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- Sunil Kumar
- Muhammed I Hussain

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THIS HANDBOOK PROVIDES COMPREHENSIVE TECHNICAL INFORMATION TO HEATING, VENTILATING, AND AIR CONDITIONING ENGINEERS, DESIGNERS AND PRACTITIONERS

HVAC: Handbook of Heating, Ventilation and Air Conditioning for Design and Implementation BY

ALI VEDAVARZ, PH.D., PE Deputy Director of Engineering, New York City Capital Projects, New York City Housing Authority and Industry Professor, Polytechnic University, Brooklyn, NY

SUNIL KUMAR, PH.D. Professor of Mechanical Engineering and Dean of Graduate School Polytechnic University, Brooklyn, NY

MUHAMMED IQBAL HUSSAIN, PE Mechanical Engineer, Department of Citywide Administrative Services New York City, NY

2007 INDUSTRIAL PRESS INC. NEW YORK

COPYRIGHT © 2007 by Industrial Press Inc., New York, NY. Library of Congress Cataloging-in-Publication Data Vedavarz, Ali. HVAC: handbook of heating ventilation and air conditioning / Ali Vedavarz, Sunil Kumar, Muhammed Hussain. p. cm. ISBN 0-8311-3163-2 ISBN13 978-0-8311-3163-0 I. Heating--Handbooks, manuals, etc. 2. Ventilation--Handbooks, manuals, etc. 3. Air conditioning-Handbooks, manuals, etc. 4. Buildings--Environmental engineering--Handbooks, manuals, etc. I. Kumar, Sunil. II. Hussain, Muhammed Iqbal. III. Title. TH7011.V46 2006 697--dc22 2006041837

Cover Photo: Image published with kind permission of CVRD and Bluhm Engineering.

INDUSTRIAL PRESS, INC. 989 Avenue of the Americas New York, New York 10018 -5410 1st Edition First Printing 10

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Printed and bound in the United States of America All rights reserved. This book or parts thereof may not be reproduced, stored in a retrieval system, or transmitted in any form without permission of the publishers.

PREFACE

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This Handbook provides comprehensive technical information in a modular form to heating, ventilating, and air conditioning (HVAC) designers and practitioners, namely engineers, architects, contractors, and plant engineers. It is also a handy reference for students mastering the intricacies of the HVAC rudiments. Each chapter is self-contained to the extent possible and emphasis is placed on graphical and tabular presentations of data that are useful for easy understanding of fundamentals and solving problems of design, installation, and operation. This Handbook draws upon the material presented in the Handbook of Air Conditioning, Heating, and Ventilating, Third Edition, Industrial Press, which forms the basis of the presentation. New topics and chapters have been introduced and previous information updated or rewritten. Examples using software solution tools have been added alongside traditional solutions using formulae from the handbook. The organization, however, remains, in the literal sense, a handbook. We gratefully acknowledge the contributors and editors of the aforementioned Handbook of Air Conditioning, Heating, and Ventilating, whose knowledge is embedded throughout the present book. We did not have the opportunity to meet any of them, but their written legacy has left an indelible imprint on the present work. An important source of information is the ASHRAE (American Society of Heating, Refrigerating, and Air-Conditioning Engineers) repertoire of publications. ASHRAE serves as the authoritative, and occasionally the sole, source of up-to-date HVAC related data and analysis. We acknowledge their permission to use material from various publications, especially the latest ASHRAE Handbook series. ASHRAE Publications 1791 Tullie Circle, NE Atlanta, GA 30329 Web Site: www.ashrae.org We also acknowledge three corporations for supplying us with material for inclusion in the Handbook. We profusely thank Mr. Michael White of Bell & Gossett (an ITT Division), Mr. Kent Silveria and Mr. Thomas Gorman of Trane Corporation, and Mr. Steven Boediarto of Preferred Utilities, for facilitating the acquisition of these materials. The Bell & Gossett corporation has graciously provided the ESP-PLUS software package to accompany the Handbook. This software, a $100 value, permits users to select components based on design or operating conditions. Bell & Gossett (ITT Fluid Handling) 8200 N. Austin Ave Morton Grove, IL 60053 Web Site: www.bellgossett.com The Trane corporation has generously allowed us to include their Trace Load 700 load calculation limited capability demonstration version software with the Handbook. Trane C.D.S. Department 3600 Pammel Creek Road La Crosse, WI 54601 Web Site: www.trane.com

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PREFACE

We are also grateful to the Preferred Utilities corporation for making available their publication on the topic of combustion analysis, and consenting to let us base our combustion chapter on it. Preferred Utilities Mfg. Corp 31-35 South Street Danbury, CT 06810 Web Site: www.preferred-mfg.com We acknowledge the input of our good friend, colleague, and HVAC critic, Mr. Naji Raad, whose experience in the profession provided a critical review of the manuscript. We thank our editors at Industrial Press, Mr. Christopher McCauley and Mr. Riccardo Heald, for their editorial input and suggestions, for reading the manuscript as it developed, and keeping the project on track; and Janet Romano for her cover design and production assistance. We acknowledge the effort of the many students at Polytechnic University who helped in researching for material, proofreading the manuscript, checking examples, and drawing figures. Those who deserve special recognition are Mr. Saurabh Shah and Mr. Christopher Bodenmiller for the graphics, Mr. Nayan Patel, Mr. Pranav Patel, and Mr. Prabodh Panindre for research, calculations, and proofing. Finally, we thank Kathleen McKenzie, freelance book editor, for her considerable contribution to this Handbook’s style, format and readability. Every effort has been made to prevent errors, but in a work of this scope it is inevitable that some may creep in. We request your forgiveness and will be grateful if you call any such errors to our attention by emailing them to [email protected]. Ali Vedavarz, Sunil Kumar, Muhammed Iqbal Hussain New York City December 2006

TABLE OF CONTENTS

1.

FUNDAMENTALS

5.

LOAD ESTIMATING FUNDAMENTALS

(Continued)

1–1 1–3 1–3 1–4 1–4 1–6 1–6 1–11 1–11 1–14 1–15 1–18 1–19

2. 2–1 2–1 2–1 2–2 2–2 2–2 2–2 2–3 2–4 2–4 2–4 2–5 2–6 2–7 2–12 2–18

3. 3–1 3–1 3–2 3–3 3–3 3–5 3–5

4. 4–1 4–1 4–5 4–6 4–8 4–10 4–10 4–10 4–10 4–10

5. 5–1 5–1 5–2 5–4 5–4 5–6 5–8 5–11

5–17 5–18 5–18 5–19 5–21 5–27 5–28 5–29

Fundamentals of Thermodynamics Conservation of Mass First Law of Thermodynamics Second Law, Reversibility, and Possible Processes Thermodynamic Cycles Fundamentals of Fluid Flow Flow in Pipes and Ducts Noise from Fluid Flow Fundamentals of Heat Transfer Overall Heat Transfer Fins and Extended Surfaces Some Details of Heat Exchange Augmentation of Heat Transfer

6. 6–1 6–1 6–2 6–2 6–3 6–3 6–4 6–4 6–6

PSYCHROMETRY Psychrometrics Ideal Gas Approximation Equation of State Humidity Ratio Relative Humidity Degree of Saturation Wet Bulb Temperature Partial Pressure of Water Vapor Dew Point Temperature Saturation Enthalpy Wet Bulb Temperature Properties of Moist Air Psychrometric Chart Presentation Thermodynamic Properties of Water at Saturation Thermodynamic Properties of Moist Air

7. 7–1 7–1 7–2 7–3 7–6 7–6 7–6 7–6 7–7 7–8 7–9 7–11 7–27 7–28 7–31 7–31 7–31 7–35 7–36 7–44 7–46 7–47 7–49 7–50 7–50 7–50

AIR CONDITIONING PROCESSES Introduction Heating and Cooling Process Cooling with Dehumidification Heating with Humidification Adiabatic Mixing of Two Air Streams Evaporative Cooling Heating and Air Conditioning System Cycles

INDOOR AIR QUALITY AND VENTILATION Indoor Air Quality Ventilation Procedure Concentration of Air Pollutants Indoor Air Quality Procedure Filters Hepa Filters Carbon Media Filters Fiber and Foam Filters Ozone Ultraviolet Light

8. 8–1 8–1 8–2 8–7 8–7 8–8 8–8 8–9 8–9 8–12 8–13 8–13 8–14 8–14 8–15 8–15 8–17

LOAD ESTIMATING FUNDAMENTALS Conduction Thermal Conductivities of Materials Convection Thermal Radiation Emissivities of Some Materials Overall Heat Transfer Coefficient Parallel Arrangement Coefficient of Transmission

iii

Relative Thermal Resistances of Building Materials Surface Conductances and Resistances Emittance Values of Various Surafces Thermal Resistances of Plane Airspaces Thermal Properties of Building and Insulating Materials Coefficients of Heat Transmission of Various Fenestrations Transmission Coefficients for Wood and Steel Doors Outdoor Air Load Components

HEATING LOAD CALCULATIONS Introduction Calculating Design Heating Loads Heat loss Through Walls, Roofs, and Glass Area Heat Loss from Walls below Grade Below-Grade Wall U-Factors Heat Loss from Basement Floor Below Grade Heat Loss Coefficients Heat Loss from Floor Slab On Grade Ventilation and Infiltration Heat Loss

COOLING LOAD CALCULATIONS Transfer Function Method (TFM) Heat Source in Conditioned Space Heat Gain from Occupants Heat Gain from Cooking Appliances Heat Gain from Medical Equipments Heat Gain from Computer Heat Gain from Office Equipments CLTD/SCL/CLF Calculation Procedure Cooling Load by CLTD/SCL/CLF Method Roof Numbers CLTD for Roofs CLTD for Walls Code Number for Wall and Roof Wall Types CLTD for Glass Zone Types for CLF Tables Zone Types for SCL and CLF Tables Residential Cooling Load Procedure SCL for Glass CLF for People and Unhooded Equipments CLF for Hooded Equipments Window GLF for Residences CLTD for Residences SC for Windows SLF for Windows Air Exchange Rates

DUCT DESIGN Introduction Pressure Head and Energy Equation Friction Loss Analysis Dynamic Losses Ductwork Sectional Losses Fan System Interface Pressure Changes System Duct System Design Design Considerations Duct Design Methods Duct Design Procedures Automated Duct Design Duct Fitting Friction Loss Example Equal Friction Method Example Resistance in Low Pressure Duct System Example Static Regain Method Example Fitting Loss Coefficients

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9.

