Area in Room, sq ft ‘% (Weight of Partitions, Floor and Ceiling, lb) ~~~~~- --- -.-----v--. - Floor Area m Room, sq ft (Weight of Outside Walls, lb) + (Weight of Floor,lb)+ %-(Weight Basement Room (Floor on ground) = ~----~ Floor Area in Room, sq ft of Partitions and Ceiling, lb) (Weight of Outside Wall, Partitions, Floors, Ceilings, Structural Members and Supports,lb) Entire Building or Zone = Air Conditioned Floor Area, sq ft With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug. Weights per sq ft of common types of construction ore contained in Tables 21 fhru 33, pages 66 fhru 76. is I l-02 , I’;\R’I‘ I . LO,\I) I~S’I~IIi.\ I’IN(i TABLE 9-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS WITH INTERNAL SHADING DEVICE* 16 Hour Operation, Constant Space Temperaturet EXPOSURE (North Lot) SUN TIME WElGHlg (lb Per rq fl of floor area) EXPOSURE (South Lot) Northeast 150 8 a”*r 100 30 Southewt East 150 & OYB, 100 30 E0tt Southeast 1 5 0 a OVB, 100 30 Northeast South 150 a oy~r 100 30 North Southwest 150 8 over 100 30 20 .08 .I9 .OB .18 .09 .I7 .09 .18 .I0 .31 .24 .46 .47 .60 .67 .66 .a1 .70 .a6 .64 .79 .50 /.60 .26 .26 .20 .I7 .I7 .I2 .I5 .08 Wart 150 & ov*r 100 30 .23 .22 .12 .23 .21 .lO .21 .I9 .10 .21 .I9 .I0 .20 .I7 .lO .I9 .16 .I0 .18 .15 .09 .25 .23 .19 .36 .36 .42 .52 .54 .65 .63 .66 .65 .68 .a5 .55 .60 .74 .22 .25 .30 .I9 .20 .I9 .17 .17 .13 150 8 OVB, 100 30 .21 .19 .12 .21 .I9 .I I .20 .I8 .I1 .I9 .I7 .1 I .ia .ia .I6 .11 .I7 .I6 .ll .16 .15 .lO .I6 .I6 .I7 .33 .34 .39 .49 .52 .63 .61 .65 .a0 .60 .23 .79 .19 .I7 .l I .28 .I7 .I5 .18 .I5 .12 .I2 -.* 150 a over 100 30 .23 .25 .07 .5a .46 .22 .75 .73 .69 .79 .7a .a0 30 .a2 .a6 .a0 .a1 .a2 .93 .a3 .94 .a2 .a4 .95 .a3 .a5 .97 .a4 .a7 .90 .86 .a8 .9a .a7 .a9 .99 .aa .90 .99 .39 .40 .35 .35 34 .23 .31 .29 .16 Northwest North ond Shade Equation: Cooling . .a1 .Ia Northwest West Southwest South , and Shade Load, Btu/hr = (Peak solar heat gain, Btu/(hr) (rq ft), (Table 6)] x [Window area, sq ft] x [Shade factor, Haze factor, etc., (Chapter 411 x [Storage factor, (above Table at desired time)]i *Internal shading device is ony type of shade located on the inside of the glass. tThese factors apply when maintaining a CONSTANT TEMPERATURE in the space allowed to swing, additional storage will result during peak load periods. Refer to $Weight during Table the operating period. Where the 13 for applicable storage factors. temperature p e r rq ft o f floor(Weight of Outside Walls, lb) + % (Weight of Partitions, Floor and Ceiling, lb) Room on Bldg Exterior (One or more outside walls) = Floor Area in Room, sq ft I,$ (Weight of Partitions, Floor and Ceiling, lb) Room in Bldg interior (No outside walls) = Floor Area in Room, sq ft (Weight of Outside Walls, lb) + (Weight of Floor, Basement Room [Floor on ground) = Entire Building or Zone = lb) + % (Weight of Partitions and Ceiling, lb) __--. .- Floor Area in Room, sq ft (Weight of Outside Wall, Partitions, Floors, Ceilings, Structural Members and Supportd Air Conditioned Floor Area, sq ft With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug. Weights per sq ft of common types of construction ore contained in Tables 21 thru 33, pages 66 thru . 76. is TABLE IO-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS W I T H B A R E G L A S S O R W I T H E X T E R N A L SHADE$ 16 Hour Operation, Constant Space Temperature? EXPOSURE (North 1.1) SUN TIME WElGHT$ ( l b persq ft of floor oreal A M 6 7 8 9 Northeast 150 8 OVB, 100 30 .28 / .37 / .42 / .41 .20 .39 33 .57 East 150 8 OYW 100 30 .29 .27 .29 .3a .3a .51 Southeast 150 8 OVB, 100 30 .24 .I9 .03 South 150 a DYBl 100 30 Southwest 150 a OVW 100 30 PM 10 11 1 . 3 a / .36 12 1 2 .31 1 .33 3 4 5 6 9 I .I2 .I0 .03 Southeast .22 .20 .I2 .20 .I8 .09 .I8 .I5 .06 .I6 .I4 .04 .I4 .I2 .03 .29 .24 .20 .32 .31 .23 .29 .27 .I8 .26 .24 .I4 .23 .21 .09 .21 .ltl .07 .19 .I6 .05 Northeast .33 .27 .Ob .31 .24 .04 .46 .53 .bl .42 .45 .42 .37 .37 .2a .33 .31 .I9 .29 .27 .I3 .26 .23 .09 North .35 .31 .ll .32 .2a .lO .09 .34 .44 .46 .30 .35 .29 .27 .29 .20 .24 .26 .14 Northwest .I0 .40 .51 .ba .46 .48 .53 .41 .4l .3a .I0 .2a .26 .I4 .30 .33 .35 .54 .ba .46 1 .53? .78 .7a West 150 8 OVW 100 30 .3a .34 .I7 .34 .31 .I4 .32 .20 .I3 .2a .25 .I 1 .26 .23 .l I .25 .22 .lO .23 .21 .I0 .25 .21 .I5 .26 .23 .29 .27 .30 .a9 .36 .40 .67 .42 .48 .76 .44 .5l .75 Northwest 150 a ov*r 100 30 .33 .30 .lS .30 .2a .14 .28 .25 .I2 .26 .23 .I2 .24 .22 .12 .23 .20 .12 .22 .19 .I2 .20 .I7 .ll .lS .17 .I3 .I7 .lP .27 .25 .29 .4a .34 .40 .65 .39 .46 .73 North and Shade 150 8 OYer 100 30 .31 .30 .04 .57 .47 .07 .64 .ba .72 .73 .73 .74 .74 .75 .76 .7a .a2 .90 Equation: Cooling Load, 6 .23 .23 .I5 t . 7 I . 2 3 I . 2 2 I 2 0 I . I 9 I . I 7 I .15 I .I4 .I9 .I7 .I6 .I4 .I2 .14 .I2 .09 .Ob .04 EXPOSURE (South Lal) .a1 + .97 Btu/hr = [Peak solar heat gain, Btu/(hr) (sq ft), (Table ? Southwert ’ South and Shade 6)] X [Window area, sq ft] x [Shade factor, Haze factor, etc., (Chapter 4)] X [Storage factor, (above Table at desired time)] *Bare glass - Any window with no inside shading device. Windows with shading devices on the outside or shaded by external projections ore considered bare glass. ?These factors apply when maintaining CI CONSTANT TEMPERATURE in the space during the operating period. Where the allowed to swing, additional storage will result during peak load periods. Refer to Table 13 for applicable storage factors. $Weight ‘. temperature p e r sq ft o f f l o o r - kcxxn on Bldg Exterior (One or more outside walls) (Weight of Outside Walls, lb) + __% (Weight of Partitions, Floor and Ceiling, lb) = __ Floor Area in Room, sq ft % (Weight of Partitions, Floor and Ceiling, lb) Room in Bldg interior (No outside walls) = --Floor Area in Room, sq ft (Weight of Outside Walls, lb) Basement Room (Floor on ground) = ~ Entire Building or Zone = + (Weight of Floor, lb) + ‘/2 (Weight of Partitions -. and Ceiling, lb) Floor Area in Room, sq ft (Weight of Outride Wall, Partitions, Floors, Ceilings, Structural __.-.. Members and Supports, 5 Air Conditioned Floor Area, sq ft With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug. Weights per sq ft of common types of construction are contained in Tables 21 thru 33, pages 66 thru 76. is TABLE II-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS 12 Hour Operation, Constant Space Temperature1 INTERNAL EXPOSURE (North La0 WEIGHT5 (lb per tq ft of floor area) BARE GLASS OR EXTERNAL SHADE1 SHADE* SUN TIME PM AM 6 1 B r9 41 11 12 1 2 3 AM 4 5 6 7 a EXPOSURE (South Lat) PM 9 IO 11 12 1 2 3 4 5 Northearl 150 a ov*r 100 30 . 4 2 . 4 7 I . 4 5 I. 4 2 I . 3 9 I . 3 6 . 3 3 1 . 3 0 . 2 9 ' . 2 6 . 2 5 .45 SO .49 ..45 .42 ..34 .30 .27 .26 .23 .20 .62 .69 .64 .48 .34 .27 .22 .I8 .16 .I4 .I2 Eort 150 a oYw 100 30 .51 .66 .71 .67 .57 .40 .29 .26 .25 .23 .2l .I9 .36 .44 SO .53 .53 SO .44 .39 .36 .34 .30 .28 .52 .67 .u .70 .58 .40 .29 .26 .24 .21 .I9 .16 .3P .44 .54 .58 .57 .51 .44 .39 .34 .31 .28 .24 . 5 3 . 7 4 . a 2 .a1 .65 .43 .25 .I9 .I6 .I4 . I 1 . 0 9 . 3 6 . 5 6 . 7 1 . 7 6 . 7 0 S 4 . 3 9 . 2 8 . 2 3 . I 8 . I 5 . I 2 Earl Southeasl 150 8 ovw 100 30 .20 .42 .59 .70 .74 .71 .61 .48 .33/.30' .26 .24 .34 .37 .43 I.50 .54 .58 .57/.55 .50 .45 .4l .37 .18 .40 .57 .70 .75 .72 .63 .49 .34 .28 .25 .21 .29 .33 .41 .51 .58 .61 .61 .56 .49 .44 .37 .33 .09 .35 .61 .78 236 .a2 .69 .50 .30 .20 .I7 .I3 .I4 .27 .47 .64 .75 .79 .73 .61 .45 .32 .23 .lf Northeast South 150 a OVB, 100 30 . 2 8 . 2 5 . 4 4 . 5 3 .b4 .72 .77 77 .73 .67 .49 .31 .47 .43 .42 .46 .51 .56 .61 .65 .66 .65 .bl .54 .26 .22 .3a .51 .64 .73 .79 .79 .77 .65 .51 .31 .44 .37 .39 .43 SO .57 .64 .68 .70 .68 .63 S? .21 .29 .48 .67 .79 .a8 .a9 .a3 .56 .50 .24 .I6 .2a .I9 .25 .38 .54 .68 .78 .a4 .a2 .76 .61 .4i 150 a OVBI 100 .31 .27 .27 .26 .25 .27 SO .63 .72 .74 .69 .54 Sl .44 .40 .37 .34 .36 .41 .47 .54 .57 .60 .5f .33 .28 .25 .23 .23 .35 .50 .64 .74 .77 .70 .55 .53 .44 .37 .35 31 .33 .39 .46 .55 .62 .64 .6( .29 .21 .18 .I5 .14 .27 .50 .69 .a2 .87 .79 .60 .48 .32 .25 .20 .I7 .I9 .39 .56 .70 30 .79 .6( Southwest .44 .39 .36 .33 .31 .31 .35 .42 .49 .5 .44 .39 .34 .31 .29 .28 .33 .43 .51 .5i .38 .28 .22 .I8 I.16 .19 .33 .52 .69 .7; Weat Northwest 30 North ond Shade 150 8 over 100 30 Equation: Cooling Load, .39 .36 33 .30 .28 .26 .26 .30 .37 .4d .41 .35 I.31 .28 .25 23 .24 .30 .39 .41 22 .33 .25 .20 ,113 .I5 .14 .13 .I9 .41 .64 .80 .75 .53 .36 28 .24 .19'.17 .I5 .I7 .30 50 .6( . 9 6 . 9 6 . 9 6 . 9 6 /. 9 6 I . 9 6 I. 9 6 I . 9 6 I . 9 6 I. 9 6 1 . 9 6 I. 9 6 . 7 5 . 7 5 . 7 9 2 3 3 . 8 4 . 8 6 . 8 8 3 8 1 . 9 1 . 9 2 . 9 3 . 9 : . 9 8 .9a .9a .98 .9a .9a .9a I .9a .9a .9a I .9a .9a .a1 .a4 .a6 .a9 .91 .93 . 9 3 . 9 4 . 9 4 . 9 5 . 9 5 . 9 : 4 I . o o - - - - +b - - __- I .oo ----- Southeast North . Northwest West Soulhwert South and Shade Btu/hr = [Peak solar heat gain, Btu/(hr) (sq ftt), (Table 6)] x [Window area, sq ft] x [Shade factor, Hare factor, etc., (Chapter 411 x [Storage factor, (above Table at desired time)] *Internal shading device is any type of shade located on the inside of the glass. $Bore glass-Any window with no inside shading device. Windows with shading devices on the outside or shaded by external projections are to considered bare gloss. tThese factors apply when maintaining a CONSTANT TEMPERATURE in the space during allowed to swing, additional storage will result during peak load periods. Refer to Table SWeight the operating period. Where the 13 for applicable storage factors. temperature p e r sq f t o f f l o o r - Room on Bidg Exterior (One or more outside walls) = Room in Bldg Interior (No outside walls) = Basement Room (Floor on ground) = Entire Building or Zone = 1% (Weight of Outside Walls, lb) + % (Weight of Partitions, Floor .__and Ceiling, lb) __-___.. Floor Area in Room, sq ft (Weight of Partitions, Floor and __Ceiling, lb1 Floor Area in Room, sq ft (Weight of Outside Walls, lb) + (Weight of Floor, lb) + % (Weight of Partitions and Ceiling, lb) Floor Area in Room, sq ft (Weight of Outside Wall, Partitions, Floors, Ceilings, Structural Members and Supports, J-j Air Conditioned Floor Area, sq ft With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug. Weights per sq ft of common types of construction are contained in TabJes 21 thru 33, pages 66 thru 76. is ,ed (:tl.\l”l‘l-l~ 3. IHI<.\’ s’I‘oII,\c;I:, I~lVEI<Sl’I’Y. .\NI) l-35 S’II~.\‘I‘II~l(:.\‘I‘IoN i”‘L / TABLE 12-STORAGE LOAD FACTORS, HEAT GAIN-LIGHTS* Lights On EQUIP. OPERATION Hour5 10 Hours-1 with Equipment WEIGHT8 ( l b per ~4 f t of floor orea) 0 / 150 8 o v e r Operating 12, NUMBER OF 16 and HOURS 24 Hours, AFTER Constant LIGHTS ARE Space TURNED i Temperature r ON f 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 i / I i I ! I I / , , , , , , 1 1.371.671 . 7 1 (.741.76j.79/.81/.83(.84/.861.87j.291.26 1.23/.20/.19/.17/.151.141.121.11I.10 . I 12 .I1 0 9 .08 0 0 “These factors apply when maintaining CI CONSTANT TEMPERATURE in the space during the operating period. Where the temperature 1s allowed t0 swing, odditionai storage will result during peak load periods. Refer to Table 13 for applicable storage factors. With lights operating the scltne number of hours os the time of equipment operation, use a load factor of .l.OO. tLights On for Shorter or Longer Period than 10 Hours 3. Equipment operating for 12 hours: rlccosionally adjustments may be required to take account of lights ating less or more than the IO hours on which the table is based. I&~= following is the procedure to adjust the load factors: Follow procedure in Step 2, except in Step 2b add values of 12th hour to that designated 0, 13th hour to the 1 st hour, etc. A - W I T H L I G H T S I N O P E R A T I O N F O R S H O R T E R P E R I O D T H A N IO HOURS and the equipment operating 12, 16 or 24 hours at the time of the overall peak load, extrapolate load factors as follows: B-WITH LIGHTS IN OPERATION FOR LONGER PERIOD THAN 10 HOURS and the equipment operating 12, 16 or 24 hours at the time of the overall peak load, extrapolate load factors OS follows: I. Equipment operating for 24 hourr: I. Equipment operating for 24 hours: o. Use the storage load factors as listed up to the time the lights care turned off. a. Use the load factors as listed through 10th hour and extrapolate beyond the 10th hour at the rate of the last 4 hours. b. Shift the load factors beyond the 10th hour (on the right of heavy line) to the left to the hour the lights are turned off. This leaves last few hours of equipment operation without designated load factors. b. Follow the same procedure as in Step lb of “A” except shift load factors beyond 10th hour now to the right, dropping off the lost few hours. c. Extrapolate the last few hours at the same rate of reduction as the end hours in the table. 2. Equipment operating for 16 hours: a. Follow the procedure in Step I, using the storage load factor v(~Iues in 24-hour equipment operation table. b. Now construct a new set of load factors by adding the new values for the 16th hour to that denoted 0, 17th hour to the I st hour, etc. C. The load factors for the hours succeeding the switching-off the lights ore as in Steps 1 b and I c. 2. Equipment operating for 16 hours or 12 hours: a. Use the load factors in 24-hour equipment operation table 0s listed through 10th hour and extrapolate beyond the 10th hour at the rate of the last 4 hours. b. Follow the procedure in Step I b of “A” except shift the load factors beyond 10th hour now to the right. C. For Ibhour equipment operation, follow the procedure in Steps 2b and 2c of “A”. d. For I?-hour equipment operation, follow the procedure in Step 3 of “A”. / t )\ 1-36 , I’:\RT I. LOAD ESTIM:\TING Example Adjust values for 24-hour equipment operation and derive new values for lb-hour equipment operation for fluorescent lights in operation 8 and 13 hours, and an enclosure of 150 Ib/sq ft of ftoor. EOUIP WEIGHT$ O P E R A T I O N ( l b p e r sq f t Hours of floor area) 24 16 150 150 NUMBER OF HOURS AFTER LIGHTS ARE TURNED ON 19 120 21 22 23 LIGHTS N Hours .17 .I5 .14 .12 .ll 13 .lO .09 .OB .07 .06 .I2 .I1 ,.I0 .09 .08 8 10 O 1 3 .37 .37 .37 .67 .71 .b7 .71 .b7 .71 .74 .76 .74 .76 .74 .76 .79 .81 .a3 .84 .79 .81 .83 .84 .79 31 .83 “84 .86 .87 .89 .90 392 .29 .26 .23 .20 .19 .29 ,215 .23 .20 .17 .86 .87 29 .26 .23 .I1 .I5 .I4 .bO .87 .90 .91 .91 .93 .93 .94 .94 .95 .51 .79 .84 .87 .88 .89 .29 .84 .84 .85 .85 .86 .60 .82 .82 .83 .84 4 .85 5 6 9 2 0 7 8 .90 10 11 12 13 .19 14 15 16 17 18 .I5 .I4 .12 .20 .19 .I7 .95 .96 .96 .97 .29 .26 y .23 .20 .19 .17 .I5 .88 %O \.32 .28 .25 .23 .19 13 6 10 $Weight per rq ft of floorRoom on Bldg Exterior (One or more outside walls) = Room in Bldg Interior (No outride walls).= Basement Room (Floor on ground) Entire Building or Zone = (Weight of Outside lb) ___--- - Walls, - - - lb) + % (Weight of Partitions, Floor and Ceiling, __Floor Area in Room, sq ft (Weight of Outside Walls, lb) + (Weight of Floor, lb) + %-.__ (Weight --__ of Partitions and Ceiling, lb) = Floor Area in Room, sq ft (Weight of Outside Partitions, Floors, Ceilings, Structural Members and Supports. lb) -~~. Wall .A--. Air Conditioned Floor Area, sq ft With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug. Weights per sq ft of common types of construction ore contained in Tables 21 thru 33, pages 66 thru 76. SPACE TEMPERATURE t l/z (Weight of Partitions, Floor and - Ceiling, lb)__Floor Area in Room, sq ft SWING In addition to the storage of radiant heat with a constant room temperature, heat is stored in the building structure when the space temperature is forced to swing. If the cooling capacity supplied to the space matches the cooling load, the temperature in the space remains constant throughout the operating period. On the other hand, if the cooling capacity supplied to the space is lower than the actual cooling load at any point, the temperature in the space will rise. As the space temperature increases, less heat is convected from the surface and more radiant heat is stored in the structure. This process of storing additional heat is illustrated in Fig. 11. The solid curve is the actua1 cooling load from the solar heat gain on a west exposure with a constant space temperature, 24-hour operation. Assume that the maximum cooling capacity available is represented by A, and that the capacity is controlled to maintain a constant temperature at partial load. When the actual cooling load exceeds the available cooling capacity, the temperature will swing as shown in the lower curve. The actual cooling load with temperature swing is shown by the dotted line. This operates in a similar manner with different periods of operation and with different types of construction. NOTE: When a system is designed for a temperature swing,. the maximum swing occurs only at the peak on design clays, which are defined as those days when all loads simultaneously peak. Under normal operating conditions, the temperature remains constant or close to constant. Basis of Table 13 - Storage Factors, Space Temperature Swing The storage factors in Table 13 were computed using essentially the same procedure as Tables 7 thru 12 with the exception that the equipment capacity available was limited and the swing in room temperature computed. FIG. 11 - ,~CTUAL COOLING L OAD W ITH V ARYING R OOM T E M P E R A T U R E The magnitude of the storage effect is determined largely by the thermal capacity or heat holding * ; CH/\I”J‘JCJ< : I . I-JJ:.\‘J‘ .S~J‘(>R;\GL’:, J)l\il:JLSI’J‘Y, 1-37 ,\ Xl> S’J‘JI,\-J‘II~I~:,\‘J‘ION TABLE 13-STORAGE FACTORS, SPACE TEMPERATURE SWING Btu/(hr) (deg F swing) (sq ft of floor area) NOTE: This reduction is to be taken at the fime of peak load only. TYPE APPLICATION Load Pattern I A VARIABLE 24-HOUR -‘dldg 150 and OV.3 Bidg Periphery, Except North Side 30 30 I CONSTANT INTERMITTENT P4-HOUR PERIOD Apartment g+\ VARIABLE 24-HOUR CONTINUOUS PERIOD HOUS.3, Hotels, Hospitals Residences (%I I 75 50 25 75 50 25 100 I Interior zonert Department storer, Factories GLASS RATIOS floor area) Type Office INTERMITTENT PERIOD HOURS WEIGHT (Ib/rq f t I 75 50 25 I OF OPERATION 16 1 12 Temperature Swing (F) 1-2 13-4 ( 5 - 6 ( l-2 1 3 - 4 ( 5 - 6 ( 1-2 ( 3 - 4 / 5 - 6 24 1.90 1 1 . 8 0 / 1.65 1 1.80 1 1.70 / 1.55 1 1.60 / 1.50 / 1.40 I 1.70 I 1.60 I 1.45 I 1.60 I 1.50 I 1.35 I 1.50 I 1.35 I 1.25 1.50 1.40 1.40 1.30 1.30 1.20 1.70 1.60 1.45 1.50 1.45 1.35 1.40 1.35 1.30 1.50 1.40 1.30 1.35 1.30 1.20 1.30 1.25 1.10 1.35 1.25 1.20 1.25 1.00 .90 1.20 .95 .70 1.20 1.10 .95 1.00 .95 1.40 1.25 1.00 .88 1.20 .95 .80 1.10 .90 .80 .90 .85 .80 .90 j .80 .70 / .85 / .75 / .60 j 30 / .70 1 .55 I I I I I I I I I 150 and OVW 100 30 - 1.60 1.40 .95 1.55 1.38 .92 1.50 1.36 .90 1.50 1.30 .90 1.45 1.28 .88 1.25 .85 1.35 1.25 .a5 1.20 .80 - 150 and Over 75 50 25 1.85 1.65 1.45 1.75 1.50 - 1.40 - - _ _ - - - 75 50 25 75 50 25 1.55 1.40 1.30 1.20 1.10 .85 1.45 1.35 1.10 .90 .70 1.40 .95 _ _ - _ - - - 100 30 30 - Equation: Reduction in Peak Cooling Load, Btu/hr = (Floor Area, sq ft) X (Desired Temp Swing, Table 4, page 20) X (Storage Factor, above table) *Weight per sq ft of floor may be obtained .from equation on pc~ge 30 iFor 12-hour operation, use a 2 degree max temp swing. $Glars ratio is the percent of glass area to the total wall area. capacity of the materials surrounding the space. It is limited by the amount of heat available for storage. Load patterns for different applications vary approximately as shown in the first column of TnOie 1.3. For instance, an office building has a rather large varying load with a high peak that occurs intermittently. An interior zone has an intermittent peak the load pattern is relatively constant. A hospital, on the other hand, has a constant base load which is present for 24 hours with an additional intermittent load occurring during daylight hours. The thermal capacity of a material is the weight times the specific heat of the material. Since the specific heat of most construction material is approximately 0.20 Btu/(lb)(F), the thermal capacity is directly proportional to the weight of the material. Therefore, the data in the tables is based on weight of the materials surrounding the space, per square foot of floor area. Use of Table 13 - Storage Factors, Space Temperature Swing . -5 Table 23 is used to determine the reduction in cooling load when the space temperature is forced to swing by reducing the equipment capacity below that required to maintain the temperature constant. This reduction is to be subtracted from the room sensible heat. NOTE: This reduction is only taken at the time of cooling load. Example 3 - Space Temperature Swing Given: Find: The actual cooling load at 4 p.m. from sun. lights, and people with 3 F temperature swing in the space. Solution: The peak sensible cooling load in this room from the sun, I igh ts, and people (neglecting transmission inliltration. ventilation and other internal heat gain) is 5700 + 5190 = 10,590 Btu/hr. (Esavtf11es 1 and 2.) NOTE: The peak cooling load in this room occurs at approximately 4 p.m. The solar and light loads are almost at their peak at 4 p.m. Although the transmission across the large glass window peaks at about 3 p,m., the peak infiltration and ventilation load also occurs at 3 p.m. and the relatively small transmission load across the wall peaks much later at about 12 midnight. The sum of these loads re- ’ Since the nortual t h e tlwrtnostat scttina i s ;~i)out 7 5 1: o r iti I: (II), tlcsijin temlxrattlr~ (78 1: = 73 T: thcrnwstat setting + :I F s w i n g ) OCCIII-s o n l y on tlcsigrl lxxk days a t t h e t i m e o[ peak loacl. Untler partial load olxzl.ation, tllc r00l11 perature is ktwcen 75 I: dl) a n d i8 I; tll), o r a t the tCI11- thertnostat s e t t i n g (7.5 1:) , tlclwn~lin~ on the loatl. PRECOOLING STORAGE AS A MEANS OF INCREASING Precooling a space below the temperature norinally desired illcl-eases the stoqe of 1~eat a t t h e time of peak load, only when the precooling temperature is maintained as the control point. This is because the potential temperature swing is increased, thus adding to the amount of heat stored at the time of peak load. Where the space is precooled to a lower temperature and the control point is reset upward to a comfortable condition when the occupants arrive, no additional storage occurs. In this situation, the cooling unit shuts off and there is no cooling during the period of warming up. When the cooling unit begins to supply cooling again, the cooling load is approximately up to the point it would have been without any precooling. Precooling is very useful in reducing the cooling load in applications such as churches, supermarkets, theaters, etc., where the precooled temperature can be maintained as the control point and the temperature swing increased to 8 F or 10 F. DIVERSITY OF COOLING LOADS Diversity of cooling load results from the probable non-occurrence of part of the cooling load on a design day. Diversity factors are applied to the refrigeration capacity in large air conditioning systems. These factors vary with location, type and size of the application, and are based entirely on the judgment of the engineer. Generally, diversity factors can be applied to people and light loads in large multi-story office, hotel or apartment buildings. The possibility of having all of the people present in the building and all of the lights operating at the time of peak load are slight. Normally, in large office buildings, sonic people will be away from the office on other I,usincss. i\lso, the lighting arrangement will freclucntly 1x2 such that the lights in the vacant ofkes will not be on. In addition to lights being elf becaiisc the people are not present, the normal maintenance procedure in large office buildings usually rest110 in some lights being inoperative. Therefore, a diversity Iactor on the people a n d light loads shoultl be applied for selecting the proper size refrigeration ecluilment, The silt of the diversity factor depends on the size of the I~uiltling and the engineer’s judgment of the circumstances involved. For example, the diversity factor on a single small office with 1 or 2 people is 1.0 or no reduction. Expanding this to one iloor of a building with 50 to 100 people, 5y0 to lO(70 may be absent at the time of peak load, and expanding to a 20, SO or 40-story building, 10% ’ to 207’ may be absent during the peak. A building with predominantly sales offices would have many people out in the normal course of business. This same concept applies to apartments and hotels. Normally, very few people are present at the time the solar and transmission loads are peaking, and the lights are normally turned on only after sundown. Therefore, in apartments and ho$els, the diversity factor can be much greater than with office building!. These reductions in cooling load are real and should be made where applicable. Table 14 lists some typical diversity factors, based on judgment and experience. TABLE I$-TYPICAL DIVERSITY FACTORS FOR LARGE BUILDINGS (Apply to Refrigeration Capacity) DIVERSITY FACTOR TYPE OF APPLICATION Office Apartment, Department Hotel Store Industrial* People Lights .75 to .90 .70 to .05 .40 to .60 .30 to .50 .BO to .90 .90 to 1.0 .85 to .95 .BO to .90 Equation: Cooling Load (for people and lights), Btu/hr = (Heat Gain, Btu/hr, Chapter 7) X (Storage Factor, Table 121% (Diversity Factor, above table) *A diversity factor should Refer to Chopfer 7. also be applied to the machinery load. Use of Table 14 - Typical Diversity Factors for Large Buildings The diversity factors listed in Table 14 are to be used as a guide in determining a diversity factor for any particular application. The final factor must necessarily be basctl on judgment of the effect of the m a n y variables irlvolvetl. STRATIFICATION OF HEAT \ There are generally two situations where heat is stratified and will reduce the cooling load on the air conditioning equipment: 1. Heat may bc stratified in rooms with high ceilings where air is exhausted through the root’ or ceiling. 2. Heat may be contained above suspended ceilings with rccessecl lighting and/or ceiling plenum return systems. The first situation generally applies to industrial applications, churches, auditoriums, and the like. The second situation applies to applications such qflice buildings, hotels, and apartments. With t,uch cases, the basic fact that hot air tends to rise makes it possible to stratify loads such as convection from the root’, convection from lights, and convection from the upper part of the walls. The convective portion of the roof load is about 25% (the rest is radiation); the light load is about 50% with fluorescent (20% with incandescent), and the wall transmission load about 40’%. In any room with a high ceiling, a large part of the convection load being released above the supply air stream will stratify at the ceiling or roof level. Some will be induced into the supply air stream. Normally, about 80% is stratified and 20y0 induced in the supply air. If air is exhausted through the ceiling or roof, this convection load released above the supply air ~riay IX subtracted from the air conditioning load. This results in a large reduction in load if the air is to be exhausted. It is not normally practical to exhaust more air than necessary, as it must be made u p by bringing outdoor air through the apparatus. This usually results in a larger increase in load than the reduction realized by exhausting air. Nominally, about a 10 F to 20 F rise in exhaust air temperature may be figured as load reduction if there is enough heat released by convection above the supply air stream. Hot air stratifies at the ceiling even with no exhaust but rapidly builds up in temperature, and no reduction in load should be taken where air is not exhausted through the ceiling or roof. With suspended ceilings, some of the convective heat from recessed lights flows into the plenum space, Also, the radiant heat within the room (sun, lights, people, etc.) striking the ceiling warms it up and causes heat to flow into the plenum space. These sources of heat increase the temperature of air in the plenum space which causes heat to flow, into the underside of the floor structure above. When the ceiling plenum is used as a return air system, some of the return air flows through and over the light fixture, carrying more of the convective heat into the plenum space. Containing heat within the ceiling plenum space tends to “flatten” both the room and equipment load. The storage factors for estimating the load with the above conditions are contained in Table 12. 141 CHAPTER 4. SOLAR HEAT GAIN THRU GLASS SOLAR HEAT - DIRECT AND DIFFUSE The solar heat on the outer edge of the earth’s nosphere is about 445 Btu/(hr)(sq ft) on December 21 when the sun is closest to the earth, and about 415 lStu/(hr)(sq ft) on June 21 when it is farthest away. The amount of solar heat outside the earth’s atmosphere varies between these limits throughout the year. The solar heat reaching the earth’s surface is reduced considerably belsw these figures because a large part of it is scattered, reflected back out into space, and absorbed by -the atmosphere. The scattered radiation is termed &/fuse or sky radiation, and is more or less evenly distributed over the earth’s surface because it is nothing more than a reflection from dust particles, water vapor and ozone in the atmosphere. The solar heat that comes directly through the atmosphere is termed direct ~ndirrtion. The relationship between the total and the direct and diffuse radiation at any point on -rth is dependent on the following two factors: _ 1. The distance traveled through the atmosphere to reach the point on the earth. 2. The amount of haze in the air. As the distance traveled or the amount of haze increases, the diffuse radiation component increases but the direct component decreases. As either or both of these factors increase, the overall effect is to reduce the total quantity of heat reaching the earth’s surface. ORDINARY dow. The direct radiation component results in a heat gain to the conditioned space only when the window is in the direct rays of the sun, whereas the diffuse radiation component results in a heat gain, even when the window is not facing the sun. Ordinary glass absorbs a small portion of the solar heat (5 % to 6y0) and reflects or transmits the rest. The amount reHected or transmitted depend? on the angle of incidence. (The angle of incidence is the angle between the perpendicular to the window surface anti the sun’s rays, Fig. 18, page 55.) At low angles of incidence, about 86% or 87oj, is transmitted and 870 or 9% is reHected, as shown in Fig. 12. As the angle of incidence increases, more solar heat’ is reflected and less is transmitted, as shown in Fig. 13. The total solar heat gain to the conditioned space consists of the transmitted heat plus about 40% of the heat that is absorbed in the glass. GLASS Ordinary glass is specified as crystal glass of single thickness and single or double strength. The solar heat gain through ordinary glass depends on its location on the earth’s surface (latitude), time of day, time of year, and facing direction of the win- FIG. 12 - REACTION ON SOLAR HEAT (R) , ORDINARY G LASS, 30” ANGLE OF INCIDENCE 142 L I’/\RT 1 . LO.\11 U’l‘l hI,\~I’IN(; i s tyl)ic:tl I’or wootl s;lsli \vintlows. For met;11 s;141 windows, the glass arca is assumed CCI~I;I~ to IOO(;{, ol’ t h e sxli ;~rca Ixxauscthe c~ontluctivity of t h e nlctal s a s h i s very high 211d tlic s o l a r heat ai~sorbctl i n the sash i s transinittctl almost instant;~neously. l‘liis .40x.O6R Heat &in to Space = (.40 x .OG R) + .42 R = ,444 R or .44 R ! ABSORBED T REFLECTED T %%!A Gkf2: 7 .42R TRANSMITTED I Frc. 13 - RUCTION ON NOTE: The sash arcn equals apl~rositnatcly 85:{, o f t h e masonry opening (or frame opening with frame walls) w i t h wootl s a s h w i n t l o w s . Soft, o f m a s o n r y o p e n i n g with tl011l)le hung metal sash w i n t l o w s , nntl 100% of masonry opening with casement wintlows. SOL~K HEA.~ (R) , ORDINARY G L A S S , 80” A N G L E OF INCIDI<NCI: 2 . S o hale iti tlie air. KOTE: of the al)sorl~ctl solar llcat ,going i n t o the space is tlcrivctl from tllc following rca\oning: The .10”,, 1. The onttloor liim coefficient is approximatelv 2.8 Utu/ (hr) (sq ft) (deg F) with a 5 mph wintl velocity <luring the SUlll”VX. 2. The inside film coefficient is approsimately 1.S ntu/ (hr) (sq ft) (deg F) because, in the average system tlesign, air velocities across the glass arc approximarely 100-200 fpm. 3. If outtloor temperature is equal to room temperature. the glass tcmpcraturc is alcove I)oth. Thcr-cfol-c al~orl)etl heat 2.8 x 100 .\l)sorl)etl heat flowing out = ___, ,‘i + o,9 = fiO.X~~,. or c;w,. I i< 4. .\s tile outtloor tempcratt~re rises. tile ghs tetllpCl‘;ltlll’C also rises. tatlsing more of the al~sorl~ctl heat to flow into the space. This (an IX ac’c‘ountctl for 1)~ atitling the trans. mission of heat across the glass (causctl I)) tcmperatIIre tlilference I)ctwccn i n s i t l c ant1 nuttlool-s) to the constant -lOc<, of the al)sorl~ctl heat going insitlc. 5. This reasoning npplics equally well when the outtloor tempcraturc is I~CIOIV the room tcmperatllrc. Basis of Table 15 - Solar Heot Gain thru Ordinary Glass Table 25 provides data for 0”, lo”, 20”, SO”, qO”, and 50” latitutlcs, for each month of the year and for each hour of the day. This table includes the clircct a n d clitfuse radiation and that portion of the heat absorbed in the glass which gets into the space. It does not incluclc the transmission of heat across the glass causctl by a temperature difference between the outdoor and inside air. (See Chcrpter 5 for “U” values.) The data in T(rble li is basccl on the following conditions: - 3. Sea level elevation. 4. A sea level tlewpoint temperature of 66.8 F (95 F clb, 55 F wb) which approximately corresponds to 4 centimeters of precipitable water vapor. Precipitable water vapor is all of the water vapor in a column of air from *a level to the outer edge of the atmosphere. If these conditions do not apply, use the correction factors at the bottom of each page of Table 15. Use of Table 15 - Sotar Heat Gain thru Ordinary Glass The hold face values in Table 15 indicate the maximum solar bent guin for the month for each exposure. The bold face values that are boxed indicate the yearly maximums for each exposure. Table 15 is used to determine the solar heat gain thru ordinary glass at any time, in any space, zone or building. To determine the actual cooling load due to the soln~ heat gain, refer to Chapter 3, “Heat Storage, Diversity and Stmtificntion.” CAUTION - Where Estimoting.Mu/ti-Exposure or Buildings Rooms If a haze factor is used on one exposure to determine the peak room or building load, the diffuse component listed for the other exposures must be divided by the haze factor to result in the actual room or building peak load. This is because the diffuse component increases with increasing haze, as explainctl on @Se $1. . (:H.\l”l‘~l< 4. SOL/\R IHk:.\‘I‘ (i.\IN ‘I‘HliU 143 (iI..\SS Example 2 - . (Bottom Table 75) .l’llc ccllltlilicur on rvllich Y‘d/~/r 1s Solar heat gain Scptcml,cr 22 \\‘cs t SOClll~ 2:oo !)!I I IO 3:oo I39 XI 7‘01;11 209 “20 2:oo 88 13; 225 3:oo ] ~>‘-> 104 Novemlwr 21 WCS t South 2:oo 74 I39 3:oo 100 I o-1 Total 21.3 204 4:oo p.m. 149 44 I!):1 Solar heat gain Octolxr 23 \\‘cst South Total 226 4:oo p.m. 117 59 156 Solar heat gain 4:oo I’.“‘. 91 59 130 The peak solar heat gain to this room occurs at 3:00 p.m. on Octolxr 23. The peak room cooling load tlocs not necessarily occur at the same time as the peak solar heat gain, - Solution: By inspection of Table 15 the I~oxed boldface valt:es for peak solar heat g a i n , occurring at 4:00 p.m. on July 23 = lG4 I!du/(hr) ci . . /5 i s Islwtl (11, not npl)ly to ’ . I:intl: I’cak solar l1cat pill. Soltltion: I:rom T,l/J/f, ? 1 * Solar Gain Correction Factors (sq fr) ;\ssumc a somewhat Ilazy condition. .\ltitutlc correction = l.OOi (Iwttom Tnblr IS) Dewpoint tlilfcrencc = 09.8 - 66.8 = 3 F Dcwpoint correction = 1 - (3/10 X .Oi) = -979 (bottom Table 15)’ Haze correction = I - .I0 = .90 (Iwttom Table 15) Steel sash correction = I /.85 (hottom Tnble 15) Solar heat gain at 4:00 p.m., July 23 = 164 X 1.007 X ,979 X .90 X I/.% = 171 Btu/(hr)(sq ft) ,’ l-44 I’.\R’I‘ I. 1.0 \I) 1-~5’l’l.\I.\‘i‘l~(; TABLE 15-SOLAR O0 0” HEAT GAIN THRU ORDINARY GLASS O0 Btu/(hr) (sq ft sash area) NORTH Time of Y e a r J U N E 21 LATITUDE Exposure 1 AM 6 7 North Northeast 0 0 East 0 Southeast South Southwest west I Northwest 0 0 101 101 101 SUN 8 45 II9 II6 9 I 37 b b 5 5 - 4; I ;; I ii I MAR 22 FEB 20 J A N 21 Southwest West Northwest lol 101 lo1 0 0 ;j 13 13 147 II II 87 52 II II II II 91 I2 118 167 IIB 12 I2 I2 t )I I2 100 I2 79 Iii 141 28 I2 t I2 t t 12 97 t - I8 14 I4 14 14 195 3 4 5 78 1 141 74 t 131 65 t 45 / 0 III 51 0 I3 IO1 I51 IO1 I 13 I3 I3 I3 I63 I3 AC; Iii I33 31 I I3 I 13 I 13 1 I50 I4 14 I4 14 lb 223 14 14 I4 I4 43 233 Ii I 1’4 I 1: I I! I ii 36 52 46 Id Cl ii 411 1 QA 139 152. I21 150 153 118 86 Iii 223 195 1 I51 1 91 1 29 Steel Sash, or No Sash X l/.85 o r I.17 Exposure I4 I4 I4 I4 I4 I3 I2 6 58 31 I4 14 I4 I3 I2 6 107 47 I4 14 I4 I3 12 5 68 31 14 I4 I4 13 I2 6 I 141 141 14 I 14 I 14 ) 13 ) 12 I 6 14 I4 I4 31 68 IO1 IIB 95 14 I4 I4 47 107, 151 1 6 7 1 3 4 14 1 4 1 4 31 68 101 II8 95 210 2 4 0 2 5 0 240 210 I63 100 32 I4 I4 I4 I4 I4 13 12 6 75 is 14 14 14 13 -12 6 ;4 ii I4 13 I2 5 IOJ ;b 102 51 24 I4 I4 13 I2 6 33 34 34 34 33 31 2% I7 I I41 i4 I 24 1 61 ( 102 1 133 I 141 I 110 1 141 141 14 i 46 ) 103 ) 148-l 163 i 129 1 141 141 14 I 15 / 35 I k-5 I 79 I 67 (2061 2341 245 1234 1206 1 I50 1 97 1 31 141 I IAI 96 124 55 14) 141 141 14 i 14 I 14 1 13 1 I I? I. 141 43 85 bb 16) 141 II 1 i4 I 14 I 14 I 13 I II 1 4 I 14 1 1 4 1 1 3 1 II I Haze Altitude - 15% ( M a x . ) +0.7% per 1000 Ft Bold Face Values - Monthly Maximums j 1 East 0 I ( 1 1 1 1 0 0 0 0 0 0 I . South I FEB 20 & OCT 23 I South Southeast Eari Northeast MAR 22 I ) North Northwest west Southwest Horizontal South southeast 28 I 0 East APR 20 Northeast & North I Northwest 1 West I Southwest 1 Horizonta! AUG 24 I I I Northeast JULY 23 I I I Northeast J U N E 21 North Northwest west Boxed Vaiues - Yearly maximums I Horizontal Dewpoint Increase From 67 F - 7 % p e r IO F M A Y 21 & North 1 Northwest I West I Southwest ( Horizontal 37 I 0 I Southwest Dewpoint Decrease From 67 F + 7% per IO F ’ & SEPT 22 6 l 0 1 South I b I o I Southeast 0 East 0 0 0 0 & N O V 21 6 6 45 II9 II6 j J A N 21 0 southeast 0 Eest 0 Northeast 0 North 0 Northwest 0 west 1 0 I southwest 0 Horizontal 0 0 0 0 ( o 0 0 0 0 0 0 0 0 0 I o 1 0 I o 1 0 DEC 22 South Southeast 5 1 0 1 south I 6 I o I Southeast 6 0 East II II 6 54 37 1 I53 1 I I8 1 152 I 121 52 1 45 ( 91 1 29 I ,j I 7; I I; I 141 14 I 14 I 14 I 13 I II 147 I ,; 4; 4i I4 I4 I4 I3 II I56 I54 I33 95 53 20 I4 13 II b5 74 78 ' 80 82 80 78 74 65 II I3 I4 20 53 95 133 154 156 II I3 14 I4 14 I 43 93 I35 147 II I 13 I 141 141 14 I 14 I I5 I 27 I 42 I 87 147 I91 217 2 2 6 217 I91 147 87 1 South Southeast East Northeast North Northwest west Southwest Horizontal Time o f Year Northeast 1 : 1 North iI Northwest 0 west 0 Southwest 1 0 1 Horizontal 34 34 33 I 31 I 28 I 17 I 24 14 14 I3 12 6 14 14 I4 I3 12 6 I4 I4 I4 13 12 6 I4 I4 14 13 I2 5 14 I5 35 65 79 67 I4 46 103 148 lb3 129 24 1 51 I102 II33 I I41 1 II0 245 234 206 I50 97 3I I4 I4 I4 I3 43 I6 I4 I3 67 56 65 51 ‘?3 ) 86 ) 124 1 150 m II I 1 3 I 1 4 1 1 4 1 1’4 I 43 I 96 I I39 I II I I3 I I41 I41 I4 I I4 I I8 I 35 I 91 ( l5i II95 ( 2231 233 1223 ( I95 1 I51 II 6 i4l 33 34 102 61 103 46 35 15 I4 I4 I4 14 14 I4 I I41 141 206 234 II I 13 I ;j I ?C, I 15; 1;; 153 150 54 61 I I ) I3 1 >I 0” SOUTH LATITUDE PM 2 14 14 14 191 36 13 13 I3 I3 I51 --I Solar Gain Correction I 14 1 14 1 14 1 13 1 II 1 6 1 0 ;i ii 14 14 14 43 20 53 95 ‘133 I54 -I&. 119 0 217 226 217 191 147 87 28 0 61 55 bb 67 55 55 bl 54. 37 0 I50 124 I ihl 43 I lb I 14 I 13 I II I 5 I 0 9 5 1 iii 14 1 I3 I I I I 5 10 I39-- .-. I ; I 14 I ;; 31 I;3 148 55 12 I! I2 I3 12 I I3 II2 1 I3 97 I50 Southwest west TIME Noon J.g 78 APR 20 II 65 1 74 1 78 t 801 82 1 80 1 156115411~11951 201 I:; 1521 M A Y 21 IO I South Lat. Dec. or Jan. + 7% CM;\1”1‘1:1< 1 . l-45 SOl..\I< lrI<,\.I‘ C.,\IN ‘l‘H1111CI,.\SS TABLE 15-SOLAR HEAT GAIN THRU ORDINARY GLASS (Contd) 10° Btu/(hr) IO” N O R T H Time of Year LATITUDE ,-- Exposure 6 North Northeast East southeast 1 J U N E 21 South Southwest west Northwest Horizontal North Northeast JULY 23 East -So”t~~east I. & South _ Southwest MAY 21 west Northwest Horizontal North Northeast AUG 24 East Southeast & South Southwest APR 20 west ’ Northwest Horizontal SEPT 22 & MAR 22 OCT 23 & FEB 20 N O V 21 & J A N 21 DEC 22 Solar Gain Correction 19 55 54 18 2 2 2 2 4 5 42 50 26 I I I I 3 I 17 25 18 I I I I 2 AM i 7 i.8, 44 131 I34 49 8 8 8 8 44 34 127 135 57 7 7 7 7 42 15 II3 138 79 7 7 7 7 38 S U N IO I I Noon 44 106 98 25 I4 I4 I4 14 205 33 109 98 32 14 14 I4 14 210 I5 80 104 60 I4 I4 I4 14 213 43 41 43 44 65 28 14 14 41 14 14 14 14 I4 14 I4 I4 14 14 I4 14 14 14 25 I4 14 41 98 I8 28 65 106 233 243 233 205 31 30 3 I 33 56 22 14 14 43 14 I4 14 I4 I4 14 14 14 14 14 14 I4 I4 14 32 I4 I4 43 98 14 22 56 109 236 247 236 210 14 14 14 15 34 I4 14 14 46 I4 I4 I4 27 I4 14 14 14 14 I4 14 I4 I4 27 60 I4 I4 46 80 1414 3 4 242 250 242 213 ‘\ 9 50 45 153 140 155 139 55 43 II 13 8 13 8_ 13 8 I3 107 166 39 35 148 133 158 142 66 56 I I 13 II I3 II I3 I/ 13 107 lb6 16 15 130 III 163 149 94 85 II 13 II 13 II 13 II I3 I05 I67 I 6 I I I3 14 I 89 103 80 45 . I 130 I64 I51 106 I 97 127 I22 94 I b 13 I9 24 I 6 II 13 14 I 6 II 13 14 I 6 II 13 14 II 31 97 lb0 1 207 North Northeast East Southeast South Southwest west ’ Northwest Horizontal North Northeast East Southeast South Southwest \ West ‘\ Northwest Horizontal North 0 5 0 58 0 II8 0 103 0 18 0 5 01 5 0 5 0 22 Northeast East Southeast South Southwest I west ’ Northwest Horizontal North Northeast East Southeast South ._ S o u t h w e s t , west ’ Northwest Horizontal Steel Sash. or No Sash x l/.85 o r I.17 0 4 9 00 0 0 0, 0 0 0 0 0 0 01 0 0 0 0 0 99 27 99 35 4 4 4 I7 4 I5 86 143 37 I53 65 9 9 9 62 9 28 137 154 74 9 9 9 66 ‘5; 4 4 4 I4 12 132 17 161 91 12, 12 12 131 12 I7 130 1631 94 I2 12 12 120 (Max.] I I M 2 3 4 5 4 5 -5& 13 II 13 II 13 II 13 II 43 55 139 155 140 1 5 3 166 107 35 39 13 Jl I3 II I3 II 13 II 56 66 142 158 133 148 Ibb 107 I5 16 I3 II I3 II I3 II 13 II 85 94 149 l b 3 III 15 130 I67 I05 I2 12 9 12 9 91 65 Ibl I53 132 143 I7 37 I31 62 12 9 12 9 I2 9 I2 9 94 74 1 6 3 I54 130 137 17 28 120 66 Altitude +0.7% Bold Face Values - Monthly per IO00 F t Maximums I2 9 Values 6 SOUTH ; 2 2 18 54 55 4 34 5 7 I 7 i 7 I I 7 I 57 26 135 50 127 42 42 3 I5 I 7 i 7 I 1 7 7 / 79 I8 I38 25 II3 I7 38 2 6 1 6 6 6 6 97 130 89 3 I , 5 5 5 5 I8 103 II8 1 58 22 4 1 I I I I 0 0 0 0 0 0 0 0 0 0 South Southeast East Northeast North Northwest ‘Nest Southwest Horizontal South Southeast East Northeast North .Northwest west Southwest Horizontal South 4 4 35 99 99 27 I7 4 4 4 4 50 99 86 I5 I4 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 East Southeast No&east North Northwest west Southwest Horizontal south Southeast East Northeast North Northwest west Southwest Horizontal - Yearly LATITUDE Exposure South Southeast East Northeast North Northwest west Southwest Horizontal South Southeast East N ortheast North \ Northwest west Southwest Horizontal South Southeast East N ortheast North Northwest west Southwest Horizontal 44 8 8 8 8 49 I I34 131 44 Dewpoint Decrease From 67 F + 7% p e r I O F Boxed I O” PM I3 93 39 14 14 14 14 14 I3 146 109 70 31 I7 96 104 106 104 96 I7 31 70 109, 146 13 I4 14 39 93 I3 14 14 I4 I3 175 202 210 202 I75 13 I4 14 14 I3 I3 14 14 I4 I3 91 42 I4 14 I3 149 121 79 3 6 ) 2 3 109 II6 120 I I6 109 23 36 79 I21 149 I3 I4 I4 42 91 I3 I4 I4 I4 13 lb7 193 202 193 167 13 14 I4 14 17 14 14 47 ‘I4 14 56 21 14 27 28 27 14 21 56 14 14 47 I4 14 I7 235 247 235 I I4 I4 14 14 14 I4 40 I4 I4 81 46 I8 71 73 71 I8 46 81 I4 I4 401 14 14 14 220 230 220 14 I4 I4 E 13 II 13 II 13 II 13 II 19 13 122 127 I51 I64 80 103 I60 97 13 IO I3 lo 13 10 I3 IO 55 40 149 147 1 4 5 I55 44 66 139 85 Haze - 15% T 14 I4 I4 14 24 94 106 45 207 I4 14 14 I4 65 I23 100 28 193 I3 IO !3 14 66 44 28 155 I45 100 147 149 I23 40 55 65 IO I3 I4 IO I3 I 14 IO I3 14 85 139 193 10° (sq ft sash area) 2 I I I Dewpoint( Increase From 67 F - 7% p e r IO F maximums Time of Year DEC 22 . J A N 21 & N O V 21 _ FEB 20 & OCT 23 I MAR 22 & SEPT 22 : L , APR 20 1 & ! AUG 24 : M A Y 21 i’ & JULY 23 JUNE 21 South Lat. Dec. or Jan. + 7% I l-4 6 I’AR’I I. l.O.\l) l~.4’l’l,\l,\‘l’lN(; TABLE 15-SOLAR HEAT GAIN THRU ORDINARY GLASS (Contd) 2o” 20” NORTH LATITUDE Time of Year M A Y 21 SEPT 22 MAR 22 7 Southeast 1 South Southwest west Northwest Horizontal I North Northeast I 0 I west NOV 21 1 I I 19 I Northwest Horizontal North Northeast East Southeast 01 41 4 : 4 0 1 I8 1 0 1 3 I ( 1 I ~~ 91 121 9 9 12 I2 68 I127 1 8 1 II 1 0 0 131 I; 1 North Northeast 1 0) 51 4 8 I 0 I IO1 7 I II 2 I I I Southeast South Southwest 0 0 0 0 Northwest Horizontal ‘74 i;i 7 II 20 7 Ii I2 7 II I2 36 92 I35 2 2 2 4 Steel Sash, or No Sash Haze l/.85 o r 1.17 I -15% (Max.] I Bold I 2 17 I 19 3 I Values 28 62 66 73 143 160 I48 122 144 154 176 I21 60 17 23 28 13 I2 8 13 12 8 13 I2 8 13 12 780 79 85 145 lb3 148 ill 138 132 175 II8 55 13 II IO 13 II 7 13 II 7 I3 /I 7 I4 II 7 I08 II3 89 149 165 142 89 118 III 167 107 48 I3 II 6 I3 II 6 13 II 6 6 13 II 1 38 1 22 I 8 I 140 1 136 1 99 ] I:49 1 lb3 1130 1 Exposure I4 I3 13 72 .9 I2 9 -4 4 Time of Year South _ 28 81 81 II 20 3 3 3 Northwest west Southwest Horizontal South Southeast East Northeast North Northwest J A N 21 & 3: 75 west 71 Southwest 8 Horizontal 6 South 2 Southeast 2 East 2 Northeast 2 North 29 Northwest 53 west 45 Southwed ; Horizontal 0 South 0 Southeast 0 East 0 Northeast 0 I North 0 1 Northwest 0 1 West 0 - _ NOV 21 . FEB 20 & OCT 23 MAR 22 . t & SEPT 22 . South Southeast 0 1 0 *#I,2 .I‘ 100 I3 I71 I41 29 127 52 ~~ b8 44 I8 0 Horizontal Ii I3 II 8 3 0 South 91 141 91 I3 13 180 I3 i; 46 136 135 4.3 13 172 I7 _ II II II 100 lb4 127 14 IO1 II II II 11 III lb7 I21 I2 92 88 8 69 144 I28 26 48 7 7 7 7 74 139 II8 18 36 3 3 3 28 73 71 24 5 2 2 2 2 25 59 56 14 4 00 0 0 0 0 0 0 0 0 0 0 0 0 0 n 6 East Southeast Northeast North Northwest west Southwest Horizontal _ I3 lb 123 158 91 13 I46 12 12 I3 I I 1; I i; I 1; 135 I36 46 ‘2 6 41 I 20” SOUTH LATITUDE _ I4 ii 60 I3 I3 I61 97 I, I2 I3 170 11*.1 I ;; I46 134 :; 132 159 85 12 135 76 7; I61 & J U L Y 23 East Northeast North Northwest J U N E 21 west -tauthwrct __..... __. Ho+n,+zl I + 7% p e r I O F I Maximums M A Y 21 South Southeast Dewrxint I Dew& 1 Decrease’From 67 F I Increase ko”,’ +0.7yo p e r IO00 F t 1 - Monthly . ~~- 5 74 1 II9 1 149 1 lb0 1 I46 1 91 I Face I41 4 25 I 33 I I APR 20 & ._ 0 West ~-~ 131 I46 172 I2 I 131 0 0 East , ;i 1 -3 1 77 1144 IIhA IlGR Horizontal X I I I3 13 14 I4 14 I4 49I4 171 1 1961 208 1196 , Solar Gain Correction 15 13 1 271 J A N 21 22 Noon 171 I 4 I 9 I 12 I I3 44 52 29 Ii I 141 IOC 99 147 I l41- I 0 Southeast South Southwest FEB 20 IO 251 I 0 East & 9 PM T I M E 4 I; 14 14 14 ii 1; 44 9 12 14 i4 14 I4 41 96 9 12 14 14 14 15 38 83 60 121 176 216 232 250 232 216 28 23 17 15 I4 14 I4 I5 132 138 III 73 31 14 I4 I4 I48 lb3 145 99 46 I4 14 14 70 85 79 57 29 I4 14 I4 8 I2 13 I4 I4 14 I4 14 8 12 I3 14 14 14 29 57 8 12 13 14 14 14 46 99 8 I2 I3 I4 1 4 14 31 73 8 55 II8 175 216 240 251 240 216 6 IO II I3 I4 14 14 14 I4 45 111 118 89 50 18 14 14 14 53 142 lb5 149 106 51 I4 14 29 89 113 I08 98 55 lo” I4 14 2 7 II 14 20 24 26 24 20 2 7 II I3 14 14 20 55 98 2 7 II 13 I4 14 I4 51 I06 2 7 II I3 I4 14 14 18 50 5 48 107 167 210 235 247 235 210 0 6 II I3 14 14 I4 14 14 0 83 87 59 22 I4 I4 I4 14 0 I30 lb3 149 104 45 I4 14 14 0 99 136 140 120 84 41 I5 I4 1 0 1 8 1 22 1 38 t 52 1 63 t 65 1 63 I 52 1 0 1 6 t I I t I3 1 I41 I51 41 1 84 1 120 t 1 0 1 6 1 I I I 13 1 14 1 I41 I4 1 45 II04 1 East O C T 23 8 3 3 3 II 20 71 75 31 3 3 3 3 Southeast South Southwest west Northwest Horizontal North Northeast APR 20 6 S U N Southwest west Northwest Horizontal East & AM 28 41 33 81 154 144 81 148 IbO- North Northeast A U G 24 1 North Northeast East North Northeast East Southeast South Southwest west Northwest Horizontal & & I Exposure J U L Y 23 DEC 2o” Btu/(hr) (sq ft sash area) Boxed 1 I Values - Yearly 7% maximums per 67 F Iii’ I I South Lat. n-r w J a n . --+:‘70 (:tl,\I”I‘ER ,k. SOl>.\l< ill:.\~I’ <;/\lN ‘I‘[-llilJ TABLE 15-SOLAR HEAT GAIN THRU ORDINARY GLASS (Contd) 3o” 30” 3o” Btu/(hr) (sq ft sash area) NORTH T i m e o f Year J U N E 21 J U L Y 23 & M A Y 21 A U G 24 & A P R 20 SEPT 22 & M A R 22 O C T 23 & FEE 147 (;l,,\SS 20 N O V 21 & J A N 21 DEC 22 Solar Gain Correction LATITUDE A M S U N T I M E b 7 8 9 North Northeast East Southeast South Southwest west Northwest Horizontal North Northeast I East Southeast South Southwest west Northwest Horizontal North Northeast East Southeast South Southwest west Northwest Horizontal North Northeast East Southeast South Southwest west Northwest Horizontal 33 105 108 42 5 5 5 5 I9 22 93 100 42 4 4 4 4 I5 6 55 66 37 2 2 2 2 b 0 ~ 0 0 0 0 0 0 0 0 29 139 156 75 IO IO IO IO 61 20 131 I55 82 9 9 9 9 bb 8 108 147 98 8 8 8 8 47 5 74 124 98 9 5 5 5 25 18 130 lb1 90 12 I2 12 12 131 I4 123 lb4 100 I2 I2 I2 12 123 II 100 lb5 127 I3 II II II 107 IO 90 158 131 18 IO IO IO 81 14 97 143 90 I4 14 14 I4 180 13 89 I45 100 I4 I3 I3 I3 I76 I3 66 148 129 27 I3 13 I3 Ibl I2 40 I.44 152 60 I2 12 12 I35 North Northeast East Southeast South Southwest west Northwest Horizontal North Northeast East Southeast South Southwest west Northwest Horizontal North 1 Northeast East Southeast South .O 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 3 33 79 73 18 3 3 3 6 I 8 27 28 IO I I I 2 0 Exposure 1 01 1 Steel Sash, or No Sash X Ii.85 o r I.17 Bold 01 0 1 0 0 IO Noon I 2 14 14 14 17 21 17 14 14 250 14 14 I4 22 30 14 14 I4 246 I4 I4 14 39 63 39 14 I4 135 14 14 I4 b7 105 67 14 14 212 14 14 14 14 19 44 44 19 240 I4 I4 14 14 27 53 44 16 236 14 14 I4 I5 58 82 46 I4 225 14 14 I4 25 98 II3 48 14 202 14 14 14 14 15 73 98 55 217 14 I4 I4 14 20 83 99 46 214 13 13 13 13 47 II2 102 27 200 I3 I3 13 13 82 I41 LO3 15 179 8 II I2 13 I4 39 I8 I2 I3 I4 135 132 94 43 14 142 lb3 I59 136 92 57 92 I21 139 145 8 II IS 47 92 8 II I2 I3 14 8 II I2 13 14 49 100 143 171 179 b 9 II I2 I2 lb 9 II I2 I2 109 II6 83 35 I2 127 lb1 lb2 143 104 68 109 137 154 159 6 9 23 64 104 b 9 II I2 I2 6 9 II I2 I2 27 71 109 l3b 145 4 9 II I2 I2 10) 91 II) 121 12) 0 ( 92 105 80 32 I2 0 114 I57 l b 2 143 108 0 64 II3 142 I59 lb3 9 28 72 108 9 II I2 I2 13 13 I3 47 139 136 43 I3 171 I2 12 I2 64 I54 143 35 I2 I36 I2 121 I2 72 I59 143 32 Haze -15% Face Altitude (Max.) Values II 14 14 55 19 98 44 73 44 I5 19 I4 I4 I4 14 14 14 217 240 14 14 46 I6 99 44 83 53 20 27 14 14 14 I4 I4 14 214 23b I3 I4 27 14 102 46 II2 82 47 58 I3 15 13 I4 I3 14 200 225 I3 I4 15 I4 103 48 I41 II3 82 98 I3 25 13 14 13 14 179 202 +0.7% p e r 1 0 0 0 F t - Monthly Maximums 30” PM 3 14 14 14 14 14 90 143 97 180 13 I3 13 I3 I4 100 145 89 I76 I3 13 I3 I3 27 129 I48 66 lbi I2 I2 I2 12 60 152 144 40 I35 I2 II 12 II 12 II I5 II 121 92 159 lb3 94 132 12 I8 143 100 II 9 II 9 I I 9 23 9 137 109 lb2 lb1 83 II6 II 9 109 71 II 9 II) 9 II 9 28 9 142 I I3 lb2 I57 80 105 Values South Southeast East Northeast North Northwest west Southwest Horizontal South Southeast East Northeast North Northwest West Souihwest Horizontal South Southeast East Northeast North Northwest 8 8 8 8 57 142 135 39 49 6 b 6 b 68 127 109 lb 271 4 4 3 3 3 3 18 73 79 33 6 I I I I IO 28 27 8 2 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 South Southeast East Northeast North Northwest west Southwest Horizontal South Southeast East Northeast North Northwest West Southwest Horizontal South Southeast 19 0 I 0 Horizontal - Yearly LATITUDE Time of Year EXpOSUre 18 29 33 12 IO 5 I2 IO 5 12 IO 5 I2 IO 5 90 75 42 lb1 I56 108 130139105 I31 61 19 14 20 22 I2 9 4 12 9 4 I2 9 4 12 9 4 100 82 42 lb4 155 100 123 131 93 123 66 I5 II 8 6 II 8 2 II 8 2 II 8 2 13 8 2 127 98 37 lb5 147 6b 100 108 55 IO7 47 b IO 5 0 IO 5 0 IO 5 0 IO 5 0 I8 9 0 I31 98 0 158 124 0 90 74 0 81 25 0 Dewpoint Decrease From 67 F + 7% p e r IO F Boxed 6 4.5 SOUTH DEC 22 J A N 21 & N O V 21 FE6 2 0 & OCT 23 w e s t Southwest Horizontal South Southeast East Northeast North Northwest west Southwest Horizontal Dewpoint Increase From 67 F - 7 % p e r IO F maximums MAR 22 & SEPT 22 APR 20 & AUG 24 MAY 21 & JULY 23 1 South Lat. Dec. or Jan. + 7% ’ l’;\l<-l‘ I, 1.0 \I) I~:4l‘IxI.\‘I 1x(; l-48 TABLE 15-SOLAR HEAT GAIN THRU ORDINARY GLASS (Contd) Btu/(hr) (sq ft sash area) 40” N O R T H--~I LATITUDE 1 S U N AM Time of Year J U N E 21 8 / JULY 23 & M A Y 21 AUG 24 & APR 20 North Northeast East Southeast &uth Southwest ‘west Northwest Horizontal North 1 32 1 Southeast ,5outh Southwest /West Northwest Horizontal North Northeast East Southeast /South Southwest /West Northwest Horizontal I 1 1 IO II Noon I 20 1 51 1 b b 6 I 61 3l\( 24 1 I 1 1 ) 9 T I M E 54 I 51 5 5 5 24 7 68 84 48 3 3 3 3 9 PM 2 3 12 13 14 14 14 14 14 13 112 73 30 I4 14 14 I4 I3 95 44 14 I4 I4 13 162 1 I4 88 I 109 [II: 99 71 34 14 I4 131 IO 12 19 35 44 54 44 35 19 IO I2 I3 14 I4 34 71 99 III IO I 12 I I3 I 14 I I41 14 I 44 I 95 I142 I IO) 121 I31 I41 14j I41 301 82 I134 1179 210 232 23,7T232 210 179 14 1 I2 / I3 14 14 14 14 14 13 6 26 14 14 14 I4 1.7 43 14 14 I3 96 I II9 i-1245 1:: 82 1442 I5 I4 I3 101 131 261 441 b3/ 691 631 441 261 IO I2 I3 14 I5 42 82 110 I25 IO I2 13 I4 I4 14 43 98 144 IO I2 13 I4 I4 I4 14 26 66 73 I25 171 203 225 233 225 203 I71 8 II 13 I4 I4 I4 I4 I4 I3 102 82 46 I6 14 14 14 I4 13 147 I62 I45 IO1 45 IS 14 14 I3 I05 I38 I46 139 107 bb 25 I4 I3 8 24 51 89 97 102 97 89 51 8 II I3 I4 25 66 107 139 I46 8 II 13 14 14 I4 45 101 145 8 II 13 14 I4 I4 I4 lb 46 ’ 47 100 150 185 205 214 205 I85 150 4 14 0 ” S O U T H L A T I T U D E 5 6 EXpOSW 12 20 32 12 IO 6 12 IO 6 12 IO 6 12 1 IO 1 6 IO 9 1 88 151 I62 1 lbl i I26 73~11211331118 134 82 31 I2 14 24 I2 IO 5 I2 IO 5 12 IO 5 131 101 51 I19 96 54 I64 lb1 118 105 127 106 126 73 24 II 8 7 II 8 3 II 8 3 II 8 3 24 8 3 138 105 48 I62 147 84 82 102 68 100 47 9 Time o f Year South Southeast East Northeast North 1 1 1 1 Northwest west Southwest Horizontal South Southeast East Northeast North Northwest west Southwest Horizontal 1 DEC 2 2 J A N 21 & 1 NOV 21 . South Southeast East Northeast FE6 2 0 North & Nokthwest Wed Southwest Horizontal OCT 23 . North Northeast East SEPT 22 1 Southeast /‘South Southwest /west Northwest Horizontal & MAR 22 North Northeast East OCT 23 Southeast rSouth & Southwest FEE 2 0 1 I 1 , west Northwest Horizontal North Northeast East Southeast /South Southwest /west N O V 21 & JAN 21 Northwest Horizontal North Northeast East southeast /South Southwest DEC 22 /west Northwest Horizontal Solar Gain Correction 1 Steel Sash, or No Sash X li.85or I.17 1 I 1 O(Il6ll491139( 991 451 0 I57 133 0 II0 122 0 5 9 I2 I4 41’ 0 5 9 I2 I3 13 0 5 9 12 13 I3 0 21 67 124 I53 I76 I41 90 140 90 14 14 183 131 131 121 91 41 I4 12’ 9 122 I10 81 44 I33 157 162 144 45 99. 139 149 I3 13 26 58 I76 I53 124 67 0 2 IO II I2 I2 I2 II IO 6 0 35 3: I2 II 12 I2 12 II IO b 01 851 11711221 881 391 121 I21 I I 1 IO1 61 0 i 81 1 I32 I Ibl I I63 I 1441 107 I 6 3 I 20 1 101 61 01 211 5911041 1371 1541 I62 II541 1371 1041 591 0 2 6 IO 20 63 107 144 I63 lb #I / 1321 0 2 b IO II 12 I2 39 88 12 211171 0 2 b IO II I2 I2 12 II I2 33 0 8 29 64 101 123 129 123 I01 64 29 0 0 3 7 9 IO II IO 9 7 3 0 0 I2 7 9 IO II IO 9 7 3 0 0 91 100 74 33 II IO 9 7 3 0 0 109 144 I56 144 II6 70 27 7 3 0 0 59 104 I39 158 I66 I58 139 104 59 0 0 3 7 27 70 II6 144 I56 144 109 0 0 3 7 9 IO II 33 74 100 91 0 0 3 7 9 IO II IO 9 7 I2 0 0 I6 43 73 92 I03 92 73 43 I6 0 0 2 6 9 IO IO IO 9 b 2 0 0 7 6 9 IO IO IO 9 b 2 01 01 72 1 861 681 311 101 101 91 61 21 0 I 0 1 88 I 1341 I48 I 1421 I I5 1 7 3 1 3 0 t 7 1 2 1 01 01 51 1 991 I341 I581 I65 I I581 1341 991 51 0 0 2 7 30 73 II5 142 148 I34 88 0 0 2 6 9 IO IO 31. 68 86 72 0 0 2 b 9 IO IO IO 9 6 7 0 0 8 32 55 76 85 76 55 32 8 Haze -15% ( M a x . ) Altitude +0.7% per 1000 Ft Bold Face Values L Monthly Maximums Dewpoint 51 5 I2 95 II6 51 21 01 East 0 0 0 0 0 0 Northeast North Northwest 1 Decrease From 67 F +7%per IOF 01 0 0 0 0 & SEPT 22 Southwest Horizontal South Southeast East Northeast North Northwest West / APR 20 1 % A U G 2 4 M A Y 21 East Northeast North & west JULY 23 Northwest Southwest Horizontal South Southeast East 0 1 Northeast 0 0 0 E l west 2 0 2 0 2 1 0 1 2 I 0 I 21 I 0 1 81 1 0 1 851 01 35 8 0 0 0 0 I 0 0 : 0 0 0 0 0 0 0 0 0 i 0 0 I 0 I 0 1 MAR 22 North Northwest west I J U N E 21 Southwest Horizontal Dewpoint Increase From 67 F - 7 % p e r IO F Boxed Values - Yearly maximums South Lat, Dec. or Jan. + 7% TABLE 15-SOLAR HEAT GAIN THRU ORDINARY GLASS (Contd) 50° 50” 50° Btu/(hr) (sq ft sash area) NORTH LATITUDE Time of Year JUNE 21 Exposure 1 1 AM 6 North Northeast East Sod east South Southwest west Northwest Horizontal 29 126 I39 64 a 8 8 a 44 i JULY 23 & M A Y 21 -AUG 24 & APR 20 SEPT 22 & MAR 22 OCT 23 & FEB 20 I-JOV 21 & J A N 21 S U N 101 I I 14 lb 94 124 b8 14 I4 14 197 14 14 41 98 87 23 I4 I4 214 North Northeast East Southeast South Southwest west Northwest Horizontal 21 I I4 I31 65 I4 I5 961 136 6 80 North Northeast East Southeast South Southwest west Northwest Horizontal North Northeast East Southeast South Southwest west Northwest Horizontal a 76 T I M E 1Noon 14 14 431 109 98 14 14 I4 bl 93 61 I4 I4 220 14 14 I41 70 I06 98 1 451 I53 I 1321 4 5 I I4 14 14 14 68 124 94 lb 197 I3 13 13 I3 39 135 136 50 173 I2 I2 I2 I2 lb I26 lb2 94 133 12 IO IO IO IO 102 lb4 125 86 I4 I4 141 14 80 I3 I3 I31 I3 50 I2 I2 I21 12 21 6 Exposure 29 8 South Southeast East Northeast North Northwest West Southwest a 8 8 64 139 126 44 Ii IO 101 IO IO North Northeast East Southeast South Southwest west Northwest Horizontal C 0 C C C C C C C North Northeast c c South Southeast East Northeast North or J A N 21 & 4 1 East 4 1 FEB 20 Northeast OCT 23 ’ I 12 1 121 12 12 1 1 12 1 IO I 8 I 4 1 0 1 Southeast MAR 22 & SEPT 22 ] 01 East 0 I 0 0 Northwest west Southwest 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 South Southeast East Northeast North Northwest west Southwest Horizontal 0 0 l 0 0 0 l 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 144 1157 1145/ill I 35 1 79 II05 / 9 9 ) 6 6 57 127 II6 21 6 6 30 5 5 47 8 8 28 127 143 b7 8 8 47 6 6 23 99 25 5 5 I9 Sash, , DEC 22 Horizontal 21 6 61 6 6 14 1 I4 1 13 1 12 1 10 1 8 89 I 40 I 13 I 12 I IO I 8 107 I II61 N o Lash l/.85 o r 1.17 ime of Year & 0 0 “0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 X 14 I4 I41 26 98 3 NOV 21 94 53 4 4 4 4 I3 North Northeast East Southeast South Southwest west Northwest Horizontal Steel - I4 I4 I4 23 87 98 41 14 214 1 50” SOUTH LATITUDE PM 2 b 6 b 33 DEC 22 Solar Gain Correction I 1 I31 62 b b 33 9 9 9 107 153 107 9 9 53 7 7 7 8 8 8 67 143 127 28 8 47 6 6 6 ( M a x . ) 1 +0.7% II6 23 6 33 107 47 5 I9 41 27 3 5 I I I I 34 62 51 5 4 0 0 Dewpoint Decrease From 67 F per 1000 Ft Bold Face Values - Monthly Maximums 4 4 4 4 70 95 64 4 I3 3 3 3 100 I 62 I 25 I 3 I I41 I31 99 31 100 7 7 40 Altitude -15% 6 6 6 21 II6 127 57 6 30 5 5 5 691 73 29 Boxed + 7 % p e r IO F Values - Yearly , l & South Southeast East Northeast North Northwest AUG 24 MAY 21 & JULY 23 J U N E 21 west Southwest Horizontal Dewpoint Increase From 67 F - 7% per 10 F maximums APR 20 South Lat. Dec. or Jan. + 7% Heat C;aitl to Space = (.4 x 52 R) + .43 R = .G38 R, or .64 R REFLECTED I .4‘3 R TRANSMITTED .OBx.51:.77R / / ITT-t-.4 x.i5x .51 x.77R . ALL GLASS TYPES-WITH AND WITHOUT SHADING DEVICES Glass, ollw tltnn odi~nm-y solar heat because it 1. May be thicker, or glass, absorbs more 2. May be specially treated to absorb solar heat (heat absorbing glass). These special glass types reduce the transmitted solar heat but increase the amount of absorbed solar heat Nowing into the space, Normally they reflect slightly less than ordinary glass because part of the reHection takes place on the inside surface. /I portion of heat reflected from the inside surface is absorbed in passing back through the glass. The overall effect, however, is to r-educe the solar heat gain to the conditioned space as shown in Fig. 15. ‘Refer t o Ilerr~ S, @ge 51, for absorptivity, reflecivity and transmissibility of common types of glass at 30” angle of incidence.) The solar heat gain factor through 52% heat absorbing glass as compared to ordinary glass is .64R/.88R = ,728 or .73. This multiplier (.73) is used with Table 15 to determine the solar heat gain thru 52(,v0 heat absorbing glass. Multipliers for various types of glass are listed in Table 16. The effectiveness of a &ndi~g device depends on its ability to keep solar heat from the conditioned space. .\I1 shading devices reflect and absorb a ma,jor portion of the solar gain, leaving a small portion to be tr-ansmitted. The outdoor shading devices are much more effective than the inside devices because all of the reflected solar heat is kept’ out and the absor!~etl heat is dissipated to the outdoor air. Insirle devices necessarily dissipate their absorbed heat within the conditioned space and . Heat Gain to Space = (.40 X .15 R) + (.37 X .77 R) + (.I!2 X .77 R) + (.08 x 31 X .77 R) + (.40 X .I5 X .51 X .77 R) = ,492 R or .49 R FIG. 16 - REACTION ON SOLAR HEAT (R) P L A T E GL A S S , W H I T E V E N E T I A N OF B L I N D, , I/~-INCH 30” IN C I D E N C E 1 AN G L E must also reflect the solar heat back through the glass (Fig. 16) wherein some of it is absorbed. (Refer to Item S, @ge 51, for absorptivity, reflectivity and transmissibility of common shading devices at 30” angle of incidence.) The solar heat gain thru glass with an inside shading device may be expressed as follows: Q = [.4ag + tg (%d f kd f r$sd f .$$,d)] ,$ where: Q = solar heat gain to space, Btu/(hr)(sq ft) R = total solar intensity, Bttt/(hr)(sq ft), (from T&k a = solar absorptivity t = solar transmissibility r = solar reflectivity g = glass sd = shading device 38 = conversion factor from Fig. I2 IS) For drapes the above formula changes as follows, caused by the hot air space bettveen glass and drapes: Q = [.?hp + tg (.85&l + tsd + rgrSd + .%,r,d)] -& The transmission factor U for glass with 100yO drape is 0.80 Rtu/(hr) (sq Et) (F). The solar heat gain factor thru the combination in Fig. fh as compared to ordinary glass is .49R/.88R =.557 or .56. (Refer to Table 16 Eor r/-inch regular plate glass with a white venetian blind.) NOTE: /\ctually the reaction on the solar heat reflcctctl Ijatk throltgh the glass from the hlintl is not always itlcntical to the lirst pass as assumed in this example. The first pass through the glass filters ant most of solar radiation that is to he al)sorhetl in the glass, and the second pass al)sorbs somewhat less. For simplicity. the react ion is assumed identical, since the quantities are normally small on the second pass. (i. Outdoor canvas awnings ventilated at sides and top. (See T/lb/e 16 I’ootnote.) 7. Since Tfr/Ile Ii is basctl on the net solar heat gain thru ordinary glass, all calculated solar heat factors are clividetl by .88 (Fig. 12). 8. ‘1‘11~ Basis of Table 16 Over-all Factors for Solar Heot Gain thru Glass, With and Without Shading Devices Use of Table 16 -Over-all Factors for Solar Heat Gain thru Glass, With and Without Shading Devices The factors in TnOle 16 are multiplied by the values in Table I5 to determine the solar heat gain The Eactors in Tcrblc 16 ~lre I,aserl on: !. ,411 o u t d o o r film coefficient o f 2 . 8 Btu/(hr) (sq ft) (deg F) at 5 mph wind velocity. 2. An inside film coefficient of 1.8 Btu/(hr)(sq ft) (deg F), 100-200 fpm. This is not 1.47 as normal!y used, since the present practice in well designed systems is to sweep the window with a stream of air. 3. A 30” angle of inciclencc which is the angle at which most exposures peak. The 30” angle oE incidence is approximately the balance point on reduction of solar heat coming through the atmosphere and the decreased transmissibility of glass. Above the 30” angle the transmissibility of glass decreases, and below the 30” angle the atmosphere absorbs or reflects more. 4. All shading devices fully drawn, except roller shades. Experience indicates that roller shades are seldom fully drawn, so the factors have been slightly increased. 5. Venetian blind slats horizontal at 45” and shading screen slats horizontal at 17”. TYPES OF GLASS OR SHADING DEVICES* Ordinary Glass Regular-Plate, IA” Glass, Heat Absorbing Venetian Blind, Light Color Medium Color Dark Color Fiberglass Cloth, Off White (5.72 - 61/58) Cotton Cloth, Beige (6.18 - 91/36) Fiberglass Cloth, Light Gray Fiberglass Cloth, Tan (7.55 - 57/29) Glass Cloth, White, Golden Stripes i Fiberglass Cloth, Dark Gray Dacron Cloth, White (I.8 - 86/81) Cotton Cloth, Dark Green, Vinyl Coated ‘-(similar to roller shade1 &ton Cloth, Dark Greek (6.06 - 91/36) I average absorptivity, reflectivity and trans- missability for common glass and shading devices at a 30” a. ngle of incidence along with shading factors appear in the table below. thru diferent combinations of glass and shading devices. The correction factors listed under Table 15 are to be used if applicable. Transmission due to temperature diffcrcnce between the inside and outdoor air must be added to the solar heat gain to determine total gain thru glass. Example 3 - blind on inside, Find: Peak solar heat gain. Solution: By inspection of Table 15, the boxed boldface values for peak solar heat gain, occurring at 4:00 p.m. on July 23 I = 164 Btu/(hr)(sq f:) Reflectivity (4 .nx is ii .51 .39 .27 .60 .51 .47 Transmissibility (9 .86 .77 (1 - .05 - a) .12 .03 .Ol .35 .23 .23 .14 .54 .42 .41 29 .0!2 .28 Jvi .-- texture: figures in parentheses are ounces per sq yd, and yarn count warp/filling. Consult manufacturers for actual values. / Shades Given: West exposure, 40” North latitude Thermopane window with white Venetian :3/, drawn. .60 rne actual drapery material may be different in color and Drawn Occasionally it is necessary to estimate the cooling load in a building where the blinds are not to be fully drawn. The procedure is illustrated in the following example: .-\bsorptivity (4 36 .iia by mfg. .37 .58 .72 .05 .2G .30 .44 .05 .02 -‘factors tor various draperies are given for guidance only since . . Partially -::: 14 .-.28 .oo .70 tcompared to ordinary glass. $For a shading device in combination Solar Facto* 1 .oo .94 - .56; .651 .75i .482 .56i .59+, .64f .65,+ .75f .76$ .88$ .76f with ordinary glass. _ 1-53 I’,\liT I. I.O:\II ES-l‘lhl.\TING Thcrnwlx~rw wintlows have n o sasll; t h e r e f o r e , s a s h arca correcli~ln = 1 /.%I (hottom Tn/r/e 15) I n t h i s vcncti;tn exnmplc, 7/, o f the wintlow i s ~ovcrctl with tllc Ijlintl anti I/ is n o t ; Illcrcforc. F a c t o r factor. eq”“ls ‘V, o f t h e ‘iol3r hc;lt pin (IlC 0vc1.311 fxtor -t- I/, of the gl;1ss F a c t o r f o r :+‘, tlrawn = (‘fi X 3)+ (IA X .80)(Tn/~k 1 6 ) = 59 59 = 164 X _Hr, Solar heat gain = II4 l)LU/ (hr) (“‘I ft) . Example 4 - Pea& Solar Heat Gain thru Solex “R” Glass Given: Weste x p o s u r e , ,40” N o r t h laritutle IA” S&x “R” glass in steel sash, double hung wintlow TABLE 16-OVER-ALL FACTORS FOR SOLAR HEAT GAIN THRU GLASS WITH AND WITHOUT SHADING DEVICES* Apply Factors to Table 15 Outdoor wind velocity, TYPE OF GLASS ORDINARY GLASS ’ REGULAR PLATE (i/4 inch] HEAT ABSQRBING, GLASStt 40 to 48%, Absorbing f ,-i 48.to 56% ~At&binq ., 56 to 76‘)” Absorbing DOUBLE PANE . .! Ordinary Glass Regular Plate 48 to 56% Absorbing outside; Ordinary Glass inside. 1-48 to 560/d Absorbing outs&: R.eguler Plate inside. 5 I mph Angle SEE _’ I .oo of incidence, - Shading 30” INSIDE ’ VENETIAN BLIND* CLASS FACTOR ..- I Medium Light _ _ _ Dark C o l o r Cnlor. --.-. I Calor .b5 1 8 .I5 .94 .56 .b5 .74 .80 .73 .56 .53 .b2 .72 .I2 .59 .b2 .b2 .51 .54 .5b .I I .I0 .90 .80 .54 .52 .bl .b7 .I4 .59 .b5 .I2 .52 ) .36 .50 .83 .b9 PAINTED GLASS .-Light color ,’ ‘M&dium Co&r Dark Color . .28 .39 .50 STAINED GLASS+* .70 36 .bO .32 .4b .43 appear on next page. A8 .47 ) .39 ) .39 .43 .%I .52 .b4 .57 f .I0 .I2 .I0 vent. sides & top Light Color Color -,I 1 __-- .I3 .22 .I5 .20 .25 .I2 .2 I .I4 .I9 .24 .I I .I0 .I2 .I I .I6 .I5 .I0 .I8 .I6 .I4 .I0 .I2 .20 .I8 .I6 .I2 .I I .20 .I8 .I4 .12 .I8 .Ib .22 .20 .I0 .I3 .I0 I window OUTSIDE 1 A W N I N G S 17O horiz. slats Inside .43’) I covering Light on Outside Medium’‘* Dark 8‘ Dark on Color Color Liaht I Color .75 I fully OUTSIDE SHADING SCREENt 45’ horir. slats or ROLLER ‘SHADE .:_ ,56 : ; devices OUTSIDE VENETIAN BLIND 45’ horiz. or vertical .I4 -TRIPLE P A N E Ordinary G&s Regulbr Plate Footnotes for Table 16 - 1 .I0 .I! .I0 .I !O .I I .I0 .I0 .I2 .I I .I0 .I8 .I5 .I2 .I0 .I6 .I4 .20 .I7 CM,\l’TI:I< 4. SOL,\li f-I1:,\7‘ C,,\IN 1-53 ‘1‘111~~1 C;I>;\SS Pig. 15 md Ih, (2) by applying the absorptivity, rellectivity and transmissibility of glass and shades listed in the table on Page 51, or determined Crom manul’acturcr, and (3) by distributing heat absorbed within the dead air space and glass panes (rig. 17). Example 5 - Approximation of Over-d Factor Given: ,A cornl)inatinn as in Fig. fh backed on the inside with another pane of IA-inch regular plate glass. Find: The G. I’i - REACTION ON S OLAR HEAT (R) , ~-INCH P LATE GLASS, WHITE V ENETIAN BLIND , ~/-INCH P LATE GLASS, 30” A NGLE OF INCIDENCE APPROXIMATION OF FACTORS FOR COMBINATIONS NOT FOUND IN TABLE 16 Occasionally combinations of shading types of glass may be encountered that ered in Table 16. These factors can be (1) by using the solar heat gain flow devices and are not covapproximated diagrams in over-all factor. Solution: FiczLr’e 17 shows the distribution of solar heat. The heat al). sorbed I,etween the glass panes (dead air space) is divided 45% and 557” respectively between the in and out flow. The heat ahsorbed within the glass panes is divided 207” in and 80%) out for the outer pane. and 7.57” in and 25% out for the inner pane. These divisions are based on reasoning partially stated in the notes under 13, which assume the outdoor film coefficient of 2.8 Btu/ (hr) (sq ft) (deg F), the indoor tilm coefficient of 1.8 Rtu/ (hr) (sq ft) (deg F). and the over-all thermal conductance of the air space of 1.37 Btu/ (hr) (sq ft) (deg F). Heat gain to space (Fig. 17) = (.75 X .I5 X .I2 X .77R) + (.77 X .I2 X .77R) + .45 [(.37 X .77R) + (.08 X .5l X .77R) + (.08 X .I2 X .77R)] + .20 [(.l5R) + (.I5 x .5l X .77R)] = .2G84R or .27R gain factor as compared to ordinary glass = .27R/.88R = .3l Solar heat Equotionr: Solar Gain Wifhouf Shades = (Solar Data from Table 15) X (Glass Factor from table) Solar Gain With Shader = (Solar Data from Table 15) X (Over-all Factor from table) Solar Gain Wifh Shades Partially Drown = (Solar Data from Table 15) X [(Fraction Drawn X Over-all Factor) + (1 - Fraction Drawn) X (Glass Factor)] Footnotes for Table 16: ding devices fully drawn except roller shades. For fully drawn roller shades, multiply light colors by .73, medium colors by .95, and dark colors by 1.08. tFactors for solar altitude angles of 40’ or greater. At solar altitudes below 40”, some direct solar rays pars thru the slats. Use following multipliewMULTIPLIERS FOR SOLAR ALTITUDES BELOW **Commerdol per inch. ttMost heat absorbing glass used in comfort air conditioning is in the 40% to 56yo range; industrial applications normally use 56yo to 7Oyo. The following table presents the absorption qualities of the most common glass types:SOLAR 40’ Trade Name or Description 6:00 a.m. 6:00 p.r”. 6~45 5:15 a.m. p.m. 7:30 4:30 a.m. p.m. 5:45 a.m. 6:15 p.m. 6:40 5:20 a.m. p.m. 7:30 4:30 a.m. p.m. 5:30 a.m. 6:30 p.m. 6:30 5:30 a.m. p.m. 7:30 4:30 a.m. p.m. 10 2.09 3.46 20 1.59 2.66 30 1.09 1.67 Aklo Aklo Coolik Coolite C.O.F. Solex R $With outside convos awnings tight against building on sides and top, multiply over-all factor by I .4. $Commerciol shade bronze. Metal slots 0.05 inches wide, I7 per inch. shade, oiuminum. Metal slats 0.057 inches wide, 17.5 $fWith RADIATION ABSORBED Manufacturer Blue Ridge Glass Corp. Mississippi Glass Co. Mississippi Glass Co. Libbey-Owens-Ford Pittsburgh Plate Glass BY HEAT 1 ABSORBING GLASS (%) -IS&r Color Pole Pale Light Light Pale Blue-Green Blue-Green Blue Blue Blue-Green Pole Green multicolor windows, ore the predominant color. Radiation Absorbed 56.6 69.7 58.4 70.4 48.2 50.9 Tnhlc 17 have been incrcxsetl to include the I /.85 multiplier in Tfll~le 15. tars in GLASS BLOCK Use of Table 17 - Solar Heat Gain Factors for Glass Block, With and Without Shading Devices The lactors in Tuhle I$ are used to dctcrminc the solar heat gain thru all types ol’ glass block. The transmission ol heat caused by a dilfercncc between the inside and outdoor temperatures must also be figured, using the appropriate “U” value, Chpte~ 5. Shading devices on the outdoor side of glass block are almost as effective as with any other kind OC g 1 ass since they keep the heat awry l’rom the g l a s s . Shading devices on the inside are not effective in reducing the heat gain because most of the heat reHccted is absorbed in the glass block. Example 6 - Peak Solar Heat Gain, Glass Block Given: Find: I’eak solar heat gain . Solution: By inspection of Basis of Table 17 - Solar Heat Gain Factors for Glass Block, With and Without Shading Devices Table 15, the peak solar heat gain occurs on July 23. Solar heat gain At 4:00 p.m. = (.39 X 164) + (.21 X 43) = 53 At 5:OO p.m. = (.39X 161) -I- (.21 X 98) = 84 At 6:00 p.m. = (.39 X 118) + (.21 X 1 4 4 ) = 76 The factors in Table I7 are the average of tests conducted by the ASHAE on several types of glass block. Peak solar heat gain occurs at 5:00 p.m. on July 23. Since glass block windows have no sash, the Eac- . TABLE II-SOLAR HEAT GAIN FACTORS FOR GLASS BLOCK WITH AND WITHOUT SHADING DEVICES* Apply Factors to Table 15 I MULTIPLYING FACTORS FOR--..-GLASS "LVIR n’npy : .,,., )., . ;:" I EXPOSURE IN NORTH LATITUDES .' ‘9 .A, I I .--, I ..““,a 4.” 2” .39 .35 74 .21 22 .27 .39 .24 .22 3.0 3.0 .-)_I?C 33 .-- ,n 4.” 97 Northeast East Southeast .- 3.0 3.0 e”“I..FYII i -.; C^..*Ls^r. I East Northeast . : .,. <: “,. :2,.... 2, North South Summart Wintert Southwest West Northwest “Factors~include ,- L A T I T U D E S ‘:’ ‘.i ..-,. . .‘.Z-, . . . ,, , 2j ,z .39 .27 I I correction for no rash with gloss block windows. Equations Solar heat gain without shading devices = (Bi X Ii) + (6. X Ia) Solar heat gain with outdoor shading devices = (Bi X Ii + Ba X 1,) X .25 Solar heat gain with inside shading devices = (Bi X Ii + Ba X Ial X .90 . .21 .24 I I 3.0 3.0 - -. . . . . . -. , Wintert I U--*L..,^-‘ . ..arl.ln~r, West .. Southwest ._ -!., i( . . ; tllse the wmmer factors for. all latitudes, North or South. Use the winter factor for intermediate seasons, 30’ to 50’ North 01 South latitude. Where: Bi = Instantaneous transmission factor from Table 17. B (I = Absorption transmission factor from Table 17. = Solar heat gain value from Table 15 for the desired time and Ii wall facing. 42 = Solar heat gain value from Table 15 for 3 hours earlier than Ii and same wall facing. (:FIi\l’TI:l< 4. SOI.,\I< 111-,\‘I’ G.\IN ‘1‘111<11 l-5.5 (;L.:\SS SHADING FROM REVEALS, OVERHANGS, FINS AND ADJACENT BUILDINGS i\II ~intl0~s arc sliatlctl t o a greater o r Icrixr tlcgrcc by tile projections close to it and by buildi n g s worii~tl i t . Tliis sll;itling retluccs the s o l a r heat gain tltror~gh tllcse w i n d o w s by keeping t h e direct rays oC the sun off part or all of the window. T h e sllarletl p o r t i o n h a s o n l y t h e tliffusc c o m ponent striking it. Shading ol’ windows is signilic a n t i n nionuincntal type b u i l d i n g s w h e r e t h e reveal may be large, even at the time of peak solar heat gain. Clrtrrl I, this chapter, i s p r e s e n t e d t o simplify the dotermination of the shading of windows by these pro jcctions. Basis of Chart 1 - Shading from Reveals, Adjacent Overhangs, Fins and e location of the sun is defined by the solar azimuth angle and the solar altitude angle as shown in Fig. IS. The solar azimuth angle is the angle in a horizontal plant between North and the vertical plane passing through the sun and the point on ’ earth. The solar altitude angle is the angle in a vertical plane between the sun and a horizontal plane through a point on earth. The location of the sun with respect to the particular wall facing is defined by the wall solar azimuth angle and the solar altitude angle. The wall solar azimuth angle is the angle in the horizontal plane between the perpendicular to the wall and the vertical plane passing through the sun and the point on earth. The shading of a window by a vertical projection alongside the window (see Fig. 13) is the tangent of the wall solar azimuth angle (B), times depth of the projection. The shading of a window by a horiZC 31 projection above the window is the tangent Or qle (X), a resultant of the combined effects of the altitude angle (A) and the wall solar azimuth angle (Is), times the depth of the projection. Tan X = Tan A, solar altitude angle Cos B, wall solar azimuth angle T h e u p p e r p a r t o f Clln1.t 1 determines the tangent of the wall solar azimuth angle and the bottom part determines tan X. Use of Chart 1 - Shading from Adjacent FIG. 18 - SOl.hll Buildings Reveals, Buildings Overhangs, Fins and The procedure to determine the top and side shading from Chnrt I is. 1. Determine the solar azimuth and altitude angles l’roin Table 18. ‘\NGI.I<S I FIG. 19 -S HADING BY 2. Locate the solar aLiniuth upper part of Cl~clr.t /. WA L L PR O J E C T I O N S angle on the scale in 3. I’roceetl hori/.on tally to the exposure clesirctl. -I. Droll vertically to “Shading from Side” scale. 5 . ~Iultiply t h e d e p t h 0C the projection (plan vielv) by the “Shading from Side.” 6. Locate the solar altitude angle iowcr part ol C/,nrt I. on the scale in nlltil the “Shatli~lg tronl 7 . ;\love Iiori~ontally Side” value (45 clcg. lines) tlctcrniinctl in Step 4 is intersected. x. DI-01) vertically t o “ S h a d i n g I’rotn 7’01”’ Cronl intersection. hlultil>ly tile d e p t h 0C tllc lxojection (cleva!I. tioil view) by the “Slla(li~lg l’rom TO]).” SUN’S RAYS PLAN FIG. 21 - .%ADINC O F bYEAL ANI) O V E R H A N G SUN’S RAYS . \\\ ’ \L \ \ ‘, \\ Length of I)uiltling in shack, L = 8.5 - 15 - (.l x 75) = 62.5 It Height of l)uildinS in shatle, H = 100 - (75 X .i) = -17.5 ft The air contlitionctl lxtiltling is -shatlctl tfl a height of 47.5 ft and 62.5 Lt along the face at 4:00 pm. on July 23. Example 8 - Shading of Window by Reveals Given: ^, L A steel casement wintlow reveal. on the west sicie with an S-inch 1 Find: Shading by the reveal at 2 p.m. on July 23, 40” North Latitude. Solution: From Table 18, ELEVATION FIG. 20 - SHADING OF IjUlLDlNG RY I-\D,JACENT solar azimuth-angle = 242” solar altitucte angle = 57” From Chart I, shading from side reveal = .G X 8 = 4.8 in. shading from top reveal = 1.8 X S = 14.4 in. RlJIL.DING Example 9 - ‘example ’ 7 - Shading of Building by Adjacent Building Given: 13uiltlings located as shown in Fig. 20. Fintl: Sharling at 4 p.m., July 23, bof I)uilding to 1)~ air conditioned. Solution: It is recommentletl fhat the hkriltling plans ancl elevations be sketcheci to scale with approximate location of the sun, Lo enable the engineer to visualize the shntling contlitions. From Ta6le 19, solar azimuth angle = 265” solar altitutle angle = 45” From C/ICZV! I, shading from sicte = ,I ft/ft a shacting from top = .7 It / ft , Shading of Window by Overhang and Reveal Given: The same window as in Esarnple 6 inches above the window. 9 with a 2 ft overhang Finch: Shading by reveal and overhang at 2 p.m. on July North Latitude. Solution: Refer to Fig. 21. Shading from sitle reveal (same as Exaurple 23, 40” 8) = 4.8 in. Shatling from overhang = 1.8 X’ (24 i- 8) = 57.6 in. Since the overhang is 6 inches above the wintlow, of window shaded = 57.6 - 6.0 = 5 1.6 in. the portion CM,-\I”l‘l<R ,I, 401..\K III,:\ I ’ (i\IS I’IIl<li l-57 (i1.\5S CHART 1 - SHADING FROM REVEALS, OVERHANGS, FINS AND ADJACENT BUILDINGS Given: Fintl: Shading I)y rcvenl and overhang nt 2 pm, July 23, 40” North Latitutlc. Solution: From Table 18, .\zimuth nnglc = 212” ’ Altitude nnglc = 57” From Clrnrt I, 1. Enter at solar nrimuth angle (242”) to west (it’) exposure shading from side = 0.G inch/ inch. 2. Enter at solar altitutlc angle (.57”) to shxling from sick (0.G inch/inch).. Shading from top = 1.8 inch/inch. 3. Shading by reveal = 0.G X 8 = 4.8 in. 4. Shading I~yoverhnng=1.8(24+8)-6=51.6 in. -I DE (INCH/II NCI \ \ \ t \ \ \, ‘\ 30 \ 35 40 45 50 55 60 65 + 70 75 8 0 ,I .I5 .2 .3 .4 3 .6 .7 .8 I:5 2 3 “SHADING’ FROM TOP ( I N C H / I N C H ) 4 5 678 IO 15 20 l-58 ,I’ART 1 . LO:\D ESTIbl.\‘I‘iNG TABLE 18-SOLAR ALTITUDE AND AZIMUTH ANGLES NORTH* LATITUDE LAT 0 0. SUN TIME Jai n . 2 1 cir AZ - Fel b. Alt i 6AM 7 a 9 IO II 12 N I PM 2 3 4 5 6 14 28 42 54 b5 70 as 54 42 28 14 15 30 44 58 71 79 71 58 44 33 I5 bAM 7 : IO II I2 N I PM 2 3 4 5 6 LA1 IO” bAM 7 a 9 IO II I2 N I PM 2 LAT 20” : 5 6 6Alu 7 a 9 IO II l2N I PM LAT 30” : 4 5 6 LAT 40’ bAt. 7 8 9 IO II I2 N I PM 2 3 4 5 b LAT 50” 6AM 7 a 9 IO II 12 N I PM 2 3 4 5 6 SOUTH* LATITUDE ‘Use months SUN TIME indicated II 13 I7 26 44 80 ‘lb !34 !43 ,47 249 IO 24 37 48 57 60 57 48 37 24 IO 113 II7 124 I36 I55 I80 205 224 236 243 247 12 27 41 54 64 69 64 54 41 27 I2 7 103 I5 108 30 I I5 44 125 59 1 4 4 72 180 80 Tii 72 59 235 245 44 252 30 I5 257 - -I 114 121 I30 142 I58 I80 202 218 230 239 246 IO 23 36 47 55 59 5: 47 3t 23 IC, 106 I I2 121 133 I52 I80 2oE 227 235 246 254 95 1: IO1 42 I08 55 I20 143 bC JC I I80 c> 217 5 5 240 4i 252 2EI 259 I4 265 - - 2 I4 24 3; 3E 4( It 3; 2’ II i I< 3C 4C 4t 4( 4t 4( 3C I< ; 125 I36 I49 lb4 I80 196 211 224 235 j I I( I tI II5 22 $ I3 3 :2 l4! 3. 7 lb: 3’? l8( 3;7 rsr 3 :2 21! 2. 4 22’ I!5 !J 2, f 1; 2. 2f 3( -5 2z I. I - - 3 IO I5 :; 19 I5 IO 3 I25 138 I51 lb5 I80 195 209 222 235 IO 17 23 27 29 27 23 17 I C, 121 134 I48 lb4 1.20 196 212 226 239 IO I9 27 34 39 4c I % 34 27 I9 IC I - luly 2 3 pug. 2 4 at top for North T-99 I IO 122 138 157 I80 203 222 238 250 261 - - - 90 92 95 99 IO6 122 180 238 254 261 265 268 270 - IO1 I: I 97 I It 2t I06 12; 3c ; I lb 141 4c I 130 15s 5; 7 ISI 1% b( 1 I80 201 T 7 209 215 4 ’I 230 23: 3t I 244 2 4 ‘ 2( 254 25: I: ; 263 115 124 134 146 lb2 I80 I98 214 226 236 245 Ma iF 78 77 74 68 53 0 iii !92 286 283 282 4 ‘8 I2 ;4 a5 ‘0 & j4 &2 18 14 54 36 0 124 106 297 293 29 I 7 81 83 84 84 84 0 7% 276 276 277 279 282 3 I7 32 46 50 73 30 73 60 46 32 I7 3 JO 72 72 72 67 53 0 307 293 288 288 288 290 4 I8 32 46 59 72 81 72 59 46 32 I8 4 19 84 89 94 IO2 117 l8C Tz 256 261 271 2Jt 281 7 20 34 48 b2 75 90 75 62 48 a I? 31 44 56 67 71 I 56 44 31 I9 b -E 8; 7 I9 30 41 51 58 .L! 58 51 41 30 I9 7 - a 15 90 I.5 30 89 30 45 8 9 44 b0 89 5 8 88 71 75 90 L 7 ,9 75 t-i ‘I b0 ;8 45 (4 30 IO< T5 2 7 0 I5 - 6 I? 30 4c 47 50 4i 4c 3c I9 6 02 03 06 II2 I27 80 GT !48 !54 157 258 - 1 lb 31 46 51 75 39 75 bi 46 31 lb 2 - Latitudes: . I 2 4 I5 102 30 I03 44 IO6 58 112 71 127 7’9 I80 71 233 58 248 44 254 30 257 15 258 4 III ‘8 II3 I2 117 ;4 126 ,5 144 ‘0 180 ,5 216 54 2 3 4 +2 243 28 247 I4 2 4 9 14 II4 27 117 41 122 53 131 62 I48 67 180 62 212 53 229 41 238 27 243 I4 2 4 6 bAM 7 IB 9 I( 0 II I: ZN I PM 2 3 4 5 6 9 II6 23 121 35 128 46 139 53 156 57 180 53 204 46 221 35 232 23 239 9 244 6AM 7 a 9 I0 I I I 2N I PM 2 3 4 5 6 -7 lb il 46 jl 75 39 ii ii 16 31 lb 2 78 81 83 84 84 84 0 276 276 276 277 279 282 I I5 IO $4 j? 72 30 72 j’? 14 30 15 I 90 92 95 99 106 122 I80 238 254 261 265 268 270 I2 103 27 108 41 II5 54 I25 64 144 69 I80 64 216 54 235 41 245 27 252 12 257 IO 24 37 48 57 10 57 48 37 24 IO I13 117 124 I36 155 180 205 2 2 4 236 243 2 4 7 -7 75 79 82 85 88 0 272 275 278 281 285 289 8 68 21 72 35 75 48 77 62 77 76 74L 87 CI 76 28t b2 28:; 4 8 2 8 :I 3 5 2%> 21 2% I 8 2% , T I8 32 4b 59 72 81 i 55 4c 32 IE 4 79 84 89 94 102 I I7 I8C 243 258 2bt 271 276 281 14 28 42 55 66 JO 66 55 42 28 I4 95 IO1 108 12c 143 I80 217 24C 252 259 265 IO 23 36 47 55 59 55 47 36 23 IO 6 19 30 40 47 50 47 40 30 19 6 114 121 130 142 158 I80 202 218 230 239 246 To -77 79 23 35 86 48 93 61 103 73 122 it!? 180 7 3 218 bl 257 48 267 35 274 23 281 IO 288 II 6’I 24 71 37 8: : 49 8l i 62 9t j 75 II: 2 83 181I 75 241 3 62 26‘4 4 9 2 7 :2 37 2713 2 4 282 4 II 29 I 2 14 24 32 38 40 38 32 24 I4 2 II5 124 II 134 21 146 29 162 35 180 1 37 198 35 214 29 226 21 236 II 245 t 7 71 20 75 34 79 48 82 62 85 75 88 90 c 75 272 62 275 48 2 7 E 34 281 2 0 285 7 289 13 74 83 24 35 93 47 104 57 II8 66 143 J&c ?c bb 217 5 7 242 47 256 3s 267 24 277 13 286 - 83 94 106 120 137 I57 l8C 203 223 24C 254 2bt 277 - I5 25 34 44 52 58 60 58 52 44 34 25 15 - I8 74 27 85 37 97 46 IIC, 55 I28 61 151 63E I bl 209 55 232 4 6 2 5 CI 37 263I 2 7 276, I8 28C use Dec. 22 Alt AZ 4 67 3 70 I8 68 I7 72 12 72 32 68 45 67 I6 72 58 61 50 67 73 53 JO 44 77 0 I 30 0 J O 316 73 307 58 299 ( 50 293 45 293 16 2 8 8 32 292 32 288 I8 2 9 2 I7 288 4 293 3 290 I 9 I 0: I I: 12’ 15 I 81 55 23 24’ 251 265 279 - 9: 90 Nov. 2 I I4lt A Z ;-.-i-i >O 89 ‘5 88 ‘0 0 ‘5 272 ,O 2 7 1 t5 2 7 1 IO 2 7 1 I5 270 t 1 t c 77 88 100 I I4 I31 152 180 208 229 246 260 272 283 - Jov. 21 months at Dec. 22 bottom for 5 IO6 112 121 I33 152 I8C 20E 227 235 24E 254 IO 72 23 75 35 81 48 93 bl IO: 73 12; 80 I8( 73 238 bl 25; 48 26; 35 27’ 23 281 IO 285 I5 25 34 44 52 58 60 58 52 44 34 25 15 77 88 I00 II4 I31 I52 I80 208 229 246 260 272 283 . Ian. 21 South 9 18 28 37 44 49 51 49 44 37 28 I8 9 83 94 IO6 120 137 157 I8C 203 223 240 254 266 277 :ab. 20 Latitudes. 12 23 33 42 48 50 48 42 33 23 I2 99 II0 122 138 157 I80 203 222 238 250 261 IO I9 27 34 39 40 39 34 27 I? IO IO1 II4 127 143 lb0 I80 200 217 233 246 259 dar. 22 T-- Oct. 23 Alt AZ 78 77 74 68 53 0 307 292 286 283 282 7 -z IF 8i 31 9E 4 4 104 51 I Ii 6; I4( 71 I 8( s 22( 24: 5t 4’ 2% 31 2bt 27: I( 28( -i t 7 81 I5 7: 2 I3 7’ 26 8( 3 24 8: I? ?I 10; 37 8’ ? 35 9 : 30 41 I I: 49 I01 3 47 IO‘ 60 II.4 5 7 v/t 51 12’ 66 14: 58 15 69 1319 l8( blI8( 73 I81 3 _ 70 58 20’ 69 22:2 51. 23l 60 24, b if 2 4: 49 261 3 47 251 41 24 37 27 I 35 26: 30 251 7 2 6 28L 2 4 27; I9 26’. I5 2 8 8 I3 286 7 279 >ct. 23 and 4r iii - i 5 ,O .4 ,8 .I ‘9 T i8 14 IO i5 2 7 IO‘ Iii I4( 18( zi 24: 251 26! 27: 28( 9 I8 28 37 44 49 51 49 44 37 28 I8 9 - IO1 II4 127 143 I60 I80 200 217 233 246 259 ;ept. 22 June 2 I 20 F - 5 I5 24 32 37 39 37 32 24 I5 5 IO I7 23 27 29 27 23 I7 IO III I II! I 13 I l4! j lb: 2 18( JI91 3 2l! 5 22’ I 24 I 25L7 I21 I,34 148 lb4 I8C , I96 212 226 2,39 Apr. 20 8 17 24 28 30 28 24 I7 8 3 IO 15 19 20 I? 15 IO 3 125 136 I49 lb4 I80 I96 211 224 235 I25 I38 I51 lb5 I80 I95 209 222 235 lJuna 21 May 5 17 28 38 44 47 44 38 28 17 5 5 I4 21 25 27 25 21 14 5 SUN TIME bAM 7 117 124 133 I45 163 I80 197 215 227 236 243 . : I0 I I I2 N I PM 2 3 126 136 149 lb4 I80 196 211 224 234 6AM 7 a 9 IO II I2 N I PM 2 3 4 5 6 127 138 I51 lb5 I80 195 209 222 233 : 6 bAk.4 7 a 1’0 II 12 N I PM 2 3 4 5 b 6AM 7 a 6 12 15 17 I5 I2 6 I39 152 lb6 I80 194 208 221 21 9 0 I 2N I PM 2 3 4 5 6 SUN TIME . l-59 CHAPTER 5. HEAT AND WATER VAPOR FLOW THRU STRUCTURES This chapter presents the methods and data for determining the sensible and latent heat gain or loss thru the outdoor structures of a building or thru a structure surrounding a space within the building. It also presents data for determining and preventing water vapor condensation on the enclosure surfaces or within the structure materials. Heat flows from one point to another whenever a t >eraturc clifference exists between the two points; the direction of flow is always towards the lower temperature. Water vapor also flows from one point to another whenever a difference in vapor pressure exists between the two points; the direction of flow is towards the point of low vapor pressure. The rate at which the heat or water vapor will flow varies with the resistance to flow between the two points in the material. If the temperature and vapor pressure of the water vapor correspond to saturation conditions at any point, condensation occurs. HEAT FLOW THRU BUILDING ‘I = UAAt, where q = heat flow, Btu/hr U = transmission coefficient, Btu/(hr)(sq ft)(dcg F temp cliff) A = arca of surface, sq ft At, = equiv temp diff F Heat loss tllru the exterior construction (walls and is normally calculated at the time of greatest bent /loru. This occurs early in the morning after a roof) few hours of very low outdoor temperatures. This approaches steady state heat flow conditions, and for all practical purposes may be assumed as such. Heat flow thru the interior construction (floors, I ceilings and partitions) is caused by a diflerence in temperature of the air on both sides of the structure. This temperature difference is essentially constant thruout the day and, therefore, the heat flow can be determined from the steady state heat flow equation, using the actual temperatures on either side. STRUCTURES Heat gain thu the exterior construction (walls and roof) is normally cnlculnted at the time of greatest heat /?OZO. It is caused by solar heat being absorbed at the exterior surface and by the temperature difference between the outdoor and indoor a:-- Both heat sources are highly variable thruout 2, one day and, therefore, result in unsteady state heat flow thru the exterior construction. This unsteady state flow is difficult to evaluate for each individual situation; however, it can be handled best by means of an equivalent temperature difference across the structure. , The equivalent temperature difference is that temperature difference which results in the total heat flow thru the structure as caused by the variable solar racliation and outdoor temperature. The equivalent temperature difference across the structure must take into account the different types of construction and exposures, time of day, location of the building (latitude), and design conditions. The heat flow thru the structure may then be calculated, using the steady state heat flow equation with the equiv--- --.__ alent temperature difference. - il EQUIVALENT TEMPERATURE DIFFERENCE SUNLIT AND SHADED WALLS AND ROOFS The process of transferring heat thru a wall under indicated unsteady state conditions may be visualized by picturing a 12-inch brick wall sliced into 12 one-inch sections. Assume that temperatures in each slice are all equal at the beginning, and that the indoor and outdoor temperatures remain constant. When the sun shines on this wall, most of the solar heat is absorbed in the first slice, Fig. 22. This raises the temperature of the first slice above that of the outdoor air and the second slice, causing heat to flow to the outdoor air and also to the second slice, Fig. 23. The amount of heat flowing in either direction depends on the resistance to heat flow within the wall and thru the outdoor air film. The heat flow into the second slice, in turn, raises its temperature, causing heat to flow into the third slice, Fig. 24. This process of absorbing heat and passing some on to the next slice continues thru the wall to the last or 12th slice where the remaining heat is transferred to the inside by convection and radiation. For this particular wall, it takes approximately 7 hours for FIG. 22 - SOL,\R HEAT ARSORBED IN FIRST S LICE FIG. 25 - BEIIAVIOR SECOND T IME HEAT 01; ABSORBED INTERVAL AUSORBED PLUS DURING S OLAR HEAT AD D I T I O N A L DURING S OLAR T HIS INTERVAL ’ . - B EHAVIOR FIG. 23 DURING OF S ECOND ABSORBED S OLAR HEAT T IME INTERVAL 3 FIG. - FIG. 24 - B EHAVIOR DURING OF T HIRD ABSORI~ED TIME S OLAR HEAT INTERVAL solar heat to pass thru the wall into the room. Because each slice must absorb some heat before passing it on, the magnitude of heat released to inside space would be reduced to about 10% of that absorbed in the slice exposed to the sun. These diagrams do not account f o r possible changes in soln~ intensity or outdoor temperature. 26 T HIRD -B EHAVIOR OF ABSORBED S OLAR HEAT T IME INTERVAL PLUS ADDITIONAL S OLAR ABSORBED DURING T HIS INTERVAL DURING HEAT The solar heat absorbed at each time interval by the outdoor surface of the wall throughout the day goes thru this same process. Figs. 25 and 26 show the total solar heat flow during the second and third time intervals. A rise in outdoor temperature reduces the amount of absorbed heat going to the outdoors and more flows thru the wall. This same process occurs with any type of wall construction to a greater or lesser degree, depending on the resistance to heat how thru the wall and the thermal capacity of the wall. NOI‘E: The thermal capacity of a wall or roof is the density of the material in the wall or rool’, times the specific heat of the material, times the voli~n~c. This progression of heat gain to the interior may occur over the full 24-hour period, and may result in a heat gain to the space during the night. If the equipment is operated less than 24 hours, i.e. either skipping the peak load requirement or as a routine procedure, the nighttime radiation to the sky and the lowering of the outdoor temperature may decrease the transmission gain and often may reverse it. Therefore, the heat gain estimate (sun and transmission thru the roof and outdoor walls), even with equipment operating less than 24 hours, may be evaluated by the use of the equivalent temperature data presented in Tables 17 and 20. Basis of Tables 19 and 20 - Equivalent Temperature Difference for Sunlit Shaded Wails and Roofs and tions using Schmidt’s method based on the following conditions: 1. Solar heat in July at 40” North latitude. 2. Outdoor daily range of dry-bulb temperatures, 20 deg F. 3 . hIaximum outdoor temperature of195 F db and a design indoor temperature of 80 F db, i.e. a design difference of 15 deg F. 4 . Dark color walls and roofs with absorptivity of 0.90. For light color, absorptivity is 0.50; for medium color, 0.70. 5 . Sun time. 3e specific heat of most construction materials is &ipproximately 0.20 Btu/(lb)(deg F); the thermal capacity of typical walls or roofs is proportional to the weight per sq ft; this permits easy interpolation. for Sunlit Civcn: .\ flat roc,f cxposc’l to the sun, w i t h I)uilt-up roofing, 1% in. insulation, 3 in. woo(l tlcck anti suspentlctl acoustical tile ceiling. Room design tcmI)erature = X0 1: tll) Olittloor tlcsign tcmpfratul-e = 95 F tll) I)aily range ‘= 20 tleg F Find: Equivalent temperature tlilfercncc at 4 p.m. July. Solution: avt/sq ft = 8 + 2 + 2 = 12 Il,/sq ft (Table -77. @ge i/i Equivalent temperature difference = 43 tleg I; (Table 20, intcrpolatetl) Example 2 - Daily Range and Design Temperature Difference Correction At times the daily range may be more or less than 20 deg F; the difference between outdoor ancl room clesign temperatures may he more or less than 15 deg F. The corrections to be applied to the equivalent temperature cliff rence for combinaf tions of these two variables are listed in the notes following Tables 19 and 20. Tubles I9 nnd 20 are analogue computer calcula- Use of Tables 19 and 20 - Equivalent Temperature Difference Shaded Walls and Roofs Example J - Equivalent Jemperafure Difference, Roof and The equivalent temperature differences in Tables 19 and 20 are multiplied by the transmission coefFicients listed in Tables 21 thru 33 to determine the heat gain thru walls and roofs per sq ft of area during the summer. The total weight per sq ft of walls and roofs is obtained by adding the weights per sq ft of each component of a given structure. These weights are shown in italics and parentheses in Tables 21 thou 33. Given: The same roof as in Exa~r~/Ae I Room design temperature = 78 F dl) Outdoor design temperature = 95 F tlh Daily range = 26 cleg F Find: Equivalent tempqrature difference I under changed conditions Solution: Design temperature ciifference = 17 deg F Daily range = 21.3 deg F Correction to eqttivalent temperature difference = -1 cleg F (TaDle 20A, interpolated) Equivalent temperature difference = 43 - 1 = 42 deg F Example 3 - Other Months and Latitudes Occasionally the heat gain thru a wall or roof must be known for months and latitudes other than those listed in Note 3 following Table 20. This equivalent temperature difference is determined from the equation in Note 3. This equation adjusts the equivalent temperature difference for solar radiation only. Additional correction may have to be made for differences Ijetween outdoor and indoor design temperatures other than 15 cleg F. Refer to Tables 19 and 20, pages 62 and 63, ancl to the correction TaBle 20A. Corrections for these differences must Ile made first: then the corrected equivalent temperature differences for both sun and shade must be applied in corrections for latitude. Given: 12 in. common brick wall facing west, with no interior finish, located in New Orleans, 30” North latitude. Find: Equivalent temperature difference in November at 12 noon. Solution: The correction for design temperature difference is as follows: ‘J Outtloor tlcsign tlry-l)ull) =!)r,-ir,=XOF lw~pcr;~turc in Novcrnlvx at 3 p.m. A’,~,,, for west wall, in sun = i (Tmhlr 1 9 ) - I I .5 = - 4,s tlcg 1; TABLE 19-EQUIVALENT TEMPERATURE DIFFERENCE (DEG F) FOR DARK COLOREDt, SUNLIT AND SHADED WALLS* / or/ -Jq Based on Dark Colored Walls;b“;“F db Outdoor Design Temp; Constant 80 F db Room Temp; 20 deg F Daily Range; 24-hour Operation; July and 40” N. Lat.i SqN WEIGHT OF WALLf ( I b / s q ft) EXPOSURE TIME I a /9 110 I 11 I 12 i 101 South % 71 61 31 2 21 Southwest 31 41.5 - 2 - 3 -4 -2 2 1 O - l 6 6 5 5 9 8 7 7 0 -1 4 9 8 13 13 -2 -3 -3 3 110 .7 7 6 12 12 12 0 -1 5 4 9 9 11 1 1 -1 -2 -2 3 3 2 8 8 7 10 10 9 11 2 8 10 1 5 10 8 0 1 7 9 1 4 9 8 1 6 9 0 4 9 8 0 -1 O-1 6 5 8 7 -1 3 8 8 -1 3 7 8 1 0 0 -1 -1 5 4 3 3 2 12 11 10 9 8 20 18 16 15 13 west Y 11 4 7 9 East Y-4 0I I i I A M PM Northeast Southeast Northwest 6 6 6 -2 1 -;;-:, 0 ) 61 7) 81 4 IO 0 6 4 LJ ;, 0 11 A M 6 a 10 ; 6 6 12. 14 ; Jl; 0 0 1 "12 1 2 2 3 7 8 5 6 7 PM I , 9 13 12 10 1; 12 12 5. 5 3 4 5 4 I I I . SUN TIME E q u a t i o n H: e a t G o i n T h r u W a l l s , B t u / h r= ( A r e a , s q f t ) x ( e q u i v a l e n t t e m p d i f f ) x ( t r a n s m i s s i o n c o e f f i c i e n t U , T a b l e s 2 1 thru 25) * A l lv a l u eosr e f o r b o t h i n s u l a t e d a n d u n i n s u l a t e d w a l l s . i F o ro t h e r c o n d i t i o n s , r e f e r t o c o r r e c t i o n p s ao g n e6 4 . f " W e i g h t p e r s q f t " v a l u e s f o r c o m m o n t y p e s o f c o n s t r u c t i o n o r e l i s t e d i n Tables 21 thru25. F o r w a l lc o n s : r u c t i o n sl e s s t h a n 2 0 I b / s qi t , u s e l i s t e d v a l u e s o f 2 0 I b / s qf t . ‘1 !. i \ C:l~i\I’~I‘El< .5. HE.\‘I‘ ,\NI) W.\~I‘I:K \‘.\I’OR Ipl.OW ‘1.1~IRlJ l-63 S’I‘lIII(:‘l‘IJl<~S TABLE 20-EQUIVALENT TEMPERATURE DIFFERENCE (DEG F) ‘71. FOR DARK COLOREDi, SUNLIT AND SHADED ROOFS* Based on 95 F db Outdoor Design Temp; Constant 80 F db Room Temp; 20 deg F Daily Range; 24-hour Operation; July and 40’ N. Lat.t CONDITION rosed to SUn WEIGHT OF ROOFS (lb/v ft) 10 20 40 60 80 Covered with Water Sprayed SUN TIME I 6 A f , M .\ AM PM 10 11 12 1 2 3 4 j 6 7 8 9 1 2 - 4 -6 - 7 -5 -1 0 - 1 - 2 - 1 2 4 3 2 3 6 9 8 6 7 8 7 9 15 24 32 38 43 46 45 41 35 20 22 16 10 7 3 16 23 30 36 41 43 43 40 35 30 25 20 16 16 I6 23 22 22 28 27 26 33 31 28 38 35 3_2 40 3% 35 41 39 37 39 38 37 35 36 35 32 34 34 28 31 34 24 28 32 15 20 25 30 12 17 22 27 8 13 18 23 19 13 22 15 20 15 18 16 16 15 14 15 12 14 10 12 6 10 2 16 10 5 10, 1 2 14 15 16 15 14 12 2 7 10 1 5 8 1 - 1 - 2 - 3 - 4 - 5 3 1 -1 - 2 - 3 - 3 6 4 3 2 1 0 I5 9 5 18 17 16 15 14 12 10 6 2 1 13 14 14 14 14 13 12 9 7 5 0 - 1 - 2 - 2 - 3 - 3 3 1 0 0 - 1 - 1 8 10 12 13 14 13 12 11 10 8 6 6 7 a 13 11 12 10 11 13 2 4 10 -I -I 0 -2,-2 - 2 - 2 5 12 20 40 60 - 5 -2 - 3 - 2 20 40 60 - 4 -2 - 2 -2 - 1 - 2 -I Shaded 9 11 0 0 2 4 8 12 - 1 -I 0 2 5 - 2 - 2 - 2 0 2 7 lop12 4 3 4 5 1 - 1 - 3 6 4 2 11 9 6 16 13 11 20 18 14 2 1 0 - 1 3 4 ” 6 7 8 9 10 11 12 1 2 3 4 AM 5 9 10 11 12 1 2 A PM M 5: :-..I SUN TIME Equation: Heat Gain Thru Roofs, sq ft) X (equivalent temp diff) X (t ransmission Btu/hr = (Area, *With attic ventilated and ceiling insulated roofs, reduce equivalent temp diff For peaked roofs, use the roof oreo projected on CI horizontal plane. coefficient U, Tables 27 or 28) 25yo. tFor other conditions, refer to corrections below and on page 64. :“Weight per sq ft” values for common types of construction are listed in Tables 27 or 28. TABLE 20A-CORRECTIONS TO EQUIVALENT TEMPERATURES (DEG F) OUTDOOR FO:E%NNTH/’ AT 3 P.M. -/MINUS ROOM TEMP (deg F) -30 -20 -10 0 5 10 15 20 25 30 35 40 DAILY RANGE (deg F) 12 14 1‘6 -41 -42 -31 -32 -2.L -22 Lll J-12 la 20 22 24 26 -a3 , - 4 4 -33 -34 - e -24 -14 -,;3 -45 -35 -25 -46 -36 -26 - 1 5 -16 -47 -37 -27 -17 -48 -38 -28 -18 2;. --;I, -12 - 7 - 2 3 8 13 18 23 -4’ . -’ - 1 4 9 14 19 24 29 ‘1’2 _Tgr \- 3.: B-;,.7 13 ia 23 28 2”, 12 17 22 27 ;J+ ‘Ll ‘b 11 16 21 26 0‘ 3” - 1 4 10 9 15 20 25 14 19 24 28 30 32 34 -50 -40 -30 -20 -51 -41 -31 -21 -52 -42 -32 -22 36 -53 -43 -33 -23 -13 -8 - 3 2 -15 -10 -5 0 -16 -17 -12 - 7 - 2 -ia -13 - a - 3 7 12 17 22 5 10 15 20 4 9 14 19 3 8 13 18 2 7 12 17 - - 4 3 2 1 9 9 9 9 6 11 16 21 -11 -6 - 1 38 - 5 4 3 2 4 4 4 4 1 6 11 16 -55 -45 -35 -25 1-64 I’AKT Corrections to Equivalent Temperature Differences in Tables 19 & 20 for Conditions Other Than Basis of Table I . LO.\D ES-I‘lMr\-I‘ING Al,,* cliffercncc for same wall or = cclilivalcnt tcmpcrattirc roof in shade at clesirctl time of (lay. corrcctecl if neccsqary for design conditions. 3 ,‘,I& = eclilivalent tcmperatnrc ciilFcrcnce for wall or roof cxl~~secl to the sun for the clesirccl I ime of clay. carrcctccl if ncccssary for tlesign conditions. Note: Light color = white, cream, etc. ;\Ieclilim color = light green, light blue, gray, etc. I)ark color = Clark Ijlue, dark red, dark brown, etc. 5. Other Iatitllcle, other month. light or metlicim color walls or roof. The coml,inecl formulae arc: Light color walls or roof 2. Shaclccl walls For sha~lccl wa1l.s on any cxpc~surc, tlse the values of equivalent temperature clifferencc listed for north (shacle), carrcctecl if necessary as shown in Correction I. 3. Latitudes other than 40” North and for other months with cllfferent solar intensitlcs. Tables 19 mtl 20 values are~approximately correct for the east or west wall in any latitude(luring the hottest weather..- In lower latituclcs when the maximum solar altitude is 80” to 90” (the maximum occurs at noon), the temperature difference for either south or north wall is approximately the same as a north or shade wall. See Table 18 for solar altitude angles, Meclium color walls or roof. R At,, = .78 j{f 1tc,,, Tab/r J9: Northeast East Southeast South Southwest West Northwest North (shade) = equivalent temperature difference for month and time of clay desired. equivalent temperature difference for same wall or roof in shade at desired time of clay, corrected if necessary for design conclitions. 4t equivalent temperature difference for wall or roof En1 exposed to the sun for desired time of clay,-corrected if necessary for design conclitions. -\ sRs! = maximum solar heat gain in I%tu/(hr)(sq ft) thru glass for wall facing or horizontal for roofs, for month and latitude desired, Table 15, page q-1, or 7 Table 6, l-‘age 29. ft) thru R,,:, = maximum solar heat gain in Btn/(hr)(sq glass for wall facing or horizontal for roofs, for July at 40” North latitude, Tnble 15. page ff, or = Table 6, jmge 39. .Qaffrfile 3 illustrates the procedure. 30 ltc = ltcs -f To (At< ,,,, - -\t(,J hfeclium TRANSMISSION COEFFICIENT U thru structure. The reciprocal 0E building structure t o t a l rcsist:lnce heat How thru the the U value for any wall is the total resistance of this wall ‘Phe a ft)(deg F temp cliff). T h e r a t e t i m e s How of heat. the to the of any wall to heat flow is s u m m a t i o n o f t h e r e s i s t a n c e i n e a c h c o m p o n e n t 0E the structure and surface in Tnb1e.s most the films. common resistances The -31 tlcrzl tyllcs of the transmission outdoor and coefficients 33 have been calculated for of construction. Basis of Tables 21 thru 33 - Transmission Coefficients U for Walls, Roofs, Partitions, Ceilings, Floors, Doors, and Windows = 55 4tc,,L + .45 4t,,* Tnbies -71 that 3 3 c o n t a i n c a l c u l a t e d U v a l u e s .70 4tca based on the resistance The where: roof transferred the tcmperaturc difference is the listed 4tc = At(,n + 3 (4\t(,,,, - 4to) = .i8 4tc,,; + .22 or heat is i n Ktu/(hr)(scl the color wall or roof: = equivalent or U value is the rate at Transmission coefficient which inside 4. Light or meclium color wall or roof Light color wall or roof: 4tc Use Exposure Value Southeast East Northeast North (shade) Northwest West + Southwest South .hutJ2 IA titlltle where 4tcr . 6. I;or South latitudes, use the following exposure values from The temperature differential Ate for any wall facing or roof and for any latitude for any month is approximatecl as follows: 4te R + (I - .78 2) -Itcs !!I temperature desired. difference for color of wall for listed in Tnble resistance of the outdoor surfxc and winter conditions and the inside film is listed in Table 3f. summer ~surface 34, j3nge 78. film coefficient Note: Tlrc tliffercrrcc ljetwecn summer and winter, tr;rrlsmissiori cocffkients for ;I typical wall is ricgligil)le. For example, with ;I trnnsrrrission coefficient of 0.3 I~tu/(hr)(sq ft) (F) _ [or w i n t e r contlitions, t.lic c o e f f i c i e n t f o r summer conditions will be: I. Thcr-mal rcsistancc zz R (winter) ol twll ) Example 4 - Transmission Coefficients Civcn: brasonry partition matlc of 8 in. I~ollow clay tile, Itoth sides finished, metal lath plasterctl on furring with ‘% in. sand plaster. Find: Transmission coelfirient Solution: Transmission coefficient U = 0.18 I%tu/(hr)(sq ft)(tleg F), Table 26, /xtge 70 1 I -= - = 3.33 0.3 I/ 2. Cktdoor film therm;\1 resistance (winter) = 0.17 (Tnble 34) 3. Thermal resistance of wall without outdoor air film (wiritcr) = 3.33 - 0.17 = 3.16 -1. Outdoor film thermal rcsist:ince = 0.25 (Table 34) 6. Transmission coefficient U of wall in summer 1 1 =----_ = 0.294 R 3.41 7. Difference between summer- and winter transmission becomes greater with larger U values and less with smaller U values. Ceilings, Floors, Doors, U and for Wails, 5 - Transmission Insulation Coefficient, Addition of The transmission coefficients listed in Tables 21 thru 30 do not include insulation (except for flat roofs, Table 27, page 71). 5. Thcrrnal rcsistancc of wall with outdoor air film (surnrncr) = 3.16 + 0.25 = 3.41 Use of T,ables 21 thru 33 - Transmission Coefficients Example (summer) Roofs, Partitions, Windows The transmission coefficients may be used for calculating the heat flow for both summer and winter conditions for the average application. Frequently, fibrous insulation or reflective insulation is included in the exterior I)uilding structure. The transmission coefficient for the typical constructions listed in Tables 21 thru 30, with insulation, may Ire determined from Tabfe3Z,page 75. Given: Masonry wall consisting of 4 in. face brick, 8 in. concrete \cinder block, metal lath plastered on furring with sh in. ’ sand plaster and 3 in. of fibrous insulation in the stud space. Find: Transmission coefficient. Solution: Refer to Tables 22 and 31. C’ value for wall without insulation = 0.24 Btu/(hr)(sq ft)(dcg F) CT value for wall with insulation = 0.07 I%tu/(hr)(sq ft)(deg F) l-66 \ I’,\RT I . LOI\I) ESl‘I~I.\‘I‘IN(; + TABLE 2 1 -TRANSMISSION COEFFICIENT U-MASONRY WALLS* FOR SUMMER AND WINTER Btu/(hr) (sq ft) (deg F temp diff) All numbers in parentheses indicate weight per sq f t . T o t a l w e i g h t p e r s q f t i s s u m o f w a l l a n d f i n i s h e s . INTERIOR %” IHICKNESS inches) and YElGHT (lb p e r sq ftl G iypsum %* Board Plaster on Well NOW ( Plaster IBoard) (21 T gq-!g SOLID BRICK Face 8 Common 8 (87) 2 (123) I6 (173) .48 2.5 .27” .41 .31 .25 .45 .33 .26 .41 .30 .25 Common Only. a (80). I2 (1201 16 (1601 .41 .31 .25 .36 .2a .23 .39 20 .24 .35 .27 .23 (100: (15-o: (200’ (300’ .67 .55 .47 .36 .55 .47 .41 .32 .63 .52 .45 .35 8 (26) 12 (40) .34 .25 .30 .23 6 8 10 12 (70) (93) (117 (140 .75 .67 .61 .55 6 140) 8 (53) 10 (66) 12 (80) 6 (15) 8 (20) FINISH I/s If Metal Lath Plastered on Furring =h/4n Sand P‘laster(l) Yin 1t wt Plaster(J) Insulating Board Plain or Plastered on Furring Gypsum or Wood Lath Plastered on Furring l/i ‘I t/$ # Sand 1t wt Plaster(7J “:, Plaster(Z) .31 .25 .21 .28 .23 .I9 .29 .23 .20 .27 .22 .I9 .22 .19 .I6 .16 .I4 .I3 .28 .23 .19 .26 .?2 .I8 .26 .22 .18 .25 .21 .I8 .21 .18 .16 .I5 .14 .I2 .53 .46 .40 .32 .39 .34 .31 .26 .34 .31 .28 .24 .35 .31 .2a .24 .32 .29 .27 .23 .26 .24 .22 .19 .I8 .17 .I6 .I5 .32 .24 .30 .23 .25 .20 .23 .18 .23 .I8 .22 .18 18 .15. .12 .I4 .55 .49 .44 .40 .69 .63 .57 .52 .58 .53 .49. .45 .41 .39 .36 .34 .36 .34 .32 .31 .37 .35 .33 .31 .34 .32 .31 .29 .27 .26 .25 .24 .18 .17 .17 .16 .31 .25 .21 .18 .2a .23 .19 .17 .30 .24 .20 .I7 .27 .23 .19 .I5 .23 .19 .17 .15 .21 .I8 .I6 .14 .22 .18 .I5 .I4 .21 .I8 .14 .14 .I8 .16 .14 .I2 .14 .12 .I 1 .lO 12 (30) .13 .lO .08 .07 .13 .lO .OE .07 .13 .I0 .08 .07 .13 .I0 .08 .07 .I2 .09 .08 .07 .I 1 .09 .07 .07 .I 1 .09 .08 .06 .I 1 .09 .07 .06 .13 .lO .08 .07 .09 .07 .06 .06 8 (43) 12 (63) .52 .47 .44 .41 .48 .45 .43 .40 .23 .22 .17 .16 a (37) 12 (53) 8 (32) 12 (43) .39 .36 .35 .33 .37 .35 .34 .32 .20 .19 .15 .I5 .35 .32 .32 .29 .34 .31 .31 .28 .19 .18 .15 .14 8 (39) 10 (44) 12 (49) .36 .32 .29 .32 .29 .27 .34 .31 .28 .32 .28 .26 .I9 .18 .17 .15 .14 .13 - . - - STONE 8 12 16 24 ADOBE-BLOCKS OR BRICK POURED CONCRETE 140 lb/w ft 80 lb/w f t 30 lb/w ft HOLLOW CONCRETE BLOCKS Sand 8 G r a v e l Agg Cinder Agg Lt Wt Agg STUCCO HOLLOW ON CLAY TILE 1 0 (25) 1 +-g-g-+ 1958 ASHAE Guide Equations: Heat Gain, Btu/hr = (Area, sq ft) X (U value) x (equivalent temp diff, Table 19) Heat Loss, Btu/hr = (Area, sq ft) x (U value) x (outdoor temp - inside temp) ‘For addition of insulation and air spacer to above walls, refer to Table 31, page 75. - CII;\I”I‘I:I: 5. 111:.\‘1‘ ,\Nl) W \‘I’I~:IC V.\I’OII I;I>O\zi ‘1f11<1J S’I’I~IJ(:~I‘I_Ill~:.S 1-67 TABLE 22-TRANSMISSION COEFFICIENT U-MASONRY VENEER WALLS* FOR SUMMER AND WINTER Btu/(hF) (sq ft) (deg F temp diff) All numbers in parentheses indicate weight per rq ft. Total weight per sq ft is sum of wall and finishes. EXTERIOR FINISH I INTERIOR BACKING -or4” stone (50’ -orPrecast concrete (Sand Age’ 4” 8 6” (39’ (58’ Ad 8” 8 10” (78’ (98’ -or- lh ” Board 1” Board Plader(7) Plaster(J) PlasterC7) Plaster(l) (21 141 .26 .23 .22 .25 .21 .21 .21 .I8 .18 .I6 .I4 .I4 L .41 .37 .30 .29 .39 ..,f2 .30 (17’ .35 .32 .34 .31 .25 .23 .24 .22 .19 .I5 (32’ (43’ .30 .28 .26 .28 .29 .27 .27 .25 .23 .21 -35 .:?.! - - - -r __, .24 .28 .23 .26 .22 .21 / .21 .20 .21 .20 .20 .I9 .17 .I7 .I4 .I3 .46 .39 .37 .41 .35 .33 .32 .28 .27 .29 .26 .25 .29 .26 .25 .27 .25 .24 .22 .21 .20 .I7 .16 .I5 Hollow Clay Tile 4 (16’ 8 .1301 ~ 12 (40’ .41 .31 .26 .25 .25 .24 .20 .I9 .19 .I8 .16 .13 .46 .34 .41 .31 .32 .25 .29 .23 .29 .24 .27 .22 .22 .I9 .16 .15 Concrete (Lt Wt Ad 80 Ib/cu ft Sand 8. Gravel Ad I (Sand & Gravel Agg’ I I I I 4 (40’ a (80’ .49 .35 .42 .31 4 (20’ 8 (37’ 12.(53’ .36 .29 .28 .33 .28 .26 .35 .29 .27 .32 .26 .25 .26 .22 .21 .24 .21 .20 .24 .21 .20 .23 .20 .I9 .I9 .I7 .I7 .I5 .14 .13 4 (17’ 8 (32’ 12 (43’ .32 .27 .25 .29 .26 .24 .30 .26 .25 .2a .25 .23 .23 .21 .20 .22 .20 .I9 .22 .20 .19 .21 .I9 .18 .18 .17 .16 .14 .13 .13 4 (23’ 8 I431 12 (631 .42 .36 .34 .38 .33 .32 .40 .35 .33 .36 .32 .30 .29 .26 .25 .26 .24 .23 .27 .24 .23 .25 .23 .22 .21 .I9 .19 .16 .15 .15 1 I I Hollow Clay Tile concrete (Lt Wt Agg’ 80 lb/w ft (Sand 8 Gravel Agg’ - o r - Common Brick Equations: Heat Gain, Btu/hr Heat l/i ” Lt wt .44 .37 .35 4w concrete Block (23’ (Sand Age) v stone (100’ (3) ‘ii “ Sand =hft Lt wt .49 .41 .3a Brick (40) _ (Sand Ad I J? * Sand 4 (23’ 8 (43). 