Issuu on Google+

HVAC SYSTEMS TESTING, ADJUSTING & BALANCING

SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSOCIATION, INC.


HVAC SYSTEMS TESTING, ADJUSTING & BALANCING

THIRD EDITION — AUGUST, 2002

SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSOCIATION, INC. 4201 Lafayette Center Drive Chantilly, VA 20151-1209


HVAC SYSTEMS TESTING, ADJUSTING & BALANCING COPYRIGHT2002 All Rights Reserved by

SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSOCIATION, INC. 4201 Lafayette Center Drive Chantilly, VA 20151 Printed in the U.S.A.

FIRST EDITION - 1983 SECOND EDITION - JULY, 1993 THIRD EDITION - AUGUST, 2002

Except as allowed in the Notice to Users and in certain licensing contracts, no part of this book may be reproduced, stored in a retrievable system, or transmitted, in any form or by any means, electronic, mechanical, photocopying, recording, or otherwise, without the prior written permission of the publisher.


FOREWORD This handbook has been extensively updated for 2002 from the original 1983 publication and includes all of the many changes that have takes place in the industry since the 1990’s. We have added many new sections covering variable frequency drives (VFD), direct digital control (DDC) systems, lab hood exhaust balancing, and the latest changes in the balancing equipment and procedures. All of the system testing, adjusting, and balancing fundamentals that make up the original text has been updated, and all helpful reference tables and charts in the Appendix have been extensively updated. This handbook will provide any SMACNA contractor already familiar with mechanical system operation basics, with the information necessary to balance most heating, ventilation, and air conditioning (HVAC) systems. Chapters on both air and water side HVAC system adjusting and balancing are included, and the chapters on system controls have been totally rewritten to reflect the trend away from pneumatic controls and towards programmable micro−processor controls. Most of today’s HVAC systems are being designed with many more individually controlled temperature zones to im− prove occupant comfort, and variable speed fans and pumps are now commonplace to provide the exact amount of heating and cooling system capacity necessary to minimize energy usage. New occupant air ventilation codes are much more restrictive, at the same time building envelopes are becoming much tighter. The combination of constantly changing HVAC flows and increased demand for fresh and filtered ventilation air for all occupants is placing much more emphasis on proper HVAC system operation and balancing. Any SMACNA contractor wanting to become part of this rapidly growing field is strongly encouraged to read other related SMACNA publications available, and take part in the many training courses offered to become a certified TAB Contractor. The International Training Institute provides a Certified Technician program for journeyman sheet metal workers who already have a basic understanding of system testing and balancing, and many of these courses are avail− able in versions for home study. The building construction industry is experiencing a major growth in demand for trained and experienced contractors who can balance today’s much more complex HVAC systems.

SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSOCIATION, INC.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

iii


TASK FORCE Bill Freese, Chairman International Testing & Balancing, Ltd. Seaford, New York

Ray Coleman Certified Testing & Balancing, Inc. Riverton, Utah

David Aldag Aldag−Honold Mechanical, Inc. Sheboygan, Wisconsin

Ben Dutton SMACNA, Inc. Chantilly, Virginia

John Brue Balancing Precision, Inc. Bloomington, Illinois

Eli P. Howard, III SMACNA, Inc. Chantilly, Virginia

OTHER CONTRIBUTORS J. R. Yago & Associates Consulting Engineers Manakin−Sabot, Virginia

iv

HVAC SYSTEMS Testing Adjusting & Balancing • Third Edition


NOTICE TO USERS OF THIS PUBLICATION

1.

DISCLAIMER OF WARRANTIES

a) The Sheet Metal and Air Conditioning Contractors’ National Association (“SMACNA”) provides its product for informational purposes. b) The product contains “Data” which is believed by SMACNA to be accurate and correct but the data, including all information, ideas and expressions therein, is provided strictly “AS IS”, with all faults. SMACNA makes no warranty either express or implied regarding the Data and SMACNA EXPRESSLY DISCLAIMS ANY IMPLIED WARRANTIES OF MERCHANTABILITY OR FITNESS FOR PARTICULAR PURPOSE. c) By using the data contained in the product user accepts the Data “AS IS” and assumes all risk of loss, harm or injury that may result from its use. User acknowledges that the Data is complex, subject to faults and requires verification by competent professionals, and that modification of parts of the Data by user may impact the results or other parts of the Data. d) IN NO EVENT SHALL SMACNA BE LIABLE TO USER, OR ANY OTHER PERSON, FOR ANY INDIRECT, SPECIAL OR CONSEQUENTIAL DAMAGES ARISING, DIRECTLY OR INDIRECTLY, OUT OF OR RELATED TO USER’S USE OF SMACNA’S PRODUCT OR MODIFICATION OF DATA THEREIN. This limitation of liability applies even if SMACNA has been advised of the possibility of such damages. IN NO EVENT SHALL SMACNA’S LIABILITY EXCEED THE AMOUNT PAID BY USER FOR ACCESS TO SMACNA’S PRODUCT OR $1,000.00, WHICHEVER IS GREATER, REGARDLESS OF LEGAL THEORY. e) User by its use of SMACNA’s product acknowledges and accepts the foregoing limitation of liability and disclaimer of warranty and agrees to indemnify and hold harmless SMACNA from and against all injuries, claims, loss or damage arising, directly or indirectly, out of user’s access to or use of SMACNA’s product or the Data contained therein.