TABLE OF CONTENTS

PIPE SIZING

10.

HYDRONIC HEATING AND COOLING SYSTEM

(Continued)

9–1 9–1 9–3 9–3 9–3 9–3 9–3 9–4 9–4 9–6 9–28 9–29 9–29 9–31 9–31 9–32 9–32 9–59 9–59 9–61 9–72 9–72 9–78 9–80 9–80 9–91 9–91 9–93 9–93 9–96 9–97 9–97 9–97 9–132 9–134 9–134 9–136 9–136 9–139 9–139 9–141 9–142 9–142 9–142 9–142 9–143 9–143 9–143 9–144 9–144 9–144 9–144 9–144 9–145 9–157 9–158

10. 10–1 10–4 10–4 10–4 10–4 10–4 10–6 10–6 10–7

Pressure Drop Equations Valve and Fitting Losses Water Piping Flow Rate Limitations Noise Generation Erosion Allowances for Aging Water Hammer Hydronic System Piping Valve and Fitting Pressure Drop Service Water Piping Plastic Pipe Cold Water Pipe Sizing Steam Flow in Pipes Steam Flow Formulas Vertical Pipes Steam Piping Gas Piping For Buildings Residential Piping Commercial-Industrial Piping Compressed Air Systems Compressed Air Viscosity of Liquids Piping Types of Materials Plastics Pipe Joining Techniques Standards for Specification and Identification Design Parameters Installation Codes and Regulations Pipe Fittings Taper Pipe Thread Laying Lengths of Pipe with Screwed Fittings Allowable Spaces for Pipes Expansion of Pipe Corrosion Resistance Pipe Support Spacing Gate, Globe, and Check Valves Operation Maintenance Methods Formulas for Sizing Control Valves To Determine Valve Size To Determine Valve Capacity For Vapors Other Than Steam Identification of Piping Systems Dangerous Materials Fire Protection Materials and Equipment Safe Materials Protective Materials Method of Identification Heat Losses in Piping Heat Losses from Bare Pipe Heat Losses from Steam Piping Heat Loss from Insulated Pipe Cold Surface Temperature

HYDRONIC HEATING AND COOLING SYSTEM Basic System Temperature Classifications Closed Hydronic System Components Design Convectors or Terminal Units Boiler Air Eliminations Methods Pressure Increase Due to Change in Temperature Expansion Tank Expansion Tank Sizing

10–8 Characteristics of Centrifugal Pumps 10–8 Operating Characteristics 10–9 Pump Laws 10–9 Change of Performance 10–10 Centrifugal Pump Selection 10–10 Total Dynamic Head 10–11 Net Positive Suction Head (NPSH) 10–11 Pumping System 10–16 Parallel Pumping 10–17 Series Pumping 10–18 Design Procedures 10–18 Preliminary Equipment Layout 10–19 Final Pipe Sizing and Pressure Drop Determination 10–19 Final Pressure Drop 10–19 Final Pump Selection 10–19 Freeze Prevention

11. 11–1 11–1 11–2 11–4 11–5 11–7 11–7 11–7 11–8 11–9

12.

ENERGY CALCULATION Degree Day 65°F as the Base Application of Degree Days Predicting Fuel Consumption Predicting Future Needs Empirical Constants Load Factor and Operating Hours Limitations Degree-Days Abroad Degree Days for Various US Locations

COMBUSTION

12–1 Combustion Basics 12–3 Efficiency Calculations 12–7 Saving Fuel with Combustion Controls 12–11 Combustion Considerations 12–11 Pressure and Flow Basic Principles 12–12 Atomizing Media Considerations 12–12 Combustion Air Considerations 12–13 Flue Gas Considerations 12–14 Gas Fuel Firing Considerations 12–14 Fuel Oil Firing Considerations 12–15 Operational Rules of Thumb 12–16 Common Application 12–20 Combustion Control Strategies 12–20 Control System Errors 12–20 Combustion Control Strategies 12–21 Parallel Positioning Systems 12–22 Fully Metered Control 12–23 Feedwater Control Systems 12–24 Draft Control 12–26 Oxygen Trim 12–27 Combustion Air Flow Control Techniques 12–28 Flue Gas Recirculation (FGR) 12–33 Fuel Oil Handling System Design 12–33 Determination of Required Flow Rate 12–34 Stand by Generator Loop Systems 12–34 Multiple Pumps 12–34 Burner Loop Systems 12–36 Maximum Inlet Suction 12–37 Pump Discharge Pressure 12–37 Piping System Design 12–37 Pump Set Control System Strategies

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TABLE OF CONTENTS

13.

AIR CONDITIONING SYSTEMS

13.

AIR CONDITIONING SYSTEMS

(Continued)

13–1 13–1 13–5 13–7 13–8 13–10 13–14 13–14 13–17 13–18 13–20 13–21 13–22 13–22 13–23 13–26 13–27 13–30 13–31 13–31 13–36 13–36 13–36 13–37 13–37 13–38 13–39 13–40 13–40 13–41 13–41 13–42 13–42 13–43 13–43 13–44 13–44 13–45 13–47 13–47 13–47 13–48 13–48 13–49 13–54 13–54 13–54 13–60 13–60 13–60 13–65 13–65 13–66 13–66 13–67 13–68 13–68 13–69 13–69 13–70 13–73 13–76 13–77 13–80 13–81 13–81 13–84 13–88 13–88 13–88

Air Conditioning Systems Single Package Units Single Package Installations Installation of Split Systems Zoning Unitary Installations Selection Procedure Evaporative Air Conditioning Permissible Air Motion Variable Volume AC System Initial Costs Cooling Considerations Overlapping Heat Recovery Heating Cooling Systems Air Systems Controls Air Water Systems Sources of Internal Heat Heat from Service Refrigeration Exhaust Air Heat Recovery Systems Heat Pumps Reverse-Cycle Principle Coefficient of Performance Heating Season Performance Factor Types of Heat Pumps Air-to-Air Heat Pumps Water-to-Water Heat Pumps Water-to-Air Heat Pumps Air-to-Water Heat Pumps Ground Source Heat Pumps Special Heat Sources Operating and Installation Factors Outdoor Temperature Effects Thermostats Heat Anticipators Equipment Arrangement Electrohydronic Heat Recovery Cooling Cycle System Design Supplementary Heat Optimized Data for Heat Pump Development of Equations Development of Tables Selecting Air Handling Units Well Water Air Conditioning Heat Pump/Solar Energy Application System Description and Operation High Velocity Dual Duct Systems Advantages and Disadvantages Dual Duct Cycles Duct Sizing Technique Large vs. Small Ducts Design Velocity Maximum Velocity Sizing High Pressure Ducts Return Air Ducts Low Pressure Ductwork Basic Arrangement Zoning Ceiling Plenum Modular Type Office Buildings Constant Volume Mixing Units Apparatus Floor Area Construction Details Automatic Control Applications Rooftop Multizone Units Multizone Unit Control Damper Control Economizer Control Cycle Unit Ventilator Control

13–91 13–94 13–95 13–95 13–97 13–97 13–99 13–102 13–104 13–105 13–105 13–106 13–107 13–107 13–107 13–107 13–107 13–107 13–108 13–108 13–108 13–108 13–108 13–108 13–108 13–108 13–110 13–111 13–111 13–111 13–112 13–112 13–112 13–112 13–113 13–113 13–113 13–114 13–114 13–114

14.