12 (63’ (Lt Wt Age’ -or- Lt wt be insulating Board Plain or Plastered on Furring (Sand 8 Gravel Atxd Concrete Btock (Cinder Agg’ Precast concrete Sand Aw J/“ Gypsum or Wood Lath Plastered on Furring Metal Lath Plastered on Furring n“ Plaster on Wall (6’ 420’ Common Brick 4” Common Gypsum Board (Plaster Board’ (2) None Concrete I” ‘(’ Block -- (Cinder Agg’ (Lt Wt 4” Face Brick (43’ THICKNESS (inches’ and WEIGHT (lb per FINISH L OSS, Btu/hr ! = (Area, sq ft) x (U value) = (Area, ~q ft) x (U v&e) 1958 ASHAE Guide x (equivalent temp diff, Table 19) x (outdoor temp - inside temp’ *For addition of insulation and air spacer to walls, refer to Table 31. page 75. /’ ,’ ’ l-68 l’;\R~I‘ TABLE 23-TRANSMISSION COEFFICIENT U-LIGHT CONSTRUCTION, I. LO,\rJ ES~I‘I~l.\TlNG INDUSiRlAL WALLS*1 FOR SUMMER AND WINTER * Btu/(hr) (sq ft) (deg F temp diff) All numbers in parentheses indicate weight per sq ft. Total weight per sq ft is rum of wali and finishes. INTERIOR WEIGHT (lb per rq ftl FINISH Insulating Board Fiat Iron III NOna Wood 2%” %n EXTERIOR FINISH =Yifr (21 (3) (2) SHEATHING %n Corrugated Transite None ‘Vi” Ins. Board %z” Ins. Board II) I21 24 Gauge Corrugated Iron NO”0 ‘h” Ins. Board ‘5%d’ Ins. Board $4” W o o d (11 3/I” W o o d Siding (2) (2) (2) 01 12) NOW 1.16 .34 .27 .55 .26 .2l .32 .19 .17 .26 .17 .15 .36 .21 .18 1.40 .36 .28 .46 .60 .27 .22 .33 .33 .20 .17 .22 .27 .17 .I5 .19 .3a .21 .18 .24 .58 .37 .25 .21 .27 1 . 1958 ASHAE Guide Equations: Heat Gain, Btu/hr = (Area, sq ft) X NJ value) X (equivalent temp diff, Table 191. Heat loss, Btu/hr = (Area, sq ft) X iU v&e) x (outdoor temp - inside temp). *For addition of air spacer and insulation to walls, refer to Table 31, page 75. $Values apply when sealed with calking compound between sheets, and ot ground and roof lines. When sheets are not sealed, increase U factors by 10%. These values may be used for roofs, heat flow up-winter; for heat flow down-summer, multiply above factors by 0.8. TABLE 24-TRANSMISSION COEFFICIENT U-LIGHTWEIGHT, PREFABRICATED FOR SUMMER AND WINTER CURTAIN TYPE .WitLLS* ., Btu/(hr) (sq ft) (deg F temp diff) All numbers in parentheses indicate weight per sq ft. Total weight per sq ft is sum of wall and finishes. METAL FACING Core INSULATING CORE MATERIAL 1 2 M E T A L F A C I N G VVlT” %” AIR SPACE (3) (3) Thickness (in.) 3 Core Thickness 4 1 2 (in.) Glass, Wood. Cotton Fibers Paper Honeycomb Paper Honeycomb with Perlite Fill, Foamglar Fiberboard Wood Shredded (Cemented in Preformed Slabs) Expanded Vermiculite 3 5 9 I5 22 7 .2l .39 .29 .36 .3l .34 .12 .23 .I7 .21 .I8 .20 .08 .17 .12 .15 .13 .14 .06 .I3 .09 .I2 .10 .I1 .19 .32 .25 .29 .25 .28 .l 1 .20 .15 .I9 .16 .18 ;; .I I .l4 .12 .13 .09 .1 1 .09 .lO Vermiculite 20 30 or Perlite concrete 40 60 .44 .5l .58 .69 .27 .32 .38 .49 .19 .24 .29 .39 .15 .19 .23 .31 .35 .39 .43 .49 .23 .27 .31 .18 .21 .25 31 .14 .17 .20 .26 Equations: Heat Gain, Btu/hr = (Area, sq ft) x (U value) X (equivalent temp diff, Table 19). Heat Loss, Btu/hr = (Area, sq ft) x [U value) X (outdoor temp - inside temp). *For addition of insulation and air spaces to walls, refer to Table 31. page 75. tTotal weight per sq ft = core density Xcore thickness 12 + 3 Ib/sq ft .3a 1 ;; . CHAPTER 5 . HEAT ;\i’iI) W,\T‘I-R V,\I’OR FLOW TABLE 25-TRANSMISSION TTIRIJ 1-69 STRIJCTURES COEFFICIENT U-FRAME WALLS AND PARTITIONS* FOR SUMMER AND WINTER Btu/(hr) (sq ft) (deg F temp diff) All numbers in parentheses indicate weight per sq ft. Total weight per sq ft is sum of component materials. INTERIOR FINISH Metal lath Plastered G %” EXTERIOR FINISH ) SHEATHING Wood Panel (2 1 I %* .91 .6a .48 .42 .32 .33 .30 .25 .23 .20 .42 .37 .30 .27 .23 .31 .29 .24 4” Face Brick V BI (43) OR ywood (I) 01. Asphalt Siding (2) None, Building Paper S/(6” Plywood (I) o r ‘A” Gyp (2) ?h” W o o d & Bldg Paper (2) %” insulating Board (2) ‘%” Insulating Board (3) .73 .57 .42 .38 .30 .30 .28 .23 .22 .I9 .37 .33 .27 .2s .21 .40. .36 .29 .27 .22 Wood Siding (3) OR Wood Shingles (2) OR =/‘I Wood Panels (3) None, Building Paper S/(6” Plywood (1) or %” Gyp (21 %z“ Wood 8, Bldg Paper 1/Z” Insulating Board (2) ‘%” Insulating Board (3) .57 .48 .36 .33 .27 .27 .25 .22 .20 .18 .33 .30 .2s .23 .20 .35 .31 .26 .24 .21 .24 .22 .19 .I8 .16 1%;53, 1;‘;;;;;;;;;~ Inrulated,Siding j! (4) %z” Insulating Board (3) Single Partition (Finish on one side only) Double Partition (Finish on both sides) .43 .24 Woo:L.th Plastered T- Insulating Board Plain 0, Plastered I” ocwd (2) Board (4) None, Building Paper S/(6” Plywood (I) o r l/S” Gyp (2) %O” Wood 8. Bldg Paper (2) ‘h ” Insulating Board (2) %” Insulating Board (31 &wood (I) or %” G y p (2) Gypsum ( OR Asbestos Cement Siding (1) OR Asphalt Roll Siding (2) %/(a” %I” .27 .25 .21 .29 .26 .22 .21 .I8 .20 .19 .17 .I6 .14 .33 .30 .25 .24 .20 .26 .24 .21 .20 .17 .19 .I8 .I6 .15 .I4 .30 .27 .23 .22 .I9 .24 .18 I .22 .I9 .18 .I6 .17 .I5 .14 .I3 , .28 .25 .22 .20 .18 .21 .I9 .17 .16 .I5 .16 .I5 .I4 .13 .I2 .60 .34 .36 .19 .23 .12 .31 .28 .24 .22 .I9 .32 .29 .24 .23 .I9 1958 ASHAE Guide Equations: Walls-Heat Gain, Btu/hr = (Area, sq ft) X IU value) X (equivalent temp diff, roble 19). -Heat Loss, Btu/hr = (Area, sq it) X (U value) X (outdoor temp-inside temp). Partitions, unconditioned rpoce adjacent--Heat Gain or Loss, Partitions, kitchen or boiler room Btu/hr = (Area sq ft) X odjocent-Heat Gain, Btu/hr = (Area sq (U value) X (outdoor temp-inside temp-5 F). ft) X (U value) x (actual temp diff or outdoor &mp-inside temp + 15 F lo 25 F). “For addition of insulation and air spacer to partitions, refer to Table 31, page 75. TABLE 26-TRANSMISSION COEFFICIENT U-MASONRY PARTITIONS* FOR SUMMER AND WINTER Btu/(hr) (sq ft) (deg F temp diff) All numbers in parentheses indicate weight per ~q ft. Total weight per sq ft is rum of moronry unit and finish X 1 or 2 (finished one or both rides). FINISH THICKNESS (inches) and Both NO. U’EIGHT S i d e s of (per UnSides sq f t ) f i n i s h e d F i n i s h e d RACKING . HOLLOW CONCRETE BLOCK Cinder 3 (17) .45 One Both 4 r2y .40 One Both a d7) .32 One Both 1 2 (531 31 Oll.2 Both 3 (151 .3a Olle Agg i ?h” Gypsum Board :Plaster Board) w %” 1% ” Sand tt wt Sand 1t w t A g g (6) A g g (3) Plaster(7) Plaster(3) ._79 .31 .29 30 .26 __1 .27 .29 .25 .26 .34 .31 I .31 .29 .27 .25 11 Wf Agg S a n d B Gravel Am HOLLOW CLAY 1%” aoord(2) Beard(4) .24 .19 .22 .17 .22 .I7 .21 .lb .lB .I2 .23.lB .21 .I6 .22 .17 .21 .15 .I7 .12 .36 .35 .34 .32 .27 / .20 .17 j .I7 .lb ( .29 .2a .27 .24 / / .22 .lE .21 .I6 / 1 .21 .I6 .20 .I5 1 .27 .2b .25 .23 1 .21 .17 .20 .15 I .20 .lb .19 .I5 1” -___ .I4 .09 ’ I( .14 .09 .20 .13 .15’ .09 .19 .13 .15 .09 .17 .I2 .I4 .09 .16 .12 .13 .OB 1 2 (43) / .25 .23 a (431 .40 One Both .36 .32 .39 .37 .35 .31 .2a .21 .26 .19 .26 .19 .25 .lB .20 .13 12 (63) .3a Otle .34 .30 .3b .35 .33 .29 .27 .21 .25 .1a .25 .19 .24 .17 .I9 .I3 .15 .l 1 .15 .09 3 (151 .46 Otle Both .40 .44 .39 .31 .2a .2B .27 .22 .lb 4 (16) .40 One Both 6 (25) .35 Olle Both 8 (30) .31 On.2 Both 3 (9) .37 One Both .35 .32 .26 .24 .24 .23 .I9 .I5 4 (13) .33 On-2 Both Both TILE H O L L O W GYPSUA TILE 1% ” 1% N Sand 1t wt Plaster/7) Plaster(l) ___~ ~ Both i % ” Plaster o n Wail Insulating Board Plain or Plastered on Furring %” Gypsum or Wood Lath Plastered on Furring Metal lath Plastered on Furring .33 / SOLID GYPSUM PLASTER 1958 ASHAE Guide Equations: Partitions, unconditioned space adjacent: Heat Gain or LOSS, atu/hr = (Area, Sq ft) X (U value) x [outdoor temp-inside temp-5 FL Partitions, kitchen or boiler room adjacent: Heat Gain or Loss, Btu/hr = (Area, sq ft) X (U value) x (actual temp diff or outdoor temp-inside temp + I5 F to 25 *For addition of insulation and air spaces to partitions, refer to . Table 31, page 75. Fl. (:11,\1”1’1:11 5. TABLE IIE.\‘I’ .\&-I) W.\‘l’l~:I~ 27-TRANSMISSION V.\I’Ol< I~l.OLV COEFFICIENT ‘1‘11111J U-FLAT l-71 S’I‘RIJ(:‘I‘URES ROOFS COVERED WITH FOR HEAT FLOW DOWN-SUMMER. FOR HEAT FLOW UP-WINTER (See Equation Btu/(hr) All numbers in parentheses indicate weight per sq THICKNESS OF DECK (inches) TYPE OF DECK (sq ft) (deg F temp ft. Total weight per sq BUILT-UP ROOFING* of Page). at Bottom diff) ft is sum of roof, finish and insulation. INSULATION ON TOP OF DECK, INCHES CEILING t NO Inrulation (3) 1% ‘/a (1) 2% (2) S u s p e n d e d Plast (lt W t A g g on Gypsum Board) 2 (91 None or Plaster (6) Suspended Plaster (5) S u s p e n d e d A c o u T i l e (2) .27 .I8 .15 .20 .14 .12 .15 .12 .l 1 .13 .lO .09 .l 1 .09 .08 .lO .09 .08 .08 .08 .07 3 (13) None or Plaster (6) Suspended Plaster (5) S u s p e n d e d A c o u T i l e (2) .21 .15 .13 .16 .12 .l 1 .I3 .l 1 .\o .l 1 .09 .OfJ .lO .08 .08 .09 .OE .07 .08 .07 .06 4 (16). None or Plaster (6) Suspended Plaster (5) Suspended Acou Tile(Z) .17 .13 .12 .14 .l 1 .lO .i 1. .lO .09 .lO .08 .07 .09 .08 .07 .08 .07 .06 .07 .06 .05 2 None or Plaster (6) Suspended Plaster (5) S u s p e n d e d A c o u T i l e (2) .32 .21 .17 .22 .17 .13 .17 .13 .I2 .14 .l 1 .lO .12 .lO .09 .lO .09 .08 .09 .08 .07 3 (15) None or Plaster (6) Suspended Plaster (5) S u s p e n d e d A c o u T i l e (2) .27 .19 .I5 .19 .15 .12 .15 .I3 .ll .13 .l1 .09 .l 1 .lO .08 .lO .09 .08 .08 .07 A (19) None or Plaster (61 Suspended Plaster (5) S u s p e n d e d A c o u T i l e (2) .23 .17 .lA .17 .13 .12 .lA .lO _ .12 .l 1 .12 .lO .09 .09 .08 .09 .OE .08 .21 .16 .13 116 .13 .ll .13 .l 1 .lO .ll .09 .09 .lO .09 .08 .09 .08 .07 N Gypsum Slab on %” Gypsum Board N Wood (11) I I i I 1 (3) None or Plaster (6) Suspended Plaster (5) S u s p e n d e d A c o u T i l e (2) 2 (5) None or Plaster (6) Suspended Plaster (5) S u s p e n d e d A c o u T i l e (2) 3 (81 None or Plaster (61 Suspended Plaster (5) S u s p e n d e d Acou T i l e (2) .oa , .oa .07 .07 .08 .07 .06 1958 ASHAE Guide Equations: Summer-(Heat F l o w Down) Heat Gain, Btu/hr = (Area, rq ft) X VJ v a l u e ) X ( e q u i v a l e n t t e m p d i f f , T a b l e Winter-(Heat Flow Up) Heat Loss, Btu/hr = (Area, rq ft) X (U *For a d d i t i o n o f a i r spaces or inrulotion 20). value x 1.1) X (outdoor temp-inside temp). to roofs, refer to Table 31, page 75. tFor suspended I%” i n s u l a t i o n b o a r d , p l a i n (.6) o r w i t h I/” sand aggregate plaster (5). use values of suspended ocou tile. TABLE 28-TRANSMISSION COEFFICIENT FOR HEAT FLOW DOWN-SUMMER. FOR HEAT FLOW Btu/(hr) (sq ft projected area) (deg All numbers in parentheses indicate weight per PITCHED U-PITCHED F temp diff) sq f t . T o t a l w e i g h t p e r sq f t i s r u m o f c o m p o n e n t m a t e r i a l s . CEILING’ ROOFS ‘/‘a” Gypsum Woo: Lath Plastered l/i ” Sand EXTERIOR SURFACE ROOFS* UP- WINTER (See Equation at Bottom of Page) ‘/a ” tt wt Plaster Plaster (5) (2) SHEATHING Insulating Board Plain or ‘Ii” S a n d A g g Plastered Acoustical Tile on Furring 01 J/O Gypsum (2) 1” Board (4) l/i n Tile (2) J/4” Tile (3) % It Board Bldg paper on %/16” Asphalt Shingler, (2) Asbestos-Cement Shingles (3) Sl .27 30 30 .23 .26 .59 .20 .34 .29 .29 .2a .22 .17 .23 .21 .25 .25 .24 .20 .I6 .21 .I9 .37 .33 .33 .3 I .25 .I8 .25 .22 .32 .27 Arph:;t Roll Roofina (1) Eldg paper on %z” wood sheathina (3) .45 .25 .29 .31 .28 .20 .27 .22 .17 .22 .20 Slates (8) T i l e (10) B l d g p a p e r o n +‘kn plywood (2) .64 .29 .36 .38 .34 .35 .47 .26 .19 .26 .23 Sheet letal (1) B l d g p a p e r o n %i wood sheathing (3) I------ .48 .25 .29 .31 .28 .28 .27 .22 .I7 .23 .20 Bldg paper on 1“ x 4” strips (I I .53 .26 .31 .33 .30 .30 .2a .23 .17 .24 .21 .41 .23 .27 .29 .26 .27 .25 .21 .16 .2l .34 .21 .24 .25 .23 .23 .22 .19 .15 .19 Wood Shingles (2) B l d g paper , ’ o n %” plywood (2) Bldg paper on ?/Ls wood sheathing (31 I------ . .19 .17 1958 ASHAE Guide E q u a t i o n s : S u m m e r ( H e a t F l o w D o w n ) H e a t G a i n , Btu/hr = ( h o r i z o n t a l p r o j e c t e d area, sq ft) X (U v a l u e ) X ( e q u i v a l e n t t e m p d i f f , T a b l e 2 0 ) . W i n t e r ( H e a t F l o w U p ) H e a t L o s s , Btu/hr = ( h o r i z o n t a l p r o j e c t e d a r e a , sq ft) X (U v a l u e X 1 . 1 ) X ( o u t d o o r t e m p - i n s i d e t e m p ) . *For addition of air spaces or insulation for above roofs, refer to Table 31, page 75. TABLE 29-TRANSMISSION COEFFICIENT U-CEILING AND FLOOR, (Heat Flow Up) Based on Still Air Both Sides, Btu/(hr) (sq ft) (deg F temp diff) All numbers in parentheses indicate weight, per sq I- ft. Total weight per sq MASONRY CEILING Suspended or Furred Not Furred FLOOR Iah” Wood Block on Slob 1CONCRETE ( SUBFLOOR ! 4 (,y Sand Agg 1 6 (60) 1 ,; Lt Wt Agg 80 Ib/ft” or 2 122) Sand Agg on 5/a” Plywood on 2” x 2” Sleepers 2 (16) 4 (29) 6 (42) Floor Tile $4” Linoleum ACOUStiCol THICKTile NESS NOW Glued (inches) or ,,*fl and %r WEIGHT Sand 1t wt % u Y-4 ” (lb per Plaster Plaster Tile Tile sq ff) 4 (42) 6 (62) 8 (82) 1 0 (102) Lt Wt Agg 80 lb/f@ 2 (19) 4 (31) 6 (44) 2 (24) %“Hordwood Sand Agg 0” %z” Subfloor on 2” x 2” Sleepers Lt Wt 80 lb/f@ Agg 4 (44) 6 (64) 8 10 2 4 6 (84) (104) (20) (33) (46) (51 .44 .41 p3f3) (3) I Ill 1 .36 .34 3; .29 .28 3;. .25 .24 2; .31 .25 .21 .28 .27 .26 .25 .24 .25 .21 .18 .23 .23 .22 .21 .20 1:; .27 .22 .19 .26 .25 .24 .23 .22 .22 .19 .16 .24 .20 .17 .23 .22 .21 .21 .20 .20 .17 .I5 .20 .I7 .I5 .20 .19 .19 .18 .17 .17 .15 .I4 .18 .I6 .14 .18 .17 .17 .16 .16 .16 .14 .13 , ?h” Gypsum or Wood Lath Plastered Metal Lath Plastered %” =A/,” 1%” Insulating Board Plain or ‘/a” Sand Agg Plastered .31 20 2; .28 2; i 1;: .25 1 .19 .31 .30 .28 .27 .26 .26 .22 .18 .25 .24 .23 .22 .21 .21 .18 .16 .18 .28 .27 .26 .25 .24 .24 .20 .17 .23 .22 .21 .21 .20 .20 .17 .I5 .16 .21 .20 .19 .19 .18 1 .32 114 .38 ’ 1;; .27 .23 .I9 .32 .30 .29 .27 .26 .26 .22 .19 .25 .24 .23 .22 .22 .22 .lB .I6 .21 Acoustical Tileon Furring Vi” :;prum ‘vi n Sand l.t wt Sand Lt wt ‘vi * Plaster Plaster Plaster Plaster By.;;d (7) (31 (51 (2) ClJ .36 .28 .23 .32 .31 .29 .28 .27 , ft is rum of ceiling and floor. 1% ” Tile (1) ?firr Tile (1) .16 .16 .22 .22 :;:, .i .19 :;; .20 .17 .I5 .I8 .lB .18 .17 .17 .17 .15 .I3 .I6 .16 .16 .I5 .15 .18 .16 .I4 .I7 .17 .16 .16 .15 .I5 .14 -12 .I5 .I5 .14 .14 .14 .14 .12 .l 1 .22 .21 1;; .32 .26 .21 .30 .28 .27 .26 .25 .25 .21 .18 .24 .23 .22 .21 .21 .21 .18 .16 1” Board (41 2; .19 .I7 .15 .18 .I8 .17 .17 .16 .16 .14 .13 .16 .16 .15 .15 .14 .14 .I3 .12 / :;: .I5 .I3 .I2 .14 .I4 .14 .I3 .I3 .13 .12 .l 1 .13 .13 .I2 .12 .12 .12 .I 1 .099 .15 .13 .12 FRAME CONSTRUCTION CEILING Not Furred Acoustical Tile Glued NolIe FLOOR 1 SUBFLOOR ‘/2 “ Tile =Yiv Tile (1) (1) Suspended or Furred Metal Lath Plastered Yirf Sand vi“ 1t wt 3/s” Gypsum Woo:L.th Plastered vi“ Sand ‘A2 ” Lt wt Insulating Board Plain or 1%” Sand Agg Plastered ‘h ” Plaster Plaster Plaster Plaster Board (7) j (3) (21 (2) (5) Acoustical Tile on Furring or J/s” Gypsum 1” I/$ ff =/qn Board (4) Tile (1) Tile II) ?h” Wood (2) l/d” Hardboard on 3/g” Insulating Board .27 .20 .21 .19 .22 .28 .20 .20 .26 .19 .20 .26 .19 .17 .38 .24 .I8 .19 .17 .19 25~” Wood (21) .24 .I8 .20 .16 .I4 .I6 .I3 .I7 .21 .I6 .l-; .19 .15 ‘s%” Wood (5) .33 .22 .24 .I7 .21 .I6 .25 .I8 .23 .I7 .23 .I7 .22 .17 .I8 .15 .15 .12 .19 .15 .17 .14 ?h” .28 .20 .20 .16 1 .21 .16 .20 .16 .I7 .14 .I4 .12 .I8 .14 .I6 .I3 Wood (5) 2” Wood (81 1958 E q u a t i o n s : H e a t f l o w u p , U n c o n d i t i o n e d s p a c e b e l o w : H e a t G a i n , Btu/hr = (Area, sq f t ) X (U K i t c h e n o r b o i l e r r o o m b e l o w : Heat G a i n , Btu/hr = (Area, sq ft) X (U value) value) ASHAE G u i d e x (outdoor temp - inside temp - 5 F). x (actual temp diff, or outdoor temp - inside temp + I5 F to 25 F). 1-74 TABLE JO-TRANSMISSION COEFFICIENT U-CEILING AND FLOOR, (Heat Flow Down) Based on Still Air Both Sides, Btu/(hr) (sq ft) (deg F temp diff) All numbers in parentheses indicate weight per sq ft. Total weight per sq ft is rum of ceiling and floor. F u r r e_d _ r--~N o t-T I THICKNESS :inches) and WEIGH1 (lb per sq ft) 2 4 6 8 IO Sand Agg or $6” linoleum or Floor Tile CEILING .- MASONRY .___~ Suspended or Furred -.-~~- ,-.-- Acoustical NOlV3 or % ” Sand lostet til .48 .44 .4l -39 i i .43 .40 .37 .35 .31 .30 .26 .25 32 .31 - ---1~~-‘~-T~I-nru a na .29 .28 .30 .28 .28 .27 .23 .22 .23 .22 .22 .21 20 .20 .20 .I9 .I9 .I8 .20 .17 .I5 .18 .16 .I4 .20 .19 .18 .18 .I7 .18 .17 .17 .16 .16 (42) .17 .15 .I4 .16 .14 .I3 (221 (42) (62) (82) (102 .20 .19 .I8 .I8 .17 .I7 .I7 .I6 .16 .I5 .17 .15 . -. .14 .17 .16 .16 .I5 .I5 .15 .I4 .I3 Tile (1)(1) j Tile (1)(II (191 (39) (59) (79) (99) .17 .I7 2 (15) 4 (28) 6 (41) 2 4 6 8 10 4cousticol Tile on Furring or ?/a” Gypsum 4 c2Ol (40) (60) (80) (100 2 (16) _-Floor Tile or l/8” Linoleum on ?‘a” Plywood on 2” x 2” Sleepers %” 4 (29) 6 Sand Agg I Lt Wt Agg 8 0 Ib/ft3 r- Hardwood 2 4 6 8 10 Sand Agg I -I VIZ” ::bfloor L--on Lt Wt Agg 2” x 2” Sleepers 8 0 Ib/ft’ 2 (19) 4 (31) 6 (441 .2l __.21 .18 .16 .I9 .19 .I6 .I4 I I%” Ceramic Tile on, ‘,I/*” rrmrnt on1 4x” Cement ‘%6” Hard ‘=/(6” Hardwood Floor or L1no or linoleum .--... on -.. %” Plywood Plvwnnd Vi” I, ‘% Linoleum on Vi” Hordboard on ‘:, % Insulating %” Insulating Board . . I . - I - - . - ; %” - . . I- -. I?! / .20 I .20 .19 t .__~ .20 .I9 .17 .15 .2l .18 (24) (44) (6-i) (84) (104 .26 .25 .24 .23 .22 .25 .24 .23 .22 .21 .20 .20 .19 .19 .10 .18 .18 .17 .17 .I6 .20 .20 .19 .I9 .18 .I9 .10 .I8 .I7 .17 2 (20) 4 (33) 6 (46) .22 .19 .I6 .21 .I8 .l6 .18 .I6 .I4 .16 .14 .13 .18 .16 .14 .I7 .I5 .13 2 4 6 0 10 I , I I SUBFLOOR NOW .20 .22 .I9 ( Tile (I) (1) 1 Tile (I) (1) .17 .13 .17 .15 .13 .13 .12 .I 1 .I6 .13 .I6 .13 .I5 .13 .15 .I2 .14 .12 --___ .14 .13 .I2 -c FRAME CONSTRUCTION CEILING / Plaster Plaster Plaster Plaster Board Board (7) j (3) / (5) (5) / (2) j (2) 1 (41 (3) (41 j (2) (2) None .21 .I5 .I2 Wood (2) 72” Wood .12 .I 1 .10 .15 .15 .14 .14 .14 .15.14 .13 .12 .12 .ll Suspended or Furred Not Furred FLOOR .21 .2l .18 .I6 .3l .20 .15 .I5 .12 .12 .I 1 (24) ‘K” Wood - - - (5) Y Wood Wnnrl 2” (7) --. -. ~-p----t %z” Wood (5) 4:: $#j;,;- (8) 2” Wood .14 .I2 .I2 .I4 .13 .I0 .12 .I I ---~___.I I .10 .14 .1 1 1958 ASHAE Equations Heot flow down, unconditioned space above: Heat Gain, Btu/hr = (Area. 54 ft) X (U value) X ( outdoor temp - inside temp - 5 FL Kitchen above: Heat Gain, Btu/hr = (Area, sq ft) X (U v&e) X (actual temp diff, or outdoor temp - inside temp + 15 F to 25 FL . .27 .I7 .14 .13 .1 1 Guide ’ l-75 CH/\I”I‘ER 5. HEAT AND WATER V.\I’OK FLOW THRU STRUCTURES TABLE Jl--TRANSMISSION COEFFICIENT U-WITH INSULATION & AIR SPACES SUMMER AND WINTER Btu/(hr) (sq ft) (deg F temp diff) Addition of Reflective Sheets to Air Space (Aluminum Foil Average Emisrivity u Value Before Adding Intul. Wall, Ceiling, Roof Floor - Direction Addition of Fibrous Insulation Thickness Winter and Summer Horizontal 4dded to one >r both sides (Inches) of Added to one or both sides Olle sheet in air Space TWO sheets in air lp.Xe Added to one or both sides One sheet in air IpCXe TWO sheets in air SPClC.3 .05 .05 .OS .05 .05 .05 35 .36 35 .34 .33 .32 .20 .20 .20 .I9 .19 .I9 .I4 .I4 .14 .I4 .14 .I3 .04 .04 .04 .04 .04 .31 .30 .29 .28 .27 .18 .18 .I8 .17 .17 .I3 .I3 .13 .13 .I2 .04 .04 .04 .04 .04 .26 .25 .24 .23 .22 .17 .16 .16 .I5 .15 .I2 .I2 .12 .11 .l 1 .04 .04 .04 .04 .04 .20 .19 .18 .16 .15 .I4 .13 .I3 .12 .l I .I0 .I0 .lO .09 .09 .04 .04 .04 .03 .03 .14 .13 .I2 .I0 .09 .l 1 .lO .09 .08 .07 .08 .08 .07 .07 .06 2 3 .60 .sa .56 .s4 .52 30 .08 .08 .08 .08 .08 .08 .38 .37 .36 .36 .35 .34 .34 .33 .32 .3l .30 .29 .I8 .18 .I8 .17 .I7 .17 .l 1 .l 1 .l 1 .l 1 .lO .I0 .12 .I2 .l 1 .I I .I I .l 1 .06 .06 .06 .06 .06 .06 .48 .46 .44 42 40 .I7 .17 .I7 .I6 .16 .I I .lO .lO .lO .lO 08 .08 .07 .07 .07 .33 .32 .31 .30 .29 .28 .28 .27 .26 .26 .I6 .16 .16 .I5 .I5 .I0 .lO .I0 .lO .lO .l I .I I .I 1 .I 1 .I0 .06 .06 .06 .06 .06 .3a .36 .34 .32 .30 .I6 .15 .15 .15 .I4 .I0 .lO .I0 .lO .09 .07 .07 .07 .07 .07 .28 .27 .26 .25 .23 .lO .I0 .I0 .lO .lO .06 .06 .06 .05 .05 .28 .26 .24 .22 .20 .14 .I3 .13 .12 .12 .09 .09’ .09 .08 .08 .07 .07 .07 .06 .b6 .22 .21 .20 .18 .17 .09 .09 .09 .08 .08 .05 .05 .05 .05 .05 .lB .16 .l 1 .lO .09 .08 .07 .08 .07 .07 .06 .06 .06 .06 .05 .05 .05 .15 .I4 .12 .I I .09 .08 .07 .07 .06 .06 .05 .05 .04 .04 .04 .14 .I2 .lO Insulation Added Air Space Added AIR SPACES INSULATION Dl<lOER .lO .09 .08 .08 .07 Reflective Sheets Added to One or Both Sides AIR SPACE REFLECTIVE SHEETS Winter UP TWO sheets in air *pWX .l 1 .l I .l 1 .ll .l 1 .I 1 .14 .12 .l 1 .I0 .08 = .05) Flow DOWll .I9 .19 .18 .I8 .18 .18 - Heat .07 .07 .06 .06 .05 : I Reflective Sheet in Air Space AIR SPACES REFLECTIVE SHEETS ‘Checked for summer conditions for up, down ond horizontal heat flow. Error from above values is less than 1 yo. 1958 ASHAE Reflective Sheets in Air Space AIR SPACES REFLECTIVE SHEETS , Guide PART I. LOAD ESTIMATING l-76 TABLE 32-TRANSMISSIONCOEFFICIENT U-FLAT ROOFS WITH ROOF-DECK INSULATION SUMMER AND WINTER Btu/(hr) (sq ft) (deg F temp diff) Addition U VALUE OF ROOF BEFORE ADDING ROOF DECK INSULATION 1% 350 .50 .40 33 .29 .26 .22 .21 .I9 24 _-L- _ --.2 1 .I9 .16 .12 .09 &J \ ~&p-.25 -20 .15 .lO of Roof-Deck Thickness (in.) .___.--. 1 1% Insulation __3 2 2% .17 .I6 .15 .I4 .14 .13 .I2 .12 .l 1 .lO .I0 .09 .18 .16 .15 .14 .13 .12 .I2 .I2 .l 1 .I0 .10 .09 .09 .09 .08 .13 .I 1 .08 .I 1 .09 .07 .I0 .08 .07 .09 .08 .06 .08 .07 .05 TABLE 33-TRANSMISSIONCOEFFICIENT U-WINDOWS, SKYLIGHTS, DOORS & GLASS BLOCK WALLS . Btu/(hr) (sq ft) (deg F temp diff) GLASS Vertical Single Air Space Thickness (in.) Horizontal Triple Double ‘A ,WWithout Sto m Windows 1.131‘ W i t h S t o r m - c&. mdows lm Glass ‘h Oh’ %-4 0.55 ‘A 0.53 1% 0.4 1 0.36 Single %-4 0.34 Summer 0.86 0.43 Glass Double (%“) Winter Summer Winter 1.40 0.64 0.50 0.70 DOORS Nominal Thickness of Wood (inches) u Exposed Door U With Storm Door 1 1% 1% 1% 0.69 0.59 0.52 0.51 0.35 0.32 0.30 0.30 2 0.46 0.38 0.33 1.05 0.28 0.25 0.23 0.43 2’/2 3 Glare (%” Herculite) HOLLOW GLASS BLOCK WALLS Description* U 5%x5%x3%” Thick-Nominal Size 6x6x4 (14) 7%/7%x3%” Thick-Nominal Size 8x8~4 (14) 11 %x1 1 %x3%” Thick-Nominal Size 12x12~4 (16) 7%x7%x3%” Thick with gloss fiber screen dividing the cavity (14) . 11 %x1 1 %x3%” Thick with glass fiber screen dividing the cavity (16) - 0.60 0.56 0.52 0 . 4 8 0.44 1958 ASHAE Equation: Heat Gain or Loss, Btu/hr = (Area, rq ft) X (U *Italicized numbers in parentheses indicate weight in lb per . value) x (outdoor temp sq ft. ,,_ - inside temp) Guide CHAI’TEK ,5. HEA~I‘ .\NI) CV.\~I‘I<K V,\I’OK I:I.OW .I‘HKIJ CALCULATION OF TRANSMISSION COEFFICIENT U For types of construction not listed in Tabtes 21 thru 33, calculate the U value as follows: 1. Determine the resistance of each component of a given structure and also the inside and outdoor air surface films from Table 34. 2. Add these resistances together, l-77 S’I‘I~IJ~:‘I’IJI~I‘S Example 6 - Calculation of U Value Given: A wall as per Fig. 27 R = rl + yI + T,~ + . . . . . r, FIG. 27 3. Take the reciprocal, U =k Basis of Table 34 - Thermal Resistance R, Building and Insulating Materials Find: Transmission Table 34 was extracted from the 1958 ASHAE Guide and the column “weight per sq ft” added. Solution: Refer to Table 34. &o of Table 34 lermql Resistance R, Building and Insulating Materials The thermal resistances for building materials are listed in two columns. One column lists the thermal resistance per inch thickness, based on conductivity, while the other column lists the thermal resistance for a given thickness or construction, based on conductance. I. 2. 3. 4. 5. coefficient - OUTDOOR W ALL in summer. Construction Outdoor air surface (7% mph wind) Stone facing, 2 in. (2 X .OS) Hollow clay tile, 8” Sand aggregate plaster, 2 in. (2 X .20) Inside air surface (still air) -.. Total Resistance U = + = & = 0.30 Rtu/(hr)(sq Resistance R 0.25 0.16 1.85 0.40 0.68 ft)(deg F) I’;\R’I‘ l-78 TABLE 34-THERMAL RESISTANCES R-BUILDING AND I . INSULATING LOAD ESTIM,\TING MATERIALS (deg F per Btu) / (hr) (sq ft) RESISTANCE R THICKNESS (in.) DESCRIPTION MATERIAL DENSITY (lb per C” f0 WEIGHT (lb per l I ft) Per Inch Thickness 1 -iF For Listed Thickness 1 e BUILDING MATERIALS BUILDING BOARD Boards, Panels, Sheathing, etc Asbestos-Cement Board Asbestos-Cement Board Gypsum or Plaster Board Gypsum or Plaster Board Plywood Plywood Plywood Plywood Plywood or Wood Panels W o o d F i b e r B o a r d , Lominoted Wood Wood Wood, Wood, Fiber, Fiber, Fir or Fir or 1% % ‘55 % % % % or H o m o g e n e o u s Hardboard Type Hardboard Type Pine Sheathing Pine ‘/I ?YJl 1% - - 0.25 - - 120 120 50 50 34 34 34 34 34 26 31 1.25 1.58 2.08 0.71 1.06 1.42 2.13 - 1.25 2.38 2.00 0.03 0.32 0.45 0.31 0.47 0.63 0.94 - 65 65 32 32 1.35 2.08 4.34 0.72 - 0.18 0.98 2.03 - - BUILDING PAPER Vapor Permeable Felt Vapor Seal, 2 Layers of Mopped 15 lb felt Vapor Seal, Plastic Film - - - 0.06 0.12 Negl WOODS Maple, Oak, and Similar Hardwoods Fir. Pine. and Similar Softwoods 4.5 32 - 0.9 1 1.25 - MASONRY UNITS Brick, Common Brick, Face Clay Tile, Hollow: I Cell Deep 1 Cell Deep 2 Cells Deep 2 Cells Deep 2 Cells Deep 3 Cells Deep Concrete Blecks, Three Oval Core Sand 8. Gravel Aggregate Cinder Aggregate Lightweight Aggregate ( E x p a n d e d S h a l e , C l a y , S l a t e or Slag; Pumice) Gypsum Partition Tile: 3”xl2”x30” solid 3”xl2”x30” 4-cell 4”xl2”x30” 3-cell L Stone, Lime or Sand . 4 4 120 130 40 43 3 4 6 a 10 I2 60 40 50 45 42 40 15 16 25 30 35 40 3 4 6 8 12 76 69 64 64 63 19 23 32 43 63 3 4 6 8 12 68 60 54 56 53 17 20 27 37 53 3 4 8 12 60 52 48 43 15 17 32 43 - 3 3 4 45 35 30 11 9 13 - 1.26 1.35 1.67 0.08 - 150 - - 30 .44 0.80 1.1 ? 1.52 1.85 2.22 2.50 0.40 0.71 0.91 1.11 1.28 0.86 1.1 I 1 so 1.72 1.89 1.27 1.50 2.00 2.27 . (:M.\I”I‘l~li 5. HL\‘I’ TABLE .\NI) LV.\ I I;li V.\I’OI< I:I.OW 34-THERMAL RESISTANCES ‘I’IlllIJ l-79 S’I‘IIIJ(:‘I’tIIII~S R-BUILDING AND INSULATING MATERIALS (Contd) (deg F per Btu) / (hr) (sq ft) RESISTANCE MATERIAL THICKNESS (in.) DESCRIPTION DENSITY (lb p e r C” ft) WEIGHT (lb p e r sq ft) Per Inch Thickness 1 -ii- BUILDING MASONRY MATERIALS Concretes PL .TERIN6 MATERIALS ROOFING SIDING MATERIALS (On Flat Surface) tement Mortar Gypsum-Fiber Concrete 12’/a’$& w o o d c h i p s MATERIALS , ((:ONT.) 116 Bi’Yx% gypsum, 51 - .ightweight Aggregates Including Expanded Shale, Clay or Slate Expanded Slag; Cinders Pumice; Perlite; Vermiculite Also, Cellular Concretes 120 100 80 60 40 30 20 S a n d & Gravel or Stone Aggregate (Oven Dried) S a n d & Gravel or S t o n e A g g r e g a t e ( N o t D r i e d ) stucco 140 140 116 Cement Plaster, Sand Aggregate Sand Aggregate Sand Aggregate 116 116 116 4.8 7.2 Gypsum Plaster: lightweight Aggregate Lightweight Aggregate lightweight Aggregate on Metal Lath Perlite Aggregate Sand Aggregate Sand Aggregate Sand Aggregate Sand Aggregate on Metal Lath Sand Aggregate on W o o d L a t h Vermiculite Aggregate 45 45 45 45 105 105 105 105 105 45 ~.BB 2.34 2.80 4.4 5.5 6.6 - Asbestos-Cement Shingles Asphalt Roll Roofing Asphalt Shingles Built-up Roofing Slate Sheet Metal Wood Shingles 120 70 70 70 201 40 - Shingles W o o d , 16”, 7%” exposure W o o d , D o u b l e , 16”. 12” exposure W o o d , P l u s lnsul Backer Board, ?&” Siding A s b e s t o s - C e m e n t , ‘A” lapped Asphalt Roll Siding A s p h a l t lnrul S i d i n g ‘An Board Wood Drop 1”xB” Wood: Bevel ‘/z”xB”;,lapped W o o d I Bevel, Y/x10 , lapped W o o d , P l y w o o d , 3/s”, lapped % ‘Ii Asphalt Tile Carpet and Fibrous Pad Carpet and Rubber Pad Ceramic Tile Cork Tile Cork Tile Felt, Flooring Floor Tile Linoleum Plywood Subfloor Rubber or Plastic Tile TfXraZIO Wood Subfloor Wood, Hardwood Finish 2.2 8.4 - - - - - - Structural Glass FLOORING MATERIALS - 120 25 25 - 1.25 0.26 - 80 34 110 140 32 45 0.83 1.77 1.15 11.7 2.08 2.81 R For Listed Thicknert 1 c c- - 0.20 0.60 0.19 0.28 0.40 0.59 0.86 1.11 1.43 0.11 0.08 0.20 0.20 - 0.67 0.1B 0.59 - 1 1- 610 &5 0.32 0.39 0.47 0.09 0.11 0.13 0.40 - Negl - 0.21 0.15 0.44 0.33 0.05 0.94 - 0.87 1.19 1.40 2.22 - 0.21 0.15 1.45 0.79 0.81 1.05 0.59 0.10 0.04 2.08 1.23 0.08 0.28 0.06 0.05 0.08 0.78 0.02 0.08 0.98 0.68 I l'/\R'I I. LoAD I:s'I‘IM.\'I‘INC; l-80 TABLE 34-THERMAL RESISTANCES R-BUILDING AND INSULATING MATERIALS (Contd) (deg F per Btu) / (hr) (sq ft) -I I I- RESISTANCE MATERIAL THICKNESS (in.) DESCRIPTION DENSITY (lb p e r cu ft) / WEIGHT (lb per ‘ 4 ffl Per Inch Thickness 1 k I INSULATING MATERIALS BLANKET AND BOARD AND BATT* SLABS LOOSE FILL ROOF INSULATION I L -7 AIR SPACES AIR FILM Still Air Cotton Fiber I.8 - 2 . 0 3.85 Mineral Wool, Fibrous Form Processed From Rock, Slag, or Gloss 1.5 _ 4.0 3.70 4.00 Wood Fiber Wood Fiber, Multi-layer Stitched Expanded 3.70 Glass Fiber 4.00 Wood or Cane Fiber Acoustical Tile Acoustical Tile Interior Finish (Tile, Lath, Plank) Interior Finish (Tile, Lath, Plonk) 22.4 22.4 15.0 Roof Deck Slab Sheathing (Impreg or Coated) Sheathing (Impreg or Coated) Sheathing (Impreg or Coated) % “h 0.62 20.0 20.0 20.0 0.83 9.0 6.5 _ 8.0 8.5 Macerated Paper or Pulp Products Wood Fiber: Redwood, Hemlock, or Fir Mineral Wool (Glass, Slag, or Rock) Sawdust or Shavings Vermiculite (Expanded) 2.5 2.0 2.0 8.0 1 .b2 22.0 _ 3.5 - 3.5 - 5.0 _ 15.0 7.0 All Types Preformed, for use above deck Approximately Approximately Approximately Approximately Approximately Approximatley AIR POSITION Horizontal Horizontal Horizontal Horizontal Horizontal Horizontal Horizontal Horizontal Horizontal Sloping 45’ Sloping 45’ Vertical Vertical HEAT FLOW Up (Winter) Up (Summer) Down (Winter) Down (Winter) Down (Winter) Down (Winter) Down (Summer) Down (Summer) Down (Summer) Up (Winter) Down (Summer) Horiz. (Winter) Horiz. (Summer) POSITION Horizontal Sloping 45’ Vertical Sloping 45” Horizontal HEAT FLOW UP UP Horizontal Down CiOW” 15 Mph Wind Any Position (For Winter) Any Direction 7% M p h W i n d Any Position (For Summer) Any Direction TI i ‘Ii 11: 2 2% - 15.0 Cellular Gloss Cork Board (Without Added Binder) Hog Hair (With Asphalt Binder) Plastic (Foamed) Wood Shredded (Cemented in Preformed Slabs) 2.86 - - 2.63 - I .3 I 2.50 3.70 3.00 3.45 I .82 - 3.57 3.33 3.33 2.22 2.08 3 3A . 4 74 - 4 9.4 1% 4 8 =/4 1 Y2 -4i =/4 - 4 =/4 - 4 =/4 . 4 =/4 - 4 - - - - - - - - .7 - 15.6 15.6 15.6 15.6 15.6 15.6 1.3 1.9 2.6 3.2 3.9 - - R For listed Thickness 1 c ~1.19 1.78 - * 1.32 2.06 l - 1.39 2.78 4.17 5.26 6.67 8.33 0.85 0.78 1.02 1.15 I .23 1.25 0.85 0.93 0.99 0.90 0.89 I 0.97 0.86 0.61 0.62 0.68 0.76 0.92 _ 0.17 k 0.25 ‘Includes paper backing and facing if any. In corer where the insulation forms a boundary (highly reflective) of an air space, refer to Table 31, page 75 :/ CHAI’TEK 5 . HEA’I. AND W,\~I‘tCK V/\I’OK I;I.OW HEAT LOSS THRU BASEMENT WALLS AND FLOORS BELOW THE GROUND LEVEL The fess through the floor is normally small and relatively constant year round because the ground temperature under the floor varies only a little throughout the year. The ground is a very good heat sink and can absorb or lose a large amount of heat without an appreciable change in temperature at about the 8 Et level. Above the 8 ft level, the ground temperature varies with the outdoor temperature, with the greatest variation at the surface and a decreasing variation down to the 8 ft depth. The heat loss thru a basement wall may be appreciable and it is difficult to calculate because the ground temperature varies with depth. Tables 35 thru 37 have been empirically calculated to simplify the evaluation of heat loss thru basement walls and Hoors. ‘he heat loss thru a slab floor is large around the perimeter and small in the center. This is because the ground temperature around the perimeter varies with the outdoor temperature, whereas the ground temperature in the middle remains relatively constant, as with basement floors. Basis of Tables 35 thru 37 -Heat Loss thru Masonry -1‘HKU 1-81 SI‘KUCI‘UK1SS Example 7 - Heat Loss in CI Basement Given: Basement - 100’ X 40’ X 9 Basement temp- 65 F tll), heated continuously Outdoor temp - 0” F tlb Grade line -G ft above basement floor Walls and floors - 12 in. concrete (80 lb/cu Solution: I. Heat loss above ground H = UA, (ta - to,) = 0.18 x (ZOO + 80) X 3 X (65 - 0) = 9828 Btu/hr 2. Heat loss ground. thru walls and and Walls in Use of Tables 35 thru 37 - Heat Loss thru Masonry Floors and Walls in Ground Yhe transmission coefficients listed in Table 35 .may be used for any thickness of uninsulated masonry floors where there is good contact between the floor and the ground. The perimeter factors listed in Table 36 are used for estimating heat loss thru basement walls and the outside strip of basement Hoors. This factor can be used only when the space is heated continuously. If there is only occasional heating, calculate the heat 10~s using the wall or floor transmission coefficients as listed in Tnbles 21 tllru 33 and the temperature difference between the basement and outdoor air or ground as listed in Table 37, The heat loss in a basement is determined by adding the heat transferred thru the floor, the walls and the outside strip of the Hoor and the portion of the wall above the ground level. strip of floor below = 19,100 Btu/hr 3. Heat loss thru floor H = UAi (th - to) = 0.05 x (100 X 40) X (65 - 55) Total Heat Loss where Ground Tables 35 thru 37 are based on empirical data. The perimeter factors listed in Table ?6 were developed by calculating the heat transmitted for each foot of wall to an 8 ft depth. The ground was assumed to decrease the transmission coefficient, thus adding resistance between the wall and the outdoor air. The transmission coefficients were then added to arrive at the perimeter factors. outside H = Lp 4 (b - toa) = (200 + 80) X 1.05 X (65 - 0) = 2 0 0 0 Btu/hr = 30,928 Btu/hr U = Heat transmission coefficient of wall above ground ( T a b l e 21) and floor ( T a b l e 3 5 ) i n Btu/(hr) (sq ft) (deg F) I A, = Area of wall above ground, sq ft A, = Entire floor area, Floors ft) Find: Heat loss from basement sq ft Lp = Perimeter of wall, ft Q = Perimeter factor (Table 36) tb = Basement dry-bulb temp, F tg = Ground temp, F, (Table 37) t oa = Outdoor design dry-bulb temp, F TABLE 35-TRANSMISSION COEFFICIENT UMASONRY FLOORS AND WALLS IN GROUND (Use only in conjunction with Table 36) Portion of Wall exceeding B feet below ground level .08 *Some additional floor loss is included in perimeter factor, see Table 36. Equations: H e a t l o s s t h r o u g h f l o o r , Btu/hr = (area of floor, sq X (U value) x (basement - ground temp). ftt) H e a t l o s s t h r o u g h w a l l b e l o w 8 f o o t l i n e , Btu/hr = (area of wall below 8 ft line, sq ft) X (U value) X (basement - ground temp). NOTE: The factors in Tables 35 and 36 may be used for ony thickness of uninsulated masonry wall or floor, but there must be a good contact (no air space which may connect to the outdoors) between the ground and the floor or wall. Where the ground is dry and sandy, or where there is cinder fill along wall or where the wall has a low heat transmission coefficient, the perimeter factor may be reduced slightly. l-82 I’AKT I. LOAD ESTIMATING TABLE 36-PERIMETER TRANSMISSION COEFFICIENTS PIPES IN WATER OR BRINE FACTORS FOR ESTIMATING HEAT LOSS THROUGH BASEMENT WALLS AND OUTSIDE STRIP OF BASEMENT FLOOR Heat transmission coefftcients for copper and steel pipes are listed in Tables 38 and 39. These coeflicicnts may be ~~scful in applications such as cold water or brine storage systems and ice skating rinks. (Use only in conjunction with Table 35) Distance of Floor From Ground Level Perimeter (91 2 Feet above Factor Basis of Tables 38 and 39 -Transmission Coefficients, Pipes in Water or Brine .90 .60 .75 .90 At ground level 2 below 4 Feet below 6 Feet below 8 Feet below Feet Table 35 is for ice coated pipes in water, based on a heat transfer film coefficient, inside the pipe, of 1 5 0 Btu/(hr)(sq Itinternal pipe surface)(deg F). Table 39 is Sor pipes in water or brine based on a heat transfer of 18 Btu/(hr)(sq Et external pipe surface) (deg F) in water, 14 Btu in brine. It is also based on a low rate of circulation on the outside of the pipe and 10 F to 15 F temperature difference between water or brine and refrigerant. High rates of circulation will increase the heat transfer rate. For special problems, consult heat transfer reference books. 1.05 I .20 Equation: Heat loss about perimeter, Btu/hr = (perimeter of wall, ft) x (perimeter factor) X (basement - outdoor temp). TABLE IR 37-GROUND TEMPERATURES ESTIMATING HEAT LOSS THROUGH BASEMENT FLOORS Outdoor Design Temp (F) - 30 Ground Temp (F) 1 40 - 20 - 10 / 45 1 50 0 1 55 TABLE 3B-TRANSMISSION +10 +20 1 60 / 65 COEFFICIENT U-ICE COATED PIPES IN WATER Btu/(hr) (lineal ft pipe) (deg F between 32 F db and refrig temp) . Inside film coefficient = 150 Btu/(hr) (sq Copper Pipe Size (Inches O.D.1 % % % 1% Copper Pipe With Ice Thickness (Inches) ‘vi 6.1 7.1 8.0 9.8 I ’ 1% 2 4.5 5.1 5.7 6.7 3.8 3.4 4.2 4.7 5.4 if 417 it) (deg F) Steel Pipe Size Nominal (Inches) Steel Pipe With Ice Thickness (Inches) % 1% % 7.2 8.7 1 1% 10.6 1 2: 7.2 8.6 13.0 1% 2 3 4.4 5.1 5.8 3.9 4.5 5.1 6.8 5.9 3.4 3.8 4.2 4.8 TABLE 39-TRANSMISSION COEFFICIENT U-PIPES IMMERSED IN WATER OR BRINE Btu/(hr) (lineal ft pipe) (deg F between 32 F db and refrig Outside water film coefficient= 1 B Btu/(hr) lsq ftt) Outside brine film coefficient= 14 Btu/(hr) (sq Water refrigerant temp= 10 F to 15 F (deg F) ft) (deg F) temp) . CH;\I”I‘lCK 5 . HE.\‘I’ .\NI) W\~l‘lili V.\I’OK I;l.OW ‘I’HKU WATER VAPOR FLOW THRU BUILDING STRUCTURES Water vapor [lows thru building structures; resulting in a latent load whenever a vapor pressure difference exists across a structure. The latent load from this source is usually insignificant in comfort applications and need be considcrcd only in low or high tlewpoint applications. Water vapor flows from high to lower vapor pressure at a rate determined by the permeability of the structure. This process is quite similar to heat flow, except that there is transfer of mass with water vapor flow. As heat flow can be reduced by adding insulation, vapor How can be reduced by vapor barriers. The vapor barrier may be paint (aluminum or asphalt), aluminum foil or galvanized iron. It sF--lid always be placed on the side of a structure h ng the higher vapor pressure, to prevent the water vapor from flowing up to the barrier and condensing within the wall. Basis of Table 40 -Water Vapor Transmission thru Various Materials The values for walls, floors, ceilings and partitions have been estimated from the source references listed in the bibliography. The resistance of a homogeneous material to water vapor transmission has been assumed to be directly proportional to the thickness, and it also has been assumed that there is no surface resistance to wa.ter vapor flow. The values for, permeability of miscellaneous materials are based on test results. NOTE: Some of the values for walls, roofs, etc., l-83 S’I’I~IJ(:‘I‘IJl~I:S have been increased by a safety factor because conclusive data is not available. Use of Table 40 -Water Vapor Transmission thru Various Materials T&le 40 is used to determine latent heat gain from water vapor tiransmission thru building structures in the high and low dewpoint applications where the air moisture content must be maintained. Example 8 - Wafer Vapor Transmission Given: iZ 40 ft X 40 ft X 8 ft Inhoratory on secontl floor requiring insicle design contlitions of 40 F tll), 50y0 rh, with the olltdoor design contlitions ;It 99 I: tll). 75 F WI,. The outdoor wall is 12 inch Ijrick with no wintlows. The partitions are metal lath and plaster on both sides of studs. Floor and ceiling are 4 inch concrete. Find: The latent heat gain from the water vapor transmission. Solution: Gr/lh at 95 F db, 75 F WI, = 99 (psych chart) Gr/lh at 40 F db, 5O70 rh = 18 (psych chart) Moisture content difference = 81 gr/lb Assume that the dewpoint in the areas surrounding the laboratory is uniform and equal to the outdoor dewpoint. Latent heat gain: 40 x 8 Outdoor wall = 100 X 81 X .04 (Table 40) = 10.4 Btu/hr 40 x 40 Floor and ceilings = 2 X - X 81 x .lO 100 = 259 Btu/hr 40 X 8 Partitions = 3 X - x 81 x 1.0 100 = 777 Btu/hr Total Latent Heat Gain = 1046.4 Btu/hr , I’Alil L 1-84 TABLE 40-WATER VAPOR TRANSMISSION I. LOi\D ESTIM;\‘1‘1N<; THRU VARIOUS MATERIALS PERMEANCE Btu/(hr) ( 1 0 0 sq ft) (gr/lb .diffl latent heat With 2 Coots Vapor-seal Paint on Smooth Inside Surface* With Aluminum Foil Mounted on One Side of Poper Cemented to Wallt .I2 .06 .04 .49 ,075 .046 .033 - .024 .020 .017 - .067 .034 .40 ,050 .029 - .02 1 .016 - Frame-with plaster interior finish -some with asphalt coated insulating board lath .79 .42 .I6 .I4 .029 .028 Tile-hollow cloy (face, glazed)-4 inches -hollow clay (common)-4 inches -hollow clay, 4 inch face and 4 inch common .013 .24 .012 ,012 .ll .Ol I .0091 .025 .0086 .lO .051 2.0 SO .40 .067 .040 .18 .14 .I3 .023 .019 .030 .028 .028 4.0 1 .o .I9 .17 .030 .029 .02 .02 1.5 .02 .02 .018 .018 .18 .018 .018 .012 .012 .29 .012 .012 .17 .027 No Vapor Seal Unless Noted Under Description DESCRIPTION OF MATERIAL OR CONSTRUCTION WALLS Brick- 4 inches - 8 inches - 1 2 inches -per inch of thickness Concrete- 6 inches - 12 inches -per inch of thickness I CEILINGS AND FLOORS Concrete-4 inches -8 inches Plaster on wood or metal lath on joist-no flooring Plaster on wood or metal lath on joist-flooring Plaster on wood or metal lath on joists-double flooring PARTITIONS Insulating Board % inch on both sides of studding Wood or metal lath and plaster on both sides of studding ROOFS Concrete-2 inches, plus 3 layer felt roofing -6 inches, plus 3 layer felt roofing Shingler, sheathing, rafters-plus plaster on wood Wood-l inch, plus 3 layer felt roofing -2 inches, plus 3 layer felt roofing or metal lath MISCELLANEOUS Air Space, still air 3% inch 1 inch Building Materials Masonite-l thickness,?% inch -5 thicknesses Plaster on wood lath -plus 2 coats aluminum point Plaster on gypsum lath -ditto plus primer and 2 coats lead and oil paint Plywood--% inch Douglas fir (3 ply) -ditto plus 2 coats asphalt paint -ditto plus 2 coats aluminum paint --% inch Douglas fir (5 ply) -ditto plus 2 coats asphalt paint -ditto plus 2 coats aluminum paint Wood-Pine .508 inch -ditto plus 2 coats aluminum paint -spruce, .508 inch Insulating Materials Corkboard, 1 inch thick Interior finish insulating board, l/i“ -ditto plus 2 coats water emulsion paint -ditto plus 2 coats varnish base paint -ditto plus 2 coots lead and oil paint -ditto plus wall linoleum 3.6 13.0 1.1 .32 1.1 1.95 .63 - .12 .13 .087 .I3 .27 -- .04 1 .12 .33 .20 .046 .63 5.0 3.0 .1 .17 .03 - 7.0 4.0 1.0 .06 <:FI.\1”1‘1~1< i. ill~.\‘l‘ .\NI) LV.\‘1‘1-I< V.\I’OI< I~I.OCV 1’111111 l-85 s’I’l1Il(:‘l’lJl<I:s TABLE 40-WATER VAPOR TRANSMISSION THRU VARIOUS MATERIALS (Contd) PERMEANCE Btu/(hr) (100 sq ft) (gr/lb diff) latent heat DESCRIPTION OF MATERIAL OR CONSTRUCTION No Vapor Seal Uliless Noted Under Description , With 2 Coats Vapor-real Paint on Smooth Inside Surfate* With Aluminum Foil Mounted on One Side of Paper Cemented to wallt MISCELLANEOUS Insulating Materials. cont. Insulating board lath -ditto plus %” plaster -ditto plus 1/2” plaster, sealer, rind flat coat of point Insulating board sheathing, %I” -ditto plus asphalt coating both sides Mineral wool (3% inches thick), unprotected Packaging materials Cellophane, moisture proof Glarsine (1 ply waxed or 3 ply plain) Kraft paper waked with parafin wax, 4.5 Ibs per 100 rq ft liofilm Paint Films 2 coots aluminum paint, estimated 2 coats asphalt point, estimated 2 coats lead and oil paint, estimated 2 co& water emulsion, estimated Papers Duplex or asphalt lominae (untreated) 30-30-30, 3.1 lb per 100 sq ft -ditto 30-60-30,4.2 lb per 100 sq ft Kraft pope+1 sheet -2 sheets -aluminum foil on one side of sheet --aluminum foil on both sides of sheet Sheathing paper Asphalt impregnated and coated, 7 lb per 100 sq ft Sloterr felt, 6 lb per 100 sq ft, 50% saturated with tar Roofing Felt, saturated and coated with ospholt 25 lb. per sq ft 50 lb. per sq ft Tin sheet with 4 holes l/(6 diameter Crack 12 inches long by ‘/Lo inches wide (approximated from above) .05 .05 .l 5.0 - .2 - .I - .6 - 8.0 I .I5 - .27 .051 - ,091 8.1 5.1 .016 .012 .02 - .lO 1.4 .015 .Ol 1 .17 5.2 / I *Pointed surfacer: Two coots of o good vapor seal paint on o smooth surface give o fair vapor barrier. More surface treatment is required on o rough surface than on o smooth surface. Doto indicates that either asphalt or aluminum paint ore good for vapor seals. tPluminum Foil on Paper: This material should also be applied over o smooth surface and joints lapped and sealed with asphalt. vapor barrier should always be placed on the side of the wall having the higher vapor pressure if condensation of moisture in wall is possible. Application: The heat gain due to water vapor transmission through walls moy be neglected for the normal oir conditioning or refrigeration job. This latent gain should be considered for air conditioning jobs where there is a great vapor pressure difference between the room and the outside, particularly when the dewpoint inside must be low. Note that moisture gain due to infiltration usually is of much greater magnitude than moisture transmission t,hrough building structures. Conversion Factors: To convert above table values tc: grain/(hr) (sq ft) (inch mercury vapor pressure difference), multiply by 9.8. grain/(hr) (rq ft) (pounds per sq inch vapor pressure difference), multiply by 20.0. TO convert Btu latent heat to grains, multiply by 7000/1060= 6.6. I’,\li’I‘ I. l,O,\D ESTII\l.\TINC l-86 CONDENSATION OF WATER VAPOR Whenever there is a difference of temperature and pressure of water vapor across a structure, conditions may tlevelop tllat lead to a condensation of moisture. This condensation occurs at the point of saturation temperature and pressure. As water vapor flows thru the structure, its temperature decreases and, if at my point it reaches the clewpoint or saturation temperature, condensation begins. r\s condensation occurs, the vapor pressure decreases, thereby lowering the dewpoint or saturation temperature until it corresponds to the actual temperature. The rate at which condensation occurs is determined by the rate at which heat is removed from the point of condensation. As the vapor continues to condense, latent heat of condensation is released, causing the dry-bulb temperature of the material to rise. To illustrate this, assume a frame wall with wood sheathing.and shingles on the outside, plasterboard on the inside and fibrous insulation between the two. Also, assume that the inside conditions are 75 F db and 50% rh and the outdoor conditions are 0°F db and 80% rh. Refer to Fig. 28. The temperature and vapor pressure gradient decreases approximately as shown by the solid and dashed lines until condensation begins (saturation point). At this point, the latent heat of condensation decreases the rate of temperature drop thru the insulation. This is approximately indicated by the dotted line. Another cause of concealed condensation may be evaporation of water from the ground or damp locations. This water vapor may condense on the under-, side of the floor joints (usually near the edges where FIG . it is coldest) or may llow up thru the outdoor side of the walls because of stack effect and/or vapor pressure dilfercnces. Concealed condensation may cause wood, iron and brickwork to deteriorate and insulation to lose its insulating value. These effects may be corrected by the following methods: 1. Provide unpor bar-?-ien on the high vapor pressure side. 2. In winter, ventilate the building to reduce the vapor pressure within. No great volume of air change is necessary, and normal infiltration alone is frequently all that is required. 3. In winter, ventilate the structure cavities to remove vapor that has entered. Outdoor air thru vents shielded from entrance of rain and insects may be used. . Condensation may also form on the surface of a building structure. Visible condensation occurs when the surface of any material is colder than the dewpoint temperature of the surrounding air. In winter, the condensation may collect on cold closet walls and attic roofs and is commonly observed as frost on window panes. Fig. 29 illustrates the condensation on a window with inside winter design conditions of 70 F db and 40% rh. Point A represents the room conditions; point B, the dewpoint temperature of the thin film of water vapor adjacent to the window surface; and point C, the point at which frost or ice appears on the window. Once the temperature drops below the dewpoint, the vapor pressure at the window surface is also reduced, thereby establishing a gradient of vapor pressure from the room air to the window surface. This gradient operates, in conjunction with the convec- 28 - CONDENSATION WITHIN FRAME WALL I;rc. 29 - C ONDENSATION tive action within the room, to move water Vapor continuously to the window surface to be condensed, as long as the concentration of the water vapor is m ained in a space. Visible condensation is objectionable as it causes staining of surfaces, dripping on machinery and furnishings, and damage to materials in process of manufacture. Condensation of this type may be corrected by the following methods: 1. Increase the thermal resistance of walls, roofs and Hoors by adding kulation with vapor barriers to prevent condensation within the structures. 2. Increase the thermal resistance of glass by installing two or three panes with air space(s) between. In extreme cases, controlled heat, electric or other, may be applied between the glass of double glazed windows. 3. Maintain a room dewpoint lower than the lowest expected surface temperature in the room. 4. Decrease surface resistance by increasing the velocity of air passing over the surface. Decreas,ing the surface resistance increases the window surface temperature and brings it closer to the room dry-bulb temperature. Basis of Chart 2 -Maximum Room Condensation RH; No Wall, Roof or Glass Chnrt 2 has been calculated from the equation u s e d to determine the maximum room dewpoint temperature that can exist with condensation. tap = trn, - U(LH - b,,) fi where t,, = dewpoint temp of room air, F db t,.,n = room temp, F U = transmission coefficient, Btu/(hr)(sq ft)(deg F) t,, = outdoor temp, F ON WINL~OW S URFACE fi = inside air film or surEace Btu/(hr)(sq ft)(deg F) conductance, Chad -3 is based upon a room dry-bulb temperature of ‘i0 F db and an inside film conductance of I.46 Btu/(hr)(sq ft)(deg F). Use of Chart 2 -Maximum Room Condensation RH; No Wall, Roof or Gloss Chart 2 gives a rapid means of determining the maximum room relative humidity which can be maintained and yet avoid condensation with a 70 F db room. Example ? - Moisture Condensation Given: 12 in. stone wall with ?h in. sana aggregate plaster Room temp - 70 F db Outdoor temp - 0” F db Find: hiaxirnum room rh without wall condensation. Solution: Transmission coefficient U = 0.52 Btu/(hr)(sq ft)(degF) (Table 21, page 66) Maximum room rh = 40.05y0, (Chart 2) Corrections in room relative humidity for room temperatures other than 70 F clb are listed in the table under Ciuzrt 2. Values other than those listed may be interpolated. Example J0 - Moisture Condensation Given: Same as Exorn$e 9, except room temp is 75 F db Find: Maximum room rh without wall condensation Solution: Transmission coefficient U = 0.52 Btrl/(hr)(sq ft)(deg F) (E.uanaple 9) Xfaximnm room rh for 70 F tlh room temp = 40.05rr/, (Example 9) Rh correction for room temp of 75 F tll) with U factor of 0.52 = --1.57rr/, (I,ottom CIlarl 3). Maximum room rh = 40.05 - 1.57 = 38.48% or 38.5% 1-88 PAI<* CHART 2-MAXIMUM ROOM RELATIVE HUMIDITY WITHOUT I . LOr\D ES’I‘IM.\~L‘INC; CONDENSATION NO WALL, ROOF OR GLASS CONDENSATION . - 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 WALL, ROOF OR GLASS TRANSMISSION COEFFICIENT “U” BTU/(HRl(SO FT)(DEG F) CORRECTION IN ROOM RH I.1 (70, For Wall, Roof or Glass Transmission Coefficient U u = .65 Outdoor Temp IF db) u = .35 Room Temp (F db) 60 -30 -20 - 1 0 0 10 20 30 40 +2.0 +3.5 +5.0 f7.0 +9.0 +12.0 - 2.0 - 2.5 - 3.5 - 4.0 -7.5 - 9.5 +3.5 f4.0 f5.0 +6.5 f8.5 f9.5 - 3.0 -4.0 - 4.5 - 5.0 - 6.0 -7.5 +2.5yo f3.0 f3.0 i3.5 f4.0 f4.5 f5.0 f6.0 80 - 2.0 y . -2.0 - 2.0 - 2.5 - 3.0 - 3.5 - 4.0 - 4.5 1-89 CHAPTER 6. INFILTRATION AND VENTILATION The data in this chapter is based on ASHAE tests evaluating the infiltration and ventilation quantities OE outdoor air. These outdoor air quantities normally have a different heat content than the air within the conditioned space and, therefore, impose a load on the air conditioning equipment. In the case of infiltration, the load manifests itself directly within the conditioned space. The ventilation air, taken thru the conditioning apparatus, imposes a load both on the space thru apparatus bypass effect, and directly on the conditioning 5; :pment. ,. _ he data in this chapter is based on ASHAE tests and years of practical experience. INFILTRATION Infiltration of air and particularly moisture into conditioned space is frequently a source of sizable teat gain or loss. The quantity of infiltration air aries according to tightness of doors and windows, lorosity of the building shell, height of the buildrg, stairwells, elevators, direction and velocity of rind, and the amount of ventilation and exhaust ir. Many of these cannot be accurately evaluated ad must be based on the judgment of the estimator. Generally, infiltration may be caused by wind :locity, or stack effort, or both: . Wind Velocity - The wind velocity builds up a pressure on the windward side of the building and a slight vacuum on the leeward side. The . outdoor pressure build-up causes air to infiltrate thru crevices in the construction and cracks around the windows and doors. This, in turn, causes a slight build-up of pressure inside the building, resulting in an equal amount of exfiltration on the leeward side. !. Difference in Density or Stack Effect - The variations in temperatures and humidities produce differences in density of air between inside and outside of the building. In tall buildings this density difference causes summer and winter infiltration and exfiltration as follows: Summer - Infiltration at the top and exfiltration at the bottom. Winter-Infiltration at the bottom and exfiltration at the top. This opposite’ direction how balances at some neutral point near the mid-height of the building. Air flow thru the building openings increases proportionately between the neutral point and the top and the neutral point and bottom of the building. The infiltration from stack effect is greatly influenced by the height of the building and the presence of open stairways and elevators. The combined infiltration from wind velocity and stack effect is proportional to the square root of the sum of the heads acting on it, The increased air infiltration flow caused by stack effect is evaluated< by converting the stack effect force to an equivalent wind velocity, and tp calculating the flow from the wind velocity data in the tables. I In buildings over 100 ft tall, the equivalent wind velocity may be calculated from, the following formula, assuming a temperature difference of 70 F db (winter) and a neutral point at the mid-height of the building: Ye = v V” - 1.75a v, = v vz + 1.75b where (for upper bldgs (for lower bldgs - section of winter) section of winter) _ tall (1) tall (2) V, = equivalent wind velocity, mph V = wind velocity normally calculated for location, mph a = distance window is above midheight, ft b = distance window is below midheight, ft NOTE: The total crackage is considered when calculating infiltration from stack effect. INFILTRATION THRU WINDOWS AND DOORS, SUMMER Infiltration during the summer is caused primarily by the wind velocity creating a pressure on the windward side. Stack effect is not normally a significant factor because the density difference is slight, (0.073 lb/cu ft at 7sF db, 50% rh and 0.070 lb/cu ft at 95 F db, 75 F wb). This small stack effect in tall buildings (over 100 Et) causes air to flow in the top and out the bottom. Therefore, the air infiltrating in the top of the building, because of the wind i l-90 l’,\l<T L I . LOAI) EST1 ICI;\-I‘ING Use of Table 41 - Infiltration thru Windows and Doors, Summer pressure, tends to flow down thru the building and out the doors on the street level, thereby olfsetting some of the infiltration thru them. In low buildings, air infiltrates thru open doors on the windward side unless sufficient outdoor air is introduced thru the air conditioning equipment to offset it; refer to “0)llsetting Infiltrntion zuith Out- The data in T(lOle Jf is used to determine the infiltration thru windows and doors on the windward side with the wind blowing directly at them. When the wind direction is oblique to the windows or doors, multiply the values in Tables ?Ila, b, c, d, by 0.60 and apply to total areas. For specific locations, ntl.just the values in Table 41 to the design wind velocity: refer to Table 1, Page 10. door Air.” With doors on opposite bills, the infiltration can be considerable if the two are open at the same time. Basis of Table 41 - infiltration thru Windows and Doors, Summer During the summer, infiltration is calculated for the windward side(s) only, because stack effect is small and, therefore, causes the infiltration air to flow in a downward direction in tall buiIdings (over 100 Et). Some of the air infiltrating thru the windows will exfiltrate thru the windows on the leeward side(s), while the remaining infiltration air flows out the doors, thus offsetting some of the infiltration thru the doors. To determine the net infiltration thru the doors, determine the infiltration thru the windows on the windward side, multiply this by .80, and subtract from the door infiltration. For low builclings the door infiltration on the windward side shoulcl be included in the estimate. The data in Tables -lla, b and c is based on a wind velocity of 7.5 mph blowing directly at the window or door, a n d o n o&r_v_eLcrack widths around typical windows and doors. This data is ‘erivecl from Table 4-/ which lists infiltration thru cracks around windows and doors as established by ASHA4E tests. Table 4Id shows values to be used for doors on opposite walls for various percentages of time that each door is open. The data in Table ille is based on actual tests of typical applications. . TABLE 41 -INFILTRATION THRU WINDOWS AND DOORS-SUMMER* 7.5 mph T A B L E 4la-DOUBLE H U N G WINDOWS1 Wind Velocity? ’ C F M P E R SQ F T S A S H A R E A No W-Strip Average iABLE Wood Poorly Fitted Metal Sash Sash Wood 4lb-CASEMENT L a r g e - 5 4 " x9 6 " Small--30” x 7 2 " DESCRIPTION Sash W-Strip Storm Sash No W-Strip W-Strip .27 .I4 .25 .43 .26 .22 1.20 37 .60 .76 .I7 .24 .80 .35 .40 Sl .22 Storm Sash 36 T Y P E WINDOWS$ CFM PER DESCRIPTION 0% 25% 33% .72 - Residential .33 - Heavy Projected - - .27 .50 Rolled Section-Steel Industrial Pivoted 40% Hollow Metal-Vertically Pivoted .39 - .28 - - - - .23 .99 .82 Openable 50% 45% Sash Architectural Proiected SO F T S A S H A R E A Percent / - ROLLED SECTlON STEEL s*w WINDOWS R~ptESENTr,T,“E TYPES OF WINDOWS WIEWED F R O M O”TS,OEI Area - - .55 .49 .74 - 66Yo 60% HOLLOY METAL WlNDOW 100% - 2.6 .32 .39 - 1.2 - 2.2 1.45 - 75% . .63 . (:I-l,\I”I‘I-R 6. INl~II.‘I‘I~.\ I‘ION .\NI) l-91 Vl~N’I‘II..\‘I‘ION TABLE 41 -INFILTRATION THRU WINDOWS AND DOORS-SUMMER* (Contd) 7.5 mph Wind Velocityt TABLE 41c-DOORS O N O N E O R A D J A C E N T W A L L S , F O R C O R N E R E N T R A N C E S C F M P E R SQ F T A R E A * * DESCRIPTION No Use R e v o l v i n g D o o r s - N o r m a l Operation Crock Use No Open Vestibule Vestibule .8 5.2 - - - 1,200 4.5 10.0 700 500 1 .o 6.5 700 500 Panels Open Glass Door-%” Average CFM Standing Wood Door (3’ x 7’) 900 Small Factory Door Garage 8 Shipping Room Door Ramp Garage Door TABLE 4ld-SWINGING D O O R S ON OPPOSITE CFM PER PAIR OF DOORS y. Time 2nd Door is Open % Time 1st Door is Open . 10 25 50 75 10 100 250 500 750 1,000 25 250 625 1250 1875 2,500 50 75 500 750 1250 1875 2500 3750 5,000 3750 5625 1000 2500 5000 7500 100 I WALLS I 100 7,500 / 10,000 I I ! I T A B L E 41e-DOORS I APPLICATION CFM PER PERSON IN ROOM PER DOOR 72“ Revolving No I Bank Barber Shop Candy and Soda Cigar Store Drug (Small) Store Hospital Swinging Door i Vestibule 6.5 8.0 4.0 5.0 6.0 3.8 5.5 7.0 5.3 20.0 30.6 22.5 6.5 2.0 8.0 6.0 2.5 1.9 5.5 7.0 5.3 1 Department Store Dress Shop 1.. 36” Door Room - 3.5 2.6 Lunch Room 4.0 5.0 3.8 Men’s Restaurant 2.7 2.0 3.7 2.5 2.8 1.9 Shoe 2.7 3.5 2.6 Shop Store *All values in Table 41 are bared on the wind blowing directly at the window or door. When the wind direction is oblique to the window or door, multiply the above values by 0.60 and use the total window and door oreo on the windward ride(s). tEased on o wind velocity of 7.5 mph. For design wind velocities different from the base, multiply the above values by the ratio of velocities. $lncludes frame leakage where applicable. **Vestibules may decrease the infiltration os much os 300/, reducing infiltration. Example 7 - lnfi/tration when the door usage is light. When door uroge is heavy, the vestibule is of little in Tall Buildings, Summer Given: A 20-story building in New York City oriented true north. Building is 100 ft long and 100 ft wide with a floor-to-floor height of 12 C t . rVal1 area is 5O(r, residential casement windows having SO(z> fixed sash. There are ten 7 ft x 3 ft swinging glass doors on the street level facing south. value for Find: Infiltration into the building thru doors and windows, disregarding outside air thru the equipment and the exhaust air quantity. Solution: The prevailing wind in New York City during the summer is south, 13 mph (Table 1, page IO). Correction l‘he infiltration thru revolving doors is caused by displacement of the air in the door quadrants, is almost intlcpendent of wind velocity and, therefore, cannot be offset by outdoor air. to Table 1 values for wind velocity = 1317.5 = 1 . 7 3 Glass arm on south side = 20 x I2 x 100 x .50 = 12,000 sq ft Infiltration thru windows Basis of Table 42 - Offsetting Swinging Door Infiltration with Outdoor Air, Summer = 12,000 x .49 x 1.73 = 10,200 cfm (Table Jib) Infltration thru doors = IO x 7 X 3 X 10 X 1.73 = 3640 cfm (Table 41~) Since this building is over 100 ft tall, net infiltration thru doors = 3640 - (10,200 X .80) = - 4520 cfm. Therefore, there is no infiltration thru the doors on the street level ou tlesijin days, only exfiltration. OFFSETTING SUMMER INFILTRATION WITH OUTDOOR AIR, Completely offsettin g infiltration by the introduction of outdoor air thru the air conditioning apparatus is normally uneconomical except in buildings vith few windows and doors. The outdoor air SO introduced must develop a pressure equal to the wind velocity to offset infiltration. This pressure causes exfiltration thru the leeward walls at a rate equal to wind velocity. Therefore, in a four sided building with equal crack areas on each side and the wind blowing against one side, the amount of outdoor air introduced thru the apparatus must be a little more than three times the amount that infiltrates. Where the wind is blowing against two sides, the outdoor air must be a little more than equal to that which infiltrates. Offsetting swinging door infiltration is not quite as difficult because air takes the path of least resistance, normally an open door. Most of the outdoor air introduced thru the apparatus flows out the door when it is opened. Also, in tall buildings the window infiltration tends to flow out the door. Some of the outdoor air introduced thru the apparatus exfiltrates thru the cracks around the windows and in the construction on the leeward side. The outdoor air values have been increased by this amount for typical application as a result of experience. Use of Table 42 - Offsetting Swinging Door Infiltration with Outdoor Air, Summer Table 42 is used to determine the amount of out- . door air thru air conditioning apparatus required to offset infiltration thru swinging doors. Example 2 - Offsetting Swinging Door Infiltration Given: A restaurant with 3000 cfm outdoor air being introduced thru the air conditioning apparatus. Exhaust fans in the kitchen remove 2000 cfm. Two 7 ft x 3 ft glass swinging doors face the prevailing wind direction. At peak load conditions, there are 300 people in the restaurant. . Find: The net infiltration thru the outside doors. Solution: ‘. Infiltration thru doors = 300 X 2.5 = 750 cfm (Table 41e) Net outdoor air = 3000 - 2000 = 1000 cfm Only 975 cfm of outdoor air is required to offset 750 cfm of door infiltration (Table 42). Therefore, there will be no net infiltration thru the outside doors unless there are windows on the leeward side. If there are windows in the building, calculate as outlined in Example 1. TABLE 42-OFFSETTING SWINGING DOOR INFILTRATION WITH OUTDOOR AIR-SUMMER Net Outdoor Air* (cfm) Door Infiltration (cfm) Net Outdoor Air* (cfm) Door Infiltration (cfm) 140 100 1370 1100 270 200 300 400 500 1480 1560 1670 1760 1200 600 700 800 900 1000 1890 2070 2250 2450 2656 1600 1800 2000 2200 2400 410 530 660 790 920 1030 1150 1260 *Net outdoor air is equal to the outdoor air qwantity introduced thru the apparatus minus the exhaust air quantity. 1300 1400 1500 l-93 CHAPTER r,. INFlLTR,\~I’ION AND VENTIL.\TION INFILTRATIQN THRU WINDOWS AND DOORS, WINTER Infiltration thru windows and doors during the winter is caused by the wind velocity and also stack effect. The temperature clifferenccs during the winter arc consideral~ly greater than in summer and, therefore, the density difference is greater; at 75 F db nncl SO% rh, tlensity is ,073s; at 0°F db, 40% t-h, density is .0865. Stack effect causes air to flow in at the bottom and out at the top, and in many cases requires spot heating at the doors on the street level to maintain conditions. In applications where there is considerable infiltration on the street level, much of the infiltration thru the windows in the upper levels will be offset. Basis of Table 43 - Infiltration thru Windows and Doors, Winter The data in Table 43 is based on a wind velocity 15 mph blowing directly at the window or door and on observed crack widths around typical windows and doors. The infiltration thru these cracks is calculated from Table 44 which is based on ASHAE tests. Use of Table 43 - Infiltration thru Windows and Dbors, Winter Table 43 is used to determine the infiltration of air thru windows and doors on the windward side during the winter. The stack effect in tall buildings increases the infiltration thru the doors and windows on the lower levels and decreases it on the upper levels. Therefore, whenever the door infiltration is increased, the infiltration thru the upper levels must be decreased by 80% of the net increase in door infiltration. The infiltration from stack effect on the leeward sides of the building is determined by using the difference between the equivalent velocity (V,) -d the actual velocity (V) as outlined in Example 3. .le data in Table 43 is based on the wind blowing directly at the windows and doors. When the wind direction is oblique to the windows and doors, multiply the values by 0.60 and use the total window and door area on the windward sides. Example 3 - Infiltration in Tall Buildings, Winter Given: The I>uilcling descrilted in Example 1. Find: The infiltration thru the doors and windows. Solution: The prevailing wind in New York City during the winter is NW at 16.8 mph (Table I, page 10) Correction on Table 43 for wind velocity is lf.H/15 Since the wind is coming from the Northwest, the on the north and west sides will allow infiltration wind is only 603, cffcctive. Correction for wind is 6. = 1.12. crackage Ijut the direction Since this huiltling is over 100 Et tall, stack effect causes inliltration on all sides at the lower Icvels and exfiltration at the upper levels. The total infiltration on the windward sitlcs remains the same hecause the increase at the hottom is exactly equal to the decrease at the top. (For’a floor-hyfloor analysis, use cquivalcnt wind velocity formulas.) Inliltration thru windows on the windward sides of the lower levels = 12,000 X 2 X 1.12 X .G X .98 = 15,810 cfm. The total inliltration thru the windows on the leeward sides of the huiltling is equal to the difference hetwcen the equivalent velocity at the first floor and the design velocity at the midpoint of the building. ve = v-” + 1.751, II = 22.2 mph Ve - V = 22.2 - 16.8 = 5.4 mph Total (upper = = infiltration thru windows in lower half half is exfiltration) on leeward side 12,000 x 2 x IA x (5.4/15) x lh x .98 2160 cfm (Table 43) of building / NOTE: This is the total infiltration thru the windows on the leeward side. A floor-by-floor analysis should be made to balance the system to maintain proper conditions on each floor. The infiltration thru the doors on the street level (on leeward side) = 10 x 7 x 3 x (5.4/15) x 30 =2310 cfm (Table 43c, average use, 1 and 2 story building). Example 4 - Offsefting Infiltration with Outdoor Air Any outdoor air mechanically introduced into the building offsets some of the infiltration. In Example 3 all of the outdoor air is effective in reducing the window infiltration. Infiltration is reduced on two windward sides, and the air introduced thru the apparatus exfiltrates thru the other two sides. \ Given: The building described in Example I with .25 cfm/sq ft supplied thru the apparatus and 40,000 cfm being exhausted from the building. Find: The net infiltration into this building. Solution: Net outdoor air = (.25 X 10,000 X 20) - 40,000 = 10,000 cfm Net infiltration thru windows = 15,800 + 2160 - 10,000 = 7970 cfm Net infiltration thru doors = 2310 cfm (Example 3) Net infiltration into building = 7970 + 2310 = 10,280 cfm l-94 PART I. LOAD ESTIM,\TING TABLE 43-INFILTRATldN THRU WINDOWS AND DOORS-WINTER* 15 mph Wind Velocityt T A B L E 43a-DOUBLE H U N G W I N D O W S O N W I N D W A R D SIDE$ CFM PER SO FT AREA Small-30” x 72” DESCRIPTION Average Poorly Wood Fitted No Sash W-Strip Wood Sash TABLE denotes W-Strip .52 .42 35 Metol Sash NOTE: W-Strip L a r g e - 5 4 “ x 96” Storm Sash 2.4 .74 1.60 .69 No 1.2 30 W-Strip W-Strip Storm Sash .53 .33 .26 1.52 .47 .74 1.01 .44 .50 weatherstrip. 43b-CASEMENT T Y P E WINDO W S O N WINDW A R D SIDEI C F M P E R SQ FT AREA . Percent DESCRIPTION 25% 0% Roiled Section-Steel .65 - Architectural Projected Residential Hollow 4oYo 450/o 50% Area 60% 66Yo 7% ~w% 5.2 Sarh Industrial Pivoted Heavy 33Yo Ventilated Projected Metal-Vertically Pivoted 1.44 - 1.98 - .78 - - - - - .45 .98 - - - 1.64 - - - - - .56 - .54 1.19 - 1.1 - 2.9 - I.48 - - - - - 1.26 .63 .7a - - 4.3 2.4 . T A B L E 43c-DOORS O N O N E O R A D J A C E N T W I N D W A R D SIDES1 CFM PER SQ FT AREA** Average Use DESCRIPTION Infrequent Use I Revolving Door Glass D o o r - f % / Crack) Wood Door 3’ Small Garage B Shipping Room Door Romp Garage Door 10.5 30.0 2.0 I / I 1.6 9.0 x 7’ Factory Door 1.5 4.0 4.0 Tall Building (ft) 182 Story Bldg. 13.0 ! 13.0 50 1 12.6 36.0 I 14.2 40.5 15.5 I 100 17.3 49.5 17.5 I 200 21.5 I 9.0 13.5 *All values in Table 43 are based on the wind blowing directly at the window or door. When the prevailing wind direction is oblique to the window or doors, multiply the above values by 0.60 and use the total window and door oreo on the windward side(s). tBased on a wind velocity of 15 mph. For design wind velpcities different from the base, multiply the table values by the ratio of velocities. $Stock effect in toll buildings may also cause infiltration on the leeward ride. To evaluate this, determine the equivalent velocity (Ve) and subtract the design velocity (V). The equivalent velocity ir: VC = \I Vz- 1.750 (upper section) Ve=$ Vi-l- 1.75b (lower section) Where a and b are the distances above and below the mid-height of the building, respectively, in ft. Multiply the table valuer by the ratio (V.,- V)/15 for doors and one half of the windows on the leeward side of the building. (Use valuer under “1 and 2 Story Bldgt” for doors on leeward side of toll buildings.) **Doors on opposite sides increase the above values 25y* Vestibules may decrease the infiltration 03 much (IS 3Oyo when door uroge is light. If door usage is heavy, the vestibule is of little value in reducing infiltration. Heat added to the vestibule will help maintain room temperature near the door. (:H,\l’~I‘ICK fi. INI~Il.~I‘K,\‘I‘ION INFILTRATION .\NI) l-95 VEN’I‘II.,\‘I‘ION Use of Table 44 - CRACK METHOD (Summer or Winter) - Infiltration thru Windows and Doors, Crack Method The crack method ol evaluating infiltration is more accurate than the arca methods. It is difficult Tnble # is used to determine the infiltration thru the doors and windows listed. This table does not take into account winter stack effect which must be evaluated separately, using the equivalent wind velocity formulas fireviously presented. to establish the exact crack dimensions but, in certain close tolerance applications, it may be necessary to evaluate the load accurately. The crack method is applicable both summer and winter. Infiltration thru Windows, Crack Method Example 5 Basis of Table 44 - Infiltration thru Windows and Doors, Crack Method Given: A 4 ft x 7 ft rcsitlcntial The data on windows in Table 44 are based on ASHAE tests. These test results have been reduced 20% because, as infiltration occurs on one side, a certain amount of pressure builds up in the building, thereby reducing the infiltration. The data on glass and factory doors has been calculated from observed typical crack widths. Fintl: The inliltration casement wintlow facing south. thru this window. Solution: Assume the crack widths are measured as follows: Wintlow frame - none, well sealed Wintlow openable area - l/32 in. crack; length. 20 ft Assume the wintl velocity is 30 mph due south. Infiltration thru winclow = 20 X 2.1 = 42 cfm (Table 44) 4 TABLE 44--INFILTRATION THRU WINDOWS AND DOORS-CRACK METHOD-SUMMER-WINTER* I TABLE 44o-DOUBLE HUNG WINDOWS-UNLOCKED ON WINDWARD SIDE CFM PER LINEAR FOOT OF CRACK 3 Wind TYPE OF DOUBLE HUNG WINDOW 5 No W- WStrip Strip Wood Sash Average Window Poorly Fitted Window Poorly Fitted--with Storm Sash Mekd S a s h Tba1.E 44b-CASEMENT .12 .45 .23 .33 Velocity--Mph 15 10 No W Strip WStrip .07 .35 .lO 1.15 .05 .lO .57 .7a .22 .32 .16 .32 20 No W- W Strip Strip .65 1 .a5 .93 1.23 25 No W- W Strip Strip .40 .57 .29 .53 .98 2.60 1.30 1.73 30 No W- W Strip Strip .I50 .85 .43 77 1.33 3.30 1.60 2.3 .82 1.18 .59 1.00 No W- W Strip Strip 1.73 4.20 2.10 2.8 1.05 1.53 .76 1.27 TYPE WINDOWS ON WINDWARD SIDE CFM PER LINEAR FOOT OF CRACK TYPE OF CASEMENT WINDOW AND TYPICAL CRACK SIZE Wind Velocity-Mph 5 10 15 20 25 30 ‘b’s” crock ‘h” crock ‘A” crack .a7 .25 .33 1.80 .60 .87 2.9 1.03 4.1 1.43 1.47 1.93 5.1 1.06 2.5 6.2 2.3 3.0 Residential Casement R e s i d e n t i a l Casement ‘!&“ crack ‘h” crack .lO .23 .30 .53 .55 .07 1.27 1 .oo 1.67 1.23 2.10 Heavy Casement Section Projected Heavy Casement Section Projected ‘%A” crack ‘%I” crock .05 .13 .17 .40 .30 .63 .43 .90 .58 1.20 30 1.53 1.46 2.40 3.10 3.70 4.00 R o l l e d Section-Steel Sash Industrial Pivoted Architectural Projected Architectural Projected Hollow Metal-Vertically Pivoted . .50 ‘Infiltration caused by stack effect must be calculated separately during the winter. *No allowance has been made for usage. See Table 43 for infiltration due to usage. \ .7a TABLE 44-INFILTRATIONTHRU WINDOWS AND DOORS-CRACK METHOD-SUMMER-WINTER* (Contd) TABLE 44c-DOORSt ON WINDWARD SIDE CFM PER LINEAR FOOT OF CRACK TYPE OF DOOR Wind Velocity- mph 5 10 15 20 25 30 3.2 4.8 6.4 6.4 1go 13.0 9.6 14.0 19.0 13.0 20.0 26.0 16.0 24.0 26.0 19.0 29.0 38.0 .45 .90 .90 .60 1.2 2.3 .90 1.8 3.7 1.3 2.4 5.2 1.7 3.3 6.6 2.1 4.2 8.4 6.4 9.6 13.0 16.0 19.0 G l o s s Door-Herculite Good Installation ‘h “ crack Average Installation ?&‘I crock Poor Installation l/i” crock Ordinary Wood or Metal Well Fitted-W-Strip Well Fitted-No W-Strip Poorly Fitted-No W-Strip Factory Door ‘/‘a” crock 3.2 / . VENTILATION VENTILATION STANDARDS The introduction of outdoor air for ventilation of conditioned sp?ces is necessary to dilute the odors given off by people, smoking and other internal air contaminants. The amount of ventilation required varies primarily with the total number of people, the ceiling height and the number of people smoking. People give off body odors which require a minimum of 5 cfm pel person for satisfactory dilution. J an and one half cfm per person is recommended. This is based on a population density of 5J to 75 J s-r person and a typical ceiling height of 8 ft. With greater population densities, the ventilation quantity should be increased. When people smoke, the additional odors given off by cigarettes or cigars .equire a minimum of 15 to 25 cfm per person. In special gathering rooms with heavy smoking, 30 to 50 cfm per person is recommended. Basis of Table 45 -Ventilation Standards The data in Table Ji is based on test observation of the clean outdoor air required to maintain satisfactory odor levels with people smoking and not smoking. These test results were then extrapolated for typical concentrations of people, both smoking and not smoking, for the applications listed. Use of Table 45 -Ventilation Standards Table 1’5 is used to determine the minimum and recommended ventilation air quantity for the listed a applications. In applications where the minimum values are used and the minimum cfm per person and cfm per sq It of Noor area are listed, use the larger minimum quantity. Where the crowd density is greater than normal or where better than satisfactory conditions are desired, use the recommended values. * SCHEDULED VENTILATION In comfort applications, where local codes permit, it is possible to reduce the capacity requirements of the installed equipment by reducing the ventilation air quantity at the time of peak load. This quantity can be reduced at times of peak to, in effect, minimize the outdoor air load. At times other than peak load, the calcuiated outdoor air quantity is used. Scheduled ventilation is recommended only for installations operating more than 12 hours or 3 hours longer than occupancy, to allow some time for flush- ing out the building when no odors are being generated. It has been found, by tests, that few complaints of stuffiness are encountered when the outdoor air quantity is reduced for short periods of time, provided the flushing period is available. It is recommended that the outdoor air quantity be reduced to no less than 40y0 of the recommended quantity as listed in TnOle 45. The procedure for estimating and controlling scheduled ventilation is as follows: 1. In estimating the cooling load, reduce the air quantity at design conditions to a minimum of 40y0 of the recommended air quantity. 2. Use a dry-bulb thermostat following the cooling and dehumidifying apparatus to control the leaving dewpoint such that: \ (:H,\I’TEK fi. INFILTKA’I‘ION i\Nl) l-97 VENTIL;\‘rION a. With the dewpoint at design, the damper motor closes the outdoor air damper to 40Cj’0 of the design ventilation air quantity. . b. As the dewpoint decreases below design, the outdoor air damper opens to the tlesign setting. Example 6 - Solution: The population density is typical, 100 sq ft per person, but the number of snlokers is considerable. Recommended ventilation = 50 X 15 = 750 cfm (Tnble 45) Minimum Ventilation Air Quantity, Office Space Given: A 5000 sq ft office with a ceiling height of 8 ft and 50 people. Approximately 4070 of the people smoke. Find: The ventilation air ventilation = 50 X 10 = 500 cfm (Tnble $5) 500 cfm will more than likely not main&in satisfactory conditions within the space because the number of smokers is considerable. Therefore, 750 cfm should be used in this apphcation. NOTE: Many applications have exhaust fans. This means that the outdoor air quantity must at least equal the exhausted air, otherwise the infiltration rate will increase. Tables -/6 and 47 list the approximate capacities of typical exhaust fans. The data in these tables were obtained from published ratings of several manufacturers of exhaust fans. quantity. TABLE 45-VENTILATION STANDARDS CFM APPLICATION Banking Space S M O K I N G . 15 25 15 10 Occosionol Very Heavy Corridors (Supply or Exhaust) t Department Stores Directors Rooms Drug Stores t Heavy - 50 30 - N o n e Extreme 50 7% Five and Ten Cent Storer NO”= Funeral Parlors NO”.? - GO’O& Some Meeting Rooms VeryHeavy Restaurant ‘Ofeteriat Dining Room t Some NO”= Considerable Considerable Considerable- School Rooms $ N o n e Shop Retail N0’le Theate’S NO”8 Theater Some .33 - 7% 30 25 5 30 .25 .05 - 7% 10 30 20 Heavy I /‘/ / 10 N o n e None None I _ -- 7% 10 I CFM PER SQ F T O F F L O O R Minimum* IO Considerable None Factories$$ R e s i d e n c e 20 30 10 Occasional Broker’s Board Rooms Cocktail Bars Laboratoriest Minimum* Considerable Beauty Parlors PERSON Recommended Some Some I Barber Shops Operating RoomsS** Hospitals Private Rooms t Words Hotel Rooms I PER I I 1 - 30 - 25 15 25 - 20 15 50 30 IO 15, 25 10 12 5 25 30 7% 1.25 .29 .25 - Toil&f (Exhaust) *When minimum is used, use the larger. fSee local codes which moy govern. tMay be governed by exhaust. @Jse these values unless governed by other sources of contamination or by local codes. **All outdoor air is recommended to overcome explosion hazard of anesthetics. \ I’/\KT I. LOAD ESTlhI.\-I‘ING 1-98 TABLE 47-PROPELLER FAN CAPACITIESFREE DELIVERY TABLE 46-CENTRIFUGALFAN CAPACITIES Motor Ou1let Horrepower Range Velocity Ranaa (fD”l) Fan Diameter (ill.) a Speed (rpm) Capacity* kfm) 1500 500 l/70-1 /20 800-2000 l/20-1/6 500-2500 12 12 l/20-% MO-2900 l/5-2 950-4300 1000-2000 18 850 1800 1000-2000 18 1140 2350 1000-2000 20 850 2400 1000-2000 20 1140 2750 20 1620 3300 *These typical air capacities were obtained from published ratings of several manufacturers of notionally known exhaust fans, single width, single inlet. Range of static pressures 1% to 1 ‘A inches. Fans with inlet diameter 10 inches and smaller are direct connected. *The capacity of these fans has been arbitrarily taken at 1000 fpm minimum and 2000 fpm maximum outlet velocity. Far there fans the usual selection probably is approximately 1500 fpm outlet velocity for ventilation. 1140 025 1725 1100 16 055 1000 16 1140 1500 *The capacities of fans of various manufacturers may from the values given above. vary + 10% . II l-99 CHAPTER 7. INTERNAL AND SYSTEM HEAT GAIN INTERNAL HEAT GAIN Internal heat gain is the sensible and latent heat released within the air conditioned space by the occupants, lights, appliances, machines, pipes, etc. This chapter outlines the procedures for determining the instantaneous heat gain from these sources. A portion of the heat gain from internal sources is radiant heat which is partially absorbed in the building structure, thereby reducing the instantaneous heat gain. Chapter 3, “Heat Storage, Diversity and Stratification,” contains the data and methods for estimating the actual cooling load from .’ 7 heat sources referred to in the following text. PEOPLE Heat is generated within the human body by oxidation, commonly called metabolic rate. The metabolic rate varies with the individual and with his activity level. The normal body processes are performed most efficiently at a deep tissue temperature of about 98.6 F; this temperature may vary only thru a narrow range. However, the human body is capable of maintaining this temperature, thru a wide ambient temperature range, by conserving or dissipating the heat generated within itself. This heat is carried to the surface of the body by the blood stream and is dissipated by: 1. Radiation from the body surface to the surrounding surfaces. 2. Convection from the body surface and the respiratory tract to the surrounding air. . Evaporation of moisture from the body surface and in the respiratory tract to the surrounding air. The amount of heat dissipated by radiation and convection is determined by the difference in temperature between the body surface and its surroundings. The body surface temperature is regulated by the quantity of blood being pumped to the surface; the more blood, the higher the surface temperature up to a limit of about 96 F. The heat dissipated by evaporation is determined by the difference in vapor pressure berween the body and the air. Basis of Table 48 - Heat Gain from People Table 48 is based on the metabolic rate of an average adult male, weighing 150 pounds, at different levels of activity, and generally for occupancies longer than S hours. These have been adjusted for typical compositions of mixed groups of males and females for the listed applications. The metabolic rate of women is about 85% of that for a male, and for children about 75yo. The heat gain for restaurant applications has been increased 30 Btu/hr sensible and 30 Btu/hr latent heat per person to include the food served. The data in Table 48 as noted are for continuous occupancy. The excess heat and moisture brought in by people, where short time occupancy is occurring (under 15 minutes), may increase the heat gain from people by as much as 10%. Use of Table 48 - Heat Gain from People To establish the proper heat gain, the room design temperature and the activity level of the occupants, must be known. Example 1 - Bowling Alley Given: A 10 lane bowling alley, 50 people, with a room design dry-bulb temperature of 75 F. Estimate one person per alley bowling, 20 of the remainder seated, and 20 standing. Find: Sensible and latent heat gain from people. Solution: Sensible heat gain = (10 X 525) + (20 X 240) + (20 X 280) = 15,650 Rtu/hr Latent heat gain = (10 X 925) + (20 x 160) + (20 X 270) = 17,850 Btu/hr LIGHTS Lights generate sensible heat by the conversion of the electrical power input into light and heat. The heat is dissipated by radiation to the surrounding surfaces, by conduction into the adjacent materials and by convection to the surrounding air. The radiant portion of the light load is partially stored, and the convection portion may be stratified as described on page 39. Refer to Table 12, page 35, to determine the actual cooling load. Incandescent lights convert approximately 10% of the power input into light with the rest being generated as heat within the bulb and dissipated by radiation, convection and conduction. About 80yo of the power input is dissipated by radiation and only about 10% by convection and conduction, Fig. 30. . 3 1 -CONVERSIONOFELECTRICPOWERTO F IG . H EAT AND LIGHT WITH FLUORESCENT LIGHTS, ,~PI'ROXI~IATE F1c.30 -CONVERSION OF ELECTRICPOWERTO HEAT ANDLIGHTWITHJNCANDESCENTLIGHTS, /~I'I'ROxIMATE Fluorescent lights convert about 25y0 of the power input into light, with about 25y0 being dissipated by radiation to the surrounding surfaces. The other 50% is dissipated by conduction and convection. In’ addition to this, approximately 257’, more heat is generated as heat in the ballast of the fluorescent lamp, Fig. 31. Table 49 indicates the basis for arriving at the gross heat gain from fluorescent or incandescent lights. . TABLE 4B-HEAT GAIN FROM PEOPLE Aver- DEGREE OF ACTIVITY Seated at rest Seated, very work Office Theater, Grade High walking Walking, School 390 seated 350 175 175 195 155 210 140 230 120 260 90 School 450 400 Offices, Apts., Hotels, College 180 220 195 205 215 185 240 160 275 125 475 - Dept., Retail, or Variety Stoic 450 180 270 200 250 215 235 2.F 205 285 165 5 5 0 Drug Store =j Standing, slowly 1 light worker Standing, slowly TYPICAL APPLICATION “ROOM DRY-BULB TEMPERATURE age M e t - Adobolic justed 80 F ‘, 78 F 75 F ’ 02 F 70 F Rate Met( A d u l t abolic Btu/hr Btu/hr Btu/hr Btu/hr Btu/hr Rate* Mole) I Btu/hr Btu/hr Sensible Latent Sensible Latent Sensible Latent Sensible ( Latent Sensible Latent 500 Bank 550 Sedentary work Rertaurantt 500 590 light Factory. 800 900 bench Moderate Walking, 1 180 320 1 200 300 / 220 280 1 255, 245 1 290 210 walking work dancing 3 mph Dance light work Hall Factory, fairly heavv work 1000 190 360 220 330 240 310 280 270 320 230 750 190 560 220 530 245 505 295 455 365 385 850 220 630 245 605 275 575 325 525 400 450 1000 270 730 300 700 330 670 380 620 460 540 605 845 I Heavy work B o w l i n g Alleyt, Factory 1500 1450 4 5 0 ’ 1000 *Adjusted Metabolic Rote is the metabolic rote to be applied to o mixed group of people with o typical percent composition based on the following foctorr: Metabolic rate, adult femole=Metabolic rote, adult male X 0.85 Metabolic rote, children =Metabolic rote, adult mole X 0.75 I 465 985 / 485 965 / I 525 921 tllestouront-Values for thir application include 60 Btu per hr for food per Individual (30 8tu reralble and 30 Btu latent heat per hrl. fBowling-Assume one person per alley actually bowling and oil others sitting, metabolic rote 400 Btu per hr; or standing, 550 Btu per hr. I i ! (;H/\I”I’EI< ,‘< TYPE HEAT GAIN* Btu/hr Fluoresc+ Total Light Watts X i.‘%T X 3.4 Incandescent Total Light Watts X 3.4 *Refer to Tables 12 and 13, pager 35-37 to determine actual cooling load. tfluorescent light wattage is multiplied by 1.25 to include heat gain in ballast. APPLIANCES iLfost appliances contribute both sensible and latent heat to a space. Electric appliances contribute latent heat, only by virtue of the function they l-101 ,-,’ 7 . INTERN,\I./IN11 SYS’I‘ICXI Flli.\~I‘ (;.\IN TABLE 49-HEAT GAIN FROM LIGHTS ,’ perform, that is, drying, cooking, etc., whereas gas burning appliances contribute additio4al moisture as a product of combustion. A properly designed hood with a positive exhaust system removes a considerable amount of the generated heat and moisture from most types of appliances. Basis of Tables 50 thru 52 - Heat Gain from Restaurant Miscellaneous Appliances Appliances and The data in these tables have been determined from manufacturers data, the American Gas Association data, Directory of Approved Gas Appliances and actual tests by Carrier Corporation. TABLE 50-HEAT GAIN FROM RESTAURANT APPLIANCES NOT OVERALL DIMENSIONS Less Legs and Handles (In.) APPLIANCE HOODED*-ELECTRIC TYPE OF CONTROL Coffee Brewer--‘/i gal Man. Wormer-!/i g a l Man. 4 Coffee Brewing Units MISCELLANEOUS DATA RECOM HEAT GAIN’ MAINMFR FOR AVG USE TAIN- r MAX ING Sensible Latent Total RATING RATE Heat Heat Heat Btu/hr Btu/hr Btu/hr Btu/hr Btu/hr 2240 306 25x 30 x 26H Auto. Water heater-2000 watts Brewers-2960 WOWS 16900 Coffee Urn-3 gal - 3 gal -5 gal 1 5 D i o x 34H 12 x 23 oval x 2 1 H 18 Dio x 37H Man. Auto. Auto. Black finish Nickel plated Nickel plated 11900 15300 17000 Doughnut 22 x 22 x 57H Auto. Exhaust system t o W ho motor 16000 with 4% g a l T a n k Machine outdoors- 10 x 13 x 25H Egg Boiler Food Warmer with Plate Warmer, per sq ft top surface Auto. \od Warmer without Plate W a r m e r , per sq ft top Man. Auto. surface Med. ht.-550 watts ._ L o w h t - 2 7 5 watts Insulated, separate heating unit for each pot. Plate warmer in base Ditto, without plate wormer Fry Kettle-l 1% lb fat 12 Dia x 14H Auto. Fry Kettle’-25 lb fat 16xlBx12H Auto. Frying area 12” x 14” Griddle, 18 x 18 x 8H Auto. F r y i n g t o p 1 B” x 14” G r i l l e , Meqt - 14 x 14 X~ iOH Auto. Cooking orea 10” x 12” Grille, 13 x 14 x 10H Auto. Roll Warmer ( 26xl7x13H Toaster, 15 x 15 x 2BH Toaster, - Frying . Sandwich Continuous Continuous Toaster, P o p - U p Waffle Iron , Waffle Iron for Ice Cream Sandwich . 1 3740 306 306 3000 2600 3600 900 230 220 90 1120 320 4800 1200 6000 2600 2200 3400 1700 1500 2300 4300 3700 5700 5000 1 1 1200 5aon / BOO 1 2000 1350 500 350 350 700 1020 400 200 350 550 8840 1100 1600 2400 4000 23800 2000 3800 5700 9500 8000 3100 1700 4800 10200 2800 *' 1900 3900 2100 6000 Grill area 12” x 12” 5600 1900 2700 700 3400 Auto. One 1500 400 1100 100 1200 Auto. 2 Slices wide360 slices/hr 7500 5000 5100 1300 6400 drawer 20 x 15 x 2BH Auto. 4 Slices wide720 slicer/hr 10200 6000 6100 2600 8700 6xllx9H Auto. 2 Slices 4150 1000 2450 450 2900 12 x 13 x 10H Auto. One waffle 7” dia 2480 600 1100 750 1850 14 x 13 x 10H Auto. 12 Cakes, 7500 1500 3100 2100 * I f p r o erly d e s i g n e d p o s i t i v e e x h a u s t h o o d i s u s e d , m u l t i p l y r e c o m m e n d e d P each 2%” x 3sqn value by .50. I 5200 ,. l-102 PAR-I‘ I. I.O/\D ES-I‘ihl,\‘l‘ING . Use of Tables 50 thru 52 - Heat Gain from Restaurant Miscellaneous Appliances Appliances and The M~~intnining Rate is the heat generated when the appliance is being maintained at operating temperature but not being used. The Recommended for Avel-nge Use values arc those which the appliance generates under normal use. These appliances seldom operate at maximum capacity during peak load since they are normally warmed up prior to the peak. The values in Tables 50 thru 52 are for unhooded appliances. If the appliance has a properly designed positive exhnust hood, reduce the sensible and the latent heat gains by 50%. A hood, to be effective, should extend beyond the appliance approximately 4 inches per foot of height between the appliance and the face of the hood. The lower edge should not bc higher than 4 feet above the appliance and the average fact velocity across the hood should not be less than 70 fpm. TABLE 51 -HEAT GAIN FROM RESTAURANT APPLIANCES NOT HOODED*-GAS BURNING AND STEAM HEATED OVERALL DIMENSIONS Less Legs and Handles (In.) APPLIANCE TYPE OF CONTROL GAS Coffee Brewer-% gal wormer--/i gal Mon. Man. RECOM HEAT GAIN FOR AVG USE MAINMFR TAINMAX ING RATING RATE Btu/hr Btu/hr MISCELLANEOUS DATA . Sensible Lotent Heat Heat Btu/hr Btu/hr Total Heat Btu/hr BURNING Combination brewer and wormer 3400 500 500 3200 3900 4 Brewers and 4% gal tank 1350 400 350 100 1700 500 7200 1BOO 9000 2900 2900 5800 Coffee Brewer Unit with Tank 19 x 30 x 26H C o f f e e u r n - 3 gal 15” Dia x 34H Auto. Block finish Coffee U r n - 3 g a l 1 2 x 23 oval x 2lH Auto. Nickel plated 3400 2500 2500. 5000 Coffee u r n - 5 g a l 18 D i a x 37H Auto. Nickel plated 4700 3900 3900 7800 / Food Warmer, Values SQ . fl too surface per I Man. I W a t e r b o t h t;pe I- , 2000 / 900 Fry Kettle-15 lb fat 12 x 20 x 1BH Auto. Frying area 10 x 10 14250 3000 Fry Kettle-28 lb fat 15 x 35 x 1lH Auto. Frying orea 11 x 16 24000 4500 22 x 14 x.17H s (1.4 $4 ft -grill surface) Man. Insulated 2 2 , 0 0 0 Btu/hr 1 5 , 0 0 0 Btu/hr 37000 .3 1 Man. Ring type burners 12000 to 22000 Btu/ea Grill-Brqil-O-Grill Top Burner Bottom Burner : Stoves, Short OrderOpen Top. Valuer per tq ft top surface Stoves, Short OrderC l o s e d T o p . Valuer per sq ft top surface Toaster, Continuous . 15 x 15 x 2BH Man. Auto. STEAM ( 850 1 450 14400-1 3600 / 1300 1 BOO0 1 14000 4200 4200 8400 Ring type burners 10000 to 12000 Btu/ea 11000 3300 3300 6600 2 Slices wide360 slices/hr 12000 7700 3300 11000 10000 HEATED C o f f e e u r n - 3 gal - 3 gal - 5 gal 1 5 D i o x 34H 12 x 23 oval x 21H 18 Dia x 37H Auto. Auto. Auto. Black finish Nickel plated Nickel plated 2900 2400 3400 1900 1600 2300 4800 4000 5700 Coffee 1 5 D i a x 34H 12 x 2 3 oval x 21H 18 Dia x 37H Man. Man. Man. Black finish Nickel plated Nickel plated 3100 2600 3700 3100 2600 3700 6200 5200 7400 500 900 Urn-3 g a l - 3 gal -5 g a l Food Warmer, per sq ft t o p surface Auto. 400 Food Warmer, per sq ft top surface Man. 450 *If properly designed positive exhaust hood is used, multiply recommended value by .50. 1150 1500 (;~-l/\l”1‘1111 7 . IN’I‘EI~N,\I, ,\Nl> SYS~I‘I-~c1 l-103 I rb:.\‘i (;,\I& TABLE 52-HEAT GAIN FROM MISCELLANEOUS APPLIANCES NOT HOODED* TYPE OF CONTROL APPLIANCE MFR MAX RATING MISCELLANEOUS DATA RECOM HEAT GAIN FOR AVG USE Sensible Heat Btu/hr Btu/hr totent Heat Btu/hr Total Heat Btu/hr ELECTRIC Hair Dryer, Blower Type 15 amps, 115 volts AC Man. Fan 165 watts, (low 915 wat+s, high 1580 watts) 5,370 2,300 400 2,700 Hair Dryer, helmet type, 6.5 amps, 115 volts AC Man. Fan 80 watts, (low 300 watts, high 710 watts) 2,400 1,870 330 2,200 Mon. 60 heaters at 25 watts each, 36 in normal ure 5,100 150 1,000 23,460 35,460 Permanent Wave Machine i Pressurized Instrument Washer and Sterilizer 1l”X Il”X 2 2 ” Neon Sign, per linear ft tube ‘I$” outside dia %” outside dio Solution and/or Blanket Warmer 18” x 30” x 72” 18” x 24” x 72” Steriliker Dressing Sterilizer, Auto. Auto. Rectangular Sterilizer, Bulk Auto. Auto. Auto. Auto. Auto. Auto. Auto. Water ’ 850 12,000 30 60 1,200 1,050 16” x 24” 20” x 36” 24” x 24” x 24” x ~~~~~-x- 24” 24” 36” 3.6” 10 gallon 15 gallon 3,000 2,400 4,200 3,450 8,700 24,000 18,300 47,300 34,800 4 1,700 56,200 68,500. 161,700 / 184,000 2 10,000 2 1,000 27,000 36,000 45,000 97,500 140,000 180,000 55,800 68,700 92,200 113,500 259,200 324,000 390,000 4,100 6,100 16,500 24,600 9,600 23,300 x 36” x 48” x 48” x-m 36” x 42” x 84”. 42” x 48” x 96” 48” x 54” x 96” Auto. Auto. 30 60 Sterilizer, Instrument Auto. Auto. Auto. Auto. Auto. 6” x 8” x 17” 9” x 10” x 20” 10” x 12” x 22” 10”~ 12”~ 3 6 ” 12” x 16” x 24” 2,700 5,100 8,100 10,200 9,200 Sterilizer, Utensil Auto. Auto. 16” x 16” x 24” 20” x 20” x 24” 10,600 12,300 20,400 25,600 Auto. Auto. Model 120 Amer Sterilizer Co Model 100 Amer Sterilizer Co 2,000 1,200 4,200 2,100 Sterilizer, Hot Air Water Still 5 gal/hour X-ray Machines, for making pictures Physicians and Dentists office ’ X-ray Machines, for therapy Burners, Laboratory small bunsen small bunsen fishtail burner fishtail burner large bunsen Cigar Lighter Dryer System 5 helmets 10 helmets I 20,600 30,700 2,400 3,900 5,900 9,400 8,600 1,700 5,100 9,000 14,000 19,600 17,800 3 1,000 37,900 6,200 3,300 2,700 NOW 4,400 NOW NOW Heat load may be opprecioblewrite mfg for data GAS BURNING Mon. 36 dia barrel with manufactured gas 1,800 Man. Mon. 56 dia with nat gos %6 dia with nat gas Mon. Mon. 36 dio bar with nat gas I ‘/2 dio mouth, adi orifice Mon. Continuous flame type Auto. Auto. Consists of heater L fan which blows hot air thru duct system to helmets Hair ‘If properly designed positive exhaust hood is used, multiply recommended value by .50 1 960 240 3,000 3,500 1,680 1,960 420 490 2,100 2,450 5,500 6,000 3,080 3,350 770 850 3,850 4,200 2,500 1 900 / 100 4,000 6,000 j 1 1,200 1,000 19,000 27,000 ’ l-104 PART Example 2 - Restaurant Given: A restaurant with the following electric appliances with a properly designed positive exhaust hood on each: 1. Two 5.gallon coffee urns, both used in the morning, only one used either in the afternoon or evening. 2. One 20 sq Et food warmer without plate warmer. 3. Two 24 x 20 x 10 inch frying griddles. 4. One 4.slice pop-up toaster, used only in the morning. 5. Two 25 II) deep fat, fry kettles. Find: Heat gain from evening meal. Solution: Use Table these appliances during the 50. afternoon Sensible 1. Coffee Urn - only one in use: Sensible heat gain = 3400 X .50 = Latent heat gain = 2300 X .50 = 1700 2. Food Warmer: Sensible heat gain = 20 X 200 X .50 = Latent heat gain = 20 x 350 x 50 = 2000 3. and Latent 1150 3500 Frying Griddles: Sensible heat gain = 2 X 5300 X .50 = Latent heat gain = 2 x 2900 x .50 = 1. L O A D ESTIM,\TINC lf the fluid is conveyed outside of the air conditioned space, only the inefficiency of the motor driving fan or pump should be included in room sensible heat gain. If the temperature of the fluid is maintained by a separate source, these heat gains to the fluid heat of compression are a load on this separate source only. The heat gain or loss from the system sho~~lci be calculated separately (“System Heat Gain,” p. 120). Motors driving process machinery (lathe, punch press, etc.): The total power input to the machine is dissipated as heat at the machine. If the product is removed from the conditioned space at a higher temperature than it came in, some of the heat input into the machine is removed and should not be considered a heat gain to the conditioned space. . The heat added to a product is determined by multiplying the number of pounds of material handled per hour by the specific heat and temperature rise. 5300 2900 4. Toaster - not in use 5. Fry Kettles: Sensible heat gain = 2 X 3800 X .50 = Latent heat gain = 2 x 5700 x .50 = 3800 5700 Total sensible heat gain = 12,800 Total latent heat gain = 13,250 L ELECTRIC MOTORS Electric motors contribute sensible heat to a space by converting the electrical power input to heat. Some of this power input is dissipated as heat in the motor frame and can be evaluated as input X (1 - motor eff). The rest of the power input (brake horsepower ?r motor output) is dissipated by the driven machine and in the drive mechanism. The driven machine utilizes this motor output to do work which may or may not result in a heat gain to the space. Motors driving funs and pumps: The power input increases the pressure and velocity of the fluid and the temperature of the fluid. The increased energy level in the fluid is degenerated in pressure drop throughout the system and appears as a heat gain to the fluid at the point where pressure drop occurs. This heat gain does not appear as a temperature rise because, as the pressure reduces, the fluid expands. The fluid expansion is a cooling process which exactly offsets the heat generated by friction. The heat of compression required to increase the energy level is generated at the fan or pump and is a heat gain at this point. . Basis of Table 53 - Heat Gain from Electric Motors Table 53 is based on average efficiencies of squirrel cage induction open type integral horsepower and fractional horsepower motors. Power supply for fractional horsepower motors is 110 or 220 Golts, 60 cycle, single phase; for integral horsepower motors, 208, 220,‘or 440 volts, 60 cycle, 2 or 3 phase general purpose and constant speed, 1160 or 1750 rpm. This table may also be applied with reasonable accuracy to 50 cycle, single phase a-c, 50 and 60 cycle enclosed and fractional horsepower polyphase motors. Use of Table 53 - Heat Gain from Electric Motors The data in Table 53 includes the heat gain from electric motors and their driven machines when both the motor and the driven machine are in the conditioned space, or when only the driven machine is in the conditioned space, or when only the motor is in the conditioned space. Caution: The power input to electric motors does not necessarily equal the rated horsepower divided by the motor efficiency. Frequently these motors may be operating under a continuous overload, or may be operating at less than rated capacity. It is always advisable to measure the power input wherever possible. This is especially important in estimates for industrial installations where the motor-machine load is normally a major portion of the cooling load. CHAI’TER l-105 7 . INTERNAL ANIl SYSTI:M HE,\‘I’ G,\Ih- When retdings me obtairred dizrtly in watts and when both motors and driven machines 3x-c in the air conditioned space, the heat gain is equal to the number of watts times the factor 3.4 Btu/(watt)(hr). When the machine is in the contlitioncd space and the motor outside, multiply the watts by the motor efficiency ant1 by the factor 3.4 to determine heat gain to the space. When the machine is outside the conditioned multiply the watts by one minus the motor efficiency antI by the Factor 3.4. SIX’CC, ~\lthough the results are less accurate, it may be expedient to obtain power input measurements using ;I clamp-on ammeter and voltmeter. These instruments permit instantaneous readings only. They afford means for determining the load factor but the usage factor must be obtained by a careful investigation of the operating conditions. TABLE 53-HEAT GAIN FROM ELECTRIC MOTORS CONTINUOUS OPERATION* LOCATION OF EQUIPMENT WITH RESPECT TO C O N D I T I O N E D S P A C E O R A I R SlREAMt NAMEPLATE+ OR B R A K E HORSEPOWER Btu per Hour ‘I20 ‘I12 ‘vi ‘ii ‘yi 40 49 55 60 64 \ 320 430 580 710 1,000 130 210 320 430 640 190 220 260 280 360 66 70 72 79 80 1,290 1,820 2.680 3,220 4.770 850 1,280 1,930 2,540 3,820 440 540 750 680 950 2 3 5 ?‘/i 10 80 01 82 85 85 6,380 9,450 15,600 22,500 30,000 5,100 7,650 12,800 19,100 25,500 1,280 1,800 2,800 3,400 4,500 15 20 25 30: 40 06 87 88 09 89 44,500 58,500 72,400 85,800 1 15,000 38,200 51,000 63,600 76,400 102,000 6,300 7,500 8,800 9,400 13,000 50 60 75 100 125 89 09 90 90 90 143,000 172,000 212,000 284,000 354,000 127,000 153,000 191,000 255,000 3 18,000 16,000 19,000 2 1,000 29,000 36,000 150 200 250 91 91 91 420,000 ’ 560,000 700,000 382,000 5 10,000 636,000 38,000 50,000 64,000 m 1% ‘h % 1 1 ‘/a . I ‘For intermittent operation, an appropriate usage factor should be used, preferably measured. tlf motors are overloaded and amount of overloading is unknown, multiply the above heat gain factors by the following maximum service fadorr: Maximum Service Factors Horsepower AC Open Type DC Open Type ‘/20-H 1.4 - ‘h-‘/i 1% - ?4 1 1 N-2 3-250 1.35 - I .25 - 1.25 1.15 1.20 1.15 1.15 1.15 No overload is allowable with enclosed motors. SFor a fan or pump in air conditioned space, exhausting air and pumping fluid to outside of space, use values in last column. l-106 I’AKT I . LOAD ESTliV,\TING \ following is a conversion table which can be used to determine load Eactors from measurements: The TO FIND + HP OUTPUT KILOWATTS INPUT Direct Current IXEXeff I X E 746 1,000 1 IXEXpfXeff IXEXpf Phase 746 1,000 3 or 4 Wire I X E X pf X elf X 1.73 I X E X pf X 1.73 3 Phase 746 1,000 4 Wire IXEXpfXeffXZ IXEXZXpf 2 Phase 746 Where I = amperes E = volts IOTE: I eff = efficiency pf = power factor For 2 phase, 3 wire circuit, common conductor current is 1.41 times that in either of the other two conductors. Example 3 - Electric Motor Heaf Gain in a Factory (Motor 8hp Established by a Survey) Given: 1. Forty-five 10 hp motors operated at 80’9” rated capacity, driving various types of machines located within air conditioned space (lathes, screw machines, etc.) Five 10 hp motors operated at 800/, rated capacity, driving screw machines, each handling 5000 11~s of bronze per hr. Both the final product and the shavings from the screw machines are removed from the space on conveyor belts. Rise in bronze temperature is 30 F; sp ht is .Ol Btu/(lb) (F). 2. Ten 5 hp motors (5 bhp) driving fans, exhausting air to the outdoors. 3. Three 20 hp motors (20 bhp) driving process water pumps, water discarded outdoors. Find: Total heat gain from motors. 8olution: Use Table 53. Sensible Heat Btu/hr 1. Machines - Heat gain to space = 45 X 30,000 X .80 = 1,080,OOO Heat gain from screw machines = 5 X 30.000 X .80 = 120,000 Btu/hr Heat removed from space from screw machine work = 5000 X 5 X 30 X .Ol = 7,500 Btu/hr Net heat gain from screw machines to space = 120,000 - 7,500 = 112,500 2. Fan exhausting air to the outdoors: Heat gain to space = 10 X 2800 = 28,000 3. Process water pumped to outside air conditioned space Heat gain to space = 3 x 7500 = 22,500 Total heat gain from motors on machines, fans, and pumps = 1,243,OOO NOTE: If the process water were to be recirculated and coolctl in the circuit from an outside source, the heat gain to the water 3 X (58,500 - 7500) = 153,000 Btu/hr would I,ecomc a load on this outside source. ! ! ! / :‘I PIPING, TANKS AND EVAPORATION OF WATER FROM A FREE SURFACE 1,000 I ! 1 Gain Hot pipes and tanks add sensible heat to a space by convection and radiation. Conversely, cold pipes remove sensible heat. All open tanks containing hot water contribute not only sensible heat but also latent heat due to evaporation. In industrial plants, furnaces or dryers are often encountered. These contribute sensible heat to the space by convection and radiation from the outside surfaces, and frequently dryers also contribute sensible and latent heat from the drying process. Basis of Tables 54 thru 58 - Heat Gain from Piping, Tanks and Evaporation of Water Table 54 is based on nominal flow in the pipe and a convection heat flow from a horizontal pipe of 1.016 x(&)2x x (&).lX’ (temp diff between hot water or steam . and room). The radiation from horizontal pipes is expressed bY 17.23 x lo-10 x emissivity x (T,” - T,“) where T, = room surface temp, deg R T, = pipe surface temp, deg R Tables 55 and 56 are based on the same equation and an insulation resistance of approximately 2.5 per inch of thickness for 85yo magnesia and 2.9 per inch of thickness with moulded type. Caution: Tables 55 and 56 do not include an allowance for fittings. A safety factor of 10% should be added for pipe runs having numerous fittings. Table 57 is based on an emissivity of 0.9 for painted metal and painted or bare wood and concrete. The emissivity of chrome, bright nickel plate, stainless steel, or galvanized ‘iron is 0.4. The resistance (r) of wood is 0.833 per inch and of concrete 0.08 per inch. The metal surface temperature has been assumed equal to the water temperature. NOTE: The heat gain from furnaces and ovens can be estimated from Table 57, using the outside temperature of furnace and oven. Table 58 is based on the following formula for still air: Heat of evaporation = 95 (vapor pressure ) / / .i . l-107 CHAPTER 7 . I N T E R N A L :\NI) S Y S T E M H E A T GAIN differential between water and air), where vapor pressure is expressed in inches of mercury, and the room conditions are 75 F db and 50% rh. , b u l b temperature (Ib/hr x t e m p diff x .45). T h e latent heat gain is equal to the pounds per hour escaping times 1050 Btu/lb. Use of Tables 54 thru 58 - Heat Gain from Piping, Tanks and Evaporation of Water MOISTURE ABSORPTION When moisture (regain) is absorbed by hygroscopic materials, sensible heat is added to the space. The heat so gained is equal to the latent heat of vaporization which is approximately 1050 Btu/lb times the pounds of water absorbed. This sensible heat is an addition to room sensible heat, and a deduction from room latent heat if the hygroscopic material is removed from the conditioned space. Example 4 - Heat Gain from Hot Water Pipe and Storage Tank Given: Room conditions - 75 F db, 5070 rh 50 ft of IO-inch uninsulated hot water (125 F) pipe. The hot water is stored in a 10 ft wide x 20 ft long x 10 ft high, painted metal tank with the top open to the atmosphere. The tank is supported on open steel framework. LATENT HEAT GAIN-CREDIT SENSIBLE HEAT Find: Sensible and latent heat gain Btu/hr Jing - Sensible heat gain = 50 X 50 X 4.76 = 11,900 Tank - Sensible heat gain, sides = (20 x 10 x 2) + (10 x 10 x 2) X50X1.8= 54,000 - Sensible heat gain, bottom = (20 x 10) x 50 x 1.5 = 15,000 ROOM Some forms of latent heat gain reduce room sensible heat. Moisture evaporating at the room wet-bulb temperature (not heated or cooled from external source) utilizes room sensible heat for heat of evaporation. This form of latent heat gain should be deducted from room sensible heat and added to room latent heat. This does not change the total room heat gain, but may have considerable effect on , the sensible heat factor. When the evaporation of moisture derives its heat from another source such as steam or electric heating coils, only the latent heat gain to the room is figured; room sensible heat is not reduced. The power input to the steam or electric coils balances the heat of evaporation except for the initial warmup of the water. Solution: Use Tables 54,57 and 58 TO Total sensible heat gain = 80,900 Total latent heat gain, top = (20 X 10) X 330 = 66,000 STEAM When steam is escaping into the conditioned space, the room sensible heat gain is only that heat represented by the difference in heat content of steam at the steam temperature and at the room dry- TABLE 54-HEAT TRANSMISSION COEFFICIENTS FOR BARE STEEL PIPES Btu/(hr) (linear ft) H O T -dOMINAL PIPE SIZE (in.) ??H 3 3H 4 5 6 0 10 12 ’ (deg F diff between pipe and surrounding air) W A T E R 300 F 100 psig 338 F 157 F 230 F 268 F 0.61 0.75 0.92 1.14 1.29 0.71 0.87 1.07 1.32 1.49 0.76 0.93 1.51 1.58 2.15 2.43 2.72 2.26 2.55 2.85 1.84 2.19 2.63 2.97 3.32 1.99 2.36 2.84 3.22 3.59 3.30 3.89 4.96 6.09 7.15 3.47 4.07 5.21 6.41 7.50 4.05 4.77 6.10 7.49 a.80 4.39 5.16 6.61 a.12 9.53 120 F 150 F 18OF 50 F 80 110 F 140 F 0.46 0.56 0.68 0.85 0.96 0.50 0.61 0.74 0.92 1.04 0.55 0.67 0.82 1.01 1.15 0.58 0.72 0.88 1.09 1.23 1.28 1.53 1.83 2.06 2.30 1.41 1.68 2.01 2.22 2.53 2.80 3.29 4.22 5.18 6.07 3.08 3.63 4.64 5.68 6.67 -‘. 1.18 1.40 1.68 1.90 2.12 2.58 3.04 ' 3.88 4.76 5.59 * A t 7 0 F d b room temperature _ 5 psig 227 F 210 F TEMPERATURE F STEAM I 5 0 psig DIFFERENCE* I.80 I 1.aa 1.15 1.43 1.63 l-108 PART 1. LOAD ESTIMATING L TABLE 55-HEAT TRANSMISSION COEFFICIENTS FOR INSULATED PIPES* Btu/(hr) (linear ft) (deg F diff between pipe and room) IRON PIPE SIZE (in.) 1 In. Thick 1% In. Thick 2 In. Thick % 0.16 % 0.18 0.20 0.14 0.15 0.17 0.12 0.13 0.15 1 ‘/!I 1 ‘/I 2 2% 0.24 0.26 0.30 0.35 0.20 0.21 0.24 0.27 0.17 0.18 0.2 I 0.24 3 31% 4 5 0.40 0.45 0.49 0.59 0.32 0.35 0.38 0.45 0.27 0.30 0.32 0.38 6 0.68 0.85 1.04 1.22 0.52 0.65 0.78 0.90 0.43 0.53 0.64 0.73 a5 PERCENT MAGNESIA INSULATIONt 1 a 10 12 . *No allowance for fittings. This table applies only to straight runs of pipe. When numerous fittings exist, a suitable safety factor must be included. This added heat gain at the fittings moy be OS much as 10%. Generally this table can be used without adding this safety factor. tOther insulation. if other types of insulation we used, multiply the above values by the factors shown in the following table: PIPE MATERIAL COVERING Corrugated Asbestos (Air Cell) 4 Ply per inch 6 Ply per inch 8 Ply per inch Laminated Asbestos (Sponge Felt) Mineral Wool Diatomaceous Silica (Super-X) Brown Asbestos Fiber (Wool Felt) FACTORS 1.36 1.23 1.19 0.98 1 .oo 1.36 0.88 TABLE 56--HEAT TRANSMISSION COEFFICIENTS FOR INSULATED COLD PIPES* MOULDED TYPEt Btu/(hr) (linear ft) (deg F diff between pipe and room) ., ,.