2.

ACCEPTANCE

This document or publication is prepared for voluntary acceptance and use within the limitations of application defined herein, and otherwise as those adopting it or applying it deem appropriate. It is not a safety standard. Its application for a specific project is contingent on a designer or other authority defining a specific use. SMACNA has no power or authority to police or enforce compliance with the contents of this document or publication and it has no role in any representations by other parties that specific components are, in fact, in compliance with it.

3.

AMENDMENTS

The Association may, from time to time, issue formal interpretations or interim amendments, which can be of significance between successive editions.

4.

PROPRIETARY PRODUCTS

SMACNA encourages technological development in the interest of improving the industry for the public benefit. SMACNA does not, however, endorse individual manufacturers or products.

5.

FORMAL INTERPRETATION

a) A formal interpretation of the literal text herein or the intent of the technical committee or task force associated with the document or publication is obtainable only on the basis of written petition, addressed to the Technical Resources Department and sent to the Association’s national office in Chantilly, Virginia. In the event that the petitioner has a substantive disagreement with the interpretation, an appeal may be filed with the Technical Resources Committee, which has technical oversight responsibility. The request must pertain to a specifically identified portion of the document that does not involve published text which provides the requested information. In considering such requests, the Association will not review or judge products or components as being in compliance with the document or publication. Oral and written interpretations otherwise obtained from anyone affiliated with the Association are unofficial. This procedure does not prevent any committee or task force chairman, member of the committee or task force, or staff liaison from expressing an opinion on a provision within the document, provided that such person clearly states that the opinion is personal and does not represent an official act of the Association in any way, and it should not be relied on as such. The Board of Directors of SMACNA shall have final authority for interpretation of this standard with such rules or procedures as they may adopt for processing same. b) SMACNA disclaims any liability for any personal injury, property damage, or other damage of any nature whatsoever, whether special, indirect, consequential or compensatory, direct or indirectly resulting from the publication, use of, or reliance upon this document. SMACNA makes no guaranty or warranty as to the accuracy or completeness of any information published herein.

6.

APPLICATION

a) Any standards contained in this publication were developed using reliable engineering principles and research plus consultation with, and information obtained from, manufacturers, users, testing laboratories, and others having specialized experience. They are

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

v


subject to revision as further experience and investigation may show is necessary or desirable. Construction and products which comply with these Standards will not necessarily be acceptable if, when examined and tested, they are found to have other features which impair the result contemplated by these requirements. The Sheet Metal and Air Conditioning Contractors’ National Association and other contributors assume no responsibility and accept no liability for the application of the principles or techniques contained in this publication. Authorities considering adoption of any standards contained herein should review all federal, state, local, and contract regulations applicable to specific installations. b) In issuing and making this document available, SMACNA is not undertaking to render professional or other services for or on behalf of any person or entity. SMACNA is not undertaking to perform any duty owed to any person or entity to someone else. Any person or organization using this document should rely on his, her or its own judgement or, as appropriate, seek the advice of a competent professional in determining the exercise of reasonable care in any given circumstance.

7.

REPRINT PERMISSION

Non-exclusive, royalty-free permission is granted to government and private sector specifying authorities to reproduce only any construction details found herein in their specifications and contract drawings prepared for receipt of bids on new construction and renovation work within the United States and its territories, provided that the material copied is unaltered in substance and that the reproducer assumes all liability for the specific application, including errors in reproduction.

8.