Hot Water System Control Mixing Box Control Rotary Air-to-Air Heat Exchanger Control Automatic Control for Dual Duct System Winterizing Chilled Water System Water Circulation to Prevent Freeze-Up Mechanical Draft Cooling Towers Atmospheric Cooling Towers Quantity of Cooling Water Required Roof is a Location for AC Equipment Advantages of Roof Disadvantages of Roof Servicing Cooling Plant Servicing Cooling Plant for Summer Use Water System Air Handling System Compressor Oil Condenser Refrigeration Unit Check Oil Compressor Air Conditioning Equipment Maintenance Air Handling Equipment Air Distribution Equipment Water-Using Equipment Cooling Equipment Air Conditioning Maintenance Schedule Unit Air Conditioners Central Systems Condensing Water Circuit Cooling Water System Filters and Ducts Air Conditioning Maintenance Procedure Refrigerant Circuit and Controls Condensing Water Circuit Cooling Water System Filters and Ducts Rotating Apparatus Unit Air Conditioners Checklist for Air Conditioning Surveys

AIR HANDLING AND VENTILATION

14–1 Terminology, Abbreviations, and Definitions 14–3 Fan Laws 14–11 Fan Performance Curves 14–16 Class Limits for Fans 14–21 Fan Selection 14–26 Fan Inlet Connections 14–27 Fan Discharge Conditions 14–31 Useful Fan Formulas 14–32 Nomographs for Fan Horsepower 14–32 Monographs for Fan Horsepower and Actual Capacity 14–34 Fan Selection Questionnaire 14–37 Air Flow in Ducts 14–40 Pitot Traverse 14–40 Friction Losses 14–40 Correction for Roughness 14–40 Rectangular Duct 14–52 Air Balancing and Air Turning Hardware 14–56 Air Distribution 14–56 Fire Dampers and Fire Protection 14–56 Duct System Design 14–59 High Velocity System Design 14–68 Step by Step Design 14–68 Main Duct 14–70 Branch Trunk Ducts 14–71 Single Branch Lines 14–72 Duct Design by Computer 14–73 Fibrous Glass Duct Construction

vi

14.

TABLE OF CONTENTS

AIR HANDLING AND VENTILATION

(Continued)

15–1 15–1 15–1 15–3 15–3 15–3 15–4 15–4 15–4 15–26 15–30 15–39 15–40 15–40 15–41 15–41 15–46 15–49 15–49 15–49 15–49 15–49 15–49 15–49 15–49 15–50 15–50 15–50 15–50 15–50 15–52 15–52 15–52 15–52 15–53 15–53 15–53 15–53 15–53

STEAM HEATING SYSTEM DESIGN

(Continued)

14–75 Determining Required Air Volume 14–75 Estimating Weight of Metal 14–77 Apparatus Casing Construction 14–77 Condensate Drains for Air Conditioning Units 14–78 Air Filters and Dust Collectors 14–78 Air Filters 14–79 Dust Collectors 14–82 Dry Centrifugal Collectors 14–82 Wet Collectors 14–82 Fabric Collectors 14–83 Electrostatic Precipitators 14–83 Breeching Design and Construction 14–83 Expansion 14–84 Aerodynamics 14–85 Access 14–85 Round Breeching Construction 14–85 Rectangular Breeching Construction 14–90 Chimney Draft and Velocities 14–92 Forced Draft and Draft Control 14–94 Sizing of Large Chimneys 14–95 Chimney Design and Construction 14–96 Balancing Small Air Conditioning Systems 14–97 Balancing Medium and Large Systems 14–98 Balancing Duct Distribution 14–98 Balancing Systems Using Booster Fans 14–99 Air Balancing by Balancing and Testing Engineers

15.

15.

STEAM HEATING SYSTEM DESIGN Large Systems Equivalent Direct Radiation Piping Connections to Boilers Direct Return Connection Common Return Header Two Boilers with Common Return Header and Hartford Connection Two Boilers with Separate Direct Return Connections from Below Separate Direct Return Connections Connections to Steam Using Equipment Piping Application Industrial and Commercial Steam Requirements Flash Steam Calculations Sizing of Vertical Flash Tanks To Size Flash Tank To Size Float Trap Airbinding Estimating Friction in Hot Water Piping Hot Water Heating Systems Service Water Heating Operating Water Temperature Air Removal from System Water Flow Velocity Prevention of Freezing Water Circulation below Mains Limitation of Pressure System Adaptability Use of Waste Steam Heat Heat from District Steam System Summer Cooling Types of Water Heating Systems Design Recommendations for Hot Water Systems Water Velocity Pump Location Air Venting Balancing Circuits Filling Pressure Preventing Backflow Connecting Returns to Boiler Locating the Circulating Pump

15–53 15–54 15–54 15–55 15–58 15–58 15–58 15–59 15–59 15–59 15–59 15–60 15–60 15–60 15–60 15–61 15–61 15–63 15–63 15–63 15–63 15–63 15–64 15–64 15–64 15–65 15–65 15–65 15–65 15–66 15–66 15–66 15–67 15–67 15–67 15–68 15–68 15–68 15–68 15–69 15–69 15–69 15–69 15–69 15–70 15–70 15–70 15–71 15–73 15–73 15–74 15–76 15–81 15–81 15–81 15–81 15–81 15–82 15–82 15–83 15–85 15–85 15–86 15–87 15–87 15–87 15–88 15–88 15–88 15–89 15–90

Sizing the Expansion Tank Compressed Air to Reduce Tank Size Piping Details Design of Piping Systems Design of Two Pipe Reversed Return System Final Check of Pipe Sizes Design of Two Pipe Direct Return System Piping for One-Pipe Diversion System Sizing Piping for Main Sizing Piping for Branches Pipe Size Check Piping for One-pipe Series System Combination of Piping Systems Sizing Hot Water Expansion Tanks Conditions Affecting Design Sizing Hot Water Expansion Tanks High Temperature Water Systems High Temperature Drop Heat Storage Limitation of Corrosion Pressurization of HTW System Steam Pressurization Gas Pressurization Air Pressurization Nitrogen Pressurization Expansion Tanks Expansion Conditions Determining Expansion Tank Size Location of Steam Pressurizing Tank Nitrogen Pressurizing Tanks Application of HTW for Process Steam Circulating Pumps Pumps for HTW Systems Manufacturer’s Information Pump Specifications Net Positive Suction Head Effect of Cavitation Within Pump Pump Construction for HTW Systems Circulating Pump Seals Boiler Recirculating Pump Boilers for HTW Systems Boiler Emergency Protection Pipe, Valves, and Fittings for HTW Systems Valve Installation Welded Joints Venting of Piping Effect of Load Variation on Operation Pipe Sizing for HTW Systems Ratings of Steel Boilers Ratings Ratings for Steel Boilers Stack Dimensions Heating and Cooling Media Brine Glycerine Glycol Other Media Warm Air Heating Early Types Current Types Furnace Performance Testing and Rating of Furnaces Acceptable Limits Selection of Furnace Rule for Selection Blower Characteristics Blower Sizes Duct System Characteristics Trends Warm Air Registers Return Air Intakes

vii

TABLE OF CONTENTS

15.

STEAM HEATING SYSTEM DESIGN

(Continued)

15–91 15–94 15–94 15–95 15–95 15–109 15–109 15–112 15–117 15–124

16. 16–1 16–1 16–2 16–2 16–2 16–7 16–7 16–7 16–9 16–13 16–13 16–13 16–14 16–14 16–14 16–15 16–17 16–20 16–23 16–23 16–24 16–24 16–24 16–25 16–26 16–27 16–27 16–31 16–31 16–32 16–36 16–36 16–37 16–37 16–37 16–38 16–39 16–39 16–39 16–39 16–39 16–40 16–40 16–40 16–40 16–42 16–42 16–42 16–42 16–43 16–43 16–43 16–43 16–43 16–44 16–44 16–44

16.

NOISE AND VIBRATION CONTROL

(Continued)

Arrangement of Furnace and Ducts Basic Thermostatic Controls Continuous Air Circulation Continuous Blower Operation Intermittent Blower Operation Steam Supplied Unit Heater Gas Fired Radiant Heaters Sizing of Steam Traps Unit Heaters Checklist for Heating System Servicing

NOISE AND VIBRATION CONTROL Noise and Vibration Definitions and Terminology Noise Criteria Speech Interference Criteria Sound Levels of Sources Ratings and Standards Airborne Sound Transmission Vibration Isolation Isolation Mount Selection Airborne Noise Through Ducts Regenerated Noise Other Mechanical Noise Sources Calculation of Sound Levels from HVAC Systems Description of Decibels Addition of Decibels The Sabin Determination of Sound Pressure Level Noise in Ducted Systems Fan Noise Generation Estimating Fan Noise Distribution of Sound Power at Branch Takeoffs Attenuation of Untreated Duct Duct Lining Attenuation Sound Attenuation of Plenums Duct Lining and Elbows Open End Reflection Loss Air Flow Noise Flow Noise Generation of Silencers Sound Transmission Through Duct Walls Calculation of Sound Levels in Ducted Systems Control of Cooling Tower Noise Fan Noise Water Noise Drive Components External Noise Sources Configuration Factors Location Reducing Sound Generated Half-Speed Operation Oversizing of the Tower Changing Leaving Conditions Sound Absorbers Obtaining Desired Sound Levels Acoustical Problems in High Velocity Air Distribution System Noise Air Handling Apparatus Rooms Selection of Fan Isolation Bases Apparatus Casings Dampers and Air Valves Flexible Connectors Air Distributing Systems Duct Velocities Choice of Duct Design Method Ductwork Adjacent to Apparatus Room Duct Connections to Apparatus Casings Type Duct Construction Fittings for High Velocity Ductwork

16–45 16–45 16–46 16–46 16–46 16–46 16–47 16–47 16–48 16–49

17. 17–1 17–1 17–2 17–2 17–3 17–3 17–6 17–6 17–8 17–8 17–9 17–10 17–11 17–12 17–12 17–12 17–15 17–15 17–16 17–17 17–18 17–18 17–18 17–19 17–19 17–20 17–20 17–20 17–21 17–21 17–21 17–23 17–24 17–24 17–25 17–25 17–29 17–29 17–29 17–30 17–31 17–33 17–33 17–33 17–33 17–36 17–36 17–37 17–39 17–39