“,,,.” ,. v’““;I/ ‘,I ICE WATER litbN:tiIPE,SiZE. .,:.” (in.) 1,’ .,.” 1”:1 _)_ :,.;;; ‘/?2 :’ 3/! ‘. :‘:;y, ‘, ..~ “;..::.. ,i i ‘A, ,: T. : :’ 1 ‘A ,,,I,” 2. 1 : ,, ‘* .2’/a {‘I, BRINE .,v+ ;:,HEAyY j B R I N E ‘ i Coefficient Actual Thickness of Insulation (In.1 Coefficient 0.11 0.12 0.14 2.0 2.0 2.0 0.10 0.1 I 0.12 2.8 2.9 3.0 0.09 0.09 0.10 1.6 1.5 1.5 1.5 0.16 0.17 0.20 0.23 2.4 2.5 2.5 2.6 0.13 0.13 0.15 0.17 3.1 3.2 3.3 3.3 0.1 1 0.12 0.13 0.15 1.5 1.5 I .7 1.7 0.27 0.29 0.30 0.35 2.7 2.9 2.9 3.0 0.19 0.19 0.2 I 0.24 3.4 3.5 3.7 3.9 0.16 0.18 1.7 0.40 3.0 0.26 4.0 0.23 II .9 .9 1.9 0.46 0.56 0.65 3.0 3.0 3.0 0.32 0.38 0.44 4.0 4.0 4.0 0.26 0.31 0.36 Actual Thickness of Insulation (In.) 1.5 1.6 1.6 .il Actual Thicknesi. :, :* :. ! _I’ of Insulation (In.) Coefficient 0.18 0.20 *No allowance for fittings. This table applier only to straight runs of pipe. When numerous fittings exist, a suitable safety factor must be included. This added heat goin at the fittings moy be OS much ~1s 10%. G enerally this table can be used without adding this safety factor. tlnrulation material. Valuer in this table are based on CI material having o conductivity kE0.30.However, o 15% safety factor war added to this k value to compensate for seams and imperfect workmanship. The table applies to either cork covering (kE0.291, or mineral wool board (k=0.32). The thickness given above is for molded mineral wool board which is usually some 5 to 10% greater than molded cork board. CH/II’TEK 7 . IN-I’EKN;\I. .\NI> TABLE SYS’I‘ICRI FJ1:.\‘1‘ 57-HEAT TRANSMJSSION l-109 C..\IN COEFFICIENTS FOR UNINSULATED TANKS SENSIBLE HEAT GAIN* Btu/(hr) (sq ft) (deg F diff between liquid and room) METAL Painted CONSTRUCTION Temp Vertical(Sider) TOP l3attmn Bright (Nickel) Diff 50 F 100 F 150 F 1.8 2.1 1.5 2.0 2.4 1.7 2.3 2.7 2.0 Temp WOOD 2 % in. Thick CONCRETE 6 in. Thick Painted or Bare Painted or Bore Diff Temp Tamp Diff Diff *To estimate latent heat load if water is being evaporated, see Table 58 TABLE 58-EVAPORATION FROM A FREE WATER SURFACE-LATENT HEAT GAIN STILL AIR, ROOM AT 75 F db. WATER TEMP Btu/(hr)(sq ft) i 75 F 42 1 I 50% RH 100 F 1 125 F 1 150 F 1 175 F 1 200 F 140 / 330 ) 680 ( 1260 ) 2190 SYSTEM HEAT GAIN The system heat gain is considered as the heat added to or lost by the system components, such as the ducts, piping, air conditioning fan, and pump, etc. This heat gain must be estimated and included in the load estimate but can be accurately evaluated only after the system has been designed. SUPPLY AIR DUCT HEAT GAIN The supply duct normally has 50 F db to 60 F dl r flowing through it. The duct may pass through an unconditioned space having a temperature of, say, 90 F db and up. This results in a heat gain to the’ duct before it reaches the space to be conditioned. This, in effect, reduces the cooling capacity of the conditioned air. To compensate for it, the cooling capacity of the air quantity must be increased. It is recommended that long runs of ducts in unconditioned spaces be insulated to minimize heat gain. Basis or Chart 3 - Percent Room Sensible Heat to be Added for Heat Gain to Supply Duct Chart 3 is based on a difference of 30 F db between supply air and unconditioned space, a supply duct velocity of 1800 fpm in a square duct, still air on the outside of the duct and a supply air rise of 17 F db. Correction factors for different room temperatures, duct velocities and temperature differences are included below Chnrt 3. Values are plotted for use with uninsulated, furred and insulated ducts. Use of Chart 3 - Percent Room Sensible Heat to be Added for Heat Gain to Supply Duct To use this chart, evaluate the length of duct running thru the unconditioned space, the temperature of unconditioned space, the duct velocity, the suppIy air temperature, and room sensible heat subtotal. Example 5 - Heat Gain to Supply Duct Given: 20 ft of uninsulated duct in unconditioned space at 100 F dh Duct velocity - 2000 fpm Supply air temperature - 60 F db Room sensible heat gain - 100,000 Btu/hr Find: Percent addition to room sensible heat Solution: The supply air to unconditioned ence = 100 - GO = 40 F db From Chart 3, percent Correction for 40 F 2000 fpm duct velocity Actual percent addition space temperature addition = 4.5% db temperature difference and = 1.26 = 4.5 X 1.26 = 5.7% differ- I’.\K-r I. I.O.\I) I:S’I‘lhI.\~I‘ING l-110 CHART 3-HEAT GAIN TO SUPPLY DUCT Percent of Room Sensible Heat 0 0 1000 2000 DUCT 3000 4000 5000 V E L O C I T Y (FPMI . MULTIPLYING FACTORS FOR OTHER ROOM TEMPERATURES Room Temp Q = UPI X (2.162;‘6x5 Multiplying 75 1.10 76 77 78 79 80 1.06 1 .oo 0.97 0.94 Factor . 0.92 /r&A: UPI (‘3--11) where: A = duct area (sq ft) Q = duct heat gain (Btu/hr) U = duct heat transmission factor (Btu/hr-sq ft-F) V = duct velocity (fpm) P = rectangular duct perimeter (ft) tl = temperature of supply air entering duct I = duct length (ft) t3 Based on formulas in ASHRAE SUPPLY AIR DUCT LEAKAGE LOSS Air leakage from the supply duct may be a serious loss of cooling effect, except when it leaks into the conditioned space. This loss of cooling effect must be added to the room sensible and latent heat load. Experience indicates that the average air leakage from the entire length of supply ducts, whether large or small systems, averages around lo<;& of the supply air quantity. Smaller leakage per foot of length for larger perimeter ducts appears to be counterbalanced by the longer length of run. Individual work- (F) = temperature of surrounding air (F) Guide 1963, p. 184, 185. manship is the greatest variable, and duct leakages from 5yo to 30% have been found. The following is a guide to the evaluation of duct leakages under various conditions: I. Bare ducts within conditioned spice - usually not necessary to figure leakage. 2. Furred or insulated ducts within conditioned space - a matter of judgment, depending on whether the leakage air actually gets into the room. (;H,\I’~1‘~:11 7 . IN’I‘EIIN,\I, :\NI) SYS’I‘EM 1-111 HE,\‘1 C;,\IN TABLE 59-HEAT GAIN FROM AIR CONDITIONING FAN HORSEPOWER, DRAW-THRU SYSTEMIf CENTRAL STATION SYSTEMS$ APPLIED OR UNITARY SYSTEM** Temp Diff Room to Supply Air Temp Diff Room to Supply Air 10 F 15 F 20 F Fan Motor Not in Conditioned Space or Air Stream F a n Motortt in Conditioned space or Air Stream 1.25 3.9 1so 4.6 L----l- 1.75 2.00 30 F 25 F PERCENT OF ROOM 5.00 6.00 8.00 19.2 24.4 38.0 12.8 16.3 25.4 9.6 12.2 19.0 7.7 9.9 15.2 6.4 0.2 12.7 0.75 1 .oo 1.6 2.6 3.6 1.1 1.8 2.4 0.8 1.3 1.8 0.6 1.1 1.5 0.5 0.9 1.2 2.7 4.2 5.8 1.8 2.8 3.8 5.0 6.0 3.4 4.0 7.0 4.7 2.5 3.0 3.5 2.0 2.4 2.8 1.7 2.0 2.4 7.6 9.2 10.7 5.1 6.1 7.2 1.25 1 so I 2.00 3.00 4.00 30 F 5.4 4.00 0.50 L 25 F HEAT* 6.2 10.4 15.3 3.00 20 F 15 F 10 F SENSIBLE 1.75 I 1.4 2.1 2.9 I 1.1 1.7 2.3 I 0.9 1.4 1.9 8.0 13.2 19.0 *Excludes from heat gain, typical values for bearing losses, etc. which are dissipated in apparatus room. tFon Total Pressure equals fan static pressure plus velocity pressure at fan discharge. Below 1200 fpm the fan total pressuce the fan static. Above 1200 fpm the total pressure should be figured. $7Oyo is approximately equal to fan efficiency assumed. **5Oyo fan efficiency assumed. tt8Oyo motor and drive efficiency assumed. $$For draw-thru systems, this heat is on addition to the supply air heat gain and is added to the room’ sensible heat. For blow-thru systems this fan heat is added to the grand total heat; use the RSH times the percent listed and add to the GTH. 3. All ducts outside the conditioned space assume 10% leakage. This leakage is a total loss and the full amount must be included. When only part of the supply duct is outside the conditioned space, include that fraction of 10% as the leakage. (Fraction is ratio of length outside of conditioned space to total length of supply duct.) HEAT GAIN FROM AIR FAN HORSEPOWER CONDITIONING The inefficiency of the air conditioning equip= ment fan and the heat of compression adds heat to the system as described under “Electric Motors.” In the case of draw-through systems, this heat is an addition to the supply air heat gain and should be added to the room sensible heat. With blow-through systems (fan blowing air through the coil, etc.) the fan heat added is a load on the dehumidifier and, therefore, should be added to the grand total heat (see “Percent Addition to Grand Total Heat”). Basis of Table 59 - Heat Gain from Air Conditioning Fan Horsepower The air conditioning fan adds heat to the system in the following manner: 1. Immediate temperature rise in the air due to the inefficiency of the fan. 2. Energy gain in the air as a pressure and/or velocity rise. 3. With the motor and drive in the air stream or conditioned space, the heat generated by the inefficiency of the motor and drive is also an immediate heat gain. The fan efficiencies are about 707, for central station type fans and about 50% for packaged equipment fans. 1-112 PAKT Use of Table 59 - Heat Gain from Air Conditioning Fan Horsepower The approximate system pressure loss and dehumidified air rise (room minus supply air temperature) differential must be estimated from the system characteristics and type of application. These should be checked from the final system design. The normal comfort application has a dehumidified air rise of between 15 F db and 25 F db and the fan total pressure depends on the amount of ductwork involved, the number of fittings (elbows, etc.) in the ductwork and the type of air distribution system used. Normally, the fan total pressure can be approximated as follows: 1. No ductwork (packaged equipment) - 0.5 to 1.00 inches of water. 2. Moderate amount of ductwork, low velocity systems - 0.75 to 1.50 inches of water. 3. Considerable ductwork, low velocity system 1.25 to 2.00 inches of water. 4. Moderate amount of ductwork, high pressure system - 2.00 to 4.00 inches of water. 5. Considerable ductwork, high pressure system - 3.00 to 6.00 inches of water. Example 6 -Heat Gain from Air Conditioning Fan Horsepower Given: Same data as Example 5 80 ft of supply duct in conditioned space Find: Percent addition to room sensible heat. Solution: Assume 1.50 inches of water, fan total pressure, and 20 F db dehumidifier rise. Refer to Table 59. Heat gain from fan horsepower = 2.3’% SAFETY FACTOR AND PERCENT ADDITIONS TO ROOM SENSIBLE AND LATENT HEAT A safety factor to be added to the room sensible heat sub-total should be considered as strictly a factor of probable error in the survey or estimate, and should usually be between 0% and 5yo. The total room sensible heat is the sub-total plus percentage additions to allow for (1) supply duct heat gain, (2) supply duct leakage losses, (3) fan horsepower and (4) safety factor, as explained in the preceding paragraph. Example 7 - Percent Addition to Room Sensible Heat Given: Same data as Examples 5 and 6 Find: Percent addition to room sensible heat gain sub-total . I. LOAD ESTIMATING Solution: = Supply duct heat gain Supply duct leakage (20 ft duct of total 100 ft) = = Fan horsepower = Safety factor 5.7% 2.07, 2.37” 0.07, = 10.0% Total percent addition to RSH The percent additions to room latent heat for supply duct leakage loss and safety factor should be the same as the corresponding percent additions to room sensible heat. RETURN AIR DUCT HEAT AND LEAKAGE GAIN The evaluation of heat and leakage effects on return air ducts is made in the same manner as for supply air ducts, except that the process is reversed; there is inward gain of hot moist air instead of loss of cooling effect. . Chart 3 can be used to approximate heat gain to the return duct system in terms of percent of RSH, using the following procedure: 1. Using RSH and the length of return air duct, use Chart 3 to establish the percent heat gain. 2. Use the multiplying factor from table below Chart 3 to adjust the percent heat gam for actual temperature difference between the air surrounding the return air duct and the air inside the duct, and also for the actual velocity. 3. Multiply the resulting percentage of heat gain by the ratio of RSH to GTH. 4. Apply the resulting heat gain percentage to GTH. To determine the return air duct leakage, apply the following reasoning: 1. Bare duct within conditioned space - no inleakage. 2. Furred duct within conditioned space or furred space used for return air - a matter of judgment, depending on whether the furred space may connect to unconditioned space. 3. Ducts outside conditioned space - assume up to 3’7, inleakage, depending on the length of duct. If there is only a short connection between conditioned space and apparatus, inleakage may be disreg?rded. If there is a long run of duct, then apply judgment as to the amount of inleakage. HEAT GAIN FROM DEHUMIDIFIER PUMP HORSEPOWER With dehumidifier systems, the horsepower required to pump the water adds heat to the system as outlined under “Electric Motors”. This heat will be an addition to the grand total heat. 1-113 TABLE 60-HEAT GAIN FROM DEHUMIDIFIER PUMP HORSEPOWER L A R GE PUMPS+ SMALL PUMPS* O-100 GPM CHILLED 5 F 7 F WATER TEMP 10 F RISE 12 F PUMP HEAD (ftl 35 70 100 *Efficiency 50% CHILLED 15 F 5 F 7 F 100 GPM AND LARGER WATER TEMP RISE 10 F 12 F 15 P 0.5 1.5 2.0 0.5 I .o I .5 0.5 I .o I .o PERCENT OF GRAND TOTAL HEAT 2.0 3.5 5.0 1.5 2.5 4.0 tEfficiency 1 .o 2.0 2.5 I .o 1.5 2.0 0.5 1 .o 1.5 1.5 2.5 4.0 ! 1 .o 2.0 3.0 70% Basis of Table 60 - Heat Gain from Dehumidifier Pump Horsepower Table 60 is based on pump efficiencies of 50% for small pumps a n d 70% Eor large pumps. Small pumps are considered to have a capacity of less than 100 gallons; large pumps, more than 100 gallons. u. I Table 60 - Heat Gain from Dehumidifier Pump Horsepower The chilled water temperature rise in the dehumidifier and the pump head must be approximated to use Table 60. 1. Large systems with considerable piping and fittings may require up to 100 ft pump head; normally, 70 ft head is the average. 2. The normal water temperature rise in the dehumidifier is between 7 F and 12 F. Applications using large amounts of water have a lower rise; those using small amounts of water have a higher rise. PERCENT ADDITION TO GRAND TOTAL HEAT The percent additions to the grand total heat to compensate for various external losses consist of heat and leakage gain to return air ducts, heat gain from the dehumidifier pump horsepower, and the heat gain to the dehumidifier and piping system. These heat gains can be estirpated as follows: 1. Heat and leakage gain to return air ducts, see above. 2. Heat gain from dehumidifier pump horsepower, Table 60. 3. Dehumidifier and piping losses: a. Very little external piping - 1% of GTH. b. Average external piping - 2% of GTH. c. Extensive external piping - 4% of GTH. 4. Blow-through fan system - add percent room sensible heat from Table 59 to GTH. 5. Dehumidifier in conditioned apparatus room reduce the above percentages by one half. I 1-115 CHAPTER 8. APPLIED PSYCHROMETRICS The lxcccclingchapters contztin the t o lxol)erly evnlu:tte the he:tting ant1 They also reconlnlcnd ottt(loor air ventilzttion l>url~oses i n :tre;ts where locnl cocks tlo not exist. 2 . Air u~rttli~ionir~~ :~)+wmtus - ktctors Afccting cot~t~non lxocesses xntl the clfcct oC these lactors on selection of xir contlitionitig ccluilxncrtt. prncticnl clata cooling loacls. quantities for state, city or II/ pa7~tinl lorrtl co77 trd - t h e clfect ol lxtrtial lo;~ct on ecluilxncnt s e l e c t i o n and on the cotntnon lxocesses. 3. P.ryrl~7~omeI~ic.r T h i s challter tlcscrilxx lxxticallxychrometrics as al~l~lietl to ~tl~lxttxttts selection. It is tlivitlctl into three parts: 1. Descliplior7 of m ’ w r s , p7wesses encoitnterec! cntions. and factors 7’0 hells rccognizc ternis, lactors z~ntl lxocesses dcscrilxxl in this chapter, a brief tlefinition of lxychronletrics is offeretl at this point, along with an illusttxtion aticl tlcfinition 0E terms nl>l>earing on ;1 stanclartl lxychrotnetric chart (Fig. 32). - as in nortnztl air contlitioning appli- rv.bulb Temperature -The temperature of air as registered by .n orditiazy thermometer. Specific dry air. Temperature -The temperature registered by a thermometer whose bulb is covered by a wetted wick and exposed to a current of rapidly moving air. Wet-bulb Temperature-The temperature at which tion of moisture begins when the air is cooled. Dewpoint Volume -The cubic feet of the mixture per pound of Sensible Heat Factor-The - Located at 80 F db and 50% rh and used in conjunction with the sensible heat factor to plot the various air conditioning process lines. Alignment Circle condensa- of Dry Air-The basis for all psychrometric calculations. Remains constant during all psychrometric processes. Pounds - Ratio of the actual water vapor pressure of the air to the saturated water vqpor pressure of the air at the same temperature. Relative Humidity The dry-bulb, wet-bulb, and clewpoint temperatures and the relative humidity are so related that, if two properties are known, all other properties shown may then be determined. When air is saturated, dry-bulb, wet-bulb, and dewpoint temperatures are all equal. or Moisture Content-The weight of water vapor in grains or pounds of moisture per pound of dry air. Specific Humidity Enthalpy - A thermal property indicating the quantity of heat in the air above an arbitrary datum, in Btu per pound of dry air. The datum for dry air is 0°F and, for the moisture content, 32 F water. - Enthalpy indicated above, for any given condition, is the enthalpy of saturation. It should be corrected by the enthalpy deviation due to the air not being in the saturated state. Enthalpy deviation is in Btu per pound of dry air. Enthalpy deviation is qplied where extreme accuracy is required; how. er, on normal air conditioning estimates it is omitted. Entholpy Deviation Dry-Bulb FIG. 32 ratio of sensible to total heat. -SKELETON Temperature PSYCHROMETRIC CHART PSYCHROMETRIC CHART Normal Temperatures AIR CONDITIONING PROCESS I. RETURN AIR FROM THE ROOM @ IS MIXED WITH OUTDOOR 2 THIS AIR M,XT”RE @ REOUIRED OF OUTDOOR FOR AND VENTILATION. RETURN AIR ENTERS THE APPARATUS @ WHERE IT IS CONDITIONED 3. T,,EN THE TO A,R @ CYCLE AND IS Flc. SUPPLIED TO REPEATED AGAIN. 33 THE SPACE - TYI~ICAL 0. AIR CONDITIONING P ROCESS T RACED ON A S TANDARD PSYCHROMETRIC . CHART l-117 CH,\I’-I‘EK 8 . ,\I’I’LIED I’SYC:HKO~ll:.~T’liI(:S DEFINITION Psychrometrics is the science involving thermodynamic properties of moist air and the effect of atmospheric moisture on materials and human comlort. it applies to this chapter, the definition must be broadened to include the method ol controlling the thermal properties of moist air. AIR CONDITIONING PROCESSES’ Fig. 33 shows a typical air conditioning process traced on a psychrometric chart. Outdoor air (2)* is mixed with return air from the room (I) and enters the apparatus (3). Air flows through the conditioning apparatus (3 - 4) and is supplied to the space (4). The air supplied to the space moves along line (4 - 1) as it picks up the room Ioads, and the cycle is re- peated. Normally most o[ the air supplied to the space by the air conditioning system is returned to the conditioning apparatus. There, it is lnixetl with outdoor air required Lor ventilation. The mixture then passes tliru tile apparatus where heat and moisture are added or removed, as required, to maintain the desired conditions. The selection of proper equipment to accomplish this conclitioning and to control the thcrmotlynanlic p r o p e r t i e s ot the air depends upon a variety ofelements. However, only those which affect the psychromctric properties of air will bc discussed in this chapter. These elements are: room sensible heat factor (RSHFj’)t , grand sensible heat factor (GSHF), effective surface temperature (tCJ, bypass factor (UF), and effective sensible heat factor (UHF). DESCRIPTION OF TERMS, PROCESSES AND FACTORS SENSIBLE HEAT FACTOR The thermal properties of air can be separated into latent and sensible heat. The term sensible heat factor is the ratio of sensible to total heat, where total heat is the sum of sensible and latent heat. This ratio may be expressed as: SHF= where: SHF SH LH TH R’ = = = = SH SH =SHfLH T H sensible heat factor sensible heat latent heat total heat Fig. 34. This line represents the psychrometric process of the supply air within the conditioned space and is called the room sensible heat factor line. The slope of the RSHF line illustrates the ratio of sensible to latent loads within the space and is illustrated in Fig. 34 by ~h,~ (sensible heat) and AA, (latent heat). Thus, if adequate air is supplied to offset these room loads, the room requirements will M SENSIBLE HEAT FACTOR (RSHF) The room sensible heat factor is the ratio of room sensible heat to the summation of room sensible and room latent heat. This ratio is expressed in the following formula: RSHF = RSH RSH RSH+RLH=RTH The supply air to a conditioned space must have the capacity to offset simultaneously both the room sensible and room latent heat loads. The room and the supply air conditions to the space may be plotted on the standard psychrometric chart and these points connected with a straight line (1 - 2), *One italic numljer in parentheses represents a point, and two italic numl,ers in parentheses represent a line, plotted on the accompanying psychrometric chart examples. , DRY-BULB TEMPERATURE / FIG. M - RSHF LINE P LOTTED BETWEEN ROOM :$ AND S UPPLY A IR CONDITIONS tRefer to page 119 for a tlescription of all al)lneviations :yml>ols tlsecl in this chapter. ant1 l-118, l’.\l<‘I‘ I . LO,\11 ES~I‘IM.\‘I‘IN~; L 1x2 satislicd, !)rovitl(Yl I)otll tllc dry- and wet-!)ulh temperatures ol the supply air fall on this line. T h e ro0111 scnsiblc heat Iactor line cai1 a l s o bc drawn 011 the psychrometric chart without knowing the condition ol supply air. The lollowing procetlure illustrates how to plot this line, using the cnlculated RSHF, the room clesign contlitions, the scnsible heat factor scale in the upper right ham! corner of the psychrometric chart, and the alignment circle at 80 F dry-!)ulb and 5070 relative humidity: 1. Draw a base line thru the alignment circle and the calculated RSHF s1~ow11 on the sensible heat factor scale in the upper right corner of psychromctric chart (I - _3J, rig. 35. 2. Draw the actual room sensible heat factor line thru the room design conditions parallel to the base line in .Step 1 (3 - -/), Fig. 35. As shown, this line may be drawn to the saturation line 011 the psychrometric chart. of the air cnterillg the apparatus (mixture condition ol outdoor and return room air) ant! the contlition of the air leaving tlie apparatus may !,e plotted on the psychrometric chart anr! connected by a straight litic (/ 21, /‘ix. 36. T h i s line rcl)rcscnts tile psycllt-ometric process ol the air as it passes through the conditioning apparatus, ant! is relerred to as the grand scnsi!)lc heat Iactor line. ‘I-lie slope 0C the GSI-11; l i n e represents tlic ratio oC sensible anrl latent heat that the apparatus must hantlle. This is illustrated in Pig. 36 !)y A//,? (sensible heat) ;1nc1 Ah, (Iatcnt lieatj. / DESIGN FROM APPARTUS DRY-BULB TEMPERATURE . CALCULATEC RSHF FIG. 36 - GSHF LINE PLOTTED BETWEEN MIXTURE CONI~ITIONS T O A PPARATUS AND L E A V I N G C O N D I T I O N F R O M A PPARATUS I / SOFdb FIG, 35 - RSHF LINE PLOTTED PSYCHROMETRIC ON S KELETON CHART GRAND SENSIBLE HEAT FACTOR (GSHF) The grand sensible heat factor is the ratio of the total sensible heat to the grand total heat load that the conditioning apparatus must handle, including the outdoor air heat loads. This ratio is determined from the following equation: GSHF = TSH TSH T L H + T S H =GTH Air passing thru the conditioning apparatus increases or decreases in temperature and/or moisture content. The amount of rise or fall is determined by the total sensible and latent heat loads that the conditioning apparatus must handle. The condition The grand sensible heat factor line can be plotted on the psychrometric chart without knowing the condition of supply air, in much the same manner as the RSHF line. Fig. 3i, Step I (I - 2) and Step 2 (3 -4) show the procedure, using the calculated GSHF, the mixture condition of air to the apparatus, the sensible heat factor scale, and’the alighment circle on the psychrometric chart. The resulting GSHF line is plotted thru the mixture conditions of the air to the apparatus. REQUIRED AIR QUANTITY The air quantity required to offset simultaneously the room sensible and latent loads and the air quantity required thru the apparatus to handle the total sensible and latent loads may be calculated, using the conditions on their respective RSHF and GSHF lines. For a particular application, when both the RSHF and GSHF ratio lines are plotted 011 the psychrometric chart, the intersection of the two lines (I) Fig. 38, rcprescnts the condition ol‘ the supply air to CHAPTEK l-119 8 . ,\I’I’t,IEI> I’SY(:HIIOIVIE’I‘I~I(:S these supplementary loads are considered in plotting the RSHF and GSHF lines. Point (I) is the condition of air lcaving the apparatus and point (2) is the condition of supply air to the space. Line (1 - 2) represents the temperature rise of the air stream resulting from fan horsepower and heat gain to the duct. OUTDOOR DESIGN / DRY-BULB r IG. I MIXTURE / CONDITION TO APPARATUSt, I c is I 1 BOF TEMPERATURE 37 - GSHF L INE P LOTTED PSYCHROMETRIC ON A-1CONDlTlON OF OF AIR AIR LEAVING LEAVING APPARATUSt APPARATUSt S KELETON 2 8 % f/&,1 CH A R T DRY-BULB 1 I TEMPERATURE * OUTDOOR FIG. 39 - RSHF WITH AND GSHF L INES P LOTTED S UPPLEMENTARY L OAD L INE The air quantity required to satisfy the room load may be calculated from the following equation: cfm,, = ROOM AND AIR LEAVING APPARATUS DRY-BULB FIG. 38 - RSHF S KELETON AND I TEMPERATURE GSHF L INES P LOTTED P SYCHROMETRIC RSH 1.08 (Ln - LJ The air quantity required thru the conditioning apparatus to satisfy the total air conditioning load (including the supplementary loads) is calculated from the following equation: ON CHART the space. It is also the condition of the air leaving the apparatus. This neglects fan and duct heat gain, duct leakage losses, etc. In actual practice, these heat gains and losses are taken into account in estimating the cooling load. Chapter 7 gives the necessary data for evaluating these supplementary loads. Therefore, the temperature of the air leaving the apparatus is not necessarily equal to the temperature of the air supplied to the space as indicated in Fig. 38. Fig. 39 illustrates what actually happens when Cfmda = TSH 1.08 (L - trd The required air quantity supplied to the space is equal to the air quantity required thru the apparatus, neglecting leakage losses. The above equation contains the term t,,, which is the mixture condition of air entering the apparatus. With the exception of an all outdoor air application, the term t, can only be determined by trial and error. One possible procedure to determine the mixture temperature and the air quantities is outlined below. This procedure illustrates one method of apparatus selection and is presented to show how cumbersome and time consuming it may be. A l-120 1. I’:\RT I . LOAD a rise (trm - t,,J in the supply air to the and calculate the supply air quantity (cfm,,) to the space. Assume SpaCC, 2. Use this air quantity to calculate the mixture condition of the air (t,,,) to the space, (Equation 1, p~lge 150). 3. Substitute this supply air quantity and mixture condition of the air in the formula for air quantity thru the apparatus (cfm,,) and determine the leaving condition of the air from the conditioning apparatus (t,,,). 4. The rise between the leaving condition from the apparatus and supply air condition to the space (L - tl,,) must be able to handle the supplementary loads (duct heat gain and fan heat). These temperatures (t,,,, t,,) may be plotted on their respective GSHF and RSHF lines (Fig. 39) to determine if these conditions can handle the supplementary loads. If they cannot, a new rise in supply air is assumed and the trial-and-error procedure repeated. In a normal, well designed, tight system this difference in supply air temperature and the condition of the air leaving the apparatus (t,, - t,& is usually not more than a few degrees. To simplify the discussion on the interrelationship of RSHF and GSHF, the supplementary loads have been neglected in the various discussions, formulas and problems in the remainder of this chapter. It can not be overemphasized, however, that these supplementary loads must be recognized when estimating the cooling and heating loads. These loads are taken into account on the air conditioning load estimate in Chapter 1, and are evaluated in Chapter 7. The RSHF ratio will be constant (at full load) under a specified set of conditions; however, the GSHF ratio may increase or decrease as the outdoor air quantity and mixture conditions are varied for design purposes. As the GSHF ratio changes, the supply air condition to the space varies along the RSHF line (Fig. 38). The difference in temperature between the room and the air supply to the room determines the air quantity required to satisfy the room sensible and room latent loads. As this temperature difference increases (surlplying colder air, since the room conditions are fixed), the required air quantity to the space decreases. This temperature difference can increase up to a limit where the RSHF line crosses the saturation line on the psychrometric chart, Fig. 38; assuming, of course, that the available conditioning equipment is able to take the air to 100% ES’I’IMATING saturation. Since this is impossible, the condition of the air normally falls on the RSHF line close to the saturation line. How close to the saturation line depends on the physical operating characteristics and the efficiency of the conditioning equipment. In determining the required air quantity, when neglecting the supplementary loads, the supply air temperature is assumed to equal the condition of the air leaving the apparatus (tY,l - tldb). This is illustrated in Fig. 38. The calculation for the required air quantity still remains a trial-and-error procedure, since the mixture temperature of the air (t,,,) entering the apparatus is dependent on the required air quantity. The same procedure previously described for determining the air quantity is used. Assume a supply air rise and calculate the supply air quantity and the mixture temperature to the conditioning apparatus. Substitute the supply l air quantity and mixture temperature in the equation for determining the air quantity thru the apparatus, and calculate the leaving condition of the air frbm the apparatus. This temperature must equal the supply air temperature; if it does not, a new supply air rise is assumed and the procedure r e p e a t e d . Determining the required air quantity by either method previously described is a tedioui process, since it involves a trial-and-error procedure, plotting the RSHI; and GSHF ratios on a pspchrometric chart, and in actual practice accounting for the supplementary loads in determining the supply air, mixture and leaving air temperatures. This procedure has been simplified, however, by relating all the conditioning loads to the physical performance of the conditioning equipment, and then including this equipment performance in the actual calculation of the load. This relationship is generally recognized as a psychrometric correlation of loads to equipment performance. The corrtilation is accomplished by calculating the “effective surface temperature,” “bypass factor” and “effective sensible heat factor.” These alone will permit the simplified calculation of stiipply air quantity. EFFECTIVE SURFACE TEMPERATURE (fJ The surface temperature of the conditioning equipment varies throughout the surface of the apparatus as the air comes in contact with it. However, the effective surface temperature can be considered to be the uniform surface temperature which would produce the same leaving air conditions as the nonuniform surface temperature that actually occurs when the apparatus is in operation. This is more clearly understood by illustrating the heat transfer effect between the air and the cooling (or heating) medium. Fig. 40 illustrates this process and is applicable to a chilled water cooling medium with the supply air counterflow in relation to the chilled dewpoint (adp). The term is used exclusively in this chapter when relcrring to cooling and dchumidifying applications. The psychrometrics of air can be applied equally well to other types of heat transfer applications such as sensible heating, evaporative cooling, sensible cooling, etc., but for these applications the effective surface temperature will not necessarily fall on the saturation line. BYPASS FACTOR (BF) Bypass factor is a function of the physical and operating characteristics of the conditioning apparatus and, as such, represents that portion of the air which is considered to pass through the conditioning apparatus completely unaltered. SURFACE AREA FIG. 40 - RELATIONSHIP OF E FFECTIVE SURFACE TEMP TO SUPPLY A IR AND CHILLED W ATER The relationship shown in Fig. 40 may also be illustrated for heating, direct. expansion cooling and for air *flowing parallel to the cooling or heating medium. The direction, slope and position of th,e lines change, but the theory is identical. Since conditioning the air thru the apparatus reduces to the basic principle of heat transfer between the heating or cooling media of the conditioning apparatus and the air thru that apparatus, there must be a common referen$e point. This point is the effective surface temperature of the apparatus. The two heat traiisfers are relatively independent of e2-h other, but are quantitatively equal when re1. d to the effective surface temperature. Therefore, to obtain the most ‘economical apparatus selection, the effective surface temperature is used in calculating the required air quantity and in selecting the apparatus. For applications involving cooling and dehumidification, the effective surface temperature is at the point where the GSHF line crosses the saturation line on the psychrometric chart (Fig. 36). As such, this effective surface temperature is considered to be the dewpoint of the apparatus, and hence the term apparatus dewpoint (adp) has come into common usage for cooling and dehumidifying processes. Since cooling and dehumidification is one of the most common applications for central station apparatus, the “Air Conditiouing Load Estimate” form, Fig. 44, is designed around the term apparatus The physical and operating characteristics affecting the bypass factor are as follows: 1. A decreasing amount of available apparatus heat transfer surface results in an increase in bypass factor, i.e. less rows of coil, less coil surface area, wider spacing of coil tubes. 2. A decrease in the velocity of air through the I conditioning apparatus results in a decrease in bypass factor, i.e. more time for the air to contact the heat transfer surface. Decreasing or increasing the amount of heat transfer surface has a greater effect on bypass factor than varying the velocity of air through the apparatus. There is a psychrometric relationship of bypass factor to GSHF and RS$IF. Under specified room, outdoor design conditions and quantity of outdoor air, RSHF and GSHF are fixed. The position of RSHF is also fixed, but the relative position of GSHF may vary as the supply air quantity and supply air condition change. To properly maintain room design conditions, the air must be supplied to the space at some point along the RSHF lihe. Therefore, as the bypass factor varies, the relative position of GSHF to RSHF changes, as shown by the dotted lines in Fig. 41. As the position of GSHF changes, the entering and leaving air conditions at the apparatus, the required air quantity, bypass factor and the apparatus dewpoint also change. The effect of varying the bypass factor on the conditioning equipment is as follows: 1. Smaller bypass factor a. Higher adp - DX equipment selected for higher refrigerant temperature and chilled water equipment would be selected for less b&Y or higher temperature chilled water. Possibly smaller refrigeration machine. l-122 l’,jRT 1,. Less air - smaller fan and fan motor. c. More heat transfer surface - more rows of coil or more coil surface available. d. Smaller piping if less chilled water is used. 2. Larger bypass factor a. Lower adp - Lower refrigerant temperature to select DX equipment, and more water or lower temperature for chilled water equipment. Possibly larger refrigeration machine. b. Marc air -larger fan and fan motor. c. Less heat transfer surface - less rows of coil or less coil surface available. d. Larger piping if more chilled water is used. I. LOAD ESTIMATING rically to the bypass factor. Although it is recognized c.hat bypass factor is not a true straight line function, it can be accurately evaluated mathematically from the following equations: BF= ’ Idb - hdp t edb - &do fl,a = - hndp he, wla - Wadp - hadp = $,a - wad, and NOTE: The quantity (I--UF) is frequently called contact factor and is considered to be that portion of the air leaving the apparatus at the adp. EFFECTIVE SENSIBLE HEAT FACTOR (ESHF) OUTDOOR DESIGN / AIR DRY-BULB FIG . To relate bypass factor and apparatus dewpoint to the load calculation, the eflective sensible heat factor term was developed. ESHF is interwoven with BF and adp, and thus greatly simplifies the calculation of air quantity and apparatus selection. The effective sensible heat factor is the ratio of effective room sensible heat to the effective room sensible and latent heats. Effective room sensible heat is composed of room~sensible heat (seePSHF) plus that portion of the outdoor air<&ible load which is considered a’s being bypassed, unaltered, thru the conditioning apparatus. The effective room latent heat is composed of the room latent heat (see RSHF) plus that portion of the outdoor air latent heat load which is considered as being bypassed, unaltered, thru the conditioning apparatus. This ratio is expressed in the following formula: LEAVING TEMPERATURE 41 - RSHF AND GSHF LINES P LOTTED SKELETON PSYCHROMETRICCHART ON It is, therefore, an economic balance of first cost ’ and operating cost in selecting the proper bypass factor for a particular application. Table 62, page 127, lists suggested bypass factors for various applications and is a guide for the engineer to proper bypass factor selection for use in load calculations. Tables have also been prepared to illustrate the various configurations of heat transfer surfaces and the resulting bypass factor for different air velocities. Table 61, pnge 127, lists bypass factors for various coil surfaces. Spray washer equipment is normally rated in terms of saturation efficiency which is the complement of bypass factor (1 - BF). Table 63, page 136, is a guide to representative saturation efficiencies for various spray arrangements. As previously indicated, the entering and leaving air conditions at the conditioning apparatus and the apparatus dewpoint are related psychromet. / ESHF = ERSH ERSH + ERLH ERSH =ERTH The bypassed outdoor air loads that are included in the calculation of ESHF are, in effect, loads imposed on the conditioned space in exactly the same manner as the infiltration load. The infiltration load comes thru the doors- and windows; the bypassed outdoor air load is supplied to the space thru the air distribution system. Plotting RSHF and GSHF on the psychrometric chart defines the adp and BF as explained previously. Drawing a straight line between the adp and room design conditions (1 - 2), Fig. 42 represents the ESHF ratio. The interrelationship of RSHF and GSHF to BF, adp and ESHF is graphically illustrated in Fig. 42. The effective sensible heat factor line may also be drawn on the psychrometric chart without initialIy knowing the adp. The procedure is identical to the one described for RSHF on @ge IIN. The cal- l 1-123 (:H/\I'TEK 8. .\l'l'Ll131) I'SYC:l-IROILIE'I‘I~I<:S c&ted ESHF, however, is plotted thru the room design conditions to the saturation line (I - 2), I;ig. 43, thus indicating the adp. Tables have been prepared to simplify the method of determining adp from ESHF. Adp can be obtained by entering Table 65 at room design conditions and at the calculated ESHF. It is not necessary to plot ESHF on a psychrometric chart. AIR QUANTITY USING ESHF, ADP AND BF A simplified approach for determining the required air quantity is to use the psychrometric corre- OUTDOOR lation of effective sensible heat factor, apparatus dewpoint and bypass factor. Previously in this chapter, the interrelationship of ESHF, BF and adp was shown with GSHF and RSHF. These two factors need not be calculated to determine the required air quantity, since the use of ESHF, BF and adp results in the same air quantity. The formula for calculating air quantity, using BF and tndp, is: Cf%a = ERSH 1.08 (trm - tad (1 - I-3 (ESHF is used to determine tadp.) This air quantity simultaneously offsets the room sensible and room latent loads, and also handles the total sensible and latent loads for which the conditioning apparatus is designed, including the outdoor air loads and the supplementary loads. AIR CONDITIONING LOAD ESTIMATE FORM FIG. 42 - RSHF, GSHF AND ESHF LINES PLOTTED ONSKELETON PSYCHROMETRIC CHART The “Air Conditioning Load Estimate” form is designed for cooling and dehumidifying applications, and may be used for psychrometric calculations. Normally, only ESHF, BF and adp are required to determine air quantity and to select the apparatus. But for those instances when it is desirable to know RSHF and GSHF, this form is designed so that these factors may also be calculated. Fig. 44, in conjunction with the following items, explains how each factor is calculated. (The circled numbers correspond to numbers in Fig. 44.) ._ 1. RSHF = RSH 0 =0+0 RSH + RLH 2. G S H F = = = @+O GTH 3. ESHF = 0 ERSH ERSH.+ E R L H ERSH =ERTH 4. Adp located where ESHF crosses the saturation line, or from Table 65. ESHF @ and room conditions @ give adp @. SOF DRY-BULB FIG. TEMPERATURE 43-ESHF LINEPLOTTEDONSKELETON PSYCHROMETRICCHART 5. BF $J used in the outdoor air calculations is obtained from the equipment performance table or charts. Typical bypass factors for different surfaces and for various applications are given on page 127. These are to guide the engineer and may be used in the outdoor air calculation when the actual equipment per- . ’ formance tables are not readily available. I 1-124 P.411'I‘ 1. LO,\11 ES'I‘IM.\'I‘ING i DATE----~. SHEET P R E P A R E D NAME OF -yOFFlCE BY-- PROP NO B ------JOaNO. JOB L O C A T I O N - - - SPACE USED FOR S o FT % ..-‘LISS GLASS _ So FT GLASS SKVLIGHT X So FT x WALL s a FT x x WALL Sa FT y x WALL SQ FT x x Sa FT X x Sa FT X x PLOPLE JLL G LASS S o FT x x JmITION Sa FT x x _CgLING Sa FLOOR SP FT x Fr -~ CFM OUTDOOR AIR EFFECTlYE SENS HEIT = FACTOR ._ HPoaKW X HE/T SELECTED ADP = AIR .._ -.. F QUANTITY EFFECTIYE R O O M SENS. H E A T __v.,w-~CFM~~ ____1.08 x - - 0 - - - W A T T S x 1.4 Y LlGHTS -..-.---.-CFMo~ 0 (1 --RF) X tT,,-&F - T,,,v.@vF) = -F 8 POWER ..- F DEHUMIDIFIED INTERNAL HEAT PEOPLE q APPARATUS EFFECTIVE R O O M T O T A L H E A T = .-- 0 INDICATED ADP = 1.08 x THRU APPARATUS DEWPOINT EFFECT,YE O O M SENS. 8 --.-- R _-... .- -x PEOPLE ADDITION&L = --_ - x x cm x APPLIANCES. lxxx! .-x.-- TRANS. GAIN-EXCEPT WALLS b ROOF INFlLTRATlON lxxx OUTDOOR AIR x -CFY,PLIJOH iv ROOF WALL ROOF-SHADED lxxx -.X Y s o FT Y &_OF-SUN I x SOLAR & TRANS. GAIN-WALLS - YCE -x- SC) F T x - x @ F TLMP ~1st ROOW SENT. H E A T - ~-= --F~I+,UTLETLII~,. x @ CFM D. - 1.08 x ETC . SUPPLY AIR QUANTITY x HE A T G A I N S R O O M SCNS. H E A T - -52 SUB TOTA- 1.08 x F DESIRED - - =--CFM5, ~,r, 43 8 PCFM 5* - -----CFMDA =-.-CFMo* EDa LDS RESULTING ENT t LVG CONDITIONS AT APPARATUS @ CFM,,& @ - T,,-F) 0 f------ X CT,,-F = T,,,-F 0 T,,-F @oO.@cFM+ @ 0 T,,--.-Ff-aF FROhi LATENT HEAT CFM X lNF!LTRATlON PEOPLE PlOPLE STEAM LB,“” APPLIAHCES. Aoo,r,ow~~ G”,Lrn CHART: x 0.m @ X (T,D,-F TAD-F) = T,,,-F T,,.- F. T LWF----F NOTES x x 1030 E TC . HE A T G A I N S so FT x lilO0 VAPOR TRINS. x GIiLB x ___-- SUB T SAFETY PSYCH. F ACTOR O T A L % ROOM LATENT HEAT S U P P L Y Ourooo~ D U C T L E A K A G E EFFECTlVE E F F E C T I V E SENSIBLE: LATENT: RLT”l” DUCT “CAT Gl,H L E . - p - % GR,LB CFH x AIR x 0 BF x 0.68 R O O M T O T A L H E A T n OUTDOOR AIR HEAT Fx(,---@aF)Xl.OS. CFM x CFM x RET”“” DUCT % + LLLS. Gil” p ,@ ROOM LATENT HEAT G”,UX (,- II HP ?/. + PUMP G R A N D BF) / /@ ,-l-x Y 0.6&-AL;k+y) SUB T OTAL DC”““. * a/& + PlPL LOIS %./@ T O T A L H E A T I Form E 20 NOTE: The circled numbers are explained on the previous page under “Air Conditioning Load Estimate” FIG. 44 -AAIR~ONDITIONING LOAD ESTIMATE form. , l-125 CHAI’~I‘IIlI 8 . .\I’I’LIED I’SY(;~lIIO~lE’I‘KI(:S leaving air conditions are easily determined. The calculations for the entering and leaving dry-bulb temperatures at the apparatus are illustrated in Fig. 44. ERSH 6. cfm,,,= ___ 1.08 (r,, - cd (1 - B6 The entering dry-bulb calculation contains the This air quantity “cfmt” determ “cfmt”‘. pends on whether a mixture of outdoor and return air or return air only is bypassed around the conditioning apparatus. ( Once the dehumidified air quantity is calcu\ lated, the conditioning apparatus may be se’ lected. The usual procedure is to use the grand total heat @ , dehumidified air quantity :a, and the apparatus dewpoint @$ , to 4 select the apparatus. The total supply air quantity cfm,q, @ is used for “cfmt” when bypassing a mixture of outdoor and return air. Fig. 45 is a schematic sketch of a system bypassing a mixture of outdoor and return air. Since guides are available, the bypass factor of the apparatus selected is usually in close agreement with the originally assumed bypass factor. If, because of some peculiarity in loading in a particular application, there is a wide divergence in bypass factor, that portion of the load estimate form involving bypass factor should be adjusted accordingly. 7. Outlet temperature difference - Fig. 44 shows a calculation for determining the temperature difference between room design dry-bulb and the supply air dry-bulb to the room. Frequently a maximum temperature difference is established for the application involved. If the outlet temperature difference calculation is larger than desired, the total air quantity in the system is increased by bypassing air around the conditioning apparatus. This temperature difference calculation is: Outlet temp diff = = 8. CONDlTlONEO ( BYPASSING - = The amount of air that must be bypassed around the conditioning apparatus to maintain this desired temperature difference (At) is the difference between cfm,, and cfmda. 9. Entering and leaving conditions at the apparatus - Often it is desired to specify the selected conditioning apparatus in terms of entering and leaving air conditions at the apparatus. Once the apparatus has been selected from ESHF, adp, BF and GTH, the entering and f FAN ENT CONOITIONING COND FIG . 45 - BYPASSING MIXTURE RETURN OF OUTDOOR AND AIR When bypassing a mixture of return air only or when there is no need for a bypass around the apparatus, use the cfmda @ for the value of “cfmt”. Fig. 46 is a schematic sketch of a system bypassing room return air only. 0 x @ RSH “\ @ 1.08 X At ?’ 1.08 X At c 1 Total air quantity when outlet temperature difference is greater than desired - The calculation for the total supply air quantity for a desired temperature difference (between room and outlet) is: cfm,, . i.1 ’ t RSH 1.08 x cfmda 1.08 SUPPLY AIR MIXTURE OF OUTDOOR AN0 RETURN AIR CONDlTlONED SPACE . SUPPLY AIR BYPASSING RETURN AIR - t FAN OUTDOOR NR - FIG . ENT CON0 CONDITIONING APPARATUS 46 - BYPASSING RETURN A IR ONLY OR No FIXED BYPASS *“cfmt” is a symbol appearing in the equation next to 0 in Fig. 44. I ’ 1-126 I’AK’I I. LOAD ESTIMATING this point. (This point delincs the intcrsection of the RSHF and GSHF as described previously.) The entering and leaving wet-bulb temperatures at the apparatus are determined on the standard psychrometric chart, once the entering and leaving dry-bulb temperatures are calculated. The procedure for determining the wet-bulb temperatures at the apparatus is illustrated in Fig. 47 and described in the following i terns: a. Draw a straight line connecting room design conditions and outdoor design conditions. b. The point at which entering dry-bulbcrosses the line plotted in Step a defines the entering conditions to the apparatus. The entering wet-bulb is read on the psychrometric chart. c. Draw a straight line from the adp @ to the entering mixture conditions at the apparatus (Step IJ.) (This line defines the GSHF line of the apparatus.) d. The point at which the leaving dry-bulb crosses the line drawn in Step c defines the leaving conditions of the apparatus. Read the leaving wet-bulb from the apparatus at CALCULATED LEAWNG DRY BULB TEMP CALCULATED ENTERING DRY BULB TEMP FIG. 47 - ENTERING AT AND L EAVING CONDITIONS APPARATUS AIR CONDITIONING APPARATUS The following section describes the characteristic pspchrometric performance of air conditioning equipment. Coils; sprays and sorbent dehumidifiers are the three basic types of heat transfer equipment required for air conditioning applications. These components may be used singly or in combination to control the psychrometric properties of the air passing thru them. The selection of this equipment is normally determined by the requirements of the specific application. The components must be selected and integrated to result in a practical system; that is, one having the most economical owning and operating cost. An economical system requires the optimum combination of air conditioning components. It also requires an air distribution system that provides good air distribution within the conditioned space, using a practical rise between supply air and room air temperatures. Since the only known items are the load in the space and the conditions to be maintained within . the space, the selection of the various components is based on these items. Normally, performance requirements are established and then equipment is selected to meet the requirements. COIL CHARACTERISTICS In the operation of coils, air is drawn or forced over a series of tubes thru which chilled water, brine, volatile refrigerant, hot water or steam is flowing. As the air passes over the surface of the coil, it is cooled, cooled and dehumidified, or heated, depending upon the temperature of the media flowing thru the tubes. The media in turn is heated or cooled in the process. The amount of coil surface not only affects the heat transfer but also the bypass factor of the coil. The bypass factor, as previously explained, is the measure of air side performance. Consequently, it is a function of the type and amount of coil surface and the time available for contact as the air passes thru the coil. Table 61 gives approximate bypass factors for various finned coil surfaces and air velocities. CHAI'TEK 1-127 8 . /~I'PI,IEI) I'SYCIiROMETKIC:S TABLE 6 1 -TYPICAL BYPASS FACTORS (For Finned Coils) DEPTH OF COILS WITHOUT SPRAYS SPRAYS; 8 lins/in. 14 fins/in. 300-700 300-700 8 fins/in. 14 fins/in. Velocity (rows) WITH (fpm) 300 -700 300 - 700 3 .42 - .55 .21 * .40 .22 - .38 .I0 - .23 4 .I9 - .30 .05 - .14 .12 - .23 .02 - .09 .I2 - .22 .os - .14 .03 - .lO 5 6 .08 - .18 .Ol - .06 .06 - .ll .Ol - ;05 8 .03 - .08 2 .Ol - .08 .02 - .05 *The bypass factor with spray-coils is decreased because spray provides more surface for contacting the air. the DRY-BULB These bypass factors apply to coils with s/s in. c . tubes and spaced on approximately 11/d in. centers. The values are approximate. Bypass factors for coils with plate fins, or for combinations other than those shown, should be obtained from the coil manufacturer. Table 61 contains bypass factors for a wide range of coils. This range is offered to provide sufficient latitude in selecting coils for the most economical system. Table 62 lists some of the more common applications with representative coil bypass factors. This table is intended only as a guide for the design engineer. TABLE 62-TYPICAL BYPASS FACTORS (For - COIL BYPASS - ’ CTOR Various Applications) TYPE OF APPLICATION EXAMPLE FIG . 48 - COIL P ROCESSES COIL PROCESSES Coils are capable of heating or cooling air at a , constant moisture content, or simultaneously cooling and dehumidifying the air. They are used to control dry-bulb temperature and maximum relative humidity at peak load conditions. Since coils alone cannot raise the moisture content of the air, a water spray on the coil surface must be added if humidification is required. If this spray water is recirculated, it will not materially affect the psychrometric process when the air is being cooled and dehumidified. Fig. 48 illustrates the various processes that can be accomplished by using coils. Sensible 0.30 to 0.50 A small total load or a load that is somewhat larger with a low sensible heat factor (high latent load). 0.20 to 0.30 Typical comfort application ~ Residence, with a relatively small total Small load or a low sensible heat factor with a somewhat larger Retail Shop, Factory load. 0.10 to 0.20 Typical 0.05 to 0.10 Applications with high internal sensible loads or requiring a large amount of outdoor air for ventilation. 0 to 0.10 All comfort outdoor air application. applications. I I I<esidence c I Dept. Store, Bank, Factory Dept. Store, Restaurant, Factory Hospital Operating Room, Factory TEMPERATURE Cooling The first process, illustrated by line (1 - Z), represents a sensible cooling application in which the heat is removed from the air at a constant moisture content. Cooling and Dehumidification Line (1 - 3) represents a cooling and dehumidification process in which there is a simultaneous removal of heat and moisture from the air. For practical considerations, line (1 - 3) has been plotted as a straight line. It is, in effect, a line that starts at point (I) and curves toward the saturation line below point (3). This is indicated by line (I - 5). Sensible Heating Sensible heating is illustrated by line (1 - 4); heat is added to the air at constant moisture content. PAR?’ 1. LOAD ESTIMATINd 1-128 COIL PROCESS EXAMPLES To better understand these processes and their variations, ;1 description of each with illustrated examples is presented in the following: (Refer to page 149 for tle finit’ 1011 of symbols and abbreviations.) Cooling and Dehumidification Cooling and dehumidification is the simultaneous removal of the heat and moisture from the air, line (1 - 3), Fig. 48. Cooling and dehumidification occurs when the ESHF and GSHF are less than 1.0. The ESHF for these applications can vary from 0.95, where the load is predominantly sensible, to 0.45 where the load is predominantly latent. RSH - 200.000 Btu/hr RLH - 50,000 Btu/hr Ventilation - 2,000 cfm,, Find: 1. 2. 3. 4. 5. 6. Outdoor air load (OATH) Grand total heat (GTH) Effective sensible heat factor (ESHF) Apparatus dewpoint temperature (tad,,) Dehumidified air quantity (cfm,,) Entering and leaving conditions at the apparatus (t e d b ’ et w b ' tldb’ hub) Solution: 1. OASH = 1.08 X 2000 X (95 - 75) = 43200 Btu/hr OALH = 68 X 2000 X (99 - 65) = 46,200 Btu/hr OATH = 43,200 -I- 46,200 = 89,400 Btu/hr (14) (‘5) (17) 2. TSH = 200,000 +,43,200 = 243,200 Btu/hr TLH = 50,000 + 46,200 = 96,200 Btu/hr G T H = 243,200 + 96,200 = 339,400 Btu/hr The air conditioning load estimate form illus&rated in I’ig. 44 presents the procedure that is used .o determine the ESHF, dehumidified air quantity, and entering and leaving air conditions at the apparatus. Example 1 illustrates the psychrometrics involved in establishing these values. (7) (8) (9) l 3. Assume a bypass factor of 0.15 from Table 62. 200,000 + (.15) (43,200) ESHF = 200,000 + (.15) (43,200) + 50,000 f (.15) (46,200) = .785 (26) 4. Determine the apparatus dewpoint from the room design conditions and the ESHF, by either plotting on the psychrometric chart or using Table 65. Fig. 49 illustrates the ESHF plotted on the psychrometric chart. Example 1 - Cooling and Dehumidification Given: Application - 5P & lO$ Store Location - Bloomfield, N. J. t a@ =50F 1 NOTE: Numbers in parentheses at right edge of column refer to.,equations beginning on page 150. Summer design - 95 F db, 75 F wb Inside design - 75 F db, 50% rh OUTDOOR 75 Fdb 54.4 F db FIG.~~ -- COOLING AND 79.45 F db I 9 S F db DEHUMIDIFICATION (I:H/\I”I‘k:K 8 . /\I’l’I,lEl) l-129 I’SY~:l-lKO~I~‘1‘K1~:S 6. /\sstcmc lor this example that the 9,000 t fm, 50 F adp, and G’rII = factor that is equal, or nearly equal, 0.15. :\lso, assume that it is not I~ypass air around the ;ippara”Is, apl>aratus selected for 339,400, has a I)ypass lo the assumctl BF = necessary to physically ‘( t cd6 -- (2000 x 95) + (7000 x 75) (31) ,3. I:, I* (Ill 9000 Read tolo where the lcdb crosses tile straigtlt lint plottetl I)etuvcctl tllc 0ut(l0or and roan, tlcsign contlitions on tllc psychromctric chart, Fig. t ewh I’). = f35.5 F WI) Ild6 = 50 + .I5 (79.45 - 50) = 54.4 F tll, (32) lktcrminc the tizob I)y drawing a straight line between the atlp and the entering conditions at the apparatus. (This is the GSHF line.) Whcrc lldh intersects this lint, read tlwb. t ltuh Dehumidification Example - High Latent Load cIli some applications a special situation exists if the ESHF and GSHF lines do not intersect the saturation line when plotted on the psychrometric chart or if they do the adp is absurdly low. This may occur where the latent load is high with respect to the total loads (dance halls, etc.). In such applications, an appropriate apparatus dewpoint is selected and the air is reheated to the RSHF line. Occasionally, altering the room design conditions eliminates the need for reheat, or reduces the quantity of reheat required. Similarly, the utilization of a large air side surface (low bypass factor) coil may eliminate the need for reheat or reduce the required reheat. Once the ventilation air requirement ,is determined, and if the supply air quantity is not fixed, the best approach to determining the apparatus dewnoint is to assume a maximum allowable temPel. .re difference between the supply air and the room. Then, calculate the supply air conditions to the space. The supply air conditions to the space must fall on the RSHF line to properly offset the sensible and latent loads in the space. There are four criteria which should be examined, to aid in establishing the supply air requirements to the space. These are: 1. Air movement in the space. 2. Maximum temperature difference between the supply air and the room. 3. The selected adp should provide an economical refrigeration machine selection. 4. In some cases, the ventilation air quantity required may result in an all outcloor air application. 2 - Cooling and Dehumidification - High L a t e n t load Givcll: i\pplication - Lai)oramry Location - Bangor, Maine Summer design - 90 F (II), 73 F wl) Insitlc &sign - 75 1: (II). .50C,‘:, rh RSH - 120,000 Iltr+r RLII - 65,000 Btu/hr Ventilation - 2,500 cfj?l,, Temp. tliff. Ixtwccn room and supply air, 20 F maximum Find: I. Outdoor air load (O:\TH) 2. Efkctivc scnsiblc heat factor (ESHF) 3. = 52 . j F WI) Cooling and Ar ‘:ation Extrtnple _Q is ;I laboratory application with a high latent load. In this example the ESHF intersects the saturation line, but the resulting atlp is too low. /\pparatL!S LkwpOin~ (t&p) 4. Reheat required 5. Supply air quantity (cfttlsa) 6. Entering Conditions to C o i l (t&b, telob, 6ven) 7. Leaving conditions from coil (tldb, tllub) 8. Supply air condition to the space (fsn, 1%) 9. Grand total heat (GTH) Solution: 1. OASH = I.08 X 2.500 X (90.75) = 40,500 Btu/hr OALH = .68 X 2500 X (95-65) = 51,000 Btu/hr Oi\l‘H = 40,500 f51,000 = 91,500 Btu/hr 2. r\ssume (‘4) (15) (17) a bypass factor of 0.05 because of high latent load. 120,000 + (.05) (40,500) 120,000 + (.05) (40,500) + 65,000 + (.05) (51,000) = ,645 (26) ESHF = When plotted on the psychrometric chart, this ESHF (.G45 intersects the saturation curve at 35 F. With such a low adp an appropriate apparatus dewpoint s h o u l d he selected and the air reheated to the RSHF line. 3. Refer to Table 65. For inside design conditions of 75 F db, SOSr, rh, an ESHF of .74 results in an adp of 48 F which is a reasonable minimum figure. 4. Determine amount of reheat (Btu/hr) required to produce an ESHF of .74. ESHF (74) = 120,000 + .05 (40.500) + reheat 1 2 0 , 0 0 0 + .05 (40,500) + reheat + 65,000 + (.05) 51,000 ,74= 122,025 + reheat (25) 189,575 + reheat reheat = 70,230 Btu/hr 5. Determine clrhumidilicr air quantity (‘cf?!lda) ERSH cfmda = 1 .OY X (I - BF) (tr,,b - t”,,,,) 122,025 + 70,230 = = 6940 cfm 1.08 (1 - .05) (75-48) when no air is to be physically byCfm& iS ak0 Cfmsa passed around the cooling coil. G. t db = (2500 X 90) + (4440 X 75) e 6940 = 80.4 (31) SOTE: Numbers in parentheses at right edge of column refel to equa’ions beginning on fxzge 150. PART I. L O A D E S T I M A T I N G l-130 i Cooling and Dehumidification - Using All Outdoor Air I n some applications it may be necessary to sup- pIy a l l o u t d o o r air; for example, a hospital operating room, or an area that requires large quantities of ventilation air. For such applications, the ventilation or code requirements may be equal to, or more than, the air quantity required to handle the room loads. Items I thrz~ 5 explain the procedure for determining the dehumidified air requirements using the “Air Conditioning Load Estimate” form when all outdoor air is required. 1. Calculate the various loads and determine the apparatus dewpoint and dehumidified air quantity. 2. If the dehumidified air quantity is equal to the outdoor air requirements, the solution is self- . evident. Frc.50 - COOL~X ANDDEHUMIDIFICATION 'WITH HIGEI LATENT LOAD Read t,& where the t&b crosses the straight line plotted between the outdoor air and room design conditions on the psychrometric chart, Fig. 50. t,,b = 6G.6 F The moisture content at the entering conditions coil is read from the psychrometric chart. to the w,, = 75.9 gr/lll 7.Determine leaving conditions of air from cooling coil. hdb = tadp SBF (kdb - hdp) (32) = 48 f.05 (80.4 - 48) = 49.6 hsa = hadl, + BF (hw - hadp) (34) = 19.21 + .05 (31.3 - 19.21) = 19.82 t[,,,b = 49.1 F 8. Determine supply air temperature to space RSH tsa = hn - 1 .os (cfmsa) (120,000) = 75 - 1.08 (6940) =59F t,,, should also equal (35) t,,,,, + reheat 1.08 (Cf?f&) W,, =51.1 gr/lb between room and supply air =I,.~ - tsa = 75 - 59 = 16 F =<20 F 9. GTH = 4.45 x 6940 (31.3 - 19.82) = 354,500 Btu/hr 5. Use the recalculated outdoor air loads to determine a new apparatus dewpoint and dehumidified air quantity. This new dehumidified air quantity should check reasonably close to the cfmda in 1 tern 1. A special situation may arise when the condition explained in Item 4 occurs. This happens when the ESHF, as plotted on the psychrometric chart., does not intersect the saturation line. This situation is handled in a manner similar to that previously described under “Cooling and Dehumidification -High Latent Load Application.” Example 3 illustrates an application where codes specify that al1 outdoor air be supplied to the space. Example 7 0 2 3 0 = 4g’6 + 1.08 (6940) = 59 F Temp. difE 3. If the dehumidified air quantity is less than the outdoor air requirements, a coil with a larger bypass factor should be investigated when the difference in air quantities is small. If a large difference exists, however, reheat is required. This situation sometimes occurs when the application requires large exhaust air quantities. . 4. If the dehumidified air quantity is greater than the outdoor air requirements, substitute cfmda for cfm,, in the outdoor air load calculations. (24) 3 - Cooling ond Dehumidification All Outdoor Air Given: Application - Laboratory Location - Wheeling, West Virginia Summer design - 95 F db, 75 F ~11 Inside design - 75 F db, 55% rh RSH - 50,000 Btu/hr RLH - 11,000 Btu/hr Ventilation - 1600 cfm,, All outdoor air to be supplied to space. C H A P T E R 8. AI’I’LIED 1-131 I’SYCHKOMETKICS times referred to as a “split system.” The moisture is introduced into the space by using steam or electric humidifiers or auxiliary sprays. Find: I. 2. 3. 4. Outdoor air load (OATH) Elfective sensil)le heat factor (ESHF) Apparatus dcwpoint (tad,,) Dehumidified air quantity (cfm,,J When humidification is performed in the space, the room sensible load is decreased by an amount equal to the latent heat added, since the process is merely an interchange of heat. The humidifier motor adds sensible heat to the room but the amount is negligible and is usually ignored. 5. Recalculatetl outtloor air load (OATH) 6. Recalculated effective sensil)lc heat factor (UHF) 7. Final apparatus dewpoint temperature (tad& 8. Recalculated dehumidified air quantity (cfm& Solution: 1. OASH = 1.08 X 1600 X (95 - 75) = 34,600 Btu/hr OALH = 68 X 1600 X (98.5 - 71) = 30,000 Btu/hr O A T H = 3 4 , 6 0 0 + 30,000 = 64,600 Btu/hr ('4) (15) ('7) 2. Assume a bypass factor of 0.05 from Tnbles 61 and 62. ESHF = = 50,000+(.05)(34,600) 50,000 + (.05) (34,600) + 11,000 + C.05) (30,000) .81 W) 3. Table 65 shows that, at the given room design conditions 2nd effective sensible heat factor, tadp = 54.5 F. 50,000 +(.05)(34.600) 4. cf?nda= 1.08 (1 - .05) (75 - 54.5) = 2450 cfm (36) Since 2450 cfm is larger than the ventilation requirements, and by code all OA is required, the 0.4 loads, the adp, and the dehumidified air quantity m u s t be recalculated using 2450 cfm as the OA requirements. 5. Recalculating outdoor air load OASH = 1.08 X 2450 X (95 - 75) = 53,000 Btu/hr OALH = .68 X 2450 X (98.5 - 71) = 46,000 Btu/hr OATH = 53,000 + 46,000 = 99,000 Btu/hr (14) (15) (17) 50,000 + (.05)(53,000) (50,000) + (.05) (53,000) + 11,000 + (.05) (46,000) = .80 (26) 6. ESHF= 7. tadp =54F This checks reasonably close to the value in .ecalculation is not necessary. Cooling With Step 4, and Where humidification is required at design to reduce the air quantity, then a credit to the room sensible heat should be taken in the amount of the latent heat from the added moisture. No credit to the room sensible load is taken when humidification is usecl to make up a deficiency in the room latent load during partial load operation. When the humidifiers and sprays are used to reduce the required air quantity, the latent load introduced into the space is added to the room latent load. When the humidifier or sprays are operated only to make up the room deficiency, the latent load introduced into the room by the humidifier or auxiliary sprays in the space is not added to the ’ room latent load. The introduction of this moisture into the space to reduce the required air quantity decreases the RSHF, ESHF and the apparatus dewpoint. This method of reducing the required air quantity is normally advantageous when designing for high room relative humidities. The method of determining the amount of moisture necessary to reduce the required air quantity results in a trial-and-error procedure. The method is outlined in the following steps: 1. Humidification Cooling with humidification may be required at partial load operation to make up a deficiency in the room latent load. It may also be used at design conditions for industrial applications having relatively high sensible loads and high room relative humidity requirements. Without humidification, excessively high supply air quantities may be required. This not only creates air distribution problems but also is often economically unsound. Excessive supply air quantity requirements can be avoided by introducing moisture into the space to convert sensible heat to latent heat. This is someNOTE: Numbers in parentheses at right edge of column refer to equations beginning on page 150. 3 3. Assume an amount of moisture to be added and determine the latent heat available from this moisture. Table 64 gives the maximum moisture that may be added to a space without causing condensation on supply air ducts and equipment. Deduct this assumed latent heat from the orignal effective room sensible heat and use the difference in the following equation for ERSH to determine tndp. ndp = t,m - ERSH 1.08 x (1 - RF) cfmda is the reduced air quantity permissible in the air distribution system. Cfmda 3 . The ESHF is obtained from a psychrometric chart or Table 65, using the apparatus dewpoint (from Step 2) and room design conditions. P A R T 1. LOAD ESTIMA+lNG 1-132 4. The new effective room latent load is determined from the Lollowing equation: EKLH = EKSH x 3. Assume a bypass factor of 0.05 from Tables 61 and 62. I- ESHF ESHF ESHF The ERSH is from Step 2 and ESHF is from Step 3. 5. Deduct the original EKLH (before adding sprays or humidifier in the space) from the new effective room latent heat in Step 4. The result is equal to the latent heat from the added moisture, a n d must check with the value assumed in Step 1. If it does not check, assume another value and repeat the procedure. Example 4 illustrates the procedure for investigating an application where humidification is accomplished within the space to reduce the air quantity. lrxample 4 - Cooling With Humidification Given: Application - A high humidity chamber Location - St. Louis, Missouri Summer design - 95 F db, 78 F WI) Inside design - 70 F db, 70y0 rh RSH - 160,000 Btu/hr RLH - 10,000 Btu/hr RSHF - .94 Ventilation - 4000 cfrrzoa in the 4. Apparatus dewpoint (t& 5. Dehumidified air quantity (cfm,,) 6. Entering and leaving conditions at the apparatus tcwb’ 160.000+(.05)(108.000) = l?~~0~,~(.05)(108,000)+(.05)(109,000) = .92 (2~) 4. Plot the ESHF on a psychrometric chart and read the atlp (dotted line in Fig. 51). t a0 = 59.5 F 160,000 + (.05)(108,000) = 15 4oo cfm ’ (36) = (4000 x 95) i- (1’,400 x 70) = T6,7 F d,) 15,400 (31) cfrnda = 1.08(1-.05)(70-59.5) 5. 6. tedb Read teuzb where the tedb crosses the straight line plotted between the outdoor and room design conciitions on the psychrometric chart (Fig. 51). t elcb= 67.9 F wl, t ldb = 59.5 + .05 (76.7 - 59.5) = 60.4 F dl, (32) Determine the tlwb by drawing a straight line between the adp and the entering conditions to the apparatus (the GSHF line). Where tldb intersects this line, read the tlwb (Fig. 51). Space Find: A. When space humidification is not used: 1. Outdoor air load (OATH) 2. Grand total heat (GTH) 3. Effective sensible heat factor (ESHF) (tedb’ 2. GTH = 160,000 + 10,000 + 108,000 + 109,000 = 387,000 Btu/hr t Iwb = 60 F WI, B. When humidification is used in the space: 1. Assume, for the purposeof illustration in this problem, that the maximum air quantity permitted in the air distribution system is 10,000 cfm. Assume 5. grains of moisture per pound of dry air is to be added to convert sensible to latent heat. The latent heat is calcula.ied by m u l t i p l y i n g t h e a i r q u a n t i t y t i m e s t h e moisture added times the factor .68. 5 X 10,000 X .68 = 34,000 Btu/hr 2. New ERSH = Original ERSH - latent heat of added moisture = [lSO,OOO + (.05 X 108,000)] - 34,000 = 131,400 Btu/hr $db’ bob) B. When humidification is used in the space: 1. Determine maximum air quantity and assume an amount of moisture added to the space and latent heat from this moisture. 2. New effective room sensible heat (ERSH) 3. New apparatus dewpoint (tad& 4. New effective sensible heat factor (ESHF) 5. New effective room latent heat (ERLH) 6. Check calculated latent heat from the moisture added with amount assumed in Item 1. 7. Theoretical conditions of the air entering the evaporative humidifier before humidification. 8. Entering and leaving conditions at the apparatus edb’ tezabJ t,db’ $,ab) Solution: A. When space humidification is not used: 1. OASH = 1.08 X 4000 X ( 95-50) = 108,000 Btu/hr (14) OALH = 68 x 4000 x (117-77) = 109,000 Btu/hr (15) = 217,000 Btu/hr (17) OATH= 108,000 + 109,000 NOTE: Numbers in parentheses at right edge of column refer to equations beginning on page 150. 3. tadp 131,400 = ” - 1.08 (1 - .05) (10,000) = 57’2 F (36) 4 . ESHF is read from the psychrometric chart as .73 (dotted line in Fig. 52). > I- ESHF 5. New ERLH = New ERSH X ESHF 1 - .73 = 131,400 x ~ .73 = 48,600 Btu/hr 6. Check for latent heat of added moisture. Latent heat of added moisture = New ERLH - Original ERLH = 48,600 - [IO,000 + (.05 X 109,000)] = 33,200 Btu/hr This checks reasonably close in Step 2 (34,000 Btu/hr). i. Psychrometrically, it can ized water from the spray part of the room sensible vapor at the final room theoretical dry-bulb of the with the assumed value be assumed that the atomheads in the space absorbs heat and turns into water wet-bull, temperature. The air entering the sprays is , at the intersection of the anti the moisture content of ‘I‘his moisture content is the moisture atltletl I)y the design moisture content. room design wet-Ilull) line the air entering the sprays. tlcrerminetl I)y sril)tracting room sprays Irom the room Moisture content of air entering = 77 - 5 = 52 p/lb. humitlilier ‘I’hc theorctic;~l tlry-bull) i s tlctcrminctl from t h e pychromctric chart x i3.3 (II), illustratctl on I;ig. 52. 8. t,,d,, = (4000 x 9.5) + (6000 x i0) = 8. F (,,, 10,000 (31) where the Ipdh crosses the straight line Read l<,,,.b plotted between the outdoor and room design con. tlitions on the psychrometric chart (Fig. 52). QMFdb 70Fdb 76.7 F db 95 Fdb FIG. 51 -COOLING AND DEHUMIDIFICATION AI)DING No MOISTURETOTHE SPACE t F”, b t ldh = 69.8 I< WI) = 57.2 + (.05)(80 - 57.2) = 58.4 1: tll, (8’2) Determine tlu,* by drawing a straight line Ixtween the atlp and the entering conditions to the apparatus (GSHF line). Where tld,, intersects this line, read the llw,, (Fig. 52). t lwb = 58 F wh The straight line connecting the leaving conditions at the apparatus with the theoretical condition of the air entering the evaporative humidifier rcpresents the theoretical process line of the air. This theoretical condition of the air entering the humidifier represents what the room conditions are it’ the humidifier is not operating. The slope of this theorctical process line is the same as RSHF (.94). The heavy lines on Fig. 52 illustrate the theoretical air cycle as air passes through the conditioning apparatus to the evaporative humidifier, then to the room, and finally back to the apparatus where the return air is mixed with the ventilation air. Actually, if a straight line were drawn from the leaving conditions of the apparatus (58.4 F db, 58 F wb) to the room design conditions, this line would be the RSHF line and would be the process line for the supply air as it picks up the sensible and latent loads in the space (including the latent heat added by the sprays). The following two methods of. laying out the system are recommended when the humidifier is to be used for both partial load control and reducing the air quantity. 1. Use two humidifiers; one to operate continuously, adding the moisture to reduce the air quantity, and the other to operate intermittently to control the humidity. The humidifier used for partial load is sized for the effective room latent load, not including that produced by the other humidifier. If the winter requirements for moisture addition are larger than summer requirements, then the humidifier is selected for these conditions. This method of using two humidifiers gives the best control. F IG. 52 ADDING -COOLING MOISTURE DEHUMIDIFICATION INTO THE SPACE AND 2. Use one humidifier of sufficient capacity to handle the effective room latent heat plus the calculated amount of latent heat from the added moisture required to reduce the air quantity. In Part B, Step 5, the humidifier would be sized for a latent load of 48,600 Btu/hr. I’AK?‘ 1-134 Sensible Cooling A sensible cooling process is one that removes heat from the air at a constant moisture content, line (I - 2), Fig. 48. Sensible cooling occurs when either of the following conditions exist: 1. LOAD ESTIMATING Example 5 illustrates the method of determining the effective surface temperature for a sensible cooling application. Example 5 - Sensible Cooling Given: 1. The ESHF calculated on the air conditioning load estimate form is equal to 1.0. OT 2. The entering and leaving conditions at the apparatus, as checked or plotted on the psy chrometric chart, indicate a GSHF equal to 1 .O. In a sensible cooling application the ESHF, GSHF and RSHF all equal 1.0. When the RSHF equals 1.0, however, it does not necessarily indicate a sensible cooling process because latent load, introduced I outdoor air, can give a GSHF less than 1. The apparatus dewpoint is referred to as the effective surface temperature (tes) in sensible cooling applications. The effective surface temperature must be equal to, or higher than, the dewpoint temperature. In most instances, the t,, does not lie on the saturation line and, therefore, will not be the dewpoint of the apparatus. Whether or not t,, lies on the saturation line depends entirely on the bypass factor of the coil selected for the application. However, the calculations for ESHF, adp and cfmda may still be performed on the air conditioning load estimate form by substituting the term t,, for tadp. The use of the term cfmda in a sensible cooling application should not be construed to indicate that dehumidification is occurring. It is used in the “Air Conditioning Load Estimate” form and in Example 5 to determine the air quantity required thru the apparatus to offset the conditioning loads. ~, The leaving air conditions from the coil are dictated by the room design conditions, the load and the required air quantity. The entering and leaving dry-bulb temperatures at the apparatus should always be used to determine the effective surface temperature of the coil. Since this is strictly a sensible heat process, it is straight line function occurring at constant moisture content. Introducing wet-bulb into the calculation results in an erroneously low t,,. If this erroneous t,, is used to select the conditioning equipment, then 1. Direct expansion equipment would be selected at a lower refrigerant temperature than actually required. 2. Chilled water equipment would be selected for colder or more chilled water than actually required. . Location - Bakersfield, California Summer design - IO5 F dh, 70 F WI) Inside design - 75 F dh, 50’;:, maximum rh RSH - 200,000 Btu/hr RLH - none Ventilation - 13,000 c/m,,, Find: 1. 2. 3. 4. 5. 6. Outdoor air load (OATH) Effective sensible heat factor (ESHF) Effective surface temperature (tea)* Dehumidified air quantity (cfm,,) Effective surface temperature (tes)* Supply air temperature (tJ Solution: OASH = 1 . 0 8 X 1 3 , 0 0 0 (105 - 75) = 420,000 Btu/hr (14) OALH = .68 X 1 3 , 0 0 0 (54 - 54) = 0 OATH = 420.000 Btu/hr (15) (17) Assume a bypass factor of 0.05 from Tables 61 and 62. 200,000 + (.05) (420,000) ESHF = 200,000 + (.05) (420,000) = “ ’ Plot the ESHF to the saturation line on the psychrometric chart. The effective surface temperature is read as tcs = 50.3 F, Fig. 53. 200,000 + (.05) (420,000) cf “Ida = 1.08 x (1 - .05) X (75 - 50.3) = 8650 cfm 36) Since the dehumidified air quantity is less than the outdoor ventilation requirements, substitute the cfn,, for in Step 4. This results in a new effective surface Cfnda temperature which does not lie on the saturation line. This is illustrated in the following equation: tes = 75 - 200,000 + (.05) (420,000) 1.08 X (1 - .05) X 13,000 = 58.4 F (36) This temperature falls on the ESHF line, which is also the GSHF and RSHF lines in a sensible cooling appli4 cation. 6. Substitute tea for tadp in equation (28), supply air condition tsa as follows: t 8LI = 105 - (1 - .05) (105 - 58.4) tda is the same as the tldb which and calculate the = 60.7 F db (28) is the leaving air condition from the coil. The wet-bulb temperature of the air supplied to the space is 54 F wb. This is read at the intersection of the supply air dry-bull, temperatnre and the ESHF line, Fig. 53. . In Example 5, the assumed .05 bypass factor is used to determine t,, and dehumidified air quantity. Since the dehumidified air quantity is less than *The terms tep is substituted for tadp on the air conditioning load estimate‘form for sensible cooling application. NOTE: Numbers in parentheses at right edge of column refer to equations heginning on page 150. CHAI’I‘EK 8. AI’I’LIEI~ 1-135 I’SYCI-IROME’I‘RICS ventilation air rcquircment, the .05 bypass factor is used again to determine a new t,,, substituting the ventilation air requirement for the dchumidified air quantity. The new L,, is 58.4 F. ciency. It can be considered to represent that portion of the air passing thru the spray chamber which contacts the spray water surface. This contacted air is considered to be leaving the spray chamber at the effective surface temperature of the spray water. This effective surface temperature is the temperature at complete saturation of the air. Though not a straight line function, the e,rcct of saturation efficiency on the leaving air conditions from a spray chamber may be determined with a sufficient degree of accuracy from the following equation: 54 gr/lb Ll 58.4 F db 64.7 Fdb 75Fdb 105 Fdb FIG . 53 - S ENSIBLE COOLING If a coil with a higher bypass factor is substituted in Example 5, a lotier t,, results. Under these conditions, it becomes a question of economic balance when determining which coil selection and which refrigerant temperature is the best for the application. For instance, the maximum possible coil bypass factor that can be used is .19. This still results in a t,, above 50.3, and at the same time maintains a dehumidified air cfm of 13,000 which equals the ventilation requirements. SPRAY CHARACTERISTICS J, the operation of spray type equipment, air is c-,.wn or forced thru a chamber where water is sprayed thru nozzles into the air stream. The spray nozzles may be arranged within the chamber to spray the water counter to air flow, parallel to air flow, or in a pattern that is a combination of these two. Generally, the counter-flow sprays are the most efficient; parallel flow sprays are the least efficient; and when both are employed, the efficiency falls somewhere in between these extremes. SATURATION EFFICIENCY In a spray chamber, air is brought into contact with a dense spray of water. The air approaches the state of complete saturation. The degree of saturation is termed saturation efficiency (sometimes called contact or performance factor). Saturation efficiency is, therefore, a measure of the spray chamber effi- The saturation efficiency is the complement of bypass factor, and with spray equipment the bypass factor is used in the calculation of the cooling load. Bypass factor, therefore, represents that portion of the air passing thru the spray equipment which is considered to be leaving the spray chamber completely unaltered from its entering condition. This efficiency of the sprays in the spray chamber is dependent on the spray surface available and on the time available for the air to contact the spray water surface. The available surface is determined by the water particle size in the spray mist (pressure at the spray nozzle and the nozzle size), the quantity of water sprayed, number of banks of nozzles, and the number of.nozzles in each bank. The time available for contact depends on the velocity of the air thru the chamber, the length of the effective spray chamber, and the direction of the sprays relative to the air flow. As the available surface decreases or as the time available for contact decreases, the saturation efficiency of the spray chamber decreases. Table 63 illustrates the relative efficiency of different spray chamber arrangements. The relationship of the spray water temperatures to the air temperatures is essential in understanding the psychrometrics of the various spray processes. It can be assumed that the leaving water temperature from a spray chamber, after it has contacted the air, is equal to the leaving air wet-bulb temperature. The leaving water temperature will not usually vary more than a degree from the leaving air wet-bulb temperature. Then the entering water temperature is, therefore, dependent on the water quantity and the heat required to be added or removed from the air. Table 63 illustrates the relative efficiency of different spray chamber arrangements. 1-136 TABLE I’:\K’l‘ 63-TYPICAL SATURATION I. IDi\D ESTIMi\‘l‘lNG EFFICIENCY* For Spray Chambers l/q” NO. OF BANKS DIRECTION OF WATER SPRAY NOZZLE (25 psig Nozzle Pressure 3 gem/sq ftt) Velocityf I/~” NOZZLE (30 psig Nozzle Pressure 2.5 gpm/sq ftt) (fpm) 300 700 300 700 1 Parallel Counter 70% 75% 50% 65% 80% 82% 60% 70% 2 Parallel Opposing Counter 90% 98% 99% 85% 92% 93% 92% 98% 99% 87% 93% 94% ‘Saturation efficiency = 1 - BF tCpm/sq ft of chamber face area TVelocities above 700 fpm and below 300 fpm normally do not permit eliminators to adequately remove moisture from .ie air. Reference to manufacturers’ data is suggested for .miting velocity and performance. SPRAY PROCESSES Sprays are capable of cooling and dehumidifying, sensible cooling, cooling and humidifying, and heating and humidifying. Sensible cooling may be accomplished only when the entering air dewpoint is the same as the effective surface temperature of the spray water. The various spray processes are represented on the psychrometric chart in Fig. 54. All process lines must go toward the saturation line, in order to be at or near saturation. Adiabatic Saturation or Evaporative Cooling Line (1 - 2) represents the evaporative cooling process. This process occurs when air passes thru a spray chamber where heat has not been added to -.r removed from the spray water. (This does not .clude heat gain from the water pump and thru the apparatus casing.) When plotted on the psychrometric chart, this line approximately follows up the line of the wet-bulb temperature of the air entering the spray chamber. The spray water temperature remains essentially constant at this wetbulb temperature. Cooling and Humidification - With Chilled Spray Water If the spray water receives limited cooling before it is sprayed into the air stream, the sIope of the process line will move down from the evaporative cooling line. This process is represented by line (I - 3). Limited cooling causes the leaving air to be lower in dry- and wet-bulb temperatures, but higher in moisture content, than the air entering the spray chamber. . . DRY-BULB TEMPERATURE FIG . 54 - S PRAY P ROCESSES Sensible Cooling If the spray water is cooled further, sensible cooling occurs. This process is represented by line (1 - 4). Sensible cooling occurs only when tho entering air dewpoint is equal to the effective surface temperature of the spray water; this condition is rare. In a sensible cooling process, the air leaving the spray chamber is lower in dry- and wet-bulb temperatures but equal in moisture content to the entering air. Cooling and Dehumidification If the spray water is cooled still further, cooling and dehumidification takes place. This is illustrated by line (1 - 5). The leaving air is lower in dry- and wet-bulb temperatures and in moisture content than the air entering the spray chamber. Cooling and Humidification - With Heated Spray Water When the spray water is heated to a limited degree before it is sprayed into the air stream, the slope of the process line rises to a point above the evaporative cooling line. This is illustrated by line (1 - 6). Note that the leaving air is lower in dry-bulb temperature, but higher in wet-bulb temperature and moisture content, than the air entering the spray chamber. Heating and Humidification If the spray water is sufficiently heated, a heating and humidification process results. This is represented by line (1 - 7). In this process the dry-bulb CH,\I”I‘EK 8 . AI’I’LIEI) temperature, wet-bulb temperature, and moisture content of the leaving air is greater than that of the entering air. SPRAY PROCESS EXAMPLES The following descriptions and examples provide a better understanding of the various psychrometric processes involved in spray washer equipment. Cooling and Dehumidification w h e n a 1-137 I’SYCHKOME’I’KICS S p r a y chanlberiS t o b e used h coohg and dehumidification, the procedure for estimating the load and selecting the equipment is identical to the procedure described on page 128 for coils. The “Air Conditioning Load Estimate” form is used to evaluate the load; bypass factor is determined by subtracting the selected saturation efficiency from one. Spray chamber dehumidifiers may not be rated in ’ -ms o f a p p a r a t u s dewpoint but in terms of ent. -“g arid leaving wet-bulb temperatures at the apparatus. The apparatus dewpoint must still be determined, however, to evaluate properly the entering and leaving wet-bulb temperatures and the dehumidified air quantity. Although originally prepared to exemplify the is also operation of a coil, typical of the cooling and &humidifying process using sprays. Cooling and Dehumidification - Usiqg All Outdoor Air When a spray chamber is to be used for cooling and dehumidifying with all. outdoor air, the procedure for determining adp, entering and leaving conditions at the chamber, ESHF and cfmda is identical to the procedure for determining these items for coils using all outdoor air. Therefore, the description on page 130 and Example 3 may be used to ’ lyze this type of application. temI)erature is to he maintained during the winter or intermediate season, heat must be available to the system. This is usually accomplished by adding a reheat coil. When relative humidity is to be maintained in addition to room dry-bulb during the winter or intermediate season, a combination of preheat and reheat coils, or a reheat coil and spray water heating, is required. The latter method changes the process from evaporative cooling to one of the humidification processes illustrated by lines (1 - 6) or (I - 7) in Fig. 54. Evaporative cooling may be used in industrial applications where the humidity alone is critical, and also in dry climates where evaporative cooling gives some measure of relief by removing sensible heat. L;xnmple 6 illustrates an industrial application designed to maintain the space relative humidity only. Example 6 - Evaporative Cooling Given: An industrial application Location - Columbia, South Carolina Summer design - 95 F db. 75 F wh Inside design - 55~~ rh RSH - ?,lOO.OOO Btu/hr RSHF - 1.0 Use all outdoor air at design load conditions Find: 1. Room dry-bulb temperature at design (f,,) 2. Supply air quantity (cftn,,) Evaporative Cooling An evaporative cooling application is the simultaneous removal of sensible heat and the addition of moisture to the air, line (1 - 2), Fig. 54. The spray water temperature remains essentially constant at the wet-bulb temperature of the air. This is a process in which heat is not added to or removed from the spray water. (Heat gain from the water pump and heat gain thru the apparatus casing are not included.) Evaporative cooling is commonly used for those applications where the relative humidity is to be controlled but where no control is required for the room dry-bulb temperature, except to hold it above a predetermined minimum. When the dry-bulb b-92.3 Fdb C-94.0 F db B-93.2 F db D-95.0 F db ABC0 FIG. 55 - EVAPORATIVE COOLING, W ITH VARYING SATURATION EFFICIENCY 1-138 PAKT I. LO:\D E S T I M A T I N G L Solution: I . Dcterminc the room dry-l)ulb temperature by comp r o m i s i n g I)ctwcen the spray saturation efficiency, the acccl~tablc room dry-l)ull) tempcraturr, and the supply air quantity. To evaluate these items, use the following equation to tlctcrrninc the leaving conditions from the spray for variorls saturation efficiencies: 1 te d b - (Sat Elf) (lEdb - tczub)” Tllc I-oom dry-l~ulb tcrnperatllrc in the following table results from various spray saturation efficiencies and is dctennined by plotting the RSHF thru the various leaving conditions, to the design relative humidity, Fig. 55. Note that the supply air temperature rise clecreases more rapidly than the room dry-bulb temperature. Correspondingly, as the supply air temperature rise decreases, the supply air quantity increases in t h e s a m e proportion. DRY-BULB TEMP LEAVING SPRAYS S.4T EFF (%o) (tldh) 2. 100 75 95 90 85 80 76 77 78 79 SUPPLY AIR TEMP RISE (At) with the auxiliary sprays in the space, becomes a problem of economics which shouitl be analyzed for each particular application. When a split system is used, supplemental spray heads arc usually added to the straight evaporative cooling system. These spray heads atomize water and add supplementary moisture directly to the room. This added moisture is evaporated at the final room wet-bulb temperature, and the room sensible heat is reduced by the amount of heat required to evaporate the sprayed water. Table 64 gives the recommended maximum moisture to be added, based on a 65 F db room temperature or over, without causing condensation on the ductwork. ROOM DRY-BULB TEMP AT 55% RH it,,) I9 17.6 16.2 14.7 13.3 various Without ROOM DESIGN RH 94 93.6 93.2 92.7 92.3 Calculate the supply air quantity for the perature rises from the following equation: TABLE 64-MAXIMUM RECOMMENDED MOISTURE ADDED TO SUPPLY AIR 85 80 75 tem- Causing /MOISTURE Gr/Cu Ft Dry Air 1.25 1.30 1.35 1.40 70 Condensation on Ductct 1ROOM / MOISTURE DESIGN Gr/Cu Ft RH Dry Air 65 60 55 1.50 1.60 1.70 1.80 50 tThese are arbitrary limits which have been established by a combination of theory and field experience. These limits apply where the room dry-bulb temperature is 65 F db or over. SUPPLY AIR TEMP RISE (tWTl SUPPLY AIR QUANTITY - ‘Mb) (cfm,,) 19 102,400 The spray chamber and supply air quantity should then be selected to result in the best owning and operating costs. The selection is based primarily on economic considerations. Evaporative Cooling Used With A Split System There are occasions when using straight evaporative cooling results in excessive air quantity requirements and an unsatisfactory air distribution system. This situation usually arises in applications that are to be maintained at higher relative humidities (70% or more). To use straight evaporat’ve cooling with the large air quantity, or to use a s t lit system *This equation is applicable only to evaporative cooling applications where the entering air wet-bulb temperature, the leaving air wet-bulb temperature, and the entering and leaving water temperature to the sprays are all equal. . 4s a rule of thumb, the air is reduced in temperature approximately 8.3 F for every grain of moisture per cubic foot added. This value is often used as a check on the final room temperature as read from the psychrometric chart. Example 7 illustrates an evaporative cooling application with supplemental spray heads used in the space. Example 74voporafive Cooling-With Auxiliary Sprays Given: An industrial application Location - Columbia, South Carolina Summer design - 95 F db, 75 F wb Inside design - TOY0 rh RSH - 2.100.000 Btu/hr RSHF - 1.0 Moisture added by auxiliary sp’ray heads - 19 gr/lb (13.9 cu ft/lb X 1.4 gr/cu ft) Use all outtloor air thru a spray chamber with 90y0 saturation efficiency. Find: 1. 2. 3. 4. Leaving Room Supply Supply conditions from spray chamber pldb, tIwb) dry-bulb temperature (tr,,) air quantity (cfm,,) with auxiliary sprays air quantity (cfm,,) without auxiliary sprays , 1-139 CH,\I'J‘EK 8. i\l'I'L.Il~I~ I'SYCHKOME'I‘KICS Cfffl,, = -- 1.08 X F EVAP COOLING RSH t-i& temp = 2,100,000 1.08 X 23.75 - = 82,000 cfm 4. If no auxiliary sprays were to I)e used, the room design dry-l)ulb would be where the RSHF line intersects the room tlesign relative humidity. From Fig. 56, the room dry-bulb is read t rn”A - 84.7 I; tlb The supply air quantity required to maintain the room design relative humidity is determined from the following equation: \_THE~RETICAL db TEMP RISE WITHOUT AUXILIARY SPRAYS cf ItIle = Heating FIG. 56 89.2 Fdb 95Fdb\ 94.7F db 100.75F db -EVAPORATIVE COOLING, WITH AUXILIARY SPRAYS WITHINTHE SPACE Solution: 1. tldb = ted,-- (Sat Eff) Ctedb - tewb) = 95 - .90 (95 - 75) = 77 F db t Iwb is the same as the tewb process, Fig. 56. 2. in an evaporative cooling Room dry-bulb temperature is evaluated the moisture content of the space. W rm = Wsa by determining + 19 = 128 + 19 = 147 gr/lb The 19 gr/lb is the moisture added to the space by the auxiliary spray heads. T h e trm is the point on the psychrometric chart where h e Wrm intersects the 70% design relative humidity line, Fig. 56. t t 3. rm = 89.2 F db Psychrometrically, it can be assumed that the atomized water from the spray heads absorbs part of the room sensible heat and turns into water vapor at the final room wet-bulb temperature. The intersection of this wetbull, temperature with the moisture content of the air leaving the evaporative cooler is the theoretical dry-bulb equivalent temperature if the auxiliary sprays were not operating. The difference between this theoretical drybulb equivalent temperature and the temperature of the spray chamber, tldb, is used to determine the supply air quantity. tldO (from spray chamber) = 77 F. The theoretical dry-bulb temp is 100.75 Temp rise = 23.75 F dh 2.100,000 = 1.08 (84.7 - 77) = 253,000 cfm This air required However, quantity, 84.7 F to I 77Fdb RSH 1.08 Cfrm - tld*l I;, Fig. 56. and quantity is over three times the air quantity when auxiliary sprays are used in the space. it should be noted that, by reducing the air the room dry-bulb temperature increased from 89.2 F. Humidification -With Sprays A heating and humidifying application is one in which heat and moisture are simultaneously added to the air, line (1 - 7), Fig. 54. This may be required during the intermediate and winter seasons or during partial loads where both the dry-bulb tempera- / ture and relative humidity are to be maintained. Heating and humidification may be accomplished by either of the following methods: 1. Add heat to the spray water before it is sprayed into the air stream. 2. Preheat the air with a steam or hot water coil and then evaporatively cool it in the spray chamber. Spray water is heated, by a steam to water interchanger or by direct injection of steam into the water system. Since the supply air quantity and the spray water quantity have been determined from the summer design conditions, the only other requirement is to determine the amount of heat to be added to the spray water or to the preheater. For applications requiring humidification, the room latent load is usually not calculated and the room sensible heat factor is assumed to be 1 .O. Example 8 illustrates the psychrometric calculations for a heating and humidifying application when the spray water is heated. It should be noted that this type of application occurs only when the quantity of outdoor air required is large in relation to the total air quantity. Example 8 - Heating and Humidification With Heated Spray Water Given: 4n industrial application Location - Richmond, Virginia l-140 PART the wet-hulh line, Fig. 54. Winter tlesign - 15 F dh Inside design - 72 F tll), 35”/” rh Ventilation - 50,000 cffnon (see explanation above) Supply air - 85,000 ~/rn,~ Design room heat loss - 2,500,OOO Btu/hr Spray saturation efficiency - 95% RSHF (winter conditions) - 1.0 Make-up water - 65 F t 1. tsn = ~~xnt = wkI - wea Sat Eff + We=a 41 - 17 =- + 17 = 42.3 gr/lb .95 The heating and humidification process line is plotted on the psychrometric chart between the moisture content of saturated air (42.3 gr/lb) and the entering conditions to the spray chamber (38.5 F db and 32.4 F wb), Fig. 57. + t,n, To determine the wet-bulb temperature, plot the RSHF line on the psychrometric chart and read the wet-bulb at the point where t,@ crosses this line (Fig. 57). Supply air wet-bull, to the space = 65.8 F wb. The leaving conditions chart where the room intersects the heating Fig. 57. = 38,5 F d,, (15 X 50,000) + (72 x 35,000) 85,000 -.- tlwb = 43.4 F wb The temperature of the leaving spray water is approximately equal to the wet-bulb temperature of the air leaving the spray chamber. t Iw = 43.4 F (31) To determine wet-bull) temperature of the air entering the spray chamber, plot the mixture line of outdoor and return room air on the psychrometric chart, and read are read from the psychrometric moisture content line (41 gr/lb) and humidification process line, t Idb = 43.6 F db 2. To determine the entering and leaving spray water temperature, calculate the entering and leaving air conditions at the spray chamber: = = 32.4 F WI, tllL.) 2,500,000 +72=99.2 F d h = 1.08 X 85,000 t edb NOTE: Numbers in parentheses at right edge of column. refer to equations heginning on page IJO. SUPPLY w.. .- AIR I/ REHEAT 42.3 y/lb 41 grllb 1 7 gr/lb ’ I3 Fdb FIG. 57 crosses the mixture W )T,L = Wla = 41 gr/lh Since the spray chamber has a saturation efficiency of ‘3~5%~ the moisture content of completely saturated air is calculated as follows: Solution: design room heat loss 1.08 x ~ffn,~ temperature where tedb The air leaving the spray chamber must have the same moisture content as the air in the room. Find: 1. Supply air conditions to the space (t3 2. Entering and leaving spray water temperature jtCw, 3. Heat added to spray water to select water heater, rwb I. L O A D E S T I M A T I N G 38.5 Fdb 4L6 Fdb - HEATING AND . ?ZFdb 99.2 Fdb HUMIDIFICATION, W ITH HEATING S PRAY W ATER l 1-141 CHAPTER 8. APPLIED I’SYCHROMETRICS The temperature of the entering spray water is dependent on the water quantity and the heat to Ix! added or removed from the air. In this type of application, the water quantity is us~~ally dictated hy the cooling t&l design requirements. ,\ssume, for illustration purposes, that this spray washer is selected for 110 gpm for cooling. The heat added to the air as it passes through the washer = cfn,, x 4.45 x (llL1. - ‘2,J = 85,000 X 4.45 X (16.85 - 12) = 1,830,OOO Btu/hr The entering water following equation: t ew = tlw temperature is determined from the heat added to air 500 x gpm + I ,830,000 = 43.4 + 500 x 110 = 76.8 F 3. The heat added to the spray water (for selecting spray -%er heater) is equal to the heat added to the air plus heat ,added to the make-up water. The amount of make-up water is equal to the amount of moisture evaporated into the air and is determined from the following equation: outdoor air and mixing it with the return air from the space. This mixture must then be evaporativcly cooled to the room dewpoint (or room moisture content). And finally, the air leaving the spray chamber must be reheated to the required supply air temperature. SORBENT DEHUMIDIFIERS Sorbent dehumidifiers contain liquid absorbent or solid adsorbent which are either sprayed directly into, or located in, the path of the air stream. The liquid absorbent changes either physically or chemically, or both, during the sorption process. The solid adsorbent does not change during the sorption process. As moist air comes in contact with either the liquid absorbent or solid adsorbent, moisture is removed from the air by the difference in vapor pressure between the air stream and the sorbent. As cfm,, P1, - WcJ Make-up where: water = wea, Wza 7000 12.7 8.34 7000 X 12.7 X 8.34 I = moisture content of the air entering and leaving the spray washer in grains per pound of dry air = grains of moisture per pound of dry air = volume of the mixture in cubic feet per pound of dry air, determined from psychrometric chart = water in pounds per gallon 85,000 (41 - 17) Make-up water = 7ooo x 12.7 x 8.34 = 2.8 gpm The heat added to the make-up spray water is determined from the following equation: Heat added to make-up water . = gpm X 500 (tew - make-up water temp) = 2.8 X 500 (76.8 - 65) = 16,200 Btu/hr To select a water heater, the total amount of heat added to the spray water is determined by totaling the heat added to the air and the heat added to the make-up spray water. Heat added to spray water = 1,830,OOO + 16200 = 1,846,200 Btu/hr If the make-up water was at a higher temperature than the required entering water temperature to the sprays, then a credit to the heat added to the spray water may be taken. In this example a reheat coil is required to heat the air leaving the spray chamber, at 43.6 F db and at a constant moisture content of 41 gr/lb, to the required supply air temperature of 99.2 F db. The requirements of the application illustrated in Example 8 can also be met by preheating the DRY-BULB F IG. TEMPERATURE 58 - SORBENT DEHUMIDIFICATION PROCESSES this moisture condenses, latent heat of condensation is liberated, causing a rise in the temperature of the air stream and the sorbent material. This process occurs at a wet-bulb temperature that is approximately constant. However, instead of adding moisture to the air as in an evaporative cooling process, the reverse occurs. Heat is added to the air and moisture is removed from the air stream; thus it is a dehumidification and heating process as illustrated in Fig. 58. Line (I - 2) is the theoretical process and the dotted line (I -3) approximates what actually happens. Line (I - 3) can vary, depending on the type of sorbent used. I’AK-I‘ 1-142 I. LOAD ESI’IM,\~I‘ING PSYCHROMETRICS OF PARTIAL LOAD CONTROL The apparatus required to maintain proper space conditions is normally selected for peak load operation. Actually, peak load occurs but a few times each year and operation is predominantly at partial load conditions. Partial load may be caused by a reduction in sensible or latent loads in the space, or in the outdoor air load. It may also be caused by a reduction in these loads in any combination. PARTIAL LOAD ANALYSIS Since the system operates at partial load most of the time and must maintain conditions commensurate with job requirements, partial load analysis is at least as important as the selection of equipment. Partial load analysis should include study of resultant room conditions at minimum total load. Usually this will be sufficient. Certain applications, however, should be evaluated at minimum latent load with design sensible load, or minimum sensible load and full latent load. Realistic minimum and maximum loads should be assumed for the particular application so that, psychrometrically, the resulting room conditions are properly analyzed. Figure 57 illustrates the psychrometrics of reheat control. The solid lines represent the process at design load, and the broken lines indicate the resulting process at partial load. The RSHF value, plotted from room design conditions to point (2), must be calculated for the minimum practical room sensible load. The room thermostat then controls the temperature of the air leaving the reheat coil along line (1 - 2). This type of control is applicable for any RSHF ratio that intersects line (I - 2). If the internal latent loads decrease, the resulting room conditions are at point (3), and the new RSHF process line is along line (2 - 3). However, if humidity is to be maintained within the space, the reduced latent load is compensated by humidifying, thus returning to the design room conditions. The six most common methods, used singly or in combination, of controlling space conditions for cooling applications at partial load are the following: (SENSIBLE 1. Reheat the supply air. 2. Bypass the heat transfer equipment. 3. Control the volume of the supply air. 4. Use on-off control of the air handling equipment. HEATING) \ ‘NEW RSHF LINE i WITH REDUCED ROOM SIGN DRY-BULB I SENSIBLE HEAT J DRY-BULB TEMPERATURE 5. Use on-off control of the refrigeration machine. 6. Control the refrigeration capacity. The type of control selected for a specific application depends on the nature of the loads, the conditions to be maintained within the space, and available plant facilities. REHEAT CONTROL Reheat control maintains the dry-bulb temperature within the space by replacing any decrease in the sensible loads by an artificial load. As the internal latent load and/or the outdoor latent load decreases, the space relative humidity decreases. If humidity is to be maintained, rehumidifying is reqy.ired in addition to reheat. This was described previously under “Spray Process, Heating and Humidifying.” . FIG. 59 - PSYCHROMETRICS OF REHEAT CONTROL BYPASS CONTROL Bypass control maintains the dry-bulb temperature within the space by modulating the amount of air to be cooled, thus varying the supply air temperature to the space. Fig. 60 illustrates one method of bypass control when bypassing return air only. Bypass control may also be accomplished by bypassing a mixture of outdoor and return air around the heat transfer equipment. This method of control is inferior to bypassing return air only since it introduces raw unconditioned air into the space, thus allowing an increase in room relative humidity. . 1-143 CHAPTICK 8. AI'I'LIEI) I'SY(:l-IKO~lI1'TKICS DESIGN SUPPLY CON& DESIGN DESIGN RSHF NEW RSHF LINE WITH REDUCED ROOM SENSIBLE HEAT. AROUNO APPARATUS MIXTURE OF BYPASSED AIR AND AIR THRU OEHUMIDIFIER DRY-BULB TEMPERATURE FIG. 60 -PSYCHROMETRICSOFBYPASSCONTROLWITH A reduction in room sensible load causes the bypass control to reduce the amount of air thru the dehumidifier. This reduced air quantity results in equipment operation at a lower apparatus dewpoint. Also, the air leaves the dehumidifier at a 1C - temperature so that there is a tendency to ac,.“st for a decrease in sensible load that is proportionately greater than the decrease in latent load. Bypass control maintains the room dry-bulb temperature but does not prevent the relative humidity from rising above design. With bypass control, therefore, increased relative humidity occurs under conditions of decreasing room sensib!e load and relatively constant room latent load and outdoor air load. The heavy lines in Fig. 60 represent the cycle for design conditions. The light lines illustrate the initial cycle of the air when bypass control first begins to function. The new room conditions, mixture conditions and apparatus dewpoint continue KETURNAIRONLY to change until the equilibrium point is reached. Point (2) on Figs. 60 and 61 is the condition of air leaving the dehumidifier. This is a result of a smaller bypass factor and lower apparatus dewpoint caused by less air thru the cooling equipment and a smaller load on the equipment. Line (2 - 3 - 4) represents the new RSHF line caused by the reduced room sensible load. Point (3) falls on the new RSHF line when bypassing return air only. Bypassing a mixture of outdoor and return air causes the mixture point (3) to fall on the GSHF line, I;ig. 60. The air is then supplied to the space along the new RSHF line (not shown in Fig. 60) at a higher moisture content than the air supplied when bypassing return air only. Thus it can be readily observed that humidity control is further hindered with the introduction of unconditioned outdoor air into the space. VOLUME CONTROL Volume control of the supply air quantity provides essentially the same type of control that results I’AKT I. LOAD ES’I‘IMATING 1-144 Erom bypassing return air around the heat transfer equipment, I;ig. 60. However, this type of control may produce problems in air distribution within the space and, therefore, the required air quantity at partial load sl~ould be evaluated for proper air distribution. - R E T U R N AIR CONDITIONED SPACE @ c ’ OUTDOOR AIR - / BYPASS AIR 4 F - FAN APPARATUS 2 3 &- FIG. 61- SCHEMATIC SKETCH OF B YPASS CONTROL W ITH BYPASS OF RETURN A IR O NLY ON-OFF CONTROL OF AIR HANDLING EQUIPMENT On-off control oE air handling equipment (Eancoil units) results in a fluctuating room temperature and space relative humidity. During the “off” operation the ventilation air supply is shut off, but chilled water continues to flow thru the coils. This method of control is not recommended for high latent load applications, as control of humidity may be lost at reduced room sensible loads. .- ON-OFF CONTROL OF REFRIGERATION EQUIPMENT On-off control of refrigeration equipment (large packaged equipment) results in a fluctuating room temperature and space relative humidity. During the “oW operation air is available for ventiIation purposes but the coil does not provide cooling. Thus, any outdoor air in the system is introduced into the space unconditioned. Also the condensed moisture that remains on the cooling coil, when the reErigeration equipment is turned ofE, is re-evaporated in the warm air stream. This is known as re-evaporation. Both oE these conditions increase the space latent load, and excessive humidity results. This method of control is not rccommendcd for high latent load applications since control of humidity may be lost at decreased room sensible loads. REFRIGERATION CAPACITY CONTROL Relrigeration capacity control may be used on either chilled water or direct expansion ret‘rigeration equipment. Partial load control is accomplished on chilled water equipment by bypassing the chilled water around the air side equipment (Ean-coil units). Direct expansion refrigeration equipment is controlled either by unloading the compressor cylinders or by back pressure regulation ’ in the refrigerant suction line. Refrigeration capacity control is normally used in combination with bypass or reheat control. When used in combination, resuIts are excellent. When used alone, results are not as effective. For example, temperature can be maintained reasonably well, but relative humidity will rise above design- at partial load conditions, because the latent load may not reduce in proportion to the sensible load. * PARTIAL’ LOAD CONTROL Generally, reheat control is more expensive but provides the best control of conditions in the space. Bypass control, volume control and refrigeration capacity control provide reasonably good humidity control in average or high sensible heat factor applications, and poor humidity control in low sensible heat Eactor applications. On-off control usually results in the least desirable method of maintaining space conditions. However, this type of control is frequently used for high sensible heat factor applications with reasonably satisfactory results. CH,\I”I‘I-K 8 . /\I’I’I,IEI~ I’SY~:FI11OILlE’I‘I~I~:S l-145 TABLE 65-APPARATUSDEWPOINTS EFFECTIVE SENSIBLE HEAT FACTOR AND APPARATUS DEWPOINT” 9 0 - 80 F DB ROOM CONDITIONS DBlRHl WBI W EFFECTIVE SENSIBLE HEAT FACTOR AND APPARATUS DEWPOINT* (F) I / (%) / (F) // (fir/lb) 45 bb7. 73S. ESHF ADP 1 . 0 0 .91 .87 .80 .75 .72 .bB’.bS .43 58.5 57 56 54 52 50 46 41 53 Ss b9. 8 9.. 2 ESHF ADP 1 . 0 0 .90 .83 .74 .bB .b4 . b l .58 .5b 6 4 . 2 63 62 60 58 56 54 50 44 bs 72. 8 ,07 . o I , ,4 9 ,ob 4 ESHF . . ADP 1 . 0 0 .92 60.9 68 .?8 66 .b4 .bO S8 6 1 5 8 56 63 Sb .54 53 47 55 7b . 7 ,,,. S ESHF 1 . 0 0 .92 .7b .bB .b4 .57 .54 .52 .50 ADP 7 1 . 6 71 69 67 66 62 59 57 50 , b. 78 . 4 128 . 4 ESHF 1 . 0 0 .Bb .bB .bO Sb .52 .SO .48 .4b ADP 7 4 . 2 73 71 69 67 64 62 59 50 bs 8.. o 139 . b ESHF ADP 1 . 0 0 .75 .bB .b2 .SS .SO .47 .45 ,43 7 6 . 8 75 74 73 71 69 66 64 59 7. 81 . b lsl . o ESHF ADP 1 . 0 0 .78 .$b .bO .S2 .47 .43 .41 39 7 9 . 0 78 7 7 , 7 6 74 72 69 66 58 / I ESHF ADP 1 . 0 0 .Bb .71 .b3 .58 .54 .S2 .Sl .49 69.1 68 66 64 62 60 58 56 51 7o 74. 2 ,ls. s ESHF ADP 1 . 0 0 .80 .71 .bS .bO .54 Sl .48 .4b 7 1 . 2 7 0 69 68 67 65 63 60 56 3s b25 . Ss2 . ESHF ADP 1 . 0 0 .94 .89 .84 .81 .77 .75 .73 .71 5 0 . 8 49 47 45 43 39 36 32 21 4. b4 . 2 b3. 2 ESHF ADP 1 . 0 0 .94 .87 .02 .78 .75 .72 .b9 .b7 5 4 . 4 53 51 49 47 45 41 36 23 4S bs9. 71 . 2 ESHF ADP 1 . 0 0 .9b .91 .83 .78 .74 .70 .b7 .b4 5 7 . 6 57 56 54 52 50 47 43 36 I 81 ss b9 o 874 E S H F 1 . 0 0 .90 .77 .71 . b b .b2 . b O .58 .56’ . . ADP 63.2 62 60 58 56 53 51 47 35 b. 7.. s bS 9s . 4 ESHF ADP 71 . 9 ,037. ESHF ADP I I I 7.. 8 90,. ESHF ADP ” 72-3i 99-4 ESHF ADP 1 . 0 0 .85 .76 .71 .bb .bO .Sb .S2 SO 6 8 . 2 6 7 66 65 64 62 60 56 52 I 7. 73 . 3 ,,, . 9 ESHF ADP so 1 . 0 0 .92 .77 .68 .63 .59 Sb .54 .53’ 6 5 . 8 65 63 61 59 56 53 50 4 6 / I I I I / I I 1 . 0 0 .80 .71 .bl SS .52 .48 .47 .46 7 0 . 3 69 68 66 64 62 58 56 52 1 . 0 0 .92 .80 .73 .bB .b4 . b l .S9 .57 6 4 . 2 63 61 59 57 54 51 40 39 1 . 0 0 .92 .83 .73 .b7 .bO .S7 .Sb .54 6 6 . 9 66 65 63 61 57 54 52 4 7 bs 7s . S 1182 ESHF 1 . 0 0 .8B1 .b9 .bl .Sb .53 .SO .48 .47 . ADP 7 1 . 9 71 69 67 65 63 61 58 54 / 7. 77 . o ,27. b ESHF/ 1 . 0 0 .81 .b3 .SS ADP ! 7 4 . 0 73 71 69 82 _ 3S b33 . 570 ESHF . ADP 1 . 0 0 .92 .88 .84 .80 .7b .74 .72 .71 5 1 . 6 49 48 4 6 4 3 3 9 3 6 31 27 4o “*’ ESHF ADP 1 . 0 0 .90 .87 .82 .78 .74 .71 .b9 .b7 5 5 . 2 53 52 50 48 45 41 38 31 “*’ See page 147 for notes. ESHF ADP 1 . 0 0 .78 .71 .bS . b l .SS .52 .49 .476 9 . 4 68 6 7 , 66 65 63 61 58 5 3 , I 1-146 PART TABLE 79 - 72 F DB 65-APPARATUS DEWPOINTS (Continued) ROOM CONDITIONS EFFECTIVE SENSIBLE HEAT FACTOR AND APPARATUS DEWPOINT* 35 5e’9 EFFECTIVE SENSIBLE HEAT FACTOR A N D A P P A R A T U S DEWPOlNT* 46-7 1 . 0 0 .96 .91 .87 .83 .79 .77 .75 .73 40.2 4 7 4 5 4 3 41 3 7 3 5 31 2 2 4. 6, . 9 573 . ESHF ADP 1 . 0 0 .93 ,07 .82 5 1 . 7 50 4 8 46 1 I 1 5’ 65’o 78 s5 I 66 . 6 I I I I I I I I 1 . 0 0 .94 .06 .81 .77 .74 .71 .69 .67 53.2 52 50 48 4 6 4 4 4 0 3 7 31 67’4 ESHF ADP 1 . 0 0 .93 ‘.83 .77 .73 .69 .67 .65, .63 56.2 5 5 5 3 5 1 4 9 46 4 3 4 0 3 2 693 . 64’9 74’o ESHF ADP 1 . 0 0 .94 .82 .75 .70 .67 5 8 . 7 58 56 5 4 5 2 50 6o 66’2 ao’9 ESHF ADP 1 . 0 0 .90 .77 .70 .66 .62 .60 .58 .57 61.1 6 0 5 8 56 5 4 5 2 4 9 4 6 4 3 65 67 . 6 87 . 6 ESHF ADP 1 . 0 0 .84 .72 .65 .61 .58 .56 .54 .53 63.4 6 2 6 0 5 8 5 6 5 4 5 2 4 8 4 3 TO 68’9 94’6 ESHF ADP 1 . 0 0 .a0 .67 .60 .56 .54 .52 .51 .50 65.5 64 6 2 6 0 5 8 56 5 4 5 2 4 9 ! .64 .62 71’9 ESHFADP 57.9 1.00 .94 57.03 55.76 53.73 51 .70 49.67 47 42 36 ! 6. 67 9 86 4 ESHF 1.00 .90 .82 .76 .69 .64 .60 .57 -55 . . ADP 63.0 62 61 60 58 56 53 49 42 65 55 ! 79 . 2 ESHF 1 . 0 0 ,96 .83 .75 .70 .65 .62 .60 .59 ADP 60.5 6 0 5 8 5 6 5 4 51 48 4 4 41. I I I / / I / I / r 93 . 8 ESHF ADP 1 . 0 0 .85 .77 .71 .67 .62 .58 .54 .52 6 4 6 3 6 2 61 5 9 5 7 5 3 4 8 7. 7.. 6 ,o, . 2 ESHF ADP 1 . 0 0 .71 .66 .62 .59 .55 .52 .50 .48 67.5 6 5 6 4 6 3 62 6 0 5 8 5 5 4 8 75 7. 68 o 91 2 . . 601 6 7 . 1 1 83.61 70 69.8 97. 1 . 0 0 .96 .89 .84 .81 .78 .76 .72 .70 4 9 . 9 49 47 45 43 41 39 32 22 604 . ESHF ADP 6, . 9 .79 .77 .73 .71 .69 44 42 38 34 25 I ESHF ADP 45 I 5oo . ESHF ADP 1 . 0 0 .96 .91 . a 7 .84 .8i .79 .77 .74 46.3 4 5 4 3 41 39 37 3 4 31 21 537 . 76 6.f3 ESHF ADP 4. 604 . c:-’ 63’4 35 I. LOAD ESTIMATING .65 .62 .60 48 44 38 , ~H,\I’~I‘EII 1-147 H. ;\l’l’l.lEI) I’SY(:HKOMETRICS TABLE CONDITIONS 65-APPARATUS DEWPOINTS (Continued) ROOM CONDITIONS EFFECTIVE SENSIBLE HEAT FACTOR A N D A P P A R A T U S DEWPOINT* DB/RH WB/ (F) (F) I(%) 1(F) 1(er/lb) ” 64.0 76.3 7o 82.3 72 65.2 ESHF ADP (%I (F) 1 . 0 0 .84 .73 .67 .63 .61 .59 .S8 5 9 . 5 58 56 54 52 50 48 47 ESHF 1 . 0 0 .80 .69 .62 .59 .56 .54 .53 ADP 161.61 6 0 58 56 54 51 48 44 W 7 2 - 5 5 F DB EFFECTIVE SENSIBLE HEAT FACTOR A N D A P P A R A T U S DEWPOINT+ (w/lb) 6. S23 . E S H F ’ l . O O .94 .89 .81 .77 .74 .72 .70 .68 462 . ADP 46.0 45 44 42 40 38 36 34 28 ” 50*o 53*3 I70 114.31 53.91 ESHF ADP / 50.1 1 * 00 i1. 49 89 1. 48 83 1. 74 46 1. 701.67 44 142 I 75 E S H F 1 . 0 0 .91 .86 .78 .74!.70 .69 .67 .65 A D P 4 8 . 1 47 4 6 , 44 421 4 0 39 36 31 I 55’3 57.8 I ESHF ADP I I I I 1.001.79I.74 .71 .68 .64 .62 .60 .59 5 2 . 0 / 50 14 9 48 47 45 43 / 40 37 *O 56+3 “e7 ESHF ADP 1 . 0 0 .85 .76 I.70 .66 .61 .59 .57 .56 5 3 . 8 i 5 3 , 52 51 50 48 46 44 41 9o 58’2 69*4 ESHF ADP 95 59-1 73*5 E S H F 1 . 0 0 / .69 1 . 5 5 .49 .47 .46 .45 ADP 58.5 / 58 57 56 55 54 52 1 . 0 0 .72 .62 .57 .54 .52 S O .49 5 7 . 0 56 55 54 53 52 50 47 - T 70 55 i ES 9o ES 66 . 8 937. ESHF ADP 1 . 0 0 .71 .56 .52 .50 .48 .47 .46 .45 6 5 . 3 64 62 61 60 59 58 57 54 9. 67 . 9 99 . 3 ESHF ADP 1 . 0 0 .66 .56 .50 .47 .45 .43 .42 .41 6 6 . 9 66 65 64 63 62 61 60 56 ESHF 5’*2 ADP *O 5’-5 S24 . S4 . S ESHF 53*2 s7*7 1 . 0 0 .88 .79 .74 .67 .64 .62 .6l .60 48.8 48 47 46 44 42 40 39 37; ADP 1 . 0 0 .77 .70 .66 .63 .60 .58 .57 5 0 . 4 49 48 47 46 44 42 40 ESHF ADP 1 . 0 0 .76 .67 .61 SE .55 .54 .53 5 2 . 0 51 50 49 48 46 44 41 i * T h e v a l u e s shown i n t h e g r a y areas i n d i c a t e t h e l o w e s t e f f e c t i v e sensible heat factor possible without the use of reheat. This limiting condition is the lowest effective sensible heat factor line that intersects the saturation cuwc. Note that the r~~rn dewpoint is equal to the required apparatus dewpoint for on effective sensible heat factor of 1.0. 6s 65 57 . 7 59 7 ESHF . ADP 1 . 0 0 .92 .85 .80 .73 .69 .66 .64 .62 5 2 . 9 52 51 50 48 46 44 41 37 7. SE . 9 64 . S ESHF 1 . 0 0 .89 .80 .76 .69 .65 .62 .60’.58 ADP 5 5 . 0 54 53 52 50 48 46 4 3 3 7 75 S9 . 9 69 . 2 ESHF ADP 1 . 0 0 .88 .78 .72 .651.61 .58 .56 .55 5 6 . 9 56 55 54 5 2 5 0 48 45 41 8. S, . o 73 . 8 ESHF ADP 1 . 0 0 .75 .68 .63 5 8 . 7 57 56 55 / 620. I 83 . 2 ESHF ADP 1 . 0 0 .70 .58 .53 .50 .48 .46 .45 6 1 . 9 61 60 59 58 57 55 53 95 880 . ESHF ADP 1 . 0 0 .69 .51 .46 .43 .42 .41 6 3 . 5 63 62 61 60 59 58 = 1 , +.62a (wm- wad,) (trm - tad,,) 786 . ESHF 1 . 0 0 .71 .63 .58 .55 .52 .50 .49 ADP 6 0 . 3 59 58 57 56 54 52 50 9. 63 . o 640 . 1 . F o r R o o m C o n d i t i o n s N o t G i v e n ; T h e a p p a r a t u s dewpoint m a y be determined from the scale on the chart, or may be calculated CIS shown in the following equation: ESHF / / ES NOTES FOR TABLE 65: This equation in mwe familiar form il: E S H F 0 . 2 4 4 (frm - t.dp) = - 0 . 2 4 4 (trm -todp) + g (wr, - W.dp) (Cont.) 1-148 i 1’.4KT I. LOAD ESTIMATING 7000 = groins per pound. w h e r e W,, = room moisture content, gr/lb of dry air W.dp = moisture content at apparatus dewpoint, gr/lb of dry air 2. frm = room dry-bulb temperature todp =apporatus 3 . F o r A p p a r a t u s Dewpoint B e l o w F r e e z i n g . T h e l a t e n t h e a t o f fusion of the moisture removed is not included in the calculation of apparatus dewpoint below freezing or in the calculation of room load, in order to simplify estimating procedures. Use the some equation as in Note 1. The selection of equipment on CI basis of 16 to 1 B hour operating time provides a safety factor large enough to cover the omission of this latent heat of fusion, which is a small part of the total load. dewpoint temperature 0.244 = specific heat of moist air at 55 F dewpoint, Btu per deg F per lb of dry air 1076 = average heat removal required to condense one pound of water vapor from the room air For High Elevations. For elevations, see Table 66. effective sensible heat factors at high TABLE 66-EQUIVALENT EFFECTIVE SENSIBLE HEAT FACTORS FOR VARIOUS ELEVATIONS* For use with sea level psychrometric chart or tables Effective Sensible Heat Factor from Air 1000 Conditioning (28.86) Lo(ld E s t i m a t e - .60 .55 .50 Elevation 2000 (27.82) Equivalent .61 .56 .51 .62 .57 .52 (Feet) 3000 (26.82) Effective Sensible .63 .58 .53 end Barometric 4000 (25.84) Heat 5000 (24.89) Factor .64 .59 .54 Referred .6A .60 55 1 ESHF== IpI) (I-ESHF) +, (pc.1 W-IF) = barometric pressure at sea level = barometric pressure at high elevation PI ESHF = ESHF obtained from air conditioning load estimate ESHF,= 5TES FOR equivalent ESHF referred to a sea psychrometric chart or Table 66 TABLE level 66: 1. The required apparatus dewpoint for the high elevation is determined from the sea level chart or Table 65 .by use of the equivalent effective sensible heat factor. The relative humidity and dry-bulb temperature must be used to define the room condition when using this table because the above equation was derived on this basis. The room wet-bulb temperature must not be used because the wet- (Inches 6000 (23.98) to o of Hg) 7000 (23.09) et Installation 8000 (22.12) Sea Level Psychrometric .65 .61 .56 .66 .61 .57 .67 .62 .57 . 9000 (21.39) Chart or 10000 (20.57) Tables .68 .63 .5B - .69 .64 .59 bulb temperature corresponding to any particular condition, for example, 75 F db, 40% rh, at o high elevation is lower (except for saturation) than that corresponding to the same condition (75 F db, 4Oa/o rh) at sea level. For the same value of room relative humidity and dry-bulb temperature, and the some apparatus dewpoint, there is a greater difference in moisture content between the two conditions at high elevation than at sea level. Therefore, o higher apparatus dewpoint is required at high elevation for a given effective sensible heat factor. *Values obtained by use of equation W h e r e p0 Pressure 2. Air conditioning load estimate (See Fig. 44). The fack 1.08 ond .68 on the air conditioning load estimate should be multiplied by (Pl) the direct ratio of the barometric pressures -. Using this method, (P01 it is assumed that the air quantity (cfm) is measured at actual conditions rather than at standard air conditions. The outdoor and room moisture contents, grains per pound, must also be corrected for high elevations. 3. Reheat-Where the equivalent effective sensible heat factor lower than the shaded values in Table 65, reheat is required. is <;HAITER 8 . XI’I’LIED 1-149 I’SYCHKOMETRICS SYMBOLS ABBREVIATIONS adp apparatus tlcwpoint r5F (IZF) (OALH) (RF) (OASH) (RF) (OATH) lStu/hr bypass factor bypassed outdoor air latent heat bypassed outdoor air sensible heat bypassed outdoor air total heat british thermal units per hour crm cubic feet per minute (lb dp dry-bulb dcwpoint ERLH ERSH ERTH ESHF effective effective effective effective F fprn Fahrenheit degrees feet per minute SP 9-P GSHF GTH GTHS gallons per minute grains per pound grand sensible heat factor grand total heat grand total heat supplement OALH OASH OATH outdoor air latent heat outdoor air sensible heat outdoor air total heat rh RLH RLHS RSH RSHF RSHS R’ relative humidity room latent heat room latent heat supplement room sensible heat room sensible heat factor room sensible heat supplement room total heat Sat Elf SHF saturation efficiency of sprays sensible heat factor TLH TSH total latent heat total sensible heat wb wet-bulb In ba cim da ch,,, cfm,., c/m,, bypassed air quantity around apparatus dehumidified air quantity outdoor air quantity return air quantity supply air quantity specific enthalpy apparatus dcwpoint enthalpy effective surface temperature enthalpy entering air enthalpy lcaving air enthalpy mixture of outdoor and return air cnthalpy outdoor air enthalpy room air enthalpy supply air enthalpy room latent heat room sensible heat room total heat sensible heat factor t t Wfp t rffh t. L L~h tldb t Ito t itch t 1,L t Ml t,.,,, tm W Wodp WC, We, WI, W, W,* W,?ll W,, temperature apparatus dewpoint temperature entering dry-bulb temperature effective surface temperature entering water temperature entering wet-bulb temperature leaving dry-bulb temperature leaving water temperature leaving wet-bulb temperature mixture of outdoor and return air dry-bulb temperature outdoor air dry-bulb temperature room dry-bulb temperature supply air dry-bulb temperature moisture content or specific humidity apparatus dewpoint moisture content entering air moisture content effective surface temperature moisture content leaving air moisture content mixtureof outdoor and return air moisture content outdoor air moisture content room moisture content supply air moisture content l-150 PART 1. LOAD ESTIMATING PSYCHROMETRIC A. FORMULAS AIR MIXING EQUATIONS (Outdoor and Return Air) h, = (cfm”a x hoa) + (cfm,, x km) Cf%a w n = (cfm,, x WA + (cfs x WA cfm,, (2) TSH TLH GTH (3) RSHF = ESHF = ERSH = RSH + (BF) (OASH) + RSHS* (4) ERLH = RLH + (BF) (OALH) + RLHS’ (5) ERTH = ERLH + ERSH (6) TSH TLH GTH = RSH + OASH + RSHS” = RLH + OALH + RLHS” = TSH + TLH + GTHS” RSH RLH RTH RTH = 1.08-f = .68t x cfm, x (tr,,, - t,,) x cfm,, x (W,, - W,,) 4.45-t x cfm,, x (h,, - haa) = RSH $ RLH OASH = 1.08 x cfm,, (t,, - t,,) O A L H = .68 x cfm,, ( W , , - W,,) OATH = 4.45 x cfm,, (h,, - h,) (7) (8) c-9 (14) (15) (16) OATH = OASH + OALH (17) (BF) (OATH) = (BF) (OASH) + (BF) (OALH) (18) ERSH = 1.08 x ERLH = .68 x ERTH = 4.45 x cfm& x (t, - tadp) (1 - BF) cfq,J x cfmaa$ X (19) (W,, - Wadp) (1 - BF) (20) (h, - ha& (1 - BF) (21) *RSHS, RLHS and GTHS are supplementary loads due to duct heat gain, duct leakage loss, fan and pump horsepower gains, etc. To simplify the various examples, these plementary loads have not been used in the calculations. However, in actual practice, thesesupplementary loads should be used where appropriate. Chapter 7 gives the values for the various supplementary loads. Fig. 1, Chapter I, illustrates the method of accounting for these supplementary loads on the air conditioning load estimate. fItem H, page 151, gives the derivation of these air constants. fWhen no air is to he physically bypassed around the conditioning apparatus, cfnd, = cfm,,. . RSH RSH R S H + R L H =--RTH ERSH D. ERSH ERTH ERSH + ERLH = ~ GSHF = (23) (24) TSH =GTH TSH TSHfTLH (25) (26) (27) . BYPASS FACTOR EQUATIONS t edb BF = hdb - todp ; (1 _ BF) = (10) (11) (12) (13) (22) C. SENSIBLE HEAT FACTOR EQUATIONS B. COOLING LOAD EQUATIONS = = 1.08 X cfm& X (tcdb - tJdt,)“* = .68 x kfmaaz X (W,, - W,,)“* = 4.45 x cfm& x (hoa - h,,)** t edb - bdp - hdb t edb - bdp (28) BF = wl, - Wadp ; (1 - RF) = pI;za we, - Wadp ea a& (29) . BF = 4, - hadp ; (1 - BF) = ,“eaI,“ia (30) ea : he, - hodp a& E. TEMPERATURE EQUATIONS AT APPARATUS x toa) + (cfm, x LJ t edb tZdb (31) cfm,,f = bdp + BF (kdb - kdp) t ewb and hzob correspond to the (32) calculated values of h,, and hl, on the psychrometric chart. h ea xI = (cfm,, X hoa) + (cfm, x h,) (33) cf md h la = hadp + BF (he, - had,) (34) F. TEMPERATURE EQUATIONS FOR SUPPLY AIR t,, = bn - RSH 1.08 (cfm,,U (35) **When t,,,, W,,, and hm are equal to the entering conditions at the cooling apparatus, they may be substituted for tedb. W,, and hea respectively. (;t-lAI”I‘El< x. /\l’l’I.IIcl) I’SY(:HKOIZlE~I’III~:S 1-151 G. AIR QUANTITY EQUATIONS \_ _ Cf%,z ERSH = 1.08 x (1 - BF) (trm - t,,,) cfmcta = ~ ERLH ~33 x (I - RF) (JJ’,,,, - Wertp) cf%in ERTH = 4.45 x (1 - BF) (A,, - h,,J cfmd = cfmdd = cffn& = TSH I.08 (t,,b - t,rid TLH .@3 (We, - W,“) GTH 4.45 (II,.,‘ - h,,,) drn bn = cf7n,, Note: cftnd,, (,.flfl ,“R = ffm,,,, + cfttt,.,, (36) (37) H. Cf%,, = \ x (t,.,,, - Lzl .x’ .,’ ,’ ) ‘1. -. ~.-RLH- -.. /-‘Cfms,L = .68 x (W,.,,, - W,,,) cfm,, = 1.08 (38) RTH 4.45 x (h,.,,‘ - h,,) where .244 = specific heat of moist air at 70 F db ant1 50% rh, Btu/(deg F) (lb dry air) 60 = min/hr 13.5 = specific volume of moist ail at 70 F c!b and 50% rh .&3 Ix 60- x - 13.5 where (41) (42) (44) Ct*G) .244 x 60 13.5 (39) (43) (‘9 will bc less than cfrn,, only when air is physically bypassed around the conditioning apparatus. ,/.’ ,------a < cfmd,, DERIVATION OF AIR CONSTANTS 1.08 = (40) - 4.45 =K 13.5 where 1076 7000 60 = min/hr 13.5 = specific volume of moist air at 70 F db and 50% rh 1076 = average heat removal required to condense one pound of water vapor from the room air 7000 = grains per pound 60 = min/hr 13.5 = specific volume of moist air at 70 F db and 50% rh l-152 BIBLIOGRAPHY CHAPTER 2 5. Air Conditioning and Refrigeration Institute, Application Engineering Standard 530-56 Air Conditioning. 1956. 2. Heating, Ventilating and Air Conditioning Guide, Chapters 12 and 13, 1956. 3. Summer Weather Data - Marley Company. 6. TABLE 1 1. TABLE 4 1. Conditions for Comfort, by C . S. Leopold: H e a t i n g , Piping and Air Conditioning, June 1947, p. 117. 2. The Mechanism of Heat Loss and Temperature Regulation, by E. F. DuBois; Transactions of the Association of American Physicians, Vol. 51, 1936, p. 252. 3. The Relative Influence of Radiation and Convection upon the Temperature Regulation of the Clothed Body, by C. E. A. Winslow, L. P. Herrington, and A. P. Gagge; American Journal of Physiology, Vol. 124, October 1938, p. 51. 4. Reactions of Office Workers to Air Conditioning in South Texas, by A. J. Rummel, F. E. Giesecke, W. H. Badgett, and A. T. Moses; ASHVE Trans., Vol. 45, 1939, p. 459. 5. Shock Experiences of 275 Workers After Entering and Leaving Cooled and Air Conditioned Offices, by A. B. Newton, F. C. Houghten, C. Gutherlet, R. W. Qualley, and M. C. W. Tomlinson; ASHVE Trans., Vol. 44, 1938, p. 571. 6. How to Use the Effective Temperature Index and Comfort Charts, by C. P. Yaglou, W. H. Carrier, Dr. E. V. Hill, F. C. Houghten, and J. H. Walker; ASHVE Trans., 1932, p. 411. 7. Heat and IMoisture Losses from the Human Body and Their Relation to Air Conditioning Problems, by F. C. Houghten, W. W. Teague, W. E. Miller and W. P. Yant; ASHVE Trans., 1929, p. 245. 8. Thermal Exchanges Between the Human Body and Its Atmospheric Environment, by F. C. Houghten, W. W. Teague, W. E. Miller and W. P. Yant; American J o ~ w n a l of Physiology, Vol. 88, 1929, p. 386. TABLE 5 1. Heating, Ventilating and Air Conditioning Engineers Guide, Chapter 45, 1956. 2. Air Conditioning and Refrigerating Data Rook, 1955. 7. 8. 9. 10. Circuit i\naIysis Applied to Load Estimating, Part IIInfluence of Transmitted Solar Radiation, by H. B. Nottagc and C. V. Parmelee; ASHAE Transactions, Vol. 61, 1955, pp. 128-139. Thermal Circuit Analysis for Developing Application Engineering Information, hy Stanley F. Gilman and 0. \Villiam Clausen; Ifeating, Piping and Air Conditioning, June 1957, pp. 153-160. Temperature Changes in Refrigerated Rooms During I’ulldown P e r i o d , hy J. L. Threlkeld and T. Kusada; Refrigerating Engineering, July 1956, p. 35. Heat Transmission as influenced by Heat Capacity and Solar Radiation by F. C. Houghton, J. L. Blackshaw, E. M. Pugh, and P. McDemott; ASHAE Transactions, Vol. 38, 1932. p. 263. Cooling Load from Sunlit Glass, hy C. 0. Mackey and N. R. Gay: ASHAE Transactions, Vol. 58, 1952, p. 321. Analysis of an Air Conditioning Thermal Circuit by an, Electronic Differential Analyzer, by G. V. Parmelee, P. Vance, and A. N. Cerny; Heating, Piping and Air Conditioning, Sept. 1956, p. 117. TABLE 15 1. The Transmission of Solar Radiation Through Flat Glass Under Summer Conditions, by G. V. Parmelee; Heating, Piping and Air Conditioning, October-November 1945. Also ASHVE Trans., 1945, Vol. 51, p. 317. 2. Measurements of Solar Radiation Intensity and Determinations of its Depletion by the Atmosphere,aby H. H. Kimball; Monthly Weather Review, February 1930, Vol. 58, p. 52. 3. Review of United States Weather Bureau Solar Radiat i o n I n v e s t i g a t i o n s , b y I . F . H a n d ; Monthly Weather Review, December 1937, Vol. 65, p. 430. 4. Smithsonian p. LXXXIV. Meteorological Tables, 5th Revised edition, 5. Pyrheliometers and Pyrheliometric Measurements, by I. F. Hand; U. S. Weather Bureau, November 1946. 6. Proposed Standard Solar Radiation Curves for Engineering Use, by Parry Moon; Journal of the Fra nklin Institute, November 1940, Vol. 230, No. 5, pp. 586-617. 7. Performance of Flat Plate Solar Heat Collectors, by H. C. Hottel and B. B. Woertz; ASME Trans., February 1942, Vol. 64, pp. 91-104. 8. Where is the Sun?, by M. J. Wilson and J. M. Van Swaay; Heating and Ventilating, May and June 1942. 9. CHAPTER 3 TABLES 6 THRU 12 I. Heat Transfer, hy Max Jakoh, Vol. 1, John Wiley 9c Sons, Inc., New York, N. Y., 1949. 2. The Solution of Transient Heat Conduction Problems by Finite Differences, by G. A. Hawkins and J. T. Agnew. Purdue Univ., Eng. Exp. Sta. Bulletin No. 98, 1946. 3. Hydraulic Analogue for the Solution of Problems of Thermal Storage, Radiation, Convection and Conduction, by C. S. Leopold; ASHAE Transactions, Vol. 54, 1948, p. 389. 4. Circuit Analysis Applied to Load Estimating, by H. B. Nottage and G. V. Parmelee; ASHAE Transactions, Vol. 60, 1954, pp. 59-102. . Summer Weather Data and Sol-Air Temperature Study of Data for New York City, by C. 0. Mackey and E. B. Watson: Heating, Piping and Air Conditioning, Nov. 1944, p. 651. Also ASHVE Trans., 1945, Vol. 51, p. 75. 10. Summer Weather Data and Sol-Air Temperature Study of Data for Lincoln, Nebraska, by C. 0. Mackey and E. B. Watson; Heating, Piping and Air Conditioning, January 1945, p. 42. Also ASHVE Trans., 1945, Vol. 51, p. 93. TABLE 16 1. An Experimental S t u d y of Flat-type Sun Shades, by G. V. Parmalee, W. W. Aubele and D. J. Vild; Heating, Piping and Air Conditioning, January 1953. BIBLIOGRAPHY l-153 2. Design Data for Slat-type Sun Shades for Use in Load Estimating, by G. V. Pnrmalee and D. J. Vild; Heating, Piping and Air Conditioning, September 1953. 3. The Transmission of Solar Radiation Through Flat G l a s s u n d e r S u m m e r C o n d i t i o n s , b y G . V . Parmalee; Heating, Piping and Air Conditioning, October, November 1945. Also ASHVE Trans., 1945, Vol. 51, p. 317. 4. Heat Gain Through Western Windows With and Without Shading, by F. C. Houghten and David Shore; Heating, Piping and Air Conditioning, April 1941, p, 256. Also ASHVE Tmns., 1941, pp. 251.274. 5. Studies of Solar Radiation Through Bare and Shaded Windows, by F. C. Houghten, C. Gutberlet, and J. Blackshaw; ASHVE Trans., 1934, Vol. 40, pp. 101-116. 6. Solar Heat Gain Factors For Windows With Drapes, by N. Ozisik and L. F. Schutrum; paper presented at ASHR,lE meeting, Dallas, Texas, February, l-4, 1960. TABLES 35 AND 36 I. Houghten, S. I. Taimuty, C. Gutberlet and C. J. Brown; ASN VE Trans., Vol. 48, 1942, p. 369. 2. Measurements of Heat Losses From Slab Floors, by R. S. Dill, W. C. Robinson and H. E. Robinson; U. S. National Bureau of Standards, Report BMSIOJ, 1945. 3. Heat Losses Through Floors of Basementless Building; Heating and Ventilating, Vol. 42, Septemher 1945, p. 89. TABLE 38 1. Ice Formation on Pipe Surfaces by S. Lewis Elmer, Jr.; Re/rigerating Engineering, July 1332, p. 17. 2. Notes on the Formation of Ice on Pipe Surfaces by F. Raseri; Refrigerating Engineering, January 1933, p. 21. TABLE 40 TABLE 17 1. Heat Gain Through Glass Blocks by Solar Radiation and Transmittance, by F. C. Houghten, David Shore, H. J. Olson and Burt Gunst; ASHVE Trans., 1940, pp. 83-107: 1. 2. CHART 1 AND TABLE 18 1 Tables of Computed Altitude and Azimuth, Volume I to V inclusive (0” to 50” Latitude, 10” per volume) ; U. S. Navy Department - Hydrographic Office, No. 214. 2. Where is the Sun?, b y M . J. W i l s o n a n d J . M . V a n Swaay; Heating and Ventilating, May, June 1942. 3. 4. CHAPTER 5 TABLES 19 AND 20 1. Periodic Heat Flow - Homogeneous Walls or Roofs, by C. 0. Mackey and L. T. Wright, Jr.; Heating, Piping and Air Conditioning, September 1944, p. 546. Also ASHVE Trans., 1944, Vol. 50, p. 293. 5. 2. Periodic Heat Flow - Composite Walls or Roofs, by C. 0. Mackey and L. T. Wright, Jr.; Heating, Piping and Air Conditioning, June 1946, p. 107. 6. 3. Summer Cooling Load as Affected by Heat Gain Through Dry, Sprinkled. and Water Covered Roofs, by F. C. Houghten, H. T. Olson, and Carl Gutberlet; Heating, ?ping and Air Conditioning, July 1940. Also ASHVE :rans., 1940, p. 231. 4. Summer Weather Data and Sol-Air Temperature - Study of Data for New York City, by C. 0. Mackey and E. B. Watson; Heating, Piping and Air Conditioning, November 1944, p. 651. Also ASHVE Trans., 1945, Vol. 51, p. 75. 5. Summer Weather Data and Sol-Air Temperature - Study of Data for Lincoln, Nebraska, by C. 0. Mackey and E. B. Watson; Heating, Piping and Air Conditioning, January 1945, p. 42. Also ASHVE Trans., 1945, Vol. 51, p. 93. 6. Estimating Heat Flow Through Sunlit Walls, by C. 0. Mackey and L. T. Wright: Heating and Ventilating, March, April, May 1940. 7. Heat Transmission As Influenced by Solar Radiation, by F. C. Houghten Trans., 1932, pp. 231-284. Heat Capacity and and others; ASHVE 8. Heat Gain Through Walls and‘RooEs as Affected by Solar Radiation, by F. C. Houghten and others; ASHVE Trans., 1942, Vol. 48, pp. 21-105. 9. Heat Loss Through Basement Walls and Floors, by F. C. Solar Heat Gain Through Walls and Roofs for Cooling Load Calculations, by James P. Stewart; Heating, Piping and Air Conditioning, August 1948. 7. 8. 9. 10. Comparative Resistance to Vapor Transmission of Various Building Materials, by L. V. Teesdale; ASHVE Trans., Vol. 49, 1943, p. 124. The Relation of Wall Construction to Moisture Accumulation in Fill Type Insulation, by Henry J. Barre; Research Bulletin 271, Agricultural Experiment Station, Iowa State College: Ames, Iowa, April 1940. Vapor Transmission Analysis of Structural Insulating Board, by F. P. Rowley and C. E. Lund: Bulletin No. 22,, University of Minnesota Engineering Experiment Station, October 1944. The Diffusion of Water Vapor Through Various Building Materials by J. D. Babbitt; Canadian Journal of Research, Vol. 17, Sec. A, pp. 15-32, February 1939. Also Permeability of Building Paper to Water Vapor, by J. D. Babbitt; Canadian Journal of Research, Vol. 18, Sect. A, May 1940, pp. 90-97. Comparison of Methods for the Determination of Water Vapor Permeability, by Sears, Schlagenhauf, Givens and Yett; Paper Trade Journal, Vol. 118, TAPPI Section, pp. 27-28, January 20, 1944. Comparison of Methods for the Determination of Water Vapor Permeability, by C. J. Weber; Paper Trade Journal, Vol. 118, TAPPI Section, pp. 24-26, January 20, 1944. International Critical Tables. Vapor Barriers with Annotated Bibliography, by J. Louis York, University of Michigan, for Office of Production, Research and Development, War Production Board, Washington, D. C., February 1, 1945. How to Overcome Condensation in Building Walls and Attics, by L. V. Teesdale; Heating and Ventilating, April 1939. Moisture Condensation in Building Walls, by H. W. Wooley; National Bureau of Standards, Report BMS63. 1940. CHART 2 1. Preventing Condensation on Interior Building Surfaces, by P. D. Close: ASHVE Trans., Vol. 36. 1930. 2. Permissible Relative Humidities in Humidified Buildings, by P. D. Close; Heating, Piping and Air Conditioning, December 1939, p. 766. 3. Methods of Moisture Control and Their Application to Building Construction, by F. B. Rowley, A. B. Algren, and C. E. Lund: University of Minnesota Engineering Experiment Station, Bulletin No. 17. 4. Condensation Within Walls, by F. B. Rowley, A. B. Algren, and C. E. Lund; ASHVE Trans., Vol. 44, 1938. l-154 I’:\K’I’ CHAPTER 6 2. Thermal Exchanges Bctwcen the Human Body ant1 Its .\tmospheric Environment, hy F. C. Houghten, W. W. Tcague, W. E. Miller and W. I’. Yant; American Journal of Physiolog)~, Vol. 88, 1929, p. 386. 3. Heat and Moisture Losses From Men at Work and Application to Air Conditioning Problems, by F. C. Hough. ten, W. W. Tcague, W. E. Miller and W. 1’. Yant; ASHTIE Tf-arts., Vol. 37, 1931, p. 54. Air Contlitioning in Industry, by W. L. Fleischer, A. E. Staccy, Jr., F. C. Houghten and M. B. Ferberber; ASHVE Trans., Vol. 45, 1939, p. 59. Pl~psiological Basis of Medical Practice, Best and Taylor. TABLES 41, 43 AND 44 I. The Infiltration Problem of Multiple Entrances, Iiy j\. M. S i m p s o n and K . I\. A t k i n s o n ; ff-lcntitig, Pii)ing nntl A i r L’orulilio?xifq, .Jttne 1936. p, 345. 2 . Air Lcakagc Sttttlics on Metal Windows in a Modern OfFice Building, l)y F. C. Houghten and M. E. O’Connell; A.SI-IJ’l< Trnns., Vol. 34, 1928, p. 321. of Rolled Section Steel Windows, ‘.I The ~Vcathcrtightncss Iiy J, E . llmswilcr ant1 IV. C. R a n d a l l ; A S H V E Tra?rs., Vol. 34. 1928, p. 527. 4. El&t of Frame Calking and Storm Windows on Inliltmt i o n Aronntl a n d T h r o u g h W i n d o w s , b y \V. ;\I. Richtmann and C. Bmatz; ASHVE Trans., Vol. 34, 1928, p. 547. 5 . Air Inliltration T h r o u g h D o u b l e - H u n g W o o d !Vindows, by G . L . L a r s o n , D . IV. N e l s o n a n d R . W . Kubasta; ASHVE Trans., Vol. 37, 1931. p. 571. 6. I’rcssurc Differences :\cross Windows in Relation to Wind Velocity, by J. E. Emswiler and W. C. Randall: ASHVE Trans.. Vol. 36, 1930, p. 83. 7 . Air Infltration Through Steel Frame Windows, by D. 0. Rusk, V. H. Cherry and L. Boeltcr; ASHVE Trans., Vol. 39, 1933, p. 169. 4. 5. 6. Tnlrlcs, Factors and Formulas for Computing Respiratory Exchange and Biological Transformations of Energy, by Thornc M. Carpenter: Carnegie Institution, Washington, I). c . TABLE 49 1. \Vestinghouse current. data (1952) for 110~ to 125v, 60 cycle, a.c. . TABLE 50 I. Exhaust Hoods, by J. M. Dalla Valle. 2. TABLE 45 1. Code of Minimum Requirements for Comfort Air Conditioning; ASHVE Trans., Vol. 44. 1938, p. 27. 2 . Ventilation Requirements, by C. I?. Yaglou, E. C. Riley and D. J. Coggins; ASHVE Trans., Vol. 42, 1936, p. 133. 3. Control of Physical Hazards of Anesthesia, by R. M. Tovell and A. W. Friend; Canadian Medical Association Journal, 46560, 1942. 4 . Air Conditioning Requirements of An Operating Room and Recovery Ward, by F. C. Houghten and W. Leigh Cook Jr.; ASHVE Trans., Vol. 45, 1939, p. 161. 5. Code of Minimum Requirements for Heating and Ventilating Garages; ASHVE Trans., Vol. 41, 1935, p. 30. I. LOAD ESTIM:\‘I‘ING Reducing Heat Loads in Industrial Air Conditioning, by L. R. St. Onge; Refrigerating Engineering, January 1946, p. 35. TABLE 51 1. Helpful Hints to Fried Food Fame, by the Edison General Electric Appliance Company. . TABLE 53 1. hIotors and Generators, National Electric Manufacturers Associa,tion Standards Publication, No. MG-1 - 1955, Part 4, p. 10. TABLE 54 1. Heat Loss from Copper Piping, by R. H. Heilman; Heating, PiPing and Ai Conditioning, September 1935, p. 458. CHAPTER 7 TABLE 48 1. Heat and hloisture Losses From the Human Body and Their Relation to Air Conditioning Problems, by F. C. Houghten, W. W. Teague, \V. E. Miller and W. P. Yant; ASHVE Trans., Vol. 35, 1929, p. 245. CHAPTER 8 Rational l’sychrometric Formulae, by Willis H. Carrier; ASME Trans.. Vol. 23, 1911, p. 1005. Psychrometric Factors in the Air Conditioning C. M. Ashley; ASHVE Trans., Vol. 55, 1949. Estimate, by Part 2 AIR DISTRIBUTION 2-1 CHAPTER 1. AIR HANDLING APPARATUS This chapter describes the location and layout of air handling apparatus from the outdoor air intake thru the fan discharge on a standard air conditioning system. Construction details arc also included for convenience. Air handling apparatus can be of three types: (1) built-up app aratus where the casing for the conditioning equipment is fabricated and installed at or near the job site; (2) fan coil equipment that is manufactured and shipped to the job site, either completely or partially assembled; and (3) self-cont?. sd equipment which is shipped to the job site CL sJietely assembled. This chapter is primarily concerned with built-up apparatus; fan coil and self-contained equipment are discussed in Part 6. In addition to the built-up apparatus, items such as outdoor air louvers, dampers, and fan discharge connections are also discussed in this chapter. These items are applied to all types . of apparatus. Equipment location and equipment layout must be carefully studied when designing air handling apparatus. These two items are discussed in detail in the following pages. LOCATION The location of the air handling apparatus directly influences the economic and sound level asr.pcts of any system. ECONOMIC CONSIDERATION The air handling apparatus should be centrally located to obtain a minimum-first-cost system. In a few instances, however, it may be necessary to locate the apparatus, refrigeration machine, and cooling .tower in one area, to achieve optimum system cost. When the three components are grouped in one location, the cost of extra ductwork is offset by the reduced piping cost. In addition, when the complete system becomes large enough to require niore than one’ refrigeration machine, grouping the mechanical equipment on more than one floor becomes practical. This design is often used in large buildings. The upper floor equipment handles approximately the top 20 to 30 floors, and the lower floor equipment is used for the lower 20 to 30 floors. Occasionally a system is designed requiring a grouping of several units in one location, and the use of a single unit in a remote location. This condition should be carefully studied to obtain the op timum coil selection-versus-piping cost for the remotely located unit. Often, the cost of extra coil surface is more than offset by the lower pipe cost for the smaller water quantity resulting when the extra surface coil is used. SOUND LEVEL CONSIDERATIONS It is extremely important to locate the air handling apparatus in areas where reasonable sound levels can be tolerated. Locating apparatus adjacent to conference rooms, sleeping quarters and broadcasting studios is not recommended. The following items point up the conditions that are usually created by improper location; these conditions can be eliminated by careful planning when making the initial placement of equipment: 1. The cost oE correcting a sound or vibration problem after installation is much more than than the original cost of preventing it. 2. It may be impossible to completely correct the sound level, once the job is installed. 3. The owner may not, be convinced even after the trouble has been corrected. The following practices are recommended to help avoid sound problems for equipment rooms located on upper floors. 1. In new construction, locate the steel floor framing to match equipment supports designed for weights, reactions and speeds to be used. This transfers the loads to the building columns. 2. In existing buildings, use of existing floor slabs should be avoided. Floor deflection can, at times, magnify vibration’ in the building structure. Supplemental steel framing is often necessary to avoid this problem. 3. Equipment rooms adjacent to occupied spaces should be acoustically treated. 4. In apartments, hotels, hospitals and similar buildings, non-bearing partition walls should be separated at the floor and ceilings adjoining occupied spaces by resilient materials to avoid transmission of noise vibration. 2-2 . 5. Hearing walls adjacent to equipment rooms should be acoustically treated on the occupied side of the wall. LAYOUT Package equipment is usually factory shipped with all of the major equipment elements in one unit. With this arrangement, the installation can be completed by merely connecting the ductwork and assembling and installing the accessories. In a central station system, however, a complete, workable and pleasing layout must be made of all major components. This involves considerations usually not present in the unitary equipment installation. The shape and cross-sectional area of the air handling equipment are the factors that determine the INSULATE ,----CASINO - PART 2. AIK DISTRIBUTION dimensions oE the layout. The dehumidifier asscmbly or the air cleaning equipment usually dictates the overall shape and dimensions. A superior air handling system design has a regular shape. A typical apparatus is shown in I;ig. 1. The shape shown allows for a saving in sheet metal Cabrication time and, therefore, is considered to be better industrial design. Its clean lines give a more workmanlike appearance. From a functional standpoint, an irregular shaped casing tends to cause air stratification and irregular flow patterns. The most important rule in locating the equip ment for the air handling apparatus is to arrange the equipment alon,u a center line for the best air flow conditions. This arrangement keeps plenum pressure losses to a minimum, and is illustrated in Fig. 1. . - . ELEVATION -. F1c.1 -TYPICALCENTRALSTATIONEQUIPMENT CH/\I’TER I . AIR H/\NDl,ING 23 :\l’I’.\R,\TUS EQUIPMENT This section describes available central station apparatus cquipmcnt ant1 recommends suitable application of the various components. OUTDOOR AIR LOUVERS AND SCREEN I;ig. 2 illustrates outdoor air louvers that tninimizc the entry of snow and water into the equipment. It is impossible to completely eliminate all moisture with vertical louvers, and this is usually not neccssary. The screen is added to arrest most foreign materials such as paper, trash and birds. Often the type of screen required and the rncsh are specified hy codes. The screen and louver is located sufficiently above the roof to minimize the pickup of roof dust and the probability of snow piling up and subsequently entering the louver during winter operation. This height is tlcterminctl by the annual snowfall. HOWever, a minimum of 2.5 feet is recommended for most areas. In some locations, doors are added outside the louver for closure during extreme in- SCREEN AND BRACES MATERIAL SPECIFICATIONS Over-all Height Maximum Over-all Width Blades Frame Screen Screen Frame Braces *Equivalent strength aluminum may be Maximum 91%” 95” 22 U.S. gage steel* 18 U.S. gage steel* I/*” #16 wire mesh 1” x 1” x l/g” angle 1” x l/g” band iron substituted. LOUVER WIDTH (in.) 0 - 30 31-47 48 - 60 61 -95 Over 95 NUMBER O F SCREENSt NUMBER O F BRACES1 1 0 1 1 2 1 2 2 2 equal length louvers j-Screens over 60” high have center horizontal stiffening braces of 1” x 1” x l/g” angle. JBraces spaced evenly on front and back of louver and tack welded to blade edges. FIG. 2 - O UTDOOR A IR LOUVER AND SCREEN clement weather suc11 as hurricanes and blizzards. It is best to locate the outdoor air louver in such a manner that cross contamination from exhaust fan to louver does not occur, specifically toilet and kitchen exhaust. In addition, the outdoor air intake is located to minimize pulling air over a long stretch of roof since this increases the outdoor air load during summer operation. Chnrt I is used to estimate the air pressure loss at various face velocities when the outdoor louvers are constructed, as shown in Fig. 2. There are occasions when outdoor air must be drawn into the apparatus thru the roof. One convenient method of accomplishing this is shown in Fig. 3. The gooseneck arrangement shown in this figure is also useful for exhaust systems. .OUVER DAMPERS The louver damper is used for three important functions in the air handling apparatus: (1) to control and mix outdoor and return air; (2) to bypass heat transfer equipment; and (3) to control air quantities handled by the fan. Fig. 4 shows two damper blade arrangements. The single action damper is used in locations where the damper is either fully open or fully closed. A double-acting damper is used where control of air flow is required. This arrangement is superior since the air flow is tl~rotticd more or less in proportion to the blade position, whereas the single action type damper tends to divert the air and does little or no throttling until the blades are nearly closed. Outdoor and return air dampers are located that good mixing of the two air streams occurs. installations that operate 24 hours a day and located in a mild climate, the outdoor damper occasionally omitted. With the fan operating and the damper fully closed, leakage cannot be completely eliminated. C/tart 2 is used to approximate the leakage that occurs, based on an anticipated pressure difference across the closed damper. Table 1 gives recommendations for various louver dampers according to function, application, velocities and type of action required. PITTSBURGH PITTSBURGH LOCK SEAM ON 12” CENTERS ABOVE ROOF DEPENDS ON ANTICIPATED FLASHING WITH WELD CORNER SEAMS NOTE: Supplemental wind bracing may be required on larger intakes. FIG. 3 so On are is - GOOSENECK OUTSIDE AIR INTAKE l CI-i,\I'TlCI< I. AIR H;\NI~I.IN(; 2-s :\I'I',\I<.\'I'US CHART l--LOUVER PRESSURE DROP CHART 2-LOUVER DAMPER LEAKAGE ‘“I / / .2 / 1.0 / .9 / / / .6 / / Al- .02 .Ol 200 250 300 300 400 500 600 7w 7w 800 800 FACE VELOCITY (FPM) 1000 .. IO EXAMPLE 20 30 L E A K A G E R A T E (CFM/SO 40 FT) Solution; Pressure loss=.067 in. rg TABLE 1 --LOUVER DAMPERS FUNCTION OR LOCATION t I urn Outdoor Air APPLICATION I VELOCITY* (fom) Ventilation 500-800 The higher limit may be used with short outdoor air duct connection and long return air duct. May be single acting damper. Permissible system resistance and balance 500-800 Should be double acting when used for throttling. 500-800 Single acting damper may be used. 800-1200 May b e h i g h e r v e l o c i t y w i t h s h o r t r e t u r n duct and long outdoor air duct. Should be double octina damoer. 400-800 Should equal cross-sectional area of dehumidifier. Should be a double acting damper. 15DO-2500 Should balance resistance of dehumidifier plus humidifier face damper. Should be double actina. 1000-1500 Should balance resistance be double acting. I Maximum Outdoor Air All Outdoor Air I Dehumidifier Permissible system resistance and balance I Permissible system resistance and balance Return Air Control space Face conditions I Dehumidifier Heater Bypass System balance Bypass Balance I Fan Suction or Discharge or located in Duct REMARKS I I Available duct area * Recommend velocity through (I fully open damper. same or duct Use double acting at damper. heater. Should 2-6 I’:\lIT 7 “. .\IK DISTRIl~lJTlON CLEARANCE=;BLADE W I T H P L U S I$ CONNECTOR BLADE LINKAGE ROD ( 2 REQ’D FOR DAMPERS OVER 38" WIDE) ATTACHMENT HOLES FOR DAMPER MOTOR1 LINKAGE OR QUADRANT DOUBLE ACTING PARTIALLY OPEN SINGLE ACTING CLOSED SINGLE LOUVER DAMPER SET AT ANGLE 45' L 7 US. GAGE STEEL PLATE v=\ CONTINUOUS. REOUIRED FOR SPANS OF 12’0R MORE HAT CHANNEL SECTION B -6 SECTION A-A MULTIPLE LOUVER DAMPER ASSEMBLY (FOR ASSEMBLY EXCEEDING MAXIMUM DIMENSIONS) BLADES MATERIAL Maximum Over-all Height Maximum Over-all Width Maximum Blade Width Frame - Top and Bottom - Sides Blades Bearing Blade Linkage Rod Trunnion Blade Link (hfulti-section) SPECIFICATIONS 911/?” 50” 12” 3” x l/s” flat bar 3” x 7/s” x i/8” hat channel 16 U.S. gage steel Oil-retaining porous bronze 5/ 16” dia. CRS Die-formed steel Stainless steel bar FIG. DAMPER HEIGHT (in.) NUMBER OF BLADES To and incl. 12.11/16 12% thru 211/ 219/,, thru 311/s 31s/,, thru 41% 41o/rs thru 511/2 1 2 3 4 5 519& Sltj/,, 719/& 81D/is 6 7 8 9 thru thru thru thru 611/s 711/s 811/s 911/s 4--LOUVERDAMPERARRANGEMENTS CHAI’TEli I . ;\IK H A N D L I N G AI’I’.\KA’TUS \ STOP 2-7 HAT CHANNEL /’ /, I i ‘--- A L U M I N U M S P A C E R I WASHER I ,-LEAF CONNECTOR INTERCONNECTING - STOPS \= “4, DIA. ATTACHMENT HOLES SINGLE RELIEF DAMPER INTERCONNECTING SPOT WELD TO 1IBLADE BLAD E CONNECTION 7 U.S. GAGE PLATE CONTINUOUS. REQUIRED ONLY FOR SPANS OF 12’ OR MORE SECTION / 1 ALUMINUM %A$$ 1 LHAT CHANNEL SECTION A-A B-B MULTIPLE RELIEF DAMPER ASSEMBLY ( FOR ASSEMBLY EXCEEDING MAXIMUM DIMENSION) PRESSURE MATERIAL Maximum Over-all Height Maximum Over-all Width Maximum Blade Width Frame - Top and Bottom - Sides Blades Blade Linkage Rod Spacer Washer DROP SPECIFICATIONS 91%” 40” 31/2’1 3” wide, 11 gage black iron 3” x 7/g” x l/g” hat channel 22 B & S gage aluminum I/~” wide, 0.050” aluminum s/s” ID x l/s” OD aluminum FIG . 5 FACE - RELIEF DAMPER VELOCITY , (fpm) PRESSURE DROP (in. wg) 400 500 600 .067 .084 ,120 700 800 900 .160 .200 .256 RELIEF DAMPERS Figz~re 5 shows a typical relief damper. This accessory is used as a check damper on exhaust systems, and to relieve excess pressure from the building. AIR CLEANING EQUIPMENT A variety of air filtering devices is available, each with its own application. The pressure drop across these devices must be included when totaling the static pressure against which the fan must operate. Filters are described in detail in Part 6. HEATING COILS Heating coils can be used with steam or hot water. They are used for preheating, and for tempering or reheating. The air velocity thru the coil is determined by the air quantity and the coil size. The size ‘ray also be determined by a space limitation or by the recommended limiting velocity of 500 to 800 fpm. The number of rows and fin spacing is determined by the required temperature rise. Manufacturer’s data lists pressure drop and capacity for easy selection. Steam coils must be installed so that a minimum of 18 in. is maintained between the condensate outlet and the floor to allow for traps and condensate piping. Preheat Coils Non-freeze coils are recommended for preheat service, particularly if air below the freezing temperature is encountered. To reduce the coil first cost, the preheater is often sized and located in only the minimum outdoor air portion of the air handling apparatus. If a coil cannot be selected at the required load and desired steam pressure, it is better o make a selection that is slightly undersize than one that is oversize. An undersized coil aids in preventing coil freeze-up. The use of two coils for preheating also minimizes the possibility of freeze-up. The first coil is deliberately selected to operate with full steam pressure at all times during winter operation. In this instance, the air is heated from outdoor design to above the freezing temperature. The second coil is selected to heat ,the air from the freezing temperature to the desired leaving temperature. The temperature of the air leaving the second coil is automatically controlled. Refer to Pm-t 3, “Freeze-up Protection.” In addition to the normal steam trap required to drain the coil return header, a steam supply trap immediately ahead of the coil is recommended. These traps must he locatcd outsitlc the apparatus casing. Most coils are manufactured with a built-in tube pitch to the return header. If the coil is not constructed in this manner, it must be pitched toward the return header when it is installed. To minitnirc coil cleaning problems, filters should be installed ahead of the preheaters. Reheat or Tempering Coils Coils selected for reheat service are usually ovcrsized. In addition to the required load, a liberal safety factor of from 15% to 25% is recommended. This allows for extra load pickup during early morning operation, and also for duct heat loss which can be particularly significant on long duct runs. These coils are similar to preheat coils in that the tubes must be pitched toward the return header. COOLING COILS Cooling coils are used with chilled water, well water or direct expansion for the purpose of precooling, cooling and dehumidifying or for aftercooling. The resulting velocity thru the cooling coil is dictated by the air quantity, coil size, available space, and the coil load. Manufacturer’s da‘ta gives recommended maximum air velocities above which water carry-over begins to occur. SPRAYS AND ELIMINATORS Spray assemblies are used for humidifying, dehumidifying or washing the air. One item often overlooked when designing this equipment is the bleeder line located on the discharge side of the pump. In addition to draining the spray heads on shutdown, this line controls the water concentrates in the spray pan. See Part 5, Water Conditioning. Eliminators are used after spray chambers to prevent entrained water from entering the duct system. AIR BYPASS An air bypass is used for two purposes: (1) to increase room air circulation and (2) to control leaving air temperature. The fixed bypass is used when increased air circulation is required in a given space. It permits return air from the room to flow thru the fan without first passing thru a heat exchange device. This arrangement prevents stagnation in the space and maintains a reasonable room circulation factor. The total airway resistance for this type system is the sum of the total resistance thru the ductwork and air handling apparatus. Therefore, the resist- . CHAI’TEK I. i\ill H/\NIlLING ante thru the bypass is normally designed to balancc the resistance of the components bypassed. This can be accomplished by using a balancing damper and by varying the size of the bypass opening. The t’ollowing formula is suggested lor use in sizing the bypass opening : A = where: 2-9 i\l’I’i\KATUS cfm ;t1 5 8 1 -\i .0707 A = damper opening (sq Et) cEm = maximum required air quantity thru bypass h = design pressure drop (in. wg) thru bypassed equipment Temperature control with bypassed air is accom:d with either a face and bypass damper or a P couLrolled bypass damper alone. However, the face and bypass damper arrangement is recommended, since the bypass area becomes very large, and it is difficult to accommodate the required air flow thru the bypass at small partial loads. Even where a controlled face and bypass damper is used, leakage approaching 5% of design air quantity passes thru the face damper when the face damper is closed. This 5y0 air quantity normally is, included when the fan is selected. See Part 6 for systems having a variable air flow to determine fan selection and brake horsepower requirements. FANS Properly designed approaches and discharges from fans are required for rated fan performance ir %tion to minimizing noise generation.Figures 6 U~L / indicate several possible layouts for varying degrees of fan performance. In addition, these figures indicate recommended location of double width fans in plenums. Fans in basement locations require vibration isolation based on the blade frequency. Usually cork or rubber isolators are satisfactory for this service. On upper floor locations, however, spring mounted concrete bases designed to absorb the lowest natural frequency are recommended. The importance of controlling sound and vibration cannot be overstressed, particularly on upper floors. The number of fans involved in one location and the horsepower required for these fans directly determines the quality of sound and vibration control needed. Small direct connected fans, due to higher operating speed, are generally satislactorily isolaLec1 by rubber or cork. In addition, all types 0C fans must have Ilexible connections to the discharge ductwork and, where required, must have flexible connections to the intake ductwork. Details of a rccommendcd flexible connection arc shown in Fig. 8. Unitary equipment should be located near columns or over main beams to limit the floor deIlection. Rubber or cork properly loaded usually gives the required deflection for efficient operation. FAN MOTOR AND DRIVE A proper motor and drive selection aids in long life and minimum service requirements. Direct drive fans are normally used on applications where exact air quantities are not required, because ample energy (steam or hot water, etc.) is available at more than enough temperature difference to compensate for any lack of air quantity that exists. This applies, for example, to a unit heater application. Direct drive fans are also used on applications where I system resistance can be accurately determined. However, most air conditioning applications use belt drives. V-belts must be applied in matched sets and used on balanced sheaves to minimize vibration problems and to assure long life. They are particularly useful on applications where adjustments may be required to obtain more exact air quantities. These adjustments can be accomplished by varying the pitch diameter on adjustable sheaves, or by changing one or both sheaves on a fixed sheave drive. Belt guards are required for safety on all V-belt drives, and coupling guards are required for direct drive equipment. Figure 9 illustrates a two-piece belt guard. The fan motor must be selected for the maximum anticipated brake horsepower requirements of the fan. The motor must be large enough to operate within its rated horsepower capacity. Since the fan motor runs continuously, the normal 15% overload allowed by NEMA should be reserved for drive losses and reductions in line voltages. Normal torque motors are used for fan duty. APPARATUS CASING The apparatus casing on central station equipment must be designed to avoid restrictions in air flow. In addition, it must have adequate strength to prevent collapse or bowing under maximum operating conditions. , 2-10 I’.\I<‘I‘ SINGLE INLET FAN SINGLE INLET FAN 2. .\lli I)Is’I’l~Il~I1 ‘ION DOUBLE INLET FAN l-FT77c’““~ l-F-/ &=J NOTES I+2 NOTES l+2’ PLAN VIEW PLAN VIEW - F PLAN VIEW 1 / . i‘E NOTE 3 ELEVATION VIEW ELEVATION VIEW DIMENSIONS : C= DIAMETER OF FAN INLET 0.1; X C ELEVATION VIEW (SECTION) E = 45O MAXIMUM 30° PREFERRED F 8 36” MINIMUM FOR ACCESS DOOR . INLET CONNECTIONS ’ BEST GOOD T R A N S F O R M A T I O N I” I N 7 ” P R E F E R R E D , I” IN 4” PERMITTED NOTE 68 7 DIMENSIONS : A.l;xB TO 23x8 6 = FAN DISCHARGE OPENING, LARGEST DIMENSION DISCHARGE CONNECTIONS NOTES: 1. Fan should be centered in casing to provide good flow conditions. 5. Use square vaned elbow for best results, with take-off in opposite direction to fan rotation. 2. All equipment should be centered for best performance. G. Slope of 1” in 4” recommended for low velocity. 3. 7. Slope of 1” in 7” recommended for high velocity. Angle “E” is used to determine “F” distance between equipment and fan. 4 . R, = G” minimum. Vane spacing determined from Chart 6. F IG. (i-SINGLEFAN INLET , AND DISCHARGECONNECTIONS (:I-I.\l”I‘l~l< I. .\II< II.\NI)I.IN(; 2-11 .\I’I’.\I<,\~I’IJ,S ‘E I r NOTE I NOTE I NO TYPICAL VANE LOCATION R, = 6: R2OETERMINED FROM CHART 6. VANE SPACING A .I$B T O 24B B = LONGEST DIMENSION OF OUTLET OPENING NOTES: 1. Transformations to supply duct have maximum slope of 1” in 7”. 3. Do not install ducts so that the air flow is counter to fan roration. If necessary, turn fan section. 2. Square elbows with double thickness vanes may be substituted. 4. Transformations and units shall be adequately supported so no weight is on the flexible fan connection. I’,\i<.I‘ 2-F 2. \ II< I~IS’I‘l~II11J~I‘ION I” FLANGE AND HEM ALTERNATE POSITION OF B O L T ~- IMPREGNATED FABRIC - 20 U.S. GAGE STEEL - RECTANGULAR (FAN OUTLET) SEALING COMPOUND APPLIED BETWEEN FLEXIBLE CONNECTION AND FAN BEFORE ASSEMBLY l 5” 16 F L A N G E BOLT ON 4” CENTERS SEALING COMPOUND APPLIED BETWEEN FLEXIBLE CONNECTION AND CASING BEFORE ASSEMBLY * SHEET METAL SCREWS ON 12” CENTERS r BAND . --‘--x I ” x ;’ B A N D I R O\N - -SEALING *REQUIRED ON HIGH PRESSURE SYSTEMS COMPOUND APPLIED BETWEEN FABRIC, BAND AND 18 US. GAGE STEEL, BEFORE ASSEMBLY * \ ONLY, ROUND (F A N INLET) Frc.8 - FLEXIBLE CONNECTIUN SHEAVE SECTION A-A A.B.C.R,.Rp,+ R3 DIMENSIONS REQUIRED FOR CONSTRUCTION EXPANDED METAL AND 1” HEM REVERSED IN THIS SECTION S T A/ R L CENTER LINE OF SHEAVE AT U P T I E;ich sheet ol m:ltcri;ll sl10uItl he l’:~l)ricatctl a s a panel and joined together, ;ks illrlstratetl i n Pig /0, by standing sc~iiiis, bolted or riveted on 12 in. kentars. Nornially, SC’:IIIISperpendicular to air flow arc i>l;lcecl outsitlc of the casing. Side walls over 6 It high ;intl roof spns over fi I t witlc require supplemental reinforcing as shown in Tol~le -3. Diagonal angle braces as illustratcrl in Pig. 1 I may also lx rcquiretl. The rccommcntled construction of apparatus casings and connections between equipment components (except when mounted in the ducts) is 18 U.S. gage steel or 16 13 PCS gage aluminum. Aluminuni in contact with galvanized steel at connections to spray type equipment requires that the inside of the casing be coat& with an isolating material for a distance of 6 in. from the point of contact. CONNECTIONS TO MASONRY I\ coricretc curl) is rccomrncntletl t o p r o t e c t insul;itiori I’rorii clcterior;kting whcrc t h e aplxir:ltits casing joins the Iloor. I t a l s o provitlcs ;I iI nilorm burface f o r ;itt:iching t h e casing; t h i s conserves 1’:ibriT cation time. Figure I- illustrxtes the rcconimcntlctl methotl OF attaching a casing to the curl). When an equipment room wall is used as one sitlc of the ;ipparxtus, the casin,cr is attachctl 2s shown in Fig. 13. The degree oE tightness required for an appratus casing tlcpentls on the air conditioning application. DE FIG. I1 - APPARATUS CASING FIG. 10 - APPARATUS CASING S EAMS TABLE 2--SUPPLEMENTAL REINFORCING REINFORCING FOR APPARATUS SIDE WALL HEIGHT OR ROOF WIDTH w NUMBER OF ANGLES* ANGLE SPACING 6t0 a Et0 12 1 2 middle ‘Y3 points - over 12 variable 4 ft centers 3 & 4 panels 5 & 6 panels 7 a 0 panels ‘For lengths up to 12 ft., use 1 % x 1 ‘/z x ‘/a in. angle. CASTING LENGTHS For lengths over 12 ft., use I 3/4 x 13/ x %6 in. angle. CASING DIAGONAL ANGLE BRACES* (pairs) 1 2 3 2-14 l’;\l<-I‘ 2 . :\IK I~IS~I’RIlIIJ-I‘ION \ For instance, 011 a pull-thru system, leakage between the tlehuniiclificr ant1 the ran cannot be tolcratetl if the aplxirattls is in a humid non-contlitionetl space. .-11~0, as the negative pressure at the fan intake in,crcascs, the less the leakage that can I)e tolcratctl. If the apparatus is located in a return air plenum, normal construction as shown in I;ig.~. 1-3 ~ntl /3 cm be used. Corresponding construction practice for equipment requiring extreme care is shown in 18 U.S. GAGE STEEL CASING Ifx Figs. 14, 15 rind Ih. EXPANSION BOLT ON 12” CENTERS FIG. 12 - CONNECTION TO M ASONRY CURB In addition to the construction required for lcakage at seams, pipes passing thru the casing at cooling coil connections must bc scaled as sl~ow~i in Fig. f7. This applies in applications where the temperature difference between the room and supply Gr temperatures is 20 F ant1 greater. DRAINS AND MARINE LIGHTS Upkeep and maintenance is better on a.11 appxthat can be illuminated and easily cleaned than on one that does not have good illumination ant1 drainage. To facilitate this maintenance, marine lights, as well as drains, arc recommended as shown in Fig. 1. As a rule of thumb, drains should be located in the air handling apparatus wherever water is likely to accumulate, either in normal operation of the equipment or because of maintenance. Specific exampIeS are: 1. In the chamber immediately after the outdoor air louver where a driving rain or snow may accumulate. 2. Before and after filters that must be periodically washed. 3. Before and after heating and cooling coils that must be periodically cleaned. 4. Before and after eliminators because of backlash and carry-over due to unusual air eddies. Drains should not normally be connected directly to sewers. Instead, an open site drain should be used as illustrated in Pn~t 3. tus CASING RIVETED ON APPROXIMATELY 12” CENTERS ANGLE INSIDE OF CASING, ‘EXPANSION FIG. 13 - BOLT ON 12” CENTERS’ CONNECTION TO M ASONRY W ALL APPLY SEALING COMPOUND TO ANGLE BEFORE RIVET ON 6” CENTERS EXPANSlON FIG. 1-I - Low DEWPOINT BOLT ON 12” CENTERS MASONRY CONNECTIONS . CURB INSULATION Insulation is required ahead of the preheater ant1 vapor sealed for condensatidn during winter operation. Normally, the section of the casing from the preheater to the dehumidifier is not insulated. The dehumidifier, the fan and connecting casing must be insulated and vapor sealed; fan access doors are not insulated, however. The bottoms ant1 sides of the dehumidifier condensate pan must also be insulated, and all parts of the building surfaces that I (:11,\1”1’1-1< 1. /\IK lI,\KI)I.IN(~ 2-15 ,\1’l’.\l</\‘I‘IJS SERVICE J$luil)iiient service is csscntial :lntl sp;lcC mllst bc provided to acco~nplish this service. It is recommcntlctl th;tt minimuln cle:lr;inccs be m a i n t a i n e d so tll;tt ;ICCCSS to dI equipment is ;lvailablc. In x&Iitioli, provision sl~oultl be made so that equipment can be removed without dismantling the complete apparatus. Access must be provitlecl for heating and cooling coils, steam tr:lps, damper motors and linkages, control valves, bearings, fan motors, lans and similar components. Scrvicc access doors as illustrated in Fig. 18 are rcconimendetl, ant1 are 1oc:itcd in casing sections as shown in Fig. 1. conserve rioor space, the entrance to the cquipment room is olten located so that coils can be removed directly thru the equipment room doors. This arrangement requires less space than otherwise possible. IL‘ the equipment room is not arranged as described, space must be allowed to clean the coil tubes mechanically. This ap’plies to installations that have removable water headers. -, APPROXlMATELY EXPANSION I ;$i-aj y BOLT ON 12” CENTERS FIG . 15 -Low DEWPOINT MASONRY WALL CONNECTIONS CENTERS OUNO BRUSHED EMBLY FIG. I6 - SK,\LJN(; C S A ACCESS PANELS WITH SHEET METAL SCREWS, SEAL WITH SEALING COMPOUND. APPLY SEALING COMPOUND TO ANGLE BEFORE R’“ET’NG RIVET OR BOLT ON APPROXIMATELY 12” STANI)IN(; Skh\Ms SEE DETA,,. “8” -WINDOW TYPE SASH. HANDLE BUTT DR E A C H - T H OR U A C C E SOS SINGLE R ACTING HANDLE I SEE OETAIL “B”l “A”, SASH LOCK DOOR FRAME EXTENSION COLLAR U S E D W H E N D U C T IS INSULATED D E T A I L “8” ~WINDOW SASH LOCK EXTENSION COLLQR LENGTH (SEE DETAIL 8, IS OETERMlNED BY INSULInON THICKNESS USED ON 0”CT OR CASING. . DOOR, /GASKET MATERIAL SPECIFICATION 1. Door - 24 U.S. gage steel or 22 B &S gage aluminum. CASING 2. Frame - 24 U.S. gage steel or 22 B &S gage aluminum. 3. Extension Collar - Same gage as duct metal. 4. Formed Protection Angle - 18 U.S. gage steel or 16 B & S gage aluminum. DETAIL “A” DOUBLE ACTING DOOR HANDLES 3. Angle Brace - ls/4” x ls/a” x i/s” angle. 6. Butt Hinges - Steel. 7. Gaskets - Felt. F O R M E D PROTECTlON 8. Fastener a. Walk-thru door: Three double acting handles. b. ACCESSDOOR ANGLE BRACE Reach-thru sash lock. Single acting Walk-thru Access Doors Normal size - 22” W x 58” H Reach-thru Access Norma! sizes CASING INSIDE SECTION door: A-A F IG. 18 - ACCESS DOORS Doors W H 10” x 12” 12” x 16” 16” x 04” b handle with window 2-17 CHAPTER 2. AIR DUCT DESIGN The function of a duct system is to transmit air from the air handling apparatus to the space to be conditioned. To fulfill this function in a practical manner, the system must be designed within the prescribed limits of available space, friction loss, velocity, sound level, heat and leakage losses and gains. This chapter discusses these practical design criteria and also considers economic balance be/. . -n first cost and operating cost. In addition, it 0. .s recommended construction for various types of duct systems. GENERAL SYSTEM DESIGN CLASSIFICATION Supply and return duct systems are classified with respect to the velocity and pressure of the air within the duct. Velocity There are two types of air transmission systems used for air conditioning applications. They are called conventional or low velocity and high velocity systems. The dividing line between these systems is rather nebulous but, for the purpose of this chapter, the following initial supply air velocities X are offered as a guide: 1. Commercial comfort air conditioning a. Low velocity - up to 2500 fpm. Normally between 1200 and 2200 fpm. b. High velocity - above 2500 fpm. 2. Factory comfort air conditioning a. Low velocity - up to 2500 fpm. Normally between 2200 and 2500 fpm. b. High velocity - 2500 to 5000 fpm. Normally, return air systems for both low and high velocity supply air systems are designed as low velocity systems. The velocity range for commercial and factory comfort application is as follows: 1. Commercial comfort air conditioning - low velocity up to 2000 fpm. Normally between 1500 and 1800 fpm. 2. Factory comfort air conditioning - low velocity up to 2500 fpm. Normally between 1800 and 2200 fpm. Pressure Air distribution systems are divided into three pressure categories; low, medium and high. These divisions have the same pressure ranges as Class I, II and III fans as indicated: 1. Low pressure - up to 3% in. wg - Class I fan 2. Medium pressure - 3s/4 to 63, in. wg - Class II fan 3. High pressure - ($4 to 12% in. wg - Class III fan These pressure ranges are total pressure, including the losses thru the air handling apparatus, ductwork and the air terminal in the space. AVAILABLE SPACE AND ARCHITECTURAL APPEARANCE The space allotted for the supply and return air conditioning ducts, and the appearance of these ducts, often dictates system layout and, in some instances, the type of system. In hotels and office buildings where space is at a premium, a high velocity system with induction units using small round ducts is often the most practical. Some applications require the ductwork to be exposed and attached to the ceiling, such as in an existing department store or existing office building. For this type of application, streamline rectangular ductwork is ideal. Streamline ductwork is constructed to give the appearance of a beam on the ceiling. It has a smooth exterior surface with the duct joints fabricated inside the* duct. This ductwork is laid out with a minimum number of reductions in size to maintain the beam appearance. Duct appearance and space allocation in factory air conditioning is usually of secondary importance. A conventional system using rectangular ductwork is probably the most economical design for this application. ECONOMIC FACTORS INFLUENCING DUCT LAYOUT The balance between first cost and operating cost must be considered in conjunction with the available space for the ductwork to determine the best air distribution system. Each application is different and must be analyzed separately; only general rules or principles can be given to guide the engineer in selecting the proper system. The PART 2. AIR DISTRIBUTION 2-18 CHART J-DUCT HEAT GAIN VS ASPECT RATIO I 2. Ducts carrying small air quantities at a low velocity have the greatest heat gain. 3. The addition of insulation to the duct dccreases duct heat gain; for example, insulating _ the duct with a material that has a U value of .12 decreases heat gain 907”. It is, therefore, good practice to design the duct system for low aspect ratios and higher velocities to minimize heat gain to the duct. If the duct is to run thru an unconditioned area, it should be insulated. Aspect Ratio ,Jlowing items directly inHuencc the first and operating cost: 1. Heat gain or loss from the duct 2. Aspect ratio of the duct 3. Duct friction rate 4. Type of fittings Heat Gain or Loss The heat gains or losses in the supply and return duct system can be considerable. This occurs not only if the duct passes thru an unconditioned space but also on long duct runs within the conditioned space. The transfer of heat takes place from the space to the air in the duct when cooling, and from the air to the space when heating. An allowance must be made for duct heat gain for that portion of the duct in the unconditioned ‘-ace when estimating the air conditioning load. e method of making this allowance is presented in Part I, Load Estimating. This allowance for duct heat gain increases the cooling capacity of the air. This increase then requires a larger air quantity or lower supply air temperature or both. TO compensate for the cooling or heating effect of the duct surface, a redistribution of the air to the supply outlets is sometimes required in the initial design of the duct system. The following general guides are offered to help the engineer understand the various factors influencing duct design: 1. Larger duct aspect ratios have more heat gain than ducts with small aspect ratios, with each carrying the s a m e air quantity. Chart 3 illustrates this relationship. . The aspect ratio is the ratio of the long side to the short side of a duct. This ratio is an important factor to be considered in the initial design. Increasing the aspect ratio increases both the installed cost and the operating cost of the system. The installed or first cost of the ductwork depends on the amount of material used and the difficulty experienced in fabricating the ducts. Table 6 reflects these factors. This table also contains duct class, cross-section area for various round duct sizes and the equivalent diameter of round duct for rectangular ducts. The large numbers in the table are the duct class. The duct construction class varies from F to 6 and depends on the maximum duct side and the semi-perimeter of the ductwork. This is illustrated as follows: DUCT CLASS 1 2 3 4 5 6 MAX. SIDE SEMI-PERIMETER (in.) 6 - 171/ 12-24 26-40 24 - 88 48 - 90 go-144 (in.) 10 24 32 48 96 96 - 23 46 46 94 176 238 Duct class is a numerical representation of relative first costs of the ductwork. The larger the duct class, the more expensive the duct. If the duct class is increased but the duct area and capacity remain the same, the following items may be increased: I. Semi-perimeter and duct surface 2. Weight of material 3. Gage of metal 4. Amount of insulation required Therefore, for best economics the duct system should be designed for the lowest duct class at the smallest aspect ratio possible. Example I illustrates the effect on first cost of varying the aspect ratio for a specified air quantity and static pressure requirement. . CHAPTEK 2. AIR DUCT 2-19 DESIGN Example 1 -Effect of Aspect Ratio on First Cost of the Ductwork CHART 4-INSTALLEDCOST VS ASPECT RATIO Given: Duct cross-section area - 5 . 8 6 xl ft Space available - unlimited Low velocity duct system Find: The duct dimensions, class, surface area, weight and gage of metal required. Solution: I. Enter Tnble 6 at 5.86 sq ft and determine the rectangular duct dimensions and duct class (see tabulation). 2. Determine recommended metal gages from TaOles 14 and 15 (see tabulation). 3. Determine weight of metal from T&/e 18 (see tabulation). DIMENSION ARE:\ (in.) 69 ft) 94x 12 84x 13 7Gx 14 42 x 22 30 x 30 32.5 (round) DIMENSION (in.) 5.86 5.86 5.86 5236 5.86 5.86 ASPECT RATIO 7.8: 1 6.5:l 5.4: 1 1.9:1 1:l - SURFACE GAGE AREA (U.S.) . (sq ft / ft) DUCT CONSTR. CLASS 6 5 4 4 4 WEIGHT (lb/f9 When the aspect ratio increases from 1: 1 to 8: 1, the surface area and insulation requirements increase 70y0 and the weight of metal increases over three and one-half times. This example also points oi*- ‘hat it is possible to construct Class 4 duct, for tl . given area, with three different sheet metal gages. Therefore, for lowest first cost, ductwork should be designed for the lowest class, smallest aspect ratio and for the lightest gage metal recommended. Chart 4 illustrates the percent increase in in.stalled cost for changing the aspect ratio of rectangular duct. The installed cost of round duct is also included in this chart. The curve is based on installed cost of 100 ft of round and rectangular duct with various aspect ratios of equal air handling capacities. The installed cost of rectangular duct with an aspect ratio of 1: 1 is used as the 100% cost. Friction Rate The operating costs of an air distribution system can be adversely influenced when the rectangular duct sizes are not determinccl from the table of circular equivalents (Table 6). This table is used to obtain rectangular duct sizes that have the same friction rate and capacity as the cquivalcnt round duct. For example, assume that the required duct area for a system is 480 sq in. and the rectangular duct dimensions arc determined directly from this area. The following tabulation shows the resulting equivalent duct diameters and friction rate when 4000 cfm of air is handled in the selected ducts: DUCT DIM. (in.) 24 x 20 30x 16 48x 10 80x 6 1 EQUIV / F R I C T I O N j .-\SI’ECT R&ND RATE DUCT DIAM R.+TIO (in.) (in. wg/lOO ft) 23.9 23.7 22.3 20.1 ,090 ,095 ,125 ,210 1.2:1 1.9:1 4.8: 1 13.3:1 If a total static pressure requirement 0C 1 in., based on 100 ft of duct and other equipment is assumed for the above system, the operating cost increases as the aspect ratio increases. This is shown in C/MU-~ 5. Therefore, the iowcst owning and operating cost occurs where round or Spired-Pipe is used. II' round 2-%.I \ CHART 5-OPERATING COST VS ASPECT RATIO DUCT LAYOUT PART 2. AIR DlSTRlBUTlON CONSIDERATIONS There are several items in duct layout that should he considered before sizing the ductwork. These include duct transformations, elbows, fittings, take-offs, duct condensation and air control. Transformations ductwork cannot be used because of space limitations, ductwork as square as possible should be used. An aspect ratio of 1: 1 is preferred. Type of Fittings In general, fittings can be divided into Class A and Class B as shown in Table 3. For the lowest first cost it is desirable to use those fittings shown as Class A since fabrication time for a Class B fitting is approximately 2.5 times that of a Class A fitting. Duct transformations are used to change the shape of a duct or to increase or decrease the duct area. When the shape of a rectangular duct is changed but the cross-sectional area remains the same, a slope of 1 in. in 7 in. is recommended for the sides of the transformation piece as shown in Fig. 19. If this slope cannot be maintained, a maximum slope of 1 in. in 4 in. should not be exceeded. Often ducts must be reduced in size to avoid obstructions. It is good practice not to reduce the duct more than 20y0 of the original area. The recommended slope of the transformation is 1 in. in 7 in. when reducing the duct area. Where it is impossible to maintain this slope, it may be increased to a maximum of 1 in. in 4 in. When the duct area is increased, the slope of the transformation is not to exceed 1 in. in 7 in. Fig. 20 illustrates a rectangular duct transformation to avoid an obstruction, and Fig. 21 shows a round-to-rectangular transformation . to avoid an obstruction. TABLE 3-DUCT FITTING CLASSES CLASS A-NO VANED FITTINGS Any fitting with constant cross-section dimensions. FIG . 19 - DUCT TRANSFORMATION Any fitting with changing radius and constant width. Fittings with straight sides and seams. 63 CLASS B-ALL VANED FITTINGS Any fitting with concentric radii, and changing width. Any fitting with eccentric radii and changing width. NOTE: 1:7 slope is recommended for high velocity, 1:1 slope for low velocity. FK. 20 - RECTANGULAR D~JCT TRANSFORMA-VION -IX) AVOII) OKSTRIICTION l 2-21 CHAPTER 2. AIR DUCT DESIGN In some air distribution systems, equipment such as heating coils is installctl in the ductwork, Normally the equipment is larger than the ductwork and the duct area must be incrcascd. The slope 01 the tr~~nsformation piece on the upstream sitlc ol the equipment is limited to 30” as shown in I;ig. 22. On the leaving sick the slope should be not more than ,15”. Duct Reduction Increments L ,\cceptcti methods of duct design visually indicate ;I reduction in duct area after each terminal and branch take-off. Unless a reduction of at least 2 in. can be made, however, it is recommended that the original duct size be maintained. Savings in installed cost of as much as 25% can he realized by running the duct at the same size for several terminals. TOP VIEW /I : I NOTE: Angles shown are for low velocities. I:7 slope is recommended for high velocities. FK. 22 - DUCT TRANSFORMA.IXON E QUIPMENT Locating pipes, electrical conduit, structural members and other items inside the ductwork should always be avoided, especially in elbows and tees. Obstruction of any kind must be avoided inside II I \ -Sk&& L MAX ELEVATION Obstructions Ii : I ’ duct sizes should be even dimensions and all reci;tctions should be in 2 in. increments, preferably in one dimension only. The recommended minimum duct size is 8 in. x 10 in. for fabricated shop ductwork. 1. MAX IN THE WITH DIJCI high velocity ducts. Obstructions cause unnecessary pressure loss and, in a high velocity system, may also be a source of noise in the air stream. I/ 1 ‘I 1 jl ’ ’ I; IY I - 1 ; I ~MAXIMUM DUCT AREA REDUCTION - 20 % NOTE: I:7 slope is recommended for high velocity, I:4 slope for low velocity. - / , PART 2-22 2. AIR DISTRIBUTION 111 those few instances in which obstructions must 1x1~s thru the duct, use the Following rccommendations: 1. Cover all pipes and circular obstructions over ‘1 in. in diameter with an easement. Two typical easements are illustrated in Fig. 23. 2. Cover any flat or irregular shapes having a width exceeding 3 in. with an easement. Hangers or stays in the duct should be parallel to the air flow. If this is not possible, they should be covered with an easement. I;ig. 24 shows a tear drop-shaped easement covering an angle. Hanger “B” also requires an easement. 3. If the easement exceeds 20% of the duct area, the duct is transformed or split into two ducts. When the duct is split or transformed, the original area should be maintained. Fig. 25 illustrates a duct transformed and a duct split to accommodate the easement. In the second case, the split duct acts as the easement. When the duct is split or transformed, slope recommendations for transformations should be followed. 4. If an obstruction restricts only the corner of the duct, that part of the duct is transformed to avoid the obstruction. The reduction in duct area-must not exceed 20y0 of the original area. Elbows A variety of elbows is available for round and rectangular duct systems. The following list gives the more common elbows: FIG. 24 - EASEMENTS COVERING IRREGULAR SHAPES Rectangular Duct 1. Full radius elbow 2. Short radius vaned elbow 3. Vaned square elbow Round Duct 1. Smooth elbow 2. S-piece elbow 3. 5-piece elbow NOTE: 1:7 slope is recommended for high velocity, I:4 slope for low velocity. FIG. 25 - DUCT T RANSFORMED FOR EASEMENW FIG. 26 - FULL RADIUS RECTANGULAR ELBOW l CHAP-I-EK 2 . 2-23 AIK DUC’I‘ DESIGN The elbows are listed in order of minimum cost. ‘l’his sequence does not necessarily indicate the minimum pressure drop thru the elbow. Tal~lk 3 thtx I2 show the losses for the various rectangular and round elbows. Full radius elbows (Fig. 26) are constructed with :I throat radius equal to s/4 of the duct dimension in the direction of the turn. An elbow having this throat radius has an K/D ratio of 1.25. This is considered to be an optimum ratio. The short radius vaned r elbow and their location is determined from Chad 6. Exnmple 2 illustrates the use of Chart 6 in deter- mining the location of the vanes in the elbow in Fin 28. HEEL RADIUS D = 20” elbow is shown in Fig. 27. This elbow can have one, two or three turning vanes. The vanes extend the full curvature of the CENTER CINE RADIUS L I THROATRADIUS=3” FIG. 28 -RECTANCULARELBOWVANELOCAIION Exu,np/e 2 L Locating Vanes in a Rectangular E/bow Example 3 - locating Vanes in a Rectangular Elbow with a Square Throat Given: Rectanglllar elbow shown in Fig. 28. Throat radius (I?,) - 3 in. Duct dimension in direction of turn - 20 in. Heel radius (2X),) - 23 in. Find: 1. Spacing for two vanes. 2. R/D ratio of ell,ow. Soluticn: Given: Elbow shown in Fig. 29. Throat radius - none Heel radius - 20 in. Duct dimension in direction of turn - 20 in. Find: Vane 1. Enter Ckrt 6 at X, = 3 in. and I?,, = 23 in. Read vane spacing for R, and R, (dotted line on chart). R, = 6 in. RL = 12 in. 2. The centerline radius R of the elbow equals 13 in. Therefore R/D = 13/20 = .65. Although a throat radius is recommended, there may be instances in which a square corner is imperative. Chart 6 can still be used to locate the vanes. A th- It radius is assumed to equal one-tenth of the ht .I adius. Example 3 illustrates vane location in an elbow with a square throat. r spacing Solution: .\ssume a throat radius equal to .l of the heel radius: .l x 20 = 2 in. Enter Cl~art 6 at R, = 2, and R, = 20 in. Read vane spacing for R, and R2. R, = 4.5 in. R, = 9.5 in. In addition a third vane is located at 2 in. which is the assumed throat radius. EEL ADIUS (I+,) D I_ Rh T H R O A T R A D I U S t R+) F1c.27 -SHORTRADIUSVANEDELBOW F1c.29 -RECTANGULARELBOWWITH No THROATRADIUS CHART 6-VANE LOCATION FOR RECTANGULAR ELBOWS 2 3 4 5 6 THROAT RADIUS (Rt) (IN.) 7 8910 20 30 40 ej0 Y 60 70 t30 90 lO( .UU hPLE 2 NO. I OF 3 I i I \ ++ btnttttt +-I- l++H+n NO. I OF 2 ., , NO. 2 OF 3 / -: NO. I OF I ++ tttttfttt *-tnttm ++ +*tttt +t +tUt+Hi n7 Tnrrrm . . I 2 3 4 5 6 7 8 910 HEEL RADIUS (R”) (IN3 20 ’ 30 40 50 60 70 609OlW From Fan Engineering, Buffalo Forge Co. A vaned square elbow has either double or single thickness closely spaced vanes. Fig. 30 illustrates double thickness vanes in a square elbow. These lbows are used where space limitation prevents the use of curved elbows and where square corner elbows are required. The vaned square elbow is expensive to construct and usuaily has a higher pressure drop than the vanecl short radius elbow and the standard elbow (R/D = 1.25). Smooth elbows are recommended for round or .$pil-n-Pipe systems. Fig. 31 illustrates a 90” smooth elbow with a R/D ratio of 1.5. This R/D ratio is standard for all elbows used with round or SpzmPipe duct. FIG. 31-90” SMOOTH ELBOW CHAI’TEK 2 . AIR DUC’I 2-25 DESIGN ,4 S-piece elbow (I;;g. 32) has the same R/D ratio as a smooth elbow but has the highest pressure drop oE either the smooth or 5-piece elbow (Fig. 33). This elbow is second in construction costs and should be used when smooth elbows are unavailable. 5-piece elbow (I;i,.(7 33) has the highest first cost of all three types, It is used only when it is necessary to reduce the pressure drop below that of the 3-piece elbow, and when smooth elbows are not available. A 45” elbow is either smooth (Fig. 34) or 3-piece (Fig. 35). A smooth 45” elbow is lower in first cost and pressure drop than the S-piece 45” elbow. A 3-piece elbow is used when smooth elbows are not available. Take-offs There are several types of take-offs commonly used in rectangular duct systems. The recommendations given for rectangular elbows apply to takeoffs. Fig. 36 illustrates the more common take-offs. Fig. 36A is a take-off using a full radius elbow. In Figs. 36A and 36B the heel and throat radii originate from two different points since D is larger than D,. e principal difference in Figs. 36A and 36B is thar the take-off extends into the duct in Fig. 36B and there is no reduction in the main duct. Figure 36C illustrates a tap-in take-off with no part of the take-off extending into the duct. This FIG . 33 - 90" !&PIECE ELBOW FIG. 34 -,15” SMOOTH E~uow type is often used when the quantity of air to be taken into the branch is small. The square elbow take-off (Fig. 36D) is the least desirable from a cost and pressure drop standpoint. It is limited in application to the condition in which space limitation prevents the use of a full radius elbow take-off. A straight take-off .(l;ig. 37) is seldom used for duct branches. Its use is quite common, however, when a branch has only one outlet. In this instance it is called a collar. A splitter damper can be added for better control of the air to the take-off. There are two varieties of take-offs for round and Spira-Pipe duct systems: the 90” tee (Fig. 38) and the 90” conical tee (Fig. 39). A 90” conical tee is used when the air velocity in the branch exceeds 4000 fpm or when a smaller p r e s s u r e d r o p t h a n t h e straight take-off is required. Crosses with the takeoffs located at ISO”, 135” and 90” to each other are shown in Fig. 40. When the duct system is designed, it may be necessary to reduce the duct size at certain take-offs. The reduction may be accomplished at the take-off (Fig. 41) or immediately after the take-off (Fig. 42). Reduction at the take-off is recommended since it eliminates one fitting. FIG . 35 - 45’ ~-PIECE E LBOW l'.\l<'l‘ 2. ,\Il< I)IS'I'KII~UTION 2-s i FIG. 36 - TYPICALTAKE-OFFS I, -7 AIR FLOW c I, FIG. 37 -OUTLETCOLLAR r FIG. 38 - 90” TEE (;FI~\1’T1:II 2. \II< I>IJ(:‘I‘ I)I~:SI(;N 2-27 TABLE 4-MAXIMUM DIFFERENCE BETWEEN SUPPLY AIR TEMPERATURE AND ROOM DEWPOINT WITHOUT CONDENSING MOISTURE ON DUCTS (F) AIR VELOCITY IN STRAIGHT RUN OF DUCT (FPM)* AIR CONDlTlONS -‘tRROUNDING DUCT Bright Bright Bright P a i n t e d M e t a l P a i n t e d M e t a l P a i n t e d Metal 400 74 - 100 I 55 60 70 80 85 VALUE OF $ _ , 15 15 18 15 13 13 11 10 11 10 7 4 3 7 4 3 13 9 6 4 1 800 20 .90 1 - .66 1 .66 1200 8 7, 5 3 2 6 A 3 2 .A2 1 .49 2000 1600 8 7 4 5 4 2 2 11 10 9 ; 1 Bright Bright Bright Painted Metal Painted Metal Painted Metal 1 8 7 6 5 4 2 2 .31 1 5 5 A 3 2 2 7 6 5 A 3 2 2 1 .37 1 .2A 1 .31 3000 A A 3 3 2 5 A A 3 2 2 1 1 1 .20 3 3 2 2 2 1 1 1 1 .23 1 .15 *For elbows and other fittings, see Notes 4 and 7. E Q U A T I O N : tdp - ts, = (trm - tdp) (;where: 1 ) tdp = duct surface temp. assumed equal to room dewpoint. = supply air dry-bulb temp in duct. tsa trm = room d r y - b u l b t e m p . NOTES: 1. Exceptional Cases: Condensation will occur at a lower relative humidity than indicated in the table when f;! falls below the average value of 1.65 for painted ducts and 1.05 for bright metal ducts. The radiation component of fz will decrease when the ct is exposed to surfaces colder than the room air, as new a d wail. The convection component will decrease for the top of ducts, and also where the air flow is obstructed, such as a duct running very close to a partition. If either condition exists, use value given for CI relative humidity 5% less than the relative humidity in the room. If both conditions exist, use value given for 10% lower relative humidity. 2. Source: Calculated using film heat transmission coefficient on inside of duct ranging from 1.5 to 7.2 Btu/(hr) (sq ft) (deg F). The above equation is based on the principle that the temperature drop through any layer is directly proportional to its thermal resistance. It is assumed that the air movement surrounding the outride of the duct does not exceed approximately 50 fpm. 3. For Room Conditions Not Given: Use the above equation and the values of fz/U- 1 shown at the bottom of the table. 4. Application: Use for bare ducts, not furred or insulated. Use the values for bright metal ducts for both unpainted aluminum and unpointed galvanized ducts. Condensation at elbows, transformations and other fittings will occur at (1 higher supply air temperature because of the higher inside film heat transmission coefficient due to the air impinging against the elbow or fitting. For low velocity fittings, assume an equivalent velocity of two U = overall heat transmission coefficient of duct Etu/(hr) (sq ft) (deg F) f2 = film heat transmission coefficient on outside of duct, Btu/(hr) (sq ft) (deg F) = 1.65 for painted ducts and 1.05 for bright metal ducts. times the straight run velocity and use the above table. For higher velocity fittings where straight run velocity is 1500 fpm and above, keep the supply air temperature no more than one degree lower than the room dewpoint. Transformations having a slope less than one in six may be considered as a straight run. 5. Bypass Factor and Fan Heat: The air leaving the dehumidifier will be higher than the apparatus dewpoint temperature when the bypass factor is greater than zero. Treat this OS a mixture problem. Whenever the fan is on the leaving side of the dehumidifier, the supply air temperature is usually at least one to four degcxr higher than the air leaving the dehumidifier, and con be calculated using the fan brake horsepower. 6. Dripping: Condensation will generally not be severe enough to cause dripping unless the surface.temperature is two to three degrees below the room dewpoint. Note that the table is based on the duct surface temperature equal to the room dewpoint in estimating the possibility of dripping. It is recommended that the surface temperature be kept above the room dewpoint. 7. Elimination of Condensation: The supply air temperature must be high enough to prevent condensation at elbows and fittings. Occasionally, it might be desirable to insulate only the elbows or fittings. If moisture is expected to condense only at the fittings, apply insulation I%” thick usually sufficient) either to the inside or outside of duct at the fitting and for a distance downstream equal to 1.5 times the duct perimeter. If condensation occurs on a straight run, the thickness of insulation required ccm be found by solving the above equation for U. f 9o” REDUCING TEE Air Duct Control In low velocity air distribution systems the flow of air to the branch take-off is regulated by a splitter clampcr. The position of the splitter damper is atl,justcd by wise of the splitter rod. Splitter dampers for rectangular duct systems arc illustrated in Fig. 36. Pivot type dampers are sometimes installed in the branch line to control Bow. When these are used, splitter dampers are omitted. Splitter dampers arc preferred in low velocity systems, and pivot type or volume dampers are used in high velocity systems. Condensation D u c t s m a y “ s w e a t ” when the surface tempera’ ture of the duct is below the clewpoint of the surrounding air. Table 4 lists the maximum clifference between supply air temperature ant1 room dewpoint without condensing moisture on the duct for various duct velocities. See the notes below the table for application of the data contained in the table. Table 5 lists various U factors for common insulating material. It can be used in conjunction with Table 4 to determine required insulation to eliminate condensation. In high velocity systems, balancing or volume dampers are required at the air conditioning terminals to regulate the air quantity. TABLE 5-DUCT HEAT TRANSMISSION COEFFICIENTS TYPE DUCT INSULATION FINISH Uninsulated Sheet Metal NOW Metal lath and ploster-3~” Wood lath and plaster-%” Corkboard NOlV.2 None Plaster-%” Plaster--3/s” 2 2.9 - .12 Corrugated Asbestos Paper (air cell) None None 1 7 0.73 1 “‘ SO .34 311 Rock NOW3 None .OE .17 .27 - .21 .lO 39 .26 Cork Plaster-%” Plaster-%” Mineral Wool Blanket None None Glass NOM 1 2 NOIll? 1 05% Fiber Magnesia *Conductivity of insulating material only (per in.) tOvera U for still air outside duct and 1200 fpm inside duct. $Uninsulated Bare Duct. Air Velocity (fpm) Overoll u 1 400 1 800 1 1200 / 1600 / 2000 1 .98 1 1.08 / 1 1.19 / 1.14. 1.22 1.0 (:I I,\I”I‘I~l< 2. , \ I I< l)lJ(:‘l‘ Ifb:SI(;N 2-89 tlislril~titiott DUCT SYSTEM ACCESSORIES Pit-e th111~~33, xc ;tc(~cssories systcttt ;~ucss d o o r s ;ittd sound wltidt ttt;ty 1x2 I)itt do riot tti;itcri:tlly tlwx: ;trc sevctxl tttenl, tltc ;ttltlitioti;il twogtii~ctl htttjws ;il~ec:t ;iI~sorI~crs iii 2 tltc tlcsigtt, iii scrics. rcsist;ittcc wlieii sclectittg t-quit4 t ~Jttdcrwritet3 inst;tll;itiott duct ‘I‘lterc itttlcss ;iir Ilow tttttst I)e the r:ttI. Sotw;tlly 21itl ;tt-e I. 7‘!1e N;tlioti;tl tlcscrilx3 rltc: I)txctic-cs t w gcttetxl itt Ixttttlhlct o Ih~;ttd 01 I’irc cotlstritc~liott ;iiid Nh1;IJ ~wittcil~~tl tyjws !)O,\. 01’ fire tl;rtttlxt3 ductwork: rcctangttl;tr iti;iy IX tisetl pivot tlatttper t (I’ig. in either the vcrtic;tl I) whit It or Ilori/orit;ll Insitioii. Fire Dampers tiott ‘I‘ltc usctl i t t tu.t;tngrtI;tr l;or this :trtxttgco systetn. loc21 or st;ttc cotistructioti cotlcs di’ct;ttc tltc IISC, loc;t- 01 l i t - c d;itttIxx-s [or FUSIBLE LINK ANGLE 211 air L’. T l i e rec:t:ingular i)c usctl loliver litx d;itttpet- which only in tlic horizontal position (Fig. BAR, STOP TRUNNION BAR I (PREFERRED) I I I I I / BLADE ANGLE STOP SPRING CATCH REQ’O FOR BLADES i OPEN I CONNECTION I I I _ SLEEVE CLOSED / lMaximum Over-all Height Maximum Over-all Width Minimum Sleeve Length Sleeve Blade - Up to 18” 181/,,” - 36” 36%/’ and over Frame Bearing Support Trunnion Bar Spring Catch I I 4 I^ POSITION MATERI.\L 0 &I / ~~~~~~~ POSITION H A T SPECIFICATIONS 30” 50” 113A” 10 U.S. gage steel 16 U.S. gage steel 12 U.S. gage steel 7 U.S. gage steel 3” x 7/8” x I/B” hat channel Die cast steel 0.040” bronze spring stock SECT A-A C H A N N E L ntay JJ). ’ I’;\K-1‘ 2. 2-30 ICig7rve f i illustr;ites ;I l)ivot lirc damper Access Doors ;\ccess doors or ;~cccss pnels 3re required in duct s y s t e m s bclorc ;1nt1 :lftcr ecjuipent installctl i n ducts. Access panels ;ire dso rquirctl lor access to l’tlsiblc links in fire thnlpers. l)lS’I‘IIIl~U’l‘ION DUCT DESIGN lor round duct systcrus. ‘I‘his thnip III:L~ bc used in either the hori~ont;tl or vertic;il I,osition. r\lK I‘liis sectioll presents the ncceswry thta lor ticsig-!litig l o w :tntl high v e l o c i t y duct systetris. This ht;i inrlutlcs the stanthrcl :lir Iriction chrts, rccoii~tuentlctl tlcsign velocities, losses tllru ell~ows ;~ncl fittiligs, ;tntl the con1u1on niethotls ol designing the air distribution systenls. 1nforni;ltion is given also lor ev:~lu;~ting the effects ol duct hwt glin ;intl ;tltitutle 011 systeni tlesjgn. - SLEEVE 1” x 1” x L’ AANGLE STOP . FUSIBLE LINK BAA BLADE LINKAGE TIE ROD (2 REO’D FOR BLADES OVER 38’1 -COUNTERWEIGHT r FUSISLE LINK -LINKAGE AIR FLOW (ACCEPTABLI CLEVIS (PREFERRED) ANGLE TIE BAR - - FRAME BEARING SUPPORT HAT CHANNEL - BLADE CONNECTOR / HOLDING TAB -SPRING CATCH ( 2 REO’DFOR BL&DESOVER 38’) MATERIAL Maximum Over-all Height Maximum Over-all Width Minimum Sleeve Length Maximum Blade Width Sleeve Blade Frame Bearing Support Blade Linkage Roci Trunnion Bar Spring Catch SPECIFICATIONS 91 I/” 10” 1 l!V,” 6” 10 U.S. gage steel 16 ‘J. S. gage steel 3” x 7/g” x l/g” hat channel B/tu” die CRS Die cast steel 0.040” bronze spring stock F1c.44-- KECTANCULAK LOUVER FIKEDAMPF.K . ANNEL FRAME SECTION A-A o f these f a c t o r s i s illustrated i n t h e f o l l o w i n g equation: Friction Chart In any duct section thru which air is flowing, tllcrc is a continuous loss of pressure. This loss is callctl duct friction loss and depends on the Iollowing: wh&: AP = frjction loss (in. wg) ~~ f= interior surface roughness (0.9 for galvanized duct) L = length of duct (ft) d = duct diameter (in.), equivalent diam. for rectangular ductwork V = air velocity (fpm) 1. Air velocity 2. Duct size 5. Interior surface roughness 4. Duct length Varying any one of these four factors inlluences the friction loss in the ductwork. The relationship SHEET METAL ACCESS PANEL FUSIELE LINK BAR -_.. - PREFERRED) i (ACCEPTABLE) BLADE - SLEEVE DUCT ( 2 REP D FOR DIMPERS L MINIMUM ’ ’ ANGLE CLIPS TO FASTEN SLEEVE IN FIRE PARTITION OPEN POSITION MATERIAL CLOSED POSITION SPECIFICATIONS Maximum Diameter 48” Minimum Sleeve Length Sleeve Blade - Up to 18” 18r/,,” to 36” 364/,~” and overt Trunnion Bar Spring Catch 151/2” plus wall thickness* 10 U. S. gage steel 16 U.S. gage steel 12 U.S. gage steel 7 U. S. gage steel Die cast steel 0.040” bronze spring stock “.4ccess panel in sleeve. Length 8” plus wall thickness when access panel is in duct. tRequires s” x v,” x l/s” angle blade stiffener. I;IG. 45 - ROUNLI PIVOT FIRE DAMPER SECTION A-A ( BEARING ASSEMBLY) The u2tut’ll ducts Ior a high velocity supply systeiii I‘his equation is used to construct the standard Ii;tve tllc S;IIIIC tlcsigll \,elocity reroilllllcncl~~tions as I’rictiorl chart (Cli,r~‘l 7) IMsctl o n galv;tiliKtl duct listed ill I-‘trl~le 7 lor a low velocity systcln, unless antI air at i0 1; ;IIICI ‘L!).!)‘L ill. I-lg. *l‘his chart nlay I)c wsctl I’or systciils h;intllirlg ;iir Iroin Y) t o I20 I; csteiisive sound trcatnlcnt is provitlctl to iise liighc1 velocities. ;intl lor altitiitlcs ril) t o 2000 It without correcting tlrc air tleflsity. 1’0clge 5 9 c o n t a i n s tlie data l’or tleFriction Rate signing high altitude air tlistril)ution systcllls. ‘l‘hc friction rate 011 the Uriction chart is given in terins 0L inches of w a t e r per 1 0 0 l‘t ol equivalent Air Quantity length of duct. To cletcrmine the loss in any section The total s u p p l y air qiiailtity and t h e q u a n t i t y of ductwork, the total equivalent length in tilat required for each spiw is deterniincd from the air section is multiplictl by the friction rate which gives contlitioniilg load estiniatc in P0rl /. the friction loss. The total equivalent length ol duct includes all elbows and fittings that may IX in the Duct Diameter duct section. Tnbles 9 thm 2,3 w e used to cvaluatc Trrble 6 gives the rectangular duct sizes for the the losses thrii various duct systein elements in terms various equivalent duct diameters shown on C1tnl.t 7. of equivalent length. The duct sections including Next to the tliametcr is the cross-section area of the these elements arc measured to the ccnterlinc of the . round duct. The rectangular ducts shown for this elbows in the duct section as illustrated in Fig. -Ih. cross-section area handle the same air quantity at The fittings are measured as part of the duct section the same friction rate as the equivalent round duct having the largest single dimension. listed. Therefore, this cross-section area is less than Velocity Pressure the actual cross-section area of the rectangular duct The friction chart shows a velocity pressure condetermined by multiplying the duct dimensions. In version line. The velocity pressure may be obtained selecting the rectangular duct sizes from Tnble 6, by reading vertically upward from the intersection the duct diameter from the friction chart or the of the conversion line and the desired velocity. duct area as determined from the air quantity and TnOle 8 contains velocity pressures Eor ‘the corvelocity may be used. responding velocity. However, rectangular duct sizes should not be determined directly from the duct area without Flexible Metal Conduit using Table 6. If this is done, the resulting duct Flexible metal conduit is oEten used to transmit sizes will be smaller, and velocity and friction loss the air from the riser or branch headers to the air will be greater, for a given air quantity than the conditioning terminal in a high velocity system. design values. The friction loss thru this conduit is higher than Air Velocity The design velocity for an air distribution system depends primarily on sound level requirements, first cost and operating cost. Table 7 lists the recommended velocities for supply and return ducts in a low velocity system. These velocities are based on experience. In high velocity systems, the supply ducts arc normally limited to a maximum duct velocity Of 5000 fpm. Above this velocity, the sound level may become objectionable and the operating cost (friction rate) may become excessive. Selecting the duct velocity, therefore, is a question of economics. A very high velocity results in smaller ducts and lower duct material cost but it requires a higher operating cost and possibly a larger fan motor and a higher class fan. If a lower duct velocity is used, the ducts must be larger but the operating cost decreases and the fan motor and fan class may be smaller. \ . thru round duct. Chart 8 gives the friction rate for 3 and 4 in. flexible metal conduit. (Cotttinzud on page 38) OFFSETS XOTE; FIG. *All measurements are center line. Fittings are measured as part of the duct having largest single dimension. 46 - GUIDE FOR b~/IEA.SURINC l>uc:-I’ LEN(;.I.HS (:t-I.\I’-I‘EI< 2. .\II< I)u(:~I‘ 2-33 i)kSl<;N CHART 7-FRICTION LOSS FOR ROUND DUCT .02 .03 0 4 .05 0 6 08 0 . 1 0.15 0.2 0.3 0.4 0.5 0.6 0.8 1.0 1.5 2.0 3.0 4.0 5.0 6.0 100 60000 70000 60 000 50000 40000 30000 20000 15 0 0 0 IO 000 IO 0 0 0 JO0 8000 7000 7000 6000 6000 5000 5000 4 0 0 0 5 k 3 0 0 0 ;-’ F 2000 0 LL z 1 5 0 0 I500 5 s (I z 1000 1000 800 700 800 700 600 600 500 5 0 0 4 0 0 4 0 0 80 70 60 50 “” .02 .03 .04 .05 .06 .08 0.1 0.15 0.2 0.3 0.4 0.5 0.6 0.8 I.0 I.5 2.0 FRICTION LOSS (IN. WG PER 100 FT OF EQUIVALENT LENGTH) 3.0 4 . 0 5.0 6.0 I TABLE 7 SIDE 10 6-DUCT 6 8 1 wea Diam q ft in. DIMENSIONS, 1 SECTION AREA, CIRCULAR AND DUCT CLASSt 12 14 Area Diem *q ft in. 4rea Diam sq ft in. lo AleCi Diam sq ft in. Art?0 rqft Diam in. .39 .45 8.4 9.8 12 9.1 0.7 .65 .77 10.9 11.9 14 .52 9.8 1.5 .91 12.9 .94’ 1.09 EQUIVALENT 16 4rect 1q ft Diam in. DIAMETER,* 18 20 22 ArMI Diam sq ft in. Area Diam sq ft in. ArMI Di.¶m sq ft in. 13.1 14.2 1.20 15.3 16 1.24 .I 1.45 16.3 1.67 17.5 18 20 1.40 .O 1 .b3 17.3 1.87 18.5 2.12 19.7 1.54 .8 1.81 18.2 2.07 19.5 2.34 20.7 2.61 21.9 15.9 16.6 1 .b9 17.6 1.99 19.1 2.27 20.4 2.57 21.7 2.86 22.9 3.17 14.6 1.38 1.50 1.83 18.3 2.14 19.8 2.47 21.3 2.78 22.6 3.11 23.9 3.43 24.1 25.11 1.26 / 15.2 1.61 17.2 1.97 19.0 2.31 20.6 2.66 22.1 3.01 23.5 3.35 24.8 3.71 26.1 1.33 15.6 1.71 7.7 2.09 19.6 2.47 21.3 2.86 22.9 3.25 24.4 3.60 25.7 4.00 27.1 13.6 1.41 16.1 1.8 .3 2.22 20.2 2.64 22.0 3.06 3.46 25.2 3.89 26.7 4.27 28.0 14.0 1.48 16.5 1.93 a8.8 2.36 20.8 2.81 22.7 3.25 23.7 24.4 3.68 26.0 4.12 27.5 4.55 28.9 14.4 1.58 17.0 2.03 2.49 21.4 2.96 23.3 3.43 25.1 26.7 4.37 28.3 4.81 29.7 2.61 21.9 3.11 23.9 3.63 25.6 3.89 4.09 29.0 5.07 30.5 :::: 2.76 22.5 3.27 24.5 3.80 26.4 4.30 27.4 28.1 4.58 :f:; I ::;: 4.84 29.8 5.37 31.4 40 2.88 23.0 3.43 25.1 3.97 27.1 4.52 28.8 5.07 30.5 5.62 32.1 42 44 2.98 23.4 4.15 4.33 27.t 28.: 29.4 23.9 25.6 26.1 4.71 3.11 3.57 3.71 4.90 30.0 5.31 5.55 31.2 31.9 5.86 6.12 32.8 33.5 46 3.22 24.3 3.88 24.8 4.49 4.65 28.2 29.: 30.6 3.35 26.7 27.2 5.10 48 50 5.30 31.2 5.76 5.97 32.5 33.1 6.37 6.64 34.2 34.9 22 24 .78 12.0 1.08 14.1 .84 12.4 1.16 26 .89 12.8 28 .95 13.2 30 32 .Ol I.07 34 I.13 36 I.18 38 1.23 4 19.3 II%-% 3.46 25.2 4.03 4.15 27.6 5.51 31.8 6.19 33.7 6.87 35.5 52 2.22 20.2 2 . 9 1 23.1 3.57 25.6 430 28.1 5.72 32.4 6.41 34.3 7.14 36.0 54 2.29 20.5 2.98 23.4 28.5 5.17 30.1 5.90 32.9 6.64 34.9 7.38 36.8 2.38 20.9 3.09 23.8 26.1 26.5 4.43 56 3.71 3.83 4.55 28.9 5.31 31.: 6.08 33.4 6.87 35.5 7.62 37.4 29.3 5.48 31.; 6.26 33.9 7.06 36.0 7.87 38.0 58 60 3.94 4.06 26.9 27.3 4.68 4.84 29.8 5.65 32.: 6.50 34.5 7.26 36.5 8.12 38.6 64 4.24 27.9 5.10 30.6 5.91 33.. 6.87 35.5 7.71 37.6 8.59 39.7 68 4.49 28.7 5.37 31.4 6.26 33.! 7.18 36.3 8.12 38.6 9.03 40.7 72 76 4.71 29.4 5.69 4.91 30.0. 5.86 32.3 32.8 6.60 6.83 34.q 35.. 7.54 7.95 37.2 38.2 8.50 8.90 39.5 40.4 5.17 30.8 6.15 5.41 31.5 7.22 7.54 36.. 37.. 8.55 39.0 39.6 9.21 9.75 88 5.58 32.0 6.41 6.64 33.6 34.5 8.29 84 34.9 7.87 38.1 8.94 40.5 10.1 92 5.79 32.6 38.t 9.39 41.5 10.4 43.8 5.90 33.0 35.6 36.2 8.12 96 100 6.91 7.14 36.9 8.40 8.50 39.. 39.. 9.70 9.80 42.1 7.40 10.8 11.3 44.5 45.5 12.1 12.3 47.2 47.6 104 7.60 37.4 8.90 40.. 10.3 43.5 11.6 46.2 13.0 48.8 108 7.90 112 8.10 38.0 38.6 9.20 9.50 41. 41 .I 10.6 10.9 44.0 44.7 12.0 12.3 47.0 47.5 13.4 13.8 49.6 50.3 9.80 10.0 42. 42.E 11.3 48.1 49.1 14.3 51.3 13.1 14.4 51.5 10.3 43.5 .5 0 .o 6 11.9 46.7 12.6 11. 13.4 49.6 15.0 52.4 10.6 44.1 12.1 47.1 13.8 50.4 15.5 53.3 12.5 12.8 47.9 48.5 14.1 14.5 50.9 51.6 15.8 16.2 53.9 54.5 48.8 49.4 14.7 15.2 52.0 52.9 lb.5 55.0 55.6 4.15 80 27.6 116 120 124 128 I I 132 136 140 144 L *Circular equivalent diameter (d,). Calculated from d, = 1.3 $b) 10.4 43.8 10.8 44.6 X’ 13.0 13.3 b) “’ M I Large numbers in table ore duct class. 42.5 , lb.8 . (:f f.\l”f‘El< 2. :\IK DUC’1’ I~BSIGN 2-35 TABLE 6. DUCT DIMENSIONS, SECTION AREA, CIRCULAR EQUIVALENT DIAMETER.* AND DUCT CLASS? (Cont.) 24 26 AWXI Diem s q f t in. Alea *q ft 28 Diam A r e a in. rqft 36 32 Diam in. 4rea '4 ft Diam in. 4rea sq ft Diam in. Al=30 Diam s q f t in. 30 AK70 Dism s q f t in. 40 AlfK! sq ft Diam in. ArMI Diam sq ft in. 16 18 20 22 24 3.74 26.2 26 4.03 27.2 4.40 4.33 28.2 4.74 29.5 5.10 4.68 29.3 5.07 30.5 5.44 30.6 31.6 5.86 32.8 4.94 30.1 5.37 31.4 5.79 32.6 6.23 33.8 5.68 5.24 33.6 6.60 b.99 34.8 7.06 36.0 7.54 37.2 35.8 36.7 7.46 37.0 7.95 38.2 8.46 7.87 38.0 8.37 39.2 8.89 a.29 39.0 8.81 40.2 9.34 0.68 9.21 9.61 41 .l 10.1 10.5 28.4 36 5.58 31 .o 32.0 5.69 5.94 32.3 6.15 33.0 6.51 3a 5.86 32.8 6.38 3C.2 6.07 40 6.15 33.6 6.71 35.1 7.22 42 6.45 34.4 7.03 35.9 7.58 36.4 37.3 44 6.75 35.2 7.34 36.7 7.91 38.1 8.50 39.5 9.07 39.9 40.8 46 7.03 35.9 37.4 a.25 36.6 38.2 8.59 8.85 9.25 40.3 41.2 41.7 7.30 7.58 38.9 39.7 9.48 48 50 7.63 7.95 37.3 a.25 38.9. 0.90 40.4 9.61 42.0 9.89 10.3 42.6 43.5 52 54 7.87 38.0 a.55 39.6 9.25 41.2 9.98 11.4 38.7 8.85 40.3 9.61 42.0 10.4 10.7 11.0 44.3 a.16 a.42 42.8 43.6 45.0 11.8 39.3 9.16 41.0 9.94 42.7 10.7 44.3 11.4 45.8 58 60 8.63 39.8 9.48 41.7 1 0 . 3 43.4 11.0 a.a9 9.75 42.3 10.5 44.0 11.4 11.8 12.2 64 9.43 40.4 41.6 45.0 45.8 46.6 10.3 43.5 11.2 45.4 12.1 47.2 56 68 9.98 72 10.4 42.8 43.8 76 10.8 44.9 80 84 -- 11.5 12.0 46.0 - _ 34.6 35.5 7.34 9.43 41.6 9.89 42.6 9.80 41.4 42.4 10.4 43.6 10.5 11.0 43.8 44.8 42.0 10.3 43.4 10.8 44.6 11.4 45.8 43.0 10.7 11.3 45.6 43.9 44.8 11.1 11.6 44.3 45.2 46.5 47.4 46.8 47.8 46.1 11.8 12.2 11.9 12.4 13.0 48.8 45.7 46.5 12.1 12.6 47.1 12.7 13.5 49.7 48.0 13.2 48.3 49.2 12.2 47.3 13.0 48.8 13.7 50.1 Es? 12.6 48.1 13.4 49.6 14.2 51.0 15.0 52.4 47.3 13.0 48.9 13.8 50.4 14.6 51.8 12.9 48.7 13.8 50.4 14.7 52.0 15.5 53.4 15.5 16.5 53.3 55.0 14.6 51.8 56.6 58.0 10.9 10.9 44.7 11.8 46.6 12.8 48.4 13.7 50.2 15.6 53.5 16.5 55.0 17.5 11.5 12.0 45.9 12.4 47.0 13.1 47.8 13.5 49.7 14.4 51.5 16.4 17.4 56.5 49.0 14.1 50.8 15.1 52.7 17.3 54.9 56.3 18.3 57.9 18.3 19.3 12.6 13.2 48.0 13.7 49.2 14.2 50.1 14.7 52.0 15.8 53.9 17.0 55.8 18.1 46.9 17.3 17.7 18.5 57.0 58.2 18.9 19.7 59.3 60.7 50.1 1 4 . 8 55.0 56.3 19,2 20.1 13.7 53.2 54.3 57.6 58.9 47.9 15.4 16.1 16.5 12.5 51.1 52.2 60.1 20.9 62.0 12.9 48.7 14.2 51.1 15.5 53.4 16.7 55.4 18.0 57.4 19.2 59.4 20.5 61.3 21.8 63.2 56.2 57.3 la.6 19.2 58.5 59.4 19.7 20.6 60.2 61.5 21.1 62.2 22.7 64.5 21.6 63.0 23.4 65.5 60.5 21.4 62.6 22.7 64.5 24.1 66.5 61.4 22.0 63.5 23.5 65.7 24.8 67.5 c .6 2 . 3 22.5 64.3 24.5 67.0 25.7 96 13.3 49.5 14.8 52.2 15.9 54.0 17.2 100 13.9 50.6 15.0 52.5 16.7 55.3 17.9 104 14.6 51.8 15.8 53.9 17.1 56.0 18.6 108 14.8 52.1 16.2 54.6 17.6 56.8 19.2 58.5 59.4 19.9 20.5 112 15.1 52.7 16.8 55.5 18.3 58.0 19.7 60.1 21.1 116 15.8 120 124 16.2 16.6 128 17.1 17.4 56.0 132 136 *Circular 39.4 40.4 61 .O 21.2 62.4 2 2 23.0 . 1 65.0 24.0 24.8 66.3 67.5 25.6 26.5 68.5 69.7 68.7 27.1 70.5 26.2 69.4 20.2 71.9 27.2 28.2 70.6 71.9 29.0 73.0 74.0 56.4 la.9 .9 20.3 61 .l 22.0 63.6 23.5 65.7 24.8 17.8 18.4 57.1 19.4 o .6 5 8 . 1 1 9 . 8 c 3. 3 20.9 62.0 22.7 64.5 21.6 63.0 23.2 65.4 24.2 25.2 66.7 68.0 26.1 26.5 la.8 19.3 58.8 20.3 22.3 22.6 64.0 64.4 25.6 26.3 68.6 69.5 27.3 28.2 72.0 28.7 29.8 72.6 74.0 30.2 24.5 66.0 67.0 70.8 59.5 20.8 61.1 61.8 23.7 56.5 32.0 74.5 76.6 17.9 57.3 19.7 60.2 21.4 62.7 23.0 65.0 25.1 67.9 26.9 70.3 28.7 72.6 30.5 74.8 32.6 77.3 18.5 58.2 20.3 61.0 22.3 64.0 24.1 66.5 25.9 69.0 27.5 71.1 29.4 73.5 31.5 76.0 33.4 78.3 la.8 58.7 20.6 61.5 22.7 64.5 24.8 .Gr.; 67.5 26.3 69.5 28.2 72.0 29.9 74.1 32.0 76.6 34.0 79.0 55.2 diameter (d,.). Calculated from d, = 1.3 2 lb) T ttorge numbers in table are duct class. 67.5 69.2 20.3 17.3 equivalent 53.9 54.6 59.5 69.8 29.8 , 2-06 i’,\R’I‘ 2. i\IK I)IS-I‘KIBU?‘ION TABLE 6. DUCT DIMENSIONS, SECTION AREA, CIRCULAR EQUIVALENT DIAMETER,* AND DUCT CLASS1 (Cont.) . I 60 I 64 I 68 I I 72 76 42 44 46 48 50 52 54 56 58 60 23.5 65.7 64 25.0 67.7 26.7 70.0 28.3 72.1 30.2 74.4 31.8 33.5 76.4 78.4 26.5 69.7 72 28.0 76 129.5 74.1 71.7 29.9 73.6 I 31.q6.1 80 31.0 75.4 33. 84 32.5 77.2 34.8 68 88 96 *Circular 34.0 137.0 79.0 82.4 equivalent 36.3 139.8 diameter 78.8 80.9 37.7 83.2 8.1 135.2 79.9 37.0 80.4 I 37.4 82.8 39.6 85.3 82.4 39.2 84.8 41.4 81.6 84.2 41.1 86.8 43.4 87.2 89.3 47.5 93.4 85.5 (d,.). 142.1 Colculoted . 87.9/44.6 from I II I / 33.8 35.7 18.6 I d,. = 90.5/ 1.3 (q + b).‘” 41.7 ttarge I 87.5 numbers in table ore ducr &~I. I I I (:I-I.