Take-off Fittings Dual Duct Area Ratio Dampers as a Noise Generating Source Sound Barrier for High Velocity Ductwork Sound Traps Cross Over of Horizontal Dual Duct Mains Testing of High Pressure Ductwork Terminal Devices Radiation Protection at Wall Openings for Duct or Pipe Medical Installations

MOTORS AND STARTERS NEMA Motor Classifications Locked Rotor Torque Classification of Single-Phase, Induction Motors by Design Letter Torque, Speed, and Horsepower Ratings for Single-Phase Induction Motors Classification by Environmental Protection and Method of Cooling Standard Voltages and Frequencies for Motors The National Electrical Code Grounding Motor and Load Dynamics, and Motor Heating Torque Speed Relationships Torque, Inertia, and Acceleration Time Dynamics of the Motor and the Load Motor Heating and Motor Life Rotor Heating During Starting Single Phase Motors Types of Motors Repulsion-Induction Large Single-Phase Motors Application Loading Motor Protection Motor Selection Analysis of Application Polyphase Motors Enclosure Bearings Quietness Polyphase, Squirrel Cage Induction Motors Speed Control Two-Speed Polyphase, Squirrel Cage Induction Motors Two Speed Motors Come in Two Types Wound-Rotor Polyphase Induction Motors Variable Speed Synchronous Motors Hermetic Type Motor Compressors Hermetic Compressors to 5 hp Starters Motor Controllers Overcurrent Protection Overload Protection Starters for Large AC Motors Winding and Reduced-voltage Starting Electric Utility Limitations Minimizing Mechanical Shocks Application Types of Starters Open Circuit Transition Advantages and Disadvantages Useful Formulas Electric Motor Maintenance

viii

18.

TABLE OF CONTENTS

DESIGN PROCEDURE, ABBREVIATIONS, SYMBOLS

20.

UNITS AND CONVERSIONS

(Continued)

18–1 18–1 18–1 18–1 18–5 18–6 18–7 18–8 18–9

19.

Design Procedure Contract and Mechanical Drawings HVAC Drawings Floor Plans Valve Symbols Piping Symbols Pipe Fittings Symbols Abbreviations for Scientific and Engineering Terms Lists of Abbreviations and Symbols

CLIMATIC DESIGN INFORMATION

19–1 Climatic Design Conditions 19–1 Applicability and Characteristics of the Design Conditions 19–27 Dry Bulb and Wet Bulb Temperature for US Locations

20. 20–1 20–1 20–1 20–1 20–1 20–1 20–1 20–2 20–2 20–2 20–2 20–2 20–2 20–2 20–2 20–3

UNITS AND CONVERSIONS U.S. Customary Unit System Linear Measures Surveyor's Measure Nautical Measure Square Measure Cubic Measure Shipping Measure Dry Measure Liquid Measure Old Liquid Measure Apothecaries' Fluid Measure Avoirdupois or Commercial Weight Troy Weight, Used for Weighing Gold and Silver Apothecaries' Weight Measures of Pressure Miscellaneous

20–4 U.S. System And Metric System Conversion 20–4 Length and Area 20–4 Mass and Density 20–5 Volume and Flow 20–6 Force, Energy, Work, Torque and Power Conversion 20–7 Velocity and Acceleration 20–8 Metric Systems Of Measurement 20–8 Measures of Length 20–8 Square Measure 20–8 Surveyors Square Measure 20–8 Cubic Measure 20–8 Dry and Liquid Measure 20–8 Measures of Weight 20–10 Binary Multiples 20–10 Terminology of Sheet Metal

21.

INDEX

FUNDAMENTALS OF THERMODYNAMICS actual quantity of the material under consideration, such as volume. The thermodynamic state of a substance is defined by listing its intrinsic properties. The most common intrinsic thermodynamic properties are temperature T, pressure p, specific volume v (which is the inverse of density ρ), specific entropy s, specific enthalpy h, and specific internal energy u. It has been established that a thermodynamic state of a substance can be uniquely identified by two independent intrinsic properties. For example, temperature and pressure are such properties, except in the saturation region where the same temperature and pressure can have an infinite number of states corresponding to quality from 0 to 100%. Enthalpy (h) is defined by

FUNDAMENTALS This chapter covers the fundamentals of thermodynamics, fluid mechanics, and heat transfer as related to the theory and practice of air conditioning, heating and ventilation. Basic concepts needed for the HVAC professional are presented, while advanced topics are not considered. Excellent resources, such as the ASHRAE Handbook series and many great textbooks in the areas of thermal fluid sciences, treat the details and advanced topics that have been omitted from this focused chapter. Fundamentals of Thermodynamics Thermodynamics is the study of energy, its transformations, and its relation to states of matter or substance. Thermodynamics deals with equilibrium conditions that are typical of steady state, and any changes are considered to be quasi-equilibrium processes where change occurs slowly and incrementally so as to allow each incremental intermediate state to reach equilibrium before it advances further.

h = u + pv

(1)

where u =internal energy per unit mass. Each property in a given state has only one definite value, and any property always has the same value for a given state, regardless of how the substance arrived at that state. The thermodynamic property entropy s measures the molecular disorder of a system. The more mixed a system, the greater its entropy; conversely, an orderly or unmixed configuration is one of low entropy. Figs. 1-1a and 1-1b schematically show the liquid and vapor states of water using two properties to uniquely identify the states. Fig. 1-1a shows the states in (p, h) coordinates and a line showing the states along a constant temperature is also presented. Fig. 1-1b shows the same information in (T, s) coordinates and a constant pressure line is shown. Within the dome where the saturated states exist it is possible to increase the specific volume (and other intrinsic, per unit mass based properties) even when keeping the pressure and temperature constant (for example line B-AC in Figs. 1-1a and 1-1b).

A pure substance has a homogeneous and invariable chemical composition. It can exist in more than one phase, but chemical composition is the same in all phases. If a substance exists as liquid at saturation temperature and pressure, it is called a saturated liquid. If the temperature of the liquid is lower than the saturation temperature for the existing pressure, it is called either a subcooled liquid (the temperature is lower than the saturation temperature for the given pressure) or a compressed liquid (the pressure is greater than the saturation pressure for the given temperature). When a substance exists as part liquid and part vapor at the saturation temperature, its quality is defined as the ratio of the mass of vapor to the total mass. Quality has meaning only when the substance is in a saturated state; i.e., at saturation pressure and temperature. If a substance exists as vapor at the saturation temperature, it is called saturated vapor. The term dry saturated vapor is used to emphasize that the quality of the substance is 100%. When the vapor is at a temperature greater than the saturation temperature, it is superheated vapor. The pressure and temperature of superheated vapor are independent properties, since the temperature can increase while the pressure remains constant. Gases are highly superheated vapors. A property of a substance is any observable characteristic of the material. An intrinsic property of the material is one that does not depend on the shape or volume or mass of the material. Such properties are pressure and temperature, or properties that are usually expressed in the units of per unit mass, such as specific volume. This is in contrast with extrinsic properties which depend on the

Fig 1-1a. Thermodynamic states of water in p-h coordinates

1–1

1–2

FUNDAMENTALS OF THERMODYNAMICS

Fig 1-1b. Thermodynamic states of water in T-s coordinates

The specific heats at constant pressure Cp and volume Cv are defined as the heat added to a unit mass at constant pressure or volume, respectively, to cause temperature of the mass to increase by one unit. Mathematically, by using the first law for closed systems and the definition of work (described later in the chapter), these are defined as ∂u C v = ------ ∂T v

∂h C p = ------ ∂T p

(2)

For ideal gases the following holds true rigorously du = C v dT

dh = C p dT

(3)

and for real gases this is a good engineering approximation. For incompressible substances the two specific heats are the same and the above expressions also holds. Thus, liquids and solids, which are treated as incompressible, are described by Equation (3). The basic entity analyzed by thermodynamics is termed a thermodynamic system. A thermodynamic system is an identifiable or specific region in space or an identifiable or specific quantity of matter. It is bounded by the system surface or the system boundaries. The surroundings include everything external to the system, and the system is separated from the surroundings by the system boundaries. These boundaries can be movable or fixed, real or imaginary. There are two types of systems: closed and open. In a closed system, no mass enters or leaves the system. A closed system is thus a fixed mass system. In an open system, mass may enter or leave the system and therefore the boundaries of such a system are imaginary in some parts where the mass is exchanged with the surroundings. Energy can always cross the system boundaries, whether open or closed, except for an isolated closed system, which is one that does not allow energy to cross its boundaries. Almost all HVAC systems that will be studied here fall in the open system category.