\l”l‘~:K 2. ,\lli I)lJ(:‘I’ DI3I(;N 237 TABLE 7-RECOMMENDED MAXIMUM DUCT VELOCITIES FOR LOW VELOCITY SYSTEMS (FPM) CONTROLLING CONTROLLING FACTOR NOISE GENERATION Main Ducts APPLICATION FACTOR-DUCT Main Ducts I FRICTION Branch Ducts Return SUPPlY SUPPlY Return 600 1000 800 600 600 Average Stores Cafeterias 1800 2000 1500 1600 1200 Industrial 2500 3000 1800 2200 1500 Residences Theatres Auditoriums General Offices High Class Restaurants High Class Stores Banks TABLE 8-VELOCITY PRESSURES VELOCITY PRESSURE (in. wg) .Ol .02 .03 .04 .OS .06 .07 08 .09 .lO .11 .12 .13 .14 .15 .16 .17 .lB .19 .20 .21 .22 .23 .24 .25 .26 .27 .2a VELOCITY (Ft/Min) 400 565 695 800 895 980 1060 1130 1200 1270 1330 1390 1440 1500 1550 1600 1650 1700 1740 1790 1830 1880 1920 1960 2000 2040 2080 2120 VELOCITY PRESSURE (in. wg) .29 .30 .31 .32 .33 .34 .35 .36 .37 .38 VELOCITY (Ft/Min) 2150 2190 2230 2260 2300 2330 2370 2400 2440 2470 I .39 .40 .41 .42 .46 .47 .48 , __ __ 2500 2530 2560 2590 I 2710 2740 2770 .53 .54 .55 .56 NOTES: 1. Data for standard air (ZP.W in. Hg and 70 2. Data derived from the following equation: 2910 2940 2970 2990 I I VELOCITY PRESSURE (in. wg) .58 .60 .62 .64 .66 .68 .70 .72 .74 .76 .7a .BO .a2 .a4 .86 .68 .90 .92 .94 .96 .9a 1.00 1.04 1.08 1.12 1.16 1.20 1.24 VELOCITY (Ft/Min) I I 3050 3100 3150 3200 3250 3300 3350 3390 3440 3490 3530 3580 3620 3670 3710 3750 3790 3840 3880 3920 3960 4000 4080 4160 4230 4310 4380 4460 I I I VELOCITY PRESSURE (in. wg.) 1.28 1.32 1.36 1.40 1.44 1.48 1.52 1.56 1.60 1.64 1.68 1.72 1.76 1.80 1.84 1.88 1.92 1.96 2.00 2.04 2.08 2.12 2.16 2.20 2.24 2.28 VELOCITY (Ft/MinI .I A530 I I I I 4600 4670 4730 4800 A870 49% 5000 5060 5120 5190 5250 5310 5370 5430 5490 5550 5600 5660 5710 5770 5830 5080 -I-5940 i690 6040 FJ where: V = velocity in fpm. h, = pressure difference termed “velocity head” (in. wg!. CHART 8-PRESSURE DROP THRU FLEXIBLE CONDUIT FAN CONVERSION l.OSS OR GAIN In addition to the calculations shown for deter- mining the required static pressure at the fan discharge in l<xnrtzple -t, a fan conversion loss or gain must be included. This conversion quantity can be a significant amount, particularly on a high velocity system. It is determined by the following equations. if the velocity in the duct is higher than the fan outlet velocity, use the following lormuIa for the additional static pressure required: Loss = I.][(&)” -(&)‘] where T’,t = duct velocity V, = fan outlet velocity Loss = in. wg If the fan discharge velocity is higher than the duct velocity, use the following formula for the credit taken to the static pressure required: Cain = 35 [(,&>l - (‘$&)L’] DUCT SYSTEM ELEMENT FRICTION LOSS Friction loss thru any fitting is cspt~essetl in terms of equivalent length of duct. This method provides units that can be used with the friction chart to tletertnitic the loss in ;I section of duct containing elbows and fittings. ToOk l-7 gives the friction losses for rectangular elbows, and TuOle /I gives the losses for standard round elbows. The friction losses in TU/ll6% If o,rd 12 arc given in terttts of additional equivalent length of straight duct. This loss for the elbow is added to the straight run of duct to obtain the total equivalent lctigth of duct. ‘l‘lie straight run . of duct is measured to the intersection of the center line of the fittings. Fig. 46 gives the guides for mcasuring duct lengths. Tables 9 and 10 list the friction losses for other size elbows or other R/D ratios. Table 10 presents the friction losses of rectangular elbows and elbow combinations in terms of L/D. Table 10 also includes the losses and regains for various duct shapes, entrances and exits, and items located in the air stream of the duct. This loss or regain is expressed in the number of velocity heads and is represented by “n”. This loss or regain may be converted into equivalent length of duct by the equation at the end of the table and added or subtracted from the actua1 duct length. Table 9 gives the loss of round elbows in terms of L/D, the additional equivalent length to the diam- . eter of the elbow. The loss for round tees and crosses are in terms of the number of velocity heads (“n”). The equation for converting the loss in velocity head to additional equivalent length of duct is located at the bottom of the table. In high velocity systems it is often desirable to have the pressure drop in round elbows, tees, and crosses in inches of water. These losses may be obtained from Chart 9 for standard round fittings. DESIGN., METHODS The general procedure for designing any duct system is to keep the layout as simple as possible and make the duct runs symmetrical. Supply terminals are located to provide proper room air distribution (Chapter j), and ducts are laid out to connect these outlets. The ductwork should be located to avoid structural members and equipment. The design of a low velocity supply air system may be accomplished by any one of the three following methods: 1. Velocity reduction 2. Equal friction 3. Static regain The three methods result in diBerent levels of accuracy, economy and use. The equal friction meti~od is rccommencled return and exhaust air systems. for LOW VELOCITY DUCT SYSTEMS Velocity Reduction Method The procedure for designing the duct system by this method is to select a starting velocity at the fan discharge and make arbitrary reductions in velocity down the duct run. The starting velocity selected sl~oulcl not exceed those in Table 7. Equiva- CHAPTER 2. AlIt DIJCT DESIGN 2-39 TABLE O-FRICTION OF ROUND DUCT SYSTEM ELEMENTS ELEMENT CONDITION L/D RATIO’ R/D = 1.5 9 90’ 5-Piece Elbow R/D = 1.5 12 45’ j-Piece Elbow \ R/D = 1.5 6 45’ Smooth Elbow R/D = 1.5 90’ Smooth Elbow 90’ j-Piece Elbow i fv-.;/ 90’ Miter Elbow r-l ELEMENT 90’ Tae$ and 90°, 135O B 180” Voned Not Vaned Thru 22 65 CONDITION VALUE OF n f 0.2 0.5 Fl- :*; 1 . 4.0 Crossf v2 _ Pressure Loss 4.5 I ga B r a n c h = nhv, .lO 44 i.21 1.47 Pressure Loss Thru B r a n c h = nhvs 90’ Conical Tee and 180” Conical Cross v2 -= Vl N o t e s o n page 42. I. 0.5 1.0 g 0.2 0.5 1.0 1.2 .;g$ PAKT 2. AIR DISTRIRUTION 2-m TABLE IO-FRICTION OF RECTANGULAR DUCT SYSTEM ELEMENTS CONDITIONS ELEMENT Rectangular Radius Elbow .5 W/D .5 1 3 Voned X0 Elbow No Rectangular Square Elbow Elbow Double 9 11 14 18 5 7 8 ‘12 j 1 4 4 5 7 18 12 10 Radius Elbow 10 8 7 8 7 7 7 7 6 :/PO times value for imilar 90’ elbow 60 Single Thickness Turning Vanes 15 Double Thickness Turning Vanes ‘0. s=o 15 S=D 10 22 15 Elbow S=D 1.25* For Both Elbow Elbow 14 18 30 40 Vanes Elbow W/D = 2. RI/D 1 1.50 1 .oo 20 W/D = 1, R/D = 1.25* Double or Unvaried S = D W/D = 1. R/D, = 1.25* W/D = 1, R/D = .75 Elbow 1 2 3 Double / 33 45 80 125 6 Double 1.25* t L/D R a t i o hc ’ QY! Double RATIO R/D r P Rectangular Vonad Radius 1 L/D 16 Direction of Arrow = 1.25*, RZ/D = .5 Reverse Direction w W / D = 4. R/D = 1.25* for both elbows Direction of Arrow Reverse Direction 18 l TABLE lo-FRICTIONOF RECTANGULAR DUCT SYSTEM ELEMENTS (Contd) ELEMENT CONDITIONS Transformer V A L U E O F nl I .15 v:! = Vl S . P . Loss = n h v , ’ Expansion 5O A3 .89 .93 ~ZPI .20 .40 .bO “il” Angle “a” 15O 10” .74 33 .87 S.P. Regain = n(hvt 20” .b0 .78 .a4 .b2 .74 A2 30° .52 40” .45 .b8 .79 .b4 .77 - a 1 3o” 1 45’ 1 60” n 1 1.02tt 1 1.04 1 1.07 S . P . L o s s = n(hvl -hvl) ttslope 1 0 i n 4“ Abrupt .35 Entrance S . P . Loss = n h v , Bellmouth Abrupt Entrance Exit -4 S.P. Loss or Regain Considered Zero I Bellmouth Re-Entrant Exit Entrance . S.P. Loss = nhv, Sharp Edge Round Orifice 0 *I -*2 .85 A,/A, 1 0 1 ” 1 2.5 ) .25 2.3 1 ) SO 1.9 1 1 .75 1.1 S.P. Loss = nhvl Abrupt Contraction v,/v~ ” 1 1 1 1 1.00 0 T -5 1 1 0 1.34 .25 1.24 1 1 .50 .9b 1 1 .75 .52 S.P. Loss = nhv? Abrupt Expansion V~/VI n 1 1 1 1 .20 .32 .40 .48 .bO .48 1 1 1 1 S . P . R e g a i n = nhv, Pipe Running Thru Duct E/D n 1 I .lO 1 .25 1 .50 .20 I .55 I 2.00 .lO / .25 I .50 .7 1 1.4 I 4.00 S.P. Loss = nhv, Bar Running Thru Duct E/D ” 1 1 S.P. Loss = nhv, Easement Over Obstruction f-1 E/D n 1 .I0 1 .25 .07 1 .23 S.P. Loss = nhv, Notes on page 42. 1 1 .50 .90 .80 .32 NOTES FOR TABLE 9 NOTES FOR TABLE 10: *L and D are in feet. 0 is the elbow diameter. L is the additional equivalent length of duct added to the measured length. The equivalent length L equals D in feet times the ratio listed. tThe value of n is the loss in velocity heads and may be converted to additional equivalent length of duct by the following equation. $1.25 is standard for on unvoned full radius elbow. tl and D are in feet. D is the duct dimension illustrated in ths drawing. L is the additional equivalent length of duct added to the measured duct. The equivalent length 1 equals D in feet times the ratio listed. Lcnxhv;flOO where: 1 = h, = . tThe value n is the number of velocity heads or differences in velocity heads lost or gained at a fitting, and may be converted to additional equivalent length of duct by the following equationr additional equivalent length, ft velocity pressure at Vz, in. wg (conversion line on Chart 7 or Table 8). h t = f r i c t i o n loss/100 (Chart 7). f t , d u c t d i a m e t e r a t VZ, i n . wg where: L hv = additional = velocity pressure for VI or V2, line on Chart 7 or Table 8). ” = value for tee or cross equivalent length, hf = friction loss/100 ft. duct (Chart 7). ITee or cross may be either reduced or the some size in the straight thru portion n = cross ft. in. wg section at (conversion h,, in. wg value for particular fitting. . TABLE 11-FRICTION OF ROUND ELBOWS 90’ S M O O T H R / D = 1.5 90’ 5-PIECE R/D = 1.5 A D D I T I O N A L EGUIVALENT 3 4 5 3 4 5 6 2.3 3 3.8 4.5 7 5.3 8 6 7 : 9 10 11 12 14 16 18 20 22 24 6 10 - - 90° 3-PIECE R/D = 6 8 10 12 14 lb 18 20 11 12 14 22 24 20 32 18 20 22 24 36 lb 1.5 I LENGTH OF STRAIGHT DUCT 40 44 48 .’ 45’ O - P I E C E 45’ S M O O T H R / D = 1.5 R / D = 1.5 (FT) . 1.5 2 2.5 3 1.1 1.5 1.9 2.3 3.5 4 4.5 5 2.6 5.5 6 7 0 9 10 11 12 3 ( I (:1l,\l”l‘lcI~ II. .\lII l)li(: I‘ 2-40 Ill-SIGN TABLE 12-FRICTION OF RECTANGULAR ELBOWS RADIUS ELBOW NO VANES DUCT W D RADIUS Radius Ratio+ R/D = 1.25 ELBOW-WITH VANES1 R t = 6” (Recommended) SQUARE ELEOWSf I Rt = 3” (Acceptable) Double Thickness Turning Vanes Single Thickness Turning Vanes ADDITIONAL EQUIVALENT LENGTH OF STRAIGHT DUCT (FT) 96 45 36 31 33 28 2 2 2 1 43 31 38 29 25 3 3 2 2 2 40 30 25 20 17 60 45 37 30 25 28 23 21 1715 13 12' 44 33 28 29 23 18 2 2 2 1 1 1 41 29 33 25 19 16 15 3 3 2 2 2 2 1 35 29 25 21 18 15 11 60 45 37 30 2 20 15 48 36 30 24 20 16 12 27 22 19 16 14 12 10 41 31 25 27 22 16 2 2 2 1 1 1 39 27 31 26 21 15. 14 3 3 2 2 2 2 1 33 27 23 20 17 13 10 60 45 37 30 25 20 15 96* 48 36 30 24 20 16 12 10 a 45 26 20 la 15 14 11 9 a a 35 35 26 23 24 19 15 3 2 2 2 1 1 1 34 22 28 21 17 14 13 11 9 3 3 2 2 2 2 1 1 1 29 23 21 ia 15 12 10 a 7 60 A5 37 30 25 20 15 12 10 42 36 30 24 20 16 12 10 23 20 17 15 13 11 9 a 28 24 21 21 la 14 2 2 2 1 1 1 26 21 26 19 16 13 13 10 8 3 3 2 2 2 2 1 1 I 24 22 20 16 14 12 9 a 6 53 45 37 30 25 20 15 12 10 36 72* 36 30 24 34 19 16 14 27 22 19 20 3 2 2 1 34 ii. 'ix 1 37 1 19 22 22 15 12 12 9 a 3 2 2 2 2 1 1 1 20 la 15 13 11 9 a 6 45 37 30 25 20 15 12 10 19 ia 19 1612 - 2 2 1 1 1 16 21 17 14 12 12 9 a 3 2 2 2 2 1 1 1 17 17 15 12 11 a 7 6 40 37 30 25 20 15 12 10 72 60 48 42 ~._ 48 36 30 24 31 25 22 19 48 36 30 .24. 20 16 12 a I 10 a 32 32 30 24 20 16 12 10 a I I * a 7 17 16 14 -l2 10 a, 7 6, 1 s 5 / TABLE 12-FRICTION OF RECTANGULAR ELBOWS (CONT.) DUCT DIMENSIONS (in.) RADIUS ELBOW NO VANES RADIUS ELBOW-WITH VANES1 SQUARE ELBOWS$ R’ W Radius Ratio+ R/D = 1.25 D R, = 3” (Acceptable) R+ = 6” (Recommended) Double Thickness Turning Vanes Single Thickness Turning Vanes ADDITIONAL EQUIVALENT LENGTH OF STRAIGHT DUCT (FT) Vfltle5 28 28 24 20 16 12 10 8 24 20 15 13 12 10 a 14 17 15 11 2 1 1 1 17 15 13 11 11 9 8 7 6 96* 72% 40* 24 20 16 12 10 8 6 38 32 22 19 17 20 lb 13 11 3 3 2 1 1 1 80* 60* 40* 20 16 12 10 8 6 32 26 16 19 15 12 9 3 2 2 1 1 64* 48* 32* 16 26 21 15 8 5 4 48* 36* 24* 12 10 19 16 11 7 6 5 4 16 v 6 20 14 12 10 10 8 7 3 2 2 1 9 12 11 8 Vall.3 $J 2 2 2 2 1 1 I 3 2 2 2 1 1 1 14 10 9 9 8 7 3 2 2 ;1 1 1 12 9 8 0 6 6 3 3 2 1 1 1 14 13 12 10 0 7 6 34 30 25 20 15 12 10 23 21 18 12 10 9 8 7 6 4 80 72 62 30 25 20 15 12 10 8 19 17 14 10 8 7 6 5 4 66 58 14 12 11 7 6 5 5 4 48 43 38 20 15 12 10 8 . 49 25 20 15 12 10 0 8 2 2 1 8 7 8 7 5 5 3 3 2 1 1 1 10 9 8 5 5 4 3 33 30 26 15 12 10 8 6 2 2 1 6 8 6 5 3 2 2 1 6 19 13 9 5 4 4 8 7 6 4 4 3 27 24 21 12 10 8 32* 24* 16' 8 6 13 11 8 4 3 5 6 4 2 1 1 21 19 16 10 6 24% 18* 12* 6 10 4 3 1 1 15 13 11 8 12 a 6 10 40* 30*. 20* 10 a 7 8 8 6 I 3 t F o ro t h e r r a d i u s r a t i o s , s e e T a b l e I O . - D e n o t e sH a r d B e n d s a s s h o w n Hard Bend 3 Easy Bend . J\' I6 I' . I F o ro t h e r s i z e s , s e e T a b l e1 0 . V a n e s m u s t b e l o c a t e d a s i l l u s t r a t e d i n C h a r t6, page 24, to have these minimum losses. I ; I CHART 9-LOSSES FOR ROUND FITTINGS Elbows, Tees and Crosses NOTES: 1. Loss for tee or cross is CI function of the velocity in the branch. This represents the loss in static pressure from the main upstream to the branch. Qlr is the ratio of air quantity of the branch to the main upstream. 2. Loss for 45’ smooth elbow is equal to one-half the loss for a 90’ smooth elbow. 3. Loss for 45’ 3-piece elbow is equal to one-half the loss for a 90’ 5.piece elbow. 2-46 ~ I’,IliT 2 . A I R DISTIIIBUTION lent round diameters may be obtained from Cllnrl i using air velocity and air quantity. Tab/e 6 is usctl with the equivnlcnt round diameter to select the rectangular duct sizes. The fan static pressure rcquiretl for the supply is detcrminctl by calc~llation, using the longest run of duct including all elbows and tittings. TnOle.5 /I) ~161 12 are used to obtain the losses thru the rectangular elbows and littings. The longest run is not necessarily the run with the greatest friction loss, as shorter runs may have more elbows, fittings and restrictions. This method is not normally used, as it requires a broad background of duct design experience and knowledge to be within reasonable accuracy. It should be used only for the most simple layouts. Splitter dampers should be included for balancing purposes. metrical layouts. II’ a design has a mixture of short and long rumi, the shortest run requires consicleral)le tlampering. Such a system is difficult to balance since the equal friction method makes no provision for cqnali/:ing pressure drops in branches or for providing the same static pressure behind each air terminal. The usual I)rocedure is to select an initial velocity in the main duct near the fan. This velocity should be selcctetl from Tcrble 7 with sound level being the limiting factor. Clrnrt 7 is ~csctl with this initial velocity and air quantity to determine the friction rate. This same friction loss is then maintained th~~ghout the system and the equivalent round duct tliametcr is selected from Clrt~~.t 7. Equal Friction Method To expedite equal friction calculations, Table 13 is often used instead of the friction chart; this results in the same duct sizes. Tliis method oE sizing is used for supply, exhaust and return air duct systems and employs the sa,me friction loss per foot of length for the entire system. The equal friction method is superior to velocity reduction since it requires less balancing for sym- The duct areas determrned from Table 13 or the equivalent round diameters from C/NW 7 are used to select the rectangular duct sizes from Table ci. This procedure of sizing duct automatically reduces the air velocity in the clirection of flow. TABLE CFM CAPACITY % I~-PERCENT DUCT AREA % SECTION PREA IN BRANCHES FOR,MAINTAINING / EQUAL CFM / CAPACITY % DUCT AREA % CFM j CAPACITY % DUCT AREA % CFM / CAPACAY % FRICTION 1 2 3 4 5 2.0 3.5 5.5 7.0 9.0 26 27 28 29 30 33.5 34.5 35.5 36.5 37.5 51 52 53 54 55 59.0 60.0 61.0 62.0 63.0 76 77 78 79 80 DUCT AREA % 81.0 82.0 83.0 84.0 84.5 6 7 8 9 10 10.5 11.5 13.0 14.5 16.5 31 32 33 34 35 39.0 40.0 41.0, 42.0 43.0 56 57 58 59 60 64.0 65.0 65.5 66.5 67.5 81 82 83 84 85 85.5 86.0 87.0 87.5 88.5 11 12 13 14 15 17.5 18.5 19.5 20.5 21.5 36 37 38 39 40 44.0 45.0 46.0 47.0 48.0 61 62 63 64 65 68.0 69.0 70.0 71.0 71.5 86 87 88 89 90 89.5 90.0 90.5 91.5 92.0 16 17 18 19 20 23.0 24.0 25.0 26.0 27.0 41 42 43 44 45 49.0 50.0 51.0 52.0 53.0 66 67 68 69 70 72.5 73.5 74.5 75.5 76.5 91 92 93 94 95 93.0 94.0 94.5 95.0 96.0 24 25 31.5 32.5 59 -50 57.0 58.0 74 75 80.0 80.5 99 100 99.0 100.0 ’ l CN/\I’TER 2 . AIR 247 VTSIGN 7’0 determine the total sYste*ll that the fan m u s t 0, ,oss in t,,e t,Lict to dwhte t h e l o s s i n the ‘G,,.t is necessary highest resistance. The friction 1, havi,,l: the and fittings in the section must be ,L,all ei,,ows Example 4 - Equal Friction Method of &“}, ‘\ Given: Duct systems for general office (rig. 47). Total air quan’ity - 5400 cfm 18 air terminals - 300 cfm each Operating pressure for all tcrlninals - 0.15 Radius ell)ows, R/D = 1.25 ‘?lCfS ‘_. DUCT SECTION 1 in, wg Solution: 1. From P) .4DD. EQUIV ENCTH ’ (f9 60 Duct A- B ” B - 1 3 - Duct Elbow Duct Duct T&/r 7 selecx an initial velocity of 1700 fpm. ;;>I Duct 17 _ 18\‘.Uct cfm 5400 Duct area = s = 3.18 sq ft - From \ \ Find: 1. Initial duct velocity, area, sire and friction r a t e in fhp duct section from the fan to the hrst branch. 2. Size of remaining duct runs. 3. Total equivalent length of duct run with highest resistance. otal sratic pressure required at fan discharge. I LENGTH ITEM Table 6,selec; a duct size - 22 in. x 22 in. Initial friction rate is determined from Chart 7 using the air quantity (5400), and the equivalent round duct diameter from Table 6. Equivalent round duct diameter = 24.1 in. i Friction rate = ,145 in. wg 2. The duct areas are sizes are determined lat.% thh design information: DUCT SECTION AIR QUANTITY CFM* CAPACITY (cfm) (%) 5400 3600 100 \ ToA A-B B- 13 13-14 14-15 15-16 1800 1580 1200 900 16 - 17 17 - 18 600 300 I DUCT SECTION ToA A-B B- 13 13 - 14 14 - 15 15 - 16 16. 17 17 - 18 (%) 69 4 3.18 loo.0 73.5 7’ 2.43 1.3 41.0 35.5 29.5 24.0 1.12 .94 .76 .56 .33 17.5 10.5 i Nair 6 DUCT SIZE: (in.) 22x22 22x 16 22x 10 18 x 14 x 12 x 8x 8x 10 10 10 10 10 quantity in duct section total air quantitv tDuct area = percent of area times initial duct area (fan to A) :Refer to page 21 for reducing duct sizes. y&!‘/.[f- The total frictioni \<he ductwork from the fan to last terminal I8 is sho LOSS = total equiv length.ihe f”‘lowing’ ‘&tion r a t e = 229ftx .I45 in. wg i 100ft =&Jo, i .33 in. wg Total static pressure required?&. ‘an discharge is the sum of the terminalLoperating_pr~~ eye and the loss in the ductwork. Credit can be taken fo r% velocity regain between the first and last sections of ducq 11 I AREA+ : a; t‘hA 4. 210 67 33 28 22 17 DUCT AREA *Percent of cfm = Total \ f+...’ J”, ( 5400 CFM $3 900 C2-ti 9 0 0 / 4 22 10 600 16 (26’ 6 0 0 I4 :20 ;jJ,,, 1 ,2dF I 12 26 9 0 0 :bC -, ..’ 26 600. ,, 300 j, -s. r, ‘, &- ,- _-1.. 2d 3& / ‘6, Frc.47 - DUCT LAYOUTFOR Low VEI.OCITY SYSTEM (EXAMPLES~ AND 4) Velocity in initial section = 1700 Cpm Velocity in last section = 590 fpm ITsin a 75’,‘{, regain coellicient, = .75 (.I8 - AX) = .I’)N in. wg Thcrcfore, the total static pressure at fan discharge: = duct friction + terminal pressure - regain = .33 + .I5 - .I2 = .36 in. wg The equal friction method does not satisfy the design criteria of uniform static pressure at all branches and air terminals. To obtain the proper air quantity at the beginning of each branch, it is necessary to include a splitter damper to regulate the flow to the branch. It may also be necessary to ave a control device (vanes, volume damper, or .djustable terminal volume control) to regulate the flow at each terminal for proper air distribution. In Example 4, if the fan selected has a discharge velocity of 2000 fpm, the net credit to the total static pressure required is determined as described under “Fan Conuersion Loss or Gain:” Gain = .75 [(gy -($j&r] = .75 (.25 - .lS) = .05 in. wg Static Regain Method The basic principle of the static regain method is to size a duct run so that the increase in static pressure (regain due to reduction in velocity) at each branch or air terminal just offsets the friction loss in the succeeding section of duct. The static presqure is then the same before each terminal and at lch branch. The following procedure is used to design a duct system by this method: Select a starting velocity at the fan discharge from Table 7 and size the initial duct section from Table 6. The remaining sections of duct are sized from Chart 10 (L/Q Ratio) and Chart II (Low Velocity Static Regain). Chart 10 is used to determine the L/Q ratio knowing the air quantity (Q) and length (L) between outlets or branches in the duct section to be sized by static regain. This length (L) is the equivalent length between the outlets or branches, including elbows, except transformations. The effect of the transformation section is accounted for in “Chart II - Static Regain.” This assumes that the transformation section is laid out according to the recommendation presented in this chapter. Chart ff is used to determine the velocity in the duct section that is being sized. The values oE the L/Qrntio(Chnrt IO) and the velocity (V,) in the duct section immediately before the one being sized are used in Chart Il. The velocity (VJ determined from Chart II is used with the air quantity to arrive at the duct area. This duct arcx is used in Table 6 to size the rectangular duct and to obtain the equivnlent round duct size. By using this duct size, the friction loss thru the length of cluct equals the increase in static pressure due to the velocity change after each branch take-off and outlet. However, there are instances when the reduction in area is too small to warrant a change in duct size after the outlet, or possibly when the duct area is reduced more than is called for. This gives a gain or loss for the particular duct section that the fan must handle. Normally, this loss or gain is small and, in most instances, can be neglected. Instead of designing a duct system for zero gain or loss, it is possible to design for a constant loss or gain thru all or part of the system. Designing for a constant loss increases operating cost and balancing time and may increase the fan motor size. Although not normally recommended, sizing for a constant . loss reduces the duct size. Example 3, - Static Regain Method of Designing Ducts Given: Duct layout (Example 4 and Fig. 47) Total air quantity - 5400 cfm Velocity in initial duct section - 1700 fpm (Example 4) Unvaned radius elbow, R/D = 1.25 18 air terminals - 300 cfm each Operating pressure for all terminals - 0.15 in. wg Find: 1. Duct sizes. 2. Total static pressure required at fan discharge. Solution: 1. Using an initial velocity of 1700 fpm and knowing the air quantity (5400 cfm), the initial duct area after the fan discharge equals 3.18 sq ft. From Table 6, a duct size of 22” x 22” is selected. The equivalent round duct size from Table 6 is 24.1 in. and the friction rate from Chart 7 is 0.145 in. wg per 100 ft of equivalent length. The equivalent length of duct from the fan discharge to the first branch: = duct length i- additional length due to fittings = 60 + 12 = 72 ft The friction loss in the duct section up to the branch: = equiv length of duct X friction rate 0.145 =72x100= first .104 in. wg The remaining duct sections are now sized. The longest duct run (A to outlet IS, Fig. 47) should be sized first. In this example, it is desirable to have the i . 249 CHAPTER 2. AIR DUCT DESIGN CHART IO--L/QRATIO .OS .04 .03 I .01 a6 -I 0 13 1 ’ 2 II’ 3 “““’ 4 567810 I! 20 1.8, I,, 30 40 5060 80 100 AIR O”ANT,TY AFTER TAKE-OFF, 0 t 100 CFMI From Form E-147A CHART II-LOW VELOCITY STATIC REGAIN VELOCITY AFTER TAKE-OFF,V2 ( FPM) From Form E-147A 2750 I’.ART 2 . A I R DLSTRII~UTION static prcsswe in the duct immediately before outlets f and 7 equal 10 the static pressure Iwfore outlet 13. is a savings in building space normally allotted to the air conditioning ducts. Figure -48 tal~ulatcs the duct sizes. Usually Class 11 fans are rcquirctl for the increased static pressure in a high velocity system and extra care must be taken in duct layout and construction. Ducts are normally sealed to prevent leakage of air which may cause objectionable noise. Round ducts are preferred to rectnngular because of greater rigidity. Spirn-Pipe should be used whenever possible, since it is made of lighter gage metal than corresponding round and rectangular ducts, and does not require bracing. 2. The total pressure required at the fan discharge is equal to the sum of the friction loss in the initial tlucl section plus the terminal operating pressure. Fan discharge pressure: = friction loss + terminal pressure = ,104 + .I5 = 2.5 in. wg It is good design practice to include splitter dampers to regulate the flow to the branches, even though the static pressure at each terminal is nearly equal. Comparison of Static Regain and Equal Friction Methods Examples 4 and 5 show that the header duct sizes determined by the equal friction or static regain method are the same. However, the branch, ducts, ized by static regain, are larger than the branch ducts sized by equal friction. Figure 49 shows a comparison of duct sizes and weights established by the two methods. The weight of sheet metal required for the system designed by static regain is approximately 13% more than the system designed by equal friction. However, this increase in first cost is offset by reduced balancing time and operating cost. If it is assumed that a low velocity air handling system is used in Examples 3 and 4 and that a design air flow of 5400 cfm requires a static pressure of 1.5 in. wg, the increased horsepower required for other equal friction design is determined in the following manner. Symmetry is a very important consideration when designing a duct system. Maintaining as symmetrical a system as possible recluccs balancing time, design time and layout. Using the maximum amount of symmetrical duct runs also reduces construction and installation costs. Particular care must be given to the selection and location of fittings to avoid excessive pressure drops and possible noise problems. Figure 50 illustrates the minimum distance of six duct diameters between elbows and 90” tees. If a 90” conical tee is used, the next fitting in the direction of air flow may be located a minimum of one-half duct diameter away (Fig. 51). The use of a conical tee is l&ted to header ductwork and then only for increased initial velocitiei in the riser. When laying out the header ductwork for a high velocity system, there are certain factori that must be considered: 1. The design friction losses from the fan discharge to a point immediately upstream of the first riser take-off from each branch header should be as nearly equal as possible. These points of the same friction loss are shown in Fig. 52. Additional hp = 1*861y51*75 = 6.3’% approx. A 6% increase in horsepower often indicates a larger fan motor and subsequent increased electrical transmission costs. HIGH VELOCITY DUCT SYSTEMS A high velocity air distribution system uses higher air velocities and static pressures than a conventional system. The design of a high velocity system involves a compromise between reduced duct sizes and higher fan horsepower. The reduced duct size 2. To satisfy the above principle when applied to multiple headers leaving the fan, and to take maximum advantage of allowable high velocity, adhere to the following basic rule wherever possible: Make as nearly equal as possible the ratio of the total equivalent length of each header run (fan discharge to the first riser take-off) to the initial header diameter (L/D ratio). Thus the longest Spiru-Pipe header run should preferably have the highest air quantity s o that the highest velocities can be used throughout. 3. Unless space conditions dictate otherwise, the take-off from the header should be made using a 90” tee or 90” conical tee rather ‘than a 45” Cr-l;\I”l‘EK 2-51 2. AIK I>U(:‘I’ D E S I G N rota1 S.P. Loss for SUI3~1~ . duct system = S.P. for critical duct _ in. wg plus air outlet S.1’. loss - in. wg = _ in. 3 4 I<()UlV. I Q 1 5 L I 1 VELOCITY 1 Indicated I 6 7 ARE,4 DIJCT DI.\M. OR RECT. srzlq 1 (in.) Fan to A .A - B B -13 13 - 14 14 - 15 5400 3600 1800 1500 1200 22 x 22x 22x 22x 22x 15 - 16 16 - 17 17 - 18 B- 7 7- 8 900 600 300 1800 1500 20x 10 16 x 10 10 x 10 22x 10 22x 10 5- 9 -10 iO- 11 ll- 12 A- 1 1200 900 600 300 1800 22x 10 20 x 10 16 x 10 10 x 10 22 x 10 1500 1200 900 600 300 22x 10 22x10 20x 10 16 x 10 10 x 10 l2345. 2 3 4 5 6 D U C T SECTIOF 48 - DUCT S IZING C ALCULATION FORA< Duct Dimensions (in.) Duct Weight (‘b) 592 179 394 411 2-3,8-g, 14-15 3-4, 9-10, 15-16 4-5, 10-11, 16.17 5-6, 11-12, 17-18 14 x 12 x 8x 8x 10 10 10 10 360 321 270 270 Total weight of duct* Allow 15y0 for scrap Total wt of sheet metal 22 x 22x 22x 22x ’ 2797 420 3217 transformation and I Duct Dimensions. (in.) 22 16 10 10 includes 0.104 STATIC REGAIN METHOD 22 x 22x 22x 18x weight 0.104 EQUAL FRICTION METHOD ToA A to B A-l, B-7, B-13 l-2, 7-8, 13-14 *Total “‘. 2 16 10 10 10 From Form E-147 not symmetrical and handle different air quantities, an initial velocity is assumed at the beginning of the branch. This velocity is somewhat less than the velocity in the header before take-off. *Duct size is assumed to determine loss thru elbow. tDuct sizes from Table 6. Longest duct run is sized first. Remaining duct sections are the same size, as they are , symmetrical to branch B tkm 18. If other branches are F IG. 9 1 F R I C T I O N 1TOT‘\L LOSS OR S.P. TAKE-OFF LOSS TO IN TAKE-OFF. DUCT S.P. CHANGE (in. wg) (in. wg) 1 Selected 1 Indicated j Selected (cfm) 8 Wg. elbows. F IG. 49 - COMPARISON OF DUCT S I Z I N G M E T H O D S Duct Weight (lb) 22 16 10 10 592 179 394 438 22x 1 0 20 x 10 16 x 10 1 0 x 10 438 435 384 297 3157 475 3632 2-52 I’,\RT tee. 1Sy using 90” fittings, the pressure drop to the branch throughout the system is more uniform. In addition, two fittings are normally required when a 45” tee is used and only one when a 90” fitting is used, resulting in lower first cost. The design of a high velocity system is basically the same as a low velocity duct system ciesigneci for static regain. The air velocity is reduced at each take-off to the riser and air terminals. This reduction in velocity results in a recovery of static pressure (velocity regain) which offsets the friction loss in the succeeding duct section. The initial starting velocity in the supply header depends on the number of hours of operation. TO achieve an economic balance between first cost and 90’TEES FIG . 50 - S PACING OF FITTINGS IN DUCT RUN operating cost, lower air velocities in the header are recommended for 24-hour operation where space permits. When a 90” conical tee is used instead of a 90” tee for the header to branch take-off, a higher initial starting velocity in the branch is recommended. The following table suggests initial veiocities for header and branch duct sizing: RECOMMENDED INITIAL VELOCITIES USED WITH CHARTS 12 AND 13 (fpm) HEADER 12 hr. operation 24 hr. operation 3000 - 4000 2000 - 3500 BRANCH* 90” conical tee 90” tee - 4000 - 5000 3500 - 4000 TAKE-OFFS TO TERMINALS 5 1 - SPAC:ING OF FI-JXNGS 2000 maximum *Branches are defined as a branch header or riser having 4 to 5 or more take-offs to terminals. Static regain charts are presented for the design of high velocity systems. Chart 12 is used for designing branches and Chart 13 is used for header design. The basic difference in the two charts is the air quantity for the duct sections. Chart 12 is used for sizing risers and branch headers handling 6000 cfm or less. The chart is based on 12 ft increments between take-offs to the air terminals in the branches or take-offs to the risers in branch headers. A scale is provided to correct for 1 spacings more or less than 12 ft. Chart 13 is used to size headers, and has an air quantity range of 1000 to 40,000 cfm. The chart is based on 20 ft increments between branches. A correction scale at the top of the chart is used when take-off to branch is more or less than 20 ft. Examples 6 and 7 are presented to illustrate the use of these two charts. Example 6 is a branch sizing problem for the duct layout in Fig. 53, and Example 7 is a header layout (Fig. 55). 90’ CONICAL TEE -, FIG . 2 . AIR DIS?‘RIBUTION WHEN USING 90" h+& CONICAL TEE . 2-53 CHAPTER 2. AIR DUCT DESIGN CHART 12-BRANCH HIGH VELOCITY STATIC REGAIN R O U N D DUCT EOUIVALENT LENGTH FOR L * 12 FT . .REFERENCE LINE L = 12 FT BRANCH TAKE-OFF LOSSI IN WC DUCT DIAMETER (IN , Form E-l 48A 2-54 PAKT 2. AIK DISTRIBUTION CHART 13-HEADER HIGH VELOCITY STATIC REGAIN - - R E F E R E N C E L I N E L = 2OFT 60 55 50 3.0 45 40 36 32 26 26 24 22 20 16 ; 16 14 12 I’ II IO ’ ” ,,,,I,! 0 55 50 45 40 36 32 9 ’ 8 ’ I, 7 I ” i/I i, / /I I II 26 26 24 2 OUCT DIAMETER (IN.) Form E-149A CHAPTER 2. AIR DUCT 2 - s DESIGN CONICAL 90° TEE SMOOTH t SECT- I 1200 CFM A.T. - 1.5 IN. ug SMOOTH 90’ ELL S E C T - 2 IOOOCFM * POINT IN BRANCH HEADER BEFORE A.T. . 1.5 IN. wg 12’ SECT-3 800 CFM A.T. = 1.5 IN. vrg 12’ SECT-4 600 CFM 12’ SECT-5 A.T. 8 1.5 IN. w(l MAIN HEADER 400 CFM A.T. = 1.5 IN. wg 12’ SECT-6 200 CFM A.T. = 1.5 IN. wg k IG. 52 - HIGH V ELOCITY HEADERS Example 6 - Use of Branch Duct Sizing . FIG. BRANCHES AND FOR E XAMPLE 6 2. Enter CAnrt 12 at the velocity range recommended for branch risers with a 90” conical tee, Page 52. Chart Given: Office building riser as shown in Fig. 23 12 air terminals - 100 cfm each . Total air quantity - 1200 cfm .\ir terminal static pressure - 1.5 in. wg Find: Duct sizes for Sections 1 thru 6, Fig. J3. Solution: 1. Sketch branch as shown in Fig. 23. values in columns 2, 3 and 8, Fig. 51. 53 - BRANCH DUCT 3. Intersect the initial branch air quantity, 1200 cfm, as shown at point A. Read 7 in. duct size and 3.8 in. wg loss per 100 ft of equivalent pipe and 4500 fpm velocity. Enter these values on the high velocity calculations form (Fig. 54). 4. From point A, determine header take-off loss by projecting horizontally to the left of point A and read 1.25 in. wg. 5. Enter 1.25 in. wg in Fig. 5f for section 1. Enter appropriate 6. Determine equivalent length from the header to the first air terminal take-off: INITIAL CONDITIONS: Cfm 1200; Duct Size 7 in.; Velocity 4500 fpm. 4 1 I 5 PRESSURE READING (in. wg) Initial Selected I 6 7 8 TAKE-OFF TO TAKE-OFF S.P. CHANGE (4 minus 5) (in. wg) S.P. AHEAD OF TAKE-OFF AIR TERMINAL PRESS. (in. wg) (in. wg) Branch T a k e - O f f F. L. = 1.25 Duct Friction Loss = .89 1.0 1.25 - 0.25 0.84 0.84 0.0 0.57 0.47 + 0.1 0.32 0.40 - 0.085 0.26 0.24 + 0.02 Maximum S. P . i s at Section 2 : f (in.) @Pm) 2.14 I .5 7 4500 2.39 2.39 2.29 2.37 2.35 1.5 1.5 1.5 1.5 I .5 7 7 7 6 5 3700 3050 2300 2050 1475 2.39 + F IG. 54 - HIGH V ELOCITY BRANCH S IZING CALCULATIONS 1.5 I + .19 = 4.08 From Form E-148 PAKT 2. AIR DISTRIBUTION 256 Length of pipe = 6 + 12 = 18 ft. One 7 in. smooth ell = 5.3 ft. Total equivalent length = 18 + 5.3 = 23.3 ft. Pressure drop = 23.3 X 3.8/ 100 = .89 in. wg. 7. Determine duct size for section 2: From point ,4 on Ctzart 12, project thru points B and C to the 1000 cfm line at point D. 8. Determine equivalent length for section 2: Actual duct length = 12 + 2 = 14 ft. Two smooth 90 ells = 2 X 5.3 = 10.6 ft. Total equivalent length = 14 f 10.6 = 24.6 ft. 9. Determine pressure loss in section 2: Project vertically from point D to reference line, then to point E. Proceed on the guide lines to 24.6 ft equivalent length, point F. Project vertically from I; to 1000 cfm line at point C, then along the 1000 cfm line to point H. Enter point H (1.25 in. wg) and point G (1.0 in. wg) in Fig. 54, columns 4 and 5. The net loss is “point H point C” = 1.25 - 1.00 = .25 in. wg. This is entered in column G of Fig. 54. Enter 7 in. diameter in column 9. Example 7 - Use of Header Sizing Chart Given: Office building, 12.hour operation Header as shown in Fig. 55 10 branches - 1200 cfm each Total air quantity - 12,000 cfm Find: Header size for sections 1 thru 10 Solution: I. Sketch header as shown in Fig. 55. values in Fig. 56, columns 1, 2, 3 and 8. Enter appropriate 2. Enter Chart 13 at the velocity range recommended headers in a system operating 12 hours, page 5~‘. 3. Intersect the initial header air quantity 12,000 cfm as shown, at point A. Read 24 in. duct size and .62 in. wg loss per 100 ft of equivalent pipe and 3800 fpm. Enter these values on high velocity calculation form (Fig. 56). 10. Determine duct size for section 3: Project downward on the 7 in. diameter line to the 1000 cfm line, points H to I. 4. Calculate the equivalent length of section 1 and record in column 3; straight duct = 20 feet, no fittings; pressure* drop = 20 X .62 = ,124 in. wg. 11. Project along guide lines at the right side of the chart from I to the 800 cfm line at point J. Duct size is 7 in. Enter appropriate values from the chart in columns 4, 5, 6 and 9 of Fig. 54. 5. Size duct section 2: From point A on chart, project thru points B and C to the 10,800 cfm line at point D. 12. Determine duct size for section 4: Project downward on the 7 in. diameter line to the 800 cfm line, points J to K. 13. Project along guide lines at the right side of the chart from point I< to the 600 cfm line at point L. Project along the 600 cfm line to the 7 in. diameter line, point L to M. This results in a static regain of .57 - .47 = .I0 in. wg. Duct size for section 4 is 7 in. Enter appropriate values in Fig. 54, columns 4,5, 6, 7 and 9. NOTE: If the 600 cfm line is projected from point L to the 6 in. diameter line, a net loss of .88 - .45 = .43 in. wg results. This friction loss unnecessarily penalizes the system. Therefore, the projection from L is made to the 7 in. diameter line. 14. Determine duct size for section 5: Project downward from M to 600 cfm line, point N. Project along guide lines to 400 cfm line, point 0. Continue along the 400 cfm line to the 6 in. diameter line, point 0 to P. This results in a static pressure loss of .40 - .315 = .085 in. wg. Duct size is 6 in. Enter the appropriate values in Fig. 54, columns 4, 5, 6, 7 and 9. for 6. Determine equivalent length for section 2: Actual length = 20 ft. One j-piece 90” ell = 24 ft. Total equivalent length = 20 + 24 = 44 ft. 7. Determine pressure loss in section 2: Project vertically , from point D to reference line, point E. Proceed on the . guide lines to 44 ft equivalent length, point F. Project vertically from F to 10,800 cfm line at point G, then along the 10,800 cfm line to point H . Enter the net loss., read from point G (.84) and from point H (.90) in columns 4 and 5, Fig. 56. The net loss is “point H - point G” = .90 - .84 = .06 in. wg. This is entered in column 6, Fig. 56. Enter 24 in. diameter in column 9. 8. Determine duct size for section 3: Project downward on the 24 in. line to the 10,800 cfm line, point H to I. Project along the guide lines at the right side of the chart from I to the 9600 cfm line at point J. Enter appropriate values from the chart in columns 4, 5, 6 and . 9, Fig. 56. NOTE: If the 400 cfm is projected from point 0 to the 7 in. diameter line, a net regain of .315 20 = ,115 in. wg results. Therefore, the 6 in. size is used to save on first cost since the net loss using the 6 in. size is insignificant. .- 15. Determine duct size for section 6: Duct size is 5 in. as determined-from point S. 16. Determine velocities for duct sections l-6 from points A, I, K, N. Q and T respectively;, enter in column 10. 17. Determine take-off and runout pressure drop by entering upper right hand portion of Chart 12 at 100 cfm and read a pressure drop of .19 in. wg for a 4 in. runout size. 18. Add 2.39 in. wg (maximum from column 7) plus.1.5 in. wg (column 8) plus .I9 (take-off and runout drop) to find 4.08 in. wg (total branch S.P.). FIG. 55 - HIGH VELOCITY D UCT SYSTEM - HEADER STATIC REGAIN M ETHOD SIZING j i CIl.2I’TER 2. AIK DUCT 2-257 DESIGN INITIAL CONDITIONS: Cfm 12,000; Duct Size 24 in.; Velocity 3800 fpm. 4 1 2 3 5 HEADER SECT. NO. AIR QUANTITY Q EQUIV. DUCT LENGTH L (cfm) w 12000 10800 9600 8400 7200 20 44 20 20 20 0.84 0.74 0.57 0.42 6000 4800 3600 2400 1200 44 20 20 20 20 0.31 0.22 0.195 0.165 0.165 PRESSURE READING (in. wg) Initial Selected 8 9 10 S.P. AHEAD OF TAKE-OFF BRANCH S.P. DUCT SIZE VELOCITY V (in. wg) (in. wg) (in.) @pm) 6 TAKE-OFF TO TAKE-OFF S.P. CHANGE (4 minus 5) (in. wg) Duct Friction = 0.124 0.90 -0.06 0.70 $0.04 0.55 $0.02 0.42 0.0 0.30 0.26 0.23 0.24 0.2 1 10.01 -0.04 -0.035 -0.075 -0.045 0.124 0.184 0.144 0.124 0.124 4.08 4.08 4.08 4.08 4.08 24 24 24 24 24 3800 3400 3000 2600 2250 0.114 0.154 0.189 0.264 0.309 4.08 4.08 4.08 4.08 4.08 24 22 20 16 12 1900 1800 1650 1650 1500 Maximum S.P. at Section 10 = 0.31 f 4.08 = 4.39 From Form E-149 FIG. 56 - HIGH V ELOCITY HEADER S IZING CALCULATIONS 9. Determine duct sizes for sections 4 thru 10 in a manner similar to Step 8, using the listed air quantities and equivalent lengths. One exception is duct section 6. Since its equivalent length is 44 feet, use the method outlined in Steps 5, 6 and 7 to determine the pressure drop. In addition, see Exampk 5, Steps 13 and 14, for explanation when the chart indicates a duct diameter other than those listed, for instance 23 inches. &CT HEAT GAIN AND AIR LEAKAGE Whenever the air inside the duct system is at a temperature different than the air surrounding the duct, heat flows in or out of the duct. As the load is calculated, an allowance is made for this heat gain or loss. In addition, air leakage is also included in the calculated load. The load allowance required and guides to conditions under which an allowance should be made for both heat gain or loss and duct leakage are included in Part I, System Heat Gain. Chart 14 is used to determine the temperature rise or drop for bare duct that has an aspect ratio of 2: 1. In addition, correction factors for other aspect ratios and insulated duct are given in the notes to the chart. Example 8 - Calculations for Supply Duct Given: Supply air quantity from load estimate form - 1650 cfm Supply duct heat gain from load estimate form - 57, Supply duct leakage from load estimate form - 5% Unconditioned space temp - 95 F Room air temperature - 78 F Duct insulation U value - .24 Duct shown in Fig. 57. Find: Air quantities at each outlet Solution: 1. Room air quantity required at 60 F 1650 = 1 + .05 + .05 = 1500 cfm FIG. 57 - DUCT HEAT GAIN AND AIR L EAKAGE 2-58 PART 2. AIR DISTRIBUTION Normally a 10 c;/o leakage allowance is used if the complete duct is outside the room. Since a large portion of the duct is within the room. 5(;;IS used in this example. 2. Determine the temperature rise from A to f1: Select an initial starting velocity from Tab/e 7 (assume 1400 fpm). Calculate the temperature rise from the fan to the room. Enter Cltczrt If at 1500 cfm; project vertically to 1400 fpm and read .25 degrees temperature c h a n g e per 100 ft per degree F tlillerence. Using aspect ratio of 2: I, temperature rise Outler =522-(1540X&) =492&n 3. Determine cfm for outlet C: Use equal friction method to determine velocity in second section of duct, with 1540 - 492 = 10-18 cfm; velocity = 1280 fpm. Dctcrmine temperature rise at outlet: From C/tart 14, read 32 for 1280 fpm and 1040 cfm. Temperature rise = 32 X 17.2 X i’& = .83 F 30 ft = 100 X .27 F change X ,185 X (95 - 60) = 52 F Supply air temperature tliff = 17.2 - .8 = 16.4 F Outlet cfm adjusted for temperature rise :\ir temperature entering room = GO.52 F Actual air quantity entering room = II cfm with allowance for duct cooling 18 = 500 X 1~.4 = 550 cfm Allowance for duct cooling 78 - GO 7s - GO.52 X 1500 = 1540 cfm =550-(1048X&)=498cfm Air temperature rise from A to B 4. Determine cfm for outlet D: Use equal friction method to determine velocity in third ’ section of duct with 1048 - 498 = 550 cfm; velocity = 1180 fpm. Determine temperature rise at outlet: From CI~art IJ, read .43 F for 1180 fpm and 550 cfm. Temperature rise = & X 17.48 X .2i = .33 F Supply air temperature cliff to outlet B = i8 - (60.32 + .33) = 17.15 F Required air quantity to outlet B 18 = 500 x - = 522 cfm 17.2 = .43 X 16.4 X -& = 1.06 F with no allowance for cooling from the duct. Supply air temperature diff = 16.4 - 1.1 = 15.3 F . CHART 14-DUCT HEAT GAIN OR LOSS 300 4 0 0 500 600 6 0 0 1000 CFM AIR 2000 NOTES: 3000 4000 6000 6ooo 10000 Aspect Ratio Correction 1. with Baseda on duct 2:l bare rectangular ratio. aspect Aspect Ratio 1 Round 1 1:l ]3:1 \ 4:l Correction 1 1.92 j 1.18) 2. If duct is furred-in or insulated, use the following correction factors: .83 1 1.1 1 5:l 1 6:l / 7:l 1 8:l I 911 / 10:1 1.26 / 1.35 1 1.43 j 1.5 / 1.58 1 1.65 Furred-in duct - .45 I n s u l a t e d (U = .27) - ,185 Insulated (U = .13) - .lO 3. For air quantities greater than 10,000 cf m, divide air quantity by 100 and multiply degree change by 0. 1 CHAI’TEK 2. AIR DUCT 259 DESIGN Outlet cfm adjusted for tcmperaturc HIGH ALTITUDE DUCT DESIGN rise 18 = 500 x __ = 5X8 cfm 15.3 Allowance for duct cooling =i88-(i8ax &) =54Gcfm. 5. Check for total cfm: 492 + 498 + 546 = 1536 cfm This compares favorably with the 1540 cfm entering room. Fig. 57 shows original and corrected outlet air quantities. CHART 15-AIR 0 - 50 2000 0 4000 50 When an air distribution system is designed to operate above 2000 feet altitude, below 30 F, or above 120 F temperature, the friction factor obtained from Clrart 7, page 33 using the actual air quantity at final conditions must be corrected for air density. Chart 15 presents correction factors for temperature and altitude. The factors are multiplied together when a system is at high altitude and also operates outside the temperature range. DENSITY CORRECTION FACTORS 6000 6000 ALTITUDE (FT) 100 I50 AIR TEMPERATURE (Fl 10000 12000 200 250 14000 300 I’AKT 2. AIK DISTRIBUTION 2-60 DUCT Tables 15 md Jh which apply for low and high pressure systems. Fig. 58 illustrates the more common seams and joints used in low pressure systems. CONSTRUCTION The sheet metal gage used in the ducts and the reinforcing required depends on the pressure conditions of the system. There is also a wide variety of joints and seams used to form the ducts which also depend on pressure conditions in the duct system. TABLE 16-MATERIAL GAGE FOR SPIRA-PIPE DUCT Low and High Pressure Systems Low Pressure Systems Table 14 lists the recommended construction for rectangular ducts made of aluminum or steel. The method of bracing and reinforcing and types of joints and seams are included in the table. Round duct and Spira-Pipe construction are included in TABLE 14-RECOMMENDED CONSTRUCTION FOR RECTANGULAR SHEET METAL DUCTS . Low Pressure Systems MATER DUCT DIMENSION (in.) L GAGE I Aluminum B 8 S Gage Steel U.S. Gage RECOMMENDED CONSTRUCTION* Transverse Joints, Bracing and Reinforcing I Duct 24 Slip 24 Duct 1 Slip Up to 24 22 I 20 24 t o 3 0 24 24 22 20 3 1 to 60 22 22 20 18 6 1 to 72 20 20 18 16 Reinforced pocket slip? or reinforced Bar-St, spaced not more than four feet apart. 1 H” x 1 l/i” x x” dio’gonal angle reinforcing$ or 1 I/211 x 1 ‘/a” x gf’ girth angle reinforcing1 located midway between joints. 16 Reinforced pocket slipt or reinforced Bar-S slip+ spaced not more than four feet apart. 1 l/i” x 1 l/5” x %” diagonal angle reinforcing? or 1 ‘/a” x 1 %fl x l/e girth angle reinforcingt located midway between joints. 1 Ya” x ‘/.” band iron stay bracing for duct width 73” to 90”. 16 Reinforced pocket slipt or reinforced Bar-S slipt spaced not more than four feet apart. 1 y2” x 1 lh” x H” diagonal angle reinforcingf or 1 %” x I’%” x %” girth angle reinforcing$ lo&d midway between joints. 1 l/q” x 1/” band iron stay bracing for duct width 91 M to 120”. 11%” x I/” b a n d i r o n s t a y b r a c i n g s p a c e d 4 8 ” a p a r t f o r d u c t w i d t h s 121” and uo. 73 to 90 20 91 and Up 20 18 20 18 16 I i Pocket slio or Bar-S slio. soaced not more than eight feet aoart. Pocket slip or Bar-S slip, spaced not more than four feet apart. I I *All ducts over 18” in either dimension ore cross-broken, except those to which rigid board insulation is applied or area of duct where outlet or duct connection is to be installed. Duct seams are either Pittsburg lock seam or longitudinal seam. IReinforce joint with 1 l/q” x I/l” band iron. IAngles ore attached to duct by tack welding, sheet metal screws, or rivets on 6” centers. TABLE 15-RECOMMENDED CONSTRUCTION FOR ROUND SHEET METAL DUCT Low and High Pressure Systems MATERIAL DUCT DIMENSION Steel (in.) Up to 8 9 to 24 I RECOMMENDED GAGE U.S. Gage B 8 S Gage 24 22 22 20 25 to 36 20 CONSTRUCTION Aluminum 18 3 7 to 48 20 18 4 9 to 72 18 16 7 3 and Up 16 14 Reinforcing I . Joints and Seams I 1 l/q” x 1 l/q” x ‘/(” girth angle reinforcing spaced o n 8’ centers. 1 fi” x 1 l/q” x %” girth angle reinforcina soaced on 6’ centers.’ 1 IA” x 1 ‘A” x ‘A” girth angle reinforcing spaced on 4’ centers. Round duct sections ore ioined together by welding, by a coupling, or by belling out one end of duct. T h e seams o n r o u n d d u c t mov b e continuous welded or grooved longitudinal seam. ’ A- DRIVE SLIP o _ REINFORCED EAR-S SLIP FIG. 58 -JOINTS Pressure F-POCKET JOtNT SECTION AT CLIP PUNCH E -SLIDING SEAM G-STANDING High C- INSIDE GROOVE SEAM e- s S L I P SEAM AND S EAMS FOR Ii -PITTSBURGH SEAM Low P RESSURE SYSTEM Systems Table I7 contains the construction recommendations for rectangular duct made of aluminum or steel. The table includes the required reinforcing and bracing and types of joints and seams used in high pressure duct systems. TACK WELDEO OR RIVETED Fig. 59 shows the common joint used for rectangular ducts in high pressure systems. The ducts are constructed with a Pittsburg lock or grooved longitudinal seams (Fig. 58). Table 15 shows the recommended duct construction for round ducts. The data applies for either high or low pressure systems. Fig. 60 illustrates the seams and joints used in round duct systems. The duct materials for Spira-Pipe are given in TaOle 16. 1 - FIG. i BOLTED 59 -JOINT FOR H IGH P RESSURE S YSTEM PART 2. AIR DISTRIBUTION 2-62 TABLE 17-RECOMMENDED CONSTRUCTION FOR RECTANGULAR SHEET METAL DUCTS High Pressure Systems DUCT MATERIAL GAGE Steel U.S. Gage Aluminum B B 5 Gage Up to 24 22 20 25to 48 20 18 49to 6 0 18 16 DIM (in.) 61 and Up 18 16 SHEET METAL SCREWS RECOMMENDED CONSTRUCTION* Tranrvarse J o i n t s Bracing and Reinforcing n F l a n g e d a n g l e garketed loint or butt welded joint with girth angle, spaced not more than twelve feet apart. Angles are 1 I%” x 1 1/l” x %“t. 1 j/z” x 1 l/z” x l/l” girth angle reinforcing spaced 38” to 40” apartt. F l a n g e d a n g l e gasketed joint or butt welded joint with girth angle, spaced not more than twelve feet apart. Angler are 1 l/2* x 1 ‘/a” x ?&“i. 1 H” x 1 l/i” x 9%” g i r t h angle reinforcing spaced 38” x 40” aportt. FIG . G1 -JOINTAND~AM FORSPIRA-PIPE . Fittings are normally used to ,join sections of Spim-Pipe as shown in Fig. 61. Seahng compound is u s e d to join Spira-Pipe to fittings. *All ducts over 18” in either dimension are cross-broken except those to which rigid board insulation is applied or orea where outlets are installed. Seams are either Pittsburg lock sec~m or longitudinal sec~m. tAngle ? “SPIRA-PIPE”SEAM, are attached to duct by tack welding or rivets on 6” centers. WEIGHTS OF DUCT MATERIALS Table 18 gives the weights of various materials used for duct systems. COUPLING SLEEVE JOINT CONTINUOUS WELDED SEAM CONTINUOUS BUTT WELDED JOINT GROOVED LONGITUDINAL SEAM BELL JOINT Frc.60 -ROUND DUCT JOINTSAND~EAMS CHAPTER 2. AIR DUCT DESIGN 2-63 TABLE II-WEIGHTS OF DUCT MATERIAL WEIGHT GAGE (THICKNESS) (in.) (lb/v ft) WEIGHT PER SHEET (lb) 36 x 96 48 x 96 48 x 120 GALVANIZED STEEL, U.S. GAGE .906 1.156 1.406 1.656 2.156 2.656 3.281 ~ HOT ROLLED STEEL, U.S. GAGE ,750 1.000 1.250 1.500 2.000 2.500 3.125 5.625 26 24 22 20 18 16 14 10 ga. ga. 90. 9". 9-z. ga. ga. ga. 1.0179) I.02391 t.0299) C.0359) t.0478) f.0596) (.0747) (.1345) 18.0 24.0 30.0 36.0 48.0 60.0 78.0 135.0 24.0 32.0 40.0 48.0 64.0 80.0 104.0 180.0 ALUMINUM, El 8 5 GAGE ( 3 5 ) .355 .456 .575 .724 .914 1.03 24 22 20 18 16 14 12 ga. go. ga. gel. 9". ga. go. (.020) t.025) f.032) f.040) f.051) t.064) (.071) 6.9 8.6 11.0 13.8 17.4 22.0 24.7 9.2 11.3 14.6 18.4 23.2 29.2 33.0 11.5 14.2 10.2 23.0 29.0 36.6 41.3 STAINLESSSTEEL,U.S.GAGE (302) . .66 .79 1.05 1.31 1.58 2.10 2.63 3.28 15.8 la.9 25.2 31.5 37.8 50.4 63.0 70.7 21.1 25.2 33.6 42.0 50.4 61.2 84.0 104.9 26.4 31.6 42.0 52.5 63.0 84.0 105.0 131.2 32.0 40.0 48.0 64.0 72.0 80.0 40.0 50.0 64.0 80.0 90.0 100.0 COPPER, OZjSO FT 1.00 1.25 1.50 2.00 2.25 2.50 16 20 24 32 36 40 OZ. oz. oz. OZ. or.. oz. f.0216) t.027) f.0323) (.0432) t.0486) (.0540) 24.0 30.0 36.0 48.0 54.0 60.0 2-65 CHAPTER 3. ROOM AIR This chapter discusses the distribution of conditioned air alter it has been transmitted to the room. The discussion includes proper room air distribution, principles of air distribution, and types and lkation of outlets. TABLE 19-OCCUPIED ZONE ROOM AIR VELOCITIES ROOM AIR VELOCITY (fpm) O-16 REQUIREMENTS NECESSARY FOR GOOD AIR DISTRIBUTION 25 TEMPERATURE Recommended standards for room design conditions are listed in Pal-t I, Chupter 2. The air distr:‘>llting system m u s t b e d e s i g n e d t o h o l d t h e tc ,crature within tolerable limits of the above recokmendations. In a single space a variation of 2 F at different locations in the occupied zone is about the maximum that is tolerated without complaints. For a group of rooms located within a space, a maximum of 3 F between rooms is not unusual. Temperature variations are generally more objectionable in the heating season than in the cooling season. Temperature fluctuations are niore noticeable than temperature variaions. These fluctuations are usually a function of the temperature control system. When they are accompanied by air movements on the high end of the recommended velocities, they may result in complaints of drafts. DISTRIBUTION 1 RECOMMENDED APPLICATION REACTION Complaints air Ideal about stagnant design-favorable all commercial applications all commercial applications 25-50 Probably favorable but 50 fpm is approaching maximum tolerable velocity for seated peWXS 65 Unfavorable-light papers are blown off a desk 75 Upper limit for people moving aboutslowly-favorable retail and dept. store 75-300 Some factory air conditioning installations-favorable factory air conditioning higher velocities for spot cooling FAIR t AIR VELOCITY Tnble I9 shows room air velocities. It also il- Ius~~tes occupant reaction to various room air ve Lies in the occupied zone. AIR DIRECTION Table 19 shows that air motion is desirable and FAIR actually necessary. Fig. 62 is a guide to the most desirable air direction tor a seated person. I PRINCIPLES OF AIR DISTRIBUTION air The following section describes the principles of distribution. BLOW Blow is the horizontal distance that an air stream travels on leaving an outlet. This distance is measured from the outlet to a point at which the velocity of the air stream has reached a definite minimum value. This velocity is 50 fpm and is measured at 6.5 ft. above the floor. F1c.62 -DESIRABLE AIR DIRECTION I 2-66 I’.\K.I‘ 2 . .\ I K I~ISTKIBUTION I%low is 1~roportional to the velocity of the primary air as it leaves the outlet, and is independent of the temperature difference bctwecn the supply air and the room air. DROP Drop, or rise, is the vertical distance the air moves bctwccn the time it leaves the outlet and the time it reaches the end of its blow. 1NDUCTlON Induction is the entrainment of room air by the air ejected from the outlet and is a result of the velocity of the outlet air. The air coming directly from the outlet is called primary air. The room air which is picked up and carried along by the primary air is crtllcd secondary air. The entire stream, corn- , posed of a mixture of primary and secondary air, is called total air. Induction is expressed by the momentum equation: M, V, + M, V, = (M, + MJ x V.1 where M, = mass of the primary air M, = mass of the secondary air V, = velocity of the primary air V, = velocity of the secondary air V,) = velocity of the total air Induction ratio (R) is defined as the ratio of total air to primary air; total air primary + secondary air R= = primary air primary air IMPORTANCE OF INDUCTION Since blow is a function of velocity and since the rate of decrease of velocity is dependent on the rate )f induction, the length of blow is dependent on the amount of induction that occurs. The amount of induction for an outlet is a direct function of the perimeter of the primary air stream cross-section. For two outlets having the same area, the outlet with the larger perimeter has the greatest induction and, therefore, the shortest blow. Thus, for a given air quantity discharged into a room with a given pressure, the minimum induction and maximum blow is obtained by a single outlet with a round cross-section. Conversely, the greatest induction and the shortest blow occur with a single outlet in the form of a long narrow slot. SPREAD Spread is the angle of divergence of the air stream after it leaves the outlet. Horizontal spread is di- vcrgcnce in the horiLonta1 plane and vertical is divergence in the vertical plane. Spread incluclccl angle measured in dcgrccs. Spread is the result ol the momentum law. is an illustration of the effect of induction on area and air velocity. spread is the Fig. 63 strcatn O U T L E T , 1 S O FT 1000 1000 CFM FPM 2000 CFM 500 FPM VEL . F1c.63 -EFFECTOF~NDUCTION Example 7 - Effect of Induction Given: 1000 cfm primary air 1000 cfm secondary air 1000 fpm primary air velocity 0 fpm secondary air velocity Find: . The velocity and area of the total air stream when 1000 cfm of primary and 1000 cfm of secondary air are mixed. Solution: Area of the initial primary air stream before induction 1000 Ml =---=~=lSqft VI Substituting in the momentum equation (1000 x 1000) + (1000 x 0) = (1000 + 1000) v, v, = 500 Area of the total air stream = M, + M, v, 1000 + 1000 = 500 = 4 sq ft An outlet discharging air uniformly forward, no diverging or converging vane setting, results in a spread of about an 18” to 20” included angle in both planes. This is equal to a spread of about one foot in every six feet of blow. Type and shape of outlet has an influence on this included angle, but for nearly all outlets it holds to somewhere between Ijo and 23”. INFLUENCE OF VANES ON OUTLET PERFORMANCE Straight Vanes Outlets with vanes set at a straight angle result in a spread of approximately 19” in both the horizontal and vertical plane (Fig. 64). FIG. 65 - SPKEAI) FIG. 64 c - S PREAD WI.~H S TRAIGHT WI’I‘H CONVIXGING V ANES VANES erging Vanes Outlets with vanes set to direct the discharge air (Fig. 65j result in approximately the same spread (19”) as when the vanes are set straight. However, the resulting blow is approximately 15% longer than the straight vane setting. Diverging Vanes Outlets with vanes set to give an angular spread to the discharge air have a marked effect on direction and distance of travel. Vertical vanes with the end vanes set at a 45” angle, and all other vanes set at intermediate angles to give a fanning effect, produce an air stream with a horizontal included angle of approximately 60” (Fig. 66). Under this condition the blow is reduced about 50y,.‘Outlets with end vanes set at angles less than 45”, and all other vanes set at intermediate angles to give a fanning effect, a blow correspondingly larger than the 45” hv. setting, but less than a straight vane setting. Where diverging vanes are used, the free outlet area is reduced; therefore, the air quantity is less than for straight vanes unless the pressure is increased. To miss an obstruction or to direct the air in a particular direction, all vanes can be set for a specific angle as illustrated in Fig. 67. Notice that the spread angle is still approximately 19”. FIG. 66 - S PREAD WITFI DIVERGING V ANES INFLUENCE OF DUCT VELOCITY ON OUTLET PERFORMANCE An outlet is designed to distribute air that has been supplied to it with velocity, pressure and direction, within limits that enable it to completely perform its function. However, an outlet is not designed to correct unreasonable conditions of flow in the air supplied to it. FIG. 67 - S PREAD WITH S TRAIGHT V ANES S ET A T AN A N G L E 2-68 I’.\R’l‘ 2 I I LJ VA = D U C T V E L O C I T Y Vg= V E L O C I T Y D U E T O P R E S S U R E DIFFERENCE ACROSS OUTLET vC = R E S U L T A N T O U T L E T VELOCITY -It F1c.68 - OUTLETLOCATEDIN WITHOUT VANES WITH .\IIi DIS’I‘l<IIIU-I‘ION Wl~erc an outlet without vanes is located directly against the side ol‘ a duct, tllc tlirection oE blow 0E [lie air from the outlet is the vector sum 0E the duct velocity and the olttlct velocity (Fig. 68). This may he motlifictl by the 1)cculiarity 0C the duct opening. Wlicrc an outlet is applictl Lo the ktcc 0C the duct, the resultant velocity V. can be modified by adjustable V;LIICS behind the outlet. Whether they shbulti be applied or not depends on the amount of divcrgeticc Irom straight blow that is acceptable. OEtcn outlets are mounted on short extension collars away from the lace of the duct. Whenever the duct vciocity exceeds the outlet discharge velocity, vanes should bc used where the collar joins the duct. Results are indicated in Fig. 69. DUCT . IMPORTANCE OF CORRECT BLOW - VANES F1c.69 -COLLARFOR~UTLETS SUPPLY AIR WARMER THAN ROOM AIR . \ Normally it is not necessary to blow the entire length or width of a room. A good rule oE thumb to follow is to blow 3/4 oE the distance to the opposite wall. Exceptions occur, however, when there are local sources of heat at the end of the room opposite the outlet. These sources can be equipment heat and open doors. Under these circumstances, overblow may be required and caution must be gxercised co prevent draft conditions. SUPPLY TEMPERATURE DIFFERENTIAL The allowable supply temperature difference that can be tolerated between the room and the supply air depends to a great extent on (1) outlet induction ratio, (2) obstructions in the path of the primary air, and (3) the ceiling height. Fig. 70 indicates the effect of changing the supply air temperature from warm to cold. Since induction depends on the outlet velocity, there is a supply temperature differential which must be specified to give satisfactory results. TOTAL ROOM AIR MOVEMENT SUPPLY AIR EQUALS ROOM AIR TEMP SUPPLY AIR COOLER THAN ROOM AIR The object of room air distribution is to provide satisfactory room air motion within the occupied zone, and is accomplished by relating the outlet characteristics and performance to the room air motion as follows: 1. Total air in circulation = outlet cfm x induction ratio. 2. Average room velocity FIG. 70 -AIR STREAM PATTERNS TEAWERATURE FOR DIFFERENTIALS VARIOUS 1.4 X total cfm in circulation = area of wall opposite outlet(s) 3. K= average room velocity 1.4 X induction ratio outlet cfm = clear area of wall opposite outlet(s) where K is the room circulation factor expressed in primary air cfm/sq ft of wall opposite the outlet. The multiplier 1.4 allows for the blocking caused by the air stream. Note that the clear wall area is indicated in the equation and all obstruction must be deducted. See Note 8, Table 21. Table 19 indicates that the average room air movement should be kept between 15 and 50 fpm for most applications. Tests have been performed on outlets at various outlet velocities to determine perCc ante characteristics. The results of such tests 01. . jpecifi’c series of wall outlets (Fig. 93) are shown in the rating tables at the end of this chapter. This rating data can be successfully used for outlets having the nominal dimensions and free area indicated in Table 21. An example illustrating outlet selection accompanies the table. The K factor as indicated in Item 3 is shown at the bottom of the rating table as maximum and minimum cfm/sq ft of outlet wall area. TYPES OF OUTLETS PERFORATED GRILLE This grille has a small vane ratio (usually from 0.05 to 0.20) and, therefore, has little directional effect. Consequently, it is used principally as an exhaust or return grille but seldom as a supply grille. When a manual shut-off damper backs up this gT;‘b) it becomes a register. Fl);cti BAR GRILLE The fixed bar grille is used satisfactorily in locations where flow direction is not critical or can be predetermined. A vane ratio of one or more is desirable. To obstruct the line of sight into the duct interior, closely spaced vanes are preferred. ADJUSTABLE BAR GRILLE This grille is the most desirable for side wall location. Since it is available with. both horizontal and vertical adjustable bars, minor air motion problems can be quickly corrected by adjusting the vanes. SLOTTED OUTLET This outlet may have multiple slots widely spaced, resulting in about 10% free area. Performance is about. the same as for a bar grille of the same cfm and static pressure, but the blow is shorter because of greater induction at the outlet face. Another design to effect early completion of induction is the long single, or double, horizontal slot. It is particularly advantageous where low ceiling heights exist and outlet height is limited, or where objections to the appearance of grilles are raised. EJECTOR OUTLET The ejector outlet operates at a high pressure to obtain a high induction ratio and is primarily used for industrial work and spot cooling. When applied to spot cooling, a high degree of ejector flexibility is desired. INTERNAL INDUCTION OUTLET Where a sufficiently high air pressure is used, room air is induced thru auxiliary openings into the outlet. Here it is mixed with primary air, and discharged into the room at a lower temperature differential than the primary stream. Induction progresses in two steps, one in the outlet casing and the other after the air leaves the outlet. CEILING OUTLETS Pan Outlet This simple design of ceiling distribution makes use of a duct collar with a pan under it. Air passes from the plenum thru the duct collar and splashes against the pan. The pan should be of sufficient diameter to hide the duct opening from sight, and also should be adjustable in distance from the ceiling. Pans may be perforated to permit part of the air to diffuse downward. Advantages of the pan outlet are low cost and ability to hide the air opening. Disadvantages are lack of uniform air direction because of poor approach conditions and the tendency to streak ceilings. Ceiling Diffuser These outlets are improvements over the pan type. They hasten induction somewhat by supplying air in multiple layers. Approach conditions must be good to secure even distribution. Frequently they are combined with lighting fixtures, and are available with an internal induction feature. See Fig. 71. Perforated Ceilings and Panels Various types of perforated ceilings for the introduction of conditioned air for comfort and industrial systems are available. The principal feature of this method of handling air is that a greater volume of air per square foot of floor area can be introduced at a lower temperature, with a minimum of movement in the occupied zone and with less I I’\ll 2-70 FIG. 71 - INTERNAL~NDWXION danger of draft. Since discharge velocity is low, induction is low. Therefore, care must be taken to provide adequate room air motion in excess of 15 fpm. Duct designed for a perforated ceiling is the same as duct designed for a standard ceiling. To obtain adequate supply to all areas, the same care necessary for conventional systems must be taken in laying out ducts for the perforated ceiling. The ceiling panels should not be depended upon to obtain proper air distribution, since they cannot convey air to areas not otherwise properly supplied. Perforated panels do assist in “spreading out” the air supply and, therefore, comparatively large temperature differentials may be used, even with low ceiling heights. APPLICATION OF CEILING DIFFUSERS Installations using ceiling diffusers normally result in fewer complaints of drafts than those using side wall terminals. To eliminate or minimize these complaints, the following recommendations should be considered when applying ceiling diffusers. BLOW Select ceiling diffusers for a conservative blow, generally not over 75% of the tabulated value. Overblow may cause problems on many installations; under-blow seldom does. . CULISG 1‘ ‘2. \II< DISTKIBUTION DI~FUSEK PRESSURE DROPS Most rating tables express the pressure drop thru the outlet only and do not include the pressure drop necessary to force the air out of the duct thru the collar and outlet and into the room. Therefore, it is recommended that rated pressure drops be carefully investigated and the proper safety factor applied Tvhen necessary. DIFFUSER APPROACH ,111 important criterion for good diffuser performance is the proper approach condition. This means either a coilar of at lcast -1 times the duct diameter, or good turning vmcs. If vanes are used, they must be placed perpendicular to the air flow at the upper end of the collar 2nd spaced approximately 2 in. apart. OBSTRUCTIONS \\%el-e obstructions to the flow of air from the clilf~laer occur, blank off a small portion of the diffuser at the point at whicll the obstruction is located. Clip-on baffles arc‘ us11:111y provided for this purpose. OUTLET NOISE LIMITATIONS One important criterion affecting the choice of an outlet is its sound level. Table 20 shows recommen&d outlet velocities that result in acceptable jountl levels for various types of applications. (:l-l,\l’~l‘l~K 3. I1001Ll .\II< TABLE 20 -RECOMMENDED OUTLET VELOCITIES TERMINAL VELOCITY IFPM) APPLICATION Broadcast studios Residences Apartments Churches Hotel bedrooms Legitimate theaters Private officer, acoustically Private offices, not treated Motion picture theaters General offices Dept. storer, upper f!oors Dept. stores, main floor OUTLET 300-500 500-750 500-750 500-750 500-750 500-750 500-750 500-800 1000 treated 1000-l 250 1500 2000 LOCATIONS Interior architecture, building construction rend streaking possibilities necessarily influence the 1‘. , .ut and location of the outlet. However desirable (! FIG. 72 - DOWNDRAFT W 2-71 I>IS’I‘RIIIIJ’I’IOU I N D O W FROM COLD it may I)e t o loc;lte 211 outlet in 2 given .sl)ot, these items may prevent such locxtion. A f t e r 211 the forcgoin,g l i m i t a t i o n s have been successfully tle;tlt with, the 2ir distribution principles whic,h relate to flow, drop, c;ipicity xntl r o o m air circul;ltion create further limitations in tlesigning an xcccpt;~l)le air distribution system. These arc t;lbulatetl i n t h e lating t:kt)ies a t the end o f t h e chapter. L o c a l loads due t o people concentration, equipment heat, outside walls ant1 window locations frequently modify the choice of outlet location. The downdraft from a cold wall or a glass window (Pig. 72) can reach velocities of over 200 fpm, causing discomfort to occupants. Unless this downdraft is overcome, complaints of cold feet result. In northern climates this is accomplished by supplementary radiation, or by an outlet located under a window as illustrated in Fig. 73. FIG. 73 -DISCHARGE A IR OFFSETTING WINDOW DOWNDRAFT I’:\ li.1‘ 2. 2-72, Another item to consider wllcn choosing an outlet location is the radiant ellcct from cold or warm surfaces. During the heating season an outlet tlischarging warm air under a cold window raises the surface temperature and rcduccs the Lceling of discomfort. The following describe four typical applications of specific outlet types. CEILING DIFFUSERS Ceiling diffusers may be applied to exposed duct, furred duct, or duct concealed in a ceiling. Although wall outlets are installed on exposed and furred duct, they are seldom applied to blow directly downward unless complete mixing is accomplished before the air reaches the occupied zone. WALL OUTLETS A high location for wall outlets is preferred where a ceiling is free from obstructions. Where beams are encountered, move the outlet down so that the air stream is horizontal and free from obstruction. If this is not done and if vanes are used to direct the air stream downward, the air enters the occupied zone at an angle and strikes the occupants too quickly. This is shown in Fig. 74. Wall outlets located near the floor (Fig. 75) are suitable for heating but not for cooling, unless the air is directed upward at a steep angle. The angle must be such that either the air does not strike occupants directly or the secondary induced stream does not cause an objectionable draft. WINDOW OUTLETS Where single glass is used, window outlets are r AIR STREAM 2 BLOW HAS A THAN .\IR I)IS’I‘KlI~UI‘ION prcl’erretl to cithcr wall or ceiling distribution to oll‘set the pronoiinccd tlowntlraTt during Lhc winter. 7‘11~ air slwuIt1 be directed with vanes at an angle ol 15” or 20’ from the vertical into tile room. FLOOR OUTLETS Where people are seated as in :I theater, floor outlet distribution is not permissible. Whcrc people are walking about, it is possible to introduce air at the floor level; for cxainplc, in stores where air is directed horizontally thru a slot under a counter. In this application, liowevcr, a very low temperaturc tliffercntial of not more than 5 or 6 degrees must be used. Maintaining this maximum is usually uncconornical b e c a u s e o f the large air volume rcquired. However, if air is directed upward behind the counter and diffused at an elevation above the occupied zone, the temperature differential may be increased approximately 5 times. Another disadvantage is that floor outlets become dirt collectors. SPECIFIC APPLICATIONS If the principles described in the previous paragraphs are properly applied, problems after installation will be at a minimum. Basically, the higher the ceiling the fewer the number of problems encountered, and consequently liberties may be taken at little or no risk when designing the system. However, wiih ceiling heights of approximately 12 feet or less, greater care must be exercised to minimize problems. Experience has shown that ceiling diffusers are easier to apply than side wall outlets, and are preferred when air quantities approach 2 cfm/sq ft of floor area. The following general remarks about specific ap- OBSTRUCTION LONGER AIR EFFECTIVE STREAM 1 ,. F L O O R . j ‘- :, : .’ FIG. 74 -WALL OUTLET IN ROOM O BSTRUCTION W ITH CEILING FIG. 75 - W ALL OUTLET NEAR FLOOR l (:H:\I’TEK 3. ROOM 2-73 ,\IK I~IS’I‘RII1U’I‘I<)N plicatiotls arc the result of thousands of installations and are olfercd ;is a guide for better air distribution. Apartiileiits, hotels and olfice I)uildings are discussed ii1 relation to specilic location of s o u r c e s of air supply con~mon to these types of buildings. 12anks, restaurants, and department and specialty stores are discussed in more general terms, although the common sources of locations of outlets previously dis- 5 . Ilctiirii grille: ixturn xir tllrri tile c.orritlor arid return ducts arc not used, CVlierc u s e r e l i e f grilles o r t o untlcrcut i s I)crltlissil)le it is nccc~ssary to olliw doors. In apartments and hotels, cotlcs must 1)~’ checked bcforc using the corridor, as a return plcnuni. Even if codes permit, it is not good engineering practice to me the corridor as a return plenum. cussctl c a n be a p p l i e d . APARTMENTS, HOTELS AND OFFICE BUILDINGS 1. Corridor Supply - No direct radiation (Fig. 76): A dvur1 loge - Low cost. Lhadvantage - Very poor in winter. Downdraft under the window accentuated by the outlet blow. Precazllion - Blow must be not more than 75y0 of the room length. 2 Corridor Supply - Direct radiation under win.ows (Fig. 77): Advnntuge - Offset of downdraft under window in winter when the heat is on. Disadvantage - Slight downdraft still occurs during intermediate season or whenever radiation is shut off during cool weather. Precaution - Do not blow more than 75% of . room length. 3. Duct above window blowing toward corridor (Fig. 78): Advantage - Somewhat better than corridor distribution but does not prevent winter downdraft unless supplemented by direct radiation. Disadvantage - Nearly as expensive (considering building alterations) as window outlet which results in better air distribution. 4. Window outlet (Fig. 79): Advantage - Eliminates winter downdraft - by “ar the best method of distribution. 3isndvantage - May not be economical for multiple windows. ELEVATION FIG. 77 - CORRIDOR AIR S UPPLY WITH . I FIG. 78 - DC’CT ABOVE W INDOW , B LOWING TO~VARD CO R R I D O R _ r-l E L E V A T I O N ELEVATION 7 6 - CORRKJOR AIR S UPPLY 1 ELEVATION - I DIRECT RADIATION Flc. 59 - WIMOW OUTLET , 2-74 PART BANKS (FIG. 80) Often the center bank space has a high ceiling with an electrical load. In this case, use of side wall outlets part of the way up the wall may result in segregating some of the ceiling load and keeping it out of the occupied zone, thus permitting some reduction in cooling load. This location of outlets part way up the wall is used with ceiling heights in excess of 20 ft. DEPARTMENT STORES (FIG. 81) There is nothing critical about air distribution in department stores if ordinary precautions are observed, provided the ceiling is high enough. Care should be taken in conditioning a mezzanine since the outlet is likely to overblow and not cool its occupants. Longitudinal distribution is preferred. Base- 2 . .\lR DISTRIHUTION merits may give trouble clue to low ceilings and pipe obstructions. Main floors usually rquire more air near doors. RESTAURANTS (FIG. 82) Great care must be taken in locating outlets with respect to exhaust hoods or kitchen pass-thru windows. Usually the air velocities over such openings are low and excessive disturbance due to direct blow or induction from outlets may pull air out of them into the conditioned space. STORES 1. Outlets at rear blowing toward door (Fig. 83): Requirements - Unobstructed ceiling. Disadvanta,q - May result in high room circulation factor K. Precaution - Blow m’ust be sized for the entire KITCHEN WRONG . ELEVATION FIG. 80 t - A IR DISTRIBUTION WITH HIGH CEILING -RIGHT I I PLAN WRONG FIG . 82 - RESTAURANT A IR DISTRIBUTION RIGHT - ELEVATION FIG. 81 - MEZZANINE A IR DISTRIBUTION PLAN FIG . 83 - A IR DISTRIBUTION FROM REAR OF S TORE . (:H.\I’TEK 2. 3. 4. -_ 3. R O O M 2-75 .\IR DISTRIRUTION length of the room; otherwise a hot zone may occur due to infiltration at the doorway. Care must be taken to avoid downdrafts near walls. Outlets over door blowing toward rear (Fig. 84): Requirements - Unobstructed ceiling. Disadvnnlnge - May result in high room circulation. Precaution - Excessive infiltration may occur due to induction from doorway. Outlets blowing from each end toward center (Fig. 85): Advantage - Moderate room circulation. Pwcazdion - There may be a downdraft in the center. Outlets should be sized for blow not greater than 40 percent of the total length of the room. Center outlets blowing toward each end (Fig. 86): 4duantage - Moderate room circulation. Juct along side wall blowing across the store Adwmtnge - ISest air distribution. I1is/lcl71(ln&rge - High cost. THEATERS 1. Ejector system for sma!l theaters, no baIcony (Fig. 89): Re~Iuirements - Unobstructed ceiling 2nd ability to locate outlets in the rear wall. Acl7xiritage - Low cost. Precaution - Possibility of clcatl spots at front back of theater. Use mushrooms for return air under seats if excavated. In northern climates direct radiation may be advisable along the sides. xltl - (Fig. ZT): Advantage - Moderate room circulation. Precaution - Overblow may cause downdrafts on the opposite wall. 6. Ceiling diffusers (Fig. 88): Requil-ements - Necessary where ceiling is badly cut u p . FIG . 86 - A IR DISTRIBUTION FIG. 87 - A IR DISTRIBUTION CENTER FROM S IDEWALL OUTLETS FROM CEILING DIFFUSERS OF S TORE PLAN PLAN FIG. 84 FROM FROM OVER THE - A IR DISTRIBUTION DOOR - PLAN FIG . 85 - A IR DISTRIBUTION OF STORE FROM E ACH E ND FIG. 88 - A IR DISTRIBUTION 2-76 I’ART 2. .\IR DISTRII3UTION L d &jnyf tnge - ConiplcLc coverage, no tlcatl spots. Ilistrtl7urrrtfrge - I-1 igher first cost. ~~-ecll~l/io~~ - i\ir must not stnikc obstructions wit11 a velocity force that causes tlellcction and drafts in the occupiccl zone. Temperature tlillerentials must be limited in regions of low ceiling heights. USC low outlet vclocitics. ELEVATION FIG. 89 -AIR DISTRIUUTION FOR SMALL THEA?XES 2. Ejector system for large theaters with balcony (Fig. 90): Requiwrnents - Unobstructed ceiling. - Low cost. Precaution - Balcony and orchestra should have separate returns. Preferred location, under seats; acceptable location, along sides or rear of theater. Return at front of theater generally not acceptable. Outlets under balcony should be sized for distribution and blow to cover only the area directly beneath the balcony. Orchestra area under balcony should be conditioned by the balcony system. Allow additional outlets in rear for standees when necessary. 3. Overhead system (Fig. 91): Requirements - Necessary when ceiling is obstructed. A dunn tage RETURN Velocities thru return grilles depend on (I) the static pressure loss allowed and (2) the effect on occupants or materials in the room. In determining the pressure loss, computations should be based on the free velocity thru the grille, not on the fact velocity, since the orifice coefficient may approach 0.7. In general the following velocities may bc used: GRILLE LOCATION FPM OVER GROSS AREA Commercial Above occupied zone 800 and above Within occupied zone not nea; seats Within occupied zone near seats Door or wall louvers Undercutting of doors Industrial Residential 600-800 400-600 poo-1000 600x 800 and ahove 400 * *Thru L- GRILLES undercut area LOCATION BALCONY ELEVATION F1c.90 -AIRDISTRIBUTION FOR LARGE THEATRES WITHBALCONY THEATER ELEVATION FIG. 9 1 - OVERHEAD AIR DISTRIBUTION Even though relatively high velocities are used thru the face of the return grille, the approach velocity drops markedly just a few inches in front of the grille. This means that the location of a return grille is much less critical than a supply grille. Also a relatively large air quantity can be handled thru a return grille without causing drafts. General drift toward the return grille must be within acceptable limits of less than 50 fpm; otherwise complaints resulting from drafts may result. Fig. 92 indicates the fall-off in velocity as distance from the return grille is increased. It also illustrates the approximate velocities at various distances from the grille, returning 500 cfm at a face velocity of 500 fpm. Ceiling Return These returns are not normally recommended. Difficulty may be expected when the room circulation due to low induction is insufficient to cause warm air to flow to the floor in winter. Also, a poorly located ceiling return is likely to bypass the cold air in summer or warm air in winter before it has time to accomplish its work. , (:H:\I”1‘EK 3. l<OOM .\IR 2-77 I~IS’I‘KIIIIJ~I‘ION OUTLET SELECTION The following example describes a method of selecting a wall outlet using the rating tables on page 78. Example 2 - Wall Outlet Selection Given: Small store Dimensions - 32 ft x 23 ft x 16 ft Ceiling - flat Load - equally distributed Air quantity - 2000 cfm Temp diff - 25 F Find: Number of outlets Size of outlets Location PLAN FIG. 92-VELocrrY FALL-OFF FROM GRILLE Wall PER DISTANCE Return A wall return near the fldor is the best location. Wall returns near the ceiling are almost as undesirable as ceiling returns. Differences due to poor mixing in the winter are counteracted by a low return since the cool floor air is withdrawn first and is replaced by the warmer upper air strata. Floor Return These should be avoided wherever possible because they are a catch-all for dirt and impose a severe strain both on the filter and cooling coils. In7 -never floor returns are used, a low velocity sett: J chamber should be incorporated, Solution: First, find the required blow in feet and the wall outlet area (air movement K factor). The minimum blow is 75% of the room width for the given condition of equally distributed heat load. Therefore, the minimum blow necessary is s/d X 23 = 17.3 ft. The maximum blow is the width of the room. The outlet wall area K factor is- equal to the cfm supplied divided by the outlet wall area: 2000 i = 3.9 primary air cfm/sq ft wall area 32 X 16 Enter Table 21 and select one or more outlets to give a blow of at least 17.3 ft. Air movement must be such that the value of K equals 3.9 primary air cfm/sq ft and that this value falls within the maximum and minimum values which are shown at the bottom of the rating tables. The tables indicate that, to best satisfy conditions, four outlets, nominal size 24 in. x 6 in., are to be used. By interpolating, it is found that the four 24 in. x 6 in. outlets at 500 cfm have a range in blow of 17.5 to 34 ft. By adjusting the vanes the proper blow can be obtained. Also the velocity of the outlet is found to be about 775 fpm. This is well within the recommended maximum velocity of 1500 fpm, Table 20. The minimum ceiling height from the table is just over 9 ft. This is less than the height of the room: therefore, the outlet selection is satisfactory. The top of the outlets should be installed at least 12 in. from the ceiling, (Note 8, Table 21). FIG. 93-bVAu OUTLET ON WHICH RATINGS ARE BASED TABLE 21-WALL OUTLET RATINGS, FOR COOLING ONLY For Flat Ceilings OUTLET 2 5 0 FPM VELOCITY STATIC PRESSURE WITH M E T E R I F4G, P L A T E Nom. Size of Outlet VWle (and Free Setting Area) 500 FPM Str B = ,013, 22%” = ,015 45O = ,019 750 FPM Str B = ,024, 22%’ = ,028 45" = ,035 S t r B = . O l , 2 2 ’ / 2 ’ = , 0 1 5 1S t r B = , 0 2 4 , 2 2 ’ / i ” = . 0 4 3 1 S t r B = . O $ l , 45'.= .02B 4 5 " z-z I Air Air rem Difl Dif f( Tern - bJan .> ouan15 20 15 20 my _t i t y Mi (cfml Mi Clg It (cfm) - Clg - t- 22%’ = .OB2 a.0 Straight 22x0 450 30 -z 2.5 1.8 6.5 6.5 6.0 7.0 6.5 6.0 - -E5.1 3.5 7.5 6.5 6.5 7s 7.0 6.5 7.0 7.0 10 x 4 (21.7) Straight 22'/2O 450 37 3.5 2.5 I.8 6.5 6.5 6.0 7.0 6.5 6.0 - 7.4 5.5 3.7 7.5 6.5 6.5 7.5 7.0 6.5 7.0 7.0 Straight 221/2O 45O 44 3.5 2.5 1.8 7.5 7.0 6.5 7.5 7.0 6.5 7.5 7.0 Straight 22x0 45O 61 3.7 2.7 2.0 7.0 6.5 6.5 7.0 6.5 6.5 - 7.5 5.5 3.9 16 x 4 (35.9) 6.5 6.5 6.0 7.0 6.5 6.0 - 7.9 6.0 4.0 7.5 7.0 6.5 7.5 7.0 6.5 7.5 7.0 20 x 4 (45.5) Straight 22x0 45O 77 4.0 3.0 2.0 7.0 6.5 6.0 a.0 6.0 4.0 7.5 7.0 6.5 a.0 7.0 6.5 7.5 7.0 24 x 4 (55.0) Straight 22%" 45O 93 4.1 3.1 2.0 7.0 6.5 6.0 7.0 6.5 6.5 7.0 7.0 6.5 a.0 6.0 4.0 7.5 7.0 6.5 30 x 4 (68.3) Straight 22Yi0 45O 116 4.2 3.1 2.1 7.0 6.5 6.0 7.0 7.0 6.5 - a.0 6.0 4.0 36 x 4 (83.5) Straight 22s0 45O 140 4.4 3.3 2.2 7.0 6.5 6.0 7.5 7.0 6.5 8x6 (26.5) Straight 22'/2O 450 52 0 3.8 2.5 10 x 6 (34.0) Straight 22%" 45O 66 12 x 6 (41.6) Straight 22'h0 45O 16 x 6 (56.6) ran Diff (F) 15 5rpc M in Clg Ht FPM s t r B = , 0 5 1 , 2 2 ' / 2 0= . 0 6 1 45O = .OB Str B = .175, 22%” = .I9 45" = .27 - r 8x4 (16.9) Air . Clum - I tity IIcfm) - 3low vu r Tern0 D i f f z (F) 25 Ht 59 lo.0 7.5 5.0 7s 7.0 6.5 a.0 7.5 6.5 8.5 7.5 7.0 a9 17.0 13.0 9.0 8.5 9.0 7.57.5 6.5 7.0 9.0 a.0 7.0 75 10.5 8.0 5.4 7.5 7.0 6.5 a.0 a.5 7.5 7.5 6.5 7.0 112 18.0 13.0 9.0 8.5 9.0 7.58.0 6.5 7.0 x7 a.0 7.0 91 II.0 8.1 5.5 a.0 7.0 6.5 8.0 7.5 7.0 a.5 7.5 7.0 136 I a.0 13.0 9.0 9.5 a.5 7.0 122 11.0 8.1 5.5 iz 7.0 6.5 8.0 7.5 7.0 a.5 7.5 7.0 183 19.0 14.0 10.0 9.5 a.5 7.5 154 1I . 5 8.5 6.0 G 7.5 6.5 8.0 7.5 7.0 8.5 a.0 7.0 231 20.0 15.0 10.0 9.5 a.5 7.5 s.0 7.0 6.5 8.0 7.5 185 7.0 11.5 a.5 6.0 ii 7.5 6.5 8.0 7.5 7.0 8.5 a.0 7.0 276 20.0 15.0 10.6 10.0 a.5 7.5 7.5 7.0 6.5 iii 7.5 6.5 8.0 7.5 233 7.0 12.0 9.0 6.0 8.0 7.5 6.5 a.0 7.5 7.0 8.5 a.0 7.0 349 21.0 16.0 II.0 Gi a.5 7.5 8.0 6.0 4.0 7.5 7.0 6.5 7.5 6.5 7.5 279 7.0 a.0 12.0 9.0 6.0 7.5 6.5 420 21.0 16.0 II.0 10.0 8.5 7.5 G7.0 4.8 8.0 7.0 6.5 - G 7.5 7.0 a.5 8.0 7.0 lb3 KY 10.0 6.0 s.f, 7.5 7.0 155 24.0 I a.0 12.0 10.5 9.5 a.0 5.5 4.1 2.8 7.5 r.s 7.0 7.0 6.0 6.5 7.5 8.0 7.0 7.5 6.5 7.0 10.0 7.5 5.0 a.0 7.5 7 .0 - a.r, 8.0 7.0 9.0 8.5 7.5 15.0 11.0 7.0 G 131 196 27.0 I 20.0 14.0 II.5 10.0 8.0 7.5 7.0 6.5 - 8.0 7.5 7.0 - 11.0 a.1 5.5 8.0 7.5 7.0 9.0 a.5 238 7.5 15.0 II.0 7.0 G a.0 9.5 80 6.0 4.5 3.0 28.0 1 21.0 ? 4.0 11.5 10.0 - a-.-0L Straight 221/2O 45O 8.0 7.5 7.0 12.0 9.0 6.0 a.5 a.0 7.0 9.5 8.5 214 7.5 16.0 12.0 a.0 321 7.5 30.0 1 22.0 15.0 12.5 10.5 a.5 20 x 6 (71.5) Straight 22Yi0 45O 8.5 7.5 7.0 12.0 9.0 6.6 9.0 a.0 7.0 - 9.5 0.0 9.0 7.5 17.0 13.0 9.0 c5 8.5 7.5 1 135 8.0 7.0 6.5 8.0 7.5 7.0 - 9.0 107 6.2 4.7 3.2 403 32.0 1 24.0 16.0 iiT 11.0 9.0 24 x 6 (86.5) Straight 22%" 450 8.5 8.0 7.0 - 13.0 10.0 6.5 0.0 9.0 324 i2 a.0 a.5 7.5 I 0.511.0 9.010.0 a.0 a.5 466 7.5 Ia . 0 13.0 9.0 33.0 1 25.0 1 17.0 13.0 11.0 9.1 30 x 6 (109.0) Straight 22%" 45O 203 8.5 7.5 7.0 a.5 a.0 0.0 9.0 7.5 0.5 9.0 406 19.0 14.0 10.0 0.0 9.0 7.5 1 1.0 II.5 9.510.0 a.0 8.5 609 7.0 13.0 IO.0 6.5 9.0 8.0 7.0 9.0 a.0 7 .5 - 95 162 8.0 7.5 7.0 - 34.0 I 25.0 1 17.0 13.5 11.5 9.0 36 x 6 (131.3) Straight 22%" 45O 245 8.5 7.5 7.0 - 9.0 8.0 7.5 - 13.0 IO.0 6.5 - 9.5 a.5 7.5 - iz 9.0 0.5 9.5 490 19.0 14.0 10.0 - iz 1 1.0 12.0 9.0 9.510.0 8 . 01 a . 0 a . 5 - 735 35.0 1 26.0 1 18.0 - 14.0 11.5 9.5 - 12 x 4 (24.6) '. 375 Str B = .Ol 22’/1” = .Ol 4;o = .Ol STATIC PRESSURE STANDARD OUTLET - - 6.6 5.0 3.2 7.5 7.0 139 6.5 9.0 8.0 202 7.5 K a.0 7.0 a.0 7.0 a.5 7.5 a.5 a.0 - a.0 a.0 a.0 a.0 159 269 a.0 a.0 a.0 - a.0 7.0 - a.0 7.0 9.5 a.5 FACTOR M a x C f m / S qF t O u t l e tW a l l Area 29.0 19.0 14.0 M i n C f m / S qF t O u t l e tW a l l Area a.7 5.7 4.2 1 c - 9.6 2.9 ' (:M/\I’rI‘EK 3. ROOM .\II< I>IS’l‘IllIllJ’l‘ION TABLE 2-79 2LWALL OUTLET RATINGS, FOR COOLING ONLY (Cont.) For Fiat Ceilings OUTLET VELOCITY 2000 FPM STATIC PRESSURE STANDARD OUTLET S t r 8 = .375, 22%” = .42 45’ = .565 STATIC PRESSURE WITt METERING PLATE Nom. Sire of Outlet (and Free Area) _, VatI* Setting 8 x 4 (16.9) Straight 22%” 45O 10 x 4 (21.7) Straight 221/2O 45O 12 x 4 (24.6) Straight 22%O 45O Air -C hantity I :cfm) 118 237 - 150 224 299 181 10.0 9.0 7.5 272 362 x 4 .5.9) Straight 22vi” 45O 244 10.5 9.5 7.5 366 488 65 49 33 20 x 4 (45.5) Straight 22x0 4s” 308 462 616 67 50 34 24 x 4 (55.0) Straight 22x0 45O 556 740 30 x 4 (68.3) Straight 22s” 45O 36 x 4 (83.5) Straight 22’/20 4s” 558 8 x 6 (26.5) Straight 22x0 45O 206 36 27 18 9.5 9.0 7.5 10 x 6 (34.0) Straight 22%” 45” 262 40 30 20 12 x 6 (41.6) Straight 22%” 45O 318 41 16 x 6 ‘26.6) Straight 22%” 45O 428 20 x 6 (71.5) Straight 22x0 45O 24 x 6 (86.5) Straight 22x9 45” 648 30 x 6 (109.0) Straight 22’/2O 45O 812 36 x 6 (131.3) Straight 22%” 45O 980 Cfm/Sq Ft Outlet Will krea Min Cfm/Sq F t Outlet Wall Area 12.0 10.0 7.5 11.5 9.5 7.5 12.5 10.0 12.5 10.5 8.0 68 51 34 11.0 12.0 9.0 10.0 7.5 7.5 11.0 12.0 9.5 10.0 7.5 8.0 11.5 12.0 9.5 10.0 7.5 8.0 - 932 70 53 35 11.5 9.5 7.5 - 12.5 10.0 8.0 13.5 11.0 8.5 71 53 36 11.5 12.5 9.5 10.5 7.5 8.0 12.5 14.0 10.5 11.5 9.0 8.5 ! 7.0 ! 7.5 370 30 22 15 9.5 8.5 7.0 10.0 9.0 7.5 10.5 9.5 7.5 466 30 22 15’ 9.5 8.5 7.0 10.0 9.0 7.5 10.5 9.5 7.5 31 9.5 10.0 23 8.5 9.0 16 ! 7.0 ! 7.5 11.0 9.5 8.0 840 1116 11.0 9.5 8.0 12.0 10.0 8.0 310 412 11.0 9.5 8.0 12.0 10.0 8.5 13.0 11.0 9.0 392 - 524 t 31 10.0 10.5 11.5 12.5 13.5 11.0 9.0 476 636 44 12.0 13.0 3 3 I1 0 . 0 I 1 1 . 0 22 ! 8.0 ! 9.0 14.0 11.5 9.0 642 856 47 12.5 13.5 3 5 I1 0 . 5 I 1 1 . 5 14.5 12.0 9.5 t 21 538 ! 8.0 ! 8.5 H 36 24 10.5 8.5 806 15.5 12.5 9.5 076 972 296 5 0 1 3 . 0 1 4 . 5 15.5 38 11.0 12.cI 12.5 25 9.0 9.5 I 10.0 1218 1624 5 1 1 3 . 5 1 5 . c I 16.0 38 11.0 12.cI 13.0 26 9.0 9.5 110.0 A 1470 1960 7.2 2.2 11.5 9.5 698 K ? 6 0 10.5 45 9.0 30 7.0 62 IO.5 47 9.0 31 7.5 - - 82 62 41 92 14.0 6 9 11.5 46 9.0 - 4.8 has vertical lowres straight f o r w a r d i n the center, with uniformly incrsosing ongulor d e f l e c t i o n lo o moximum ot each end. The 4.5’ div e r g e n c e signifies o n ongulor deflection at each end of the outlet of 45’. and simitorly for 22%’ divergence. 5. Velocity 1.4 6. on effective Static Pressure is that pressure required to produce the indi- The Minimum Ceiling Height (table) is the minimum ceiling height which will give proper operation of the outlet for rho 17.0 13.5 10.0 18.0 14.5 108 16.0 81 13.0 5 4 IO.0 - 17.5 14.0 10.5 19.0 15.0 III 83 56 I15 86 58 - 18.5 14.5 10.5 19.5 15.5 11.5 17.5I 19.c 1 3 . 1> 15.c 1 OS1 11.c 20.5 16.0 12.0 1.1 is bossd face ar*o. cared velocities ond is mea,13.5 wed in incher of water. 11.0 8.5 G 7. Meorure ceiling height in the CLEAR only. This is the distonca from the floor lo the towes, <oil12:o ing beam or obstruction. 9.0 17.0 13.5 10.5 3.6 lo- 4 . Divergent Blow blow, and cfm. The octuol meoswed c e i l i n g h e i g h t m u s t b e equal t o o r greater t h a n t h e minimum ceiling height for the selection mode. Preferably the top of on outlet should be not ICII t h a n t w i c e t h e outlet’. height below the minimum ceiling height. 11.0 11.0 9. CfmlS. Ft Outlet Wall Area is thi stbndord for judging total room air movcmcnt. T h e moxi- a&&d thot furniturc,‘peopla, etc.. obstruct 10% of the room w&-section. If I’& obstructions vary w i d e l y f r o m IO%, the “.I”.I o f the cfm/sq ft oY1l.t wall orco should be tampered accordingly. 21.0 16.5 12.0 IO. I upward U n d e r b l o w . I t is n o t olwoy, nscasrary t o b l o w t h e anpi,. length of the room unl.,, there a<* hoot l o a d sovrce, ot t h a t a n d , squipmant load, open doors, sun-glass, etc. Considering the concentration o f r o o m hoot load on ths basis of Btu/(hr) (rq ft), the outlet blow rhould cover 75% of the heat load. - 15.5 12.5 9.5 FACTOR I 3. 15.5 12.5 9.5 119 18.C 1 19.: 09 14.c 1 15.c 60 10.: i 11.0 I- oh indicoter dktonca from out. let lo Iha p o i n t where Ihs air stream is substantially dirsipatsd. 13.0 10.5 8.0 9 4 14.5 7 0 11.5 47 9.0 102 15.5 7 7 12.5 51 9.5 17.0 13.5 IO.0 the 2. Blow 8.0 13.0 10.5 8.5 _ deflect word the celling. I- 11.5 9.5 7.5 10.5 9.5 7.5 15 to /np Tet 15 20 2s (fl) M lir I CI( H t - - - 58 10.5 11.0 12.0 44 8.5 9.0 9.5 29 7.0 7.0 7.5 a SIOW 10.0 8.5 7.5 Max I. W h e n e m p l o y i n g ratings ,o, flat ceiltngs, i t i* undsrrtaod that the front IOUYI.I ore ,et Str B = 1 . 3 6 Air Puon. tity (cfm) NOTES: F o r appticotions reqetring o limiting sound level-the out. 1st v e l o c i t y is l i m i t e d b y t h e sound generotcd by the outlet. 2-W !'.\Kl' L'. \I11 I~ISl‘l~ll~~J'I‘lON / TABLE 2I-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.) For Flat Ceilings OUTLET VELOCITY STATIC PRESSURE STANDARD OUTLET S T A T I C P R E !SSl I R E W I T H METERING PLATE Nom. Size of Outlet (and Free Area) VlYn.3 Setting 250 FPM 375 FPM I str B = .Ol, 2 2 % ” = .Ol 4 5 ” = .Ol I str E = ,013, 2 2 % ” = , 0 1 5 4 5 ” = .019 500 FPM S t r B = . 0 1 , 2 2 % '= . 0 1 5 1 S t r B = . 0 2 4 , 2 2 '?O / = , 0 4 3 I str E = .061. 22x0 = .082 450x tl?R I 45’ = .065 45” = .I 1, I Air C luanmy I lcfm) T e m p Dil Fpi? E M Air (F) Quan. _ 25 t i t y _ i CIS H t 9.0 9 . 5 - fHow ‘” (f0 (cfm) 12 x 8 (56.7) 113 16x a (77.1) Straight 22x0 45" 155 8.0 6.0 4.0 9.0 8.0 7.5 - 9.5 8.5 7.5 10.0 9.0 231 8.0 20 x 8 (97.6) Straight 22'hO 45O 192 8.5 6.5 4.3 9.5 8.5 7.5 - 10.0 9.0 8.0 - 10.5 9.5 287 8.0 24 x 8 (118.0) Straight 22'/2O 45O 231 9.0 6.9 4.5 9.5 10.0 8.5 9.0 7.5 8.0 - 30 x 8 (149.0) Straight 22'/2O 45O 209 9 . 5 1I O . 0 7.0 9.0 4.7 8.0 10.5 9.5 8.0 36 x 8 (179.0) Straight 221/2O 45O 350 9 . 9 1I O . 0 7.5 9.0 5.0 8.0 11.0 11.5 9.5 10.0 8.5 9.0 16 x 10 (97.7) Straight 22x0 450 198 9.a 7.1 5.0 95 9.0 8.0 - Ki 9.5 8.5 - 2 0 x 10 (124.0) Straight 22'/i0 45O 249 10.5 8.0 5.2 Il.0 10.0 8.5 - 2 4 x 10 (150.0) Straight 22%" 45O 300 11.0 8.4 5.5 1'0.5 9.5 8.0 1I I . 0 1 0.0 9.0 - 12.0 10.5 9.0 - 21.012.5 16.0 11.0 10.5 9.0 3 0 x 10 (195.0) Straight 22x0 45O 364 12.0 9.0 6.0 I' 2 . 0 1 0.5 9.5 - :2.5 Il.0 9.0 22.013.5 16.0 11.5 11.0 9.5 3 6 x 10 (227.0) Straight 22'/20 45O 453 12.4 9.1 6.1 I 2.0 1 0.5 9.5 13.0 II.5 9.0 16 x 12 (118.0) Straight 22x0 450 244 Il.0 I 8.1 1 5.5 20x12 (150.0) Straight 22%" 45O 307 24 x 12 (181.0) Straight 22x0 45" 30x12 (228.0) 36 x 12 (275.0) a.s 7.5 7.0 - 8.0 7.5 , lowTen E (fl) _1J_ M , Diff (F) zig 22.0 16.0 Il.5 463 40 13.5 30 11.0 20 9.0 14.515.5 12.012.5 9.510.0 16.0 10.5 12.0 9.0 8.0 8.0 24.0 18.0 12.5 575 43 14.0 32 11.5 22 9.5 15.516.5 12.513.5 10.0 10.5 10.5 9.5 346 8.5 17.0 10.5 13.0 9.5 8.5 8.0 25.0 19.0 13.0 - 692 45 14.5 34 12.0 23 9.5 16.0 17.0 , 13.0 14.0 10.010.5 11.0 10.0 435 8.5 18.0 11.0 13.0 9.5 9.0 8.5 26.0 19.0 13.5 868 46 15.5 35 12.5 23 10.0 17.018.0 13.515.0 10.511.0 18.0 11.5 13.0 9.5 9.0 8.5 27.0 20.0 14.0 1048 48 16.0 36 13.0 24 10.0 18.0 19.0 14.015.0 10.5 11.5 18.0 11.5 13.0 10.0 9.0 8.5 27.0 13.0 14.0 15.5 20.0 11.0 12.0 12.5 14.5 9.0 9.5 10.0 - 595 48 15.5 36 13.0 24 10.0 18.019.0 14.515.0 10.5 11.5 29.0 22.0 15.0 13.5 15.0 16.0 12.0 12.5 13.5 9.5 10.0 10.5 746 51 17.0 38 13.5 26 10.5 18.520.0 15.016.5 11.012.0 30.0 22.0 15.5 14.5 16.0 17.0 12.0 13.0 14.0 10.0 10.5 11.0 a99 55 18.5 41 14.5 28 11.0 20.021.0 15.517.0 12.012.5 14.516.0 12.0 13.0 751 10.010.5 32.0 24.0 16.5 - 15.0 17.0 18.5 13.0 14.0 15.0 10.0 10.5 11.5 1126 58 19.5 44 15.5 29 11.5 21.523.0 17.018.5 12.513.5 22.014.0 16.0 12.0 11.0 9.5 15.016.0 12.5 13.5 904 10.010.5 33.0 25.0 17.0 15.0 17.5 19.0 13.0 14.0 15.5 11.0 11.0 12.0 1355 60 20.0 45 16.0 30 12.0 22.023.5 17.519.0 12.513.5 21.012.5 16.0 11.0 11.0 9.0 13.515.0 11.5 12.5 488 9.510.0 Il.0 23.0 16.0 14.5 16.0 17.0 12.0 13.0 14.0 10.0 10.5 11.0 733 55 18.5 41 14.5 28 11.0 20.021.0 16.017.0 12.012.5 918 60 45 30 2o.c 16.c 12.C 22.0 23.5 17.5 19.0 12.513.5 1110 64 21.1 48 17S 32 12.C 24.025.0 18.520.0 13.514.5 1388 68 51 34 23.c 18.C 13.c 26.027.5 20.021.5 14.015.0 1673 7124.: 53 19.c 36 13.C 27.529.0 21.0 22.5 14.515.5 525 13s 1 1.c 9.c - 370 13.0 10.0 6.5 13.: 12.c 9.: - Straight 221%" 45O 462 13.9 10.0 7.0 14.! 12.: 10s Straight 22%" 45O 560 14.5 11.0 8.0-L - 15.: 13.c 1o.c I 8.7 GG 14.0 9.5 10.0 8.5 7.0 7.5 I 13.514.0 11.012.0 9.0 9.5 170 12.1 9.1 6.0 29.0 Air Temp Diff (F) ( auanTi-pipr tity Min Clg Ht (cfm) 339 8.5 7.5 12.5 11.0 367 9.5 M i n C f m / S qF t O u t l e tW a l l A r e a et) str 8 = .175, 22%” = .I9 45’ = .27 36 12.0 27 10.0 18 8.5 12.0 IO.5 9.0 M a x C f m / S qF t O u t l e tW a l l A r e a e IlOH & M Straight 22%" 45O 750 FPM Str B = ,024, 22%" = .028 4 5 ” = ,035 15.0 1II.0 13.0 14.0 15.5 11.0 12.0 12.5 9.0 9.5 10.0 14.0 12.0 460 9.5 22.0 14.c 15.016.0 16.0 12.C 12.513.5 11.0 9.2 10.010.5 613 33s 25.C 17.c 16.0 17.5 19.0 13.0 14.0 15.5 1 1 . 0 1 1 . 01 2 . 0 14.7 12.5 555 10.0 24.0 14.: 16.018.0 18.0 12.C 13.0 14.0 740 1 2 . 0 1 o . c 10.0 11.0 35.c 26.C 18.C - 17.0 18.5 20.0 14.0 15.0 16.0 11.0 11.0 12.5 15.5 13.0 695 10.5 25.0 15.: 17.018.0 19.0 13.c 14.015.0 12.0 1o.c 10.511.0 37.c 28.C 19.c 18.0 20.0 21.5 14.5 16.0 17.0 16.5 14.0 10.5 27.0 16.C 20.0 13.5 13.0 10.5 836 t K 925 39.c 29.C 2o.c - FACTOR 19.0 14.0 9.6 5.7 4.2 2.9 (;~-I.\I’IxR 3. itoobr .\III I)Is’r‘1~II:II’r.IoN 2-81 TABLE 21-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.) For Flat Ceilings 1000 FPM OUTLET VELOCITY ~._.. - - STATIC PRESSURE T A N-D A R D O U T L E T - S___--~ STATIC PRESSURE WITH Str B = .33, 22'/2" = METERING PLATE L - r Nom. Size Air of Outlet Vane C2uan- e (and Free Setting tity Area) (cfm) 1500 FPM -.__....~ S t r B = , 2 1 1 , 22%" = - 2000 FPM .24 StrB = , 3 7 5 , 2 2 % " = . 4 2 45" = ,565 B = ,715, 22X0= .74 45O = 1.15 .33 # D i f f (F) . E 201 Clg Ht 12 x 8 (56.7) Straight 22%" 45" 452 16 x 8 (77.1) Straight 22%" 45" 616 20 x 8 (97.6) Straight 22%" 45O 770 62 16.0 47 13.0 31 10.0 ?4 x a 18.0) Straight 22x0 450 920 65 17.0 49 13.5 33 10.5 19.020.0 15.0 16.0 1384 11.012.0 30 x 8 (149.0) Straight 22'/z0 45" 1160 60 17.5 51 14.0 34 10.5 19.521.0 15.5 17.0 1736 11.512.5 36 x 8 (179.0) Straight 22x0 45O 1404 71 18.5 53 14.5 36 11.0 16x10 197.7) Straight 22%O 45O 20x10 (124.0) 15.016.5 12.513.5 10.010.0 18.0 19.0 14.015.5 10.511.5 678 16.518.0 1 3 . 5I 1 4 . 5i926 10.01 1 l.O! I . W h e n e m p l o y i n rg a t i n g s ‘ o r Str B = 1.36 flat Blov (ft) 904 121 91 62 TY I Temp Diff (F) I5~ 2. p5 18.0 20.0 21.5 14.0 15.5 17.0 10.5 11.5 12.0 13a 100 67 21.023.0 15.5 18.0 12.013.0 1540 144 108 72 21.0 23.5 26.0 16.5 18.0 20.0 12.0 13.0 14.0 1 0 72 0 . 0 80 15.5 54 11.5 22.524.5 16.518.5 12.513.0 1840 151 113 76 22.0 25.0 27.5 17.5 19.0 20.5 12.0 13.5 14.5 111 21.0 83 16.5 56 11.5 t 23.525.5 17.519.5 13.013.5 2320 157 118 79 23.5 26.0 29.5 18.0 20.0 21.5 13.0 14.0 15.0 20.522.0 16.0 17.5 2096 12.012.5 116 21.5 87 17.0 58 12.0 25.026.5 18.020.0 13.014.0 2808 164 123 82 24.5 27.5 31.0 19.0 21.0 22.5 13.0 14.5 15.5 792 71 18.0 53 14.5 3611.0 20.522.0 16.0 17.5 1190 12.012.5 116 21.0 87 17.0 58 12.0 25.026.5 19.020.0 13.014.0 1584 164 123 82 Straight 22%" 450 994 75 19.5 56 15.5 38 11.5 22.024.0 17.0 18.5 1492 12.513.5 1 2 22 3 . 0 92 18.0 61 13.0 26.529.0 19.021.5 14.015.0 1988 174 130 a7 2 4 x 10 (150.0) Straight 22Yi" 45" 1200 a021.0 60 16.5 40 12.0 24.025.5 18.5 19.5 1798 13.014.0 <l 24.5 98 19.0 66 13.5 28.531.0 20.022.5 14.515.5 2400 185 139 93 3 0 x 10 (195.0) Straight 22%" 45" 1502 25.527.5 19.5 21.0 2252 14.015.0 139 26.5 104 20.5 70 14.0 30.534.0 21.524.5 15.516.5 3004 196 147 98 36x 10 (227.0) Straight 221/2O 45O I808 87 23.5 65 18.5 44 13.0 26.528.5 20.0 21.5 2710 14.515.0 142 27.0 106 21.0 71 14.5 31.535.0 23.025.0 16.017.0 16 x 12 (118.0) Straight 22%O 45O 976 81 21.0 61 16.5 41 12.0 24.025.5 18.5 19.5 1466 13.014.0 131 24.5 98 19.0 66 13.5 28.031.0 21.522.5 14.515.5 -0x12 (150.0) Straight 22w 45O 1226 142 27.0 106 21.0 71 14.5 2 4 x 12 (181.0) Straight 22x9 45O 1480 153 29.0 115 22.0 77 15.0 34.0 37.5 25.026.5 17.0 18.0 30x12 (228.0) Straight 22'/2O 45O 1850 163 31.5 122 24.0 8216.0 36.5 40.5 27.0 28.5 18.0 19.0 36x 12 (275.0) Straight 22Yi" 45" 2230 172 33.5 129 25.0 0616.5 38.042.5 28.030.0 18.5 20.0 1 6 . 0 /1 7 . 5 1 -L K + Blow mdicoter Min Clg Ht 1232 t ceilings, it is undwrteod the front louvrer are re, to deflect the air upword toward the ceiling. that Air Quantity (cfm) 19.521.0 14.017.0 11.512.0 43 12.5 29 9.5 NOTES: dirtonts from out- pat-d. 20.0 22.0 24.0 15.5 17.0 18.5 11.0 12.0 13.0 4. Divergent S l o w ha vertical lowres Itmight forward in the center, with uniformly increasing angular d e f l e c t i o n 10 a m a x i mum 01 each end. The 45” div e r g e n c e signifisr a n a n g u l a r dsflection 01 each e n d o f t h e outlet o f 450, and similarly for 22%0 divsrg.ma. Velocity is foes area. bored on effective Measure ceiling height in the CLEAR only. This is the dirtonce from Ihs floor to the lowest ceiling beam or obstruction. The Minimum Ceiling Height (table) is the minimum ceiling height which will give proper operation of the outlet for the given outlet velocity, vane setting, temperature difference, blow, and cfm. The odwl mcorwed c e i l i n g h e i g h t m u s t be equal to 01 g r e a t e r t h a n t h e minimum ceiling height for the selection made. Preferably the top of an outlet should be not less than t w i c e t h e outlet’s height below the minimum cciling height. 28.0 32.0 36.5 2 1 . 5 I2 4 . 0I 2 5 . 5 1 4 . 5 /1 6 . 5 !1 7 . 0 31.535.0 I’21 36 .. 50 21 57 .. 00 9. Cfm/Sq Ft Outlet Wail Area is the rtondard for judging total room air m o v e m e n t . The m a x i m u m value shown results i n an air mwement in Ihe zone of ocCUPD~CY o f about 5 0 fem. It i s aswmeb t h a t furniture,‘peopls, etc., obstruct IO% of the room UOII-section. If the obstructions vary w i d e l y f r o m IO’%. t h e values o f the cfm/rq f t outlet w a l l area should be tempered accordingly. 10. For applicqtionr requiring a limiting round level--the out. FACTOR Max Cfm/Sq F t O u t l e tW a l l Area 7.2 4.8 3.6 Min Cfm/Sq Ft Outkt Wall Area 2.2 1.4 I.1 l e t v e l o c i t y i s l i m i t e d b y the sound generated by the outlet. TABLE 21-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.) For Beamed Ceilings I OUTLET STATIC PRESSURE STANDARD OUTLET STATIC PRESSURE WITH METERING PLATE Nom. Size of Outlet (and Free Area) 375 250 FPM VELOCITY Str B = .024, 22%" = .043 450= .065 str B = .OI, 22'/b0 = 015 450= 03A . - 8x4 (16.91 Straight 22Yl" 45O 10 x 4 (21.7) Straight 22%" 45O 3.5 2.5 1.8 8.2 7.6 57 7.0 f 7.4 5.5 3.7 8.2 8.8 7.5 8.0 6.9 7.2 9.2 8.2 7.5 12 x 4 (24.6) Straight 22'/rt0 4s" 3.5 2.5 1.8 8.2 7.6 7.0 7.5 5.5 3.9 - 8.3 8.8 7.6 8.0 7.0 7.2 16 x 4 (35.9) Straight 22%" 45O 20 x 4 (45.5) Straight 22%" 45O 7.9 6.0 4.0 8.0 6.0 4.0 - 24 x 4 (55.0) Straight 22J/l" 45O 93 30 x 4 (68.3) Straight 22x0 45O 36 x 4 (83.5) Straight 22Yl" 45O 8x6 (26.5) Straight 221x0 45O 5: 10 x 6 (34.0) Straight 22x0 45O 6t 12 x 6 (41.6) Straight 22%" 45O a( 16 x 6 (56.6) Straight 22x0 45O 10; 20 x 6 (71.5) Straight 22x0 45O 5 3 2 24 x 6 186.5) Straight 22%" 45O 3 1 5 30 x 6 (109.0) 36 x 6 (131.3) 61 emp Diff (IF) t5 %-Kg t str B = . 061I 22'/z0 = .082 45O= ,118 Air cavan. tity (cfm) 8.1 7.5 7.0 8.2 8.7 7.5 7.9 ,5 . 5 6 . 5 44 68 3.7 2.7 2.0 10.0 7.5 5.0 (cfm) I- 9.3 0.4 7.5 tity 9.8 a.8 a9 7.0 17.0 13.0 9.0 9.7 8.5 7.4 9.2 8.3 7.5 91 11.0 8.1 5.5 8.9 8.1 7.2 9.4 9.9 8.5 8.9 7.67.9 136 18.0 9.9 13.0 8.6 9.0 7.5 8.4 8.9 7.7 8.1 7.0 7.3 9.4 8.4 7.6 122 11.0 8.1 5.5 9.0 8.2 7.3 9.610.0 8.6 9.0 7.77.9 la3 19.010.0 14.0 8.8 10.0 7.6 8.4 9.0 7.8 8.2 7.07.4 9.4 8.5 154 7.6 11.5 8.5 6.0 9.1 8.3 7.4 - 9.6 10.1 8.7 9.1 7.8 8.0 231 20.010.2 15.0 8.9 10.0 7.6 11.5 8.5 6.0 9 . i, 8 . 3I 7 . 4I 9.610.2 8.8 9.1 7.8 8.0 278 20.010.2 15.0 8.9 10.0 7.6 9.: 8 . LI 7 . 4I 9.710.3 8.5 9.0 7.88.2 7.17.4 9.5 8.6 7.6 116 4.2 3.1 2.1 8.5 7.9 7.2 175 8.0 6.0 4.0 - 8.5 9.0 7.9 8.3 7.17.5 9.6 0.6 233 7.7 12.0 9.0 4.6 i4a 4.4 3.3 2.2 - 8.5 7.9 210 7.2 8.0 6.0 4.0 8.5 9.1 7.9 8.3 7.17.5 9.6 8.6 279 7.7 12.0 9.0 6.0 - 5 . c) 3 . c3 2 . :i 9.0 8.3 7.4 77 9.5 7.0 4.8 9.0 9.6 8.3 8.8 7.37.8 1IO.2 : 9.2 103 8.0 5.15 4 . 1I 2 . fI __ 6.c1 4 . :5 3.c) 9.6 8.7 7.8 98 9 . 6 1 0 . 2'1 1 0 . 9 8.7 9.3 9.8 7.6 8.0 8.4 ia5 131 1 l . CI 9 . 7 1 0 . 4 I 11.1 8.8 9.4 9.8 159 8.1 8.5 5 . 51 7 . 7 8 . 1 1 2 s) 1 0 . 1 1 0 . EI 1 1 . 6 9s) 9.1 9.7 10.2 214 6 . CI 7.9 8.4 I 8.8 (F) 15 i 20 1 25 Mill Clg Ht 18.0 9.8 13.0 8.6 9.0 7.4 8.0 6.0 4.0 Diff tft) 112 139 114 in Clg HI 8.7 8.0 7.2 i 8.4 7.8 7.1 , .t:, _ Blow Temp 75 4.1 3.1 2.0 , %p Diff (F) 1 20 / 25 9.4 9.9 8.5 8.8 7.6 7.0 115 I:: 7.8 str B = 1 . 7 5, 2 2 ' / 1 O = . I 9 45" = .27 8.8 10.5 8.0 8.0 5 ._ 4 7.2 _ 8.4 7.8 7.1 - FPM str B = ,051 22'/Iz0 = ,061 45O'= .oa I Air hmn tity (cfml 3.5 30 2.5 1.8 Vane setting 750 500 FPM FPM S,r rJ = .Ol, 22x0 = .Ol -- str B = ,013, 22'Lz0 = .015 Str 8 = ,024 22'/1' = 028 45-o; .035 . 45O= .Ol 45" = ,019 1 21.010.3 --I- 9 . :3 9 . 1I 7 . 1I I L 21 2 . 0 9.7 10.3 8.28.6 12.2 13.1 10.411.2 8.6 9.2 13.0 7? 1 0 . 0 8 . 13 6.0 7 . ;7 15.0 11.0 7.0 1 0 . 5: 9 . :3 8 . 'I 13.314.2 11.2 12.1 9.1 9.7 15.0 11.0 7.0 1 0 . 61 9 . 24 8 . 'I 13.614.5 11.412.2 9.2 9.8 1 2 . c1 9 . c1 6 . ti 1 0 . 5 1 1 . :, 1 2 . 1 9 . 3 1 o . cI 1 0 . 5 2 6 9 8 . 1 8 . C, 9.0 1 6 . 0 i - L2 12.0 9.8 8.0 a . ,4_ : 17.a 13.0 9.0 8 . 6 9 . 2 /9 . 6 1 10.7 9.5 24: 8.5 1 3 s1 1 o . c1 6 . :5 1 0 . 7 1 1 . :i 1 2 . 3 9 . 5 1 0 . :t 1 0 . 7 3 2 4 8 . 2 a . 13 9.2 18.0 13.0 9.a 15.716.9 12.913.9 10.210.8 Straight 22%" 45O 11.0 9.7 30' 8.6 1 3 . cI 1 o . c1 6.5 1 1 . 0 1 1 . tI 9 . 6 1 0 . ~1 8.4 0.9 I 9.0 14.0 10.0 16.117.4 13.314.2 10.4 11.1 Straight 22x0 45" 11.1 9.8 8.7 1 3 . 0 '1I l . 21 2 . 0 1 2 . 4 10.0 9.810.6 '1 Il.1 6.5 8.5 9.1 9.5 - 19.0 14.0 10.0 - 16.618.0 13.514.6 10.611.3 2 7 2 10.2 9.0 8.1 161 368 K 12.6 10.9 406 9.4 490 14.515.5 12.012.9 9.710.3 1 1 6 . 0 19 . 3 FACTOR Max Cfm/Sq outlet Wall A Ft rea 29.0 19.0 14.0 9.6 OM ui tn lCeftWma/lSlqFAtr e a 8.7 5.7 4.2 2.9 15.216.3 12.513.5 10.010.6 TABLE 21-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.) For Beamed Ceilings OUTLET VELOCITY 2000 FPM jtr 6 = . 3 7 5 , 2 2 % " = . 4 2 A S ” = SAS STATIC PRESSURE STANDARD OUTLET S T A T I C PRti s s U R E WIT1 METERIt 4 G PLATE L Nom. Size Air o f Outlei V a n e cauanS e t t i n g tity (and Free Area) (cfm) 8x4 (16.9) Straight 22%" 45O 118 10 x 4 (21.7) Straight 22s" 45O 150 12 x 4 (24.6) Straight 22%" 45O 181 Straight 22%" 45O 244 Straight 22s" 45O 24 x 4 (55.0) Straight 22%" 45" 30 x 4 (68.3) Straight 22'h0 45O Str 8 = 1.36 Air D i f (F) B l0WTBm - Qucln. 2 5 (it) 15 2 0 _ tity Mi C l ? It 4t (cfm) 24 10.4 1.2 1 1 . 9 18 9.0 9.6 10.2 177 12 7.6 8.0 8.4 f: 26 10.6 19 9.1 13 7.7 Il.4 9.7 8.2 12.1 10.4 8.5 27 10.7. 11.6 20 9.3 9.9 14 7.8 8.2 2 8 11.0 1 1 . 9 21 9.5 1 0 . 1 14 7.8 8.4 29 11.2 22 9.7 15 7.9 466 237 299 60 13.0 45 10.4 30 8.0 14.2 15.4 11.3 12.0 8.5 8.9 362 62 13.3 47 10.6 31 8.2 14.5 15.7 11.6 12.3 8.7 9.2 65 13.7 4 9 11.0 33 8.5 15.0 16.2 12.012.8 9.0 9.5 67 14.0 50 11.3 34 8.6 15.316.6 12.3 13.1 9.29.7 68 14.2 51 11.4 34 a.7 15.516.8 12.413.3 9.3 9.9 70 14.4 53 11.6 35 8.9 15.717.0 12.613.5 9.5 10.0 272 44 12.3 33 10.2 22 8.0 13.4 14.4 1 0 . 9 11.6 8.5 8.9 12.7 10.7 8.6 366 46 12.6 35 10.4 23 8.2 13.8 14.8 Il.3 12.0 8.7 9.2 12.1 10.3 8.5 - 12.9 10.9 8.8 462 48 12.9 36 10.6 24 8.3 12.3 10.4 8.5 12.4 10.5 8.6 - 13.1 11.0 8.8 556 49 13.0 37 10.8 25 8.4 14.2 15.3 I 1 . 6 12.4 9.0 9.4 13.3 11.2 8.9 698 50 13.2 37 10.9 25 8.5 14.4 15.5 11.8 12.6 9.1 9.6 932 51 13.3 3 8 11.0 26 8.5 14.5 15.7 11.9 12.6 9.2 9.6 1116 206 36 11.6 2 7 10.8 18 8.6 Straight 22s" 45O 262 13.8 11.5 9.2 1 5 . 2, 12.6 , 9 . 5, 12 x 6 (41.6) Straight 22'h0 45O 16 x 6 (56.6) Straight 22Yi" 45O ‘-Ax6 (71.5) Straight 22%" 45O 24 x 6 (86.5) Straight 22%" 45O 14.9 12.3 9.6 310 I- M i n Clg Ht 5 8 1 2 . 7 (1 44 10.2 29 7.8 12.3 10.5 8.6 Straight 22%" 45O 16.217.5 13.114.0 10.0 10.6 15.917.2 12.813.6 9.6 10.1 412 82 16.2 62 13.0 41 9.9 17.919.3 14.215.3 10.611.3 16.5 13.5 392 10.4 66 16.1 50 13.2 3 3 10.1 17.9 19.5 14.4 15.5 10.8 11.6 S24 9 2 18.1 69 14.3 46 10.8 20.0 21.7 15.717.0 11.612.5 318 15.5 , 1 6 . 9 12.8 1 3 . 7 4 7 6 10.0 / 1 0 . 5 67 16.4 50 13.4 34 10.2 18.319.9 14.715.7 11.0 11.8 636 94 18.5 70 14.6 4 7 11.0 20.4 22.1 16.017.3 11.812.7 428 16.4 1 8 . 1 1 3 . 6, 1 4 . 6 10.5 1 1 . 1 - 642 72 17.6 54 14.3 36 10.8 19.7 21.3 15.6 16.9 11.6 12.5 856 102 20.0 77 15.6 51 11.6 22.123.9 17.118.7 12.513.5 19.0 15.3 11.5 806 77 18.6 58 14.9 39 11.2 20.822.4 16.417.8 12.113.0 1076 108 21.1 81 16.4 54 12.1 23.325.2 17.919.7 13.0 14.1 19.7 15.7 11.8 972 79 19.3 59 15.4 4 0 11.6 21.623.2 16.918.4 12.413.3 1296 111 21.9 83 17.0 56 12.4 24.2 26.1 18.520.4 13.414.5 82 20.0 62 15.9 41 11.7 22.424.0 17.419.0 12.613.7 1624 115 22.7 86 17.5 58 12.7 25.027.1 19.1 21.1 13.714.9 84 20.6 63 16.3 4 2 11.9 23.024.8 17.819.6 12.914.0 1960 119 23.4 89 18.0 60 13.0 25.828.0 19.621.7 11145..03 17.5 , 14.2 10.8 1 8 . CI 14.6 , 11.1 648 Straight 22x0 Straight 22%Q 45" (cfml 2.913.8 0 . 5 11.2 8.1 a.5 3.1 14.1 0.7 11.4 a.48.7 12.5 10.6 8.6 - 36 x 6 (131.3) :Ig H t 42 12.1 3 2 10.0 21 7.8 31 11.5 2 3 10.0 16 8.1 30 x 6 (109.0) Min 224 Straight 22'h0 45O 8x6 (26.5) Air auan. tity 18.7 20.4 15.c I 1 6 . 2 1 2 1 8 11.4 I 1 2 . 1 19.; I 15.4 1 I.6 , K FACTOR 980 Max Cfm/Sq F t o u t l e t W a l l Area 7.2 4.8 3.6 M i n Cfm/Sq F t Outlet Wall Area 2.2 1.4 1.1 i t NOTES: I. angular daflsction io a m a x i mum at eorh end. The 45” divetgsnca rignifias an angular deflection ot sach end of Ihs outlet of 45’. and ~imilorly for 22’h” divergenca. 3 . Underblow. I t i, n o t olwoy, nace,sory t o blow tha entire length of the roam unless there or., heat load sourc.,~ a t that end, squipmsnt load, open door,, wn-glars, *tc. Gmidering the toncsntralion o f room heat load on the basis of Btu/(hr) (q ft), the outl*t blow should cov*r 7536 of the hsat load. 4. Velocity is based on effective face area. 6. Measure ceiling height in the CLEAR only. Thb is the distance from the floor to the low*s1 c*iling beam or obstruction. 7. The Minimum Ceiling Height (table) is ths minimum csiling haighl which will givs proper operation of the wtl*t for the given outlet v&city. van* s*tii.%& tempsroture difference, blow, and cfm. The actual mea+ wed ceiling height must be equal t o or g,.mtc, t h a n t h e minimum ceiling height for the rsledion mad*. Preferably the top of an outlet should be not less t h a n twice the o u t l e t ’ s height below the minimum ceiling height. 8. Cfm/Sq Ft O u t l e t W a l l A r e a is the standard for judging total room air movement. The maximum value shown results in an air movement in the zone of oc. cupancy of about 50 fpm. It is orrumed that furniture. people, etc., obstruct IO% of the room cross-section. If the obrtructionr vary w i d e l y f r o m IO%, the values of the cfm/sq h outlet wall orso should be tempered accordingly. 9. For opplicationr requiring a limittng round level-the outlet velocity is limited by the sound gencr.tcd by the outlet. I’\K I ‘ 2-84 2. .\IK I~IS’1‘1~11117’1~10N TABLE 21-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.) For Beamed Ceilings OUTLET VELOCITY STATIC PRESSURE STANDARD OUTLET S T A T I C P R E S S U RWEI T b METERING PLATE Nom. S i z e ofOutlei (andFree Area) 250 FPM __-__. str B = .O', 22x0 = .Ol 45" = .O' 375 FPM 500 FPM 750 FPM I 45" = .0'9 S t r B = . O l . 2 2 % " = , 0 1 5 1S t r B = , 0 2 4 , 2 2 % ' = . 0 4 3 45"'= ,028 Diff ( F ) -C Valle Setting 5irpc 12 x 8 (56.7) Straight 22%" 45" 2.''3.0 0.7'1.2 9.2 9.8 16 x 8 (77.1) Straight 22%" 45" 3.114.0 1.411.9 9.7 10.1 Clg Ht I 81 ( ( Min Clg Ht T-; 1 1 36 27 '8 '5.4 12.6 10.0 17.0 18.3 13.7 '4.8 10.8 11.4 2 1 1 1 5.0 '6.1 1 2.7 13.7 463 1 0.4'0.9 40 30 20 16.9 '8.6 20.3 13.7 15.0 '6.2 '0.6 11.5 '2.2 Straight 22x0 45" 1 9 2 3.8'4.9 1 . 91 2 . 6 0.0 10.6 385 2 1 1 1 5.9 '7.2 1 3.414.5 I 0.811.5 57s 43 32 22 18.' 14.5 11.3 20.0 21.7 '6.0 17.3 12.1 13.0 24 x 8 (118.0) Straight 22x0 45" 231 14.3 15.6 12.3 '3.2 10.2 10.8 460 2 ! 5 . 01 5 . 2 1 9.012.8 1 3.010.4 1 6.6 18.1 1 3.9 15.1 1 1.0 '1.8 692 45 34 23 19.0 15.2 11.6 21.0 22.7 16.8 '8.2 12.5 13.4 30 x 8 (149.0) Straight 22x0 45" 209 14.9 '6.3 12.7 '3.7 10.5 11.1 580 :! 6 . 01 5 . 9 1i 9 . 01 3 . 4 1 3.510.7 1 7.319.0 I 4.515.7 1 1.4 12.2 868 46 35 23 20.0 22.0 23.9 16.0 '7.5 19.' '2.0 13.0 '4.0 36 x 8 (179.0) Straight 22s" 45O 18.0 '4.3 15.5 16.8 13.0 12.' 13.1 14.1 9.0 10.1 10.7 '1.4 702 :! 7 . 0' 6 . 5 1 i 8 . 01 9 . 7 :2 0 . 0 1 3 . 7 1 , 5 . 0 ' 6 . 2 1048 11 4 . 0 1 0 . 9 1 1I l2 . 67 1 16 x 10 (97.7) Straight 22x0 45O ' 8 . 0 1 4 . 1 I ' 5 . 5 1 6 . eI 13.0 12.1 '3.1 '4.' 9.0 10.1 '0.7 11.4 396 27.0 16.3 20.0 13.7 14.5 10.9 20 xa (97.6) '8.0 13.8 '3.0'I.8 9.0 9.9 9.6 '2.2 13.3 14.2 7 . ' i1 0 . 8 I1 1 . 7 ' 2 . 5 5.0 9.3 9.8 '0.4 297 1 0 . 5 /1 3 . 1 !1 4 . 1 8.0 11.4 '2.3 5.2 9.7 10.2 '5.2 '3.2 '0.8 374 '6.1 13.9 11.2 t t I 11 8 . 0 19.7 Il.7 12.6 11 5 . 0 1 6 . 2 595 16.5 18.C '3.8 15.C '1.1 11.5 497 19.3 21.2 15.9 17.3 12.2 13.3 746 450 '7.6 '9.1 14.6 15.5 11.5 12.: 600 20.6 22.5 '6.8 '8.3 12.6 '3.8 099 17.3 14.8 11.7 564 1 8 . 9 2 0 . :i 1 5 . 5 1 6 . 1P 1 2 . 0 1 2 . e1 751 22.1 '8.0 13.3 24.4 1 9 . 6 1126 14.6 . 17.7 15.1 11.8 680 1 9 . 5 2 1 . 1I 1 5 . 81 7 . : I ' 2 . ' 1 3 . c1 904 22.7 18.3 13.5 25.0 20.0 1355 14.9 16.1 13.9 11.2 367 1 7 . 6 1 9 .I‘ 1 4 . 6 1 5 . 1? 1 1 . 5 1 2 . :2 488 31.0'8.7 23.015.4 '6.011.9 Straight 22x0 45" '7.7 15.1 11.8 460 2 2 . 0 1 7 . ; 1 9 . 52 1 . ' 16.0 14.1 '5.8 17.: 11.011.1 '2.113.c 613 33.020.5 25.0'6.9 17.013.7 24 x 12 (181.0) Straight 221/2= 45O 18.8 16.0 '2.2 555 24.0 18.; 20.723.1 18.0IS.< 16.8 '8.: 12.0 11.1 '2.613.: 30x12 (228.0) Straight 22r/,O 45O 462 20.2 '6.9 12.7 695 36 x 12 (275.0) Straight 22x0 45" 560 21.3 '7.6 '3.1 036 2 0 x 10 (124.0) Straight 22x0 45O 2 4 x 10 (150.0) Straight 22%" 45O 3 0 x 10 (195.0) Straight 22x0 45O 3 6 x 10 (227.0) Straight 221/2O 450 16 x 12 (1'8.0) Straight 22vi" 45O MClX ctm,sq rt O u t l e tW a l l A r e a Min C f m / S qF t o u t l e tW a i l A r e a 249 364 244 -l-L '4.5 18.0 '9.8 11.0 15.0 '6.3 8.0 11.5 '2.2 19.015.; '4.0 12.5 25.02O.f '9.016., 22.224.t 17.719.2 13.2 14. 48 20.8 22.9 24.8 36 '6.4 '8.2 19.7 24 112.4/13.3/14.4 55 41 28 24.1 26.5 28.4 I1 8 . 8 I2 0 . 7 !2 2 . 4 13.7 14.8 16.1 925 23.6 25.: 18.5 20.: 13I . 6 14~ K F ACl'OR 29.0 19.0 14.0 9.6 8.7 5.7 4.2 2.9 . T A B L E 21-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.) For Beamed Ceilings OUTLET 1000 FPM V&OCITY 1500 FPM 2000 FPM STATIC PRESSURE STANDARD _- .-..STATIC PRES METERIN Nom. of Size Outlet (and Free OUTLET IRE WITH PLATE Van* Setting Area) --~-~-_-~ str 6 = 1 .33, 2 2 % ” = .33 Str B = ,715. 22%’ = .74 str i Air i cfm) Straight 12 x 8 (56.7) ~.- 22vi" 45O 16 x 8 (77.1) Straight 20 x a (97.6) Straight 1x8 .1&O) 22fi" 45O Straight 16 x 10 (97.7) Straight 20 x 10 (124.0) Straight 24 Straight 22s" 45O 22%O 45O 22%" 45O 221/2O 450 22'/i0 450 30x10 (195.0) Straight 36x10 (227.0) Straight 16 x 12 !118.0) Straight (150.0) 24 x 12 (181.0) I. 678 904 926 1232 2 9 2. Blow indicotsr dirtoncs from outl e t t o ths p o i n t w h e r e the air s t r e a m is rubstantiolly dirripotad. 1 jl1.6/12.7/13.51 3. Underblow. It i s not alwoyl nscer,ory lo b l o w the e n t i r e ,eng,h of ths room unless there ore heat load s o u r c e s ot t h a t end, e q u i p m e n t l o a d , o p e n doors, wn-gloss, etc. Considering ,he concentration of room hoot load on the barb of Btu/(hrl (sq f,), the outlet blow should cover 7S% of the heot load. 4. Velocity is foes area. I 1540 1 3 1 1 1 2 . 4 1 1 3 . 41 1 4 . 5 1 22'/i0 45O 22%" 45O 22x0 45" 33 12.9 14.0 15.1 I160 68 23.3 26.0 28.4 51 18.4 20.3 22.1 1736 34 13.3 14.5 15.7 2320 157 31.4 35.5 40.4 1 1 82 4 . 0 2 7 . 0 2 9 . 2 79 16.3 18.2 19.6 1404 71 24.3 27.2 29.5 53 19.0 21.2 22.8 2096 36 13.8 15.0 16.2 2808 164 32.0 37.3 42.2 1 2 32 4 . 8 2 8 . 1 3 0 . 2 82 16.8 18.8 20.2 792 7 12 3 . 9 2 7 . 2 2 9 . 5 53 19.0 21.2 22.8 1190 36 13.8 15.0 16.2 1584 164 32.3 37.3 42.2 123 24.8 28.1 30.2 82 16.8 18.8 20.2 994 75 26.0 29.6 32.0 56 20.6 22.7 24.6 1492 38 14.6 16.0 17.3 1988 30 x 12 1228.0) Straight 36 x 12 (275.0) Straight 22%" 45O 22'/i0 45O 35.3 40.2 46.6 130 27.0 30.3 32.7 87 17.9 20.3 21.7 1200 1798 1 I 2400 I 1502 86 30.2 34.6 37.4 65 23.6 26.3 28.2 43 16.3 18.2 19.3 2252 i t I 3004 147 31.1 35.4 38.0 98 20.1 22.8 24.5 87 31.3 36.0 38.5 65 24.2 27.0 29.0 44 16.6 18.5 19.7 2710 ) 1 3616 I 200 43.0 49.1 55.8 1808 5 1952 158 38.1 43.4 50.0 1 3 82 8 . 6 3 2 . 5 3 4 . 8 93 18.8 21.5 22.8 185 38.1 43.4 50.0 139 28.6 32.5 34.8 93 18.8 21.5 22.8 196 41.7 47.0 54.3 150 32.0 36.4 39.2 100 20.5 23.3 25.0 976 81 28.0 32.0 34.2 6 12 1 . 8 2 4 . 2 2 6 . 1 41 15.3 16.0 18.1 1466 1226 87 31.3 36.0 38.5 65 24.2 27.0 29.0 44 16.6 18.5 19.7 1836 1 1 I 2452 200 43.0 49.1 55.8 150 32.0 36.4 39.2 100 20.5 23.3 25.0 1480 93 33.6 38.6 42.0 70 25.8 28.9 31.0 47 17.4 19.5 20.8 2220 5 2 5 2960 213 47.0 53.2 59.2 1 6 03 4 . 2 3 9 . 2 4 2 . 1 107 121.7124.6126.5 1850 98 36.7 42.4 46.2 74 27.8 31.1 33.6 49 18.4 20.6 22.1 2776 Straight 22'h0 45O 174 80 28.0 32.0 34.2 60 21.8 24.2 26.1 40 15.3 16.9 18.1 Straight 22vi" 45O 1 a40 920 2230 I I 3 3700 I 1 -7 4 4460 2 103 39.4 45.1 49.8 77 29.4 33.0 35.7 3346 5 2 1 1 9 . 1( 2 1 . 0 1 2 3 . 0 1 K Divergent Blow hor verlicol ,ouvr., rtroight forword i n the ccnkr, with uniformly increasing clngular d e f l e c t i o n lo o moximu,,, a, each e n d . The 45’ divergence s i g n i f i e r o n ongulor deflection of each end of Ihs ode, o f 450, ond rimilorly for 22’b” divergence. 1 Straight 36 x a (179.0) 2 0 x 12 - ,375 1 22%" 45" Straight x 10 452 22%" 45O 30 x 8 (149.01 (150.0) = Air luan- tl tity ( Icfm) “on- lily B hosed on effective 5. S,o,lc P,errure is t h a t pressure r e q u i r e d lo p r o d u c e t h e indicored velocities o n d i s meos- I wad in inches of w&w. 6. Measure ceiling height in the CLEAR only. This b the dirtonce from the floor lo the lowest cciling beam or obstruction. 7. The Minimum Ceiling Height (table) is t h e m i n i m u m ceiling height which will give proper operation of the outlet for the given outlet velocity, vans retting, temperature difference, blow, and cfm. The cc+u.l meoswed c e i l i n g h e i g h t must b e equd t o o r g r e a t e r t h a n t h e minimum ceiling height for the selection mode. Preferably the top of an outlet should be not less than twice t h e o u t l e t ’ s height below the minimum ceiling height. 8. Cfm/Sq Ft Outlet Wall A r e a is the stondord for judging total room air movement. The maximum value shown results in on air movement in the zone of occupancy of about 50 fpm. It is orrumed that furniture, pcopls, etc.. obstruct 10% of the room cross-vxtion. If the obstructions vary widely from IO%, t h e values of the cfm/rq ft outlet wall oreo should be tempered accordingly. 9 . F o r oppticotionr requiring o limiting sound level--the outlet velocity is limited by the round generated by the outlet. FACTOR M a x Cfm/Sq F t Outlet Wall A r e a 7.2 4.8 3.6 M i n Cfm/Sq F t O u t l e t Wall A r e a 2.2 1.4 1.1 t PIPING DESIGN 3-l CHAPTER 1 m PIPING DESIGN-GENERAL GENERAL SYSTEM DESIGN Piping characteristics that are common to normal air conditioning, heating and refrigeration systems arc present4 in this chapter. The areas discussed include piping material, service limitations, cspansion, vibration, fittings, valves, and prcssurc losses. These areas are 01 prime consitlcration to the design cnginecr since they influence the piping lile, maintenance cost and first cost. The basic concepts ol fluid Ilow and design information on the more specialized fields such as high temperature water or low tcmpcrature refrigeration systems are not included; this information is avnil;“.‘p in other authoritative sources. MATERIALS Tile materials most com~~lo~lly systems arc the following: I. Steel - black and galvani~ctl used in piping 2. [\‘rought iron - black and galvanized 3. Copper - soft and hard TtrOle I illustrates the recon~n~entled materials [or the v a r i o u s scrviccs. Minimum standards, as shown, should be maintainctl. Table 2 contains the physical properties of steel pipe and Table 3 lists the physical properties 0E copper tubing. TABLE l-RECOMMENDED PIPE AND FITTING MATERIALS FOR VARIOUS SERVICES SERVICE Suction . REFRIGERANTS 12, 22, AND 500 Line l i q u i d Line Hot Gas Line CHILLED WATER Wrought copper, wrought brass or tinned cast brass Steel pipe, standard wall Lap welded or seamless for sizes larger than 2 in. IPS 150 lb welding or threaded malleable iron H&d copper tubing, Type L* Wrought copper, wrought brass or tinned cast brass 300 lb welding or threaded malleable iron Hard copper tubing, Type L* Wrought copper, wrought brass or tinned cast brass Steel pipe, standard wall Lap welded or seamless for sires larger than 2 in. IPS ~_ _~ .~__ .-~~ - --: : - -m $%&or gglvantzed steel pipe@ L-. --’ Galvanized OR WATER DRAIN OR CONDENSATE LINES STEAM OR CONDENSATE HOT WATER - - ’ Steel oiDe Extia’strong wall for sizes 1 ‘/2 in. IPS and smaller Standard wall for sizes larger than 1 ‘/a in. IPS Lap welded or seamless for sizes larger than 2 in. IPS “Hbrd copp;; CONDENSER MAKE-UP FITTINGS PIPE Hard copper tubing, Type L* tub;&& -j’ steel pipet 300 lb welding or threaded malleable iron .%elding,- galvanizer.4;., c&t, malleab!ao --“-&ckjron$ -4,. - r 9--.--^ _~ *Cast brass, wroughi &ape;pr y%wght brhsq Welding,gaIvanized;castormalleableironf Hard copper tubing+ Cast brass, wrought copper or wrought brass Galvanized Galvanized, drainage; cast or malleable ironj steel pipet Hard copper tubingi’ Cart brass, wrought copper or wrought brass Black steel pipet Welding or cast iron: Hard copper tubingt Cast brass, wrought copper or wrought brass Block steel pipe Welding or cast iron$ .Hard copper tubing+ ‘2 .’ Cast brass, wrought copper or wrought brass *Except for sizes l/4” and 3/g” OD where wall thicknesses of 0.30 and 0.32 in. are required. Soft copper refrigeration tubing may be used for sizes 1%” OD and smaller. Mechanical joints must not be used with soft copper tubing in sizes larger than %” OD. tNormolly standard wall steel pipe or Type M hard copper tubing is satisfactory for air conditioning applications. However, the piping material selected should be checked for the design temperature-pressure ratings. fNormally 125 lb cast iron and 150 lb malleable iron fittings are satisfactory for the usual air conditioning application. However, the fitting material selected should be checked for the design temperature-pressure ratings. 3-1. PAR'I- :\. I'II'II\:<; DESI(;N \ TABLE 2-PHYSICAL PROPERTIES OF STEEL PIPE INSIDE DIAM WEIGHT OF PIPE (Ib/ft) (in.) .269 .215 10 12 14 16 ,244 ,314 , ,364 ,302 ,088 .I19 ,424 ,535 .0451 .0310 ,141 ,141 .0955 .0794 .1041 .0716 ,493 .423 ,091 ,126 .567 ,738 .oa27 13609 ,177 ,177 .1295 .I106 .I910 .I405 ,622 .546 .109 .I47 .a50 1.087 .1316 .1013 ,220 .220 .1637 .1433 .3040 .2340 .a24 ,742 .113 ,154 1.130 1.473 .2301 .I875 ,275 ,275 .216a .194a .5330 .4330 1.049 .957 .I33 ,179 I.678 2.171 .3740 .3112 ,344 .344 .2740 .2520 .a640 .7190 1.380 1.278 .I40 .191 2.272 2.996 .6471 .5553 .434 ,434 .3620 .3356 1.495 1.283 1.610 1.500 ,145 .200 2.717 3.631 .8820 .7648 ,497 .497 .4213 3927 2.036 1.767 2.067 1.939 .154 ,218 3.652 5.022 1.452 1.279 ,622 ,622 s401 JO74 3.355 2.953 2.469 2.323 .203 .276 5.79 7.66 2.072 i .a34 .7.53 ,753 .6462 .6095 4.788 4.238 3.068 2.900 .216 .300 7.57 10.25 3.548 3.364 .226 .3ia 9.11 12.51 4.28 3.85 4.026 3.826 ,237 .337 10.79 14.98 5.51 4.98 5.047 4.813 .25a .375 14.62 20.78 6.065 5.761 .2ao .432 i a.97 28.57 '12.51 11.29 7.981 7.625 ,322 .500 28.55 43.39 21.6 19.8 2.24 2.26 WX) 80 10.750 10.750 10.750 10.020 9.750 9.564 ,365 .500 .593 40.48 54.70 64.33 34.1 32.4 31.1 2.81 2.81 2.81 30(S) 40 00 a0 12.750 12.750 12.750 12.750 12.090 ii.938 11.750 11.376 .330 ,406 ,500 .687 43.80 53.53 65.40 88.51 49.6 48.5 46.9 44.0 306) 40 (Xl 80 14.000 14.000 14.000 14.000 13.250 13.125 13.000 12.500 .375 .43a ,500 .750 54.60 63.37 72.10 106.31 30(S) 40(X) 80 16.000 16.000 16.000 15.250 15.000 14.314 ,375 ,500 .a43 ia . 0 0 0 17.250 17.000 lb.874 16.126 .375 .500 .562 ,937 40(S) (5) 19.250 19.000 la.814 17.938 23.250 23.000 22.626 21.564 *TO change "Wt of Water in Pipe (lb/f1 .375 ,500 ,687 1.218 ) I 1.178 1.178 1.735 1.735 2.006 45.6 3.34 3.34 3.34 3.34 3.17 3.13 3.08 2.98 115.0 111.9 108.0 101.6 59.8 58.5 55.8 51.2 3.67 3.67 3.67 3.67 3.46 3.44 3.40 3.27 138.0 135.3 133.0 122.7 62.40 82.77 136.46 79.1 76.5 69.7 4.18 4.1 a 4.18 3.99 3.93 3.75 183.0 176.7 160.9 70.60 93.50 104.75 170.75 100.8 98.3 97.2 88.5 4.71 4.71 4.71 4.71 4.52 4.45 4.42 4.22 234.0 227.0 224.0 204.2 78.60 104.20 122.91 208.87 126.7 122.5 120.4 109.4 5.24 5.24 5.24 5.24 94.60 125.50 171.17 296.36 184.6 179.0 6.28 6.28 5.65 365.2 tz 1 6 :: t o “ G a l l o n s o f W a t e r i n P i p e ( g o l / f t t r , "d k i d e d u e ti n t a b l e b y 8 . 3 4 . t Si s d e s i g n a t i o n o f s t a n d a r d w a l l p i p e . X i s d e s i g n a t i o n o f e x t r as t r o n g w a l lp i p e . c ' , TABLE 3-PHYSICAL PROPERTIES OF COPPER TUBING HARD %% Yi I--- 0 UTSIDE IDIAM STUBBS GAGE (in.) OUTSIDE SURFACE (sq fl/fl) ,106 ,144 ,203 .036 ,069 .I IO ,098 .I31 ,164 1% 21 20 19 ,032 .035 .042 .a1 1 1.055 1.291 ,516 .a74 1.309 710 600 590 ,328 ,464 ,681 ,224 ,379 .566 .229 .295 ,360 I % 2 2% 18 17 16 .049 .058 ,065 1.527 2.009 2.495 1.831 3.17 4.89 580 520 470 .94 1.46 2.03 ,793 1.372 2.120 ,425 ,556 ,687 15 14 13 ,072 .oa3 ,095 2.981 3.459 3.935 6.98 9.40 12.16 440 430 430 2.68 3.58 4.66 3.020 4.060 5.262 ,818 ,949 1.08 12 ,109 ,122 ,170 4.907 5.881 7.785 18.91 27.16 47.6 400 375 375 6.66 8.91 16.46 8.180 11.750 20.60 1.34 1.60 2.13 19 ,035 .040 ,045 ,430 ,545 ,785 .I46 .233 .484 1000 1000 1000 ,198 ,284 ,454 ,063 ,101 .209 ,131 ,164 .229 .050 ,055 ,060 1.025 1.265 1.505 .a25 1.256 1.78 880 780 720 ,653 ,882 1.14 .35a ,554 ,770 .295 ,360 .425 1 Govt.Type “K” 2 5 0 Lb Working PreSSWe WT OF WATER INTUBE* (lb/W 1000 1000 890 =/4 I-; vi I WEIGHT OF TUBE (Ib/ft) .083 ,159 ,254 A%% SOFT MINIMUM TEST PRESSURE (psi) ,325 .450 ,569 Govt.Type "L" 2 5 0 Lb Working Pl.ZSSW.2 t (in.) TRANSVERSE AREA (rq i n . ) .025 .025 .028 1 G o v t . Type “K” 4 0 0 Lb Working PreSSWe INSIDE DIAM 23 23 22 % Gavt.Type "M" 250 Lb Working WALL THICKNESS (in.) 1% 2 - .070 1.985 3.094 640 1.75 1.338 .556 .,008900 ,100 2.465 2.945 3.425 4.77 6.812 9.213 580 550 530 2 . 43 8. 3 3 4.29 2.070 2.975 4.000 .687 ,818 ,949 .l10 .I25 .140 3.905 4.875 5.845 510 460 430 5.38 7.61 10.20 21 ia I8 .032 .049 .049 .311 .402 ,527 .076 .127 ,218 1000 1000 1000 .133 ,269 ,344 ,033 ,055 .094 .098 .131 ,164 16 16 16 .065 .065 .065 .745 .995 1.245 .436 .77a 1.217 1000 780 630 .641 .a39 1.04 .189 ,336 .526 .229 .295 ,360 15 14 13 .072 .oa3 .095 1.481 1.959 2.435 1.722 3.014 4.656 580 510 470 1.36 2.06 2.92 .745 1.300 2.015 ,425 .556 .687 12 11 10 ,109 .120 ,134 2.907 3.385 3.857 6.637 a.999 il.68 450 430 420 4.00 5.12 6.51 2.870 3.890 5.05 8.18 ,949 1.08 ,160 .I92 4.805 5.741 18.13 25.88 400 400 9.67 13.87 11.97 18.67 26.83 5.180 8.090 11.610 7.80 11.20 1.08 1.34 1.60 1.34 1.60 % f/ % 21 18 18 .032 .049 .049 .311 .402 .527 .076 .I27 .218 1000 1000 1000 ,133 ,269 ,344 .033 .055 ,094 .098 .131 .I64 % - 16 16 16 .065 .065 .065 .745 ,995 1.245 ,436 ,778 1.217 1000 780 630 ,641 .a39 1.04 ,189 .336 .526 ,229 ,295 .360 15 14 13 .072 ,083 .095 1.481 1.959 2.435 1.722 3.014 4.656 580 510 470 1.36 2.06 2.92 .745 1.300 2.015 ,425 ,556 ,687 12 11 10 ,109 .120 ,134 2.907 3.385 3.857 6.637 a.999 11.68 450 430 420 4.00 5.12 6.51 2.870 3.89 5.05 .818 ,949 1.08 .I60 .I92 4.805 5.741 18.13 25.88 400 400 9.67 13.87 * T o c h a n g e " W t o f w a t e r i n T u b e ( I b / f t ) " t o " G a l l o n s o f W a t e r i n T u b e ( g a i / f t ) , " d i v i d e v & e s i n t a b l eb y a . 3 4 . 7.80 11.2 1.34 1.60 , PAR.l- 3-4 SERVICE LIMITATIONS 3. PII’Ih’(; I)I<SI(;N TABLE J-THERMAL LINEAR EXPANSION OF COPPER TUBING AND STEEL PIPE The sal’c working pressure md teml~eratuw loI steel pipe and copper tubing, including fittings, are limited by the ASA codes. Check these codes when thcrc is tiorrbt about the ability of pipe, tubing, or fittings to withstand pressures and temperatures in a given inst;~ll;ttion. In many instances cost can be reduced and over-design ciiminatctl. For example, if the working pressure is to bc 175 psi at 250 I;, a 150 psi, class A, carbon stcci flange can bc safely used since it can withstand a pressure of 225 psi at 250 I;. If the code is not checked, a 300 psi flange must be spccifieti because the 175 psi working pressure exceeds the 150 psi rating of the 150 psi flange. The safe working pressure and temperature for copper tubing is dependent on the strength of the fittings and tube, the composition of the solder used for making a joint, and on the temperature of the uid conveyed. Table 4 indicates recommended servIce limits for copper tubing. (Inches per 100 feet) TEMP RANGE COPPER (F) TUBING STEEL PIPE 0 50 100 0 .56 1.12 0 .37 .76 150 200 250 1.69 2.27 2.85 1.15 1.55 1.96 300 350 400 3.45 4.05 4.45 2.38 2.81 3.25 450 500 5.27 5.89 3.70 4.15 NOTE: Above data ore based on expansion from O’F but are rufficiently accurate for all other temperature ranges. . Chat -3 gives the sizes of offsets in steel piping for travels up to 3 inches. Expansion loop sizes may be reduced by cold-springing them into place. The pipe lines are cut short at about 50:/, of the expansion travel and the expansion loop is then sprung into place. Thus, the loops are subject to only one-half the stress when expanded or contracted. 2. Expansion joints - There are two types available, the slip type and the bellows type. The slip type expansion ,joint has several ciisadvantages: (a) It requires packing and lubrication, which dictates that it be placed in an accessible location: (b) Guides must be installed in the lines to prevent the pipes from bending and binding in the joint. Bellows type expansion joints are very satisfactory for short travels, but must be guided or EXPANSION OF PIPING All pipe lines which are subject to changes in temperature expand and contract. Where temperature changes are anticipated, piping members capable of absorbing the resultant movement must be included in the design. Table 5 gives the thermal linear expansion of copper tubing and steel pipe. There are three methods commonly used to absorb pipe expansion and contraction: 1. Expansion hops md 0Jfset.s - Table 6, page 6, shows the copper expansion loop and offset sizes required for expansion travels up to six inches, Chn~t I shows the sizes of expansion loops made of steel pipe and welding ells for expansion travels up to 10 inches. TABLE 4-RECOMMENDED MAXIMUM SERVICE PRESSURE FOR VARIOUS SOLDER JOINTS MAXIMUM SOLDER USED IN JOINTS 50-50 Tin-lead 95-5 Tin-Antimony 95-5 Tin-lead or Solders Melting At or Above 1100 F SERVICE VI SERVICE PRESSURE (PSI) Water TEMP 100 150 200 250 vi”to Incl. 200 150 100 85 100 150 200 250 500 400 300 200 350 270 1%” I l%“to2%” IWI. 175 125 90 75 Steam I 2vsfl to 4%” hl. All 150 100 75 50 - 400 350 250 175 300 275 200 150 - 190 155 120 E x t r a c t e d f r o m A m e r i c a n Standard W r o u g h t - C o p p e r a n d W r o u g h t - B r o n z e S o l d e r - J o i n t F i t t i n g s , (ASA 816.22-l 95 1 lisher, The American Society of Mechanical Engineers, 29 West 39th Street, New York 18, New York. 15 15 1, w i t h t h e p e r m i s s i o n o f t h e p u b - C.HAP-I-ER I, PIPISC; I)l-SI(;S - (;K:.Nf?RAI. CHART 1 -STEEL EXPANSION LOOPS AMOUNT OF WE V-0” I a a Q Ib 12 ,k 116 16 20 22 2l4 26 26 30 32 34 38 LENGTH OF “H” Data from Ric-Wil Co. CHART 2 -STEEL EXPANSION OFFSETS LEG IN FEET 14 16 I6 20 22 24 26 28 30 32 34 36 38 40 LEG IN FEET Data from Pittsburgh Pipe Coil L Bending Co. TABLE 6-COPPER EXPANSION LOOPS AND OFFSETS EXPANSION LOOP OFFSE? 01‘ s;~tltllcs slioultl be Llsetl. Tllc 1)i1)c s u p p o r t s illlist Ii;kve a smooth, Ilat bearing surface, free from I)urrs (jr otlicr sliarp projections which would wear or cut tlic pipe. 7‘11~ controlling factor iI1 the spacing 01 supl)orts LOI- horizontal pil)c lines is the ticllcction of piping due to its own weight, weight of the. Iluiti, pi1)ing accessories, and the insulation. Ttrble 7 lists tlic reconlnlcntlccl support spacing for Sclictiulc -IO ljipe, using tlic !istctl contiitioiis with watcl- as a fluid. The support spacing for copper tubing is given in TnDle 8 which inclutlcs the weight of the tubing fillctl with water. Data from Mueller Brass Co. itI home other way restrained to prevent Callapse. 3. I;lexible ttzetd mtd rubber I~ose - F l e x i b l e hose, to absorb expansion, is rcconmencied for smaller size pipe or tubing only. It is not recommended for larger size pipe since an excessive length is required. Where flexible hose is used to absorb expansion, it should be installed at right angles to the motion of the pipe. ?‘hc devices listed above arc not always necessary to absorb expansion and contraction of piping. In fact they can be omitted in the great majority of piping systems by taking advantage of the changes in direction normally required in the layout. Consider, for example, a heat exchanger unit and a primp located 50 ft. apart. Sufficient flexibility is normally assured by running the piping up to the ceiling at the pump and back down at the heat exchanger, provided the piping is merely hung lronl hangers and anchored only at the ends where it is attachcti to the pump and the heat es&anger. PIPING SUPPORTS AND ANCHORS ,111 piping shot~lti bc siipportcd with hangers that cali withstand the combined weight of pipe, pipe fittings, valves, iluitl in the pipe, and the insulation. They must also IX capable of keeping the pipe in proper alignment when necessary. IVlicre extreme exlxinsion or contraction exists, roll-type hangers Tables 7 clnd 8 arc for “ticad level” piping. \Vatcl and refrigerant lines arc normally run level; steam lines are pitched. Water lines are pitched when the. line must be drained. For pitched steam pipes, rcfcr t o Table 25, page 82, for support spacing when Schedule 40 pipe is used. Unless pipe lines are atlequatcly anti properly anchored, expansion may put excessive strain on fittings and equipment. Anchors arc located according to jolt conditions. For instance, on a tall builcling, i.e. 20 stories, the risers could he anchored on the 5 t h Hoor and on the 15th Hoor with an expansion device located at the 10th Hoor. This arrangement allows the riser to expand in both directions from the 5th and 15th floor, resulting in less pipe travel at headers, whether they are located at the top or bottom of the building or in both locations. TABLE 7-RECOMMENDED SUPPORT FOR SCHEDULE 40 PIPE SPACING NOMINAL PIPE SIZE DISTANCE BETWEEN SUPPORTS (in.) (ft) ‘! - 1 % 8 1 vi - 2 % 3 - 3% IO 12 4 14 _ 6 - 12 - 24 a 14 TABLE 16 20 8-RECOMMENDED SUPPORT FOR COPPER TUBING TUBE OD (in.) ‘/I vi - I % 1 % - 2 ‘/r 2% - 5% 6% - 0 % SPACING DISTANCE BETWEEN SUPPORTS (ft) 6 a 10 1 2 14 ’ CHAPTER 011 smaller I~uildillgs, i . e . 5 s t o r i e s , riseI% arc ;I~j&(jre~l but O I I C C . Usually this is done near t h e Ileader, allowing the riser to grow in one direction only, either up or down depending on the header location. eI‘ll~ main point to consider when applying pipe support anchors and expansion joints is that cxpansion takes place o n a temperature change. The greater the temperature change, the greater the expansion. The supports, anchors and guides are applied to restrain the expansion in a desired direction so that trouble does not develop because oE negligent design or installation. For example, if a takcoff connection from risers or heaclers is located close to Iloors, beams or columns as shown in Fig. a change in temperature may cause a break in the take 4 with subsequent loss of fluid and Hooding da. Je. In. this figure trouble develops when the riser expands greater than dimension “X.” Proper consideration of these items is a must when designing piping systems. VIBRATION 3-7 1. P I P I N G l>ESlGf% -GENERAL ISOLATION .OF PIPING SYSTEMS The undesirable effects caused by vibration of the piping are: 1. Physical damage to the piping, which results in the rupture of joints. For refrigerant piping, loss of refrigerant charge results. 2. Transmission of noise thru the piping itself or thru the building construction where piping comes into direct contact. It is always difficult to anticipate trouble resulting from vibration of the piping system. For this reason, recommendations made toward minimizing the effects of vibration are divided into two categories: ’ design consideration - These involve design i>recautions that can prevent trouble effeclively. FIG. 1 - T AKE -OFF Too CLOSE To FLOOK 2. liemedjes or repai7.s - ‘i‘hese are necessary where precautions arc not taken initially or, in a minority of cases, where the precautions prove to bc insuficicnt. Design Considerations for Vibration Isolation I. In all piping systems vibration has an originating source. This source is usually a moving component such as a water pump or a compressor. When designing to eliminate vibration, the method of supporting these moving components is the prime consideration. For example: a. The weight of the mass supporting the components should be heavy enough to minimize the intensity of the vibrations transmitted to the piping and to the surrounding structure. The heavier the stabilizing mass, the smaller the intensity of the vibration. b. Vibration isolators can also be used to minimize the intensity of vibration. c. A combination of both methods may be required. 2. Piping must be laid out so that none of the lines are subject to the push-pull action resulting from vibration. Push-pull vibration is best dampened in torsion or bending action. 3. The piping must be supported securely at the proper places. The supports should have a relatively wide bearing surface to avoid a swivel action and to prevent cutting action on the pipe. The support closest to the source of vibration should be an isolation hanger and the succeeding hangers should have isolation sheaths as illustrated in Fig. -7, p a g e 8. Non-isolated hangers (straps or rods attached directly to the pipe) should not be used on piping systems with machinery having moving parts. 4. The piping must not touch any part of the building when passing thru walls, Boors, or furring. Sleeves which contain isolating material should be used wherever this is anticipated. Isolation hangers should be used to suspend the piping from walls and ceilings to prevent transmission of vibration to the building. Isolation hangers are also used where access to piping is difficult after installation. 5. Flexible hose is often of value in absorbing vibration on smaller sizes of pipe. To be effective, these flexible connectors are installed at right angles to the direction of the vibration. PART 3-8 3. PIPIN(; I)ESI(;K c DRAW UP SNUG IL ,\ weight may IJC xtldecl t o tile pip bcfow the first fixed support as illustrated in Fig. 3. .l‘his weight adds mass to the pipe, reducing vibration. c. Ol,posing isolation hangers may be added. FITTINGS METAL SLEEVE Where the vibration is not limited to one plane or direction, two flexible connectors arc used and installed at right angles to each other. The flexible hose must not restrain the vibrating machine to which it is attached. At the opposite end of the hose or pair of hoses, a rigid but isolated anchor is secured to the connecting pipe to minimize vibration. Generally, Hexible hose is not recommended in systems subject to pressure conditions. Under pressure they become stiff and transmit vibration in the same manner as a straight length of pipe. Flexible hose is not particularly efficient in absorbing vibration on larger sizes of pipe. EfFiciency i s i m p a i r e d s i n c e t h e length-todiameter ratio must be relatively large to obtain full flexibility from the hose. In practice the length which can be used is often limited, thus reducing its Hexibility. Elbows are responsible for a large percentage of the pressure drop in the piping system. With equal velocities the magnitude of this pressure drop depends upon the sharpness of the turn. Long radius rather than sflort radius elbows are recommended wherever possible. When laying out offsets, 45” ells are recommended over 90” ells wherever possible. See Fig. 4. Tees should be installed to prevent “bullheading” as illustrated in Fig. 5. “Bullheading” causes tur- . bulence which adds greatly to the pressure drop and may also introduce hammering in the line. If more than one tee is installed in the line, a straight length of 10 pipe diameters between tees is recommended. This is done to reduce unnecessary turbulence. To facilitate erection and servicing, unions and Hanges are included in the piping system. They are installed where equipment and piping accessories must be removed for servicing. The .various methods of joining fittings to the piping are described on pnge 12. GENERAL PURPOSE VALVES An important consideration in the design of the piping system is the selection of valves that give proper performance, long life and low maintenance. Remedies or Repairs for Vibration Isolation After Installation 1. Relocation OF the piping supports by trial and error tends to dampen extraordinary pipe vibration. This relocation allows the piping to take up the vibration in bending and helps to correct any vibrations which cause mechanical resonance. 2. If relocation of the pipe supports does not eliminate the noise problem caused by vibration, there are several possible recommendations: a. The pipe may be isolated from the support by nleans of cork, hair felt, or pipe insulation as shown in I;;#. 2. I/////////// -FIXED HANGER WEIGHT ADDED CY PUMP FIG. 5 - WEIGHT Ameo TO DAMWN VIBRATION C:t-IAP-I’I:R I. PIPIN<; DI:SI(;N - 3-9 (;ENERAL. -T- -T- - 4i I NO. I “BULLHEADING”- RECOMMENDED NO. 2 PREFERABLE TO NO. I ACCEPTABLE *l-lit design, construction and niaterial of the valve tlcternlines whether or not it is suited for the particalar application. Table 9 is for quick reference in selecting a valve for ;I particular application. There TABLE 9-GENERAL NO.4 PREFERABLE TO NO. 3 PURPOSE I VALVES STEAM REFRIGERANT* Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory (Low Press) Satisfactory Satisfactory Satisfactory (High Press) Satisfactory Not Recommended Not Recommended Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory (non-corrosive brines) Satisfactory Satisfactory Satisfactory Not Recommended Satisfactory Satisfactory Not Recommended Satisfactory Not Recommended Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory (non-corrosive brines) Satisfactory Satisfactory Satisfactory (Low Temp) Satisfactory (Low Temp) Satisfactory Satisfactory Not Recommended R e c o m m e n d e d Recommended Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory Recommended Not Recommended Not Not Not Not G l o b e , A n g l e , “Y” V a l v e I. P l u g d i s c 2. Conventional (Narrow-seat) 3. N e e d l e v a l v e 4. Composition disc Satisfactory Satisfactory Satisfactory Satisfactory Satisfactory Not Recommended Satisfactory Satisfactory (Low Press) Satisfactory Satisfactory Satisfactory Satisfactory P l u g C o c k Valve Satisfactory Satisfactory N o t R e c o m m e n d e d l’i B. V a l v e s t e m , o p e r a t i o n 1. Rising stem, outside screw 2. Rising stem, inside screw 3. Non-rising stem, inside screw 4. C. Sliding stem Valve connections 1. Screwed 2. Welded 3. Brazed 4. Soldered 5. Flared 6. Flanged to Refrigerants 12, 22 and ...i .i ’ pipe DISC C O N S T R U C T I O N Gate V a l v e 1. Solid wedge 2. Split wedge 3. Flexible wedge 4 . D o u b l e d i s c , p a r a l l e l seat “For - 47t are six basic valves which are comnlonly used in piping systems. These are gate, globe, check, angle, “Y” and plug cock. WATER VALVE CONSTRUCTION A. Bonnet and body connections 1. Threaded 2. U n i o n 3. Bolted 4. Welded 5. Pressure-seal DO NOT USE t NO.3 PERMISSIBLE 500 only. Recommended Recommended Recommended Recommended Each valve has a definite purpose in the control of fluids in the system. ]~efore tliscussing the various valve types, the construction dctails that arc similar in all of the valves arc presented. These construction details are made avaiIabIe to familiarize the engineer with the various aspects of valve sclcction. HANDWHEEL (RISES WITH STEMI ,/-RISING (,NSlDE STEM SCREW) PACKING NUT WITH GLAND VALVE CONSTRUCTION DETAILS XING BONNE Bonnet and Body Connections ‘1%~ bonnet and body connections are normally made in five different designs, namely threaded, union, bolted, welded and pressure-seal. Each dcsign has its own particular use and advantage. 1. Threaded bonnets arc recommended for low pressure service. They should not be used in a piping system where there may be frequent dismantling and reassembly of the valve, or where vibration, shock, and other severe conditions may strain and distort the valve body. Thrcadcd bonnets are economical and very compact. Fig. 6 illustrates a threaded or scrcwcd-in bonnet and body connection in an angle valve. 2. U?rior~ I)ot~~~et and body construction is illustrated in I;ig. 7. This type of bonnet is normally not made in sizes above 2 in. because it would require an extremely large wrench to dismantle. A union bonnet connection makes a sturdy, tight joint and is easily dismantled and reassembled. ’ ’ RISING STEM (INSIDE SCREW) - a PLUG TYPE DISC SCREWED ENDS . FlC. ,i - GLCIUE VALVE 3. Bolted bonnets are used on practically all large size valves; they are also available for small sizes. This type of joint is readily taken apart or reassembled. The bolted bonnet is practical for high working pressure and is of rugged, sturdy construction. Fig. 8 is a gate valve illustrating a typical bolted bonnet a?d body valve construction. 4. Welded bonnets are used on small size steel valves only, and then usually for high pressure, high temperature steam service (Fig. 9). Welded bonnet construction is difficult to dismantle x HANDWHEEL ISE (RISES WITH STEM1 - PACKING NUT WITH OUT GLAND GLAND - SCREWED T H R E A D E D BONNET7 sot ID WEDGE D I S C . ‘BOLTED BONNET SCREWED ENDS NARROW SEAT DISC FLOW FLOW FIG. 6 -ANGLE VALVE FIG.~-GGATEVALVE(KISINCSTEM) Figures 6.10, courtesy of Crane Co. ) <:HAI’?‘EK I. PII’INC; I)lCSl(;N - (;ENERAL 3-U Valve Stem Operation In most applications the type ol stem operation does not affect fluid control. However, stem construction may be important where the need for in. dication of valve position is required or where head room is critical. There arc four types of stem construction: rising stem with outside screw; rising stem with inside screw; non-rising stem with inside screw; and sliding stem (quick opening). BOLTED GLAND - -WELDED BONNET FIG. 9 - W ELDED BONNET CONSTRUCTION and reassemble. For this reason these valves are not available in larger sizes. _ 5. Aessure-seal bonnets are for high temperature steam. Fig. 10 illustrates a pressure-seal bonnet and body construction used on a gate valve. Internal pressure keeps the bonnet joint tight. This type of bonnet construction simplifies “making” and “breaking” the bonnet joint in large high pressure valves. RISING STEM (OUTSIDE SCREW) HEEL ES NOT RISE WrrH STEM) ‘f FLEXIBLE WEDGE DISC 7 FIG . 10 - FLEXIBLE BOLTED GLAND WELDING W EDGE DISC (GATE V ALVE) 1. Rising stem with outside so‘eru is shown in Fig. 8. The gate valve illustrated in this figure has the stem threads outside of the valve body in both the open and closed position. Stem threads are, therefore, not subject to corrosion, erosion, sediment, and galling from extreme temperature changes caused by elements in the line fluid flow. However, since the valve stem is outside the valve body, it is subject to damage when the valve is open. This type of stem is well suited to steam and high temperature, high pressure water service. A rising stem requires more headroom than a non-rising stem. The position of the stem indicates the position ( of the valve disc. The stem can be easily lubricated since it is outside the valve body. 2. Rising stem with inside screw is probably the most common type found in the smaller size valves. This type of stem is illustrated in an angle valve in Fig. 6, and in a globe valve in Fig. 7. The stem turns and rises on threads inside the valve body. The position of the stem also indicates position of the valve disc. The stem extends beyond the bonnet when the valve is open and, therefore, requires more headroom. In addition it is subject to damage. 3. Non-rising stem with inside screw is generally used on gate valves. It is undesirable for use with fluids that may corrode or erode the threads since the stem is in the path of flow. Fig. 11 shows a gate valve that has a non-rising stem with the threads inside the valve body. The non-rising stem feature makes the valve ideally suited to applications where headroom is limited. Also, the stem cannot be easily damaged. The valve disc position is not indicated with this stem. 4. Sliding stern (quick opening) is useful where quick opening and closing is desirable. A lever and sliding stem is used which is suitable for both manual or power operation as illustrated in Fig. 12. The handwheel and stem threads are eliminated. 3-12 L ,-NON-RISING BOLTED S BONNET- FLOW d /~,,EwED ENDS FIG. ll- GATEVALVE(NON-KISINGSTEM) Pipe Ends and Vahe Connections It is important to specify the proper end connection for valves and fittings. There are six standard methods of joints available. These are screwed, welded, brazed, soldered, flared, and flanged ends, and are described in the following: 1. ends are widely used and are suited for all pressures. To remove screwed end valves SLIDING STEM CLAMP TYPE BOLTED BONNET FLANGED .z / E N D S A FIG. If! - SL.II)IM: hxar GATE V ALVE . PAR.I- 3. I’IPINC; I)ESI(;N and fittings Irom the line, extra fittings (unions) are required to avoid dismantling a consitlcrable portion of the piping. Screwed end connections arc normally confined to s~allcr pipe silts since it is more difficult to make up the screwed joint on large pipe sizes. I;ig. 7 is a globe valve with screwed ends that connect to pipe or other fittings. 2. Welded ends are available for steel pipe, fittings, and valves. They arc used widely for all fitting connections, but for valves they are used mainly for high temperature and high pressure services. They are also used where a tight, leak-proof connection is required over a long period of time. The welded ends are available in two designs, butt weld or socket weld. Butt weld valves and littings come in all sizes; socket weld ends are usually limited to the smaller size valves and fittings. Fig. 10 illustrates a gate valve with ends suitable for welding. 3. Bl.ated ends are designed for brazing alloys. This type of joint is similar to the solder joint but can withstand higher temperature service because of the higher melting point of the brazing alloy. Brazing joints are use< principally on brass valves and fittings. 4. Soldered ends for valve and fitting are restricted to copper pipe and also for low pressure service. The use of this type of joint for high temperature service is limited decause of the low melting point of the solder. 5. Flared end connections for valves and fittings are commonly used on metal and plastic tubing. This type of connection is limited to pipe sizes LID to 2 in. Flared connections have the advantage of being easily removed from the piping system at any time. 6. Flanged ends are higher in first cost than any of the other end connections. The installation cost is also greater because companion flanges, gaskets, nuts and bolts must be provided and installed. Flanged end connections, although made in small sizes, are generally used in larger size piping because they are easy to assemble and take down. It is v&y important to have matching flange facing for valves and fittings. Some of the common Range facings arc plain face, raised face, male and female joint, tongue and groove joint, and a ring joint. Flange facings should never be mixed in.making up a joint. Fig. 8 illustrates a gate valve with a flanged encl. l Cf-fAf’~I‘ER I . PIf’IK(; I>ESI<;N 3-13 - (;I~NER.Al. GATE VALVES A gate valve is intcndcd for use iis a stop valve. It gives the best service when used in the fully open or closed position. f~i~7l7.e~ 8 (l?ld IO 111~21 14 arc typical gate valves con~n~only used in piping practice. An important feature ol the gate valve is that there is less obstruction and turbulence within the valve and, therefore, a correspondingly lower pressure drop than other valves. With the valve wide open, the wedge or disc is lifted entirely out of the fluid stream, thus providing a straight flow area thru the valve. I’ RISING STEM 2 ( I N S I D E SCREW) -- HANDWHEEL -2l IRISES WITH STEM) ----PACKING NUT WITH GLAND SCREWED BONNET SPLIT WEDGE DISC - Disc Construction Gate valves should not be used for throttling flow except in an emergency. They are not designed lo1 this type of service and consequently it is difficult to control flow with any degree of accuracy. Vibra1 and chattering of the disc occurs when the valve 1” Jsed for throttling. This results in damage to the seating surface. Also, the lower edge of the disc may be subject to severe wire drawing effects. The wedge or disc in the gate valve is available in several forms: solid wedge, split wedge, Hexible wedge, and double disc parallel seat. These are described in the following: Solid wedge disc is the most common type. It has a strong, simple design and only one part. This type of disc is shown in Figs. 8 and 11. It can be installed in any position without danger of jamming or misalignment of parts. It is satisfactory for all types of service except where the possi&lity of extreme temperature changes exist. Under this condition it is subject to sticking. Split wedge disc is designed to prevent sticking, but it is subject to undesirable vibration intensity. Fig. I3 is a typical illustration of this type of disc. Flexible wedge disc construction is illustrated in Fig. 10. This type of disc is primarily used for high temperature, high pressure applications and where extreme temperature changes are likely to occur. It is solid in the center portion ancl flexible around the outer edge. This design helps to eliminate sticking and permits the‘disc to open easily under all conditions. LkmDle disc parallel seat (Fig. 14) has an internal wedge between parallel discs. Wedge action damage at the seats is minimized and transferred to the internal wedge where reasonable wear does not prevent tight closure. FIG. 13 -SPLIT Wmc;~ DISC (GATE VALVE) The parallel sliding motion of the discs tends to clean the seating surfaces and prevents foreign material from being wedged between : disc and seat. Since the discs are loosely supported except when wedged closed, this design is subject to vibration of the disc assembly parts when partially open. (DOES NOT RISE WITH STEM) BOLTED GLAND RISING STEM (OUTSIDE SCREW BOLTED BONNET -, PARALLEL DISCINTERNAL WEDGEFLANGED ENDS FIG. 14 --DOUBLE DISC PARALLELSEAT (GATEVALVE) Figures 11.14, courtesy of Crane Co. I’.\l<-1‘ 3. 3-14 usctl in steam service, the closed valve may trap steam between the discs where it contlenscs and creates a V~CLIUIII. This may result ii1 leakage at the valve seats. l’Il’IN(; DESIGN \Vhc~i , ‘- HANDWHEEL ,RISES WlTH STEM, ,-PACKING GLOBE, ANGLE AND “Y” VALVES These three valves arc of the same basic design, LIX and construction. They are primarily intended Lor throttling service and give close regulation 0L’ flow. The method ol valve seating reduces wire drawing and seat erosion which is prevalent in gate valves when used for throttling service The angle or “Y” valve pattern is recommended lor frdl flow service since it has a substantially lower pressure drop at this condition than the globe valve. Another advantage of the angle valve is that it can be located LO replace an elbow, thus eliminating one fitting. Z;ig. 7, pnge IO, is a typical illustration of a globe valve, and Fig. 6, pnge 10, shows an angle valve. The “Y” valve is illustrated in Fig. 15. Globe, angle and “Y” valves can be opened or closed substantially faster than a gate valve because of the shorter lift oE the disc. When valves are to be operated l’requently or continuously, the globe valve provides the more convenient operation. The seating surfaces of the globe, angle or “Y” valve are less subject to wear and the discs and seats are more easily replaced than on the gate valve. Disc Construction There are several different disc and seating arrangements for globe, angle and “Y” valves, each of which has its own use and advantage. The different types are plug disc, narrow seat (or conventional disc), needle valve, and composition disc. I. The p/zdg disc has a wide bearing surface on a long, tapered disc and matching seat. This type of construction offers the greatest resistance to the cutting effects of dirt, scale and other foreign matter. The plug type disc is ideal for the toughest flow control service such as throttling, drip and drain lines, blow-off, and boiler feed lines. It is available in a wide variety of pressure temperature ranges. Fig. 7, pnge 10, shows a plug disc seating arrangement in a globe valve. 2. Nnrro~ sent (or conventional disc) is illustrated in an angle valve in Fig. 6. This type of disc does not resist wire drawing or erosion in closely throttled high velocity flow. It is also subject to erosion from hard particles. The narrow seat disc design is not applicable for close throttling. NUT , RISING STEM , IINSIDE S C R E W , COMPOSITION DISC SCREWED BONNET\-_ SCREWED ENDS> Courtesy of Jenkins Bras. FIG. 15 - “Y” VALVE S. Needle v&es, sometimes referred to as expansion valves, are designed to give fine control of ’ flow in small diameter piping. The disc is normally an integral part of the stem and has a sharp point which fits into the reduced area seat opening. I;&. 16 is an angle valve with a needle type seating arrangement. 4. Composition disc is adaptable to many services by simply changing the material of the disc. It has the advantage of being able to seat tight with less power than the metal type discs. It is also less likely to be damaged by dirt or foreign material than the metal disc. A composition disc is suitable to all moderate pressure services but not for close regulating and throttling. Fig. 15 shows a composition disc in a “Y” valve. This type of seating design is also illustrated in I;&. 19 in a swing check valve and in Fig. 20 in a lift check valve. RISING STEM I INSIDE SCREW) HANDWHEEL ( RISES WITH STEM ( w,THD”T GLAND) NEEDLE DISC FLUOW Y- FK;. 16 - SCREWED END ANGLE V A L V E b\‘lTH r\iEEDLE DI S C CHAPTER 1. PlPlK(; PLUG COCKS I)ESI(;K 3-15 - (;ENEIIAL. I plug (.ocks are primarily used lor ixilancing in ;I l)il)itlg system not subject to Ireclucnt changes in 110~~. They are normally less expensive than globe tyl)e valves and the setting cannot be tampered with as easily as a globe valve. Plug cocks have approximately the same line loss as ;I gate valve when in the fully open position. \\%en lxirtially closed for balancing, this line loss increases substantially. Z;ig. 17 is a lubricated type plug valve. HANDWHEEL ,RISES WITH STEM, RISING STEM lOUTSIDE SCREW, -BOLTED BONNET WIT” OIAPHRLGMSEAL, FLANGED ENDS -- FLOW d REFRIGERANT VALVES Refrigerant valves are back-seating globe valves of either the packed or diaphragm packless type. X’he packed valves are available with either a hancl wheel or a wing type seal cap. The wing type seal c“? is preferable since it provides the safety of an . itional seal. Where frequent operation of the valve is required, the diaphragm packless type is used. The diaphragm acts as a seal ant1 is illustrated in the “E‘” valve construction in I;ig. 18. The refrigerant valve is available in sizes up to IS/~ in. OD. For larger sizes the seal cap type packed valve is used. -’ COMPOSITION FIG. 18 - "Y" V,\LVLC (DIAI~HKAGM I‘u15) BOLTED CHECK VALVES ‘COMPOSITION There are two basic designs of check valves, the swing check and the lift check. The swing check valve may be used in a horizontal or a vertical line for upward flow. A typical swing check valve is illustrated in Fig. 19. T h e f l o w thru the swing check is in a straight line and without restriction at the seat. Swing checks are generally used in combination with gate valves. FIG. 19 --SWING CHECK VALVE SCREWED UNION RING BONNET COMPOSITION DISC FLOW SCREWED END FIG. 20 - LIFT Courtesy FIG. 15 - of Walworth Co. PLUG C O C K Ckik~K VALVE The lift check operates ill a manner similar to that of a globe valve and, like the globe valve, its llow is restricted as illustrated in Fig. 20. The disc is seated by backflow or by gravity when there is no Ilow, and is free to rise and fall, depending on the pressure under it. The lift check should only be installed in horizontal pipe lines and usually in combination with globe, angle and “Y” valves. Figuws II;, 18-20. courtesy of Crane Co. VALVE AND FITTING PRESSURE LOSSES SPECIAL SERVICE VALVES ‘J‘llc1.e ;tre scver;~i tyI)cs ol valves collllllorlly usctl ii1 tliffcretit !)iping ;i!~plic;itions t!l;It do IlOt nCccssarily fall i n t o the classificatiorl 01 general purpose v;rlvcs. b~xlxlnsion, relief, ant! solcnoit! valves are COIIIC ol tlie more’ c o m m o n s!)eci;il purpose valves. ;\ rclicf valve is licltl closctl by ;I spring or some otlicr means anti is tlesignet! to ;ititom;itically rclievc tlic line or container pressure in excess of its setting. III gcncr;il ;I relief valve should Ix installet! wherever there is any danger of the fluid pressure rising above tile tlcsign working pressure of the pi!x fittings or !xessurc vcssc!s. rJ‘o properly design any type of piping system a Iluir!, t h e losses thru t h e valves and [ittings in clic system must IK realistically ev:~luatet!. Tables hvc Ixxn prepared for cletermining these losses in terms of ec!uivdcnt Icngth of pipe. These values are then used with the correct friction chart for the p;irticular fluid flowing thru the system. TaOle /O gives valve losses with screwct!, Ilanget!, llaret!, weltlcr!, soltlcret!, or Ixxzetl connections. Tul~/e I/ gives fitting losses with scrcwet!, Ilan~ctl, Il:irctl, welded, solclcretl, or brazed connections. Tnble I2 lists the Iosses for special types of fittings sometimes encounterctl in piping applications. wnvcyin~ . TABLE IO-VALVE LOSSES IN EQUIVALENT FEET OF PIPE* Screwed, Welded, Flanged, and Flared Connections GLOBE+ 60° - Y 4s0 - Y ANGLEt GATEtt i WING CHECKf IF1 CHECK NOMINAL PIPE . OR ;;,,TUBE SIZE (in.) .l% J/ H J/4 17 18 22 1% 29 38 43 15 20 24 12 15 18 2 2% 3 55 69 84 30 35 43 24 29 35 3% 4 5 100 120 140 6 8 170 220 280 1, 88 115 145 70 85 105 0.6 0.7 0.9 5 6 8 1 .o 1.5 1.8 10 14 16 24 29 35 2.3 2.8 3.2 20 25 30 41 47 58 4.0 4.5 6 35 40 50 70 85 105 12 14 16 360 410 130 155 180 18 20 24 460 520 610 200 235 265 *Losses 7 9 Globe & Vertical Lift Same as Globe Valve** 60 80 100 Angle Lift S a m e OS Angle 19 22 25 165 200 240 VOlV.2 cue for all valves in fully open position. tThese l o s s e s d o n o t a p p l y t o v a l v e s w i t h n e e d l e p o i n t t y p e s e a t s . fLossa also a p p l y t o t h e i n - l i n e , b a l l t y p e c h e c k v a l v e . **For “Y” pattern globe lift check valve with seat approximately equal to the nominal pipe diameter, use values of 60” “Y” valve for loss. ttRegular and short pattern plug cock valves, when fully open, have some loss as gate valve. For valve losses of short pattern plug cocks above 6 ins. check manufacturer. clfnf’-r r,.fc I . f’II’lN(. Dk:SI(;I\‘-- (;E?Jl:RAI. 3-17 TABLE II-FITTING LOSSES IN EQUIVALENT FEET OF PIPE Screwed, Welded, hOOTH BEND 9o" street* NOMINAL Flanged, Flared, and Brazed I ELBOWS 45" 45" Std* street* Connections 180’ Std* SMOOTH BEND TEES ~ /Q 2.3 2.5 3.2 1 1% 1% I 2.6 3.3 4.0 4 5 I 10 13 1.7 2.3 2.6 6.7 8.2 0.7 0.8 0.9 2.7 3.0 4.0 4.1 5.6 6.3 1.3 1.7 2.1 5.0 7.0 8.0 8.2 10 12 2.6 3.2 4.0 10 12 15 3.3 4.1 5.0 4.7 5.6 7.0 5.0 6.0 7.5 15 17 21 4.7 5.2 6.5 21 25 I8 5.9 6.7 8.2 8.0 9.0 12 9.0 10 13 25 - 12 I 14 16 30 34 38 18 20 24 42 50 60 I * 0.9 1.0 1.4 1 1.7 2.3 2.6 - 16 18 20 68 78 19 23 26 29 33 40 - 23 26 30 85 100 115 29 33 40 60 MITRE J/s % vi 1 1% 1% I 1 ELBOWS 90° Eli 60' Eli 2.7 3.0 4.0 1.1 1.3 1.6 0.6 0.7 0.9 0.3 0.4 0.5 5.0 7.0 2.1 3.0 3.4 1.0 1.5 1.8 0.7 0.9 1.1 8.0 45' EII 30' EII 2 2% 3 10 12 15 4.5 5.2 6.4 2.3 2.8 3.2 1.3 1.7 2.0 3 % 4 5 18 21 25 7.3 0.5 11 4.0 4.5 6.0 2.4 2.7 3.2 6 8 30 40 50 13 17 21 7.0 9.0 12 4.0 5.1 7.2 10 / 2.3 3.1 3.7 1.4 1.6 2.0 1 2.6 3.3 4.0 7.9 10 13 19 23 26 NOMINAL 1.2 1.4 1.9 ~jl *R/D approximately equal to 1. tR/D approximately equal to 1.5. 1 / 26 30 35 40 44 50 I 1 30 34 38 42 50 60 ' P.\RT 3. PIPING; I)ESIGN t TABLE 12-SPECIAL SUDDEN ‘h ENLARGEMENT* ‘h d/D vi FITTING LOSSES IN EQUIVALENT FEET OF PIPE SUDDEN % CONTRACTION* d/D % ‘/a SHARP EDGE* Entrance PIPE Exit PROJECTION* Enttmce ! Exit NOM. 18 20 24 - - - - ( 18 - - 20 - - - *Enter table for losses at smallest diameter “d.” / 1.5 1.8 2.8 .8 I .o I .4 1.5 I .8 2.8 1.1 1.5 2.2 3.7 5.3 6.6 1.8 2.4 3.3 3.7 5.3 6.6 2.7 4.2 5.0 9.0 12 14 4.4 5.4 7.2 9.0 12 14 6.8 8.7 II 17 20 27 8.5 10 I4 17 20 27 13 I6 20 33 47 60 19 24 29 33 47 60 25 35 46 18 73 86 96 37 45 50 73 86 96 57 66 77 20 - 115 142 163 I ( :1: 83 I 1 1:: 163 ! 1:: 130 ( . t , 3-19 CHAPTER 2. WATER PIPING Tliis chapter presents the principles and currcntly accepted design techniques [or water piping systems i~s~tl in air conditioning applications. It also includes the various piping arrangements for air conditioning equipment and thestandard accessories lountl in most water piping systems. The principles and techniques described are applicable to chilled water and hot water heating systems. General piping principles and techniques are described in Chapter I. WATER PIPING SYSTEMS Once-Thru and Recirculating ‘he water piping systems discussed here are di..:cl int’o once-thru and recirculating types. In a dnce-thru system water passes thru the equipment only once and is discharged. In a recirculating system water is not discharged, but flows in a repeating circuit from the heat exchanger to the refrigeration equipment and back to the heat exchanger. Open and Closed Both types are further classified as open or closed systems. An open system is one in,which the water flows into a reservoir open to the atmosphere; cooling towers and air washers are examples of reservoirs oplen to the atmosphere. A closed system is one in which the flow of water is not exposed to the atmosphere at any point. This system usually contains an expansion tank that is open to the atmosphere but the water area exposed is insignificant. 011 ijew co11struction. The length of the water circuit thru the supply ;incl return piping is tllc same for aI1 uliits. Since the water circuits are equal for each unit, the major advantage of a reverse return system is that it seldom requires balancing. Fig. 21 is a schclnatic sketch of this system with units piped liori~ontally and vertically. There are installations where it is both inconvenient and economically unsound to use a complete reverse return water piping system. This somctimcs exists in a building where the first floor has prcviously been air conditioned. To avoid disturbing the lirst floor occupants, reverse return headers arc located at the top of the building and direct return risers to the units are used. I;;!{,:. 22 illustrates a reverse return header and direct return riser piping system. In this system the llow rate is not equal for all units on a direct return riser. The difference in Ilow ’ rate depends on the design pressure drop of the supply and return riser. This differcncc can be rctlucctl to practical limits. The pressure drop across the riser includes the following: (I) the loss thru the supply and return runouts from the riser to the unit, (2) the loss tliru the unit itsell, ‘he recirculating system is further classified according to water return arrangements. When two or more units are piped together, one of the following piping arrangements may be used: 1. Reverse return piping, 2. Kevcrse return header with direct return risers. 3. Direct return piping. If the units have the same or nearly the same Ill-essurc drop thru them, one of the reverse return mcthotls of piping is recommendctl. However, if the units have tliffercnt pressure drops or require balancing valves, then it is usually more ccononiical lo use ;I direct return. Reverse return piping is rcconinicntletl for most c10Setl pipilig applications; it cannot be ~~sctl 011 (JpXl systems. It is often the most economical design the loss thru E UNIT Water Return Arrangements : and (3) “NlTS PIPED VERTICALLY UNIT SUPPLY t I 1 t I 1 I I t ~RETURN UNITS PIPED HORlZONTAl.LY FIG. 2 1 - KEVE:RSE KE-IIJRN P I I I P I N G I, ii~21tlccl for a closed recirculating system where all the units require Mancing valves ant1 have different i prcssurc drops. Several Call-coil units piped tog&cl- , ;iiid rccl~iiring tlilfercnt w a t e r flow r a t e s , c;iIxicities * and prcssurc (17-01)s i s 211 es;iil~~~le 01 t h i s type ol ’ \yalclll. ; : llle direct return piping system ‘is iriiYe&tly u~ilx~la~icetl and rcquircs Ix~lancing valves or orifices; and provisions to measure the pressirrc~drop in order to meter the water flow. ~~\lthough Jnatcrial costs are lower in this system than in the two reverse return systems, engineering cost and balancing time often otfsct this advantage. Fig. 23 illustrates units piped vertically and horizontally to a direct return. the fittings and valves. Excessive unlx~lancc in the direct auI)IAy and return portion of the piping system ntay dictate the need for balancing valves or orifices. T o climinatc balancing valves, design the s u p p l y and return pressure d r o p equal to one-fourth the suni ol the pressure d r o p s 01 the preceding fteJJ7s I, 2 (llld 3. Direct retinxl piping is necessary for open piping systems and is rcconimentlcd for some closed piping systems. A reverse return arrangement on an open systcin requires piping that is normally unnecessary, siiicc the same atmospheric conditions exist at all opi points 01 the systcni. ,A direct return is rccom- -E SUPPLY RETU6-4 UNITS PIPED I I’ ‘1 -P UNIT I’ PIPED PIPING DESIGN ‘I‘herc is a friction loss in any pipe thru which water is flowing. This loss depends on the following lactors: I. Water velocity 2. Pipe diameter :i. Interior surface roughness 3. Pipe length System pressure has no effect on the head loss ol the equipment in the system. However, higher than noriiial s y s t e m pressures may dictate the use of heavier pipe, fittings and valves along with specially tlcsignctl cquipmcnt. T O properly design a wat& piping system, the cliginecr must evaluate not only the pipe friction loss hut the loss thru valves, fittings and other equipJllent. In addition to these friction losses, the USC of diversity in reducing the water quantity and VERTICALLY I I UNITS WATER CONDITIONING il:ormally all water piping systems must have aclequatc treatment to protect the various components against corrosion, scale, slime and algae. Waier treatinent should always be under the supervision of a water conditioning specialist. Periodic inspection of the water is required to maintain suitable quality. p~J1.l 5 of this manual contains a discussion Of the various aspects of water conditioning including cause, effect and remedies for corrosion, scale, slime and algae. WATER UNIT SUPPLY CODES AND REGULATIONS ^.All applicable codes and regulations should be checked to determine acceptable piping practice for the particular application. Sometimes these codes and regulations dictate piping design, limit the pressure, or qualify the selection of materials and equipment. HORIZONTALLY . l CHAPTER 2. \V.\TER 3-21 l’II’IK(; TABLE 1 I-RECOMMENDED WATER VELOCITY VELOCITY SERVICE PIPE FRICTION LOSS ‘1‘11~ l)il)c [rictioll loss water 1)~s~ \,clo(:ity, iI1 :I systelll pipe diameter, interior ;lll(l piljc lellgth. Varying tk~Kl~& slll~Llc:c (fps) 011 rough- any ollc OL t h e s e hc- tars inlluences the total friction loss in the pipe. ,\lost air conditioning ;il)piications use either steel l)ipe (jr (:oppu- tubing in the piping system. 7’0 CKlIll;ltC: the [riction loss in steel pipe or copper tubing, refer to Cl~0j’t.s j 111~71 5 in this chapter. Clrtl72.y j and f are for Schedule 40 pipe up to 2J in. in diameter. CAn7.t j shows the friction losses for closed recirculation piping systems. The friction losses in C/l& f are for open once-thru and for open reci- .!iation piping systems. ( ,,.,r~t j shbws friction losses for Types K, L and RI1 copper tubing when used in either open or closed water systems. These charts show water velocity, pipe or tube diameter, and water quantity, in addition to the friction rate per 100 ft of equivalent pipe length. Knowing any two of these factors, the other two can be easily determined from the chart. The effect of inside roughness of the pipe or tube is considered in a11 these values. The water quantity is determined from the air conditioning load and the water velocity by prcdetermined recommendations. These two factors arc used to establish pipe size and friction rate. Water RANGE Velocity The velocities recommended for water piping tlepencl on two conditions: The service for which the pipe is to be used. 2. The effects of erosion. Table 13 lists recommended velocity ranges for. different services. The design of the water piping system is limited by the maximum permissible flow velocity. The maximum values listed in Table 13 are based on established permissible sound levels of moving water and entrained air, and on the effects of erosion. Erosion in water piping systems is the impingement on the inside surface of tube or pipe of rapidly nioving water containing air bubbles, sand or other solid matter. In som&cases this may mean complete deterioration of the ‘tube or pipe walls, particularly on the bottom surface and at the elbows. Since erosion is a function of time, water velocity, and suspended materials in the water, the selection of a design water velocity is a matter of judgment. ‘1%~ ni;ixiinuIIL water velocities presented in ?‘rlOle If ;lix based on many years of cxpericncc 2nd t h e y insure the attainment of optimum equipment lilt unticr normal conditions. Friction Rate The design of a water piping system is limited by the friction loss. Systems using city water must have the piping si/.etl so as to provide the required BOW rate at a pressure loss within the pressure available at the city main. This pressure or friction loss is to include all losses in.the system, as condenser pressure ,tirop,. pipe and fitting losses, static head, and water meter drop. The total system,pressure drop must be less than the city main pressure to have design water flow. A recirculating system is sized to provide a rcasonable balance between increased pumping horsepower due to high friction loss and increased piping first cost due to large pipe sizes. In large air conditioning applications this balance point is often taken as a maximum friction rate of 10 ft of water per 100 ft of equivalent pipe length. In the average air conditioning application the installed cost of the water piping exceeds the cost of the water pumps and motors. The cost of increasing the pipe size of small pipe to reduce the friction rate is normally not too great, whereas the installed cost increases rapidly when the size of large pipe (approximately 4 in. and larger) is increased. Smaller pipes can be economically sized at lower friction rates (increasing the pipe size) than the larger pipes. In most applications economic considerations dictate that larger pipe be sized for higher flow rates TABLE 14-MAXIMUM WATER VELOCITY TO MlNlMlZE EROSION NORMAL OPERATION (hr/vd 1500 2000 3000 4000 6000 8000 WATER,VELOCITY (fPf) 12 11.5 11 10 9 8 CHART 3-FRICTION LOSS FOR CLOSED PIPING SYSTEMS Schedule 40 Pipe .I 15 \ 20000 15000 2.0 .25 .3 .4 .5 .6 .8 1.0 1.5 \ \ \ 10000 e 000 6 0 0 0 5000 4 0 0 0 3000 2000 . I500 IO00 800 \ 600 j, 600 500 400 3 0 0 200 150 100 80 6 0 60 50 5 0 4 0 4 0 3 0 2 0 15 IO 8 6 6 5 5 4 4 FRICTION LOSS (FEET OF WATER PER 100 FT) 3-23 CHAI’TER 2. WATI~K PIPING CHART 4-FRICTION LOSS FOR OPEN PIPING SYSTEMS Schedule 40 Pipe I .I5 .2 .25 .3 4 5 .6 .0 II 1.0 1.5 2 . 0 2.5 3 4 5 6 0 IO I5 2 0 25 30 4 0 60 00 100 N 10000 0 0 0 0 6 0 0 0 6 0 0 0 5000 5 0 0 0 4000 4 0 0 0 3000 2000 I500 IO00 000 600 500 300 z a 2 0 0 cl 3 s L 150 100 100 00 00 60 6 0 50 5 0 4 0 4 0 3 0 20 I5 IO IO 0 7 0 7 6 6 5 5 4 4 3 2 1.5 1.0 FRICTION LOSS (FEET OF WATER PER IOOFT.) L 3-24 CHART S-FRICTION LOSS FOR CLOSED AND Copper 4000 I .I5 .2 .25 .3 .4 .5 .0 .6 I.0 1.5 2 2.5 I’,\R’I’ 3. I’II’INC lIESIGN OPEN PIPING SYSTEMS Tubing 3 4 010 56 15 20 25 30 40 50 60 3000 2000 I ITYPk Id U+-I-I+l i 1500 1000 000 000 600 6 0 0 500 500 400 4 0 0 300 3 4 0 s L 3 0 2 0 15 6 5 4 .I .I5 .2 .25 .3 .4 .5 .6 .0 1.0 FRICTION 1.5 LOSS 2 2.5 (FEET 3 OF 4 5 6 W A TER 0 IO 15 PER loons.) 20 25 30 4 0 50 60 00 1.0 100 ’ ;IIJ([ l)r(‘ssurc drops that~ smaller l)ipc wliic.11 is si/ctl 1’0~ l~mu- prcssurc drops ant1 (low rates. l&cptiolls to tllis general guide oltcn occur. I;or appearance or physical limitations may tli(,t;lte the use 01 sriiall pil)cs. Tllis is ol’ten done I’or short runs where the total 1)ressurc tlrol) i s n o t greatly inllucnced. E;~ch s y s t e m al~oultl be alialy~ccl scparatcly t o tleternline tlic economic balance between first cost (pipe si/e, pump ;intl motor) and operating cost (1)rcssure drop, pump and motor). esmple, Pipe Length -1‘0 clctcrminc the friction loss in a water piping straight system, the engineer must calculate the lengths ol pipe and evaluate the additional cquivalent lengths of pipe due to fittings, valves and other ‘~lents in the piping system. Tn1)le.s IO, If and 12 b me the ‘additional equivalent lengths of pipe for these various components. The straight length of pipe is measured to the centerline of all fittings and valves (Fig. 2-f). The equivalent length of the components must be added to this straight length oE pipe. WATER PIPING DIVERSITY * LVhen the air conditi6ning load is determined for each exposure of a building, it isassumed that the exposure is at peak load. Since the sun load is at a maximum on one exposure at a time, not all of the units on all the exposures require maximum water How at the same time to handle the cooling load. Units on the same exposure normally require maximum How at the same time; units on the adjoining or opposite exposures do not. Therefore, if the individual units are nzrtonznticnlly controlled rary the water quantity,.the system water quantity ‘\ sicually required during normal operation is less than the total water quantity required for the peak design conditions for all the exposures. Good engineering design dictates that the water piping and the pump be sized for this reduced water quantity. The principle of diversity allows the engineer to evaluate and calculate the reduced water quantity. In all water piping systems two conditions must be satisfied before diversity can be applied: 1. The water How to the units must be automatically controlled to compensate for varying loads. 2. Diversity may only be applied to piping that supplies units on more than one exposure. I;iRlye 25 is a typical illustration of a header layout to which diversity may be applied. In this il- FK. 24- PIPL LENGTH MEASUREMENT lustration the header piping supplies all I‘our exposures. Assuming that the units supplied are automatically controlled, diversity is applied to the ’ T\‘est, south and east exposures only. The last leg or exposure is never rcducecl in water quantity or pipe size since it requires full flow at some time during operation to meet design conditions. Fi,qxye 16 illustrates another layout where diversity may be used to reduce pipe size and pump capacity. In this illustration diversity may be applied to the vertical supply and return headers and also to the supply and return branch headers at each Boor. Diversity is not applied to pipe section ‘i-8 of both the supply and return vertical headers. In addition N W FK. 25 - HEAL)EK I’II’INC 3-26 I’/\Rl’ 3. I’II’INC; I>ESI(;N 4 Solution: I . l’utnp .\ srtpplics nor111 ant1 west cxl)~~s\Irc I)llt tlivcrsity can IX applic~l tn norrh cxpos”rc only. The tolal g~“n in p~~n,p ;\ circuit is 280 gpm and the a~~~mulatctl gpn’ in the north exposure is i(iO gpm. The ratio of accllmlliatctl gpm to the total waler quantity in the circuit is: I GO 280 Enter Clrccrl factor ,785. Pump ..)I 6 at the ratio 57 and rcatl tllc t l i v c r s i t y 1% circuit has a ratio for the cast exposure of: 120 "80=.43 Entering Chart 6 at the is read as ,725. 2. ratio of .43, the diversity factor The following table illustrates how the diversity factors are applied to the maximum water quantities to oljtain the design water quantities. . I’UMI’ “A” CIRCUI-I Max Quantity (mm) Section F IG. 26 - HORIZONTAL \VVA.TER PMNG LAYOUT the south leg of the return piping and the west leg of the supply piping on each floor must be full size. In any water piping system with automatically controlled units, the water requirements and pump head pressure varies. This is true whether or not diversity is applied. However the water requirements and pump head vary considerably more in a system in which diversity is not considered. In a system in which diversity is not applied, greater emphasis is required for pump controls to ,’ revent excessive noise being created by throttling %-. alves or excessive w5ter velocities. In addition, since the system never requires the full water quantity for which it is designed, the pump must be either throttled continuously, or bypassed, or reduced in size. It is good practice, therefore, to take advantage of diversity to reduce the pipe size and pump capacity. Chn~t 6 gives the diversity factors which arc used in water piping design. Exnmple I illustrates the use of Chn~t 6. Example 1 - Diversity Factors for Wafer Piping Given: Water piping layout as illustrated in Fig. -77. Find: I. Diversity factor to I)e applied to the water quantity. 2. Water quantities in header sections. Headers Design Quantity (gpm) Diversity Factor A-RI RI-R2 R2-R3 R3-R4 R4-R5 280 260 240 220 200 .785 .785 .785 .785 ,785 220 204 188 I73 157 R.‘,-RG R6-R7 R7-R8 KS-R% R9-RIO 180 160 140 I20 100 ,785 ,785 ,785 I41 * 126 RIO-RI1 Rll-RI2 Rl2-RI3 Rl3-RI4 I 1 80 60 40 20 I I P U M P “B” I .oo 1 .oo 1.00 1 .oo I .oo I .oo I (110)120* 120 100 80 GO 40 20 CIRCUIT B-R28 R28-R27 R27-R26 R2F-R25 R25R24 R24-R23 R23-R22 R22-R2l R2l-R20 R20-RI9 280 260 240 220 200 160 140 120 100 I 1 I I .oo .oo .oo .oo 160 140 120 100 Rl9-R18 Rl8-RI7 R17-RI6 RI&RI9 80 . 1.00 I .oo I .oo I .oo 80 180 GO 40 20 ,725 .725 ,725 ,725 .725 203 188 174 ( 1 4 5 )IGo* .725 ( 1 3 0 )lGO* 160 GO 40 20 *When applying diversity, the design water quantity in the last section of the exposure is usually less than the water quantity in the first section on the adjoining exposure. When this occurs, the water quantity in the last section or last two sections is increased to equal the water quantity in the first section of the next exposure. CH.\ P~I‘Icl< 2. wA’I‘I:Iz 3-27 PI 1’1 xc; CHART .I0 .20 .30 S--DIVERSITY .40 50 FACTORS .60 .70 .80 .90 .I00 ACCUMULATED WATER FLOW FOR EXPOSURE TOTAL WATER FLOW FOR PUMP N 20 In Example 1 pump “A” is sclcctcd for 220 gpm ant1 pump “B” is selected for 203 gpm. The pipe skes in the north and east exposures are reduced using the design gpm, whereas the pipes in the south and west exposures are select4 lull size. Esrrttqle -3 077d 3 illustrate the economics involved when applying diversity. Esatttple 2 shows ;I typial header layout with one 1)ump serving ~11 (our esposures. The header is sized without diversity. Exrrtttple 3 is the sanic piping layout Ijut is used to size the heatlcr. tlivcrsity 2. ;I 20-m RIO ZO- RII 20- RI2 20- RI3 2d RI4 RI5 20 20 20 20 20 A7 R6 R5 R4 R3 ‘” 2 0 p. PLAN VIEW W 20 20 E RI6 RI7 RI8 RI9 R20 20 20 20 20 20 S R23 R2l R 2 2 20 \ 20 20 Eli,ows. R/I) - 1 Kxpcctcd Icngth of operation - 6000 Itours NORTH RI R2 \ R4 R3 U U A R5 u U 320 3 4 0 1 7.0’ I 40 20’ 760 7 FIG. Solution: I. Design water velocity for sizing the hcatlers is determined from Tnl~les 13 14. Slaximum water velocity - i fps 2 . \laximum water quantity required when no diversity is applied is 3G0 gpm. Pump is selected for 360 gpm. c 240 PLAN VIEW 20’ ;;;I 60 4’1 I:intl: I. lksign header water velocity 2. \Vater quantity for pump selection 3. Header pipe size and pump friction head I- 20’ 100 n , 1 20’ 120 rl ( 20’ 1 140 r-l , 3. The table below gives the header pipe sizes and pump Friction head when no diversity is applied. 20’ 1 I60 160 n Example 3 - Sizing Header Using Diversity 28 - SUPPLY HEADER PIW SIZING Example 2 - Sizing Header Using No Diversity Given: .\ building with a closed recirculation water piping system using a horizontal header and vertical risers as illustrated in Frrg-. 28. 1laximum llow to each riser - 20 gpm Schetlulc 40 pipe and fittings. Given: Same piping layout as in Exa+e -7 and Fig. 28. Maximum flow to each riser - 20 gpm Schedule 40 pipe and fittings Elbows, R/D = 1 Expected length of operation - 6000 hours Maximum design velocity - 7 fps (Exantple 2) Find: I. 2. 3. 4. . Diversity factor for each exposure Design gpm for each header section JVater quantity for pump selection Header pipe size and pump friction head . H E.-\DER SECTION TORI Rl-R2 R2-R3 R3-R4 R4-R5 PIPE SIzEt kP’) (in.) SF0 340 320 300 280 - 5 5 5 .i 4 ’ . LENGTF RETWEE: T.iKEOFFS w - 27 I8 20 20 20 260 41 RF-R; R7-R8 R8-R9 240 220 200 4 -I 4 20 20 20 R9-RIO 180 4 8 RIO-RI1 Rll-RI2 R12-RI3 160 140 120 3 3 3 20 20 20 R13-RI4 RI4R15 100 80 RlS-RI6 R16-R17 R17-RI8 GO -10 2 0 R5-R6 ‘- W.\TER OUAN?I-ITY 8 FITTIXGS - FITTING’ EQUIVALENT LENGTH* w TOTAL EQUIV.-\LENT LENGTH (4 FRICTION LOSS? (ft of water per 1 0 0 equiv ft) FRICTION HE.AD (ft of water) 2-ells .I-tee l-tee l-tee 1 -red. tee 26 8.2 8.2 8.2 12.0 53.0 26.2 28.2 28.2 32.0 2.3 2.0 I .8 16 4.4 1.22 .53 .dl .45 I.41 l-tee I-ell I-tee l-tee l-tee 6.7 10.0 6.7 6.7 6.7 24.7 26.7 26.7 26.7 3.8 3.2 2.7 2.3 .94 l-tee 1 -ell l:red. tee l-tee l-tee 6.7 6.0 9.0 5.0 5.0 20.7 29.0 25.0 25.0 2.1 5.5 4.6 3.2 .43 I .59 I.15 .80 l-tee l-tee 1-ell 1 -red. tee 1 -tee 1 -red. tee 5.0 !‘,.O 7.5 7.0 3.3 5.0 25.0 2.5 20.5 27.0 23.3 25.0 I.6 6.8 3.2 B.5 Pump friction .85 .72 .62 .33 I .84 .i5 I .62 headf 16.39 *Fitting losses are determined from Tnhlr II. For reducing tees enter Table I! at the larger diameter. tFriction rate and pipe size are determined from Cltnrr 3 not exceeding the maximum design water velocity (7 fps). :Pump friction head does not include losses For valves. strainers, etc., which must he included in the actual design. (:HAPTER 2 . WA’I‘ER 3-29 I’II’INC DINtiN I . Cl~ur-t 0 i s risccl with tllc ratio of accumulated gpm i n rhc exposure to the total pump gpm, in order to dctcrfollowing talk illustrates mint l h c diversity factors. The the method of determining diversity factors. (First cxposure listed is always first exposure served by pump.) l<SI’OSI!RI: North East S011th ivcst =$yJ;E ACCtihI. Gl’hl DIVERSITY FACTOR TOTAL I’UXII’ Gl’hf Q”,;NTrTY (gP’n) 100 80 100 HO 100/360 = .28 180/360 = .5O .Gi .5G 280/360= .i8 3 6 0 / 3 G O= I . o o 239 I .oo 2. The diversity factor found in SteL I is applied to the maximum water quantity in each header section to cstahlish the design gpm for sizing the header. The table at right gives the design water quantity for the various header sections. 3. 4. The design water quantity required when diversity is applied is 240 gpm. for pump HEADER SECTIOS 4 4 4 4 4 27 18 20 20 -20 115.RG 197 4 8 RF-R7 R7-R8 RS-R9 182 4 167 160 4 3 20 20 20 R9-RIO 160 3 8 Rl3-RI4 R14-RI5 90 R15-RI6 Rl6-RI7 Rli-R18 GO 80 40 20 20 20 20 20 8 20 20 20 260 240 220 200 .7G 7G .7G .iG 1 160 140 120 100 39 39 142 .89 .89 12-i 1 .oo 106 90 80 1.00 I .oo 1 .oo GO 40 20 , GO / 40 20 (lH7)197# I97 182 lG7 (152)160f 160 .89 80 / Rl5-RIG ( Rl6-Rl7 R17-R18 *lVhen applying diversity, the design water quantity in the last section of the exposure is usually less than the design water quantity in the first section of the adjoining exposure. When this occurs, the water quantity in the last section or last two sections is increased to equal the water quantity in the first section of the next exposure: (in.) 240 227 ii7 67 .G7 Ai7 IL RlO-RI 1 Rl l-R12 Rl2-RI3 Rl3-RI4 R14-RI,5 PIPE SIZEt To Rl RI-R2 R2-R3 R3-R4 R4-R3 2 14 201 197 R5-RB RG-R7 R7-RH R&J-R9 180R9-RlO 340 320 300 280 selection ‘rhe design water quantity found in Strp -3 is used in sbing the header pipe and in establishing the pump friction head. The talk Ijelow illustrates the header pipe sizing: DESIGK W’.\TER QUANTITY (gpm) RI-R2 R?-KS RS-R4 R4-R5 \V.\‘I‘ER (,)LJ.\N’[‘I’I‘Y kI”“) 240 ““7 ..h 2 I4 201 2-ells l-tee l-tee 1 -tee l-tee 20.0 6.7 6.7 6.5 6.7 l-e11 l-tee l-tee l-tee 1 -red. tee 10.0 6.7 6.7 6.7 9.0 1 -ell l-tee l-tee 1-tee l-tee 7.5 5.0 5.0 5.0 5.0 l-tee 1-ell l-tee l-red. tee l-tee 1 -red. tee 5.0 7.5 4.0 7.0 3.3 5.0 i !I TOTAL EQUIVALENT LENGTH (W FRICTION Losst (ft of water per 100 equiv ft) FRICTION HEAD 47.0 24.7 26.7 26.7 26.7 3.4 3.0 2.7 2.3 2.3 1.60 24.7 26.7 26.7 29.0 2.3 2.0 1.8 5.6 57 20.5 25.Q 25.P 25.0 5.6 4.5 3.5 2.7 (ft of water) .74 .72 .61 .61 .53 .48 1.62 1.12.. .8i 68 25.0 20.5 27.0 23.3 25.0 Pump friction head: 16.35 ‘Fitting losses are determined from Tnble 11. For reducing tee enter Table II at the larger diameter. tFriction rate and pipe size are determined from Chart 3 with the water velocity not exceeding 7 fps. :Prlmp friction head does not include losses for valves, strainers, etc., which must be included in the actual design. 3l30 i lhcitrtp1e.r 2 clrlfl 3 iridicalc that the following reductions iI1 pipe ant! fitting siLc c;lIl be made when diversity is used: I. 5 j ft of 5 in. pipe replaced with 4 in. pipe. 2. 28 ft of ‘1 in. pipe replaced with 3 in. pipe. 3. 8 fittings rccluced 1 size. In addition the pump can be selected for 240 gpm instead of 360 gpm which is approximately a l/s reduction. Other areas whcrc a reduction in size is possible are: 1. Pipe and fitting5 in the return piping header. 2. Valves, unions, couplings, strainers and other clcmcnts located in the supply and return licadcrs. PUMP SELECTION Pumps are selected so that there is no sustained ise in pressure when the water Row is throttled. Systems having considerable throttling have the pump selected on the flat portion of the “headversus-flow” curve. Normally, new installed pipe has less than design friction and, therefore, the pump delivers greater gpm than clesign and requires more horsepower. For this reason a centrifugal pump is always selected for the calculated pump head without the addition of safety factors. If the pump is selected for the calculated head plus safety factors, the pump must handle a larger water quantity. When this occurs and provision is not made to throttle or bypass the excess water flow, the possibility of pump motor overload exists. Again, if the pump is selected for maximum water quantity and diversity is not applied, the water flow must be throttled. This increases the ->ump head. SYSTEM ACCESSORIES AND LAYOUT I’i\K’I‘ 3. PIPING DESIGN The open and closed expansion tanks are the two types used in water piping systems. Open expansion tanks are open to the atmosphere and are located on the suction side of the pump above the highest unit in the system. At this location the tank provides atmospheric pressure at or above the pump suction, thus preventing air leakage into the system. ‘I’he static head on the pump due to the expansion tank must be greater than the friction drop oE the water in the pipe from the expansion line connection to the pump suction. In Pig. 139 the static head AB must be greater than the friction loss in line AC. Adding any accessories such as a strainer in line AC increases the friction drop in AC and results in raising the height of the expansion tank. ‘To keep the height of the tank at a reasonable level, accessories should be placed at points 1 and 2 in . I;is. < 29. At these designations the friction loss in line AC is not affected. The following procedure may be used to determine the capacity of an open expansion tank: 1. Calculate the volume of water in the piping, from Tables 2 and 3, pages 2 and 3. 2. Calculate the volume of water in the coils and heat exchangers. 3. Determine the percent increase in the volume of water due to operating at increased temperatures from Table 15. 4. Expansion tank capacity is equal to the percent increase from Table 15 times the total volume of water in the system. The closed expansion tank is used for small or residential hot water heating systems and for high temperature water systems. Closed expansion tanks are not open to the atmosphere and operate above EXPANSION TANK This section discusses the function and selection of piping accessories and describes piping layout techniques for coils, condensers, coolers, air washers, cooling towers, pumps and expansion tanks. ACCESSORIES Expansion Tanks An expansion tank is used to maintain system ljressure by allowing the water to expand when the water temperature increases, and by providing a method of adding water to the system. It is normally required in a closed system but not in an open system; the reservoir in an open system acts as the expansion tank. , FIG. 29 - STRAINER LOCATION SYSTEM IN W ATER PIPING CH,\I”1’1:K 2 . W.\~I‘EK 1’Il’INO 3-31 TABLE 15-EXPANSION OF WATER v = (0.00041 1 - 0.0‘166) v, t P,‘ P, I’I PO (Above 40 F) TEMP W) VOLUME INCREASE (%) (F) VOLUME INCREASE (%I 100 125 150 175 200 225 250 .6 1.2 I .a 2.8 3.5 4.5 5.6 275 300 325 350 375 400 6.0 8.3 9.8 11.5 13.0 15.0 TEMP atmospheric pressure. Air vents must be installed in the system to vent the air. Closed expansion tanks arc located on the pump suction side of the system to permit the pump suction to operate at or near constant pressure. Locating the expansion tank at the p~mip discharge is usually not satisEactory. iill ‘ssure changes caused by pump operation are subtractcd from the original static pressure. If the pressure drop below the original static is great enough, the system pressure may drop to the boiling point, causing unstable water circulation and possible pump cavitation. If the system pressure drops blow atmospheric, air sucked in at the air vents can collect in pockets and stop tvater circulation. The capacity of a closed expansion tank is larger than an open expansion tank operating under the same conditions. ASME has standardized the calculation of the capacity of closed expansion tanks. The capacity depends on whether the system is operating above or below 160 F water temperature. Water temperatures below 160 F use the following formula to determine the tank capacity: J, = E x vs I- P P 2-2 Pf p,, where: V, = minimum capacity of the tank (gallons). E = percent increase in the volume oE water in the system (Table 15). V, = total volume of water in the system (gallons). P, = pressure in the expansion tank when water first enters, usually atmospheric pressure (feet of water absolute). P, = initial fill or minimum pressure at the expansion tank (feet of water absolute). P, = maximum operating pressure at the expansion tank (feet of dater absolute). When the system water temperature is between 160 and 280 F, the following equation is used to determine the expansion tank capacity: where 1 = maximum average operating temp (F). Strainers The primary function of a strainer is to protect the equipment. Normally strainers are placed in the line at the inlet to pumps, control valves or other types of equipment that should be protected against damage. The strainer is selected for the design capacity of the system at the point where it is to be inserted in the line. Strainers for pump protection should be not less than 40 mesh and be made of bronze. On equipment other than pumps the manufacturer should be contacted to determine the degree of strainer protection necessary. For example, a control valve needs greater protection than a pump and, therefore, requires a finer mesh strainer. Thermometers and Gages Thermometers and gages are required in the system wherever the design engineer considers it important to know the water temperature or pres- : sure. The following temperatures and pressures are usually considered essential: 1. Water temperature entering and leaving the cooler and condenser. 2. Pump suction and discharge pressure. 3. Spray water temperature and pressure entering the air washer. Water thermometers are usually selected for an approximate range of 30 F to 200 F; they should be equipped with separable wells and located where they can be easily read. Pressure gages are selected so that the normal reading of the gage is near the midpoint of the pressure scale. Air Vents Air venting is an important aspect in the design of any water system. The major portion of the air is normally vented thru the open expansion tank. Air vents should be installed in the high points of any water system which cannot vent back to the open expansion tank. Systems using a closed expansion tank require vents at ail high points. Runoff drains should be provided at each vent to carry possible water leakage to a suitable drain line. PIPING LAYOUT Each installation has its own problems regarding location of equipment, interference with structural members, water and drain locations, and provision t 3-x P./\K’I‘ 3. I’IPINC; DESIGN c lor’ service and replacement. ‘l‘he following guides are presented to familiarize the cnginccr with accepted piping practice: I. Shut-ofE valves are installed in the entering and leaving piping to equipment. These are normally gate valves. This arrangement permits servicing or replacing the equipment without draining the entire system. Occasionally a globe valve is instaJlec1 in the system to bcrvc as one of the shut-off valves and in addition is used to balance water flow. Most often it is located at the primp discharge. In a close coupled system the shut-of valves may be omitted if the time and expense required to drain tile system is not excessive. This is a matter of economics, the first cost of the valves versus the cost of new water treatment and time spent in draining the system. ~i$Y 2. Systems using screwed, weided or soldered joints require unions to permit removal of the equipment for servicing or replacement. If gate valves are used to isolate the equipment in the system, unions are placed between the equipment and each gate valve. Unions arc also located before and after control valves, and in the branch of a three-way control valve. It is good practice to locate the control valve between the equipment and the gate valve used to shut,off flow to the equipment. This permits removal of the control valve from the system without draining the system. By locating the control valve properly, it is possible to eliminate the unions required for removal of the equipment. If the system &es flanged valves and fittings, the need for unions is eliminated. 3. Strainers, thermometers and gages are normally located between the equipment and the gate valves used to shut off the water flow to the equipment. Thd following piping diagrams are illustrated with screwed connections, However, flanged, welded or soldered connections may be used. These layouts have been simplified to show various principles involved in piping practice. Ilowing thru the coil or thru the bypass. It is rcgulatcd by a temperature controller. Gage cocks are usually installed in both the supply and return lines to the coil, This permits pressure gages to be connected to determine pressure drop thru the coil. The plug cock is manually adjusted t o s e t the pressure drop thru the coil. Figure 31 illustrates an alternate method of piping a water coil. The plug cock shown is used to adjust manually the water Ilow for a set pressure drop thru the coil. The pressure drop is determined by connecting pressure gages to the gage cocks. In this piping layout control of the leaving air tcmperaturc from the coil is maintained within a required range since normally the entering water is controlled to a set temperature. Often an air bypass around the coil is used to maintain final air temperature. Figz~l-e 3-3 illustrates a multiple coil arrangement. Piping connections for drain and vent lines for the coil arc included and should be I,$ in. nominal pipe size. The same principles covered in Figs. 30 and 31 are applicable to multiple coil arrangements. A globe valve may be substituted for the plug cock and gate valve combination in the return lines in Fig. 30,31 and 32. In this arrangement the globe valve is used to balance the pressure drop thru the coil, and also to shut off the water when szrvicing is required. However, it has disadvantages that AUTOMATIC AIR VENT THREE pt -WAY Water Coils F~~ZIWS 30 thou 36 illustrate typical piping layouts for chilled water coils in a closed piping system. The coil layout illustrated in 30 contains a three-way mixing valve. This valve, located at the cooling coil outlet, maintains a desired temperature by proportioning n?~tomnticnZZy the amount of water NOTE: Flange or union is located removed.’ FIG. 30-- SO coil may be CHILLED W ATER P IPING FOR COILS (A U T O M A T I C: CO N T R O L ) ‘ 3-34 PAR.1’ 3. I’II’ING DESIGN \ I’ig~es 34, 35 old 36 show typical piping layouts for multiple units in a horizontal installation. The principle difference in the three systems is the numbcr of shut-off valves (gate) and take-offs from the hcadcr. Since the header is located under the floor, each take-ofE must pass thru the floor. ThereEore, it is a matter of economics to determine the number ol shut-ofE valves required for servicing. Fig. 35 shows the minimum number of valves that may be used, anti Pig. 36 shows valves at each unit. Cooler A typical chilled water piping diagram lor a water cooler is illustrated in Fig. 37. In a close coupled system most of the gate valves can be omitted. If they are omitted, all 0T the water is drained Irom the system thru the drain valve when a component requires servicing. In an extensive piping system the gate valves arc used to isolate the equipment requiring servicing or replacement. I“igu~e 37 illustrates the recommended water piping and accessories associated with a cooler. HUT-OFF LAYOUT FOR Figure 38 sliows a water-cooled condenser using city, well or river water. The return is run higher than tlic condenser so that the condcnscr is always l’ull OC water. Water flow thru the condenser is modulated by the control valve in the su~~ply line. I;igure 39 is an illustration of an alternate drain arrangcmcnt for a con denscr discharging waste water. Drain connections oE all types must bc checked for compliance with local codes. Codes usually require that a check valve be installed in the supply lint when city water is used. I;igzwe -/O illustrates a condenser pipecl up with a cooling tower. If the cooling tower and condenser arc close coupled, most of the gate valves can be eliminated. If the piping system is extensive, the gate valves as shown are recommended for isolating ’ the equipment when servicing. When more than one condenser is to be used in the same circuit, the flow thru the condensers must be equalized as closely as possible. This is complicated by the following: SHUT-OFF VALVES ( GATE VALVE ) VALVE NOTES: 1. Though not shown, control valves (automatic or manual) may be required to control flow thru eacll unit. 2. A shut-off valve may be installed in the supply ant1 return branch heatlers when headers serve 3 to 5 units. 3. Supply ant1 return runouts to the coil shoultl have flared connections if the runouts are soft copper. Otherwise unions or Hanges are installect to facilitate servicing units. FIG. 34 - PIPING Condenser HOKIZONTAL MULTIPLE COILS (4 UNITS - 4 S HUT-OFF V ALVES) . SOTES: 1. Though not shown, control valves (automatic or manual) may he requirecl to control flow thru each unit. 2. Supply ant1 return runouts to the coil shoultl have flared connections if the runouts are soft copper. Otherwise unions or flanges are installed to facilitate servicing units. FIG. 35 - PINING LAYOUT FOR HOKIZONTAL MULTIPLE COILS (4 UNITS - 2 S HUT-OFF V ALVES) SHUT-OFF VALVE t GATE VALVE1 NOTES: 1. Though not shown, control valves (automatic or manual) may be required to control Row thru each unit. 2. A shut-off valve may be installed in the supply and return branch headers when headers serve 3 to .i units. 3. Supply and return runouts to the coil should have flared connections if the runouts are soft copper. Otherwise unions or flanges. are installed to facilitate servicing units. FIG. 36 - PIPING LAYOUT FOR HORIZONTAL M ULTIPLE COILS (3 UNITS - 6 SHUT-O FF V ALVES) SUPPLY RETURN NOTES: 1. Flange or union is located to allow condenser head removal. 2. With outlet at top, condenser will be flooded even though automatic control valve is in modulating position. 3. Check valve is required by most sanitary codes (city water). 4. Required for city water only. FIG. 38 - CONDENSER PIPING S YSTEM FOR A O NCE -THRU 1. The pressure drops thru the condensers are not always equal. 2. Water entering the branch line and leaving the run of tees seldom divides equally. 3. Workmanship in the installation can affect the pressure drop. To equalize the water flow thru each condenser, the pipe should be sized as follows: 1. Size the branches for a water flow of 6 fps minimum. The branch connections to each condenser should be identical. CIRCULATING PUMP REMOVE WELL AND BUSHING NOTES: 1. Flange or union is located to allow cooler head removal. 2. Gate valves shown may be eliminated in a close coupled system. Fro. 37 - PINING AT A W ATER COOLER FOG. 39 - ;\r:rrxNArE D RAIN CONNECTION 3-06 f'AI<'r 3. l'll'lK(; I)I:SI(;N WATER OUT (5-10 tPS1 THERMOMETER FOR BLOW OUT WELL AND GUSHING GLOW O U T T U B E S . TO REDUCING TEE’ NOTES: 1. Flange or union is located to allow condenser head removal. 2. Gate valves shown may ix eliminated in a close coupled system (except drain valve). 3. When water enters bottom of condenser, air will vent naturally thrn cooling tower sprays. If it is necessary to drop piping after leaving condenser, install air vent at high point of line Ijefore drop. See dotted line in figure. FIG. 40 - CONI)WSER PII~INC FOK A C OOLING TOWER HEADER-TOTAL GPM AT 3 FPS MAX. NOTES: 1. Thermometer wells are inserted in tees. Remove wells to blow omit coils. 2. A single water rcgnlating valve mnst be used as shown. If nntier capacity, install two valves in parallel and connect pressure tribe in liquid header. 3. \Vater supply in or return ant can be at any point in the headers. FIG. 41 - iLIL’LTIllLE PIPING 2. Size the tended approsimntely 12 in. beyond the last branch LO the condenser. Size the COOLING \vhTER CONNECTIONS P ARALLEL) h e a d e r lor t h e t o t a l requird water ciuantity lor all the condensers with a velocity OC n o t m o r e t h a n 3 E p s . T h e h e a d e r i s cs- 3. CONDENSEK (KEFRIGERANT water main supplying the header lor a velocity of 5 to 10 f p s w i t h 7 Ips a good average. The water main may enter the header :tt the end or at any point along the length ol the header. Care should be used so that crosses do not result. 4. Size the return branches, header ant1 main in the same manner as the supply. 5. Install 2 single water regulating valve in the main, rather than separate valves in the hnches to the condenser (Fig. fl). Cooling Tower FifiUr.efO illustrates a cooling tower ant1 condenser piped together. Since the cooling tower is an open piece ot’ equipment, this is 211 open piping system. II the conclcnscr nntl cooling tower are on the same level, a small suction head I’or the puml) exists. The strainer should bc installed on the discharge side oE the pump t o keep the suction side ol’ the pump ns close to atmospheric :IS possible. . It is often desirable to maintain a constant water temperature to the condenser. This is done by installing a bypass around the cooling tower. When the condenser is at the same level or above the coolIng tower, a three-way diverting valve is recommended in the bypass section (Fig. f 2). ,A three-way mixing valve is not recommended since it is on the pump suction side, and tends to create VRCUUIJI conditions rather than maintain atmospheric pressures. I;if:~e f3 illustrates the bypass layout when the condenser is below the level of the cooling tower. This particular piping diagram uses a two-way automatic control valve in the bypass line. The lriction drop thru the bypass is sized for the unbalanced static head in the cooling tower with maxiinuniwater flow thru the bypass. II multiple cooling towers arc to be connected, it is recommencletl that piping be clesigned such that the loss from the tower to the pump suction is approximately equal for each tower. F i g . 44 illustrates typical layouts I’or multiple cooling towers. Equ;llizing lines are usetl to maintain the same water level in each tower. Cl-I;\I”i’I<li 3-37 2. w.\~l‘I*:I< I’Il’Ir\‘(i COOLING TOWER HEADERS tir COOLING TOWER 4y/ HEADxr=-+T~ AIf- I /THREE -WAY DIVERTING VALVE COOLING UNBALANCED HEAD COOLING /- TOWER 4, F R O M CONDENSEI RETURN -BYPASS 1 FROM , CONDENSER TO PUMP SUCTION t ____j____i NOTES: kN O T E 2 .4 three-way diverting valve is used when the condenser is at the same level as or above the cooling tower. See Fig. f3 for piping layout when condenser is below the cooling tower. 2. :\ three-way mixing valve is not recommended at this point as it imposes additional head at the pump suction. 1. A two-way automatic control valve is used when the condenser is below the cooling tower. See Fig. 42 for piping layout when the condenser is at the same level as or above the cooling tower. 2. The friction loss from “A” to “B” includes the loss thru that section of the pipe and the loss thru the two-way automatic control valve. This friction loss should be designed for the unbalanced head of the cooling tower. 3. Locate the automatic control valve close to the cooling tower to prevent pump motor overload and tower sptll-over when valve is. in full open position. FIG. 42 - COOLING 71‘~w~~ PIPING FOR CONSTANT LEAV:NG W ATER TEMI’ERATURE (CONDENSER AND TOWER AT SAME LEVEL) FIG. 43 - COOLING TOWER PIPING FOR CONSTANT LEAVING W ATER TEMPERATURE (CONDENSER BELOW TOWER) 1NO’IT.S: I. NOT RECOMMENDED TWO COOLING RECOMMENDED TOWERS RECOMMENDED NOT RECOMMENDED THREE I’K;. 4-i - h’[LJLTIl’LE COOLING TOWERS COOLING 7‘OWEK I’II’ING l’;\l<‘l‘ 3. I’II’ING DESIGN 338 THERMOMETER \. GAGE - .- ._ _ i FLOODING HEADER SUPPLY FLOAT VALVE SITE ORAIN NOTE: See 46 lrnd 47 for typical piping when an air washer is used for the dehumidifying system (section ‘LA - A”). FIG. 45 -AIR WASHEK.PIHNC Air Washer The water piping layout for an air washer used Sor humidifying is presented in Fig. 45. When the pump and air washer are on the same level, there is usually a small suction head available for the pump. Therefore, if a strainer is required in the line, it should be located on the discharge side of the pump. Normally air washers have a permanent type screen at the suction connection to the washer to remove large size foreign matter. The drain line is connected to an open-site drain similar to those illustrated in Fig. 38 and 39. The drain arrangement should always be checked for compliance to local codes. The piping layout shows a shell and tube heater for the spray water. Occasionally heat is added by a steam ejector instead of by a normal heater. Chilled water is required if accomplish dehumidification, illustrate two typical methods chilled water supply. The plug the sprays are to Figures 46 rind 47 of connecting the cock in both dia- DRAIN TO GRAVITY GRAVITY RETURN TO SURGE TANK THREE-WAY DIVERTING CHILLED VALVE WATER RECIRCULATING PUMP ! c’ _ l/-i SECTION “.A”-“A” NOTE: Adjust plug cock so that full flow thru automatic control valve is approximately 90% of recirculating water design. FIG. 46 - .&R WASH!& PIPSNC USING CONTROL VALVE A .I‘HKE:E-WAY CI-I~\I”I‘ER 2 . W.\‘I‘El< 3-39 I’Il’IN(; grams is adjusted way diverting valve control valve (Fig. recirculating water so that full llow thru the three- (Fig. 46) and thru the automatic 47) is approximately 90% of the design quantity. ~;igures -fh’ and 49 are schematic sketches of multiplr air washers with gravity returns piped to the smic header. Sprayed Coil SECTION h typical layout for a sprayed coil piping system is shown in Fig. 50. The diagram shows a water heater which may be required for humidification. If a preheat coil is used, the water heater may be ‘r, eliminated. The drain line should be fitted with a gate valve rather than a globe valve since it is less likely to “A’‘-“A” NOTE: Adjust plug cock so that full Row thru diverting valve is approximately 90% of recirculating water design. ;. Lk5 - ,‘\IR WASILK I'II'ING USING CONTROL VALVE A become clogged with sediment. Pump Piping TWO-WAY ?‘he following items illustrated in I’&. 5f should be kept in mind when designing piping for a pump: 1. Keep the suction pipe short and direct. SLOPE GRAVITY RETURN LINES TOWARD RISER A I R 2. Increase the suction pipe size to at least one size larger than the pump inlet connection. ’ 3. Keep the suction pipe free from air pockets. 4. Use an eccentric type reducer at the pump suction nozzle to prevent air pockets in the V E N T , 0 RISER suction UNIT ‘I line. 4 x’: ENTER RISER AT SHARP ANGLE IN DIRECTION OF FLOW p: 4 1 \- L IG. 48 STRAINER ELEiATION --AIR WASHER RETURN CONNECTIONS DIFFERENTELEVATIONS AT /- GATE VALVE SPRAY WATER HEATER - t. <, - HEADER ENTER HEAD& AT SHARP ANGLE IN DIRECTION OF FLOW ‘1 - SLOPE HEADER TOWARD RISER GATE VALVE NO’IX: If no strainer is installed in this location, then a strainer is recommended on the pump discharge. FIG-N-AIR WASHER RETURN CONNECTIONS !k\MELEVEL AT THE FIG. Ed--SPRAYWATERCOIL WITH WATER HEATER , I’.\l<‘L‘ 3-40 3. I’II’IN<; DESIGN x_ GAGE / GATE TV A L V E @ I I USING TWO GAGES GATE . USING ONE GAGE -GAGE - GATE FK.5i VALVE OR PET COCK -PUMP SUCTION CON'NECTIONS ,I ” a 5. Never install a horizontal elbow at the pump inlet. hy horizontal elbow in the suction line should be at a lower elevation than the pump inlet nozzle. Where possible, a vertical elbow should lead into a pipe reducer at the pump inlet. If multiple pumpsare to be interconnected to the same hcadcr, piping connections are made as illustrated in Fig. 52. This method allows each pump FIG. 53 -GAGELOCATION . ATA PUMP to handle the same water quantity. Under partial load conditions and at reduced water flow or when one pump is out of the line, the pumps still handle equal water quantities. Figz~7e 53 illustrates two methods of locating pressure gages at the pump; one method uses two gages and the other uses one. The use of one gage has the advantage of always giving the correct pressure differential across the pump. Two gages may give an incorrect pressure differential if one or both are reading high or low. A pulsating damper located before the pressure gage is shown in Fig. 53. This is an inexpensive device for dampening pressure pulsations. The same result can be obtained by using a pigtail in the line as shown in the diagram. Expansion Tank Piping Figure 5f is a suggested piping layout for an open expansion tank. Piping is enlarged at the connection to the expansion tank. This permits air entrained or carried along with the water to separate I QUICK F I L L L I N E ,-AIR V E N T d SIGHT GLASS t. fT G AAT ET E b % VALVE ’ -To- i +I I-’ TRAP ENLARGED PORTION OF RETURN LINE TO PERMIT A I R S E P A R A T I O N ---, ‘ N O T E 2) RETURN LINE ,C AT LEAST 4d x.Ly EXPANSION LINE ( I f ” MIN ) FIG . E N L A R G E D T E E FOFi AIR SEPARATION ,-NORMAL LINE SIZE CIRCULATING PUMP NOTES: 1. Do not put any valve strainer or trap in the expansion line. 2. Enlarged portion of return line and enlarged tee are each two standard pipe sizes larger than return line. FIG. 54 - OPEN EXPANSION TANK PII’INC and be vented thru the tank. The expansion tank should be located at the pump suction side at the highest point in the system. Vaives, strainers and traps must be omitted from ‘~‘5 expansion line since these may be accidentally .led off or become plugged. ,-. I;ig~re 55 illustrates the piping diagram l‘or a closed tank. Drain Line 55 - CLOSED E XPANSION TANK PII~INC sure as in a blow-thru fan-coil unit. When the system is under negative pressure as in a draw-thru unit, the trap prevents water from hanging up in the drain pan. Figure 56 illustrates the trapping of a drain line from the drain pan. The length of the water seal or trap depends on the magnitude of the positive or negative pressure on the drain water. For instance, a Z-inch negative fan pressure requires a Z-inch water seal. Normally, under-the-window fan-coil units have the drip lyan subject to atmospheric conditions only and the drain line from these units is not trapped. The drain line runout for all systems is pitched to offset. the line friction. For a single unit the runout is piped to an open site drain. Local codes and regulations must be checked to determine proper piping practice for an open site drain. The runout is run full size corresponding to the drain pan connection size. Some applications have multiple units with the drain lines connected to a common header or riser. Piping Moisture that forms on the cooling coils must be collected and carried off as waste. On factory fabricated fan-coil units a drain pan is used to collect this moisture. For built-up systems the floor or base of the system (before and after the cooling coil) is used to gather the moisture. Since, under operating conditions, the drain water is subject to pressure conditions slightly above or slightly below atmospheric pressure, the line used to carry off this water must be trapped. This trap prevents conditioned air from entering the drain lint when the drain water is llnder nositivr nrc+ I- .---- - I---- P I T C H RUNOUT TO OFFSET LINE FRICTION m m UNIT \ TRAP FOR WATER SEAL FIG. 56 - P IPING FOR DRAIN P ANS 342 To size the header or riser, the amount of moisture that is expected to form niust be determined. This moisture and the available head is used to determine the pipe size from the friction chart for open piping systems. However, in no instance is the header or riser sized smaller than the drain pan connection size. Also, as required in all water How I’AKI‘ 3 . 1’II’ING DESIGN systems, pockets traps in the risers and mains must be vented to prevent water hangup. Each system should be investigated to determine the need for drainage fittings and cleanouts for traps. These are necessary when considerable sediment may occur in the drain pan. . 3-43 CHAPTER 3. REFRIGERANT PIPING GENERAL SYSTEM DESIGN This chapter includes that practical material quired for the design and layout of a refrigerant piping system at air conditioning temperature levels, using either Refrigerant 12, 22 or 500. APPLICATION CONSIDERATIONS A refrigerant piping system requires the same general design considerations as any Huid How system. However, there are additional factors that critically influence system design: 1. The system must be designed for minimum pressure drop since pressure losses decrease the thermal capacity and increase the power requirement in a refrigeration system. 2. The fluid being piped changes in state as it circulates. 3. Since lubricating oil is miscible with Refrigerants 12, 22 and 500, some provision must be made to: a. Minimize the accumulation of liquid refrigerant in the compressor crankcase. b. Return oil to the compressor at the same rate at which it leaves. Piping practices which accomplish these objectives are discussed in the following pages. CODE REGULATIONS System design should conform to all codes, la% and regulations applying at the site of an installation. addition the Safety Code for Mechanical Refrigeration (ASA-B9.1-1958) and the Code for Refrigeration Piping (ASA-B31.5-1962) are primarily drawn up as guides to safe practice and should also be adhered to. These two codes, as they apply ‘to refrigeration, are almost identical, and are the basis of most municipal and state codes. REFRIGERANT PIPING DESIGN DESIGN PRINCIPLES Objectives Refrigerant piping systems must be designed to accomplish the following: 1. Insure proper feed to evaporators. 2. Provide practical line sizes without excessive pressure drop. 3. Protect compressors by a. Preventing excessive lubricating oil from being trapped in the system. b. Minimizing the loss of lubricating oil from the compressor at all times. c. Preventing liquid refrigerant from entering the compressor during operation and shutdown. Friction Loss and Oil Return ’ In sizing refrigerant lines it is necessary to consider the optimum size with respect to economics, friction loss and oil return. From a cost standpoint it is desirable to select the line size as small as possible. Care must be taken, however, to select a line size that does not cause excessive suction and discharge line pressure drop since this may result in loss of compressor capacity and excessive hp/ton. Too small a line size may also cause excessive liquid line pressure drop. This can result in flashing of liquid refrigerant which causes faulty expansion valve operation. The effect of excessive suction and hot gas line pressure drop on compressor capacity and horsepower is illustrated in Table 16. TABLE 16-COMPRESSOR CAPACITY VS LINE PRESSURE DROP 42 F Evaporator Temperature COMPRESSOR SUCTION AND HOT GAS LINE PRESSURE DROP Capacity (%I No Line loss 2F Suction Line Loss 2F Hot Gas Line Loss 4F Suction Line Loss 4F Hot Gas Line Loss 100 95.7 90.4 92.2 96.8 1 Hp/Ton (%I 100 103.5 103.5 106.8 106.8 Pressure drop is kept to a minimum by optimum sizing of the lines with respect to economics, making sure that refrigerant line velocities are sufficient to entrain and carry oil along at all loading conditions. For Refrigerants 12, 22 and 500, consider the requirements for oil return up vertical risers. Pressure drop in liquid lines is not as critical as in suction and discharge lines. However, the pressure drop should not be so excessive as to cause gas formation in the liquid line or insufficient liquid pressure at the liquid feed device. A system should normally be designed so that the pressure drop in the liquid line is not greater than one to two degrees ’ PART 3-44 change in saturation temperature. In terms of pressure drop, this corresponds to about 1.8 to 3.8 psi for Refrigerant 12, 2.9 to 6 psi for Refrigerant 22, and 2.2 to 4.6 psi for Refrigerant 500. Friction pressure drop in the liquid line includes accessories such as solenoid valve, strainer, drier and hand valves, as well as the actual pipe and fittings from the receiver outlet to the refrigerant feed device at the evaporator. Pressure drop in the suction line means a loss in system capacity because it forces the compressor to operate at a lower suction pressure to maintain the desired evaporator temperature. Standard practice is to size the suction line for a pressure drop of approximately two degrees change in saturation temperature. In terms of pressure loss at 40 F suction temperature, this corresponds to about 1.8 psi for Refrigerant 12, 2.9 psi for Refrigerant 22, and 2.2 psi for Refrigerant 500. Where a reduction in pipe size is necessary to provide sufficient gas velocity to entrain oil upward in vertical risers at partial loads, a greater pressure drop is imposed at full load. To keep the total pressure drop within the desired limit, excessive riser loss can be offset by properly sizing the horizontal and “down” lines. It is important to minimize the pressure loss in hot gas lines because these losses can increase the required compressor horsepower and decrease the compressor capacity. It is usual practice not to exceed a pressure drop corresponding to one to two degrees change in saturation temperature. This is equal to about 1.8 to 3.8 psi for Refrigerant 12, 2.9 to 6 psi for Refrigerant 22, and 2.2 to 4.6 psi for Refrigerant 500. where h = loss of head in feet of fluid f = friction factor L = length of pipe in feet D = diameter of pipe in feet V = velocity in fps g = acceleration of gravity = 32.17 ft/sec/sec The friction factor depends on the roughness of pipe surface and the Reynolds number of the fluid. In this case the Reynolds number and the Moody chart are used to determine the friction factor. PIPING DESIGN Use of Pipe Sizing Charts The following procedure for sizing refrigerant piping is recommended: 1. Measure the length (in feet) of straight pipe. 2. Add 5070 to obtain a trial total equivalent length. 3. If other than a rated friction loss is desired, multiply the total equivalent length by the correction factor from the table following the appropriate pipe or tubing size chart. 4. If necessary, correct for suction and condensing temperatures. 