Energy has the capacity for producing an effect on the system and is described differently depending on whether it is stored or is crossing the system boundary. Thermal (internal) energy is the energy possessed by the matter in a system caused by the internal motion of the molecules, the intermolecular forces, and other microscopic mechanisms of energy storage within the material. It is impossible to find the absolute (total) value of internal energy of a material, and it is therefore measured with reference to a standard state in which the value of the internal energy is arbitrarily set to zero. The internal energy is the sum of all these microscopic energies such that when the internal energy of an isolated closed system (fixed mass) changes, the result is a change in the temperature of the system. Internal energy is an intrinsic property of material and its value per unit mass can be used to uniquely define the state of the material. The specific internal energy u is the internal energy per unit mass. Potential energy is the energy possessed by a system due to the elevation of the system: PE = mgz

(4)

where m =mass; g =local acceleration of gravity, and; z =elevation above horizontal reference plane. Potential energy is not an intrinsic property of material. Kinetic energy is the energy possessed due to the bulk velocity of the flowing material and is expressed as 2 1 (5) KE = --- mV 2 where V =velocity of a fluid stream. Kinetic energy is not an intrinsic property of material. The types of energy that are defined only when they cross the system boundaries are heat and work. Heat (Q) and work (W) cannot be stored and they only exist when energy is being transferred across a system boundary. Heat is the mechanism that transfers energy across the boundary of systems solely through temperature difference. Heat always flows from higher to lower temperatures. Heat is considered positive by convention when energy is added to a system (see Fig. 1-2). Work is the mechanism that transfers energy across the boundary of systems through forces and movement. (Here it has the traditional characterization: it is the product of force and distance.) If the total effect produced in the system can be reduced to the raising of a weight, then nothing but work has crossed the boundary. By convention work is considered positive when energy is removed from a system, i.e., the system does work on its surroundings (see Fig. 1-2). For example, when a system comprising a cylinder and movable piston expands so that the piston is moving outwards, work is done by the system on the surroundings and is therefore considered positive in any analysis of system. Mechanical or shaft work is the work associated with a rotating shaft such as a turbine, air compressor, or inter-

1–3

FUNDAMENTALS OF THERMODYNAMICS

nal combustion engine. Flow work is energy associated with the movement of fluid in conjunction with forces of fluid pressure as it crosses the boundary of an open system. It can be more easily understood as the work done by the fluid to push itself against the other fluid particles as it forces itself to enter or exit the system. Work done by entering fluid streams is negative (pushing into the system), while work done by fluid streams exiting the system is positive (pushing the surroundings). The magnitude of flow work per unit mass is determined by the expression Flow work = pv (6) where p =pressure and; v =specific volume, or the inverse of density.

Fig 1-2. Energy flow in general thermodynamic system

A property of a system is any observable characteristic of the system. A process is a change in state that can be defined as any change in the properties of a system. A process is described by its initial and final equilibrium states, its path, and the interactions that take place across system boundaries as it goes forth. A cycle is a process or a series of processes wherein the initial and final states of the system are identical. Therefore, at the conclusion of a cycle, all properties have the same value they had at the beginning. Conservation of Mass.—Conservation of mass in a closed system is the default since no mass leaves or enters the system. For an open system the conservation of mass indicates that the difference between the mass entering and leaving is equal to the increase of mass in the system. For the general case of multiple flow streams the conservation of mass is written as

∑ min – ∑ m out

= [ m f – m i ] system

(7)

For steady flow processes the above can be modified as

∑ min = ∑ mout

(8)

For flow through a pipe or duct the mass flow rate is related to the velocity by VA m· = ρVA = ------(9) v

where ρ =density of the flowing fluid; A =cross-sectional area through which the fluid is flowing; V =velocity; v =specific volume, or the inverse of the density. i =indicate initial states; and f =indicate final states. It is assumed that the velocity is uniform across the crosssection. If it is not, then V in the above mass flow rate definition is average velocity. A is normal to the direction of the fluid flow. First Law of Thermodynamics.—The first law of thermodynamics is often called the law of the conservation of energy. After a system is defined, the conservation of energy states that energy in – energy out = increase in energy (10) Fig. 1-2 illustrates energy flows into and out of a thermodynamic system. For a closed system this is written as (11) Q – W = Uf – Ui = m ( uf – ui ) where m =mass of the system; U =internal energy of systems; and m =internal energy per unit mass; For the general case of an open system with multiple mass flows in and out of the system, the energy balance can be written 2

∑ min u + pv + -----2- + gz in – V

2

∑ m out u + pv + -----2- + gz out + Q – W V

2

=

(12)

2

V V m f u + ------ + gz – m i u + ------ + gz f i system 2 2 The steady flow process is important in engineering applications. Steady flow signifies that no quantities associated with the system vary with time. Consequently,

∑

2

V - + gz – m· in u + pv + ---- in 2

(13) 2 V · · m· out u + pv + ------ + gz + Q – W = 0 out 2 Here the addition of the dot on top of the variables m, Q, and W indicates the time derivative (d/dt) which yields the rate of mass flow, the rate of heat flow, and the rate of work done, respectively. In the above two equations the flow work of the entering and exiting streams is indicated by the pv terms, and, therefore, the W in the equations is the work done by moving system boundaries plus shaft work, and any other work not described by these categories. In the above energy conservation (Equations (12)

∑

1–4

FUNDAMENTALS OF THERMODYNAMICS

and (13)) assume that velocity is uniform across the cross section. However, in real fluid systems it is not uniform due to viscosity of the fluid. The change to the kinetic energy term is discussed in the section Fundamentals of Fluid Flow. Second Law, Reversibility, and Possible Processes.—The second law of thermodynamics can be expressed in many ways. Here it is being introduced to distinguish and quantify processes that can only proceed in one direction (irreversible) from those that are reversible. It also indicates which processes cannot exist. The second law for a closed system is written as f δQ S f – S i ≥ δQ dS ≥ ------------(14) i T T where the subscripts f and i indicate final and initial state, respectively, and the system entropy S = ms, where m is the mass of the system and s is specific entropy. The temperature T must be in absolute units (Kelvin or Rankine). The equality sign holds for reversible processes and the inequality for irreversible processes. The above form of the second law shows that the entropy of the system either increases or remains same, assuming the heat flow is positive. For adiabatic processes, that is where Q = 0, the right side of either Equation (14) vanishes, clearly indicating that entropy increases if the process is not reversible. For constant temperature processes (15) T ( Sf – Si ) ≥ Q

∫

where the equality holds for reversible processes. For an open system the above can be modified as f ------- (16) ( S f – S i ) system + ( ms ) out – ( ms ) in ≥ δQ i T It is assumed in Equation (16) that the inlet and outlet properties remain invariant with time. For steady flow systems with invariant inlet and outlet properties the second law can be rewritten in the form Q·( m· s ) out – ( m· s ) in ≥ --(17) T The open system of Equation (17) can give insights about reversible and irreversible processes if we consider a single input, single output stream process. If the process is adiabatic, then the inlet and outlet entropies will be the same for reversible processes; otherwise the outlet entropy will be greater than the inlet. If the process is isothermal, then the difference between the rate of entropy flowing out and in will be equal to the heat addition rate divided by temperature only if the process is reversible. Adiabatic reversible processes are also termed isentropic because entropy remains same. On a thermodynamic chart where entropy is the horizontal axis, the adiabatic reversible processes will be vertical lines, but the irreversible processes will always veer right (towards increasing entropy) from their starting point. Thus, for any given

∑

∑

∑

∑

∫

∫

starting point, the only thermodynamically possible adiabatic processes are those that end in the half plane to the right of the vertical line drawn from the starting point in the thermodynamic chart where specific entropy is on the horizontal axis. Thus, the second law can show processes that are reversible, irreversible and possible, or impossible. The second law for cycles is described next. It further refines the concept of possible and impossible series of processes occurring in a cycle. Thermodynamic Cycles.—Thermodynamic cycles that make it possible to remove heat from cold spaces and dump the heat in hot ambient spaces are discussed here. These cycles make it possible for air conditioning and refrigeration systems to exist and they form the basis of HVAC engineering. The performance of a refrigeration or air conditioning thermodynamic cycle is usually described by a coefficient of performance. COP is defined as the benefit of the cycle (amount of heat removed) divided by the required energy input to operate the cycle, or useful refrigerating effect (18) COP = -----------------------------------------------------------------energy from external source The first law for cycles indicates that

°∫ δQ = °∫ δW Q net = W net

(19)

Q· net = W· net where Q =heat added and; W =work done. For the individual processes that make up the cycle, the first law for open systems is used. Since cyclic devices have only one working fluid and the mass flow rate of the fluid is the same through each process due to conservation of mass, the open system Equation (13) can be modified as below for each process in the cycle: (20) m· ( h – h ) + Q· – W· = 0 out

in

The kinetic and potential energy terms are usually neglected in comparison to enthalpy in refrigeration and air conditioning thermodynamic cycles because their magnitudes are usually several orders smaller than enthalpy. The Carnot cycle is an ideal cycle that is made up of completely reversible processes and operates between two fixed temperatures. This cycle is useful since it is a thermodynamic ideal for refrigeration cycles that need to operate between two temperatures: the temperature of the conditioned space and the temperature of the external (hot) ambient. Also heat is transferred from the cold space to the ambient. The properties of the Carnot cycle are that it can be used as a refrigerator or a heat pump, as well as a

FUNDAMENTALS OF THERMODYNAMICS

work-producing device, depending on the input quantities. Fig. 1-3 shows the Carnot cycle on temperature entropy coordinates. Heat is withdrawn at the constant temperature TR from the region to be refrigerated. Heat is rejected at the constant ambient temperature To. The cycle is completed by two isentropic processes that connect the high and low temperatures at the two extremes of entropy values. From Equation (17) the energy transfers are given by QR = TR ( S2 – S1 )

Q· R = m· T R ( s 2 – s 1 )

(21)

Qo = To ( S3 – S4 )

Q· R = m· T o ( s 2 – s 1 )

(22)

Since the cycle is being run to remove the heat QR, the COP of the Carnot cycle will be TR QR QR - = ----------------- = ------------------COP = ------WR Qo – QR To – TR

(23)

where Wnet work that would be supplied through isentropic processes of the cycle. Equation (23) also shows that the COP of the ideal reversible Carnot cycle is a function only of the two absolute temperatures between which it operates.