5. Read pipe size from Charts 7 thru 21 to determine size of fittings. 6. Find equivalent length (in feet) of fittings an{ hand valves from Chapter 1 and add to the length of straight pipe (Step 1) to obtain the total equivalent length. 7. Correct as in Steps 3 and 4 if necessary. 8. Check pipe size. In some cases, particularly in liquid and suction lines, it may be necessary to find the actual pressure drop. To do this, use the procedure described in . Steps 9 thru 11: 9. Convert the friction drop (F from Step 3) to psi, using refrigerant tables or the tables in Part 4. 10. Find the pressure drop thru automatic valves and accessories from manufacturers’ catalogs. If these are given in equivalent feet, change to psi by multiplying by the ratio: step (9) step (6) REFRIGERANT PIPE SIZING Charts 7 thru 21 are used to select the proper steel pipe and copper tubing size for the refrigeration lines. They are based on the Darcy-Weisbach formula: 3. 11. Add Steps 9 and 10. In systems in which automatic valves and accessories may create a relatively high pressure drop, the line size can be increasedto minimize their effect on the system. Example I - Use of Pipe Sizing Charts Given: Refrigerant 12 system Load - 46 tons Equivalent length of piping - 65 ft Saturated suction - 30 F Condensing temperature - 100 F Type L copper tubing Find: Suction line size for pressure drop corresponding to 2 F. Actual pressure drop in terms of degrees F for size selected. CHAPTER 3. REFRIGERANT PIPING Solution: Liquid Subcooling See Chart 7. I. 345 Line sizes for 40 F saturated suction and 105 F condensing temperature are shown on Chart 7. Determine the correction factor for a 30 F suction temperature of 1.19 from table in notes following Chart 9. Where liquid subcooling is required, it is usually accomplished by one or both of the following arrangements: 1. A liquid suction heat interchanger (heat dissipates internally to suction gas). 2. Determine adjusted tons to be used in Chart 7 by multiplying correction factor in Step 1 by load in tons: 1.19 x 46 = 55 tons 3. Enter Chart 7 and project upward from 55 tons, to a 25/8 in. OD pipe size, then to a 31/, in. OD pipe size. At 25/s in. OD, a 2 F drop is obtained with 33 ft of pipe; at 3% in. OD a 2 F drop is obtained with 71 ft of pipe. Select a 31/s in. OD pipe to obtain less than a 2 F drop. 4. Use the following equation to determine actual pressure drop in terms of degrees F in the 31/8 in. OD pipe with a 46 ton load: Actual pressure drop equivalent ft of pipe = piping allowed for 2 F drop LINE The amount of liquid subcooling required may be determined by use of a nomograph, Chart 22 or by calculation. The following examples illustrate both methods. Example 2 - Liquid Subcooling from Nomograph . ’ 2 F ‘65 =71 X2=1.8F LIQUID 2. Liquid subcooling coils in evaporative condensers and air-cooled condensers (heat dissipates externally to atmosphere). DESIGN Refrigeration oil is sufficiently miscible with these refrigerants in the liquid phase to insure adequate mixing and oil return. Therefore low liquid velocities and traps in liquid lines’ do not pose oil return problems. The amount of liquid line pressure drop which can be tolerated is dependent on the number of degrees subcooling of the liquid. Usually this amounts to 2 F to 5 F as the liquid leaves the condenser. Liquid lines should not be sized for more than a 2 F drop under normal circumstances. In addition, liquid lines passing thru extremely warm spaces should be insulated. Given: Refrigerant 12 system Condensing temperature - 100 F (131.6 psia) Liquid line pressure drop (incl. liquid lift) - 29.9 psi Find: Amount of liquid subcooling in degrees F required to prevent flashing of liquid refrigerant. Solution: Use Chad 22. 1. Determine pressure at expansion valve: 131.6 - 29.9 = 101.7 psia 2. Draw line from point A (100 F cond (101.7 psia at expansion valve). 3. Draw line from point C (intersection of AB with line 2) thru point D (0% flash gas) to point E (intersection of CD with liquid subcooling line). 4. Liquid subcooling at point E = 18 F. Liquid subcooling required to prevent liquid fiashing = 18 F. >I_ ^, , I ..,. cl_ -. ‘on Drop and Static Head 4th an appreciable friction drop and/or a static head due to elevation of the liquid metering device above the condenser, it may be necessary to resort to some additional means of liquid subcooling to prevent flashing in the liquid line. Increasing the liquid line pipe size minimizes pipe friction and flashing due to friction drop. In large systems where the cost is warranted, a liquid pump may be used to overcome static head. An arrangement shown in Fig. 57 illustrates a method which may be used to overcome the effect of excessive flash gas caused by a high static head in the system. This arrangement does not prevent the forming of flash gas, but does offset the effect it might have on the operation of the evaporator and valves. temp) to point B .i ‘,.: ;. ,_. ^_&: “. FLOAT-ACTUATED VENT VALVE FIG. 57 - METHOD OF OVERCOMING ILL EFFECTS SYSTEM HIGHSTATICHEAD OF 3-46 PART 3. PIPING DESIGN CHART 7-SUCTION LINES-COPPER TUBING IEFR / 4o”/lo50 For Pressure Drop Corresponding to 2 F I 300 E -200 z WI50 E -I yoo w -1 80 a > 3 6 0 ,” 5 0 40 TONS OF REFRIGERATION CHART 8-HOT GAS LINES-COPPER TUBING For Pressure Drop Corresponding to 2 F y \ F 5200 \ \ I I i\l I 2 3 4 \ \, I \ \ IR:,? I / h4 \;l;‘i \ 5 ,,p,343\ \ NI II \\5,8 J4!& \ \ \ 1 1 \ \ \ \ E 150 ifi \ -1 + too \ \ \ 1L \ \ y \ \ \ 300 Y Y \. \ \I \ fi 1; t \ \ I. \ ‘\ I I = 8 0 z 3 6 0 ,” 5 0 5 6 8 IO TONS 20 30 40 50 60 OF REFRIGERATION 80 100 200 3 0 0 400 500 I (:H;\l”I‘I’K 3. KEI:KIC;EKI\N~~‘ 347 I’Il’lN(; CHART 9-LIQUID LINES-COPPER TUBING REFRIG. 12 For Pressure Drop Corresponding to 1 F 400/1 05O I-----. h w 18 I I x I\1 I I III1 I \n \ AP I I ,\I l\l I I \ \Illl \I \I I IUI \ \ \I \ 2 3 4 5 6 8 20 30 40 50 60 OF REFRIGERATION IO TONS Range of Chart 9: Saturated Suction Condensing Temperatures 80 -4OFto Temperatures I I 200 100 Y \ I\ 300 400 500 5OF 80 F to 120 F Pressure drop is given in equivalent degrees because of the general acceptance of this method of sizing. The corresponding pressure drop in psi may be determined by referring to the saturated refrigerant tables. To use Charts 7 and 8 for conditions other than 40 F saturated suction;, 105 F condensing, multiply the refrigeration load in tons factor below and use the product in reading the chart (S = Suction, HG = Hot Gas). -40 CO ND I’,P 1 -30 1 -20 SATURATED - 1 0 1 1 SUCTION 0 TONS S 90 100 110 120 130 140 150 160 HG 4.90 1.47 5.17 1.34 5.45 1.24 5.80 1.17 6 . 2 0 1.09 6 . 6 8 1.02 80 7.20 7.90 ) 8.70 S 3.04 4.30 1.30 4.55 1.21 4.81 1.12 3.20 1.28 2.54 1.23 2.00 3.33 1.19 2.69 1.13 2.10 3.56 1.09 2.83 1.04 2.20 3.78 1.00 3.00 .98 2.37 4.03 .95 3.21 .93 2.54 .98 5.91 .95 6.43 . 9 4/7 . 1 0 1.04 .98 .94 4.35 .91 4.74 .90 1 5.20 S HG 2.41 1.37 MULTIPLYING S HG 4.09 1.43 5.08 5.50 HG 1.41 .91 3.43 .88 3.74 . 8 71 4 . 0 6 TEMPERATURE 10 1 S 1.94 .88 2.71 .85 2.97 .84i 3.22 HG 1.34 S 1.60 1.22 1.68 1.12 ' 1.76 1.03 1.86 .97 1.96 .90 2.09 .86 2.24 .82 2.41 .81 12.62 by the (F) 30 1 20 1 40 1 50 FACTOR HG 1.32 S 1.29 HG 1.29 S 1.09 HG 1.26 1.19 1.35 1.17 1.12 1.14 1.10 1.41 1.08 1.18 1.05 1.01 1.48 1.00 1.23 .91 1.30 .85 1.39 .96 .89 .83 .94 1.58 . a 7 1.68 .83 1.80 .79 1.94 . 7 8[ 2 . 1 0 .80 1.50 .78 1.62 . 7 6 / 1.74 S .90 .94 .98 1.02 1.06 1.15 .79 1.22 . 7 6 1.31 . 7 3 1 1.41 HG 1.24 1.12 1.03 .95 .87 .81 .77 .73 S .77 .76 .80 .a4 .88 .96 1.03 1.11 .71 j 1 . 2 0 HG 1.22 1.10 1.01 .92 .85 .79 .75 .71 .69 NOTES: 1. To use suction and hot gas line charts for friction drop other than 2 F or liquid line charts for friction drop other than alent length by factor below and use product in reading chart. Friction Drop (F) Liquid Line 0.25 0.5 Hot Gas Line Suction Line 0.5 1 .o 4.0 2.0 Multiplier 2. Pipe sizer ore OD and ore for Type L copper tubing. .75 1 F, multiply equiv- 1.0 1.25 1.5 2.0 2.5 3.0 1.5 2.0 2.5 3.0 4.0 5.0 6.0 1.3 1.0 0.8 0.7 0.5 0.4 0.3 348 L PART ?). PIPING DESIGN CHART IO-SUCTION LINES-STEEL PIPE For Pressure Drop Corresponding to 2 F SCHEDULE 40 \ \ 10 \I III \ I I I - - I 2 3 4 5 6 8 x Y I \ IO 20 30 40 5060 80 100 ‘( I 200 I\I 300 400500 TONS OF REFRIGERATION . CHART II-HOT GAS LINES-STEEL PIPE 1REFRIG. 12 / For Pressure Drop Corresponding to 2 F SCHEDULE 40 h 400 300 f k200 E p0 w -I I- 100 2 80 = 5 60 ," 50 TONS OF REFRIGERATION I I (‘:H,\P-I‘ER :1. REI;RI(;ER~\NI‘ I’IPINC, 349 CHART 12-LIQUID LINES-STEEL PIPE For Pressure Drop Corresponding to 1 F I- \. y 80 3 + S C H E D U L E 80 B \I h \ SCHEDULE 40 I i\I I I\ I\ \ \ \ I. I \ I I I I\ I hI I Ihlll\ h L I 2 3 4 5 6 8 IO 20 TONS Range of Chart 12: OF 30 40 5060 80 100 200 300 400 500 REFRIGERATION Saturated Suction Temperatures Condensing Temperatures -40Fto 5OF 8OFto12OF Pressure drop is given in equivalent degrees because of the general acceptance of this method of siring. The corresponding pressure drop in psi be determined by referring to the saturated refrigerant tables. may To use Charts 10 and 11 for conditions other than 40 F saturated suction, 105 F condensing, multiply the refrigeration load in tons by the factor below and use the product in reading the chart (S = S u c t i o n , H G = H o t G a s ) . COND TEMP -2 -40 1 1 -30 -20 j SATURATED SUCTION TEMPERATURE (F) -10 / 0 10 ( 20 1 30 1 40 / 50 TONS MULTIPLYING FACTOR S HG 1 S HG 1 S HG / S 2.26 2.36 2.46 HG 1.35 1.20 1.12 1.05 .98 .91 .87 .85 .84 1 s 1.83 1.92 2.00 2.13 2.25 2.39 2.58 2.78 3.04 1.32 1.20 1.11 1.02 .95 .89 1 S 1.54 1.59 1.69 1.78 1.89 2.00 .83 .82 .81 2.16 2.32 2.52 HG HG I s 1.29 1.29 1.18 1.34 1.08 1.39 1 . 0 0 1.47 .93 1.56 .86 1.65 .80 .79 .78 1.78 1.92 2.08 HG ( S 1.27 1 1 . 0 7 1 . 1 6 1.12 1.08 1.17 1 . 0 0 1.23 .92 1.29 .84 1.38 .80 .78 .76 1.46 1.59 1.72 HG 1 S 1.24 .90 1.13 .94 1.05 .98 .97 1.02 .89 1.06 .82 1.13 .79 1.20 .76 1.30 .73 1.41 HG / S 1.22 .78 1.11 .79 1.02 .81 .94 .87 .87 .93 .81 .99 .77 1.06 .73 1.13 ’ .71 1.23 HG 1.20 1.09 1.00 .92 .85 .79 .75 .71 .69 NOTES: 1 . T o use s u c t i o n a n d h o t g a s l i n e c h a r t s f o r f r i c t i o n d r o p o t h e r t h a n 2 F o r l i q u i d l i n e c h a r t s f o r f r i c t i o n d r o p o t h e r t h a n 1 F , m u l t i p l y e q u i v a l e n t l e n g t h b y factor b e l o w a n d u s e p r o d u c t i n r e a d i n g c h a r t . liquid F r i c t i o n D r o p (F) Line Hot Gas Line Suction Line Multiplier 2. Pipe sizes are nominal and ore for steel pipe. 0.25 0.5 .75 1 .o 1.25 1.5 4.0 5.0 6.0 0.5 0.4 0.3 0.5 1.0 1.5 2.0 2.5 3.0 4.0 2.0 1.3 1.0 0.8 0.7 2.0 2.5 3.0 3-50 * PAR?’ 3. PIPING DESIGN i CHART 13-SUCTION LINES-COPPER TUBING For Pressure Drop Corresponding to 2 F i\l ill I ! I\ 500 400 ! I 300 IL - 2 0 0 $50 w” -1 blO0 w’ 8 0 ;: > 5 6 0 : 5 0 40 3c 2c I 7 3 8 56 4 IO 3 0 4 0 5 0 6 0 20 TONS OF REFRIGERATION 8 0 100 200 300 400 500 mufirsT ‘” “fiv GAS LINES--COPPER TUBING Drop Corresponding to 2 F 4 150 I ‘i i Ii WiiMi \ I 11x1 \ \ \ \ 5 I I III\ 6 8 IO 20 30 40 Y \ \ \ ‘, b 50 60 T O N S O F R E F R I GERATION 8 0 100 200 300 4 0 0 5 0 0 (‘;HAI”I‘ER 3. KEI’KI<;ER,\N’I‘ PIPING ,341 CHART 15-LIQUID LINES-COPPER TUBING For Pressure Drop Corresponding to 1 F 5 0 0 \ I I\l I \ I i I\I 4 0 0 3 0 0 i= !!z 2 0 0 z 150 i5 J 100 s w 8 0 2 ; 6 0 c 5 0 \ \, \ \ \, 3 0 I I P u 4 0 \ \, 2 3 4 5 6 8 \, \ \ \ \ , ._ ~ , - 2 0 \ \. \, \ IO 2 0 3 0 4 0 5 0 6 0 8 0 100 2 0 0 ” \, TONS OF REFIGERATION R&ge of Chart 15: Saturated Suction Condensing Temperatures -4OFto 5 0 F Temperatures 8OFto12OF Pressure drop is given in equivalent degrees because of the general acceptance of this method of sizing. The corresponding pressure drop in psi may be determined by referring to the saturated refrigerant tables. To use Charts 13 and 14 for conditions other than 40 F saturated suction, 105 F condensing, multiply refrigeration load in tons by factor below and use product in reading chart. (S = Suction-HG = Hot Gas) SATURATED SUCTION TEMPERATURE tFi CONDENSING - 4 0 1 - 3 0 / - 2 0 I - 1 0 ‘WPERATURE / 0 TON F) S 80 90 100 110 120 15.4 15.8 6.1 6.5 6.9 HG / S 1.7 / 4.3 1.6 t 4.4 1.4 4.8 1.4 5.1 1.3 5.4 HG / S 1.6 / 3.4 1.5 1 3.6 1.4 3.7 1.3 4.0 1.2 4.2 HG S HG 1.6 j 2.7 1.5 1 2.9 1.3 3.0 1.2 3.1 1.2 3.3 / 1.5 / j 10 MULTIPLYING S HG I S I 20 1 30 j 40 / 50 FACTOR HG j 2.1 1.4 1 1.7 1.4 1.4 1 2.3 1.4 1 1.8 1.3 I 1.3 ! 2.4 1.2 ’ 1.9 1.2 1.2 2.5 1.2 I 2.0 1.1 1.1 2.6 1.1 2.1 1.0 / S HG / 1.4 1.4 1.5 1.2 / 1.5 1.1 1.6 1.1 1.7 1.0 / S HG / S 1 1.1 1.3 I 0.9 1.2 1.2 1 0.9 1.1 1.2 1.1 1.0 1.3 1.0 1.0 1.3 0.9 1.1 HG 1 S HG 1.2 / 0.8 / 0.8 1.1 1.0 0.8 1.0 0.9 0.9 0.9 1.2 1.0 0.9 0.9 NOTES: 1. To use suction and hot gor line charts for friction drop other than 2 F or liquid line chart for friction drop other than 1 F, multiply equivalent length by factor below and use product in reodinq chart. Friction Drop (F) I Liquid Line 0.25 0.5 Hot Gas Line Suction Line 0.5 1.0 1.5 4.0 2.0 1.3 Multiplier 2. Pipe sizes ore OD and ore for Type L copper tubing. .75 1 .o 1 ( 3 0 0 4 0 0 5 0 0 1.25 1.5 2.0 2.5 3.0 2.0 2.5 3.0 4.0 5.0 6.0 1 .o 0.8 0.7 0.5 0.4 0.3 PART 3. 3-52 PIPING DESIGN \ CHART 16-SUCTION LINES-COPPER TUBING -l REFRIG. 22 For Pressure Drop Corresponding to 2 F 4o”/lo50 500 400 300 c k 200 I I - 150 z W -I F- 100 = 8 0 : 3 60 w” 50 4 0 30 20 2 I 500 3 rF 4 5 6 8 IO 30 40 50 60 T O N S OF?fEFR IGERATION 80 100 200 300 400 500 CHART 17-HOT GAS LINES-CdPPER TUBING For Pressure Drop Corresponding to 2 F 400 300 c k200 E (3 1 5 0 e -I c Inn 5 -1 a > 5 E . -80 I 6 0 50 I I\l \I4 i\ I \I I 4 0 30 20 2 3 4 5 6 8 IO 20 30 40 50 60 TONS OF REFRIGERATION 8 0 100 200 300 400 500 CHAPTER 3. REFRIGERANT I'II'ING 3-53 CHART 18-LIQUID LINES-COPPER TUBING For Pressure Drop Corresponding to 1 F 500 \ 400 300 \ \ \I 1 I I\II I I I N IlIla h I\ I I I\1 !\I I \i \i I I I Ihl I 80100 200 I I I\1 Xl I Y I i I- 100 = 80 z 5 60 2 3 4 5 6 8 IO 20 TONS kange of Chart 18: OF Saturated 30 Suction Condensing 40 5060 300 400500 REFRIGERATION Temperatures -4OFto Temperatures 5OF 80 F to 120 F Pressure drop is given in equivalent degrees because of the general acceptance of this method of siring. The corresponding pressure drop in psi may be determined by referring to the saturated refrigerant tables. To use Charts 16 and 17 for conditions other than 40 F saturated suction, 105 F condensing, multiply the refrigeration load in tons by the factor below and use the product in reading the chart (S = Suction, HG = Hot Gas). SATURATED COND TEMP _. (F) - 4 0 1 - 3 0 1 - 2 0 SUCTION T O N S MULTJP s HG s HG 80 90 100 4.65 4.83 5.12 1.40 1.27 1.18 3.70 3.87 4.04 1.38 1.24 1.16 110 120 130 5.42 1.08 5.75 1.01 6.20 .95 4.28 1.06 4.55 1.00 4.88 .94 TEMPERATURE (F) - 1 0 ‘INti FACT( S HG s 3.39 3.60 1.04 2.70 .98 2 . 8 5 1.02 2.20 .95 2 . 3 6 1.01 .95 1.80 1.00 1.89 .91 2.01 .82 1.45 1.53 1.63 .98 .91 .83 2.14 2.27 2.46 1.73 1.85 2.02 .78 .73 .68 .77 .73 .70 s HG 1.54 1.29 1.27 1.28 zY+&ij’ 1 . 6 1‘O1 . 1 7 ’ 1 . 3 32o1 . 1 5 ’ 1.70 1.08 1.39 1.06 3o 1.21 1.28 1.36 ’ 4o .97 1 . 0 1 .90 1 . 0 7 .83 1 . 1 3 ’ .95 .87 .80 HG .81 1.23 .845o 1.10 .88 1.01 .92 .96 1 .oo .93 .87 .81 1.07 1.16 1.27 .75 .69 .64 NOTES: 1. To use suction and hot gas line charts for friction drop other than 2 F or liquid line charts for friction drop other than 1 F, multiply equivalent length by factor below and use product in reading chart. Friction Drop (F) Liquid Line 0.25 Hot Gas Line Suction Line 0.5 1 .o 4.0 2.0 Multiplier 2. Pipe sizer are OD and are for Type L copper tubing. 0.5 .75 1 .o 1.25 1.5 2.0 2.5 3.0 1.5 2.0 2.5 3.0 4.0 5.0 6.0 1.3 1 .o 0.8 0.7 0.5 0.4 0.3 PAR?’ S. 3-54 PIPING DESIGN CHART 19-SUCTION LINES-STEEL PIPE For Pressure Drop Corresponding to 2 F SCHEDULE 40 500 400 300 f *. s .LUU . .a I I \I I z Id” W -I 5 100 W OF. 0 ” . 60 50 40 00 TONS OF REFRIGERATION . CHART 20-HOT GAS LINES-STEEL PIPE REFRIG. 22 r 500’ 4o”/1 05O - For Pressure Drop Corresponding to 3 I \ \ \ \ L I\\ L G - 200 I Ilhl I I II\ \I I IIA \I \ \ I I\ I\,,,, 400 300 SCHEDULE 40 v I I\1 \ I.,\ R I \ l\I 2 F \I I l\lII\ [\ I lYll\ I\ I\I \ \I R \ ) \ II ..\I4 I I I \3’/3 j\3” \ I\ $150 z W -I + 100 z w l2f-l 40 t30 I I I’, t- 20 lIzI E !i I l\I 4 5 6 I III 6 \ nrr41\ 2’0 TONS . I “lil‘k \’ I IO I OF 30 ;o 50 60 REFRIGERATION 1\1II 60 100 \a I, I. 200 ‘i I CHART 21-LIQUID LINES-STEEL PIPE REFRIG. 22 For Pressure Drop Corresponding to 1 F I _ S C 20 H 2 E D U L 4 5 3 E 6 8 0 8 4o”/lo50 r-7 I--- SCHEDULE 40 -1 d I IO 20 TONS Range of Chart 21: OF 30 40 5060 80 100 200 I\ 300 400; 00 REFRIGERATION -4OFto Saturated Suction Temperatures Condensing Temperatures 50F 8 0 F to 120 F Pressure drop is given in equivalent degrees because of the general acceptance of this method of sizing. be determined by referring to the saturated refrigerant tablet. The corresponding pressure drop in psi may To use Charts 19 and 20 for conditions other than 40 F saturated suction, 105 F condensing, multiply the refrigeration load in tons by the factor below and use the product in reading the chart (S = S u c t i o n , H G = H o t G a s ) . 3ND _ ‘4P . ) -40 - I 80 90 100 110 120 130 S 4.40 4.60 4.88 5.19 5.54 5.92 HG 1.38 1.26 1.18 1.09 1.00 .95 140 150 160 6.35 6.88 7.50 .90 .86 .85 -30 1 S HG 3.53 1.36 3.71 1.22 3.91 1.14 4.14 1.06 4.40 1 .oo 4.71 .94 5.05 .90 5.45 .86 5.95 .83 -20 S 2.80 2.94 3.10 3.26 3.44 3.68 3.96 4.27 4.65 HG 1.33 1.23 1.13 1.05 .99 .92 .87 .82 .80 1 SATURATED SUCTIOk 4 T E M P E R A T U R E (F) 10 I 20 -10 1 0 TONS MULTIPLYING FACTOR S HG 2.28 1.31 2.40 1.20 2.52 1.10 2.66 1.01 2.81 .94 3.00 .87 S 1.86 1.93 2.01 2.13 2.26 2.40 HG 1.29 1.15 1.07 1.00 .93 .86 S 1.54 1.60 1.68 1.77 1.87 2.00 HG 1.27 1.17 1.07 1.00 .93 .85 3.20 3.43 3.75 2.58 2.75 3.01 .80 .77 .76 2.14 2.27 2.46 .79 .75 .74 .92 .80 .78 S 1.27 1.33 1.39 1.45 1.54 1.64 1.75 1.86 2.01 30 I HG 1.26 1.15 1.05 .98 .91 .85 .79 .74 .72 S 1.06 1.10 1.15 1.21 1.28 1.37 1.46 1.54 1.67 I HG 1.24 1.13 1.03 .95 .88 .81 .77 .73 .71 40 I 50 S .90 .93 .97 HG 1.22 1.12 1.02 6 ~I< S .81 .85 .89 HG 1.21 1.11 1 .Ol 1.00 1.05 1.10 1.18 1.28 1.38 .95. .88 .81 .75 .71 .69 - .93 .96 1.00 1.08 1.17 1.27 .93 .85 .80 .75 .70 .68 NOTES: 1. To use suction and hot gas line charts for friction drop other than 2 F or liquid line charts for friction drop other than 1 F, multiply equivalent length by factor below and use product in reading chart. Liquid F r i c t i o n D r o p (F) Line Hot Gas Line Suction Line Multiplier 2. Pipe sizes ore nominal and ore for steel pipe. 0.25 0.5 .75 1 .o 1.25 1.5 2.0 2.5 3.0 0.5 1 .o I .5 2.0 2.5 3.0 4.0 5.0 6.0 4.0 2.0 1.3 I .o 0.8 0.7 0.5 0.4 0.3 CHART 22-SUBCOOLING TO COMPENSATE FOR LIQUID LINE PRESSURE DROP -51 -4 -4, -3 -3 ” -2 -2 : ? -1’ -1, 00 a ,- a ,- REFRIGERANT 22 . REFRIGERANT Example 3 - Liquid Subcooling by Calculafion Given: Refrigerant 12 system Condensing temperature - 100 F Liquid lift - 35 ft Piping friction loss - 3 psi Losses thru valves and accessories - i:i phi Find: Amount of liquid sulxooling liquid refrigerant. ‘I‘0t:ll pressure loss in liquid line 2. Condensing pressure at 100 F Pressure loss in liquid line = 116.9 psig = 29.9 psi Net pressure at liquid feed valve = wpsig 3. Saturation temperature at 57 psig = 82 F (from refrigerant property tables) 4. required to prevent flashing of Solution: I. Pressure loss due to pipe friction I’ressurc loss due to valves and accessories Pressure loss due to 35 ft liquid lift = 3.‘,/ I.t)# 500 Subcooling required = condensing temp - saturation temp at 87 psig = 100 - 82 = 18 F L i q u i d suhcooling required to prevent liquid Hashing = 18 F. = 3.0 psi = 7.4 psi = 19.5 psi = L”3.1, *.\t to 1.8 of normal liquid temperatures the static pressure loss due elevation at the top of a liquid lift is one psi for every ft of Refrigerant 12, 2.0 ft of Refrigerant 22, and 2.1 ft Refrigerant 500. Ltl.41”I‘ER 3. Sizing REI:RIGER~\N’I‘ PIPING of Condenser to Receiver Lines (Condensate Lines) piping from it conclenser to a receiver is ru*l out horizontally (same siLc as the condcnscr outlet connection) to allow for drainage of the contlenser. It is then droppctl vertically a sufficient distance to allow a liquid licad in the line to overcome line friction losses. Additional head is required for coil condensers where the receiver is vented to the inlet of the coil. ‘I‘his additional head is equivalent to the pressure drop across the condenser coil. The condensate lint is then run horizontally to the receiver. Tc~Dle I7 shows recommended sizes of the condensate line between the bottom of the liquid leg and the receiver. Licluid SUCTiON TABLE 17-CONDENSATE CONDENSER line LINE DESIGN ‘;uction lines are the most critical from a design Idpoint. The suction line must be designed to return oil from the evaporator to the compressor under minimum load conditions. Oil which leaves the compressor and readily passes thru the liquid supply lines to the evaporators is almost completely separated from the refrigerant vapor. In the evaporator a distillation process occurs and continues until an equihbrium point is reached. The result is a mixture of oil and refrigerant rich in liquid refrigerant. Therefore the mixture which is separated from the refrigerant vapor can be returned to the compressor only by entrainment with the returning gas. Oil entrainment with the return gas in a horizontal line is readily accomplished with normal design velocities. Therefore horizontal lines can and should be run “dead” level. _ ,tion 3-457 Risers i\/lost refrigeration piping systems contain a suetion riser. Oil circulating in the system can be returned up the riser only by entrainment with the returning gas. Oil returning lip a riser creeps up the inner surface of the pipe. Whether the oil moves up the inner surface is dependent upon the mass velocity of the gas at the wall surface. The larger the pipe diameter, the greater the velocity required at the center of the pipe to maintain a given velocity at the wall surface. Tables 18, 19 and 20 show the minimum tonnages required to insure oil return in upward flow suction risers and the friction drop in the risers in degrees F per 100 ft equivalent length. Vertical risers should, therefore, be given special analysis and should be sized for velocities that assure TO LINE SIZING RECEIVER (Based on Type L Copper Tubing) CONDENSATE LINE SIZE (OD. In.) 1% w ‘/r REFRIGERATION, MAX. TONS Refrigerant Refrigerant Refrigerant 22 500 12’ 1.2 2.3 6.4 I I 1 1.4 2.5 7.7 1 1.2 2.4 6.0 “X” MIN.* 8” 1 *This is the minimum elevation required between CI condenser coil outlet and a receiver inlet for the total load when receiver is vented to coil outlet header (bared on IO ft of horizontal pipe, 1 valve and 2 elbows). oil return at minimum load. A riser selected 011 this basis may be smaller in diameter than its branch or than the suction main proper and, thcrcforc, a relatively higher pressure drop may occur in the riser. This penalty should be taken into account in finding the total suction line pressure drop. The horizontal lines should bc sized to keep the total pressure drop within practical limits. Because modern compressors have capacity reduction features, it is often difficult to maintain the gas velocities required to return oil upward in vertical suction risers. When the suction riser is sized to permit oil return at the minimum operating capacity of the system, the pressure drop in this portion of the line may be too great when operating at full load. If a correctly sized suction riser imposes too great a pressure drop at full load, a double suction riser sl~ould be med (Fig. 58). TO COMPRESSOR SUCTION LINE TO COMPRESSOR C ELLS !\ --m $00sm. ELLS METHOD “A” ,U-BEND OR 2 ELLS u METHOD ‘8” FIG. 58 - DOIJBLE SUCTION RISER CONSTRUCTION l'.\l<'I‘ 3-58 3. I'II'lN(; I)I:SIGN \ TABLE 18-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP SUCTlON RISERS S T E E L P I P E - S T A N D A R D W E I G H T ( S C H E D U L E 40) T = Tons of refrigeration. F = Friction Drop, degrees F per 100 ft equivalent length at tons shown. FRICTION DROP MULTIPLIER 3.5 7.0 12.0 18.0 ‘(X/2001 1.8 x 3 . 5 300 400 500 X ‘Solve this equation to determine the friction drop multiplier for any ratio of full load to min. load, tons. . TABLE 19-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP SUCTION RISERS Refrigerant COPPER Pipe OD Area, Sq In. SuctTemp c20 +40 1% .484 .825 T -40 -20 0 % F .45 T 1.6 1% 1.256 F .88 T 1.5 1 .56 1.0 11.1 1 .66 .6 Il.3 1.77 .A II.5 1 .89 .2 Il.8 F T 1.5 1.4 .9 ( 1.8 .5 1 2.2 .3 12.5 .2 1 3.0 2% 3.094 1% 1.780 F T 22 TUBING-TYPE 1 2% 4.770 F T 3% 6.812 F T 3% 9.213 F T Area, Sq In. ?4 1 I .533 I .8 / .5 1 .3 t .2 1 t -51 2.0 ._ -. _ -70 -- I +20 f40 AR 17 .-...7A .7 .87 I .o 1% I 1.425 FIT I -40.- 0 1 1 .864 Suet T e m p , T - 1 I FIT FIT 4.6 1.2 8.0 1.2 12.3 1.1 18.1 1.1 125.0 1.0 143.6 .6 1 21.9 .6 1 30.6 .6 1 53.0 2.8 .8 1 5.6 .7 1 9.7 .7 1 15.1 3 . 3 .5 I 6 . 7 .4 1 1 1 . 4 .4 I 17.9 .4 I 26.2 .4 3 6 . 3 .4 6 3 . 3 3.9 .31 7.8 .3113.5 .3/20.8 .2/30.8 .2 4 2 . 7 .2 7 4 . 1 A.6 .2 1 9 . 1 .2 1 1 5 . 8 .2 1 2 4 . 3 .2 135.4 .l 4 9 . 4 .I 86.5 1% I 1 2.036 1 FIT 1 5% 1 18.67 2.3 1.3 STEEL PIPE-STANDARD WEIGHT (SCHEDULE IPS 1 4% 1 11.97 FIT 2 1 3.355 FIT 1 2% 1 4.788 ( 3 1 7.393 FIT 1 8% 1 46.85 FIT FIT .3 .2 .I 99.5 116.2 135.6 1 4 1 12.73 FIT 1 5 1 20.01 FIT FIT FIT I.9 1 1.8 I.6 1 2 . 7 1 . 6 1 5 . 0 1 . 5 1 7 . 9 1 . 4 1 13.5 1 . 3 1 19.4 1 . 3 1 2 6 . 7 1 . 2 1 4 7 . 2 1.21 74.3 I l. l . 1 1. I33 97 9 1 1 6--.6 .8 1 2 3 . 9 .7 1 3 3 . 0 .7 1 5 8 . 2 .71 9 2 . 1 -.- 10 ..- I33 - .. 91 A7 -.. 91 . 1 .A .7 2 . 7 .6 4 . 0 .6 7 . 4 .5 11.5 .5 1 9 . 9 .5 28.5 .5 3 9 . 2 .4 6 8 . 7 .4 110.0 T = Tons of refrigeration. F 1.6 1.8 .4 .3 3.2 3.6 .4 .2 4.7 5.3 .3 .2 8.7 10.0 .3 .2 13.6 15.7 .3 .2 23.4 26.4 .3 .2 33.6 38.6 .3 .2 46.0 53.0 = Friction Drop, degrees F per 100 ft equivalent length ot tons shown. FRICTION DROP MULTIPLIER 300 400 500 X .3 199.3 .2 2 3 4 . 0 .l 2 7 2 . 0 ) 6 1 28.99 1 -93. .5 .3 F .91 6 8 . 5 .9)137.1 .8 .61 83.6 .5 1167.0 .5 .3 .2 .l 40) ( 3’% 1 9.89 FIT 1 6’/0 1 26.83 3.5 7.0 12.0 18.0 *(x/200)‘.8 x 3.5 *Solve this equation to determine the friction drop multiplier for ratio of full load to min. load. tons. any .3 .2 80.6 93.0 .3 129.1 .2 148.0 FIT 1 1 8 50.0 F 1 . 1 1148.1 1.0 .7 1182.4 .6 . A 2 1 2 . 5 .4 .3 2 5 4 . 5 .2 292.0 .2 .l TABLE 20-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP SUCTION RISERS Refrigerant 500 Piaa no I r/, I l’/n 1 .52 1% . I, IPS ?4 Area. Sa In. 1 .533 -20 I 1.8 ) .96 I C O P P E R TUBING-TYPE 1 I 2% I 2% I 3% 1% I 3% I 4 STEEL PIPE-STANDARD WEIGHT (SCHEDULE 40) . I, ^.I 1 I c I I 1.6 1 1.9 1.5 ( 2.6 1.4 / 5.2 1.3 i 8.1 1.3 ) 13.9 1.2 120.0 1.1 1 27 = T o n s bf r e f r i g e r a t i o n . F = F r i c t i o n D r o p , d e g r e e s F p e r 1 0 0 f t e q u i v a l e n t l e n g t h a t t o n s s h o w n . FRICTION DROP MULTIPLIER 300 400 500 X 3.5 7.0 12.0 18.0 *(x/200)‘.8 x 3.5 *Solve this equation to determine the friction drop multiplier for ratio of full load to min. load, tons. Double Suction Risers There are applications in which single suction risers may bc sized for oil return at minimum load without serious penalty at design load. Where single compressors with capacity control are wised, minimum capacity corresponds to the compressor capacity at its minimum displacement. The maximum 3 minimum displacement ratio is usually three or .ur to one, depending on compressor six. The compressor capacity at minimum displaccmerit shoulcl be taken at an arbitrary figure of approximately 20 I; suction and not the design suction temperature for air conditioning applications. I\Qere multipie compressors are interconnected and controlled so that one or more may shut down while another continues to operate, the ratio of maximum to minimum tlisplacemcnt becomes much greater. In this case a double suction riser may be necessary for good operating economy at design load. The sizing and operation of a double suction riscl is described as follows: 1. In Fig. 58 the minimum load riser indicated by ‘-1 is sized so that it returns oil at the minimum possible load. any 2. The second riser B which is usually larger than riser A is sized so that the parallel pressure drop thru both risers at full loacl is satisfactory, providing this assures oil return at full load. 3. A trap is introduced between the two risers as shown in Fig. 58. During partial load operation when the gas velocity is not sufficient to return oil through both risers, the trap gradually fills with oil until the second riser B is sealed off. When this occurs, the gas travels up riser i-I only and has enough velocity to carry oil along with it back into the horizontal suction main. The fittings at the bottom of the riser must be close coupled so that the oil holding capacity of the trap is limited to a minimum. If this is n o t done, the t r a p cm accumulate cnougl~ oil on partial load operation to seriously lower the compressor crnnkcase oil level. Also, larger slug-backs of oil to the compressor occur when the trap clears out on increased load operation. Fig. TS shows that the larger riser 11 forms an inverted loop and enters the hori/ontal suction lint from the top. The purpose of rhis loop is to prevent oil drainage into this riser 1\JhiCll may be “iclle” during partial load operation. Example 4 - Determination of Riser Size Given: Refrigerant 12 system Type L copper tubing Condensing temperature - 105 F Suction temperature - 40 F Height of riser - 10 ft Equivalent length - 22 ft (10 ft pipe f 2 ells) Full load - 98.5 tons Minimum load - 8.1 tons Two of 16 cylinders operating at 20 F 2 compressors, 8 cylinders each. Find: Size of riser Suction line pressure drop Solution: 1. From Table 18 for 8.1 tons minimum load (20 F suction temperature) read 21/s in. OD tubing and .5 F minimum load pressure drop per 100 ft equivalent length of pipe. 2. Calculate minimum load pressure drop: Pressure drop for 5.6 ton =&- X 22 = .11 F = 145% of 5.6 tons x 3.5 = 1.96 Minimum load pressure drop = .11 X 1.96 = .22 F 3. Determine full load pressure drop: Full load = g = 1750% of minimum load Full load pressure drop = .11 X 174 = 19.2 F This full load pressure drop of 19.2 F together with a drop for the remainder of the suction line of approximately 1.5 F results in a 20.7 F total suction line drop. This is obviously too large and consequently a double suction riser must be used. 4. Determine size of smaller riser (same as in Step 1 for 21/, in. OD) and large riser for suitable pressure drop with a total load of 98.5 tons divided between the two risers. riser at 40 F suction temperature. Therefore, a 21/s in. OD plus a 41/~ in. OD pipe are capable of returning oil at maximum load. DISCHARGE (HOT GAS) LINE DESIGN The hot gas line should be sized for a practical drop. The effect of pressure drop is shown in Table 16, page 43. pressure Discharge Risers Even though a low loss is desired in the hot gas line, the line should be sized so that refrigerant gas velocities are able to carry along entrained oil. In the usual application this is not a problem. However, where multiple compressors are used with capacity control, hot gas risers must be sized to carry oil at minimum loading. Tables 21 and 22 show the minimum tonnages required to insure oil return in upward flow discharge risers. Friction drop in the risers in degrees F per 100 ft equivalent length is also included. Double Discharge Risers Sometimes in installations of multiple compressors having capacity control a vertical hot gas line sized to entrain oil at minimum load has an excessive pressure drop at maximum load. In such a case a double gas riser may be used in the same manner as it is used in a suction line. Fig. 59 shows the double riser principle applied to a hot gas line. Sizing of double hot gas risers is made in the same manner as double suction risers described earlier. REFRIGERANT CHARGE Table 23 is used to determine the piping system refrigerant charge required. The system charge should be equal to the sum of the charges in the refrigerant lines, compressor, evaporator, condenser and receiver (minimum operating charge). Let pressure drop = .5 F. Corrected equivalent length = equivalent length X correction factor (notes under Chart 9) = 22 X 4 = 88 ft (10 ft of pipe and 2 ells) Enter Chart 7 with a 21/8 in. OD pipe and an equivalent length of 88 ft; the capacity is 18.2 tons. “c” /lJl ‘A” “B* ALTERNATE -WHERE “6’ IS SMALLER THAN “C” The capacity for 41/, in. OD pipe at an equivalent length of 88 ft is 102 tons. Therefore, a small riser of 21/, in. OD and a large riser of 41/8 in. OD in parallel carry a load of 120.2 tons at a pressure drop of .5 F. However, since the load is only 98.5 tons and a .5 pressure drop is obtained with 125 ft of pipe (21/s in. pipe, 15 tons; 41/s in. pipe, 84 tons), the actual pressure drop (degrees F) is 88/125 X .5 = 35 F. The minimum capacity required to return oil is 34.8 tons for a 41/, in. OD riser and 6.4 tons for a 21/S in. OD FIG . 59 - DOUBLE H OT GAS RISER ’ TABLE 21-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP HOT GAS RISERS Refrigerant 12 COPPER TUBING-TYPE 1 TABLE 22-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP HOT GAS RISERS Refrigerant 22 COPPER TUBING-TYPE L I 1.1 100 .081 2.1 .ll\ 3.5 1 1.25 .071 2.5 .09( 4.1 .08 T = Tons of refriaeration. .101 F = Friction Droo. deorees 5.4 .101 10.6 ,091 18.1 .081 28.4 1 6.3 .081 12.5 .071 21.6 .071 33.9 F per 100 ft ecuivolent . .061 49 lenath at tons shown. MIN. LOA X *(x/200)‘.8 x 3.5 *Solve this equation to determine the friction drop multiplier for any ratio of full load to min. load, tons. TABLE 23-FLUID WEIGHT OF REFRIGERANT IN PIPING (Lb/l0 ft of length) PIPE SIZE* copper OD In. 1 Steel Nom. In. R12 RSOO 1 1% .013 .02 1 .043 .073 .llO .013 .02 .042 .072 .l 1 .016 .025 .05 1 .087 .13 % 2 1% 2% 3% 1 ?h 2 2% 3 .16 .27 .42 .60 .15 .27 .41 .59 .19 .33 Sl .72 3% 4 1% 5% 3% 4 5 .81 1.05 1.64 2.34 4.11 .80 1.04 1.62 2.33 4.06 .98 1.27 1.98 2.84 4.96 ‘55 5/s 7/s 1 ?h 1% 6 ‘Vi 8% K % J/4 6 8 HOT GAS LINES 100 F SAT COND TEMP LIQUID LINES 100 F TEMP SUCTION LINES 40 F SAT SUCT TEMP R22 R12 RSOO R22 .80 1.28 2.65 4.52 6.87 .70 1.13 2.33 3.98 6.06 .72 1.15 2.40 4.09 6.22 .032 .051 .105 .18 .27 .032 .052 .l 1 .18 .28 .047 .075 .16 .27 .40 9.74 16.9 26.1 37.3 8.56 14.9 23.0 32.9 8.8 1 15.3 23.6 33.7 .39 .67 1.04 1.5 .39 .69 1.1 1.5 .57 .99 1.5 2.2 2.0 2.6 4.1 5.8 10.2 2.0 2.7 4.1 6.0 10.4 3.0 3.8 6.0 8.6 15.0 R500 R12 50.5 65.5 102.0 147.0 257.0 R22 44.3 57.6 90.0 130.0 227.0 45.6 59.3 93.0 133.0 232.0 : To Correct for Temperatures Other Than Above, Multiply by the Following Factors: 12 500 22 LIQUID SUCT LINE-SAT. TEMP F REFRIGERANT - LINE-SAT. TEMP HOT F GAS LINE-SAT. TEMP F 50 30 10 - 1 0 - 3 0 40 60 80 100 120 80 90 100 110 120 1.18 1.18 1.18 .84 .84 .84 .59 .58 .58 .39 .39 .39 .2b .26 .25 1.09 1.10 1.1 1 1.06 1.07 1.08 1.03 1.04 1.04 1 .oo 1 .oo 1 .oo .97 .96 .98 .75 .74 .74 .87 .86 .86 1 .oo 1 .oo 1.00 1.15 1.16 1.16 1.33 1.33 1.35 *Refrigerants 12, 22 and 500 weights are for OD sizes of Type L copper pipe. ( ; i .I PAR’1 3 . 3-62 in Fig. 6Oh lor REFRIGERANT PIPING LAYOUT cxcs I’II’LNG D E S I G N whcrc the compressor i.s aOo7~e the e71trpmtor. EVAPORATORS Suction Line Loops Evaporator suction lines should be laid Out to accoml>lish the following objectives: 1. Prevent liquid refrigerant from draining into the compressor during shutdown. 2. Prevent oil in an active evaporator from draining into an idle evaporator. This can be done by using piping loops in the lines connecting the evaporator, the compressor and the condenser, Standard arrangements of suction line loops based on standard piping practices are illustrated in Fig. 60. Figure 6Oa shows the compressor located below a ogle evaporator. An inverted loop rising to the top i the evaporator should be made in the suction line to prevent liquid refrigerant from draining into the compressor during shutdown. ; A single euaporator below the compressor is illustrated in Fig. 6Ob. The inverted loop in the suction line is unnecessary since the evaporator traps all liquid refrigerant. Figure 6Oc shows multiple evaporators on different floor levels with the compressor below. Each individual suction line should be looped to the top of the evaporator before being connected into the suction main to prevent liquid from draining into the compressor during shutdown. Figure 60d illustrates multiple evaporators stacked on the same floor level or may represent a twocircuit single coil operated from one liquid’solenoid valve with the compressor located below the evaporator. In this arrangement it is possible to use one op to serve the purpose. Where coil banks on the same floor level have separate liquid solenoid valves feeding each coil, a separate suction’riser is required from each coil, similar to the arrangements in Fig. 6Oc and 60e for best oil return performance. Where separate suction risers are not possible, use the arrangement shown in Fig. SOf. Figure 60g shows multiple evaporators located on the same leuel and the compressor located below the evaporators. Each suction line is brought upward and looped into the top of the common suction line. The alternate arrangement shows individual suction lines out of each evaporator dropping down into a common suction header which then rises in a single loop to the top of the coils before going down to the compressor. An alternate arrangement is shown . When automatic compressor l~um1~clown control is used, evaporators are automatically kept lree of liquid by tho pumpdown operation and, therelorc, evaporators located a b o v e t h e compressor can be free-draining to the compressor without protective loops. The small trap shown in the suction lines immediatcly after the coil suction outlet is recommended to prevent erratic operation 0E the thermal expansion valve. The expansion valve bulb is located in the suction line between the coil and the trap. The trap drains the liquid from under the expansion valve bulb during compressor shutdown, thus preventing erratic operation of the valve when the compressor starts up again. A trap is required only when straight runs or risers are encountered in the suction line leaving the coil outlet. A trap is not required when the suction line from the coil outlet drops to the compressor or suction header immediately after the expansion valve bulb. Suction lines should be designed so that oil from an active evaporator does not drain into an idle one. Fig. 60e shows multiple evaporators on diflerent floor levels and the compressor above the evaporators. Each suction line is brought upward and looped into the top of the common suction line if its size is equal to the main. Otherwise it may be brought into the side of the main. The loop prevents oil from draining down into either coil that may be inactive. Figure 60f shows multiple evaporators stacked with the compressor aboue the evuporators. Oil is prevented from draining into the lowest evaporator because the common suction line drops below the outlet of the lowest evaporator before entering the suction riser. If evaporators must be located both above and below a common suction line, the lines are piped as illustrated in Figs. 60~1 and bob, with (a) piping for the evaporator above the common suction line and (b) piping for the evaporator below the common suction line. Multiple Circuit Coils All but the smallest coils are arranged with multiple circuits. The length and number of circuits are determined by the type of application. Multiple circuit coils are supplied with refrigerant thru a distributor which regulates the refrigerant distribution evenly among the circuits. Direct expansion coils can be located in any position, provided proper . EVAPORATOR ABOVE COMPRESSOR STACKED ON SAME LEVELON DlFFERENT LEVELSCO MPRESSOR ABOVE COMPRESSOR ABOVE (11 (al MIJLTIPLE E V A P O R A T O R S c, AND/OR LOOP ON DIFFERENT LEVELS COMPRESSOR BELOW (cl MULTlPLE STACKED ON SAME LEVELCOMPRESSOR BELOW Cdl E”APORATORS FIG. 60 - STANDARD ARRANGEMENTS b WHEN EQUAL NECESSARY COMPRESSOR ABOVE COMPRESSOR BELOW (hl (91 MULT,P,.E EVAPORATORS ON SAME LEVEL OF S UCTION LINE refrigerant distribution and continuous oil removal facilities are provided. In general the suction line piping principles shown in Fig. 60 should be employed to assure -oper expansion valve operation, oil return and lpressor protection. Figures 61 and 62 show direct expansion air coil piping arrangements in which the suction connections drain the coil headers effectively. Fig. 61 shows individual suction outlets joining into a common suction header below the coil level. Fig. 62 illustrates an alternate method of bringing up each suction line and looping it into the common line. The expansion valve equalizing lines are connected at the top of each suction header at the opposite end from the suction connection. Figure 63 illustrates the use of a coil having connections at the top or in the middle of each coil header, and piped so that this connection does not drain the evaporator. In this case oil may become trapped in the coil. The figure shows oil drain lines fro’m connections supplied for this purpose. The LOOPS (ONE-C IRCUIT COILS S HOWN) drain lines extend from the suction connection at the bottom end of each coil header to the common suction header below the coil level. EXPANSION 1” SUCTION LINE TO COMPRESSOR FIG . L O C A T E B U L B 45O ABOVE BOTTOM OF PIPE AND AS CLOSE AS POSSIBLE TO COIL OUTLET 61 - DX COIL USING SUCTION CONNECTIONS DRAIN COIL, SUCTION HEADER BELOW COIL TO PART 3. PIPING DESIGN 3-64 .\\ / ALTERNATE ARRANGEMENT SHOULD BE USED WHEN ,“A” A N D “8” A R E E Q U A L T O “C” w THERMAL r BULB SUCTION LINE TO COMPRESSOR / FIG. 64 - DRY EXPANSION COOLER / L LOCATE BULB 45* A B O V E BOTTOM OF PIPE AS CLOSE AS POSSIBLE TO COIL OUTLET FIG. 62 - DX COIL USING SUCTION CONNECTIONS D RAIN COIL, SUCTION HEADER A BOVE COIL TO Dry Expansion Coolers Figures 64 and 66 show typical refrigerant piping for a dry expansion cooler and a flooded cooler respectively. In a dry expansion chiller the refrigerant flows thru the tubes, and the water (or liquid) to be cooled flows transversely over the outside of the tubes. The water or liquid flow is guided by vertical baffles. Multi-circuit coolers should be used in systems in which the compressor capacity can be reduced below 50%. This is recommended since oil cannot be prop erly returned and good thermal valve control cannot be expected below this minimum loading per circuit. It is also recommended that the minimum capacity of a single circuit should be not less than 50yo of its full capacity. In addition, refrigerant solenoid valves should be used in the liquid line to each circuit of a multi-circuit cooler in a system in which , the compressor capacity can be reduced below 50’7& A liquid suction interchanger is recommended with these coolers. For the larger size DX coolers a pilot-operated refrigerant feed valve connected to a small thermostatic expansion valve (Fig. 65) may be used to advantage. The thermostatic expansion valve is a pilot device for the larger refrigerant feed valve. . Flooded Coolers In a flooded cooler the refrigerant surrounds the tubes in’the shell, and water or liquid to be cooled flows thru the tubes in one or more passes, depending on the baffle arrangement. Flooded coolers require a continuous liquid bleed line from some point below the liquid level in the cooler shell to the suction line. This continuous bleed of refrigerant liquid and oil assures the required return of oil to the compressor. It is usually LlPUlD S U C T I O N INTERCHANGER OIL I RETURN VALVE FIG. 63 - DX COIL USING OIL RETURN D RAIN CONNECTIONS TO D RAIN O IL FIG. 65 - HOOKUP EXPANSION VALVE FOR LARGE DX COOLERS CHAI’TEK 3. REFRIGERANT P[I’ING ,-SOLENOID LIQUID SUCTION .3-s VALVE BACK PRESSURE VALVE (WHEN REQUIRED1 OUBLE SUCTION RISER PILOT OPERATED / VALVE OIL BLEEDER LINE -! F1c.66 - FLOODEDCOOLER drained into the suction line so that the oil can be returned to the cooler with the suction gas. This drain line should be equipped with a hand shut-off valve, a solenoid valve and a sight glass. The solenoid valve should be wired into the control circuit in such a manner that it closes when the compressor stops. A liquid suction interchanger, installed close to the cooler, is required to evaporate any liquid refrigerant from the refrigerant oil mixture which is continuously bled into the suction line. Since flooded coolers frequently operate at light loads, double suction risers are often necessary. To avoid freeze-up the water supply to a flooded cooler should never be throttled and should never bypass the cooler. MPRESSORS Fro67 -LAYOUTOFSUCTION ANDHOTGASLINES FORMULTIPLECOMPRESSOROPERATION This allows the branch line to return oil proportionally to each of the operating compressors. Figure 67 shows suction and hot gas header arrangements for two compressors operating in parallel. Discharge Piping The branch hot gas lines from the compressors are connected into a common header. This hot gas header is run at a level below that of the compressor discharge connections which, for convenience, is often at the floor. This is equivalent to the hot gas loop for the single compressor shown in Fig. 68. The hot gas loop accomplishes two functions: Suction Piping Suction piping of parallel compressors should be designed so that all compressors run at the same suction pressure and so that oil is returned to the running compressors in equal proportions. To insure maintenance of proper oil levels, compressors of unequal sizes may be erected oh foundations at different elevations so that the recommended crankcase operating oil level is maintained at each compressor. All suction lines are brought into a common suction header which is run full size and level above the compressor suction inlets. Branch suction line take-offs to the compressors are from the side of the header and should be the same size as the header. No reduction is made in the branch suction lines to the compressors until the vertical drop is reached. GAS CONNECTION AT TOP OF CONDENSER HOT GAS LINE CONDENSER WATER-COOLED) 2 LOOP TO FLOOR -’ COMPRESS• R-/ \y FIG. 68 - HOT GAS LOOP ‘, i’.\R’I‘ 3. 3-66 PIPING DESIGN EC’UALIZER . FOR DRAIN OIL AND GAS EQUALIZERS FIG . 69 - INTERCONNECTING PIPING 1. It prevents gas, which may condense in the hot gas line during the off cycle, from draining back into the heads of the compressors. This eliminates compressor damage on start-up. 2. It prevents oil, which leaves one compressor, from draining into the head of an idle one. Interconnecting Piping In addition to suction and hot gas piping of parallel compressors, oil and gas equalization lines are required between compressors and between condensing units. An interconnecting oil equalization line is needed between all crankcases to maintain uniform oil levels and adequate lubrication in all compressors. The oil equalizer may be run level with the tappings or, for convenient access to the compressors, it may be run at floor level (Fig. 69). Under no condition should it be run at a level higher than that of the compressor tappings. Ordinarily, proper equalization takes place only if a gas equalizing line is installed above the com- FOR MULTIPLE CONDENSING UNITS pressor crankcase oil line. This line may be run level with the tappings, or may be raised to allow head room for convenient access. It should be piped level and supported as required to prevent traps from forming. Shut-off valves should be installed in both lines so that any one machine may be isolated for repair without shutting down the entire system. Both lines should be the same size as the tappings on the largest compressor. Neither line should be run directly from one crankcase into another without forming a U-bend or hairpin to absorb vibration. When multiple condensing units are interconnected as shown in Fig. 69, it is necessary to equalize the pressure in the condensers to prevent hot gas Erom blowing thru one of the condensers and into the liquid line. To do this a hot gas equalizer line is installed as shown. If the piping is looped as shown, vibration should not be a problem. The equalizer line between units must be the same size as the largest hot gas line. EVAPORATIVE CONDENSER A @URGE (114”) LOCATE HERF NOT AT TOP OF RECEIVER I SAFETY RELIEF VALVE COFINECTION ( S E E ASA-89.1) ENTIRE DRAIN LINE TO BE SAME SIZE, AS COIL OUTLET &‘HORIZONTAL LENGTH OF CONDENSATE PIPING LESS THAN 6 FEET 1 MUST CONNECT TO UPPER PART OF RECEIVER1 FIG. 70 - HOT GAS A ND LIQUID PIPING, SINGLE COIL UNIT W ITHOUT RECEIVER VENT CONDENSERS Liquid receivers are often used in systems having evapl>rative or air-cooled condensers and also with V’ r-cooled condensers where additional liquid su,‘age capacity is required to pump down the system. However, in many water-cooled condenser systems the condenser serves also as a receiver if the total refrigerant in the system does not exceed its storage capacity. When receivers are used, liquid piping from the condenser to the receiver is designed to allow free drainage of liquid from the condenser at all times. This is possible only if the pressure in the receiver is not allowed to rise to the point where a restriction in flow can occur. Liquid flow from the condenser to the receiver can be restricted by any of the following conditions: 1. Gas binding of the receiver. 2. Excessive friction drop in the condensate line. 3. Incorrect condensate line design. The following piping recommendations are made to overcome these difficulties. Evaporative Condenser to Receiver Piping Liquid receivers are used on evaporative condensers to accommodate fluctuations in refrigerant liquid level, to maintain a seal, and to provide storage facilities for pumpdown. An equalizing line from the receiver to the condenser is required to relieve gas pressure tending to develop in the receiver. Otherwise liquid hang-up in the condenser due to restricted drainage can occur. The receiver can be vented directly thru the condensate line to the condenser outlet, or by an external equalizer line to the condenser. Figure 70 shows a single evaporative condenser and receiver vented back thru the condensate drain line to the condensing coil outlet. Such an arrangement is applicable to a close coupled system. A separate vent is not required. However, it is limited to a horizontal length of condensate line of less than 3-68 L PART 3. PIPING DESIGN VENT LINETO OUTLET HEADER / COIL r / EVAPORATIVE I CONDENSER \ DETAIL HOT GAS LINE - TO \ “Y” EVAPORATOR / I f . AT TOP OF RECEIVER RECEIVER VENTfSIZE FROM TABLE 241 TO COIL OUTLET HEADER 4 I /’ SAME ,I!, AS TO SECOND ELBOW FROM TABLE 17 FOR VENTING AS SIZE CONDENSATE LINE SHOWN.OTHERWISE (FROM TABLE I7 1 CONSULT CONDENSER MANUFACTURER FIG . 71 - HOT GAS AND SAFETY RELIEF I - - - - VALVE CONNECTION & \(SEE ASA-B9.11 ,,&j ’ ‘B”~‘;O;~~‘E”E”~‘::;‘:;OVE L IF AT BOTTOM;‘X”MIN. WOULD BE MEASURED FROM COIL OUTLET TO BOTTOM CCNNECTION 1 LIQUID P IPING, S INGLE COIL UNIT WITH RE CEIVER V ENT 6 ft. The entire condensate line from the condenser to the receiver is the same size as the coil outlet. The line should be pitched as shown. Figure 71 shows the refrigerant piping for a single unit with receiver vent. Note that the condensate line from the condenser is the full size of the outlet connection and is not reduced until the second elbow is reached. This arrangement prevents trapping of liquid in the condenser coil. Table 24 lists recommended sizing of receiver vent lines. There are some systems in current use without a receiver but it must be recognized that problems can occur which can be avoided if a receiver is used. Such an arrangement is more critical with respect to refrigerant charge. An overcharged system can , waste power and cause a loss of capacity if the overcharge backs up into the condenser. An undercharged system also wastes power and causes a loss of capacity because the evaporator is being fed partially with hot gas. Therefore, if the receiver is omitted, an accurate refrigerant charge must be maintained to assure normal operation. TABLE 24-RECEIVER VENT LINE SIZING Receiver to Condenser VENT LINE SIZE BASED ON TYPE 1 COPPER TUBING (In. OD) REFRIGERATION (Tons, Max.) “‘8 .I’.F NOTE: MULTIPLE CONDENSER COMBINATIONS HAVE SAME COIL CIRCUIT LENGTHS, PURGE l/4” LOCATE HERE NOT AT TOP OF RECEIVER --..--..--..\ / ----A /‘I. RECEIVER VENT (SIZE FROM TABLE 2.4) TO COIL OUTLET HEADER SAFETY RELIEF VALVE CONNECTION (SEE ASA-89.1) LIOUID LEVEL SIGHT GLASS (OPTIONAL) FOR TOTAL LOAO ---GLASS i w SAME SIZE AS OUTLET TO FIRST TEE FIG. 72 - LOAO ( FROM TABLE 17) INLET MAY BE LOCATED AT GOTTOM SEE FIG. 71 OETAIL “Y” I IF AT BOTTOM “X” MIN WOULD SIDE TO AVOID FORMING TRAP ‘VALVE NO. I IS OPTIONAL PROVlOlNG BOTH VALVES NO. 2 ARE USED HOTGAS AND LIQUID PIPINC,MULTIPLEDOUBLE COILUNIT The advantage of such an arrangement is an ecoIx c one; equipment cost is lower since receiver anhvalves are eliminated and the system operating charge is lower if charged accurately. Figure 73 shows a subcooling coil piping ‘arrangement. The subcooling coil must be piped between the receiver and the evaporator for best liquid subcooling benefits. Figure 72 shows a piping arrangement for multiple units. Note that there are individual hot gas and vent valves for each unit. These valves permit operation of one unit while the other is shut down. These are essential because otherwise the idle unit, at lower pressure, causes hot gas to blow thru the operating unit into the. liquid line. Purge cocks are also shown, one for each unit. Multiple Shell and Tube Condensers The hot gas piping should be such that the pressure in each condenser is substantially the same. To accomplish this the branch connection from the hot gas header into each condenser should be the same size as the condenser coil connection. When two or more shell and tube condensers are applied in parallel in a single system, they should be equalized on the hot gas side and arranged as shown in Fig. SF. The elevation difference between the outlet of the condenser and the horizontal liquid header must be at least 12 in., preferably greater, to prevent gas blowing thru. The bottoms of all condensers should be at the same level to prevent backing liquid into the lowest condenser. When water-cooled condensers are interconnected ns shown, they should be fed from a common water regulating valve, if used. I’ART 3-70 3. PIPING D E S I G N HOT GAS LINE --) LIQUID L I N E TO EVAPORATOR ,I- CONDENSER COIL SUBCOOLING COIL CONNECTIONS - TURN VALVE ON SIDE TO AVOID SECOND ELBOW FIG . 73 - S UBCOOLING COIL P IPING An inverted loop of at least 6 ft is recommended in the liquid line to prevent siphoning of the liquid into the evaporator (or evaporators) during shutdown. Where a liquid line solenoid va!ve or valves are used, the loop is unnecessary. Figure 74 shows a similar loop for a single condenser with the evaporator below. 6’ MINIMUM Vibration of Piping Vibration transmitted thru or generated in refrig erant piping and the objectionable noise which results can be eliminated or greatly minimized by proper design and support of the piping. The best way to prevent compressor vibration from being transmitted to the piping is to run the suction and discharge lines at least 6 pipe diameters in each of three directions before reaching the first point of support. In this manner the piping can absorb the vibration without being overstressed. LIQUID LINE / II SUCTION FROM EVAPORATOR FIG . 74 TO EVAPORATOR BELOW - LIQUID LINE FROM CONDENSER OR RECEIVER TO E VAPORATOR L OCATED B ELOW \ TO EVAPORATOR I ( I F ABOVE C O N D E N S E R S I-4 CONDENSERS OR RECEIVERS TO EVAPORATOR CONDENSERS) LEG FIG. 55 - L IQUID PIPING TO INSURE CONDENSATE F LOW F ROM INTERCONNECTED CONDENSERS The hot gas loop from the compressor can be attached to the compressor base by means of a bracket il the base is isolated. If there is enough space in the horizontal run of the loop, two brackets are recommended to eliminate excessive rocking mc-‘ement of the pipe. Brackets should be attached a b :e point of minimum movement of the compressor assembly. The riser following the loop is supported as close as possible to the compressor. If the compressor is mounted on a resilient base, the pipe support should have a resilient isolator. The isolator is selected for four times the deflection in the spring support of the compressor base. S e e “Vihntion Isohtion of Piping Systems” in Cjicrptet- I for further discussion of the subject, REFRIGERANT PIPING ACCESSORIES LIQUID LINE liquid Suction lnterchbngers These are devices which subcool the liquid refrigerant and superheat the suction gas. The follow- ing describes four reasons for their use and the best location for each application: 1. To subcool the liquid refrigerant to compensate for excessive liquid line pressure drop. Location - near condenser. Liquid suction interchangers are not recommended for single stage applications using Refrigerant 22 because superheating of the suction gas must be limited to avoid compressor overheating. However, where they are used to prevent liquid slop-over to the compressor, superheating of the suction gas should be limited to 20 F above saturation temperature. A IiqUid suction interchanger so designed to limit the superheat of the suction gas should have a bypass so that operating adjustments may be made. 1. To act as an oil rectifier. Location - near cvaporator. 3. io prevent liquid slop-over to the compressor. Location - near evaporator. -1. To increase the effciency of the Kefrigerant I2 and 500 cycles. Location - near evaporator to avoid insulation 0E subcooled liquid line. I’:\R’r 3. P[I’ING DESIGN 3-72 GAS IN INTERCHANGER TEE OBTAINABLE FOR ANY , CONDITION LIQUID L,Q”lD O U T TEE, NOTE : SlZE THE OUTER PIPE ONE PIPE THE ,NNER SIZE IN PREFERRED CONSTRUCTION T BUSHING (LIOUIO) LARGER THAN PIPE (SUCTION). . ALTERNATE CONSTRUCTION FIG. 76 - LIQUID SUCTION INTERCHANGER . Two common types of liquid suction interchangers are: 1. The shell and coil or the shell and tube exchanger, suitable for increasing cycle efficiency and for liquid subcooling. This type is usually installed so that the suction outlet drains the shell to prevent oil trapping. 2. The tube-in-tube interchnnger (Figs. 76 and 77), a preferable type for controlling slop-over caused by erratic expansion valve feed or for “rectifying” lube oil from a refrigerant oil mixture bled from a flooded evaporator, Excessive superheating of the suction gas must be avoided with heat exchangers since it causes excessive compressor discharge temperatures. Therefore, the amount of liquid subcooling permissible by a liquid suction interchanger is limited to the amount of suction gas superheating that does not cause compressor damage when the gas is compressed to the discharge pressure. Beyond this point additional subcooling should be obtained from other sources. Charts 23 and 2-f are used to determine the length (A) of a concentric tube-in-tube interchanger (Fig. 76). The amount of liquid subcooling available is SUCTION GAS OUT SUCTION GAS IN LIQUID OUT LIQUID IN FIG. 77 - E CCENTRIC THREE -PIPE LIQUID SUCTION INTERCHANGER . CHART 23-EFFICIENCYCURVES, DOUBLE TUBE LIQUID SUCTION INTERCHANGER, Refrigerants 12 and 500 SUCTION TUBE SIZES LENGTH OF INTERCHANGER (FTl”A” CHART 24-EFFICIENCY CURVES, DOUBLE TUBE Refrigerant LIQUID 22 a3 0 2 ?I $ ii :: , iA 20 IO LENGTH OF INTERCHANGER (FT) “A” SUCTION INTERCHANGER, s-74 l’.ZK’I- L 3. PIPING DESIGN by using the ratio of the specific heats of the suction gas and of the liquid (subcooling multiplier). Example 5 illustrates the use of these charts. calculated Example 5 - Determining the Length Tube-in-Tube of a Concentric Interchanger Given: Refrigerant 12 system Load - 45 tons Suction line - Si/, in. OD Type L copper Expansion valve - 10 F superheat Suction temp - 40 F Condensing temp - 105 F Find: Length of a concentric tube-in-tube interchanger to superheat the suction gas to 6.5 F (suction gas temperature to compressor in accordance with ASRE Standard 23R on rating compressors). Amount of liquid subcooling. Solution: See Chart23. 1. Determine E= . FIG. interchanger efficiency E from PoRTL~QUID INDICATOR equation leav gas temp - ent gas temp ent liq temp - ent gas temp x 100 = s x 100 = +X 100 = 27.2% 2. With an efficiency of 27.2% a 31/8 in. OD pipe size and a 45 ton load, enter Chart Zjr as indicated per dashed line to determine a length (A) of 17 ft. 3. For Refrigerant 12 the ratio of gas to liquid specific heat is ,653. Therefore, subcooling of liquid refrigerant is ,653 X 15 F (leav gas temp - ent gas temp) or 9.8 F. An eccentric three-pipe interchanger is shown in Fig. 77. The inner pipe and the outer pipe offer two surfaces for the exchange of heat between the warm liquid refrigerant and the colder suction gas. The required length of this interchanger can be determined by using the method shown in Example 5 and by basing the required length on a ratio of relative surfaces between the liquid refrigerant and the suction gas. liquid indicators Every refrigeration system should include a means of checking for sufficient refrigerant charge. The common devices used are a liquid line sight glass, liquid IeveI test cock on condenser or receiver, or an external gage glass with equalizing connections and shut-off valves. The liquid line sight glass is one of the most convenient to install and use. A properly installed sight glass shows bubbling when there is an insufficient charge and a solid clear glass when there is sufficient charge. Sight glasses should be installed full size in the main liquid line and not in a bypass line that parallels the main line. . 78 -DOUBLE A sight glass with double ports and seal caps is preferable. The double ports allow a light to be put behind one port so that the state of the refrigerant is easily seen. The seal caps serve as an added protection against leakage or breakage since they are removed only when checking the refrigerant. Fig. 78 shows a double port liquid indicator with seal caps. Theinstallation of a double port or see-through sight glass is recommended in each of the following locations: 1. On evaporative condensing installations - in the liquid line leaving the receiver. 2. On single water-cooled condenser installations - in the liquid line leaving the condenser or, if an auxiliary receiver is used, in the liquid line leaving the receiver. 3. On multiple water-cooled condenser installations - in the main liquid line leaving the bank of condensers and also in the liquid line leaving the receiver if an auxiliary receiver is used. Strainers The installation of a strainer ahead of each automatic valve is recommended. Where multiple expansion valves with integral strainers are used, a single main liquid line strainer is sufficient to protect all of these. Fig. 79 shows an angle cartridge type strainer. A shut-off valve on each side of the strainer is desirable and should be located as close to the strainer as possible. On steel piping systems an adequate strainer should be installed in the suction line and a filterFigures 78-80, courtesy of ~Mueller Brass CO. FIG. 80 - ~INGI.IC FIG. 59 - TYPE DKIEK-STKAINKH ANGLE CARTRIDGE T YPE S TRAINER drier in the liquid line to remove the scale and rust inherent in steel pipe. Refrigerant Driers A permanent refrigerant drier is recommended ‘lost systems and is essential for all low temperafr tL,.i systems. It is also essential for all systems using hermetic compressors since the compressor motor winding is exposed to refrigerant gas. If the gas contains excessive moisture, the winding insulation may break down and cause the motor to burn out. A fullflow drier must be used for this type system. Figure 80 shows an angle type cartridge drier. The drier should be mounted vertically in the liquid line near the liquid receiver. A three-valve bypass (Fig. 81) should be used to permit isolation of the drier for servicing and to allow partial refrigerant How thru the drier. -r FIG. 8 1 - T HREE-V ALVE B YPASS DRIER FOR REFRIGERANT DRIER Reliable rnoistzlre indicator-s (Fig. 82) for liquid refrigerant lines are available. These devices indicate the proper time to replace the drier cartridge. Filter-Driers Filter-driers (Fig. 83) are more commonly used t’ 1 strainers and driers together. The drier mater;U1 actually filters the liquid refrigerant. Solenoid Valves Solenoid valves are commonly used in the following places: 1. In the liquid line of any system operating on single pump-out or pump-down compressor control. 2. In the liquid line of any single or multiple DX evaporator system. 3. In the oil bleeder lines from flooded evaporators to stop the flow of oil and refrigerant into the suction line when the system shuts down. In many cases it is desirable to use solenoid valves with opening stems. The opening stem serves as a by-pass so that the system may continue to operate in case of solenoid coil failure. FIG. 82 - COMBINATION M OISTURE AND L IQUID INDICATOR FIG. 83 - FILTER-DRIER Figures 82 and 83. COU~WS~ of Sporlnn Vnlve C O . I’,\K’I‘ 3-TG S. PIPING D E S I G N Five degrees is the usual change in superheat between a full open and closed position. This is called the operating superheat. Thus a valve which operates at 10 degrees superheat at design load balances out at 5 to 6 degrees superheat at low load. A low superheat setting at design load, therefore, does not provide sufficient margins of safety at low loads because of the 5 degrees necessary for operating superheat. The expansion valve bulb should be located immediately after the coil outlet on the suction line and 45”above the bottom of the pipe. With this arrangement the coil is the source of superheat for valve operation. The valve should be set so that overfeeding does not occur at times of partial load. heat. F I G . 8 4 - I<EFKLGERANT C H A R G I N G C O N N E C T I O N S Refrigerant Charging Connections The two usual methods of charging the refrigeration system are: 1. Charging liquid into the liquid line between the receiver shut-off valve and the expansion valve. Fig. 8-f shows a charging connection in a liquid line leaving a receiver. 2. Charging gas into the suction line. Except for very small systems this method is not practical because of the time required to evaporate the refrigerant from the drum and because of the danger of dumping raw liquid into the compressor. Expansion Valves Thermal expansion valves should be sized to avoid both the penalties of being undersized at full load and of being excessively oversized at partial load. The following items should be considered before sizing valves: 1. Refrigerant pressure drop thru the system must be properly evaluated to determine the correct pressure drop available across the valve. 2. Variations in condensing pressure during operation affect valve pressure and capacity. Condensing pressure should be controlled, therefore, to maintain required valve capacity. 3 . Oversized thermal expansion valves do not control as well at full system capacity as properly sized valves and control gets progressively worse as the coil load decreases. Capacity reduction, available in most compressors, further increases this problem and necessitates closer selection of expansion valves to match realistic loads. When sizing thermal expansion valves, make the selection on the basis of maximum load at the design operating pressure and at least 10 degrees super- The effect of condensing temperature on the capacity of an expansion valve for two differentl systems is illustrated in Example 6. Example 6 - Effect of Condensing Temperature on Expansion Valves Given: Two refrigeration systems using Refrigerant 500, one operating at 40 F suction and 90 F condensing, the other operating at 40 F suction and 130 F condensing. 218.2 psig 6.2 Coxidensing pressure Liquid line drop Pressure at thermal expansion valve inlet Suction pressure Suction line losses Coil pressure drop Distributor pressure 212.0 psig 46.2 psig 2.8 7.0 17.0 drop Pressure at thermal expansion valve outlet Pressure drop available across valve 73.0 psig 139.0 psi Assume that a valve of 27.5 ton capacity at 40 F suction and 60 psi differential is selected. Find: Capacity at the pressure drop available across the valve of systems 1 and 2. Solution: The capacities will of the pressures: vary approximately as the square root 1% For system 1 Cap. T 23 tons For system 2 Cap. = 42 tons Note that the expansion valve capacity is nearly double at the higher head pressure. <;H;\YI‘EK 3. KEFKIGEKANT l?II’INC; On certain low temperature applications and on high temperature applications where the design or partial load least temperature difference (L.T.D.) between the refrigerant and air or water is extremely small, it may become necessary to consider the use of the liquid suction interchanger as a source of superheat. This is done to increase the effective evaporator surface by allowing the liquid suction interchanger to supply the superheat function. If only one liquid suction interchanger is used for the applications just mentioned, it should be an eccentric three-pipe interchanger as shown in Fig. 77. This arrangement permits the expansion valve bulb to sense the suction gas temperature from the outside surface of the interchanger. Otherwise two tube-in-tube interchangers should be used with the thermal expansion valve bulb located between the iv -rchangers. ,‘he preferred refrigerant flow in a coil circuit to obtain superheat is illustrated in Fig. 55. SUCTION LINE Back Pressure Valves A conventional type back pressure regulating valve is used in a refrigerating system to maintain a predetermined pressure in the evaporator. A conventional type regulator controls the upstream pressure. The regulator has a spring loaded diaphragm designed to actuate a seat pilot valve. The actuating pressure comes from the evaporator or upstream side of the regulator. When the upstream pressure against the diaphragm is greater than that exerted by the spring, the pilot valve opens and a flow of gas is admitted to the power piston. The piston in turn causes the main port to open. This permits a flow of from the upstream side of the valve to the down‘b-.;earn side. When the actuating pressure becomes less than that controlled by the spring pressure, the flow of gas to the power piston is stopped and the regulator closes. There are many variations of the back pressure regulating valve. Several are described in the following: 1. The compensating type, actuated by air or electricity, varies the suction pressure in accordance with temperature or load demand. 2. The dual pressure regulator is designed to operate at two predetermined pressures without‘ resetting or adjustment; by opening and closing a pilot solenoid, either the low pressure or the high pressure head operates. Figzlre 86 shows a simple back pressure regulating 3-77 - AIR REFRIGERANT7 FN;. 85 - ~‘RI-IlXK1’:I) C1RC:I:I-r TO OIsl~AIN ~I.I:RI(;ICRA\NT SuI’I<RHI:.A2.L. FLOW IN COIL (l’I.AN VIEW) valve which is ordinarily used for one of the following purposes: 1. To control evaporator suction pressure in spite of compressor suction pressure variation. 2. To establish evaporator suction pressure when lower compressor suction pressure is demanded by another part of the same system. 3. To prevent evaporator freezing when operating near the freezing temperature. PRESSURE ADJUSTING STEM GAGE CONNECTION - / LCOPPER TUBE CONNECTION FOR REFRIGERANTS 12, 22 a 500 OPENING STEM ‘SEAL CAP B PACKING GLAND FIG. 86 - BACK P RESSURE V ALVE . P,\K’I‘ 3. PIPING 3-78 CHART 25-BACK PRESSURE VALVE APPLiCATlON CHART EXPANSION FIG. 87 -INSTALLATION USING BACK P RESSURE V ALVES DI3IGN CHAI’TEK 3. KEI~K1~~EK,\N’I‘ 3-79 I’II’LNC C/WJ.~ 135 illustrates the application ot back pressure valves for various services such as number and types of evaporators, and types of room and comprcssor control. [;igure 87 illustrates the location of back pressure valves. DISCHARGE LINE Oil Separators Oil separators reduce the rate of oil circulation. However they are not 100CJ$ efficient since some oil always circulates thru the system. Oil separators are of particular value on certain types of installations such as: 1. Systems requiring a sudden and frequent capacity variation. 2. Systems having extensive pipe lines and numerous evaporators. The large volumes inherent in s&h systems result in appreciable oil hangup. ‘I-here are several objections to oil separators: 1. Oil separators permit some oil to be carried over into the system and, therefore, proper piping design to return oil is still required even though a separator is used. 2. On start-up, gas may condense in the shell of the separator. As a result the separator delivers liquid refrigerant into the crankcase. This in turn increases crankcase foaming and oil loss from the compressor. During the “off” cycle the oil separator cools down and acts as a condenser for liquid refrigerant that evaporates in the Garmer parts of the system. Thus a cool oil separator acts as a liquid condenser during “off” cycles and also on compressor start-up until the separator has warmed up. Large amounts i ,quid refrigerant in the crankcase result in poor lubrication and may also result in removing the oil from the crankcase completely. Figwe 88 shows the recommended method for piping an oil separator. F1c;.88 - OIL SEPARATOR LOCATION Check Valves Check valves contribute a relatively large addition to a line pressure drop at full load and must be taken into account in the selection of refrigeration cquipmcnt. In addition a check valve cannot be relied upon for 100% shut-off. Whenever the receiver is warmer than the compressor during shutdown, refrigerant in the receiver tends to boil off and flow back thru the condenser and hot gas discharge line to the compressor where it condenses. If thcrc is sufficient refrigerant in the receiver, liquid refrigerant eventually enters the compressor despite the loop in the hot gas line at the base of the compressor. To prevent this, a check valve should be used (Fig. 68, page 65). In a non-automatic system a hand valve may be used at the inlet to the condenser to manually shut off the flow on shutdown, in which case the pressure drop involved will be much less than that cncountcred using a check valve. REFRIGERANT PIPING INSULATION Liquid lives should not be insulated if the surrounding temperature is lower than or equal to the temperature of the liquid. Insulation is recom- , HOT GAS LINE Mufflers II a muffler is used in the hot gas line, it should be installed in downward flow risers or in horizontal lines as close to the compressor as possible. The hot gas pulsations from the compressor can set up a condition of resonance with certain lengths of refrigerant piping in the hot gas line. A muffler installed in the compressor discharge aids in climinating such a condition. FigWe 89 shows a muffler in a hot gas line at the compressor. \ WEDGE CAP BOTH ENDS HOT GAS MUFFLER COMPRESSOR -! 3-N n~entlccl only whl the liciuid line can pick L I P a consitleral~lc amount ol heat. The lollowing areas in a refrigerant piping system should be insulated: I. .\ liquid line exposed to the direct rays ol the S L I P lor ;I considerable distance. 2. Piping in boiler rooms. 3. Piping at the outlet of a liquid suction interchanger to preserve the subcooling effect. Where liquid and suction lines cm be strapped togcthcr, a single insulating covering cm be used over both lines. This inducts an exchange of heat ant1 is desirable from the standpoint of the subcooling effect on the liquid. However excessive superheating of the suction gas can result from too much exchange oE heat. Hot gns li’nes should not be insulated. Any heat lost by these lines reduces the work to be done by the condenser. Sztclion lines should be insulated only to prevent dripping where this causes a nuisance or damage. It is generally desirable to have the suction line capable of absorbing some heat to evaporate any liquid which may have entered the suction line from the evaporator. For unusual conditions of Iligh ambient temperatures and simultaneous high PAR’1 3. I’LI’ING DESLGN relative humidities extra insulation must be applied. The thickness oE insulation required to prevent condensation on the outer surlacc is that thickness which raises the temperature of the outer surlacc ol’ the insulation to a point slightly higher than the m;iximum expected dewpoint of the surrounding air. The external vapor barrier must be made as nearly perfect as possible in order to prevent leakage of vapor into the insulation. Kcgular “ice water” thickness moulded cork covering wired on and sealed with asphalt primer is desirable l’or most work in the air conditioning range. For lower temperatures, “brine” thickness moulded cork covering should be used. ‘Insulation that is not vapor-proof soon becomes saturated with moisture and rapidly deteriorates. A cellular glass or cellular plastic type of insulation is fast becoming accepted as an ideal insulation. Its cellular structure provides exceptionally high resistance to water and water vapor. The cellular glass, being inorganic, is fire-proof. Cellular plastic which is also available is self-extinguishing. When located out of doors, insulation must be weatherprooFed unless, of course, it is inherently waterproof. . 3-81 CHAPTER 4. STEAM PIPING This chapter describes practical design and layout techniques for steam piping systems. Steam piping differs from other systems because it usually carries three fluids - steam, water and air. For this reason, steam piping design and layout require special consideration. c JERAL SYSTEM DESIGN Steam systems are classified according to piping arrangement, pressure conditions, and method of returning condensate to the boiler. These classifications are discussed in the following paragraphs. PIPING ARRANGEMENT A one- or two-pipe arrangement is standard for steam piping. The one-pipe system uses a single pipe to supply steam and to return condensate. Ordinarily, there is one connection at the heating unit for both supply and return. Some units have two connections which are used as supply and return connections to the common pipe. A two-pipe steam system is more commonly used in air conditioning, heating, and ventilating applications. This system has one pipe to carry the steam supply and another to return condensate. In a two‘: system, the heating units have separate conn.-ctions for supply and return. The piping arrangements are further classified with respect to condensate return connections to the boiler and direction of flow in the risers: 1. Condensate return to boiler a. Dry-return - condensate enters boiler above water line. b. Wet-return - condensate enters boiler below water line. 2. Steam flow in riser a. Up-feed - steam flows up riser. b. Down-feed - steam flows down riser. PiESSURE CONDITIONS Steam piping systems are normally divided into five classifications - high pressure, medium pressure, low pressure, vapor and vacuum systems. Pressure ranges for the five systems are: 1. High pressure - 100 psig and above 2. Medium pressure - 15 to 100 psig 3. Low pressure - 0 to 15 psig 4. Vapor - vacuum to 15 psig 5. Vacuum - vacuum to 15 psig Vapor and vacuum systems are identical except that a vapor system does not have a vacuum pump, but a vacuum system does. CONDENSATE RETURN The type of condensate return piping from the heating units to the boiler further identifies the steam piping system. Twoarrangements, gravity and mechanical return, are in common use. When all the units are located above the boiler or condensate receiver water line, the system is described as a gravity return since the condensate returns to the boiler by gravity. If traps or pumps are used to aid the return of condensate to the boiler, the system is classified as a mechanical return system. The vacuum return pump, condensate return pump and boiler return trap are devices used for mechanically returning condensate to the boiler. CODES AND REGULATIONS All applicable codes and regulations should be checked to determine acceptable piping practice for the pxticular application. These codes usually dictate piping design, limit the steam pressure, or qualify the selection of equipment. WATER CONDITIONING The formation of scale and sludge deposits on the boiler heating surfaces creates a problem in generating steam. Scale formation is intensified since scale-forming salts increase with an increase in temperature. Water conditioning in a steam generating system should be under the supervision of a specialist. PAR-I. 3-82 3. PIPING DESIGN L TABLE 25-RECOMMENDED HANGER SPACINGS FOR STEEL PIPE DISTANCE BETWEEN SUPPORTS (FT) NOM. PIPE SIZE (in.1 Average Gradient 1” in 10’ J/i I/?” in 10’ ’ 1 1% 1% 9 13 16 19 2 21 I ; 6 I ; 10 14 17 i; 19 40 33 ‘A” in 10’ I I 1 5 8 13 23 I 25 NOTE: Data is bared on standard wall pipe filled with water and overage number of fittings. Courtcsv or Crnne co. PIPING ONE-PIPE SYSTEM SUPPORTS All steam piping is pitched to facilitate the flow of condensate. Table -3.5 contains the recommended support spacing for piping pitched for different gradients. The data is based on Schedule 40 pipe filled with water, and an average amount of valves and fittings. PIPING DESIGN A steam system operating for air conditioning comfort conditions must distribute steam at all operating loads. These loads can be in excess of design load, such as early morning warmup, and at extreme 1’K;. (30 - &JK-I’Il’1.:, partial load, when only a minimum of heat is necessary. The pipe size to transmit the steam for a design load depends on the following: I. The initial operating pressure and the allowable pressure drop thru the system. 2. The total equivalent length of pipe in the longest run. 3. Whether the condensate flows in the same direction as the steam or in the opposite direction. The major steam piping systems used in air-conditioning applications are classified by a combination of piping arrangement and pressure conditions as follows: 1. Two-pipe high pressure 2. Two-pipe medium pressure 3. Two-pipe low pressure 4. Two-pipe vapor 5. Two-pipe vacuum 6. One-pipe low pressure IJl’F1’.1:11 (;KAVITY SYSTEM A one-pipe gravity system is primarily used on residences and small commercial establishments. Fig. 90 shows a one-pipe, upfeed gravity system. The steam supply main rises from the boiler’to a high point and is pitched downward from this point around the extremities of the basement. It is normally run full size to the last take-off and is then reduced in size after it drops down below the boiler water line. This arrangement is called a wet return. If the return main is above the boiler water line, it is called a dry return. Automatic air vents are required at all high points in the system to remove non-condensable gases. In systems that require long mains, it is necessary to check the pressure drop and make sure the last heating unit is sufficiently above the water line to prevent water backing up from. the boiler and flooding the main. During operation, steam and condensate flow in the same direction in the mains, and in opposite direction in branches and risers. This system requires larger pipe and valves than any other system. The one-pipe gravity system can also be designed as shown in Fig. 91, with each riser dripped separately. This is frequently done on more extensive systems. Another type of one-pipe gravity system is the down-feed arrangement shown in Fig. 92. Steam flows in the main riser from the boiler to the building attic and is then distributed throughout the building. FIG. I) 1 - ONI:-PII’IC GRAVITY DRIPPED S~sTl:;cf WITH I<ISI-KS TWO-PIPE SYSTEM A two-pipe gravity system is shown in Fig. 93. This system is used with indirect radiation. The addition of a thermostatic valve at each heating unit adapts it to a vapor or a mechanical vacuum system. A gravity system has each radiator separately sealed by drip loops on a dry return or by. dropping directly into a wet return main. All drips, reliefs and return risers from the steam to the return side of the system must be sealed by traps or water loops to insure satisfactory operation. If the air vent on the heating unit is omitted, and the air is vented thru the return line and a vented condensate receiver, a vapor system as illustrated in Fig. 97’ results. The addition of a vacuum pump to a vapor system classifies the system as a mechanical vacuum em. This arrangement is shown in Fig. 95. CHART 26-PIPE SIZING+ 5 6 810 2 0 4 0 60 100 200 400 1000 600 2000 4 0 0 0 1 0 0 0 0 6000 2 0000 50000 100 0 0 0 S T E A M F L O W RATE-LB/HR *Use Chart 27 to determine steam velocity at initial saturated steam pressures other than 0 psig. Charts 26 and 27 from Heating Ventilating Air Conditioning Guide J 959. Used by permission. (:~1.\l”I‘l:l< I. s’I‘l-.\,\l l’ll’lN(; 3-85 CHART 27-VELOCITY CONVERSION* RECOMMENDATIONS The following recommcnd~ttioris are for use when sizing pipe for the various systems: Two-Pipe High Pressure System 60000 40 000 3 0 0 0 0 20000 10000 0000 6000 u - 1000 0 0 0 G 2 e 600 600 - 4 0 0 4 0 0 300 300 1000 000 200 100 40 0 5 IO 2 0 100 140 100 “- 2 0 0 PkSSURE-PSIG 60RO ~~__ SATURATED *See Exomde STEAM 3, page 89, for use of chart PIPE SIZING GENERAL Charts and tables have been developed which are used to select the proper pipe to carry the required steam rate at various pressures. ,hart -36 is a universal chart for steam pressure of 0 to 200 psig and for a steam rate of from 5 to 100,000 pounds per hour. However, the velocity as read from the chart is based on a steam pressure of 0 psig and must be corrected for the desired pressure from Chart 27. The complete chart is based on the Moody friction factor and is valid where condensate and steam flow in the same direction. Tables 26 tl~~ 31 are used for c1uic.k selection at specific steam pressures. Clrtr~t _“(, has been used to tabulate the capacities shown in Tnb1r.r 2~; thm -38. The capacities in TnD1e.s 23 t/cl-u 3~ are the results of tests conducted in the ASHAE laboratories. Suggested limitations for the USC of these tables are shown as notes on each table. In addition, Table 31 shows the total pressure drop I’or two-pipe low pressure steam systems. This system is used mostly in plants and occasionally in commercial installations. 1. Size supply main and riser for a maximum drop of 25-30 psi. 2. Size supply main and risers for a maximum friction rate of 2-10 psi per 100 ft of equivalent pipe. 3. Size return main and riser for a maximum pressure drop of 20 psi. 4. Size return main and riser for a maximum friction rate of 2 psi per 100 ft of equivalent pipe. 5. Pitch supply mains G in. per 10 ft away from boiler. 6. Pitch return mains ti in. per 10 ft toward the L boiler. 7. Size pipe from Table 26. Two-Pipe Medium Pressure System This system is used mostly in plants and occasionally in commercial installations. 1. Size supply main and riser for a maximum pressure drop of 5-10 psi. 2. Size supply mains and risers for a maximum friction rate of 2 psi per 100 ft of equivalent pipe. 3. Size return main and riser for a maximum pressure drop of 5 psi. 4. Size return main and riser for a maximum friction rate of 1 psi per 100 ft of equivalent pipe. 5. Pitch supply mains ‘/4 in. per 10 ft away from the boiler. 6. Pitch return mains ‘/4 in. per 10 ft toward the boiler. 7. Size pipe from Table 27. Two-Pipe Low Pressure System This system is used for commercial, air conditioning, heating and ventilating installations. 1. Size supply main and risers for a maximum pressure drop determined from Table 31, depending on the initial system pressure. 2. Size supply main and riser for a maximum friction rate of 2 psi per 100 ft of equivalent pipe. 3. Size return main and riser for a maximum TABLE 26-HIGH PRESSURE SYSTEM PIPE CAPACITIES (150 psig) Pounds Per Hour PRESSURE DROP PER 100 FT PIPE SIZE (in.) l/s psi ‘Vi psi (4 (2 02) ‘h psi (8 ox) S/4 psi 02) psi (12 o z ) 1 SUPPLY MAINS AND RISERS =/ I 29 58 130 203 412 683 1,237 1,855 2.625 41858 7,960 16,590 30,820 48,600 1. 1% 1% 2 2% 3 3% 4 5 6 8 10 12 1 41 I 583 959 1,750 2,626 3.718 6;875 11,275 23,475 43,430 68,750 RETURN 1 156 313 % 1% 1 ‘h 2 2% 3 3% 4 5 6 , 650 1.070 21160 3,600 6,500 9,600 13,700 25,600 42,000 58 825 1,359 2,476 3,715 5.260 9;725 15,950 33,200 61,700 97,250 1,167 1,920 3,500 5,250 7,430 13;750 22,550 46,950 77,250 123,000 M A I N S AND R I S E R S 232 462 960 1.580 3;300 5,350 9,600 14,400 20,500 38,100 62,500 82 I 360 690 1,500 2,460 41950 8,200 15,000 22,300 31,600 58,500 96,000 465 910 1 1,950 3,160 6,400 10,700 19,500 28,700 40,500 76,000 125,000 pressure drop detcrminetl from Tlrble 31, dcpending on the initial system pressure. -1. Size return main and riser for a maximum friction rate of I/? psi per 100 ft of equivalent pipe. 5. Pitch mains I/~ in. per 10 ft away from the boiler. F. Pitch return mains Q$ in. per 10 Et toward the boiler. 2 psi (32 01) (16 0x1 130 - 180 de--Max 1 116 233 523 813 1,650 2,430 4,210 6,020 8.400 15;ooo 25,200 50,000 90,000 155,000 / 10 psi 5 psi Error 8% 184 369 827 1.230 2;ooo 3,300 6,000 8,500 12.300 2lj200 36,500 70,200 130,000 200,000 I 300 550 1,230 1.730 3;410 5,200 9,400 13,100 19.200 331100 56,500 120,000 210,000 320,000 I 420 790 1,720 2,600 4,820 7,600 13,500 20,000 28.000 471500 80,000 170,000 300,000 470,000 1 - 2 0 prig - Max Return Pressure 560 1,120 890 1,780 2,330 3,800 7,700 12,800 23,300 34,500 49,200 91,500 150,000 3,700 6,100 12;300 20,400 37,200 55,000 78,500 146,000 238,000 6. Pitch return mains ‘/, in. per 10 It towd boiler. : the 7. Size pipe from Tables -78 th?l 30. Two-Pipe Vacuum System This system is used in commercial installations. Two-Pipe Vapor System 1. Size supply main and riser for a masimum pressure drop of !h - 1 psi. 2. Size supply main and riser for a maximum friction rate of !h - I/? psi per 100 ft of equivalent pipe. This system is used in commercial and residential installations. 3. Size return main and riser for a maximum pressure drop of I$$ - 1 psi. 1. Size supply main and riser for a maximum pressure drop of ‘,/1,; - t/s psi. 4. Size return main and riser for a maximum friction rate of Q$ - I/? psi per 100 ft of equivalent pipe. 2. Size supply main and riser for n maximum friction rate of !/lr; - IA3 psi per 100 It of equivalent pipe. 5. Pitch supply mains !/( in. per 10 ft away from the boiler. .?. Size return main and supply for a m;1ximum pressure drop of \il; - xY psi. 4. SiLe return main and supply for ;t maximum friction raw of )i,; - I$(? psi per 100 ft of equivalent pipe. One-Pipe Low Pressure System 5. Pitcil supply 51 i n . per 10 ft away Iwiler. This system is usctl o n sm;dl commercial and rcsitlential systems. front t h e G. Pitch return mains !/r in. per 10 Et townrcl boiler. 7. Sire pipe from Tables 2S tht,~~ 30. the TABLE 27-MEDIUM PRESSURE SYSTEM PIPE CAPACITIES (30 psig) Pounds Per Hour PIPE SIZE (in.) PRESSURE DROP PER 100 FT ‘/s psi (2 02) % psi (4 02) =/4 1 1 ‘A 1% 2 2% 3 3% 4 5 6 8 10 12 15 31 69 107 217 358 651 979 1,386 2,560 4,210 8,750 16,250 25,640 ?h 1 1% 1% 2 115 230 485 790 1,575 2% 3 3% 4 5 2,650 4,850 7,200 10,200 19,000 3,900 7,100 10,550 15,000 27,750 6 3 1,000 45,500 RETURN . ‘ii psi (8 02) 9% psi (12 02) MAINS AND 170 340 710 1,155 2,355 2 1 psi (16 oz) 2 psi (32 OL) 25 - 35 psig - Max Error 8% SUPPLY MAINS AND RISERS 22 31 46 63 100 141 154 219 313 444 516 730 940 1,330 1,414 2,000 2,000 2,830 3,642 5,225 6,030 8,590 12,640 17,860 23,450 33,200 36,930 52,320 38 77 172 267 543 924 1,628 2,447 3,464 6,402 10,240 21.065 40,625 64,050 45 89 199 309 627 1,033 1,880 2,825 4,000 7,390 12,140 25,250 46,900 74,000 63 125 281 437 886 1,460 2,660 4,000 5,660 10,460 17,180 35,100 66,350 104,500 0 - 4 psig - Max Return Pressure RISERS 245 490 1,025 1,670 3,400 308 615 1,285 2,100 4,300 365 730 1,530 2,500 5,050 5,600 10,250 15,250 1,600 40,250 7,100 12,850 19,150 27,000 55,500 8,400 15,300 22,750 32,250 60,000 65,500 83,000 ,98,000 TABLE 2B-LOW PRESSURE SYSTEM PIPE CAPACITIES Pounds Per Hour CONDENSATE NOM. PIPE SIZE (in.) vi 1 1% 1% 2 2% 3 3 ‘h 4 'A6 psi (1 02) 3.5 9 17 36 56 108 174 318 462 726 12 ‘/a psi (2 02) 1 3.5 11 1 / 12 FLOWING WITH THE FLOW PRESSURE DROP PER 100 FT % psi (12 02) % psi (4 02) ‘/a psi (8 02) S A T U R A T E D P R E S S U R E (PSIG) 1 3.5 1 12 1 3.5 I 12 I 3.5 I 12 I 14 / 161 20 1 241 29 54 31 37 46 66 70 96 111 100 120 234 194 378 310 550 660 800 990 1,160 1,410 2,100 2 , 4 4 0 3,350 3 , 9 6 0 7,000 8,100 12,600 15,000 3 , 7 0 0 ( 1 6 , 5 0 0 ) 1 9 , 5 0 0 1 23;400 1 28;400 / 3 3 , 0 0 0 The weight-flow rater at 3.5 psig cctn be used +CJ cover 8 to 16 psig with an error not exceeding 8 percent. STEAM so+. press. 35 66 138 210 410 660 1,160 1,700 2,400 4,250 7,000 14,300 26,000 40.000 1 psi 3.5 I 2 psi 12 50 95 200 304 590 950 1,670 2,420 3,460 6,100 10,000 20,500 37,000 I 3.5 60 114 232 360 710 A 1,150 1,950 2,950 4,200 7.500 A 11,900 24,000 42,700 67,800 from 1 to 6 prig, and the rater at 12 prig con be used to cover sat. press. Tobk 26 thru 28 from Heating Venfilafing I 12 ” 73 137 280 430 850 1,370 2,400 3,450 4,900 8.600 A 14,200 29,500 52,000. 81,000’ from Air Conditioning Guide 1959. Used by permission. TABLE 29-RETURN MAIN AND RISER CAPACITIES FOR LOW PRESSURE STEAM SYSTEMS PIPE SIZE ‘5% psi (l/z 02) (in.) W e t * 1 D r y / VclC PRESSURE DROP PER 100 FT l/24 psi (75 02) wet* Vat DV ‘A6 psi (1 01) Wet* 1 Dry Vat ‘Al psi (2 02) Wet+ / Dry VlX RETURN 1 125 213 -.338 700 62 1.-30 206 470 145 248 393 810 71 149 236 535 42 143 244 388 815 1.180 760 1,580 868 1,360 J/4 1-,. ‘A 1% 2 2% 3 % 21' .7 580 8 0 4 3:880 5 6 1 1'970 1460 2:930 175 300 475 1,000 80 168 265 575 1,680 ?4 48 48 143 1 113 113 244 1 % 240 248 388 1% 375 375 815 2 750 750 1,360 2% 2,180 3 3,250 3% 4,480 4 7,880 5 12,600 * V a t values may be u s e d f o r w e t r e t u r n r i s e r s a n d m a i n s . 7. Size supply main and dripped 40 113 248 375 750 runouts from 200 350 600 950 2,000 283 494 848 1,340 2,830 2,350 1 1,2301 2,380 3,350 1,360 3,350 4,730 71; 5 6 0 1 , 3 0 0 15,500 27,300 143,800 RISERS 48 113 248 375 750 249 426 674 1,420 2,380 3,800 5,680 7,810 13,700 22,000 Vctltilating Air Conditioning: 48 350 113 600 248 950 375 2,000 7501 3.3501 I 5.3501 8,000 11,000 19,400 31,000 I I C;uidc, 19’19. Used by 494 848 1,340 2,830 I 41730 1 7.560 ' ,~ 11,300 15,500 27,300 43,800 permission. 8. Size undripped runouts from Table 30, Column I;. 9. Size upfeed risers from Table 30, Cqlunn D. 10. Size downfeed supply risers from Table 28. il. Pitch supply mains vi in. per 10 ft away from boiler. Pounds Per Hour CONDENSATE FLOWING AGAINST STEAM FLOW I TWO-PIPE ’ SYSTEM Vertical Horizontal B* ct 8 1 ONE-PIPE 12. Pitch return mains M in. per 10 ft toward the boiler. SYSTEM Up-feed Vertical Supply CO”Risers nectars Riser RUllouts DIt E F 14 31 48 97 - 1 1 'h 1% 2 9 i9 27 49 6 11 20 38 72 7 16 23 42 7 7 16 16 23 2'h 3 3% 4 5 159 282 387 511 1,050 99 175 288 425 788 116 200 286 380 - - 42 65 119 186 278 1,800 3,750 7,000 1 1,500 22,000 1,400 3,000 5,700 9,500 19,000 - - 545 - 6 8 10 12 16 115 241 378 825 I:ro~n Iqe:lting TABLE 30-LOW PRESSURE SYSTEM PIPE CAPACITIES % 350 600 950 2,000 175 300 475 1,000 1,680 2,680 4,000 5,500 9,680 15,500 ToDie 23. A 142 103 249 217 426 3401 674 740 1 1,420 250 425 675 1,400 2 , 1 3 0 3 , 3 0 02 , 2 0 0 1 , 5 6 03 , 2 5 0 2 , 1 8 04 , 0 0 0 2 , 6/ /2 8 0 , 5 0 0 1 , 7 5 04 , 0 0 0 2 , 6 8 05 , 5 0 0 3 , 7 5 02 , 2 5 0 3 , 2 3 03 , 8 0 0 5 , 6 8 05 , 3 5 0 8 , 0 0 02 , 5 0 0 3 , 5 8 05 , 3 5 0 8 , 0 0 0 4,580 3,350 4,500 5,500 3,750 5,500 7,750 4,830 7,810 11,000 5,380 11,000 7,880 9,680 13,700 19,400 Il2,6001 /1 5 , 5 0 0 1 /22,000 1 ~31,000~ I RETURN PIPE SIZE (in.) ‘/a psi (8 02) wet* VfVJ W MAINS 100 175 300 475 , 1,000 1,680 1 9501 ‘/i psi (4 02) Wet* Dry VW2 *DO not use Column B for pressure drops less than %6 psi/100ft of e q u i v a l e n tl e n g t h . U s eC h a r t 2 6 , page 8 4 . i P i t c h of h o rtu o n oIr u n o v t rt o r i s e r s h o u l d b e n o t l e s s t h a n ‘ % i n . / f t . W h e r e t h i s p i t c h c a n n o t b e o b t a i n e d , r u n o u tosv e r B f t i n l e n g t h s h o u l d be one pipe size larger than called for in this table. $Do not use Column D for pressure drops less than l/24 psi/100ft of equivalent run except on sizes 3 in. and larger. Use Chart 26, page 84. Fmrn Hcatina Ventilntina Air Conditioning Guide, Used by permission. 1959. Use of Table 31 Example I - Determine Pressure Drop for Sizing Supply * and Return Piping Given: Two-pipe low pressure steam system Initial steam pressure - 15 psig Approximate supply piping equivalent length - 500 ft .Approximate return piping equivalent length - 500 ft Find: 1. Pressure drop to size supply piping 2. Pressure drop to size return piping Solution: 1. Refer to TnOle 31 for an initial steam pressure of 15 psig. The total pressure drop should not exceed 3.75 psi in the supply pipe. Therefore, the supply piping is sized for a total pressure drop of 3.75 or 3/4 psi per 100 ft of equivalent pipe. 2. .\lthough ‘v, psi is intlicatctl in .Strj> 1, Item 4 under the two-pipe low pressure system recommends a maximum of I/ psi for return piping. Therefore, use I/* psi per 100 ft of equivalent pipe. ’ TABLE 31-TOTAL PRESSURE DROP FOR TWO-PIPE LOW PRESSURE STEAM PIPING SYSTEMS Friction Rate fixainple 2 illustrates the method used to tleterlriction rate for sizing pipe when the total system pressure drop recommendation (supply prcssure drop plus return pressure drop) is known and the approximate equivalent length is known. lnirie tlw ( p s i ) (prig) Example 2 2 5 - Determine Friction Rate Civcn: Four systems liquivnlent length of each system - 400 ft Total pressure drop of systems - I/?, :v,, I, and 2 psi 1% 2% 3 vi 10 I5 ‘/a 1‘/4 2 ‘/!I 3% Fintl: I. Size of largest pipe not exceeding tlcsign friction rate. 2. Swam velocity in pipe. Find: Friction rate for each system Solution: SYSTEhl NUMl3ER (psi) 1% SYSTEM EQUIV. LENGTH TOTAL SYSTEXI PRESS. DROP (4 (Psi) Solution: 1. Enter Idtorn of Clrnrt 26 at 67.50 Ib/hr anti proceed vutically to the 100 psig line (dotted line in Clinrt -16). Then move ol,liquely to the 0 psig line. From this point proceed vertically up the chart to the smallest pipe size not csccetling 2 psi per 100 ft of equivalent pipe and read S!/, in. FRICTION R,\TE FOR PIPE SIZING (per 100 ft) 400 (400/ 100) (x) = % 2. The velocity of steam at 0 psig ~1s read from Clrnrt -76 is 16,000 fpm. Enter the left side of C/r& 2i at 16,000 fpm. I’roccctl ol~liquely downward to the 100 psig line andhorizontally across to the right sitlc of the chart (dotted lint in C/la,l 37). The velocity at 100 psig is 6100 fpm. x = I/* 400 (400/ 100) (x) = :fi 400 (400/ 100) (x) = 1 400 (400/ 100) (x) = 2 x =siti x =!A Exnnzple f illustrates a design problem for sizing pipe on a low pressure, vacuum return system. x = 1/x Example 4 - Sizing Pipe for a Low Pressure, Vacuum Use of Charts 26 and 27 Example 3 -Determine Velocity Return System Steam Supply Main and Final Given: Six units Steam requirement per unit - 72 lb/hr Layout as illustrated in thru 98 Threaded pipe and fittings Low pressure system - 2 psi Given: Friction rate - 2 psi per 100 ft of equivalent pipe Initial steam pressure - 100 psig Flow rate - 6750 lb/hr 72 72 72 72 Frc.96 - Low PRE~SURESTEAM 72 SUPPLY MAIN 72 3-90 I’r\K’l‘ 3 . 1’ll’lNG I’rcssurc drop for the supply main is equal lent lcngttl times pressure tlrop per 100 It: DESIGN to the equiva- 167.5 X .25/ IO0 = .42 psi This is within the rccommcntfctl (I psi) for the sl;pply. The I)ranch connection for I;ig. ner at the same friction rate. 6’ maximum pressure drop 97 is sired in a similar man- From T&/r 30 the hori/.ontal runout pipe size for a load of i2 II) is 2l,/2 in. ant1 the vertical riser size is 2 in. Convert all the fittings to equivalent pipe lengths. and add to the actual pipe length. Equivalent F IG. 97 - Low RUNOUT PKESSURE Find: Size of pipe and total pressure Note: Total pressure drop exceed one-half the initial drop is required for quiet AND RISER drop in the system should pressure. A reasonahly operation. pipe lengths I - 2th in. 4.5” cl1 1 - 21/ in. 90” ell I - 2 in. 90” ell I - 2 in. gate valve Actual pipe length Total never small Solution: Determine the design friction rate hy totaling the pipe length and adding 50% of the length for fittings: equivalent 3.2 4.1 3.3 2.3 11.0 length 23.9 ft. I’rcssure drop for branch runout 23.9 X .23/lOO . and riser is = ,060 psi The vacuum return main is sized from Tab/e 29 by starting at the last unit “G” and adding each additional load hetween unit “G” and the boiler. Each riser - 52 lb per hr, s/11 in. II5 + 11 + 133 = 259 259x .50= 130 389 ft equiv length SECTION Check pipe sizing recommendations for maximum friction rate from “Two-Pipe Vacuum System,” Ztem 2, l/8-1/? psi. Check Tab/e 31 to determine recommended maximum pressure drop for the supply and return mains (I/? psi for each). STEAM LOAD PIPE SIZE (in.) Oh/W Design friction rate = l/3.89 X (l/2 + l/2) = l/4 psi per 100 ft. The supply main is sized hy starting at the last unit “G” and adding each additional load from unit “G” to the boiler; use Table 28. The following tabulation results: SECTION STEAM LOAD P/W F-G E-F D-E C -D IS-C A - n 72 144 216 288 360 432 PIPE SITE (in.) 1% 2 2 2’/ 3 3 Convert the supply main fittings to equivalent lengths of pipe and add to the actual pipe length, Table II, page 17. Equivalent pipe lengths 1 - 1% in. side outlet tee 2 - 1% in. ells 1 - 2 in. reducing tee i - 2 in. run of tee 2 - 2 in. ells 1 - 21/~ in. reducing tee I - 3 in. reducing tee 2 - 3 in. ells 1 - 3 in. run of tee Actual pipe length Total equivalent length 7.0 4.6 4.7 3.3 6.6 5.6 7.0 15.0 5.0 115.0 167.5 ft. Convert the return main fittings to equivalent pipe lengths and add to the actual pipe length, Table 11, page 17. ’ Equivalent pipe lengths 1 _ 3h in. run of tee 5 - % in. 90” ells 1 - I in. reducing tee 1 - 1 in. run of tee 2 - 1 in. 90” ells 1 - II/~ in. reducing tee 3 - I IA in. 90” ells 1 - 1% in. run of tee .Actual pipe length Total equivalent length 1.4 7.0 2.3 1.7 3.4 3.1 6.9 2.3 133.0 161.1 ft. Pressure drop for the return equals 161.1 X .25/100 = ,404 psi Total return pressure drop is satisfactory since it is within the recommended maximum pressure drop (t/B - 1 psi) listed in the two-pipe vacuum return system. The total system pressure drop is equal to ,420 + .060 + ,404 = ,884 psi . This total system pressure drop is within the maximum 2 psi recommended (1 psi for supply and 1 psi for return). (:I-I.\l’~I‘El< 4. S’I‘E.\XI 3-91 I’II’ING 72 72 FIG. 98 -Low 72 PRESSURE ‘ING APPLICATION The use and selection of steam traps, and condensate and vacuum return pumps are presented in this section. Also, various steam piping diagrams are illustratcd to familiarize the engineer with accepted piping practice. STEAM TRAP SELECTION The primary function of a steam trap is to hold steam in a heating apparatus or piping system and allow condensate and air to pass. The steam remains trapped until it gives up its latent heat and changes to condensate. The steam trap size depends on the following: 1. Amount of condensate to be handled by the trap, Ib/hr. 2. Pressure differential between inlet and discharge at the trap. 3. Safety factor used to select the trap. Amount of Condensate The amount of condensate depends on whether the trap is used for steam mains or risers, or for the heating apparatus. The selection of the trap for the steam mains or risers is dependent on the pipe warm-up load and the radiation load from the pipe. Warm-up load is the condensate which is formed by heating the pipe surface when the steam is first turned on. For practical purposes the final temperature of the pipe is the steam temperature. Warm-up load is determined from the following equation: c = ‘V x (I/.- ti) x .114 I A, x T 72 72 VACUUM 72 RETURN where: C, = Warm-up condensate, lb/hr W = Total weight of pipe, lb (Tables 2 and 3, pages -7 and 3) I~ = Final pipe temperature, F (steam temp) ti = Initial pipe temperature, F (usually room tern p) .114 = Specific heat constant for wrought iron or steel pipe (.092 for copper tubing) h, = Latent heat of steam, Ktu/lb (from steam tables) T = Time for warm-up, hr The radiation load is the condensate formed by unavoidable radiation loss from a bare pipe. This load is determined from the following equation and is based on still air surrounding the steam main or riser: c = L X I\ X (tf - ti) 2 f/l where: C, = Radiation condensate, lb/hr L = Linear length of pipe, ft K = Heat transmission coefficient, Btu/(hr) (linear Et) (deg F diff between pipe and surrounding air), Table jf, Part I t,, ti, II, explained previously The radiation load builds up as the warm-up load drops off under normal operating conditions. The peak occurs at the mid-point of the warm-up cycle. Therefore, one-half of the radiation load is added ‘to the warm-up load to determine the amount of condensate that the trap handles. 3-a PAR?‘ 3. PlPlNG D E S I G N L Pressure Differential T!le Ilressure tlilfercntial. across the trap is determined at dcsigri conditions. IL‘ a vacu~~m exists on the tliscllarge side of the trap, the vacuuln is added to the inlet side pressure to determine the differential. Safety Factor Good design practice dictates the use of safety factors in steam trap selection. Salety factors from 2 to 1 to as high as 8 to 1 may be requirctl, and for the following reasons: 1. The steam pressure at the trap inlet or the back prcssurc at the trap discharge may vary. This changes the steam trap capacity. 2. If the trap is sized Eor normal operating load, condensate may back up into the steam lines or apparatus during start-up or warm-up opera( tion. 3. II the steam trap is selected to discharge a full and continuous stream of water, the air could not be vented from the system. The following guide is used to determine the safety factor: DESIGN Draining steam main Draining steam riser SAFETY Before reducing valve Before shut-off valve (closed part of time) Draining coils Draining apparatus 3 to 1 3 to 1 3 to 1 5 - Steam Trap Selection for Solution: I. The warm-up equation: Dripping Main to Return Line Given: Steam main - 10 in. diam steel pilje, 50 ft long Steam pressure - 5 psig (227 F) Root11 temperature - 70 F tli) (stealn main in space) Warm-up time - 15 minutes Steam trap to drip main into vacuum rettirn line (2 in. vacuum gage design) load is determined from the following iv x pr - ti) x .I I4 c, = llL x T where: 1%‘~ 40.48 lb/ft X 50 ft (Tnble 2) tf= 227 F ti= 7OF hL = 960 Btu/lb (from steam tables) T = .25 hr c, = 2024 x (227 - TO) x ,114 960 x .?5 = 156 lb/hr of condensate 2. The radiation load is calculated by using the following equation: L x I\’ x (I, - li) c, = ‘11 where: L = 50 ft K = 6.41 Btu/(hr) (linear foot) (deg F diff between pipe and air) from Table 5-I, Part I tf = 227 F ti = 70 F tzL = 960 Rtu/ll, (from steam taD1r.s) c =50 X 6.41 x (227- 70) 2 960 = 52 lb/hr‘of condensate 3. When the steam trap is to be used in a high pressure system, determine whether or not the system is to operate und& low pressure conditions at cerain intervals such as night time or weekends. If this .-condition is likely to occur, then an additional safety factor should be considered to account for the lower pressure drop available during night time operation. Example 5 illustrates the three concepts mentioned previously in trap selection-condensate handled, pressure differential and safety factor. Example Warm-up load. Radiation load. Total condensate load. Specifications for steam trap at end of supply main. FACTOR 3 to 1 2 to 1 2 to 1 3 to 1 Between boiler and end of main Find: 1. 2. 3. 4. Supply The equal load. total condensate load for steam trap selection is to the warm-up load plus one half the radiation Total condensate load = C, + (I$$ X Cr) = 150 + (I/? x 52) = 176 lb/hr 4. Steam trap selection is dependent on three factors: condensate handled, safety factor applied to total condensate load, and pressure differential across the steam trap. The safety factor for a steam trap at the end of the main is 3 to 1 from the table on this page. Applying the 3 to 1 safety factor to the total condensate load, the steam trap would be specified to handle 3 X 176 or 528 Ib/hr of condensate. The pressure differential across the steam trap is determinecl I)y the pressure at the inlet and diskharge of the steam trap. Inlet to trap = 5 psig Discharge of trap = 2 in. vacuum (gage) \\‘hen the discharge is under vacuum conditions, the discharge vacuum is added to the inlet pressure for the total pressure differential. Pressure differential = 6 psi (approx) Therefore the steam trap is selected for a differential pressure of G psi and 528 Ib/hr of condensate. . s#- AIR VENT I- VALVE ACCESS PLUG- THERMOSTATIC DISC ELEMENT AIR PASSAGE INLET VALVE AN0 ORIFICE I STEAM TRAP TYPES The types of traps commonly used in steam systems arc: Float Flash Y~ermostatic Implllsc _ loat & thermostatic Lifting Upright bucket boiler return or inverted bucket alternating receiver The description and use of these various traps are presented in the following pages. Float Trap The discharge from the float trap is generally continuous, This type (Fig. 97) is used for draining condensate from steam headers, steam heating coils, and other similar equipment. When a float trap is used for draining a low pressure steam system, it should be equipped with a thermostatic air vent. Thermostatic Trap The discharge from this type of trap is intermittent. Thermostatic traps are used to drain condensate from radiators, convcctors, steam heating cnils, unit heaters and other similar equipment. .iners are normally installed on the inlet side of the steam trap to prevent dirt and pipe scale from OUTLET entering the trap. On traps used for radiators or convectors, the strainer is usually omitted. I;ig. 100 shows a typical thermostatic trap of the bellows type and I;ig. 101 illustrates a disc type thermostatic trap. When a thermostatic trap is used for a heating apparatus, at least 2 ft of pipe are provided ahead of the trap to cool the condensate. This permits condensate to cool in the pipe rather than in the coil, and thus maintains maximum coil efficiency. Thermostatic traps are recommended for low pressure systems up to a maximum of 15 psi. When used in medium or high pressure systems, they must be selected for the specific design temperature. In addition, the system must be operated continuously at that design temperature. Tlzis menns no night setback. Float and Thermostatic Trap This type of trap is used to drain condensate from blast heaters,,steam heating coils, unit heaters and other apparatus. This combination trap (Fig. 102) is used where there is a large volume of condensate which would not permit proper operation of a thermostatic trap. Float and thermostatic traps are used in low pressure heating systems up to a maximum of 15 psi. THERMOSTATIC DISC ELEMENT\ ERMOSTATIC INLET I OUTLET FIG. IOO-THERMOSTATIC TRAP , BELLOWSTYPE L! FLOAT L VALVE AND ORfFlCE Frc.102 -TYPIC.ALFLOATANUTHEKMOSTATICTRAP I OUTLET INLET BLOWOFF / FIG. 105 - INVERXD I~UCKET TRAP WITH GUIDE For medium and high pressure systems, the same limitations as outlinctl for thermostatic traps apply. ‘Ipright Bucket Trap The discharge of condensate from this trap (Fig. is intermittent. h tlilfcrential pressure of at least 1 psi between the inlet and the outlet of the trap is normally required to lift the condensate from the bucket to the discharge connection. Upright bucket traps are commonly used to drain condensate and air from the blast coils, steam mains, unit heaters and other equipment. This trap is well suited for systems that have pulsating pressures. 103) ing condensate from steam lines or equipment where an abnormal amount of air is to be discharged and where dirt may drain into the trap. Flash Trap The discharge from a Ilash trap (Fig. 106) is intermittent. This type of trap is used only if a pressure differential of 5 psi or more exists between the steam supply and condensate return. Flash traps may be used with unit heaters, steam heating coils, steam 4 lines and other similar equipment. Impulse Trap Inverted Bucket Trap The discharge from the inverted bucket trap (Figs. 104 and 105) is intermittent and requires a differential pressure between the inlet and discharge of the trap to lift the condensate from the bottom of the trap to the discharge connection. Bucket traps are used for draining condensate and air from blast coils, unit heaters and steam heating -oils. Inverted bucket traps are well suited for drain- t Under normal loads the discharge from this trap (Fig. IO/‘) is intermittent. When the load is heavy, however, the discharge is continuous. This type of trap may be used on any equipment where the pressure at the trap outlet does not exceed 25% of the inlet pressure. Lifting Trap The lifting trap (Fig. 108) is an adaption of the upright bucket trap. This trap can be used on all steam heating systems up to 150 psig. There is an OUTLET ADJUSTABLE ORIFICE VALVE AN0 ORIFICE OUTLET AIR VENT RE -EVAPORATING CHAMBER DRAIN PLUG t INLET F1c.104 -IINVERTEDBUCKETTRAP AND SLOWDOWN VALVE OPENING FIG. 106 -FLASH TRAP . VENT TO ATMOSPHERE- FLOAT INLET AND t Flc. 109 - UOIL.ER OUTLET RIXIJRN TRAP OR ALTERNAWNG RECEIVER FIG. 105 - 1xr~u1.s~ TRAP of condensate return pumps are the rotary, screw, turbine and reciprocating pump. auxiliary inlet for high llressure steam, as illustrated il- *he figure, to force the condensate to a point Je the trap. This steam is normally at a higher ;r pressure than the steam entering at the regular inlet. Boiler Return Trap or Alternating Receiver This type ol‘ trap is WA to return condensate to a low pressure boiler. The boiler return trap (Fig. fO9) does not hold steam as do other types, but is an adaption of the lilting.trap. It is used in conjunction with a boiler to prevent flooding return mains when excess pressure prevents condensate from returning to the boiler by gravity. The boiler trap collects condensate and equalizes the boiler and trap pressure, enabling the condensate in the trap to How back to the boiler by gravity. CONDENSATE RETURN PUMP Condensate return pumps are used for low sure, gravity return heating systems. They are mally of the motor driven centrifugal type and 2 :eiver and automatic float control. Other HIGH PRESSURE INLET FIG. 108 - LIITIW TRAP presnorhave types The condensate receiver is sized to prevent large Auctuations of the boiler water line. The storage capacity of the reccivcr is approximately 1.5 times the amount of condensate returned per minute, and the condensate pump has a capacity of 2.5 to 3 times normal flow. This relationship of pump and receiver to the condensate takes peak condensation load into account. VACUUM PUMP Vacuum pumps are used on a system where the returns are under a vacuum. The assembly consists of a receiver, separating tank and automatic controls for discharging the condensate to the boiler. Vacuum pumps are sized in the same manner as condensate pumps for a delivery of 2.5 to 3 times the design condensing rate. PIPING LAYOUT Each application has its own layout problem with regard to the equipment location, interference with structural members, steam condensate, steam trap and drip locations. The following steam piping diagrams show the various principles involved. The engineer must use judgment in applying these principles to the application. Gate valves shown in the diagrams should be used in either the open or closed position, ne?Je?’ fog’ tA~~ttlj?z~. Angle and globe valves are recommended lor throttling service. In a one-pipe system gate valves are used since they do not hinder the How of condensate. Angle valves may be used when they do not restrict the ffow of condensate. All the figures show screwed fittings. Limitations for other fittingi are described in Chnfiter 1. + STEAM MAIN RISER blRlP T O CONDENSATE RETURN FIG. 1 IO - CONNECTION TO DRIITEI)R~srx Steam Riser Figz~es / 10 c1)7d I II illustrate steam supply risers connected to mains with runouts. The runout in Fig. 110 is connected to the bottom portion of the main and is pitched toward the riser to permit condensate to drain from the main. This layout is used only when the bser is dripped. If a dry return is used, the riser is dripped thru a steam trap. If a wet return is used, the trap is omitted. Fig. 1 I I shows a piping diagram when the steam riser is not dripped. In this instance the runout is connected to the upper portion of the steam’ main and is pitched to carry condensate from the riser to the main. Prevention of Water Hammer If the steam main is pitched incorrectly when the riser is not dripped, water hammer may occur as illustrated in Fig. 112. Diagram “a” shows the runout partially filled with condensate but with enough space for steam to pass. As the amount of condensate increases and the space decreases, a wave notion is started as illustrated in diagram “b”. As the wave or slug of condensate is driven against the turn in the pipe (diagram “c”), a hammer noise is FIG. 112 -WATER HAMMER caused. This pounding may be of sufficient force to split pipe fittings and damage coils in the system. The following precautions must be taken to prevent water hammer: 1. Pitch pipes properly. . 2. Avoid undrained pockets. 3. Choose a pipe size that prevents high steam velocity when condensate flows opposite to the steam. Runout Connection to Supply Main Figure 113 illustrates two methods of connecting runouts to the steam supply main. The method using a 45” ell is somewhat better as it offers less resistance to steam flow. Expansion and Contraction Where a riser is two or more floors in height, it should be connected to the steam main as shown RISER (NOT DRlP,PED%D // STEAM MAIN FIG. 111 - ACCEPTABLE CONNECTION TO R I S E R (N O T D R I P P E D) RECOMMENDED FIG. 113 - RUNOUT CONNECTIONS MOVEMENT t RAIN AIR LINE, FIG. Ii4 - RISER CONNECTEI)TO ALLOW FOR in Fig. 114. Point (A) is subject to a twisting movemcnt as the riser moves up and down. Figure 115 shows a method of anchoring the steam riser to allow for expansion and contraction, Movement occurs at (A) and (B) when the riser moves up and down. HANDHOLE’ PLUG IN TEE FOR CLEANOUT FI G . 117 - RET~JRN MA I N LO O P Obstructions Steam supply mains may be looped over obstructions if a small pipe is run below the obstruction to take care of condensate as illustrated in Fig. 116. The reverse procedure is followed for condensate return mains as illustrated in Fig. 117. The larger pipe is carried under the obstruction. Dripping BOILER REDUCING WATER COUPLING Riser A steam supply main may be dropped abruptly to a ’ <:er level without dripping if the pitch is downWL . . . When the steam main is raised to a higher level, it must be dripped a S illustrated in Fig. 318. FIG. 118 - DRIPPING S TEAM MAIN I‘his diagram shows the steam main dripped into a wet return. Figr~x I19 is one method of dripping a riser thru RISER STEAM TRAP MOVEMENT CONDENSATETE-RETURNMAIN FI G . 115 - RISKR AN C H O R I’,\K 3-98 I I. 3. I’ll’lh‘(; ,-CONTROL VALVE (NOTE INVERTED LIFT FITTING OR ELBOWS NoTE ’ -.,\d I)bSIGN 3) r STRAINER . I TOTAL LIFT LIMIT 5 FT II f IN. PET COCK‘----h A WHEN END OF SUPPLY MAIN,SEE FIG. 125 \-LIFT PIPE $“DIA. OF VACUUM RETURN 15’ CHECK VALVE FOR BREAKING VACUUM % I VACUUM RETURN F R O M -NO JOINTS LIFT PIPE LIFT IN FITTING DIRT LEG 15’ CHECK FIG . 120 - ONE-STEP CONDENSATE Lggif+?? -y LIFT CONDENSATE RETURN MAIN a steam trap to a dry return. The runout to the return main is pitched toward the return main. Vacuum Lift As described under vacuum systems, a lift is sometimes employed to lift the condensate up to the inlet of the vacuum pump. Figs. 120 and 122 show a one-step and two-step lift respectively. The onestep lift is used for a maximum lift of 5 Et. For 5 t o 8 Et a two-step lift is required. SOTES: 1. Flange or union is located to facilitate coil removal. 2. Flash trap may be used if pressure differential between steam and condensate return exceeds 5 psi. 3. When a bypass with control is required, see F&. 126. 4. Dirt leg may be replaced with a strainer. If so, tee on drop can be replaced by a reducing ell. 5. The petcock is not necessary with a bucket trap or any trap which has provision for passing air. The great majority of high or medium pressure returns end in hot wells or deaerators which vent the air. FIG. Steam 122 -HIGH OR MEDIUM PRESSIJRE COIL PIPING Coils Fig2ive.s 112 t/lrz~ 131 s h o w methods o f p i p i n g THAN 2 STEPS 2ND STEP I ST STEP q L/2 /iI *MAXIMUM LENGTH (A) VACUUM RETURN / FIG. 12 1 - TWO-STEP CONDENSATE LIFT - steam coils in a high or low pressure or vacuum steam piping system. The following general rules are applicable to piping layout for steam coils used in all systems: 1. Use full size coil outlets and return piping to the steam trap. 2. Use thermostatic traps for venting only. 3. Use a 15’ check valve only where indicated on the layout. 4. Six the steam control valve for the steam load, not for the supply connection. 5. Provide coils with air vents as required, to eliminate non-condensable gases. 6. Do not drip the steam supply mains into coil sections. . ,-CONTROL VALVE ( NOTE 2) FLANGE OR UNION CONTROL VALVE (NOTE 21 MAIN ‘-WHEN END OF SUPPLY MAIN SEE FIG 125 AM SUPPLY MAIN 125 WHEN END M SUPPLY MAIN 15’ CHECK VALVE THERMOSTATIC TRAP FLOAT AND’ THERMOSTATIC TRAP RETURN MAIN NOTES: 1. Flange or union is located to facilitate coil removal. 2. When a bypass with control is required, see Fig. 126. 3. Check valve is necessary when more than one unit is connected to the return line. 4. Dirt pocket is the same size as unit outlet. If dirt pocket is replaced by a strainer, replace tee w i t h a reducing ell from unit outlet to trap six. CONDENSATE RETURN MAIN NOTES: 1. Flange or union is located to facilitate coil removal. 2. When a bypass with control is specified, see Fig. 126. 3. Check valve is necessary when more than one unit is connected to the return line. 124 - VACUUM FIG. FIG. 123 - SINGLE COIL P Y Low PRESSUKE SY S T E M STuni COIL I’II’IN(; PIPING GRAVITYRETURN 7. Do not pipe tempering coils and reheat coils to a common steam trap. 8 . Multiple coils m a y be piped lo a common steam trap if they have the same capacity and the same pressure drop and if the supply is regulated by a control valve. STEAM SUPPLY , FLOAT B THERMOSTATIC GATE VALVE (PLUG Piping Single TYPE) Coils Fiqu~ 122 illustrates a typical steam piping dia< gram for coils used in either a high or medium pressure system. If the return line is designed for low pressure or vacuum conditions and for a pressure differential of 5 psi or greater from steam to condensate return, a flash trap may be used. Low pressure steam piping for a single coil is illustrated in Fig. 123. This diagram shows an open air relief located after the steam trap close to the unit. This arrangement permits non-condensable gases to vent to the atmosphere. Fig1rt.e 12-f shows the f>iping layout for a steam coil in a V;ICIIII~ system. h 15” check valve is used to eq ilalile the vacuum across the steam trap. CONDENSATE GATE , DRIP L I N E - , I , TO UNIT. REOUIRED ON LOW PRESSURE GRAVITY RETURN SYSTEMS. VALVE-I CONDENSATE RETURN MAIN Y NOTES: I. A bypass is necessary around trap and vaivcs when continuous operation is necessary. 2. Bypass to be the same size as trap orifice I)ut never less than I/ inch. 1;1(;. 125 - DRIPI~ING SIXAM SIJPPL\~ KI’TURN 1‘0 C O N D EN S A T E I’ \K’l :I. I’I I’I NC; I)I:SI<;N 3-too ,-CONTROL VALVE ,-GATE VALVE (NOTE 2) ,-STRAINER CONTROL VALVE .r- G A T E V A L V E ‘, -REFER TO FIG. 125 WHEN DRIPPING STEAM SUPPLY MAIN TO CONDENSATE RETURN CONDENSATE RETURN ’ rIS” C H E C K V A L V E FOR NOTES: 1. Flange or union is located to facilitate coil removal. 2. A bypass is necessary around valves and strainer when continuous operation is necessary. 3. Bypass to be the same size as valve port hut never less than IA inch. F1c.126 - BYI~ASSWITH Dripping Steam Supply MANUALCONTROL “\~ BREAKING +“PET COCK FOR CONTINUOUS IS’ DIRT LEG VACUUM VENT CHECK VALVE (6”) GATE Main FLOAT OR BUCKET TRAP (NOTE 3 B 4) A typical method of dripping the steam supply main to the condensate return is shown in I;ig. 125. /-Y--l STEAM SUPPLY VALVE CONDENSATE NOTES: 1. Flange or union is located to facilitate coil removal. 2. When bypass control is required, see Fig. l-36. 3. Flash trap can be used if pressure differential between supply and condensate return exceeds 5 psi. 4. Coils with different presume drops require individual traps. 5. Dirt pocket may be replaced by a strainer. If so, tee on drop can be replaced by a reducing ell. 6. The petcock is not necessary with a bucket trap or any trap which has provision for passing air. The great majority of high pressure return mains terminate in hot wells or deaerators which vent the air. FIG. CHECK VALVE /,-BUCKET Steam TRAP c LOCATED BELOW OUTLET, DRAIN VALVE (GATE.NOTE 2) ’ NOTES: 1. Flange or union is located 2. To prevent water hammer, ting steam. 3. Do not exceed one foot of and return main for each ential. FIG. 1!?-&NDENSA.I.E to facilitate coil removal. drain coil before admitlift between trap discharge pound of pressure differ- LIFT TOOVERHEAD RETURN 128 -MULTIPLE Bypass COIL HIGH PRESSURE PIPING Control Frequently a bypass with a manual control valve is required on steam coils. The piping layout for a control bypass with a plug type globe valve as the manual control is shown in Fig. 126. Lifting Condensate to Return Main A typical layout for lifting condensate to an overhead return is described in I;ig. 1’7. The amount of lift possible is determined by the pressure differential between the supply and return sides of the system. The amount of lift is not to exceed one foot for each pound of pressure differential. The maximum lift sl~o~~ld not exceed 8 ft. I (:11,\1’~1‘1~~1< I . 3-101 S’I‘I~.\XI I’II’lN(. ,/-CONTROL (NOTES N O T E I -._ .k/ . \ CONTROL VALVE (NOTES283) VALVE 283) STRAlNER 8-1i’ ,I GATE VALVE STEAM m-IS’CHECKV A L V E WHEN SUPPLY SUPPLY REFER TO FIG. 125 WHEN DRIPPING SUPPLY TO RETURN. DRIPPING TO RETURN. ----THERMOSTATIC ,OPEN TO TRAP THERMOSTATIC (4”) AIR RELIEF ATMOSPHERE TRAP EQUALIZING (4”) VACUUM CONDENSATE &CHECK VALVE TRAP (NOTE TRAP (NOTE 41 4) GATE VALVE _i r/ LCONDENSATE RETURN XOTES: 1. Flange or union is located to facilitate coil removal. 2. See Fig. 131 when control valve is omitted on multiple coils in parallel air flow. 3. When bypass control is required, see Fig. 126. 4. Coils with different pressure drops require individual traps. NOTES: 1. Flange or union is located to facilitate coil removal. 2. See Fig. 131 when control valve is omitted on multiple coils in parallel air flow. 3. When bypass control is required, see Fig. 126. 4. Coils with different pressure drops require individual traps. FIG. 130 -MULTIPLE COIL SYSTEMS FIG . 129 -MULTIPLE COIL Low PRESSURE Low PRESSURE VhcuuM P~privc PIPING FREEZE-UP Piping Multiple Coils Figures 117s tl21.u 131 show piping layouts for high pressure, low pressure and vacuum systems with multiple coils. If a control valve is not used, each coil must have a separate steam trap as illustrated in Fig. /31. This particular layout may be used for a low pressure or vacuum system. If the coils have different pressure drops or capacities, separate traps are required with or without a control valve in the system. Boiler Piping I ; i g t l w Ii2 illustrates a suggested layout for a steam plant. This diagram sho\vs parallel boilers and a single boiler usirq a “Hartford Return Loop.” PROTECTION tVhen steam coils are used for tempering or preheating outdoor air, controls are required to prevent freezing of the coil. I’n high, medium, low pressure and vacuum systems, an immersion thermostat is recommended to protect the coil. This protection device controls the fan motor and the outdoor air damper. The immersion thermostat is actuated when the steam supply fails or when the condensate temperature drops below a predetermined level, usually 120 F to 150 F. ‘I‘lie thermostat location is shown in Fig. 133. The 15” check valve shown in the various pifiing diagrams provides a means of equalizing the pressure within the coil when the steam supply shuts off. This check valve is used in addition to the immersion thC‘rn1ostat. The petcock for continuous venting removes non-condensable gases from the coil. UNIT 1 / 1 I STEAM SUPPLY /’ MAIN IMMERSION THERMOSTAT ( SEE NOTE) WHEN DRIPPING ST EAM SUPPLY MAIN TO CONDENSATE RETURN 4 12”MIN. b-d CHECK VALVES KOTE: Immersion thermostat is for control of outdoor air dampers and fan motor. Thermostat closes damper and shuts ofi fan when condensate temperature drops below a predet rmined level. ” NOTE: Flange or removal. F1c;.131--Low union is located to facilitate coil FIG. 133 -FREEZE-UP PROTECTIONFORHIGH,~IEDIUM, Low PRESSURE,ANDVACUUM SYSTEMS PREWJREOR Vxuuhf SYSTEM STEAMTRAPS below 165 F. Condensate drains are limited to a 5 ib pressure. Non-condensable gases can restrict the How of condensate, causing coil freeze-up. On a low pressure and vacuum steam heating system, the immersion thermostat may be replaced b y a condensate drain with a thermal element (Fig. 134). The thermal element opens and drains the coil when the condensate temperature drops CONDENSATE OUT SCREEN 81 METAL STEAM TRAP SEE NOTE 8 DETAIL A SITE DRAIN NOTE: Condensate temperature drain drops drains below coil when condensate a predetermined level. F1c.134 - FREEZE-UP PROTECTION FOR AND VACUUM SYSTEMS Low PRESSURE ‘l‘lie I’ollowing a r e g e n e r a l Icconimcntlations ill laying out systems hntlling outdoor air hclow 35 F: I. Do n o t use ovcrluxtl returns lrom the hc;lting unit. 2. I’rcssurc c o n t r o l s arc 11ot recomn~entletl si1ic.c thy do liot n e c e s s a r i l y rellect actual contli- tions. For example, it is pmsiblc for the coil to Iwcome air huntl and have pressure hut no steani. Also, the steam trap may be pluggd. Prcssurc controls are slow acting by comprison with thermostatic controls. 3. USC ;L strainer in the supply line ahcad of the trap. ;~ntl 2 dirt leg Part 4 REFRIGERANTS, BRINES, OILS 4-l CHAPTER 1. REFRIGERANTS This chapter provides information concerning the refrigeration cycles and characteristics of the commonly used refrigerants and their selection for use in air conditioning applications. To provide .refrigeration, ,a refrigerant may be utilized either: 1. In conjunction with a compressor, condenser and evaporator in a compression cycle, or 2. With an absorbent in conjunction with an absorber, generator, evaporator, and condenser in an absorption cycle. The refrigerant absorbs heat by evaporation generally at a low temperature and pressure level. Upon condensing at a higher level, it rejects this heat to an.1 available medium, usually water or air. ~_ a compression system the refrigerant vapor is increased in pressure from evaporator to condenser pressure by the use of a compressor. In an absorption system the increase in pressure is produced by heat supplied from steam or other suitable hot fluid which circulates thru a coil of pipe. The absorber-generator is analogous to a compressor in that the absorber constitutes the suction stroke and the generator the compression stroke. The evaporator spray header corresponds to the expansion valve. The evaporator and condenser are identical for both compression and absorption systems. This chapter includes a discussion of the refrigeration cycle, refrigerant selection, and the commonly used refrigerants as well as tables indicating their characteristics and properties. REFRIGERATION ABSORPTION CYCLES CYCLE The absorption refrigeration cycle utilizes two phenomena: 1. The absorption solution (absorbent plus refrigerant) can absorb refrigerant vapor. 2. The refrigerant boils (flash cools itself) when subjected to a lower pressure. These two phenomena are used in the lithium bromide absorption machine to obtain refrigeration by using the bromide as an absorbent and water as a refrigerant. Water is sprayed in an evaporator which is maintained at a high vacuum. A portion of the water flashes and cools that which remains. The water vapor is absorbed by a lithium bromide solution in the absorber. The resulting solution is then heated in the generator to drive off the water vapor which is condensed in the condenser. The water is returned to the evaporator, completing the cycle. Figure 1 illustrates the absorption cycle. Figure 2 illustrates the cycle plotted on the equilibrium diagram with numbered points representing pressures, temperatures, concentrations in the cycle. On the lower left side of Fig. 1 is the absorber partially filled with lithium bromide solution. On the lower right side is the evaporator containing water. A pipe connecting the shells is evacuated so that no air is present. The lithium bromide begins to absorb the water vapor; as the vapor is absorbed, the water boils, generating more vapor and causing the remainder of the water to be cooled. LLEO ER I SOLUTION I I 6 \ I GENERATOR PUMP F IG . 1 - AB S O R P T I O N R E F R I G E R A T I O N Since the water can vaporize more easily if it is being sprayed, a pump is used to circulate the water from the bottom of the evaporator to a spray header at the top. An evaporator tube bundle is located under the evaporator spray header; water inside the tubes, returning from the air conditioning coils or other load, is Hash-cooled by the water on the outside of the evaporator tubes. The lithium bromide solution absorbs water vapor easier if it is sprayed; therefore, a pump is used to circulate the solution from the bottom of the absorber to a spray header at the top of the absorber. As the lithium bromide continues to absorb water vapor, it becomes diluted, and its ability to absorb additional water vapor decreases. The weak solution is pumped to the generator where heat is applied by steam or other suitable hot fluid in the . C YCLE generator tube bundle to boil off the water vapor. The solution is concentrated and returned to the absorber. Since the weak solution going to the generator must be heated and the strong solution coming from the generator must be cooled, a heat exchanger is used in the solution circuit to conserve heat. Water vapor boiled from the solution in the generator passes to the condenser to contact the relatively cold condenser tubes. The vapor condenses in the condenser and returns to the evaporator so that there is no loss of water in the cycle. Before the condenser water goes thru the condenser tubes, it passes thru a tube bundle located in the absorber. Here it picks up the heat of dilution and the heat of condensation which is generated as the solution absorbs water vapor. 20.0 . 16.0 12.0 8.0 6.0 5.0 4.0 , 3.5 . 3.c , 2.: b - ) . I . 3 . 0 . .4 0 . .3 0 . .2 5 .2 0 _I 8 . . ,T COMPRESSION CYCLE The compression refrigeration cycle utilizes two phenomena: 1. The evaporation of a liquid rekigerant absorbs heat to lower the temperature ok its surroundings. 2. The condensation of a refrigerant vapor rejects heat to raise the temperature of its surroundings. The cycle may be traced from any point in the system. Figwe 3 is a schematic and Fig. f is a pressure-enthnlpy diagram of a compression cycle. I’;\I<‘I‘ 4-4 ,I. I~E~III(;I:l~;\N’I‘S, I1KINES, O I L S WATER-COOLED LIQUID STOP FIG . .~ 3 REFRIGERANT VALVE - RECIPROCATING COMPRESSION REFRIGERATION CYCLE Starting with the liquid refrigerant ahead of the evaporator at point A in both Fig. 3 and 4, the admission of liquid to the evaporator is controlled by an automatic throttling device (expansion valve) which is actuated by temperature and pressure. The refrigerant pressure is reduced across the valve from condenser pressure, point A, to the evaporator pressure, point B. The valve acts as a boundary between the high and low sides of the system. The pressure reduction allows the refrigerant to boil or vaporize. To support boiling, heat from the air or other medium to be cooled is transmitted to the evaporator surface and into the boiling liquid at a lower temperature. The refrigerant liquid and vapor passing thru the evaporator coil continues to absorb heat until it is completely evaporated, point C. Superheating of the gas, controlled by the expansion valve, occurs from C to D. The superheated gas is drawn thru the suction line into the compressor cylinder. The downstroke of the piston pulls a cylinder of gas thru the suction valve and compresses it on the upstroke, raising its temperature and pressure to point E. The pressure produced causes the hot gas to flow to the condenser. The compressor discharge valve prevents re-entry of compressed gas into the cylinder and forms a boundry between the high and low sides. In the condenser the condensing medium (air or water) absorbs heat to condense the hot gas. Liquid refrigerant is collected in receiver which may be combined with or separate from condenser. . CONDENSING EVAPORATION ENTHALPY FIG . 4 (ETUILB) - P RESSURE-E NTHALPY DIAGRAM, COMPRESSION CYCLE The liquid is then forced thru the liquid line to the expansion valve A to repeat the cycle. Liquid-Suction Interchangers Compressor ratings for Refrigerants 12 and 500 are generally based on 65 F actual suction gas’temperatures. When this suction gas temperature is not obtained at the compressor, its rating must be lowered by an appropriate multiplier. To develop the full rating, the required superheat which is over and above that available at the evaporator outlet may be obtained by a liquid-suction interchanger. I ENTHALPY (BTU/LB) SUPERHEATCF, FIG. 5 - EFFECT ON LIQ~JID-SUCTION INTERCHANGER COMPRESSION CYCLE OF The effect of a liquid-suction interchanger on the refrigeration cycle is shown on the pressureenthalpy diagram (Fig. 5). The solid lines represent’ the basic cycle, while the dashed lines represent the same cycle with a liquid-suction interchanger. The useful refrigerating effect with an interchanger is B’C rather than BC as in the basic cycle. Superheat increases the specific volume of the suction vapor to reduce the total weight of refrigerant circulated for a given displacement. It also increases the enthalpy of the vapor and may improve compressor volumetric efficiency. Provided thn heat absorbed represents useful refrigeration L I as liquid subcooling, the refrigerating effect per pound of refrigerant circulated is increased. With Refrigerants 22 and 717 volume increases faster than refrigerating effect; hence, superheating theoretically reduces capacity. With Refrigerants 12 and 500 the reverse is true, and superheating theoretically reduces both the cfm per ton and the power per ton. Figure 6 illustrates the loss due to superheating of the refrigerant vapor and the gain due to liquid subcooling. Net gain equals the gain minus the loss. REFRIGERANT PROPERTIES Refrigerant characteristics have a bearing on system design, application and operation. A refrigerant is selected after an analysis of the required characteristics and a matching of these require- Courtesy of ASKE Data Uook FIG. 6 - EFFECT OF LIQUID-SUCTION INTERCHANGER CAPACITY ON ments with the specific properties of the available refrigerants. Significant refrigerant characteristics are: 1. Flammability and Toxicity as they pertain to the safety of a refrigerant. The refrigerants treated in this chapter are classified in ASA H9.1 as Group 1, the least hazardous relative to flammability and explosiveness. The Underwriters Laboratories classification with respect to toxicity puts these refrigerants in Groups 4 to 6. The higher numbered groups in this classification are the least toxic. FigdYe 7 shows the structural formula of the refrigerant compounds treated in this chapter. The chlorine and fluorine elements in these refrigerants make them the least hazardous and least toxic respectively. 2. Miscibility of a refrigerant with compressor oil aids in the return of oil from the evaporator to the compressor crankcase in reciprocating machine applications. Centrifugal units have separate oil and refrigerant circuits. Some refrigerants are highly miscible with compressor oil. Refrigerants 12 and 500 and lubricating oils are miscible in any proportion; Refrigerant 22 is less miscible. The effect of miscibility in a refrigeration system is illustrated in Fig. 8. If Refrigerant I2 is placed - / OIL ABSORBS ALL OF THE REFRIGERANT 1. CL Refrigerant 11 CCl,,F F F H %+ 0,L ABSORBS REFRIGERANT EQUIVALENT TO 67% OF THE OIL WEIGHT - F C t CL OCL Refrigerant 12 CC&F, CL Refrigerant 22 CHCIF., F F c F Refrigerant c ‘c 41° CL CL 113 Refrigerant 114 C,,Cl,F, - H - hydrogen F - fluorine %C13F3 C - carbon Cl - c?lorine F IG. 7 -STRUCTURAL FIG. F ‘c +---ICL CL . F F F FORMULAS FOR REFRIGERANTS in one vessel and lubricating oil in another (Fig. Sn), and if both vessels are placed in a common ambient temperature, all of the refrigerant migrates to the vessel containing oil because of the absorption head of the oil. Raising the oil temperature limits this migra. tion. For instance, if the oil is at an ambient temperature 20 degrees higher than the refrigerant or if the oil is heated by an immersion type heater to 20 degrees higher than the refrigerant temperature (Fig. Sb), only 6’iy0 of the oil weight of the refrigerant becomes dissolved in the oil. 3. Theoretical Horsepower Per 7’o~l of refrigeration for most refrigerants at air conditioning temperature (Table 1). levels is R E F R I G E R A N T 12 OIL b approximately the same 4. Rate of Lectkngc of a refrigerant gas increases directly with pressure and inversely with molecular weight. The prcssurc of a refrigerant for a given saturated temperature increases in the following order: Refrigerant 113, 11, 114, 12, 500, 22. The molecular weight of a . 8 -MISCIBILITY OF REFRIGERANT 12 AND OIL refrigerant decreases in the following order: Refrigerant 113, 114, 11, 12, 500, 22. Molecular weight is directly related to vapor specific volume; the higher the molecular weight, the higher the specific volume. 5. Leak Detection of the refrigerant should be simple and positive for purposes 01 maintenance, cost and safety. The use of a halide torch makes it possible to detect and locate minute leaks of the halogen refrigerants. 6. ~‘npol- Density influences the compressor dis- placement and pipe sizing. High vapor density accompanied by a reasonably high latent heat of vaporization (low cfm per ton) is desirable in a refrigerant. A low cfm per ton results in compact equipment and smaller refrigerant piping. Reciprocating refrigeration equipment requires a relatively high vapor density refrigerant for optimum performance. Centrilugal compressors require a low vapor density refrigerant for optimum efficiency at comparatively low tonnages. High vapor density refrigerants are used with centrifugals of large tonnage. The cfm per ton increases in the following order: Refrigerant 22, 500, 12, 114, 11, 113 (Fig. 9). Cost which is usually a consideration in all selections should not inlluence the choice of a refrig- erant since it generally has little economic bearing on the normal refrigeration system. Although Refrigerant 22 costs approximately twice as much as Refrigerarlt I?, the compressor required is smaller, tending to offset the additional cost of refrigerant. 39.5 CFM n 1.98 CFM 2.68 CFM 3.14 CFM non 500 I2 114 II NOTE: Evaporator temp, 40 F Condenser tcmp, 105 F combined with the extended surl’ace rc:sults in ;I significant increase in the over;lll heat transrcr rate. When cooling coil is designed to permit 3 normal pressure drop with Refrigerant 22, it may have a rather large pressure drop with Refrigerant I2 or 500. In such a case the decreased performance for Refrigerants 12 and 500 is due partially to the difference in their conclensing film coefficients. In addition, it is affectecl by the pressure drop and the resulting lower mean effective temperature difference. CONDENSERS Frc.9-SUCTION VOLUMES OF REFRIGERANTS (CFM/T~N) l-6. .T TRANSFER COMPARISONS The value of the evaporating and condensing film coefficient (Btu/sq ft/F) for Refrigerant 22 is greater than that for Refrigerants 12 and 500. However, it does not follow that cooling coils and condensers can, therefore, be rated for higher capacities with Refrigerant 22. The higher coefficient for Refrigerant 22 does not tell the complete story. Various other factors should be considered: 1. Whether the heat exchanger is designed for either Refrigerant 22 or Refrigerants 12 and 500. 2. Whether heat transfer is between refrigerant and air or between refrigerant and water. 3. Whether tubes in a heat exchanger are prime surface or extended surface (including the amount of extended surface.) The evaporating or condensing film coefficient is 0 f a number of factors which make up the total ovei-all transfer rate for the heat exchanger. Other factors involved are these: 1. Tube wall resistance (including extended surface, if any). 2. Air or water film coefficient. 3. Refrigerant pressure drop per circuit, which affects the mean effective temperature difference. 4. Surface ratio of tube outside surface to inside surface. 5. Fouling factors (water-cooled condensers). COOLING COILS Cooling coils using Refrigerant 22 normally provide a greater capacity than those using Refrigerant 12 or 500. A cooling coil, normally, has considerable extended surface on the air side of the tubes. The higher evaporating film coefficient of Refrigerant 2:! Water-cooled condensers using Refrigerant 22 normally provide a greater capacity than those using Refrigerant 12 or 500, depending on the fouling factor used in its selection. There are a number of reasons for this improvement other than the basically higher condensing film coefficient of the refrigerant. 1. The water film coefficient is relatively high as compared to air. 2. The extended surface or refrigerant side of the exchanger tubes assures maximum transfer rate and opt&urn balance between the inside and outside surfaces. With Refrigerant 12 or 500 the performance of water-cooled condensers is otherwise not adversely affected since the pressure drop in the shell is not a consideration. Air-cooled condensers using Refrigerant 22 normally provide a greater capacity than those using Refrigerant 12 or 500. Air-cooled condensers have considerable extended surface on the air side of the tubes. The higher evaporating film coefficient of Refrigerant 22, combined with the extended surface results in significant increase in the overall heat transfer rate. When an air-cooled condenser is designed to allow for a normal pressure drop with Refrigerant 22, it may have a considerable pressure drop when. used with Refrigerant 500 or 12. In such a case the decreased performance for Refrigerants 500 and 12 is not due entirely to the difference in their condensing film coefficients, but is affected also by pressure drop and the resulting lower mean effective temperature difference. Normally, evaporative condensers use prime surface tubing (without extended surface on the outside of tubes). Where the unit is designed for Refrigerant 12 or 500, there is no significant increase in capacity when using Refrigerant 22. If the unit is tlcsigried for Kefrigerant 22 (smaller tubing or longer circuits) , and is used with Refrigerant 12 or 500, the pressure drop is sufficient to reduce the rating. SLI& ;I chip m;iy create the impression that the increased rating for Refrigerant 22 is due to the condensing film, whereas actually the performance for Refrigerant 12 or 500 rlccreases due to coil design. REFRIGERANT COMPRESSION ABSORPTION CYCLE Water 2s a refrigerant and lithium bromide as ;in absorbent are utilized in the basic absorption refrigeration cycle. The refrigerant slw~~ltl possess the same tlcsirablc qualities as those for a compression system. In addition, it should be suitable for use with an absorbent, so selected that: SELECTION CYCLE The choice of a refrigerant for a compression system is limited by: 1. Economics, 2. Equipment type and size 3. Application The manufacturer of the refrigeration compressor generally preselects the refrigerant to result in optimum owning cost. The specific refrigerant is determined by the type and size of the equipment. To minimize the number of reciprocating compressor sizes required to fill out a line, the manufacturer rates each size for several refrigerants of relatively dense vapor such as Refrigerants 12, 500 and 22. In effect, this increases the number of units offered without adding sizes. Centrifugal compressors at comparatively low tonnages require a high vapor volume refrigerant such as Refrigerant 113 or 11 to maintain optimum efficiency. For most sizes Refrigerant 114 or 12 can be used to obtain greater capacities. Refrigerants 500 and 22 are used with specially built centriEugals to obtain the highest capacities. The refrigerant selected depends on the type of application. Air-cooled condensers may not use certain refrigerants because of the design condensing temperature required and corresponding limitations on compressor head pressure. The temperature-pressure relationship of a refrigerant is of considerable importance in low temperature applications. If the evaporator pressure is comparatively low for the required evaporator temperature, the volume of vapor to be handled by the compressor is excessive. If the evaporator pressure is comparatively high for the required evaporator temperature, the system pressures are high. The refrigerants that have been mentioned are the halogens (fluorinated hydrocarbon compounds), except for Refrigerant 500 which is an azeotropic mixture of two Auorinated hydrocarbons. The mixture does not separate into its component refrigerants with a change of temperature or pressure. It has its own fixed thermodynamic properties which are unlike either of its components. . 30 IN. 4ES. 3 0 NOTE: -7 Absorption machine at full load, tising lithilim bromide as absorbent. FIG. 10 - P RESSURES AND T EMPERATURES OF A T Y P I C A L ABSORPTION M A C H I N E 1. The vapor pressures of the refrigerant and absorbent at the generator are different. 2. The temperature-pressure relations are consistent with practical absorber and generator temperatures and pressures. Figure 10 shows absolute pressures and temperatures existing in a typical absorption machine at full load. 3. The refrigerant has a high solubility in the absorbent at absorber temperature and pressure and a low solubility at generator temperature and pressure. 4. The refrigerant and absorbent together are stable within the evaporator-generator range of temperatures. Normally, the absorbent must remain liquid at absorber and generator temperatures and pressures. It should have a low specific heat, surface tension and viscosity and must be neutral to the materials used in the equipment. ant1 tlicir c.ll;l~;icteristi(,s. ‘l’ir/~/f,.r I” l o 7 l i s t tllc temperatures. TABLE l-COMPARATIVE DATA OF REFUGERANTS REFRIGERANT NUMBER (AR1 DESIGNATION) 11 Trichloromonofluoromethane Chemical Name Chemical Formula Molecular wt Gas Constant, R !ft-lb/lb-R) Boiling Point at 1 atm (F) Dichlorodifluoromethane CCl3F CC12F2 I3 7 . 3 8 120.93 11.25 388.0 635.0 Specific Heat of liquid, 86 F .220 .235 Specific Heat of Vapor, Cp 6 0 F at 1 atm * .146 -. ific Heat of Vapor, Cv 60 F at 1 atm Triehlorotrifluoroethane CHClFz CC12F-CCIFz 86.48 187.39 .156 .171 .127 * .145 .151 1.12 Ratio of Specific Heats liquid, 105 F 2.04 1.55 2.14 1.47 . 105 F Net Refrigerating Effect (Btu/lb) 4 0 F-105 F fno subcoolina) Cycle Efficiency (yo Carnot Cycle) 40 F-105 F Solubility of Water in Refrigerant Miscibility with Oil Toxic Concentration (yo by vol) Odor Warning Properties Explosive Range (yo by vol) c ‘y G r o u p , U.I.. \I Group, ASA B9.1 To. ‘c Decomposition Products Viscositv (centiooises) Sot&ted liquid 95 F I 2.04 50 F ! liquid Circulated, (Ib/min/ton) 4( 1 F-105 F Theore tical Displacement, 40 F-105 F _,_I . (cu ftlminltont 0.52 2.55 7.03 25.7 7.12 23.05 51.67 141.25 11.74 38.79 03.72 227.65 0.84 2.66 11.58 67.56 49.13 66.44 54.54 90.5 03.2 81.8 87.5 Negligible 1.59 1.77 Negligible 1.65 1 2.10 * Miscible Above 1 0 % Negligible Miscible Above 2070 Negligible Ethereal, odorless w h e n m i x e dw i t ha i r Same asR 11 Same or R 11 SameasRll Same as R 11 SameasRll None NOW 5 Non-2 NOW NOW None 1 6 1 None NOW NOW NOW 6 NOW3 NOW 5A Yes Ye* Ye* Yes Miscible * Limited * 5A 1 4-5 1 I .0121 .0105 I , 1 Ye* 1 Ye* .3420 .3272 .0108 .0109 .OlOO .Olll .2150 .2100 * * * .0124 I * * * * * I .0046 .0051 .0063 .0040 .0435 . 0 4 21 .0056 .0057 .0059 2.96 4.07 3.02 3.66 4.62 3.35 3.14 1.98 9.16 2.69 0.736 0.75 0.70 0.722 0.747 6.39 6.29 6.74 6.52 16.1 Theoretical Horsepower Per Ton 40 F-105 F 0.676 Coefficient of Performance 40 F-105 F (4.7l/hp ton) per 6.95 Cost Compared With R 11 1.00 *Datanotavailableornotapplicable. 1.13 + 1.51 Vapor at 1 atm I 1.09 I 1.84 I 50 F T h e r m a ,C o n d . . c l : ~ : . ~I L , -254 .218 1.18 OF 40 F 170.93 9.04 30.4 -137 * 1.14 1.61 73.a% CClzF2 26.27~ CHKHFz 99.29 15.57 -28.0 .149 1.11 Saturation Pressure (psia) at: - 5 0 F Aseotrope of Dichlorodifluoromethane and Difluoroethane cZc12F4 117.6 -31 417.4 495.0 Ratio 4 =K (86 F at 1 atm) Vapor, Cp, 40 F sot. press. liquid Head (ft), 1 psi at 105 F 50/o 0 ltchlorotetraf luoroethane 8.25 -41.4 -256 204.8 716.0 .335 .130 * 114 113 17.87 -21.62 -252 233.6 597.0 -I68 22 Monochlorodifluoromethone 12.70 74.7 Freerina Point at 1 otm (F) Critical Temperature 6) C r i t i c a l PreSSWe bid / 12 . 39.5 6.3 1 I 1.57 2.77 2.15 2.97 2.00 4-10 - - - - - :1/v\ v ]'\I< I‘ 1. _-. - -~. ..-- --__- J~J~:J~1~1O1~lI.\NJ'.S, . JIl~INI:.S. OILS TABLE 2-PROPERTIES OF REFRIGERANT 12, LIQUID AND SATURATED VAPOR TEMP (F) PRESSURE W/r bmtute P -100 -9 - 9 - 9 - 9 - 9 - 8 6 4 2 0 aa - 06 - a 4 - 82 - 80 78 76 74 72 70 68 66 64 62 6 0 58 - 56 5 4 5 2 5 0 - 48 - 46 - 44 - 42 - 40 - 38 - 36 - 34 - 32 - 30 - 28 - 26 - 24 - 2 2 - 2 0 - 1.3 - 16 14 12 - 1 0 1.4280 1.5381 1.6551 1.7794 I.9112 2.0509 2.1988 2.3554 2.5210 2.6960 2.8807 3.0756 3 . 2 8 1I 3.4975 3.7254 3.9651 4 . 2 172 4.4819 4.7599 5.0516 5.3575 5.6780 6.0137 6.3650 6.7326 7.1168 7 . 5 1 a3 7.9375 a.3751 a.8316 9.3076 9.8035 10.320 10.858 11.417 11.999 12.604 13.233 13.886 14.564 15.267 15.996 16.753 17.536 18.348 1 9 . 1 a9 20.059 20.960 21.891 22.854 23.849 24.878 25.939 27.036 28.167 29.335 30.539 31.780 33.060 34.378 35.736 37.135 38.574 40.056 41.580 n.) Gage P 7.0138* '6.7896' '6.5514* #6.2984' '6.0301' 15.7456' !5.4443* !5.1255* !4.7884* !4.4321* !4.0560* !3.6592+ !3.2409* z2.8002* t2.3362' !1.8482* 11.3350* 20.7959* 20.2299' 19.6360* 19.0133* 18.3607' 17.67731 16.9619* 16.2136* 15.4313* 14.6139* 13.7603* 12.8693' 11.9399' 10.9709* 9 . 9 6 1 1' 8.909* 7.814* 6.675' 5.4901 4.259* 2.979' 1.649' 0.270+ 0.57I 1.300 2.057 2.840 3.652 4.493 5.363 6.264 7.195 8.158 9.153 lo.182 11.243 12.340 13.471 14.639 15.843 17.084 18.364 19.682 21.040 22.439 23.878 25.360 26.884 rc VOLUME (cu I Liquid vt 3) lC2p.X l- vg 0 . 0 0 9 9 8 5 2;!.I 64 . 0 1 0 0 0 2 2(I.682 . 0 1 0 0 2 0 1s I.316 . 0 1 0 0 3 7 lf I.057 . o1 0 0 5 5 Id i.895 0 ~.010073 1: i.821 .010091 111.828 .010109 1.1.908 .oio128 1:I.056 .010146 1:2.226 C I.010164 111.533 .010183 I( 1.852 . 0 1 0 2 0 2 l( I.218 .010221 ?.6290 .010240 ?.oao2 C LO10259 3.5687 .01027a 3.0916 .01029a 7.6462 .010317 7.2302 .010337 6.8412 (1.010357 6.4774 .010377 6.1367 .010397 5.8176 .010417 5.5184 .010438 5.2377 ( I.010459 4.9742 .010479 4.7267 .010500 4.4940 .010521 4.2751 .010543 4.0691 ,3.010564 3.8750 .010586 3.6922 .010607 3.5198 .010629 3.3571 .010651 3.2035 0.010674 3.0585 .010696 2.9214 .010719 2.7917 .010741 2.6691 .010764 2.5529 0.010788 2.4429 .oioali 2.3387 .010834 2.2399 .010858 2.1461 .010882 2.0572 0.010906 1.9727 .010931 1.8924 .010955 1.8161 .oio9ac 1.7436 .011005 1.6745 0.011030 1.6089 .011056 1.5463 .oiloa2 1.4867 .011107 1.4299 .011134 1.3758 0.011 I b C 1.3241 .011187 1.2748 .011214 I.2278 .Oll241 I.1828 .0112bE 1.1399 0.0 1 I 29C I.0988 .01132A 1.0596 .01135; 1.0220 .01138( 0.9861 .01140( 0.9517 'Inches of mercury below one atmosphere. DENSITY fIb/cu ft) T Liquid l/Vf Vapor 119 LO45119 .048352 .051769 .055379 .059ia9 I.063207 .067441 .071900 .076591 .081525 93.690 93.493 93.296 93.098 92.899 92.699 92.499 92.298 92.096 91.893 91.689 91.485 91.280 91.074 90.867 90.659 90.450 90.240 90.030 89.818 89.606 89.392 89.178 0.32696 .34231 .35820 .37466 .39171 0.40934 .4275a .44645 .46595 .48611 0.50693 .52843 .55063 .57354 .59718 0.62156 .64670 .67263 .69934 .72687 0.75523 .78443 .81449 .84544 .87729 88.746 88.529 88.310 88.091 87.870 87.649 Liquid hf - 12.466 - 12.055 -1 1 . 6 4 4 -11.233 -10.821 - 10.409 - 9.9971 - 9.5845 - 9.1717 - a.7586 - a.3451 ).086708 .092151 - 7.9314 - 7.5173 .097863 - 7.1029 .I0385 .I1013 - 6.6881 ).11670 - 6.2730 - 5.8574 .12359 .I3078 - 5.4416 .13831 - 5.0254 - 4.6088 .14617 - 4.1919 3.15438 - 3.7745 .16295 .17189 - 3.3567 - 2.9386 .I8121 - 2.5200 .19092 - 2.1011 0.20104 .21157 - 1.6817 .22252 - 1.2619 - 0.8417 .23391 - 0.4211 .24576 0.0000 0.25806 .27084 0.4215 .28411 0.8434 1.2659 .29788 .312lb I .baa7 00.15 99.978 99.803 99.627 99.451 99.274 99.097 98.919 98.740 98.561 98.382 98.201 98.021 97.839 97.657 97.475 97.292 97.108 96.924 96.739 96.553 96.367 96.180 95.993 95.804 95.616 95.426 95.236 95.045 94.854 94.661 94.469 94.275 94.081 93.886 88.962 t 0.91006 .94377 .97843 1.0141 1.0507 T- ENTHALPY (B stu/lb) I 2.1120 2.5358 2.9601 3.3848 3.8100 4.2357 4.661a 5.0885 5.5157 5.9434 6.3716 6.8003 7.2296 7.6594 8.0898 8.5207 1 8.9521 9.3843 T-i atent ‘+a 7 ‘a.71 4 7 '8.524 7 '8.334 7 '8.144 7 '7.954 7 '7.764 7'7.574 7 '7.384 7 '7.194 7 '7.003 7 '6.812 7'6.620 7'6.429 i '6.238 7'6.046 7 '5.853 7'5.660 i' 5 . 4 6 7 ;' 5 . 2 7 3 , '5.080 i'4.885 j7 4 . 6 9 1 74.495 74.299 74.103 73.906 73.709 73.511 73.312 7 3 . 1 12 72.913 72.712 72.511 72.309 72.106 71.903 71.698 71.494 71.288 vapor h, 66.248 66.469 66.690 66.911 67.133 67.355 67.577 67.799 68.022 68.244 68.467 68.689 68.912 69.135 69.358 69.580 69.803 70.025 70.248 70.471 70.693 70.916 71.138 71.360 71.583 70.874 70.666 70.456 70.246 70.036 69.824 69.611 69.397 69.183 68.967 68.750 68.533 68.314 68.094 67.873 71.805 72.027 72.249 72.470 72.691 72.913 73.134 73.354 73.575 73.795 74.015 74.234 74.454 74.673 74.891 75.110 75.328 75.545 75.762 75.979 76.196 76.411 76.627 76.842 77.057 77.271 77.485 77.698 77.911 78.123 67.651 67.428 67.203 66.977 66.750 66.522 66.293 66.061 65.829 65.596 78.335 78.546 78.757 78.966 79.176 79.385 79.593 79.800 80.007 80.214 71.081 ENTROPY (Btu/l R) Liquid If -0.032005 - .030866 - .029733 - .028606 - .027484 -0.026367 - .025256 - .024150 - .023049 - .021953 Vapor Ig 1.1 a683 .18623 .18565 .18508 ‘18452 I.18398 .I8345 .I8293 .18242 .18192 1.18143 -0.020862 - .019776 .18096 - .018695 .l8050 - .017619 .I8004 - .016547 .I7960 -0.015481 I.17916 - .014418 .17874 - .013361 .17833 - .oi23oa .17792 - .011259 .17753 -0.010214 3.17714 - .009174 .I7676 - .008139 .I7639 - .007107 .17603 .17568 - .006080 -0.005056 0.17533 .17500 - .004037 - .003022 .17467 .17435 - .002011 - .001003 .17403 0.17323 0.000000 .001000 .17343 .001995 .17313 .I7285 .002988 .003976 .I7257 0.004961 0.17229 .005942 .17203 .006919 .17177 .007894 .17151 .ooaa64 .17126 0.009831 0.17102 .010795 .17078 .17055 .011755 .012712 .17032 .013666 .I7010 0.014617 0.16989 .015564 .16967 .016508 .16947 .017445 .16927 .01838E .16907 0.019323 0.16888 .0202x .16869 .021184 *lb851 .02211c -16833 .02303: .I6815 0 . 0 2 3 9 5 4 0.16798 .024871 .I6782 .I6765 .02578( .02669( .16750 .0276Ot .16734 0.02851: 0.16719 .I6704 .02942C .03032: .I6690 .I6676 .031221 .0321lt .I6662 .EMP (F) t --1oo - 96 96 94 92 90 88 86 04 82 80 78 76 74 72 70 66 66 , 64 62 60 5 8 56 54 5 2 50 48 46 44 42 - 40 38 36 34 32 - 30 2 8 26 24 22 20 1 8 16 - 14 - 1 2 - 10 8 - 6 4 2 0 2 4 6 a 10 12 14 lb 18 20 22 24 26 20 (iI I.\ I’ I I,.I< I. 1<1,:L’i< 4-11 (;I-I<.\x I’S TABLE Z-PROPERTIES OF REFRlGERANT TEMP (F) PRESSURE in.) (lb/ ibrolure P I r Gage P 28.452 30.064 31.721 33.424 35.174 VOLUME (cu f 3) -I- 12, LIQUID v( o.oii43a .0114ba .011497 .011527 .011557 vg 3 . 9 1a 8 0 .a8725 .a5702 .a2803 .a0023 DENSITY ( l b / c' U f f) Liquid Vapor 1 lvt 1 i-d 37.426 1.0884 37.202 1.1271 36.977 1.1668 36.751 1.2077 16.524 1.2496 Liquid vapor -r AND SATURATED VAPOR (Contd) I-- ENTHALPY Btu/lbJ ENTROPY Ib- RI (BfU/ ‘EMP T-7 (F) Liquid hr 15.058 15.500 15.942 lb.384 lb.828 hfs 65.361 65.124 64.886 64.647 64.406 Vapor he 80.419 80.624 80.828 81.031 81.234 Liquid ‘if 0.033013 .033905 .034796 .035683 .036569 Pg 0.16648 .16635 .16622 .lbblO .16598 Latent Vapor I 30 32 34 36 38 43.148 44.760 46.4 17 48.120 49.870 40 42 44 46 48 51.667 53.513 55.407 57.352 59.347 36.971 38.817 40.711 42.656 44.65 1 0.011588 .011619 .011650 .oi 1682 .011714 3.77357 .7479a .72341 .69982 .67715 36.296 86.066 35.836 35.604 35.371 1.2927 1.3369 1.3823 1.4289 1.4768 17.273 17.718 18.164 18.611 19.059 64.163 63.919 63.673 63.426 63.177 81.436 81.637 81.837 82.037 82.236 0.037453 .038334 .039213 .040091 .040966 0.16586 .I6574 .16562 .16551 .16540 40 42 44 46 48 50 52 54 56 58 61.394 63.494 65.646 67.853 70.115 46.698 48.798 50.950 53.157 55.419 0.011746 .011779 .oiiali .oiia45 .oiia79 0.65537 .63444 .61431 .59495 .57632 35.136 34.900 34.663 34.425 34.1 a5 1.5258 1.5762 1.6278 I .6808 1.7352 19.507 19.957 20.408 20.859 21.312 62.926 62.673 62.418 62.162 61.903 82.433 82.630 82.826 83.021 83.215 0.041839 .042711 .0435ai .044449 .045316 0.16530 .16519 .I6509 .I6499 .16489 50 52 54 56 58 60 62 64 66 72.433 74.807 77.239 79.729 82.279 57.737 60.111 62.543 65.033 67.583 0.011913 .011947 0.11982 .012017 .012053 0.55839 .54112 .52450 .5084a .49305 83.944 83.701 83.457 83.212 82.965 1.7909 1.8480 1.9066 1.9666 2.0282 21.766 22.221 22.676 23.133 23.591 61.643 61.380 61.116 60.849 60.580 83.409 83.601 83.792 83.982 84.171 0.046180 .047044 .047905 .048765 .049624 0.16479 .I6470 .16460 .16451 .I6442 60 62 64 66 68 .O 72 74 76 78 84.888 87.559 90.292 93.087 95.946 70.192 72.863 75.596 78.391 81.250 0.012089 .012116 .012163 .012201 .012239 0.4781 a .46383 .45000 .43bbb .4237a 82.717 82.467 82.215 81.962 81.707 2.0913 2.1559 2.2222 2.2901 2.3597 24.050 24.511 24.973 25.435 25.899 60.309 60.035 59.759 59.481 59.20 1 84.359 84.546 84.732 84.916 85.100 0.050482 .051338 .052193 .053047 .053900 0.16434 .I6425 .16417 .16408 .16400 70 72 74 76 78 80 a2 a4 a6 aa 98.870 101.86 104.92 108.04 111.23 84.174 87.16 90.22 93.34 96.53 0.012277 .012316 .012356 0.12396 0.12437 0 . 4 11 3 5 .39935 .3a776 .37657 .36575 81.450 al.192 80.932 80.671 80.407 2.4310 2.504 I 2.5789 2.6556 2.7341 26.365 26.832 27.300 27.769 28.241 58.917 58.631 58.343 58.052 57.757 85.282 85.463 85.643 85.821 85.998 0.054751 .055602 .056452 .057301 . 0 5 81 4 9 0.16392 .I6384 .lb376 .I6368 .16360 a0 a2 a4 86 aa 90 92 94 96 98 114.49 117.82 121.22 124.70 128.24 99.79 103.12 106.52 110.00 113.54 0.012178 .012520 .Ol2562 .012605 .012649 0.35529 .345ia .33540 .32594 . 3l b 7 9 80.142 79.874 79.605 79.334 79.061 2.8146 2.8970 2.9815 3.0680 3.1566 28.713 29.187 29,663 30.140 30.619 57.461 57.161 56.858 56.551 56.242 86.174 86.348 86.521 86.691 86.861 0.058997 .059844 .060690 .061536 .062381 0.16353 .16345 .I6338 .16330 .I6323 90 92 94 96 98 1oc 102 104 1OC 106 131.86 135.56 139.33 143.18 147.11 117.16 120.86 124.63 128.48 132.41 0.012693 .012738 .012783 .012829 .012876 0.30794 .29937 .29106 .28303 .27524 78.785 78.508 78.228 77.946 77.662 3.2474 3.3404 3.4357 3.5333 3.6332 31.100 31.583 32.067 32.553 33.041 55.929 55.613 55.293 54.970 54.643 87.029 87.196 87.360 87.523 87.684 0.063227 .064072 .064916 .065761 .066606 0.16315 .16308 .16301 .16293 .I6286 100 102 104 106 108 11c 111 114 11c 118 151.11 155.19 159.36 163.61 167.94 136.41 140.49 144.66 148.91 153.24 0.012924 .012972 .013022 . O 13072 .013123 0.26769 .26037 .2532a .24641 .23974 77.376 77.087 76.795 76.501 76.205 3.7357 3.8406 3.9482 4.0584 4.1713 33.531 34.023 34.5 17 35.014 35.512 54.313 53.978 53.639 53.296 52.949 87.844 88.001 88.156 88.310 88.461 0.067451 .068296 .069141 .069987 .070833 0.16279 .16271 .16264 .I6256 .16249 110 112 114 116 118 12c 4 126 128 172.35 176.85 181.43 186.10 190.86 157.65 162.15 166.73 171.40 176.16 0.013174 .013227 .oi32ao .013335 .013390 0.23326 .22698 .22089 .21497 .20922 75.906 75.604 75.299 74,991 74.680 4.2870 4.4056 4.5272 4.6518 4.7796 36.013 36.516 37.02 1 37.529 38.040 52.597 52.241 51.881 51.515 51.144 88.610 88.757 88.902 89.044 89.184 0.071680 .072528 .073376 .074225 .075075 0.16241 .16234 .I6226 .I6218 .1621C 120 122 124 126 128 13c 13: 134 13t 13s 195.71 200.64 205.67 2 10.79 216.01 181.01 185.94 190.97 196.09 201.31 0.013447 .013504 .013563 .013623 .013684 0.20364 .I9821 .19294 .I8782 .I8283 74.367 74.050 73.729 73.406 73.079 4.9107 5.0451 5.1829 5.3244 5.4695 38.553 39.069 39.588 40.110 40.634 50.768 50.387 50.000 49.608 49.210 89.321 89.456 89.588 89.718 89.844 0.075927 .076779 .077623 .078489 .079346 0.16202 .16194 .i618 .I6177 .1616E 130 132 134 136 138 14t 14: 141 141 14z 221.32 226.72 232.22 237.82 243.51 206.62 212.02 217.52 223.12 228.81 0.013746 .oi3aio .oi3874 .013941 .014008 0.17799 .17327 .I6868 .16422 .I5987 72.748 72.413 72.075 71.732 71.386 5.6184 5.7713 5.9283 6.0895 6.2551 41.162 41.693 42.227 42.765 43.306 48.805 48.394 47.977 47.553 47.122 89.967 90.087 90.204 90.31 a 90.428 0.080205 .081065 .oai92a .082794 .083661 0.16155 .1615C .1614C .1613C .1612C 140 142 144 146 148 1% 15: 15‘ 15( 151 249.3 1 255.20 261.20 267.30 273.5 1 234.61 240.50 246.50 252.60 258.81 0.014078 .014148 .014221 .014295 .014371 0.15564 .15151 .I4750 .I4358 .I3976 71.035 70.679 70.319 69.954 69.584 6.4252 6 . 6 0 01 6.7799 6.9648 7.1551 43.850 44.399 44.951 45.508 46.068 46.684 46.238 45.784 45.322 44.852 90.534 90.637 90.735 90.830 90.920 0.084531 .085404 .086280 .087159 .088041 0.16llC .1609( .1608t .1607i b o .bl! 150 152 154 156 15a 16( 279.82 265.12 0.014449 0.13604 69.209 7.3509 46.633 44.373 9 1.006 0.088927 0.1605: 160 - - 30 32 34 36 38 TABLE TEMf (F) t -40 - 3 8 - 3 6 - 3 4 - 3 2 - 3 0 - 2 8 - 2 6 - 2 4 - 2 2 - 2 0 - 1 8 -16 - 1 4 - 12 - 1 0 8 6 - 4 2 0 2 4 6 8 10 12 14 lb 18 20 2 2 24 26 28 30 32 34 36 38 40 42 44 46 \48 50 52 54 56 58 60 61 64 66 68 7a 71 74 76 78 BC 81 84 8t a8 S-PROPERTIES OF REFRIGERANT 500, LIQUID AND SATURATED VAPOR PRESSURE (Ib/rq in.) VOL iGdE (cv 1 k/l b) .iquid Vf 15.45 16.21 17.01 17.84 18.70 19.59 20.5 I 21.47 22.46 23.49 24.55 25.65 26.79 27.96 29.18 30.43 31.73 33.06 34.45 35.88 37.35 38.86 40.42 42.03 43.69 45.40 47.15 48.96 50.82 52.74 54.71 56.72 58.80 0.75 1.51 2.3 1 3.14 4.00 4.89 5.81 6.77 7.76 a.79 9.85 10.95 12.09 13.26 14.48 15.73 17.03 18.36 19.75 21.18 22.65 24.16 25.72 27.33 28.99 30.70 32.45 34.26 36.12 38.04 40.0 1 42.02 44.10 -/ 67.71 I 53.01 70.09 15 5 . 3 9 72.52 1 57.82 75.02 1 60.32 77.57 62.87 80.22 65.52 I . 0 11 9 .OllP .0119 .0119 .0120 I.01 20 .0120 .0120 .0121 .0121 P-- vapor vg Liquid l/vt 4.0757 3.8829 3.6992 3.5276 3.3671 3.2121 3.0674 2.9302 2.8020 2.6788 84.37 84.19 84.00 83.82 83.63 I.0121 .0121 .0122 .0122 .0122 I.0123 .0123 .0123 .0123 .0124 3.0124 .0124 .0125 .0125 .0125 0.0126 .0126 .0126 .0127 .0127 0.0127 .0128 .0128 .0128 .0129 2.5622 2.4520 2.3477 2.2491 2.1548 2.0657 1.9807 1.9004 1.8238 1.7507 1.6818 1.6155 1.5530 1.4929 1.4362 1.3813 1.3292 1.2796 1.2325 1.1873 1.1440 1.1027 1.0631 1.0254 0.9892 82.51 82.32 82.13 81.94 81.75 0.0129 .0129 .0130 .0130 .0130 0.0131 .0131 .0131 .0132 .0132 0.0133 .0133 .0133 .0134 .0134 0.0135 .0135 .0135 .0136 .0136 0.0137 .0137 .0138 .0138 .0139 0.0139 .0139 .0140 .0140 .0141 0.9545 .9212 .8894 .8591 .a298 0.8017 .7747 .7490 .7241 .7002 0.6774 .6554 .6343 .6138 .5943 0.5756 .5575 s400 .5231 .5069 0.4914 .4763 .4618 .4479 .4344 0.4213 .4087 .3967 .3849 .3735 *Inches of mercury below one atmosphere. 83.45 83.26 83.07 82.89 82.70 81.56 81.37 81.17 80.98 80.78 80.59 80.39 80.20 80.00 79.80 79.60 79.40 79.20 79.00 78.80 78.59 78.39 78.18 77.98 77.77 77.56 77.35 77.14 76.93 76.72 76.50 76.29 76.07 75.86 75.64 75.42 75.20 74.97 74.75 74.52 74.30 74.07 73.04 73.61 73.38 73.14 72.91 72.67 72.43 72.19 71.95 71.70 71.46 71.21 70.96 IY ff) vapor l/Q 0.2454 .2575 .2703 .2835 .2970 0.3113 .3260 .3413 .3569 .3733 0.3903 .4078 .4260 .4446 .4641 0.4841 .5049 .5262 .5483 .5712 0.5946 .6190 .6439 .6698 .6963 0.7239 .7523 .7815 .a114 .8422 0.8741 0.9069 0.9407 0.9752 1.0109 1.0476 1.0855 1.1244 1.1640 1.2051 1.2473 1.2908 1.3352 1.3811 1.4282 1.4763 1.5258 1.5764 1.6291 1.6826 1.7374 1.7939 1.8518 1.9116 1.9727 2.0350 2.0996 2.1655 2.2328 2.3021 2.3735 2.4469 2.5210 2.5981 2.6772 r .RO ENTHALPY Mu/lb) liquid hf 0.00 0 . 51 1.02 1.54 2.05 2.57 3.09 3.60 4.13 4.65 5.17 5.70 6.22 6.75 7.27 7.80 a.33 8.87 9.40 9.94 10.47 11.01 11.55 12.09 12.63 13.17 13.71 14.26 14.81 15.35 15.91 16.45 17.01 17.56 18.12 18.67 19.22 19.79 20.34 20.91 21.47 22.03 22.60 23.16 23.75 24.31 24.88 25.46 26.04 26.59 27.19 27.76 28.33 28.92 29.52 30.10 30.68 31.27 31.87 32.47 33.06 33.63 34.24 34.84 35.44 EN1 PY (mu /lb.-R) Latent Vapor b b 89.91 89.68 89.45 89.21 88.97 88.73 88.49 88.25 88.00 87.75 87.50 87.25 87.00 86.74 84.49 06.23 85.97 85.70 85.44 85.17 84.90 84.63 84.35 84.08 83.80 83.52 83.24 82.95 02.66 82.37 82.07 81.78 81.48 81.18 80.87 80.57 80.26 79.94 79.63 79.31 78.99 78.67 78.34 78.01 77.66 77.33 76.99 76.64 76.29 75.96 75.59 75.24 74.89 74.51 74.13 73.76 73.39 73.01 72.62 72.22 71.83 71.46 71.04 70.63 70.22 89.91 90.19 90.47 90.75 91.02 91.30 91.58 91.85 92.13 92.40 92.67 92.95 93.22 93.49 93.76 Liquid St 0.00000 94.03 94.30 94.57 94.84 95.11 .00122 .00243 .00365 .00485 0.00606 .00725 .00845 .00963 .01083 0.01202 .01322 .01438 .01557 .01674 0.01793 .01909 .02027 .02144 .02259 95.37 95.64 95.90 96.17 96.43 96.69 96.95 97.2 1 97.47 97.72 97.98 98.23 98.49 98.74 98.99 99.24 99.48 99.73 99.97 100.22 100.46 100.70 100.94 101.17 101.41 101.64 101.87 102.10 102.33 102.55 102.78 103.00 103.22 103.43 103.65 103.86 104.07 104.28 104.49 104.69 104.89 105.09 105.28 105.47 105.66 0.02376 .02491 .02608 .02722 .02839 0.02954 .03067 .03182 .03296 .03411 0.03526 .03638 .03752 .03865 .03978 0.04092 .04203 .04316 .04430 .04542 0.04654 .04764 .04875 .04986 .05099 0 . 0 5 21 2 .05321 .05431 .05543 .05649 0.05763 .05872 .0597s .06091 .06203 0 . 0 6 31 1 .06415 .06527 .06637 .06749 0.0685: .0695$ .07071 .0717S .07282 IlEMP (F) vapor Ig 0.21421 .21387 .21352 .21318 .21286 0.21252 .21221 .21189 .21160 .21130 0.21101 .21072 .21044 .21017 .20991 0.20965 .20939 .20914 .20890 .20866 0.20843 .20819 .20797 .20775 .20754 0.20732 .20711 .20691 .20672 .2065i 0.2063: .20614 .2059: .2057E .2056( 0.2054; .2052: .205OC .20491 . 2 0 4 7 !i 0 . 2 0 4 5 t1 . 2 0 4 4 :, . 2 0 4 2 (i . 2 0 4 1 () . 2 0 3 9 !5 0 . 2 0 3 8 (1 . 2 0 3 6 !5 . 2 0 3 5 (1 . 2 0 3 3 :5 . 2 0 3 2 (I 0 . 2 0 3 0 15 .2029I . 2 0 2 7 ;I . 2 0 2 6 :2 . 2 0 2 4 13 0 . 2 0 2 3 ~b . 2 0 2 2 (I . 2 0 2 0 15 . 2 0 1 9 :2 . 2 0 1 7 13 0 . 2 0 1 6 s4 . 2 0 1 4 19 . 2 0 1 3 15 .20121 . 2 0 1 07: t - 4 0 - 3 8 - 3 6 - 3 4 - 3 2 - 3 0 -20 - 2 6 -14 - 2 2 - 2 0 -18 - 1 6 - 1 4 - 1 2 - 1 0 - 8 -6, 4 2 0 2 4 6 II 10 12 14 lb 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 60 70 72 74 76 70 80 82 84 86 88 TABLE 3-PROPERTIES OF REFRIGERANT 500, LIQUID AND SATURATED VAPOR (Contd) TEMP (F) t 90 92 94 96 98 ;su PRES RE ( l b / ,sq i 4 Abbsolute Gage P P 135.86 121.2 139.83 125.1 143.90 129.2 133.3 148.03 152.27 137.6 iTii VOI WE ( C U f t / ll b ) Liquid Vf Vapor r L 0.0141 .0142 .0142 .0143 .0144 vg 0.3626 .3520 .341a .3319 .3223 DENSITY (lb/t ft) Liquid l/vt lI 70.70 70.45 70.19 69.93 69.67 Vapor 11% 2.7581 2.8409 2.9261 3.0128 3.1024 Liquid hr 36.04 36.66 37.26 37.88 38.47 ENTHALPY -(Btu/lb) I 1rrtent Vapor 69.81 69.37 68.96 68.51 68.10 h, 105.85 106.03 106.22 106.39 106.57 b l- ENTROPY IBtu/lb-R) Liquid Vapor % 51 0.07395 0.20092 .07505 .2007a .076!1 .20064 .07722 .20049 .20035 .07826 L I TEMP (F) t 90 92 94 96 98 100 102 104 106 108 156.61 161.02 165.55 170.14 174.84 141.9 146.3 150.9 155.4 160.1 0.0144 .0145 .0145 .0146 .0146 0.3130 .3041 .2953 .2869 .27aa 69.41 69.14 68.87 68.60 68.33 3.1947 3.2889 3.3860 3.4850 3.5871 39.09 39.71 40.33 40.96 41.59 67.65 67.20 66.74 66.27 65.80 106.74 106.91 107.07 107.23 107.39 0.07935 .oao42 .0815-l .08261 .08367 0.20019 .20004 .I9989 .I9974 .l995a 100 102 104 106 108 110 112 114 116 118 179.62 184.51 189.47 194.55 199.71 164.9 169.8 174.8 179.9 185.0 0.0147 .014a .oi4a .0149 .0149 0.2709 .2632 .2558 .2486 .2417 68.05 67.78 67.49 67.21 66.92 3.6914 3.7990 3.9086 4.0218 4.1376 42.22 42.82 43.47 44.11 44.72 65.32 64.87 64.37 63.87 63.40 107.54 107.69 107.84 107.98 108.12 0.08479 .oa482 .08691 .oaaoo .oa905 0.19942 .I9926 .19910 .I9893 .I9876 110 112 114 116 118 120 122 124 .- . 204.99 210.40 215.88 221.44 222.13 190.3 195.7 201.2 206.7 212.4 0.0150 .0151 .0151 .0152 .0153 0.2349 .22a3 .2219 .2157 .2097 66.63 66.34 66.04 65.74 65.43 4.2566 4.3807 4.5068 4.6357 4.7690 45.35 46.02 46.69 47.33 47.98 62.90 62.35 61.81 61.29 60.75 108.25 108.37 108.50 108.62 108.73 0.09012 .09124 .09236 .09342 .09451 0.19859 .l984i .I9823 .I9805 .19786 120 122 124 126 128 -i30132 134 136 138 232.89 238.80 244.84 250.96 257.17 218.2 224.1 230.1 236.3 242.5 0.0154 .0154 .0155 .0156 .0157 0.2039 .I982 .I926 .ia72 .1a20 65.13 64.81 64.49 64.17 63.85 4.9050 5.0463 5.1924 5.3420 5.4951 48.65 49.28 49.91 50.59 51.30 60.19 59.66 59.13 58.54 57.91 108.84 108.94 109.04 109.13 109.21 0.09560 .096bb .09771 .09aa3 .09996 0.19767 .19747 .19726 .19705 .I9684 130 132 134 136 138 140 142 144 146 148 263.52 269.99 276.55 283.24 290.04 248.8 255.3 261.9 268.5 275.3 0.1769 .1719 .1670 .1623 .I577 63.52 63.18 62.84 62.49 62.14 5.6528 5.8170 5.9878 6.1626 6.3431 52.02 52.73 53.45 54.15 54.86 57.27 56.63 55.97 55.33 54.67 109.29 109.36 109.42 109.48 109.53 0.10114 .I0228 .10344 .10456 .I0569 0.19663 .I9639 .19615 .I9591 .I9565 140 142 144 146 148 150 152 154 156 158 296.97 304.01 311.18 318.47 325.87 282.3 289.3 296.5 303.8 311.2 0.0157 .ol58 .0159 .Ol60 .0161 t 0.0162 .0163 .0164 .Ol65 .0166 0.1531 .l4a7 .1444 .1402 .I360 61.78 61.42 61.05 60.66 60.28 6.5304 6.7241 6.9253 7.1342 7.3510 55.62 56.34 57.07 57.82 58.61 53.95 53.25 52.54 51.80 51.01 109.57 109.59 109.61 109.62 109.62 0.10691 .I0806 .10922 .11040 .I1163 0.19539 .19511 .I9483 .19453 .I9421 150 152 154 156 158 160 333.40 318.7 0.0167 0.1320 59.88 7.5772 59.35 50.25 109.60 0.11280 0.19389 160 TABLE 4-PROPERTIES OF REFRIGERANT 22, LIQUID AND SATURATED VAPOR PRESSURE (lb/s< 1.) TEMP (F) brolute P Gage VOL (cu f l- \E b) Liquid vt j.0102 .0103 .OlO3 .OlO3 .0104 vapor “g 88.1 46.1 14.5 90.61 72.33 DENSITY (Ib/cu 0) I -L Liquid Vapor ILiquid 1 /Vf hf l/%3 97.67 1.005316 -29.07 -27.79 97.33 JO6847 96.99 .008733 - 26.52 96.63 .01104 - 25.25 - 23.99 96.27 .Ol383 I ENTHALPY ( h/lb) Latent F ENTROPY CBfu/lb-R) l- EMP (F) T Liquid Sf - 0.0808 - .0767 - .0727 - .0687 - .0647 Vapor I Il.10 110.45 109.80 89.70 90.29 90.88 91.47 92.07 -0.0609 - .0571 - .0534 - .0497 - .0461 0.2803 .2770 .2738 .2708 .2680 -130 -125 -120 -115 -110 92.67 93.27 93.87 94.47 95.08 -0.0425 - .0390 - .0356 - .0322 - .0288 0.2653 .2627 .2602 .2579 .2556 -105 -100 - 95 - 90 - a5 “fg 115.85 115.15 114.46 113.78 113.10 vapor r + ~, h, 86.78 87.36 87.94 88.53 89.11 sg 0.2996 .2952 .2912 .2874 .2837 I -155 - 150 -145 -140 -135 0.19901 0.2605 0.3375 0.4332 0.5511 P 29.51* 29.39* 29.23* 29.04* 28.80* - 0.6949 0.8692 1.079 1.329 1.626 28.51' 28.15* 27.72* 27.21' 26.61' j.0104 .0105 .OlO5 .0106 .0106 58.21 47.23 38.60 31.77 26.33 95.91 95.53 95.15 94.76 94.37 I.01718 .02118 .02591 .03147 .03798 - - 105 -100 - 95 - 90 - 85 1.976 2.386 2.845 3.417 A.055 25.90' 25.06* 24.09+ 22.96* 21.67* I.0106 .0107 .0107 .OlO8 .OlOE 21.96 18.43 15.54 13.20 II.26 93.97 93.56 93.14 92.72 92.29 1.04554 .05427 .06433 .07578 .08884 - lb.48 -15.23 - 13.98 - 11.47 109.15 108.50 107.85 107.20 106.55 - 80 78 76 74 72 4.787 5.100 5.430 5.79 6.17 20.18' 19.55* 18.87* 1a.14* 17.37* ).01090 .01091 .01093 .01095 .01097 9.650 9.086 8.561 8.072 7.616 91.85 91.67 91.49 91.31 91.13 L1036 .I101 .I168 .1239 .I313 -10.22 - 9.72 - 9.21 - 8.70 - 8.20 105.90 105.64 105.37 105.10 104.84 95.68 95.92 96.16 96.40 96.64 -0.0255 - .0242 - .0229 - .0216 - .0203 0.2535 .2526 .2518 .2510 .2502 - - 70 68 66 64 62 6.57 6.99 7.40 7.86 8.35 16.55* 15.70+ 14.86* 13.93* 12.93* 3.01100 .01102 .OllO4 .OllOb .01109 7.192 6.795 6.426 6.079 5.755 90.95 90.77 90.58 90.39 90.21 I.1391 .I472 .I556 .I 6 4 5 .1738 -7.69 - 7.19 - 6.68 - 6.17 - 5.67 104.57 104.31 104.04 103.77 103.51 96.88 97.12 97.36 97.60 97.84 -0.0190 - .0177 - .0164 - .0151 - .Ol38 0.2494 .2487 .2479 .2472 .2465 - 70 - 60 - 66 - 64 - 62 - 60 - 58 - 56 - 54 - 52 8.86 9.39 9.94 10.51 11.11 11.89' 10.81' 9.69' 8.53' 7.31* 0.01111 .01113 .01115 .Oll18 .01120 5.452 5.166 4.900 4.650 4.415 90.03 89.84 89.65 89.46 89.27 I.! 834 .I936 .2041 .2151 .2265 - 5.16 4.65 4.13 3.61 3.09 103.24 102.97 102.69 102.41 102.13 98.08 98.32 98.56 98.80 99.04 -0.0126 - .0113 - .OlOO - .0087 - .0075 0.2458 .2451 .2444 .2438 .243l - 60 - 58 - 56 - 54 - 52 - 5c - 48 - 4c - 44 - 49 11.74 12.40 13.09 13.80 14.54 6.03* 4.68+ 3.28* 1.83' 0.326' 0.01123 .Ol125 .01128 .01130 .01133 4.192 3.986 3.793 3.611 3.440 89.08 88.88 88.68 88.49 88.30 3.2386 .2509 .2636 .2769 .2907 - 2.58 2.06 1.54 1.02 0.51: 101.86 101.58 101.30 101.02 100.74 99.28 99.52 99.76 100.00 100.23 -0.0062 - .0050 - .0037 - .0025 - .OOl2 0.2425 .241& .2412 .2406 .2400 - - 4c 3E 36 34 3: 15.31 16.12 16.97 17.85 18.77 0.610 1.42 2.27 3.15 4.07 0.01135 .01138 .Oll40 .01143 .01146 3.279 3.126 2.981 2.844 2.713 88.10 87.90 87.70 87.50 87.29 0.3050 .3199 .3355 .3517 .3686 0.00 0.53 1.05 1.58 2.10 100.46 100.17 99.88 99.59 99.30 100.46 100.70 100.93 101.17 101.40 0.0000 .0013 .0025 .0037 .0050 0.2394 .2389 .2383 .2377 .2372 - 40 - 3c - 2f - 2c - 24 - 2; 19.72 20.71 21.73 22.79 23.88 5.02 6.01 7.03 8.09 9.18 0.01148 .01151 .Oll54 .01156 .01159 2.590 2,474 2.365 2.262 2.165 87.09 86.89 86.69 86.48 86.27 3.3862 .4043 .4229 .4421 .4619 2.62 3.15 3.69 4.22 4.75 99.01 98.71 98.41 98.11 97.81 101.63 101.86 102.10 102.33 102.56 0.0062 .0074 .0086 .0099 .Olll 0.2367 .236l .2356 .2351 .2346 - 30 28 26 24 22 - 2( - 18 - 1t - 1r - 1: 25.01 26.18 27.39 28.64 29.94 IO.31 11.48 12.69 13.94 15.24 0.01162 .01165 .01168 .01171 .01174 2.074 1.987 1.905 1.827 1.752 0.4822 .5032 .5249 .5474 .5707 5.28 5.82 6.40 6.90 7.43 97.51 97.20 96.89 96.58 96.27 102.79 103.02 103.25 103.48 103.70 0.0123 .0135 .0147 .0159 .0170 0.2341 .2336 .2331 .2326 .2321 - 20 18 16 14 12 - 31.29 32.49 34.14 35.64 37.19 16.59 17.99 19.44 20.94 22.49 0.01 177 .01180 .01183 .01186 .01189 86.06 85.85 85.64 85.43 - 85.21 - 1.681 1.613 1.549 1.488 1.429 84.99 84.78 84.56 84.34 84.12 0.5948 .6198 .6456 .6723 .6997 7.96 8.49 9.02 9.55 10.09 95.96 95.65 95.34 95.03 94.71 103.92 104.14 104.36 104.58 104.80 0.0182 .0194 .0205 .0217 .0228 0.2316 .2312 .2307 .2302 .2298 - 1 0 8 6 4 2 I 38.79 40.43 42.14 43.91 45.74 24.09 25.73 27.44 29.21 31.04 0.01192 .01195 .01198 .01201 .01205 1.373 1.320 1.270 1.221 1.175 83.90 83.68 83.45 83.23 83.01 0.7282 .7574 .7877 .a191 .a514 10.63 11.17 li.70 12.23 12.76 94.39 94.07 93.75 93.43 93.11 105.02 105.24 105.45 105.66 105.87 0.0240 .0251 .0262 .0274 .0285 0.2293 .2289 .2285 .2280 .2276 0 2 4 6 a l( 1: 11 l( II 47.63 49.58 51.59 53.66 55.79 32.93 34.88 36.89 38.96 41.09 0.01208 .Ol211 .01215 .01218 .01222 1.130 1.088 1.048 1.009 0.972 82.78 82.55 82.32 82.09 81.86 0.8847 0.9191 0.9545 0.9911 1.029 13.29 13.82 14.36 14.90 15.44 92.79 92.47 92.14 91.81 91.48 106.08 106.29 106.50 106.71 106.92 0.0296 .0307 .0319 .0330 .034l 0.2272 .2268 .2264 .2260 .2257 10 12 14 16 18 21 2: 2' 2( 21 57.98 60.23 62.55 64.94 67.40 43.28 45.53 47.85 50.24 52.70 0.01225 .01229 .01232 .Ol236 .01239 0.936' 0.903' 0.870: 0.839; 0.8101 81.63 81.39 81.16 80.92 80.69 1.067 1.107 1.149 I.191 1.235 15.98 16.52 17.06 17.61 18.17 91.15 90.81 90.47 90.12 89.76 107.13 107.33 107.53 107.73 107.93 0.0352 .0364 .0375 .0379 .0398 0.2253 .2249 .2246 .2242 .2239 20 22 24 26 28 1 1 1 1 1 3 2 2 1 1 0 5 0 5 0 I( I t 1 : ( 1 t *Inches of mercury below one atmosphere. 22.73 2 I .47 20.22 18.98 - 17.73 - 12.73 112.43 i111.76 -155 -150 -145 -140 -135 80 78 76 7 4 72 50 48 46 44 42 - 3 8 - 36 - 34 - 32 . TABLE 4-PROPERTIES OF REFRIGERANT 22, LIQUID AND SATURATED VAPOR (Contd) TEMP (F) t 30 PRESSURE (lb/! in.) ibrolute P Gage P 55.23 57.83 60.51 63.27 66.11 VOI (CU liquid vt I.01243 .01247 .01250 .01254 .oi25a HE lb) vapor "9 0.7816 .7543 .7283 .7032 .6791 Liquid 1 /Vf 80.45 80.21 79.97 79.73 79.49 TY ft) vapor ENTHALPY (Btu/lb) Liquid hf 18.74 19.32 19.90 20.49 21.09 Latent Vapor 119 I.280 1.326 1.373 1.422 1.473 hf, 89.39 89.01 88.62 88.22 87.81 h, 108.13 108.33 108.52 108.71 108.90 ENTl PY (Btu/ '-l b -,R) Liquid Vapor Of Ig 0.0409 0.2235 .0421 .2232 .2228 .0433 .0445 .2225 .0457 .2222 lEMP (F) t 30 32 34 36 38 32 34 36 38 69.93 72.53 75.21 77.97 80.81 40 42 44 46 48 83.72 86.69 89.74 92.88 96.10 69.02 71.99 75.04 78.18 al.40 I.01262 .01266 .01270 .01274 .01278 0.6559 .6339 .6126 .5922 .5726 79.25 79.00 78.76 78.51 78.26 1.525 1.578 1.632 1.689 1.747 21.70 22.29 22.90 23.50 24.1 1 87.39 86.98 86.55 86.13 85.69 109.09 109.27 109.45 109.63 109.80 0.0469 .04ai .0493 .0505 .0516 0.2218 .2215 .2211 .2208 .2205 40 42 44 46 48 50 52 54 56 58 99.40 102.8 106.2 109.8 113.5 84.70 ea.10 91.5 95.1 98.8 ).oi282 .01286 .01290 .01294 .01299 0.5537 .5355 .51a4 ,501 A .A849 78.02 77.77 77.51 77.26 77.01 1.806 1.868 1.929 1.995 2.062 24.73 25.34 25.95 26.58 27.22 85.25 84.80 EA.35 83.89 83.41 109.98 110.14 110.30 110.47 110.63 0.0528 .0540 .0552 .056A .0576 0.2201 .2198 .2194 .2191 .2iaa 50 52 54 56 58 60 62 4 ,6 68 117.2 121.0 124.9 128.9 133;o 102.5 106.3 110.2 114.2 118.3 I.01303 .01307 .01312 .01316 .01320 0.4695 .4546 .4403 .4264 .4129 76.75 76.50 76.24 75.98 75.72 2.130 2.200 2.271 2.346 2.422 27.83 28.46 29.09 29.72 30.35 82.95 82.47 81.99 81.50 81.00 110.78 110.93 1 i 1.08 111.22 111.35 0.0588 .0600 .0612 .0624 .0636 0.2185 .2181 .2178 .2175 .2172 60 62 64 66 68 70 71 74 76 70 137.2 141.5 145.9 150.4 155.0 122.5 126.8 131.2 135.7 140.3 1.01325 .01330 .01334 .01339 .01344 0.4000 .3875 .3754 .3638 .3526 75.46 75.20 74.94 74.68 74.41 2.500 2.581 2.664 2.749 2.836 30.99 31.65 32.29 32.94 33.61 80.50 79.98 79.46 78.94 78.40 111.49 111.63 111.75 1 i 1.88 112.01 0.0648 .066 1 .0673 .0684 .0696 0.2168 .2165 .2162 .2l5a .2155 70 72 74 76 78 aa 159.7 164.5 169.4 174.5 179.6 3.01349 .01353 .01358 .01363 .oi3ba 0.3417 .3313 .3212 .3113 .3019 74.15 73.89 73.63 73.36 73.09 2.926 3.019 3.113 3.213 3.313 34.27 34.92 35.60 36.28 36.94 77.86 77.32 76.76 76.19 75.63 0.0708 .0720 .0732 .0744 .0756 0.2151 .214a .2144 .2140 .2137 80 a2 a4 86 aa 9c 94 94 9c 9E 184.8 190.1 195.6 201.2 206.8 145.0 149.8 154.7 159.8 164.9 170.1 175.4 180.9 186.5 192.1 -i12.13 a2 84 86 af 0.2928 .2841 .2755 .2672 .2594 72.81 72.53 72.24 71.95 71.65 3.415 3.520 3.630 3.742 3.855 37.61 38.28 38.97 39.65 40.32 75.06 74.48 73.88 73.28 72.69 -i12.67 12.76 12.85 12.93 I 13.00 0.0768 .0780 .0792 .0803 .0815 0.2133 .2130 .2126 .2122 .2119 90 92 94 96 98 1oc 101 104 lot 1O f 212.6 218.5 224.6 230.7 237.0 197.9 203.8 209.9 216.0 222.3 0.01374 .01379 .oi3a4 .01390 .01396 -. 0.01402 .oi4oa .01414 .01420 .01426 0.2517 .2443 .2370 .2301 .2233 71.35 71.05 70.74 70.42 70.11 3.973 4.094 4.220 4.347 4.479 40.98 Al.65 42.32 42.98 43.66 72.08 71.47 70.84 70.22 69.58 -i13.06 113.12 113.16 li3.20 113.24 0.0827 .oa39 .0851 .0862 .oa74 0.2115 .2111 .2107 .2lOA .2100 100 102 104 106 108 11c 11; 114 1lC ‘If 243.4 249.9 256.6 263.4 270.3 228.7 235.2 241.9 248.7 255.6 0.01433 .01440 .01447 .01454 .01461 0.2167 .2104 .2043 .I983 .I926 69.78 69.45 69.12 68.78 68.44 4.614 4.752 4.896 5.043 5.192 44.35 45.04 45.74 46.44 47.14 68.94 68.30 67.64 66.98 66.32 113.29 113.34 I 13.38 113.42 113.46 0.0886 .089a .0909 .092l .0933 0.2096 .2093 .2089 .2085 .2081 110 112 114 116 118 . ..LC 12; 124 12c 12f 277.3 284.4 291.7 299.1 306.5 262.6 269.7 277.0 284.4 291.8 0.01469 .01475 .014a3 .01491 .01498 0.1871 .ia25 .I772 .I724 .1675 68.10 67.75 67.40 67.05 66.70 5.345 5.48 5.64 5.80 5.97 47.85 48.60 49.40 50.20 50.80 65.67 64.97 64.21 63.45 62.89 113.52 113.57 113.61 113.65 113.69 0.0945 .0959 .0973 .0986 .0997 0.2078 .2076 .2073 .2070 .2067 120 122 124 126 128 1x 13: 134 13t 13f 314.0 321.8 329.9 338.3 347.0 299.3 307.1 315.2 323.6 332.3 0.01507 .01515 .01523 .01532 .01541 0.1629 .1585 .I538 .1492 .I449 66.35 66.00 65.65 65.25 64.85 6.14 6.31 6.50 6.70 6.90 51.50 52.30 53.10 53.80 54.60 62.21 61.44 60.67 59.99 59.20 113.71 113.74 113.77 113.79 i 13.80 0.1009 .1022 .I035 .1046 .1059 0.2064 .2061 .2057 .2053 .2050 130 132 134 136 138 14c 14: 14r 14f 14f 356.0 365.0 374.1 383.3 392.6 341.3 350.3 359.4 368.6 377.9 0.01551 .01561 .01571 .01581 .01591 0.1408 .I368 .1330 .I292 .1253 64.45 64.05 63.65 63.25 62.85 7.10 7.31 7.52 7.74 7.98 55.30 56.10 56.90 57.70 58.40 58.51 57.70 56.89 56.08 55.36 113.81 113.80 113.79 113.78 113.76 0.1070 .1084 .I096 .I110 .1120 0.2046 .2043 .2039 .2036 .2031 140 142 144 146 148 15( 15: 154 1% 151 401.9 411.3 420.8 430.3 439.8 387.2 396.6 406.1 415.6 425.1 0.01601 .01612 .01625 .01636 .01648 0.1216 .1179 .114l .I105 .1070 62.45 62.02 61.58 61.13 60.67 8.22 8.48 a.76 9.05 9.35 59.20 60.00 60.80 61.60 62.50 54.54 53.71 52.87 52.02 51.06 113.74 113.71 113.67 113.62 113.56 0.1132 .1145 .I156 .1168 .1181 0.2027 .2023 .2ola .2013 .2008 150 152 154 156 158 lb( 449.3 434.6 0.01661 0.1035 60.20 9.66 63.50 50.00 113.50 0.1196 0.2003 160 12.24 12.36 12.47 12.57 TABLE 5-PROPERTIES TEMP WI T T- PRESSURE (lb/w 4 i t1.) ,Gage A bsolufe P P OF REFRIGERANT 1 1, LIQUID AND SATURATED VAPOR DENSITY ft) VOLUME (cu b) (lb/ I .iquid l/Yf 1 01.25 1 01.10 1 00.96 1 00.8 I 1 00.66 t ENTHALPY Btu/ib) i* Liquid hf 0.00 0.40 0.79 1.19 1.59 Latent Vapor l/Q 1.02260 .02409 .02566 .02732 .02906 hfe 87.53 87.37 87.22 87.06 86.91 h, 87.53 87.77 88.01 88.25 88.49 liquid St 0.0000 .0009 .0019 .0028 .0038 SB 0.2086 .2081 .2077 .2073 .2070 00.52 00.37 00.22 00.07 99.92 I.03090 .03282 .03485 .03697 .03920 1.99 2.38 2.78 3.18 3.58 86.75 86.60 86.44 86.29 86.13 Vapor -40 -38 -36 -34 -32 88.74 88.98 89.22 89.47 89.71 0.0047 .0056 .0065 .0074 .0084 0.2066 .2062 .2058 .2055 .2051 -30 -28 -26 -24 -22 0.7387 0.7911 0.8466 0.9053 0.9674 2!8.42* 2!8.31f 2!8.20* 2!8.08* i !7.95* -30 -28 -26 -24 -22 1.0330 I.1023 1.1754 1.253 1.334 2 !7.82* 2 !7.68* ; !7.53* i !7.37* i !7.20' 3.00995 .00996 .00998 .00999 .01001 32.36 30.47 28.70 27.05 25.51 -20 -18 - 1 6 - 1 4 - 1 2 1.419 1.509 1.604 1.704 1.809 ; !7.031 ;!6.85* :l6.65' :!6.45* 1 : !6.24* 3.01002 .01004 .01005 .01007 .01008 24.08 22.74 21.48 20.31 19.22 99.78 99.63 99.48 99.33 99.18 I.04154 .04398 .04655 .04923 .05204 3.98 4.38 4.77 5.17 5.57 85.98 85.82 85.67 85.51 85.36 89.95 90.20 90.44 90.69 90.93 0.0093 .0102 .0111 .0120 .0129 0.2048 .2045 .2042 .2038 .2035 -20 -la -16 -14 -12 - 1 0 -8 -6 - 4 _ 2 1.918 2.034 2.155 2.282 2.415 :l6.01' :25.78; >5.53* :!5.27* ::5.00* 0.01010 .OlOll .01013 .01014 .01016 18.19 17.23 16.33 15.48 14.69 99.03 98.88 98.72 98.57 98.42 I.05497 .05804 .06123 .06458 .06807 5.98 6.38 6.77 7.19 7.59 85.20 85.05 84.89 84.74 84.58 91.19 91.43 91.67 91.92 92.17 0.0138 .0147 .0155 .0164 .0173 0.2032 .2029 .2027 .2024 .2021 -10 * 8 - 6 - 4 - 2 0 2 4 6 8 2.554 2.699 2.852 3.011 3.178 14.72* 24.42* 24.11 * 23.79* 23.45* O.OlOd8 .01019 .01021 .01022 .01024 13.95 13.25 12.59 11.97 11.38 98.27 98.11 97.96 97.81 97.65 3.07171 .07550 .07945 .08356 .08785 7.99 8.39 a.79 9.19 9.60 84.43 84.27 84.12 83.96 83.80 92.42 92.67 92.91 93.16 93.40 0.0182 .0191 .019q .0208 .0217 0.2018 .2016 .2013 .2011 .2009 0 2 4 6 a 10 12 14 16 18 3.352 3.533 3.723 3.921 4.127 23.10* 22.73* 22.34' 21.94* 21.52' 0.01026 .01027 .01029 .01031 .01032 10.83 10.32 9.828 9.367 8.932 97.50 97.35 97.19 97.04 96.88 10.00 10.41 10.81 11.22 11.62 83.65 83.49 83.33 83.18 83.02 93.65 93.90 94.15 94.39 94.64 0.0225 .0234 .0243 .0251 .0260 0.2006 .2004 .2002 .2000 .1998 10 12 14 16 la 20 22 24 26 28 4.342 4.566 4.799 5.041 5.294 21.08' 20.62* 20.15* 19.66* 19.14* 0.01034 .01036 .01037 .01039 .01041 8.521 8.133 7.766 7.418 7.090 96.72 96.57 96.41 96.25 96.10 0.09233 .09694 .1018 .1068 1120 > 0.1174 .1230 .1288 .1348 .1410 12.03 12.43 12.85 13.26 13.67 82.86 82.70 82.55 82.39 82.23 94.89 95.14 95.39 95.65 95.89 0.0268 .0277 .0285 .0294 .0302 0.1996 .1994 .1992 .1990 .1988 20 22 24 26 28 30 32 34 36 38 5.556 5.829 6.112 6.407 6.712 18.61" 18.05' 17.4a* 16.881 16.25* 0.01042 .01044 .01046 .01048 .01049 6.779 6.484 6.205 5.940 5.688 95.94 95.78 95.62 95.46 95.30 0.1475 .1542 .1612 .1684 .1758 14.07 14.48 14.89 15.30 15.71 82.07 81.91 81.75 81.59 81.42 96.14 96.39 96.64 96.89 97.14 0.0310 .0319 .0327 .0335 .0344 0.1986 .1985 .1983 .19Bl .I980 30 32 34 36 38 40 42 44 46 48 7.030 7.359 7.700 8.054 8.420 15.61' 14.94* 14.24* 13.52' 12.78* 0.01051 .01053 .01055 .01056 .01058 5.450 5.223 5.008 4.804 4.609 95.14 94.98 94.82 94.66 94.50 0.1835 .I914 .1997 .2082 .2169 16.12 16.54 16.95 17.36 17.77 81.26 81.10 80.94 80.77 80.61 97.39 97.63 97.88 98.13 98.38 0.0352 .0360 .0368 .0377 .0385 0.1978 .1977 .1975 .1974 .1973 40 42 44 46 48 50 52 54 56 58 8.800 9.193 9.600 10.02 , 10.46 12.00* 11.20* 10.37' 9.52' 8.63' 0.01060 .01062 .01064 .01065 .01067 4.425 4.249 4.081 3.922 3.770 94.34 94.18 94.02 93.86 93.69 0.2260 .2354 .2450 .2550 .2653 18.19 18.61 19.03 19.44 19.86 80.44 80.28 80.11 79.94 79.77 98.63 98.89 99.14 99.38 99.63 0.0393 .0401 .0409 .0417 .0425 0.1971 .I970 .1969 .1968 .1966 x52 54 56 58 60 62 64 66 68 10.91 11.37 11.85 12.35 12.86 7.71' 6.76* 5.79* 4.77' 3.73* 0.01069 .01071 .01073 .01075 .01077 3.625 3.487 3.355 3.229 3.109 93.53 93.37 93.20 93.04 92.87 0.2759 .2868 .2981 .3097 .3216 20.27 20.69 21.11 21.53 21.95 79.61 79.44 79.27 79.10 78.92 99.88 100.13 100.38 100.62 100.87 a.0434 .0442 .0450 .0458 .0466 0.1965 .1964 .1963 .1962 .1961 60 62 64 66 68 70 72 74 76 78 13.39 13.94 14.51 15.09 15.69 2.65" 1.54* 0.39' 0.39 0.99 0.01079 .01081 .01083 .01084 :01086 2.995, 2.885 2.781 2.681 2.585 92.71 92.54 92.38 92.21 92.05 0.3339 .3466 .3596 .3730 .3868 22.37 22.79 23.21 23.63 24.06 78.75 78.58 78.41 78.23 78.05 101.12 101.37 101.61 101.86 102.12 0.0474 .0481 .0489 .0497 '.0505 0.1960 .1959 .1959 .I958 .1957 70 72 74 76 78 a0 a2 a4 86 aa 16.31 16.94 17.60 18.28 18.97 1.61 2.25 2.91 3.58 4.28 0.01088 .01090 .01092 .01094 .01096 2.494 2.406 2.322 2.242 2.165 91.88 91.71 91.54 91.38 91.21 0.4010 .4156 .4306 .4461 .4619 24.48 24.91 25.33 25.76 26.18 77.88 77.70 77.52 77.34 77.1 6 102.36 102.61 102.85 103.10 103.34 0.0513 .0521 .0529 .0537 .0544 0.1956 .1955 .1955 .1954 .1953 a0 a2 a4 86 aa 90 19.69 20.43 21.19 21.97 22.77 4.99 5.73 6.49 7.27 8.08 0.01098 .01100 .01103 .01105 .01107 2.091 2.021 1.953 1.888 1.826 91.04 90.87 90.70 90.53 90.36 0.4781 .4949 .5121 .5297 .5477 26.61 27.03 27.46 27.89 28.32 76.98 76.80 76.62 76.43 76.25 103.59 103.83 104.08 104.32 104.56 0.0552 .0560 .056B .0575 .0583 0.1953 .1952 .1951 .1951 .I950 90 92 94 96 98 -40 -30 -36 -34 -32 ;; 96 98 *Inches Yg 44.25 41.51 38.97 36.60 34.41 1 1 1 1 of mercury below one atmosphere. , vapor T EMP r (F) ENTROPY (Bfu/lb-R) liquid vi 3.00988 .00989 .00991 .00992 .00993 t vapor r t . ~ ; 1 1 . \ 1 ” 1 ’ 1 ~ 1 <I. 1~1:l~I~I~~I~:I~.\N’I’S TABLE ii--PROPERTIES lEMs I (F) t -ii% PRESSURE W/r ;-q7i n.1 Absolute P Gage r- 417 OF REFRIGERANT 1 1, LIQUID AND SATURATED VAPOR (Contd) V O Liii A E ( c u f ‘l/l l b ) liquid Vf 0.01109 102 104 106 108 23.60 24.45 25.32 26.21 27.13 P 8.90 9.75 10.62 11.52 12.44 110 112 114 116 118 28.08 29.05 30.04 31.06 32.11 13.38 14.35 15.35 16.37 17.41 0.01119 .01122 .01124 .01126 .Oll28 120 122 124 126 128 33.18 34.29 35.42 36.57 37.76 18.49 19.59 20.72 21.88 23.06 0.01130 .01133 130 132 134 -16 8 38.97 40.22 41.49 42.80 44.13 24.28 25.52 26.80 28.10 29.44 140 142 144 146 148 45.50 46.90 48.33 49.80 51.29 30.81 32.20 33.64 35.10 36.60 0.01142 .Olldd .01146 .01149 .01151 0.01154 .01156 .0115B 150 152 154 156 158 52.83 54.39 55.99 57.63 59.30 38.13 39.71 41.31 42.94 44.61 0.01166 .01168 .01171 .01173 .Ol.l76 160 61.01 46.31 0.01178 .Ollll .01113 .01115 .01117 .01135 .01137 .01139 .01161 .01163 r DE? (lb1 1.364 1.322 1.282 1.243 1.206 1.170 1.135 1.101 1.069 1.038 1.008 0.9788 liquid l/W 90.19 90.02 89.85 89.68 89.51 89.34 89.16 88.99 88.82 88.65 88.47 88.30 88.12 87.95 87.77 87.60 87.42 87.25 87.07 86.89 0.9508 .923B ‘8977 .8725 .84B2 86.68 86.50 86.32 86.14 85.96 0.8247 JO20 .7800 .7588 .7382 0.7183 85.78 85.60 85.41 85.23 85.04 84.86 vapor VB 1.766 1.708 1.653 1.600 1.549 1.500 1.453 1.408 TY ft) Vapor -r ENTHALPY 1 /%a liquid hf 0.5663 .5854 .6049 .6250 .6455 0.6667 .6882 .7104 .7330 .7563 0.7801 .8045 .8295 .8551 .8812 0.9081 0.9355 0.9636 0.9923 1.022 1.052 1.082 1.114 1.146 1.179 1.213 1.247 1.282 1.318 1.355 1.392 28.75 29.17 29.62 30.05 30.48 30.91 31.34 31.78 32.21 32.65 33.08 33.52 33.95 34.40 34.84 35.28 35.72 36.16 36.60 37.04 37.48 37.92 38.37 38.81 39.25 39.70 40.15 40.60 41.05 41.50 41.95 ENTROPY (Em/lb-R) TEMP (F) 100 102 104 106 108 110 112 114 116 118 73.34 108.18 73.14 72.94 72.73 72.52 72.31 108.42 108.65 71.47 71.25 108.89 109.12 109.35 110.28 110.51 110.74 1 10.96 111.20 111.43 .I945 .1945 0.0704 .0712 .0719 .0726 .0733 0.1945 .1944 0.0741 0.1944 .1943 .1943 .1943 140 142 144 146 148 0.0778 .0785 .0792 0.1943 .OBOO .1943 150 152 154 156 158 .1945 .1945 .1944 .1944 .1944 --T.0749 .0756 .0763 .0770 .1943 -T- --I71.04 70.82 70.60 70.38 120 122 124 126 128 .0674 .0682 .0689 .0697 .1943 .1943 130 132 134 136 138 160 TABLE 6-PROPERTIES TEMP (F) t T- l- PRESSURE (lb/w q it1.) ,Gage A bsolute P P I OF REFRIGERANT 113, LIQUID AND SATURATED VAPOR VOLUME (cu f’ liquid 4Clp.W vg 32.26 '6.8 1 '1.71 56.99 52.63 1 1 1 1 1 1 l- DENSITY ( l b / <:” ft) L iquid Vapor I.iquid 1 /Vf ht 11% 05.64 0.01216 1.97 05.50 .01302 2.36 05.37 .01395 2.76 05.23 .01493 3.16 3.56 05.09 .01597 0.1732 .I731 .1730 .1729 .1729 -20 -1.3 -16 -14 -12 0.0137 .0146 .0155 .0164 .0173 0.1728 .1727 .1726 .I726 .I725 -10 - a -6 - 4 - 2 0.0182 .0190 .0199 .0208 .0216 0.1725 .I724 .I724 .1723 .1723 0 2 4 6 a 10.00 10.41 10.81 11.22 11.62 0.0225 .0234 .0242 .0251 .0259 0.1723 .1722 .1722 .I722 .1722 10 12 14 16 18 12.03 12.44 12.85 13.26 0.0268 .0276 .06929 13.67 .0285 .0293 .0302 0.1722 .1721 .1721 .1722 .1722 20 22 24 26 28 0.07294 .07675 .08071 .08483 .08913 14.08 14.49 14.91 15.32 15.74 0.0310 .0318 .0327 .0335 .0343 0.1722, .1722 .1722 .1722 .1722 30 32 34 36 38 .09361 .09826 .I031 .1081 .1133 16.16 lb.57 0.0352 0.1723 .I723 .1723 .1724 .1724 40 42 44 46 48 0.1187 .1243 .1302 .1362 .1425 18.24 18.66 19.08 19.50 19.93 0.0393 .0401 .0410 .0418 .0426 0.1725 .1726 .1726 .1727 .I727 50 52 54 56 58 5.889 5.640 99.05 98.89 98.73 98.58 98.42 0.1490 .1557 .I626 .1698 .I773 20.35 20.77 21.19 21.62 22.05 0.0434 .0442 .0450 .0459 .0467 0.1728 .I729 .I729 .1730 .1731 60 62 64 66 68 0.01018 .01019 .01021 .01023 .01025 5.404 5.180 4.971 4.769 4.574 98.26 98.10 97.93 97.77 97.6 1 0.1851 .I931 .2012 .2097 22.48 22.90 23.33 0.0475 .0483 .0491 .0499 .0507 0.1731 .1732 .1733 .1734 .1735 70 72 74 76 78 15.87* 15.25* 14.60* 13.93: 13.24* 0.01026 0.1028 .01030 .01031 .01033 4.392 4.218 4.05 1 3.893 3.742 97.45 97.28 97.12 0.2277 .2371 0.0515 .0523 .0531 .oi39 .0547 0.1736 .1737 .1738 .1739 .I740 a0 a2 a4 86 aa 0.01035 .01037 .01039 .01040 .01042 3.600 3.463 3.333 3.208 3.089 96.63 96.46 96.30 96.13 95.96 0.2778 .2a88 , 3 0 01 .3117 .3237 26.80 27.24 10.07 12.53* 11.79' 11.03* 10.24' 9.42* 0.0555 .0563 .0571 .0578 .0586 0.1741 .1742 .I743 .1744 .1745 90 92 94 96 98 10.48 10.91 11.35 11.81 12.28 a.591 7.71' 6.82' 5.88* 4.93: 0.01044 .01046 .01048 .01050 .01051 2.976 2.867 2.762 2.662 2.567 95.79 0.3360 .3488 .3620 28.99 29.44 29.89 30.33 30.78 0.0594 .0602 .0610 .0618 0.1746 .1747 .1748 .1750 .17.51 100 102 104 106 108 0.4288 0.4600 0.4931 0.5280 0.5652 ;! 9 . 0 5 * ;! 8 . 9 8 * ;! 8 . 9 2 " ;! 8 . 8 5 * ;! 8 . 7 7 " 1.00953 .00954 .00955 .00957 .00958 58.61 54.88 51.42 18.23 15.25 1 04.96 I 04.82 I 04.68 1 04.54 1 04.40 0.01706 .01822 .01945 .02074 .02210 3.96 4.36 0.6046 3.00959 .00960 .00962 .00963 .00964 42.48 39.92 37.54 35.31 33.24 I 04.26 1, 0 4 . 1 2 11 0 3 . 9 8 11 0 3 . 8 4 11 0 3 . 7 0 0.02354 .02505 0.7369 0.7860 ;! 8 . 6 9 * !8.60* 18.51' 18.42; !a.32* .02a32 .03009 5.96 6.36 6.76 7.17 7.57 0.8377 0.8924 0.9503 1.011 1.075 1a.21* 28.10* 27.99' ?7.86* 27.73' 0.00966 .00967 .00968 .00970 .00971 31.31 29.52 27.84 26.27 24.81 11 0 3 . 5 6 103.41 103.27 11 0 3 . 1 3 11 0 2 . 9 8 0.03 194 .03388 .03592 .03806 .04031 7.98 8.38 8.78 9.19 9.59 10 12 14 16 18 1.142 1.213 1.280 1.366 1.448 27.60' 27.45' 27.30' 27.14* 26.97+ 0.00972 .00974 .00975 .00977 .00978 23.45 22.17 20.97 19.84 18.79 102.84 102.69 102.55 102.40 102.25 0.04265 .04511 .04769 .05040 .05322 20 22 24 26 28 1.534 1.624 1.719 1.818 1.922 26.80' 26.61; 26.42' 26.22* 26.01* 0.00979 .00981 .00982 .00984 .00985 17.81 16.89 0.05616 lb.02 15.20 14.43 102.10 101.96 101.81 101.66 101.51 30 32 34 36 38 2.031 2.145 2.264 2.388 2.519 25.79' 25.55' 25.31' 25.06' 24.79+ 0.00987 .00988 .00990 .00991 .00993 13.71 13.03 12.39 11.79 11.22 101.36 101.21 101.06 100.91 100.76 40 42 44 46 48 2.655 24.52" 24.23* 23.93' 0.00994 100.60 50 52 54 56 58 3.427 3.602 3.784 3.973 4.170 2-3 . 2 9-* 22.94, 22.59' 22.22+ 21.83' 21.43* .00997 .00999 .01000 10.68 10.18 9.703 9.253 8.830 0.0 1002 .01003 .01005 .01006 .01008 8.426 8.044 7.682 7.342 7.018 99.83 99.68 99.52 99.37 99.21 60 62 64 66 68 4.374 4.586 4.807 5.036 5.275 21.02* 20.59* 20.14' 19.67" 19.18' 0.01010 .OlOll .01013 .01015 .01016 6.713 70 72 74 76 78 5.523 5.780 6.042 18.68* a0 a2 a4 86 aa 6.902 90 92 94 96 98 a.545 8.908 9.281 100 102 104 106 ioa 0 2 4 6 a vapor 0.0092 .OlOl .OllO .0119 .0128 1 !9.31' 1! 9 . 2 7 * 2! 9 . 2 2 * 2! 9 . 1 6 * :! 9 . 1 1 * - 1 0 - a - 6 - 4 - 2 liquid Sf 0.0047 .EMP W t -30 .0065 .0074 .0083 0.2987 0.3214 0.3458 0.3718 0.3995 -20 --la -16 -14 -12 ENTROPY mu/ lb -R) Ig 0.1738 .1737 .I736 .1735 .1733 v( 1.00947 .00948 .00949 .00950 .00952 -30 -28 -26 -24 -22 ENTHALPY (I h/lb) 0.6462 0.6902 2.797 2.944 3.098 3.258 6.320 6.607 7.208 7.527 7.856 8.194 9.660 23.61* 18.16f 17.62* 17.06' 16.47' .00996 6.424 6.149 100.45 100.30 100.14 99.99 96.96 96.79 95.63 95.46 95.29 95.12 .02664 .05922 .06243 .06579 .2186 .2468 .2569 2672 .3756 .3896 .0056 72.33 1 75.49 4.76 5.16 5.56 71.39 71.27 70.80 70.68 70.56 77.75 78.03 79.18 79.46 79.75 .0360 16.99 17.41 17.82 .0368 .0377 .0385 23.76 24.19 24.63 25.06 25.49 25.93 26.36 27.67 28.11 28.55 65.32 91.68 65.18 65.04 64.90 64.75 91.98 92.28 92.57 92.86 .0626 Cmrrtrsy *Inches of mercury below one atmosphere. . -28 -26 -24 -22 ' TABLE 6--PROPERTIES TEMP (F) t -Xii 112 114 116 118 120 122 124 126 128 130 132 134 136 138 140 142 144 7°C -12” 152 154 1Sb 158 160 PRESSURE (Ib/sq in.) Absolute 1 Gage -iTG-k 18.45 19.11 19.79 20.48 ( 3.74 4.41 5.09 5.78 l- OF REFRIGERANT 113, LIQUID VOLUME (CU Liquid vt 0.01053 .01055 .01057 .01059 .OlObl 0.01063 .01065 .01067 .01069 .01071 0.01073 .01075 .01077 .01079 .01081 0.01083 .01085 .01087 .01089 .01092 0.0109A .01096 .01098 .01100 .01102 0.01105 b) Vapor vg 2.477 2.391 2.308 2.228 2.151 2.078 2.008 1.941 1.876 1.814 1.754 1.697 1.642 1.590 1.540 1.491 1.444 1.399 1.355 1.313 1.273 1.234 1.197 1.162 1.128 1.094 T- DENSITY (lb. /cu ft) Liquid Vapor l/V? l/% 94.95 0.4038 94.78 .4182 94.61 .4333 94.43 .4489 94.26 .4649 94.09 0.4813 93.92 .4981 93.74 .5153 93.57 .5330 93.39 .5514 93.22 0.5702 93.04 .5894 92.86 .6091 92.69 .6290 92.51 .6494 92.33 0.6707 92.15 .6926 91.98 .7150 91.80 .7379 91.62 .7615 91.44 0.7856 91.25 .8102 91.07 .8353 90.89 .a608 90.71 .a609 90.53 0.9141 r AND SATURATED VAPOR (Contd) ENTHALPY Liquid hf 31.22 31.67 32.12 32.57 33.03 33.48 33.93 34.38 34.83 35.29 35.75 36.21 36.67 37.13 37.59 38.05 38.52 38.98 39.45 39.92 40.38 40.85 41.32 41.79 42.26 42.74 -(Btu/lb) Latent hfe 63.71 63.56 63.40 63.25 63.09 62.93 62.78 62.62 62.46 62.30 62.14 61.97 61.80 61.64 61.48 61.31 61.13 60.96 60.79 60.6 1 60.44 60.27 60.09 59.91 59.73 59.55 ENTROPY Vapor hg 94.93 95.23 95.52 95.82 96.12 96.4 1 96.71 97.00 97.29 97.59 97.89 98.18 98.47 98.77 99.06 99.36 99.65 99.94 100.24 100.53 100.82 101.11 101.41 101.70 101.99 102.29 Sf 0.0634 .0641 .0649 .0657 .0665 0.0673 .0680 .0688 .0696 .0704 0.0712 .0719 .0727 .0735 .0742 0.0750 .0758 .0765 .0773 .0781 0.0789 .0796 .0804 .0812 .0819 0.0827 vapor Ig 0.1752 .1753 .I755 .1756 .I757 0.1758 .1760 .I761 .1763 .I764 0.1765 .I767 .1768 .1770 .I771 0.1773 .I774 .1775 .I777 .I778 0.1780 .1782 .1783 .1785 .I786 0.1788 l- TEMP (F) t 110 112 114 116 118 120 122 124 126 128 130 132 134 136 138 140 142 144 146 148 150 152 154 156 158 160 TABLE 7-PROPERTIES TEMP (F) t PRES (lb/s ,brolute P RE n.) GLlge l- OF REFRIGERANT 114, LIQUID VOLUME (cu lb) liquid vt 3.009469 .009484 .009506 .009513 .009528 “cl j1.26 19.83 14.87 11.69 18.92 Vapor r DENSITY (lb/ ff) L liquid Vapor 1 lVf l/v, 105.603 0.01951 105.398 .02007 .02229 105.193 105.058 .0240 104.924 .0257 l- AND SATURATED VAPOR ENTHALPY liquid ht -8.73 -8.30 -7.87 -7.44 -7 . 0 1 Latent hfa ENTROPY vapor f 69.17 69.01 68.85 68.69 68.53 h, 60.44 60.71 60.98 61.25 61.52 mu/ liquid If -R) l- Vapor TEMP (F) t - 80 -70 -76 -74 -72 0.464 0.50 0.535 0.58 0.62 P. !8.97" !8.90* !8.85* !8.73* 18.66' -70 -68 -66 -64 -62 0.670 0.72 0.775 0.833 0.895 !8.59* !8.46* !8.33* !8.23* !8.10* 3.009543 .009558 .009573 .009589 .009604 $6.40 34.20 $1.74 19.77 27.79 104.790 104.624 104.458 104.292 104.126 0.02747 .02925 .0315 .0336 .0360 -6.57 -6.14 -5 . 7 1 -5.28 -4.84 68.36 68.20 68.04 67.88 67.72 61.79 62.06 62.33 62.60 62.88 -60 -58 -56 -54 -52 0.959 1.028 1.10 1.175 1.26 27.99; !7.83* t7.68' !7.52* t7.35* 0.009619 .009635 .009651 .009666 .009682 26.06 24.55 22.94 21.50 20.16 103.960 103.792 103.622 103.452 103.282 0.03838 .04075 .0436 .0465 .0496 -4.40 - 3.96 -3.52 -3.08 -2.64 67.56 67.39 67.22 67.05 66.89 63.16 63.43 63.70 63.97 64.25 -50 -48 -46 -44 -42 1.349 1.438 1.535 1.635 1.745 17.20' 17.0 * 16.8 * 26.6 * 26.4 * 0.009698 .009714 .009731 .009747 .009764 18.96 17.85 16.80 15.75 14.87 103.113 102.938 102.766 102.594 102.422 0.05274 .0560 .0595 .0635 .06725 -2.20 -1.76 -1.32 -0.88 -0.44 66.73 66.56 66.39 66.23 66.07 64.53 64.80 65.07 65.35 65.63 -0.0054 - .0043 - .0032 - .0021 - .OOlO 0.1575 .I574 .1573 .1572 .I571 -50 -48 -46 -44 -42 -40 -38 -36 -34 -32 1.866 1.990 2.121 2.259 2.404 26.12* 25.87* 25.60* 25.32" 25.03' 0.00978 .00980 .00981 .00983 .00985 14.02 13.20 12.44 11.73 11.07 102.25 102.08 101.90 101.72 101.55 0.07132 .07574 .08038 .08524 .09034 0.00 0.45 0.91 1.36 1.81 65.91 65.74 65.56 65.38 65.21 65.91 66.19 66.47 66.74 67.02 0.0000 .OOll .0021 .0032 .0042 0.1571 .1570 .1569 .1568 .I567 -40 -38 -36 -34 -32 -30 -28 -26 -24 -22 2,557 2.718 2.887 3.064 3.249 24.72* 24.39" 24.04* 23.68' 23.31' 0.00987 .00988 .00990 .00992 .00994 10.45 9.877 9.338 8,.833 8.362 101.37 101.19 101.01 100.83 100.65 0.09568 .1013 .1071 .1132 .1196 2.27 2.72 3.17 3.63 4.08 65.03 64.86 64.68 64.51 64.34 67.30 67.58 67.86 68.14 68.42 0.0053 .0063 .0074 .0084 .0095 0.1567 .I567 .1566 .1565 .1565 -30 -20 -26 -24 -22 -20 -18 -16 -14 -12 3.444 3.648 3.862 4.085 4.319 22.91' 22.49* 22.06* 21.61* 21.13* 0.00995 .00997 .00999 .OlOOl .01003 7.921 7.508 7.121 6.757 6.416 100:47 100.29 100.11 99.92 99.74 0.1263 .I332 .I404 .1480 .1559 4.54 4.99 5.44 5.90. 6.35 64.16 63.99 63.81 63.64 63.46 68.70 68.98 69.26 69.54 69.82 0.0105 .0116 .0126 .0136 .0146 0.156; .1565 .1564 .I564 .I564 -20 -18 - 1 6 - 1 4 - 1 2 -10 - 8 - 6 - 4 - 2 4.564 4.819 5.086 5.365 5.655 20.63' 20.11* 19.57' 19.00' 18.41* 0.01005 .01006 .01008 .01010 .01012 6.095 5.794 5.510 5.244 4.992 99.56 99.37 99.19 99.00 98.81 0.1641 .1726 .I815 .1907 .2003 6.81 7.26 7.72 8.18 8.63 63.29 63.11 62.94 62.77 62.59 70.10 70.38 70.66 70.94 71.22 0.0157 .0167 .0177 .0187 .0197 0.1564 .1564 .1564 .1564 .I565 - 1 0 -8 - 6 - 4 - 2 0 5.958 6.274 6.603 6.945 7.301 17.79* 17.15" 16.48' 15.78* 15.06* 0.01014 .01016 .01018 .01020 .01022 4.756 4.533 4.322 4.123 3.935 98.62' 98.44 98.25 98.06 97.87 0.2103 .2206 .2314 .2425 .2541 9.09 9.54 10.00 10.46 IO.91 62.42 62.24 62.07 61.89 61.71 71.50 71.78 72.07 72.35 72.63 0.0207 .0217 .0227 .0236 .0246 0.1565 .1565 .1565 .1566 .1566 0 2 4 6 0 15 16 18 7.671 8.057 8.457 8.873 9.305 14.31* 13.52' 12.7! * 11.86* 10.98s 0.01024 .01026 .01028 .01030 .01032 3.758 3.59J 3.432 3.282 3.140 97.68 97.48 97.29 97.10 96.90 0.2661 .2785 .2914 .3047 .3185 11.37 11.83 12.29 12.75 13.20 61.54 61.36 61.19 61.01 60.83 72.91 73.19 73.47 73.75 74.04 0.0256 .0266 .0275 .0285 .0295 0.1566 .1567 .1567 .I568 .1568 10 12 14 16 18 20 22 24 26 28 9.753 10.22 10.70 II.20 11.72 10.07* 9.12' a.14* 7.12' 6.07* 0.01034 .01036 .01038 .01040 .01043 3.005 2.877 2.756 2.641 2.532 96.71 96.51 96.32 96.12 95.92 0.3328 .3476 .3629 .3786 .3949 13.66 14.12 14.58 15.05 15.51 60.65 60.48 60.30 60.12 59.94 74.32 74.60 74.88 75.17 75.45 0.0304 .0314 .0323 .0333 .0342 0.1569 .1569 .1570 .I571 .1571 20 22 24 26 28 30 32 34 36 38 12.25 12.81 13.38 13.98 14.59 4.99' 3.05* 2.69' 1.47* 0!221 0.01045 .01047 .01049 .01051 .01053 2.429 2.330 2.236 2.147 2.062 95.73 95.53 95.33 95.13 94.93 0.4118 .4292 .4472 .4658 .4849 15.97 16.43’ 16.89 17.36 17.82 59.76 59.58 59.40 59.22 59.04 75.73 76.01 76.29 76.58 76.86 0.0352 .036l .0370 .0380 .0389 0.1572 .1573 .I574 .1575 .1575 30 32 32 36 38 40 42 44 46 48 15.22 15.88 16.56 17.26 17.98 0.52 1.18 1.86 2.56 3.28 0.01056 .01058 .01060 .01063 .01065 1.982 1.905 1.832 1.762 1.695 94.73 94.52 94.32 94.12 93.91 0.5047 .5250 .5460 .5676 .5899 18.28 18.75 19.21 19.68 20.14 58.86 58.67 58.49 58.31 58.12 77.14 77.42 77.70 77.99 78.27 0.0398 .0408 .0417 .0426 .0435 0.1576 .I577 .1578 .I579 .1580 40 42 44 46 48 50 52 54 56 58 18.73 19.50 20.29 21.11 21.96 4.03 4.80 5.59 6.41 7.26 0.01067 .01070 .01072 .01074 .01077 1.632 1.571 1.513 1.458 1.405 93.71 93.50 93.30 93.09 92.88 0.6129 .6365 .6609 .6859 .7117 20.61 21.08 21.54 22.01 22.48 57.94 57.75 57.56 57.38 57.19 78.55 78.83 79.11 79.39 79.67 0.0444 .0453 .0463 90472 .0481 0.1581 .1582 .1583 .1584 .1585 50 52 54 56 58 2 4 6 a 10 12 *Inches of mercury below one atmosphere. -0.0227 - .0215 - .0203 - .Ol91 - .0179 Ig 0.1595 .I593 .1592 .1590 .1589 - 80 -78 -76 -74 -72 -0.0167 - .0155 - .Ol43 - .0132 - . 0121 ~ 0.1587 .1586 .1585 .I583 .1582 -70 -60 -66 -64 -62 -0.0110 .-. .0098 - .0087 - .0075 - .0064 0.1580 .1579 .1578 .1577 .l576 -60 -58 -56 -54 -52 l (:I I \I’1 I~.I< I, I~I~:l~‘I~I(;~:l~.\s’I’s 4-2 1 TABLE ‘/-PROPERTIES OF REFRIGERANT 114, LIQUID AND SATURATED VAPOR (Conic!) TEMl W I 60 62 64 66 68 70 72 74 76 78 PRESSURE (lb/w .brolute P 22.83 23.72 24.64 25.59 26.57 P a.13 9.02 9.94 lo.89 II.87 27.57 28.61 29.67 12.87 13.91 14.97 1.354 1.306 1.260 1.216 1.174 0.9869 0.01104 .01107 .OlllO .01112 .01115 0.9541 90.56 .a353 90.34 90.13 89.91 89.69 42.02 43.44 44:a9 24.59 25.94 27.32 28.74 30.19 0.01 i ia .01120 .01123 .01126 .01129 0.8084 .7a27 .7579 .7340 .7111 46.39 31.69 47.92 49.48 51.09 52.73 33.22 34.78 0.01132 .01135 .01137 .01140 .01143 0.6890 A677 54.41 39.71 41.44 43.20 45.00 46.85 0.01146 48.74 0.01162 52.65 .01165 .oliba 30.76 31.88 90 92 94 -_ 39.29 120 122 124 126 128 Vapor vg 36.69 37.97 40.64 56.14 57.90 59.70 61.55 63.44 65.37 67.35 1.133 1.094 1.057 1.021 17.18 0.01091 .01094 .01097 .01099 .01102 icr 19.52 20.74 21.99 23.27 DENSITY (lb/ ff) Liquid Vapor 1 IVf l/h 92.68 0.7383 92.47 .7655 92.26 .7936 92.05 .a225 91.84 .a521 0.8826 0.9140 33.04 34.22 35.44 110 112 114 116 ila Liquid Vf 0.01079 .oioa2 .oioa4 .01086 .oioa9 HE lb) 91.63 91.41 91.20 90.99 90.77 a0 02 a4 06 aa -100 102 104 106 108 Gage vo (CU 16.06 36.39 38.03 .01149 .01152 .OllSb .01159 0.9462 0.9793 1.013 Liquid hf 22.95 23.42 23.89 24.36 24.83 1atem 56.43 56.24 h, 79.95 80.23 80.51 80.79 al.07 25.30 25.78 56.04 al.35 55.65 al.90 82.18 82.46 26.25 26.73 hfg 57.00 56.81 56.62 55.85 55.45 27.20 55.26 1.048 I.084 1.121 1.159 1.197 27.68 28.15 28.63 29.11 29.58 55.06 89.47 89.25 89.03 88.81 88.59 1.237 1.278 1.320 1.406 30.06 30.54 31.02 31.50 31.99 .boa4 88.37 88.15 87.93 87.70 87.48 1.452 1.498 1.545 1.594 1.644 0.5901 .5724 .5554 .53a9 .5230 87.25 87.01 86.77 86.54 86.31 0.5077 .4929 .47a7 .9226 .a923 .a632 .6472 .6274 ENTROPY ( B t u / lbl R) ENTHALPY h/lb) V C!por 81.62 liquid Sf 0.0490 .0499 .05oa .0517 .0526 0.0534 .0543 .0552 .0561 .0570 54.25 82.73 83.01 83.29 83.56 83.84 0.0579 .05a7 .0596 .0605 .Obl3 54.05 53.84 84.1 i 84.39 0.0622 53.64 84.66 32.47 32.95 33.43 33.92 34.40 53.01 52.80 52.59 52.37 85.48 85.75 86.02 1.695 1.747 1.801 i.a56 1.912 34.89 35.38 35.87 1.362 54.86 54.66 54.46 53.43 53.22 84.93 85.21 .063l .0639 .0648 .0656 0.0665 .0674 .0682 ‘EMP (F) vapor 60 62 64 66 68 0.1593 .I594 .1595 .1596 .1597 70 72 74 76 78 0.1599 i .I600 a0 a2 a4 86 aa .I601 .1603 *lb04 0.1605 .1607 .I608 .1609 .lbll 90 92 94 96 98 0.1612 .I614 .1615 .1617 .1618 100 102 104 106 108 110 112 114 116 11.9 86.29 86.56 .0691 86.83 87.09 87.36 87.89 0.0708 .0716 .0725 .0733 .0741 0.1619 .1621 .1622 36.35 36.84 51.94 51.72 51.50 51.28 51.05 1.970 2.029 2.089 2.151 2.215 37.33 37.83 38.32 38.81 39.31 50.83 50.60 50.37 50.14 49.91 88.16 88.42 88.69 88.95 89.21 0.0750 .075a .0767 .0775 .0783 0.1627 0.1634 .1635 52.16 87.63 .0699 t )B 0.1587 .i5aa .15a9 .1590 .1591 .1624 .1625 .01171 .01175 .4649 .4515 86.08 as.85 85.61 85.37 85.13 0.01178 .oiiai .oi185 .oiiaa .01192 0.4387 .4262 .4142 .4025 .3912 84.89 84.65 84.41 84.16 83.91 2.280 2.346 .2.414 2.484 2,556 39.80 40.30 40.80 41.29 41.79 49.67 77.90 80.15 82.44 60.99 63.20 65.45 67.74 49.44 49.20 48.96 48.72 89.47 89.73 89.99 90.25 90.51 0.0792 .oaoo .oaoa .0816 .oa25 .1637 .163a .1640 130 132 134 136 138 84.79 70.09 0.01195 0.3803 83.66 2.629 42.29 48.47 90.76 0.0833 0.1641 140 50.67 69.37 71.43 54.67 130 132 134 136 138 73.54 58.84 75.69 140 56.73 .1628 .1630 .1631 .I633 120 122 124 126 128 4-B CHAPTER 2. BRINES This chapter provides information to guide the engineer in the selection of brines, and includes the properties of the commonly used brines. At temperatures above 32 F, water is the most commonly used heat transfer medium for conveying a refrigeration load to an evaporator. At temperatures below 32 F, brines are used. They may be: 1. An aqueous solution of inorganic salts, i.e. sodium chloride or calcium chloride. For low temperatures, a eutectic mixture may be used. An aqueous solution of organic compounds, i.e. alcohols or glycols. Ethanol water, methanol water, ethylene glycol and propylene glycol are examples. 3. Chlorinated or fluorinated hydrocarbons and halocarbons. A solution of any salt in water, or in general any solution, has a certain concentration at which the freezing point is at a minimum. A solution of such a concentration is called a eutectic mixture. The temperature at which it freezes is the eutectic temperature. A solution at any other concentration starts to freeze at a higher temperature. Figure 11 illustrates the relationship between the freezing point (temperature) of a brine mixture and the percent of solute in the mixture (concentration). Chart 18 covers a range of temperatures wide enough to reveal the two freezing point curves. -“hen the temperature of a brine with a concentrr_-,on below the eutectic falls below the freezing point, ice crystals form and the concentration of the residual solution increases until at the eutectic temperature the remaining solution reaches a eutectic concentration. Below this temperature the mix-’ ture solidifies to form a mechanical mixture of ice and frozen eutectic solution. When the temperature of a brine with a concentration above the eutectic falls below the freezing point, salt crystallizes out and the concentration of the residual solution decreases until at the eutectic temperature the remaining solution reaches a eutectic concentration. Below this temperature the mixture solidifies to form a mechanical mixture of salt and frozen eutectic solution. This chapter includes a discussion of these brines, also tables and charts indicating properties. I FREEZlNG POINT A\\\\W LlOUlD SOLUTION OF EUTECTIC - CONCENTRATION,PERCENT SOLUTE IN MIXTURE Courtesy of ASHRAE Guide and Data Book 1963 F IG . 11 BRINE - BR I N E M I X T U R E SELECTION The selection of a brine is based upon a consideration of the following factors: 1. Freezing Point - Brine must be suitable for the lowest operating temperature. .2. Application - When using an open piping system, the possibility of product contamination by the brine should be checked. 3. Cost - The initial charge and quantity of make-up required are factors in the determination of costs. 4. Safety - Toxicity and flammability of brine. 5. Thermal Performance - Viscosity, specific gravity, specific heat and thermal conductivity are utilized to determine thermal performance. 6. Suitability - Piping and system equipment material require a stable and relatively corrosive-free brine. 7. Codes - Brine must be acceptable by codes, ordinances, regulatory agencies and insuror. 4-24 l’.\IC’l‘ I . l~1:I~I~I(;I:I~.\N’I’S, I1I<INI:S. 0II.S TABLE I-TYPICAL BRINE APPLICATIONS Snow Melting X X X X Low Temperature (Special) lea Cream The most common brines are aqueous solutions ^ calcium chloride or sodium chloride. Although ,,oth of these brines have the advantage of low cost, they have the disadvantage of being corrosive. To overcome corrosion, an inhibitor may be added to the brine. Sodium dichromate is a satisfactory and economical inhibitor. Sodium hydroxide is added to keep the brine slightly alkaline. Sodium chloride is cheaper than calcium chloride brine; however, it cannot be used below its eutectic point of -6 F. It is preferred where contact with calcium chloride brine cannot be tolerated, for example, with unsealed foodstuffs. The use of calcium chloride of commercial grade is not satisfactory below -40 F. Systems using aqueous solutions of alcohol or glycol are more susceptible to leakage than those using salts. A disadvantage of alcohol is its flatimability. It is utilized mainly in industrial proc, =qses where similar hazards already exist, and in e same temperature range as the salts (down to -40 F). The toxicity of methanol water (wood alcohol) is a disadvantage. Conversely, the nontoxicity of ethanol water (denatured grain alcohol) is an advantage. Corrosion inhibitors should b’e used with alcohol type brines as recommended by the manufacturer of the alcohol. Aqueous solutions of glycol are utilized mainly in commercial applications as opposed to industrial processes. Ethylene and propylene glycol possess equal corrosiveness which an inhibitor c&n neutralize. Galvanized surfaces are particularly prone to attack by the glycols and should be avoided. An inhibitor and potable water are recommended for making up glycol brines. The glycol manufacturer sho~~ltl be consulted for inhibitor recommen. X X X X dations. Some manufacturers have a brine sample analysis service to assist in maintaining a satisfactory brine condition in the system. Heat, transfer glycols are available with nonoily inhibitors which do not penalize heat transfer qualities (sodium nitrite or borox). Glycols can be used as heat transfer media at relatively high temperatures. With stabilizers, glycol oxidization in air at high temperatures is eliminated for all practical purposes. Ethylene glycol is more toxic than propyle;e glycol, but less toxic than methanol water. Propylene glycol is pieferred to ethylene glycol in food freezing for example. Chlorinated and fluorinated hydrocarbons are expensive and are used in very low temperature work (below -40 F) . Table 8 presents typical applications for the vario u s brives. Load, brine quantity and temperature rise are all related to each other so that, when any two are known, the third may he found by the formula: Load (tons) = g-pm X temp rise (F) X sp gr X Cp 24 where: sp gr = specific gravity of I)rine Cp = specific heat of brine (fitu,/ll)-F) PIPING All materials in the piping system including flange gaskets, valve seats and packing, pump seals and other specialities must be compatible with the brine. Copper tubing (except for the salt brines) and standard steel pipe are suitable for general use. The pump rating and motor horsepower should be based on the particular brine used and the actual operating temperature. :HAPTER 4-25 2. BRINES FRICTION LOSS To determine the friction loss in a brine piping system, the engineer should first calculate the loss as if water were being used. A multiplier is then used to convert the calculated loss to the actual loss for the brine system. The multiplier is calculated as follows: Refer to Chart I. For a Reynolds number of 9.52 X 1Oa and a relative roughness of .000104, the chart indicates friction factor fb = .031. Specilic gravity of fresh 55 F = 1.00 water at a mean temperature of Refer to Chart 28. Viscosity of fresh water at a mean temperature of 55 F = 1.2 centipoises. 7740 x ,575 x 4 . 2 9 x 1.00 Multiplier = sp gr X 7 w where: sp gr = specific gravity of brine fb = friction I actor for the brine f, = friction factor for water at the brine Re (water) = velocity fb Friction multiplier = sp gr (brine) X r ,031 W = 1.05 x .o’L7 = 1.21 Brine friction loss = 1.21 X 7.5 psi = 9.08 psi or 7540 X d X v X sp gr = P’ where: 9.08 x 2.31 1.05 = 2 0 . 0 ft brine PUMP BRAKE HORSEPOWER d = inside pipe diameter (in.) v = brine velocity (ftjsec) To determine the horsepower required by a pump with brine, the following formula may be used: lb/cu ft sp gr = specific gravity of brine =62.5 p’ = viscosity (centipoises) = = 1 5 , 9 0 0 = 1.59 x 10’ Refer again to Chart 1. For a Reynolds number of 1.59 X 10’ and a relative roughness of .000104, the chart indicates a friction factor f, = ,027. Friction factor is determined from the Reynolds number. The Reynolds number is: Re = 1.2 ahsolute viscosity, lb/(hr) gpm X total head (ft brine) X sp gr (ft) bhp = 2.42 3960 X eff where: Example 1 illustrates the use of the multiplier to determine the brine friction loss thru a heat transfer coil. am = gallons/min. of brine total head = total pump head (ft brine) = specific gravity of brine sp gr eff = pump efficiency Example 7 - Friction Loss Multiplier Given: A 5/8 in. copper tube coil with a circuit water velocity of 4.29 ft/sec and a pressure drop of 7.5 psi. Mean water temperature = 55 F. Find: Fric’ ‘II loss multiplier and pressure drop when using et],. ie glycol at a mean brine temperature of 92.5 F and 417; sblution by weight at the same circuit liquid velocity. Solution: Refer to Chart 1. E .00006 T = 7ji5 = .000104 where: f = absolute roughness of drawn tubing d = inside diameter of sh in. copper tubing Refer to Chart 19. Specific gravity of ethylene glycol at a mean brine temperature of 92.5 F and 41% solution by weight is 1.05. Refer to C h a r t 18. Viscosity of ethylene 2.1 centipoises. 7740 x ,575 Re= glycol at the same conditions x 4 . 2 9 x 1.05 2.1 = 9520 = 9 . 5 2 x 108 equals BRINE PROPERTIES Specific gravity, viscosity, conductivity, specific heat, concentration, and freezing and boiling points are important factors in the consideration of liquids other than water suitable for cooling and heating purposes. High values of specific gravity, conductivity aixd specific heat, and low values of viscosity, promote a high rate of heat transfer. High values of specific gravity and viscosity result in high pumping head and consequentIy high pumping costs. High specific heats are desirable in that they reduce the quantity of liquid required to be circulated or stored for a given duty. Low viscosities are desirable from a standpoint of both rate of heat transfer and low pumping costs. They are particularly desirable at the lower temperatures where the viscosity increases. Table 9 is a tabulation of the various brines covered in this chapter, giving the properties of these brines at different temperatures and suitable concentrations. Charts 2 to 28 present the viscosity, specific gravity, specific heat and thermal conduc- TABLE 9-BRINE PROPERTIES olutior I bv WI) I )entity (%I 12 12 15 20 25 30 (1 b/cu ft) 68.2 30 Sodium Chloride Calcium Chloride Methonot Water Ethanol Water Ethylene Glycol Propylene Glycol 69.2 61.5 61.0 64.7 64.5 .a3 I .oo 1.04 .92 .94 21 20 22 25 35 40 72.0 74.8 15 Sodium Chloride Calcium Chloride Methanol Water Ethanol Water Ethylene Glycol Propylene Glycol 30 .72 .97 1.02 - 3 0 Calcium Chloride Methanol Water Ethanol Water Ethylene Glycol Propylene Glycol 25 35 36 45 50 30 45 52 55 60 78.4 - 5 Calcium Chloride Methanol Water Ethanol Water Ethylene Glycol Propylene Glycol Temp. (F) Brine 60.4 61.0 66.0 65.3 60.0 60.6 67.4 66.5 82. I 60.0 59.5 69.0 67.2 Specific Heat Btu/lb-F .8b .8b .09 .b7 .89 .95 .79 .a3 .b3 30 .a1 .73 .77 Thermal Cond. (Btu/hr rq ft-F/ft) .28 .32 .28 .27 .30 .26 .25 .31 .26 .25 .28 .24 ‘29 .23 .22 .25 .23 -28 .22 .19 .22 .21 =coefficient of heat transfer between brine and surface (Btu/hr-rq ft-F), (ft/sec), at Re= 3500 for ,554 in. ID tubing. t Vb = minimum brine velocity IAbove IO ftjsec. I; hb tivity of the brines for various mean brine temperntures and compositions. Note that specific gravity for propylene glycoI (Chart 23) in the composition range of 507” t o ircorlty ,eering Point ,oiling Point 2.2 2.4 3.2 5.5 3.7 8.0 W) 17.5 19.0 13.5 12.0 12.9 13.0 (F) 215. 213. 187. 189. 217. 216. 4.2 4.8 5.3 8.2 6.8 20.0 I .o 1.0 4.5 4.5 0.0 - 4.2 10.3 9.9 13.5 17.2 80.0 [centi- hb * vb t Relative Cart per Gal. of Solution 2.55 2.62 2.45 2.37 2.52 2.47 Tii971 781 621 775 525 1.61 1.78 2.63 4 60 2.92 6.35 1 3 13 20 42 43 216. 214. 182. 187. 2 I 9. 218. 2.57 2.77 693 730 599 504 576 103 2 . 9 0 3.28 4.44 6.05 5.25 1 5 19 25 60 58 -21.0 - 22.0 - 16.0 - 15.5 - 29.0 215. 176. 183. 223. 222. 2.85 2.82 2.62 2.82 2.72 - 216. 2.90 3.13 3.1 1 2.98 2.90 soiser) 27.8 18.0 20.2 75.0 700.0 at 7 T0 ipm/ton 47.0 45.0 50.0 43.0 55.0 velocity I 110 171. 179. 227. 227. r .554 iin. I D deg rise 2.56 2.41 2.65 2.58 513 98 97 103 98 110 91 83 93 91 -L 6.75 8.40 A 6 30 35 78 75 8 39 50 97 90 tubing. IOOY” (same mean brine temperature, F) is the same for two compositions. Specific gravity alone, therefore, is not a reliable method of determining the solution composition of this brine. . . -.. .09 .OB i iiiii-liiii I/III II I I Ill1 i I I //iii IliCi II /I/~-H~--I-.i~~I~~~I---l~l 02 ,015 .Ol .OOB !!!!!!!>!I !! I L I !!!‘-+““-i’..’ III1 TrtT I / 006 l-m ,004 ,001 0008 0006 IIIIIIII I I I NWII I IIIII ! I, I,, II, , 106 2 3 4 5 6 8 IO’ 2 3 4 5 6 8 \ ‘6 7740(d) Y (sg) 2 =.000.005 : i .000,001 = P’ 4 REYNOLDS NUMBER (f?t?) 56 8 d = inside pipe diameter Surface c (in.) Drawn tubing (very smooth surfaces of all kinds) 0.00006 sg = specific gravity Commercial steel or wrought iron 0.0018 /.L’ = viscosity (centipoises) Galvanized 0.006 iron v = brine velocity (ft/sec) E = absolute roughness I( WIII z w $ ” : n CHART 2-SODIUM CHLORIDE-VISCOSITY 20% -10 0 IO MEAN SOLUTION BY 20 BRINE TEMPERATURE 30 (F) 40 50 CHART J-SODIUM CHLORIDE-SPECIFIC GRAVITY I .I8 I. 16 0 *With reference to F water. 20 MEAN BRINE 40 TEMPERATURE 60 (F) -I--30 I’\I<‘l‘ I. I~l~:I~‘I~l~~1~:l1.\N’I’S. l~l<I~lI:.S. OILS CHART 4-SODIUM CHLORIDE-SPECIFIC HEAT . % . I j -1.. -4 ,I / / ,..I . . .,... , .9l . .9c .8C .8E I- j- . . . . i ,\6-.. . i . . i. .{ I .81 !O 0 + 20 MEAN, BRINE 40 60 TEMPERATURE (F) 80 Courtesy of Dow Chcmid Co. CHART 5--SODIUM CHLORIDE-THERMAL CONDUCTIVITY z cu. 0 m I a I \ c ? m / .30 , e .25 . MEAN BRINE TEMPERATURE(F) 30 50 70 CHART 6-CALCIUM CHLORIDE-VISCOSITY MEAN B R I N E T E M P E R A T U R E (F) 30 20 cn u fn 0 a t 2 w u > I: .. 9-. 8 7 6 M E A N B R I N E T E M P E R A T U R E (F) . CHART 7-CALCIUM CHLORIDE-SPECIFIC GRAVITY . __ i-- - 4 0 - 20 0 MEAN *With reference to 60 F water. 20 BRINE 4 0 60 80 T E M P E R A T U R E (F) Courtesy of Dow Chemiral Co. CHART S-CALCIUM CHLORIDE-SPECIFIC HEAT -40 -20 0 MEAN BRINE 20 40 TEMPERATURE (F) 60 80 Courtesy of Dow Chemical Co. CHART 9-CALCIUM CHLORIDE-THERMAL CONDUCTIVITY .33 . ^;,. . ------+.,-- .28 .27 -40 -30 -20 -10 0 IO 20 30 MEAN BRINE TEMPERATURE 4 0 ( F) 5 0 6 0 7 0 80 CHART IO-METHANOL BRINE-VISCOSITY 70 60 50 I - 50 - 4 0 - 3 0 - 2 0 - IO MEAN BRINE 0 IO 2 0 TEMPERATURE Courtesy of 30 4 0 50 (F) Carbide and Carbon Clmnicals Corporation (:H,\1”I‘EIi 2. 4-37 11IIINES - - - - CHART ll--METHANOL BRINE-SPECIFIC GRAVITY .8 .8 - 5 0 - 4 0 -30 ‘W ith reference to 60 F water. - 2 0 MEAN -10 BRINE 0 IO TEMPERATURE Courtesy 20 (F) 30 4 0 5 0 of Carbide and Carbon Chemicals Corporatim l'.\I<'I‘ 4-38 CHART 12--METHANOL I. IIEl~lil(;l-li,\N~1‘S, BRINE--SPECIFIC HEAT 1.10 1.05 I.OC c ! 0.95 m -I '= 0.9( im " 0.8! 2 g 0.81 0 ic E 0.7! w a m 0.7f 0.6: 0.6( 0.5: -60 -50 -40 -30 MEAN -20 -IV B R I N E TEMPER:T”RE’~F~ zv I 3v IIRINES. OILS CHART 13--METHANOL BRINE-THERMAL CONDUCTIVITY .32 .I0 - 5 0 - 4