Fig 1-3. Carnot cycle

Fig 1-4. Possible thermodynamic realization of a Carnot cycle using a pure substance

The second law for cycles states that (1) no refrigerating cycle may have a COP higher than that for a reversible

1–5

cycle operated between the same temperature limits, and (2) all reversible cycles, when operated between the same temperature limits, have the same COP. Proofs of these statements are not presented here. They can be found in standard thermodynamics textbooks. Although the first law states that net heat flow into the cycle is equal to the net work done by the cycle (see Equation (19)), constraints are placed by the second law. For example, it is impossible to put heat into a system and convert it completely to work, a perfect conversion. The second law forces heat to be rejected (which is wasted), and the heat rejection should be to a different temperature reservoir than the temperature of the reservoir from which the heat is added for the cycle to exist. For refrigerating cycles it also establishes the only possible cycles are the ones in which the COP is less than the ideal COP, which is based only on the two extreme temperatures of the cycle. The Carnot cycle can be created by exploiting the constant temperature and constant pressure characteristics of the phase change saturation region of pure substances for constant temperature heat addition and rejection processes. The isentropic processes that link the states between the high and low temperatures can now be processes between high and low pressures within the saturation dome. Movement from a low to high pressure is easy since pumps and compressors are available, and from high to low pressure can be achieved via turbines. This is shown in Fig. 1-4. However, in a real device the Carnot cycle shown in Fig. 1-4 is difficult to obtain because of practical considerations. It is difficult to stop process heat addition/removal precisely at state 2. The compression from pressure of 2 to the higher pressure at 3 involves a substance that is mixed liquid and vapor at entry for which reliable compression devices are difficult to make (it is easy to either pump liquid alone or compress vapor). Similar problems occur at the expansion from 4 to 1, perhaps via a turbine, where the inlet is saturated liquid and the outlet is a wet saturated mixture of liquid and vapor. To overcome these problems the standard practical refrigeration cycle is shown in Figs. 1-5a and 1-5b. In this the constant temperature (and constant pressure) heat addition process ends when the entire saturated mixture has turned to saturated vapor at state 2 and a vapor compressor is used from pressure 2 to 3. From 3 to 4 the heat rejection is isobaric, and is only isothermal when in the saturation dome region. The isentropic expansion process from 4 to 1 of the Carnot cycle is replaced by a simple expansion valve that does not preserve entropy but preserves enthalpy. The advantage is that a moving device is avoided and a passive valve is replaced instead.

1–6

FUNDAMENTALS OF FLUID FLOW

Fundamentals of Fluid Flow In addition to density discussed in the previous section, the other property that is of importance in fluid flow is viscosity. The viscosity of the fluid links the shear stress on a fluid layer to the local velocity gradient across the layer. In the terminology of Fig. 1-7, the following relation defines viscosity Fig 1-5a. Practical refrigeration cycle in (T,s) coordinates

dV (24) τ = µ ------dy where τ =shear stress (tangential force per unit area); and µ =viscosity.

Fig 1-7. Velocity profiles and gradients in viscous flows

Fig 1-5b. Practical refrigeration cycle in P-h coordinates

A pure refrigerant or an azeotropic mixture can be used as the working fluid in the practical refrigerator to maintain constant temperature during phase changes through maintenance at constant pressure. Because of such concerns as high initial cost and increased maintenance requirements, the practical machine has one compressor, and the expander (engine or turbine) is replaced by a simple expansion valve. The valve throttles the refrigerant from high pressure to low pressure. The highest temperature in the cycle is at state 3. Thus, for ideal COP the temperature of 3 has to be used, and not that of state 4. The components of the refrigerator are sketched in Fig. 1-6. Because of the use of a compressor to go from the low temperature state 2 to high temperature state 3, the cycles are called vapor compression cycles.

Fig 1-6. Components of the refrigerator or air conditioner

The density of air and water at standard conditions of 68°F and 14.696 psi (sea level atmospheric pressure) are 0.075 lbm/ft3 and 62.3 lbm/ft3, respectively. All fluids are compressible to some degree; fluid density depends on pressure. Steady liquid flow may ordinarily be treated as incompressible, and incompressible flow analysis is satisfactory for gases and vapors at velocities below about 4000 to 8000 fpm, except in very long conduits. The units of viscosity are (force × time)/length2. At standard conditions the viscosities of air and water are 1.2×10−5 lbm/ft-s (= 3.7×10−7 lbf-s/ft2) and 6.7×10−4 lbm/ft-s (= 2.1×10−5 lbf-s/ft2), respectively. Kinematic viscosity is viscosity divided by density. The values of kinematic viscosity at standard conditions for air and water are 1.7×10−4 ft2/s and 1.08×10−5 ft2/s, respectively. In general if the Mach number of the fluid is less than 0.30 the fluid can be considered incompressible. In this chapter the fluid mechanics of pipe and duct flows are emphasized. This is because in HVAC systems flows in pipes and ducts are dominant methods of fluid transport, both airflow to the conditioned space and liquids that are the working fluids in the systems. External flows, such as flows of large fluid volumes over submerged objects, compressible flows such as in turbines and compressors, and the like are not discussed for the sake of brevity and relevance (or lack thereof). Flow in Pipes and Ducts.—The conservation of mass (or continuity equation) follows the same formulation as described before in the thermodynamics section, Equations (8) and (9). For conservation of energy some simplifications and modifications can be made. For fluid mechanics analyses only conservation of mechanical energy is considered. If steady flow is considered along a

1–7

FUNDAMENTALS OF FLUID FLOW

single path system then the conservation of energy for open systems, Equation (13), can be rewritten as 2

2

V - + gz V - + gz – u + pv + ---- u + pv + ---- in out 2 2 Q· + ---·- = 0 m

2

(25)

2

(26a)

2

--p- + V ------ + gz = constant ρ 2

(26b)

This is also the Bernoulli’s equation, which is the conservation of mechanical energy in an ideal fluid with no losses, no heat additions/removal, and no work. Bernoulli’s equation shows that the sum of pressure energy, kinetic energy, and potential energy is constant along a streamline of an ideal fluid. Even when temperature variations are present or the internal energies are not constant or heat transfer is present, the above mechanical energy conservation (and modifications presented below in case of losses) can be separately considered before the entire conservation of energy Equation (13) is eventually applied if needed. In the case of real fluids, where viscosity introduces drag, loss in mechanical energy occurs. This loss of mechanical energy is usually converted to increase in internal energy or is dissipated as heat, as per general conservation of energy as described in the thermodynamics section. In such a case of losses Equation (26a) of conservation of mechanical energy (presented in the form of energy per unit mass) is rewritten as 2

pv + V ------ + gz 2

2

V – pv + ------ + gz 2 in

out

= E loss

2

V + – pv + ------ + gz 2 in

out

(28)

+E mech = E loss

Here it is assumed that no work producing shafts or work producing moving boundaries are present and that the conservation of mass then permits the inlet and outlet mass flow rates to be the same. Pressure energy, kinetic energy, and potential energy are the mechanical energies of the flow. Since it was shown in the previous thermodynamics section that for most fluids, especially away from the saturation dome, the internal energy u is primarily a function of temperature, u is not considered a mechanical energy. The conservation of mechanical energy is then stated as V pv + V ------ + gz – pv + ------ + gz = 0 in out 2 2

2

pv + V ------ + gz 2

(27)

where Eloss = mechanical energy loss per unit mass, or rate of mechanical energy loss per unit mass flow rate. If mechanical energy is added to the flow through appropriate forms of work such as shaft work, the above can be modified to include a source term for mechanical energy. This is shown in

where Emech = mechanical energy added into the pipe or duct flow via a device such as fan, blower, or pump. Another consideration in the case of real fluids is the velocity V in Equation (28). The velocity V in the above is the average velocity through the pipe and is obtained from the mass flow rate. Because of the viscosity for real fluids, the velocity varies from zero at the duct or pipe walls to a maximum along the centerline. Since it is not mathematically true that the square of the average velocity is equal to the average of squared velocity unless the velocity profile is flat, this introduces an additional kinetic energy factor α in Equation (27) as follows: 2

2

VV pv + α ----+ gz – pv + α ------ + gz = E loss (29) in out 2 2 The kinetic energy factor is the ratio of the true kinetic energy of the flow to the kinetic energy represented by the mean or average velocity. It is shown that the value of this kinetic energy factor is 2.0 for laminar flow in circular pipes and 1.54 for laminar flow in wide rectangular channels. For turbulent flow the value is close to 1.0. Thus, for most HVAC applications, where the flow is turbulent, the original Bernoulli equation Equation (27), with losses is sufficient. Reynolds number is an important unitless parameter in fluid mechanics. It is defined for flow in pipes or ducts as ρVD (30) Re = -----------µ where D = diameter for circular pipes or a characteristic length for non-circular pipe. For other cross-sections it can be the hydraulic diameter Dh which is defined as 4A (31) D h = ------P where A = cross-sectional area, and P = perimeter. For external flows, characteristic length is usually the length of the object along the direction of the flow, or the value of the length’s local coordinate depending on the application. For pipe and duct flows the flow is laminar if the Reynolds number is less than 2300 and turbulent for greater values. This value of the critical Reynolds number is observed via experiments. For external flows parallel to a surface the flow remains laminar until the local Reynolds number reaches 5 ×105. Thus, the critical Reynolds number for internal flow is 2300 and for the external flow is 5 ×105. If the Reynolds number exceeds the critical value, the flow transitions to turbulence; below the critical value, the flow is laminar.

1–8

FUNDAMENTALS OF FLUID FLOW

Laminar flows are those where viscous effects dominate and the flow is orderly and layered. Fluid particles travel in smooth trajectories, without fluctuations. In steady internal flows, such as in pipes and ducts, the velocity profile is of a form that the forces due to pressure are balanced by viscous shear forces introduced by the presence of a wall (the velocity has to be zero at the wall). This gives rise to parabolic velocity profiles in the pipe or duct, where the velocity vanishes at the walls and is a maximum at the centerline. Turbulent flows which typically have higher velocities than laminar flows, involve random perturbations or fluctuations of the velocity and pressure, characterized by an extensive hierarchy of scales or frequencies. Only flows involving random perturbations without order or periodicity are turbulent; the velocity in such a flow varies with time or locale of measurement (Fig. 1-8). Turbulence can be quantified by statistical factors. The velocity most often used in velocity profiles is the temporal average velocity, and the strength of the turbulence is characterized by the root mean square of the instantaneous variation in velocity about the temporal average velocity. The effects of turbulence cause fluid to diffuse momentum, heat, and mass very rapidly across the flow. Because of the rapid fluctuations, fluid particles do not travel in smooth trajectories. The fluctuation allows particles to cross layers and thus cause a greater uniformity of flow properties and characteristics than in the laminar case. Because of this the turbulent velocity profiles in pipes or ducts are flatter throughout the core of the flow around the centerline and only fall off to zero velocity at the walls in a small region near the walls. It is because of the flatter velocity profile that the kinetic energy factor in Equation (29) is near unity, usually ranging from 1.01 to 1.10.

Fig 1-8. Velocity fluctuations with respect to time in turbulent flows

Time average velocity is defined by the Equation (32) T

1 V = --- V dt T

∫

(32)

0

The laminar and turbulent velocity profiles in a pipe flow are schematically compared in Fig. 1-9. Turbulent flow profiles are flat compared to the more pointed profiles of laminar flow. Near the wall, velocities of the tur-

bulent profile must drop to zero more rapidly than those of the laminar profile, so the shear stress and friction are much greater in the turbulent flow case. The velocity profiles shown in Fig. 1-9 correspond to fully developed flows. Fully developed flow regions are far away from inlets, from sections of sudden changes in cross-section, and from sources of mechanical energy input. The entrance flow region is the region from an inlet to the location where the fully developed region starts. It corresponds to the length Le of the pipe or duct needed for the flow to gradually change from its inlet profile to the new conditions.

Fig 1-9. Laminar and turbulent velocity profiles in a circular duct

Note that if the cross-sectional area remains constant then by the conservation of mass the average velocity throughout the pipe, including the inlet and fully developed regions, remains the same although the shape of the profile changes. With laminar flow following a rounded entrance, the entrance length Le depends on the Reynolds number: Le ----- ≅ 0.06Re D

(33)

At Re = 2000, a length of 120 diameters is needed to establish the fully developed parabolic velocity profile. However, the pressure gradient reaches the developed value of much sooner. With turbulent flow, a length of 80 to 100 diameters following the rounded entrance are needed for the velocity profile to become fully developed, but the friction loss per unit length reaches a value close to that of the fully developed flow value more quickly. After six diameters, the loss rate at a Reynolds number of 105 is only 14% above that of fully developed flow in the same length, while at 107 it is only 10% higher. For a sharp entrance, the flow separation causes a greater disturbance, but fully developed flow is achieved in about half the length required for a rounded entrance. With sudden expansion, the pressure change settles out in about eight times the diameter change (D2−D1), while the velocity profile takes at least a 50% greater distance to return to fully developed pipe flow. The mechanical energy loss in a duct or pipe of constant cross-section due to friction is given by the following:

1–9

FUNDAMENTALS OF FLUID FLOW

where f =friction factor.

2

LV E major-loss = E loss-friction = f ---- -----D2

(34)

Fig 1-10. Friction factor (Moody’s chart)

Losses due to friction are traditionally termed major losses in a pipe or duct system because for long pipes this loss dominates. However, for HVAC applications this may not be the case, and so the terminology of major losses may not be as appropriate as simple friction loss. Here L is the length of the pipe and D is the diameter. For noncircular ducts, the diameter may be replaced by the hydraulic diameter Dh. Friction factor is a function of the Reynolds number and the relative roughness of the pipe or duct walls. For large Reynolds numbers its value is fairly constant and is only a function of the roughness ε/D, where ε is the average height of roughness. This region is called the fully rough region. The value of friction factor is obtained via experiments and the values are given in Fig. 1-10 and in Equations (35) to (38). In the laminar region, where the Reynolds number is less than 2300, the friction factor is independent of the roughness. This is because the dominant viscous effects suppress any fluctuations introduced by roughness. The value of f for this region is given as

64f = ----Re

(35)

for circular cross-section pipes. In non circular pipes exact values for f can be derived for laminar flow, and the use of hydraulic diameter in Equation (35) may lead to significant errors. In turbulent flows for smooth pipes the friction factor is empirically correlated as f = 0.3164 ---------------0.25 Re 0.221 f = 0.0032 + ----------------0.237 Re

for Re < 10

5

(36)

5

for 10 < Re < 3 × 10

6

(37)

Another correlation for smooth pipes is 0.184f = -------------0.20 Re

for Re > 2 × 10

4

(38)

1–10

FUNDAMENTALS OF FLUID FLOW

For pipes and ducts that are not smooth, where a wall roughness ε is defined (in the units of length), the relative roughness ε/D plays an important role. In the fully rough region where the influence of Reynolds number is very small, the turbulent friction factor has been empirically correlated by D 1 ----- = 1.14 + 2 log ---- (39) ε f In the other turbulent regime between the smooth walled pipes and the fully rough region several empirical correlations exist. A commonly used correlation is the Colebrook function given by D 9.3D 1 (40) ----- = 1.14 + 2 log ---- – 2 log 1 + -------------- ε εRe f f The use of hydraulic diameter Dh in the above correlations for turbulent flows is quite acceptable since errors up to 5% may be introduced as observed experimentally. Mechanical energy losses other than friction loss described above in a pipe or duct are usually correlated by using a loss coefficient K: 2

V (41) E loss-minor = E loss-fixture = K -----2 For inlets, the V in Equation (41) is velocity in the pipe after the entrance, and for outlets it is the velocity in the pipe before the exit. Thus, which velocity to use is a function of the geometry or of the type of the fitting or fixture. For a sudden expansion it is velocity before expansion and for a sudden contraction of cross sectional area it is velocity after contraction. The values of K are compiled via experiments and will be discussed in later chapters. It may be noted that although the true loss of mechanical energy occurs over a distance the formula lumps its effects at one location. For example, the inlet loss occurs over the entrance region but is lumped at the entrance by this formula. Note that these effects do not include the friction loss over the corresponding length, but only the additional losses associated with disturbances of flow and transitions. For systems with long pipe lengths these losses are a small portion of the total losses where the friction losses dominate and are traditionally called minor losses. However, for HVAC systems this is not the case and the multiple bends and junctions, size changes, and the presence of various fixtures and fittings ensure that these effects are a very significant part of the total loss. Some Details of Relevant Flow Fields.—T h e p r e s ence of walls in fluid flows introduces velocity gradients in the flow field since the fluid necessarily has to be at zero velocity at the wall. This effect, termed friction, usually is present in the form of a boundary layer in fast moving flows. A boundary layer is the slender region near the wall where the velocity goes from the zero wall velocity to the free stream velocity. All the viscous effects are con-

centrated in the boundary layer. For flow around bodies, this layer (which is quite thin relative to distances in the flow direction) encompasses all viscous or turbulent actions, causing the velocity in it to vary rapidly from zero at the wall to that of the outer flow at its edge. Boundary layers are generally laminar near the start of their formation but may become turbulent downstream of the transition point. For conduit or pipe flows, spacing between adjacent walls or within the pipe diameter is generally small compared with distances in the flow direction. As a result, layers from the walls meet at the centerline to fill the conduit. The region after the layers meet is the fully developed flow region that was discussed earlier. The length of the pipe or conduit where the boundary layers are still growing and have not yet met at the centerline is called the entrance region. Near the start of the straight conduit or pipe, the layer is very thin (and laminar in all probability), so the uniform velocity core outside has a velocity only slightly greater than the average velocity. As the layer grows in thickness, the slower velocity near the wall needs to increase in the uniform core to satisfy continuity. As the flow proceeds, the wall layers grow (and the centerline velocity increases) until they join, after an entrance length Le. Application of the Bernoulli relation to the core flow indicates a decrease in pressure along the layer. However, if the cross-sectional area is increasing along the flow so that average velocity decreases along the flow, the adverse pressure gradient can lead to flow separation. The development of the boundary layer in an adversepressure gradient situation (velocity at edge of layer decreasing in flow direction) causes the separation of the boundary layer where downstream from the separation point the fluid backflows near the wall. Separation is due to flow near the wall no longer having energy to move into the higher pressure imposed by the decrease in velocity at the edge of the layer. The locale of this separation is difficult to predict, especially for the turbulent boundary layer. Analyses verify the experimental observation that a turbulent boundary layer is less subject to separation than a laminar one because the flow in the turbulent layer has greater kinetic energy.

Fig 1-11a. Flow through orifice

Fig 1-11b. Flow through sudden contraction or expansion

FUNDAMENTALS OF HEAT TRANSFER

Flow separation is observed in several situations in HVAC equipment. Examples of flow through an orifice and through contractions and expansions are shown in Figs. 1-11a and 1-11b. The manifestation of separation is sudden pressure drop that is not recovered in the flow. In liquid flow, gas- or vapor-filled pockets can occur if the absolute pressure is reduced to vapor pressure or less. In this case, a cavity or series of cavities form. This lowering of pressure may be due to flow separation or poor selection of operating parameters or equipment design. This is called cavitation and initial evidence of cavitation is the collapse noise of many small bubbles as they are carried by the flow into regions of higher pressure. The severity of cavitation increases as velocity increases or pressure decreases. Collapse of the cavities on or near solid boundaries becomes so frequent that the cumulative impact in time results in damage in the form of cavitations erosion of the surface or excessive vibration. As a result, pumps can lose efficiency or their parts may erode locally. Noise from Fluid Flow.—Noise from fluid flow is especially important in HVAC systems where, for example, noisy ducts inside buildings can significantly cause complaints. Noise in flowing fluids results from unsteady flow fields and can be at discrete frequencies or broadly distributed over the audible range. With liquid flow, cavitation results in noise through the collapse of vapor bubbles. Noise in pumps or fittings (such as valves) can be easily eliminated by raising the system pressure. With severe cavitation, the resulting unsteady flow can produce by-product noise from induced vibration of adjacent parts. Noise produced in pipes and ducts is especially associated with the loss through the valves and fittings. The sound pressure of noise in water pipe flow increases linearly with the head loss; the broadband noise increases, but only in the lower frequency range. Fitting-produced noise levels also increase with fitting loss (even without cavitation) and significantly exceed noise levels pipe flow. The relation between noise and loss associated with is because both involve excessive flow perturbations. A valve’s pressure flow characteristics and structural elasticity may be such that for some operating point it oscillates, perhaps in resonance with part of the piping system, to produce excessive noise. A change in operating point conditions or details of valve geometry can result in significant noise reduction. Pumps and blowers are strong potential noise sources. Turbo machinery noise is associated with blade-flow occurrences. Broadband noise results from vortex and turbulence interaction with walls and is primarily a function of the operating point of the machine. For blowers, noise is a minimum at the peak efficiency point. Narrowband noise also occurs at the blade, crossing frequency and its harmonics. To reduce this noise, a designer

1–11

increases the clearances between impeller and housing, and spaces impeller blades unevenly around the circumference. However, this may be impractical because they lead to other problems. Fundamentals of Heat Transfer Heat is transfer of energy due to temperature difference. Solution of conservation of energy Equations (11) to (13) require the accurate assessment of heat transported into a closed or open system. Heat transfer theory is needed to quantify the rate at which heat will transfer for a given set of conditions. Although application of thermodynamic law requires the value of heat transfer be known, it does not provide any method to ascertain whether the required or prescribed heat transfer rate is feasible for the given system and available temperature differences. Thermal energy transfer always occurs in the direction of decreasing temperature. Thermal conduction is the mechanism of heat transfer whereby energy is transported between parts of a continuum by the transfer of kinetic energy between particles or groups of particles at the atomic and subatomic levels. In liquids and gases, conduction is caused by elastic collision of molecules; in liquids and electrically nonconducting solids, it is caused by oscillations of the lattice structure (or phonons); in metals it occurs due to the motion of free electrons. In the study of engineering heat transfer, in contrast with physics, the underlying microscopic mechanism is not of interest and the entire conduction effect is lumped into understanding thermophysical properties. Thermal conductivity k is defined as the ratio of heat flow rate per unit area to the local temperature gradient. This is given by Fourier’s law, dT Q· = – kA -----dx

(42)

where A = cross-sectional area (ft2) normal to the heat flow; and Q· = rate of heat flow (Btu/hr). In Equation (42) steady heat flow along the x direction is considered. The units for k are Btu-ft/hr-ft2-°F and units of temperature T and distance x are °F and ft, respectively. The minus sign indicates that heat flow is positive in the direction of decreasing temperature (i.e., negative gradient). Thermal conductivity is an intrinsic property of a material and is tabulated in handbooks and property tables. In solid opaque bodies, thermal conduction is the most significant heat transfer mechanism because there is no flow of material, and transport by other heat transfer mechanisms such as convection and radiation are not present. In fluids, thermal conduction dominates in the region very close to the flows solid boundary. However,

1–12

FUNDAMENTALS OF HEAT TRANSFER

in fluids the effects of conduction near the boundary are lumped into heat convection coefficients. The distinction made between thermal convection and thermal conduction is this: Heat convection involves energy transfer or exchange at an interface, notably, the solid-fluid interface— the walls. Heat conduction occurs throughout the entire body of the transport mechanism. The heat transfer coefficient h is defined at the interface in the fluid region so that the rate of heat flow is related to the temperature difference by –T ) Q· = hA ( T (43) surface

fluid

where Tsurface = temperature of the solid surface; and Tfluid = bulk or free stream temperature of the fluid (not the local temperature of the fluid near the surface) Equation (43) relates the heat flow rate from the surface to the fluid. The units for h are Btu/hr-ft2-°F. For pipe flows the fluid temperature in Equation (43) is the bulk mean temperature at the cross-section where the convective heat transport is being considered. The heat transfer coefficient is an engineering concept that combines all microscopic energy transport processes near a solid-fluid interface into a single coefficient. The heat transfer coefficient is strongly a function of flow conditions and increases as flow velocities increase. It is not an intrinsic property of any material and it must computed or measured at the required flow configurations and geometries. Thermal convection can be classified into two types: forced and natural. Forced convection occurs when fluid is flowing due to external mechanisms such as the case of it being driven by fans or pumps. Natural convection occurs in fluid that is normally quiescent, but is forced into motion by thermal buoyancy effects. For example, hot fluid near a hot wall rises because its density is lowered and colder fluid moves in to replace it, thereby setting up a loop that removes heat from the hot wall. In this case the value of h is strongly a function of temperature, the variations of the density with temperature, and the geometry of the system. For internal flows through pipes and ducts the convective heat transfer coefficient stays constant along the direction of the flow in the fully developed region, i.e., after the entrance length region. For external flows the local convective heat transfer coefficient changes along the direction of the flow because the boundary layer keeps on growing, as opposed to internal flow where the boundary layers meet at the centerline at the end of the entrance region and no further changes in flow occur. The forced flow heat transfer correlations can be written generally as n

m

(44) Nu = CRe Pr where C, m, n = constants whose values are functions of the operating range and geometry, and

Nu = Nusselt number. Nusselt number is given by hL Nu = -----k

(45)

where L =characteristic length for the configuration; and k =the thermal conductivity of the liquid. The length in the Reynolds number is the same characteristic length as in the Nusselt number. For pipes the characteristic length is the diameter D or the hydraulic diameter Dh for non-circular pipes. For flat surfaces the length to be used is the entire length along the direction of flow for computing average heat transfer coefficient, or the local coordinate along the direction of the flow to compute the local heat transfer coefficient. The Prandtl number is a ratio of viscous and thermal diffusivities of the fluid and given by µC p µ- = --------Pr = -----ρα k

(46)

where α =thermal diffusivity. The thermal diffusivity is given by kα = --------ρC p

(47)

The general forced convection Equation (44) remains valid for flows inside pipe and ducts, as well as over flat surfaces and other objects. The constants C, n, and m for the above forced convection correlation are given in Table 1-1. Table 1-1. Forced Convection Factors Characteristic Length

C

n

m

Laminar, fully developed flow in a circular pipe, constant surface temperature

Diameter D

3.66

0

0

Laminar, fully developed flow in a circular pipe, constant surface heat flux

Diameter D

4.36

0

0

Laminar, fully developed flow in a duct of square cross-section, constant surface temperature

Hydraulic diameter Dh

2.98

0

0

Laminar, fully developed flow in a duct of square cross-section, constant surface heat flux

Hydraulic diameter Dh

3.61

0

0

Laminar, fully developed flow in a duct of triangular cross-section, constant surface temperature

Hydraulic diameter Dh

2.47

0

0

Laminar, fully developed flow in a duct of triangular cross-section, constant surface heat flux

Hydraulic diameter Dh

3.11

0

0

Turbulent, fully developed flow in a pipe, heat added to fluid

Hydraulic diameter Dh

0.023

0.8

0.4

Turbulent, fully developed flow in a pipe, heat removed from fluid

Hydraulic diameter Dh

0.023

0.8

0.3

Description of Flow

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FUNDAMENTALS OF HEAT TRANSFER

Table 1-1. (Continued) Forced Convection Factors Characteristic Length

Description of Flow

C

n

m

Local coordinate in flow direction x

0.332

0.5

0.33

Laminar flow parallel to flat surface, constant surface temperature, average heat transfer coefficient

Length in flow direction L

0.664

0.5

0.33

Laminar flow parallel to flat surface, constant surface heat flux, local heat transfer coefficient

Local coordinate in flow direction x

0.453

0.5

0.33

Laminar flow parallel to flat surface, constant surface heat flux, average heat transfer coefficient

Length in flow direction L

0.68

0.5

0.33

Turbulent flow parallel to flat surface, constant surface temperature, local heat transfer coefficient

Local coordinate in flow direction x

0.0296

0.8

0.33

Turbulent flow parallel to flat surface, constant surface heat flux, local heat transfer coefficient

Local coordinate in flow direction x

0.0308

0.8

0.33

Cross-flow over a circular cylinder, average heat transfer coefficient, 0.4< Re