THE SMACNA LOGO

The SMACNA logo is registered as a membership identification mark. The Association prescribes acceptable use of the logo and expressly forbids the use of it to represent anything other than possession of membership. Possession of membership and use of the logo in no way constitutes or reflects SMACNA approval of any product, method, or component. Furthermore, compliance of any such item with standards published or recognized by SMACNA is not indicated by presence of the logo.

vi

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


TABLE OF CONTENTS


TABLE OF CONTENTS FOREWORD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

iii

TASK FORCE . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

iv

NOTICE TO USERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

v

TABLE OF CONTENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

vii

CHAPTER 1 1.1 1.2 1.3

INTRODUCTION INTRODUCTION TO TAB WORK . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . THE TAB TECHNICIAN/TEAM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . GENERAL REQUIREMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1.1 1.1 1.2

CHAPTER 2 HVAC FUNDAMENTALS 2.1 2.2 2.3 CHAPTER 3 3.1 3.2 3.3 3.4 3.5 3.6 CHAPTER 4 4.1 4.2 4.3 4.4 4.5 4.6 4.7 CHAPTER 5 5.1 5.2 5.3 5.4 5.5 CHAPTER 6 6.1 6.2 6.3 6.4 6.5

HEAT FLOW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PSYCHROMETRICS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FLUID MECHANICS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

2.1 2.6 2.19

ELECTRICAL EQUIPMENT AND CONTROLS ELECTRICAL SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ELECTRICAL SERVICES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . TRANSFORMERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . MOTORS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . MOTOR CONTROLS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . VARIABLE FREQUENCY DRIVES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

3.1 3.1 3.5 3.5 3.8 3.9

TEMPERATURE CONTROL AUTOMATIC TEMPERATURE CONTROL SYSTEMS . . . . . . . . . . . . . . . . . . . . CONTROL LOOPS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . CONTROL DIAGRAMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . CONTROL RELATIONSHIPS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ATC SYSTEM ADJUSTMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . TAB/ATC RELATIONSHIP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . CENTRALIZED CONTROL SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

4.1 4.2 4.5 4.5 4.6 4.6 4.7

FANS FAN CHARACTERISTICS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN CONSTRUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN AIRFLOW AND PRESSURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN/SYSTEM CURVE RELATIONSHIP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN CAPACITY RATINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

5.1 5.4 5.10 5.13 5.17

AIR DISTRIBUTION AND DEVICES AIR TERMINAL BOXES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . VARIABLE AIR VOLUME (VAV) TERMINAL BOXES . . . . . . . . . . . . . . . . . . . . . OTHER AIRFLOW DEVICES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . AIR DISTRIBUTION BASICS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ROOM AIR DISTRIBUTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.1 6.3 6.3 6.6 6.9

vii


CHAPTER 7 7.1 7.2 7.3 7.4 7.5 CHAPTER 8 8.1 8.2 8.3 8.4 8.5 CHAPTER 9 9.1 9.2 9.3 9.4

AIR SYSTEMS INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . TYPES OF AIR SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . AIR SYSTEM DESIGN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . DUCT SIZING EXAMPLES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SUMMARY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

7.1 7.2 7.9 7.11 7.14

HYDRONIC EQUIPMENT PUMPS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PUMP / SYSTEM CURVE RELATIONSHIP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PUMP INSTALLATION CRITERIA . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HYDRONIC HEATING AND COOLING SOURCES . . . . . . . . . . . . . . . . . . . . . . TERMINAL HEATING AND COOLING UNITS . . . . . . . . . . . . . . . . . . . . . . . . . . .

8.1 8.7 8.11 8.13 8.14

HYDRONIC SYSTEMS HYDRONIC SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HYDRONIC SYSTEM DESIGN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HYDRONIC DESIGN PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . STEAM SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

9.1 9.8 9.13 9.14

CHAPTER 10 REFRIGERATION SYSTEMS 10.1 10.2 10.3 10.4 10.5 10.6 10.7 CHAPTER 11 11.1 11.2 11.3 11.4 11.5 11.6 11.7 11.8

REFRIGERATION SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . REFRIGERATION TERMS AND COMPONENTS . . . . . . . . . . . . . . . . . . . . . . . . SAFETY CONTROLS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . OPERATING CONTROLS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . REFRIGERANTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . THERMAL BULBS AND SUPERHEAT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . COMPRESSOR SHORT CYCLING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

10.1 10.2 10.4 10.4 10.4 10.4 10.6

TAB INSTRUMENTS INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . AIRFLOW MEASURING INSTRUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PRESSURE GAGE, CALIBRATED . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ROTATION MEASURING INSTRUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . TEMPERATURE FUNCTION TACHOMETER MEASURING INSTRUMENTS ELECTRICAL MEASURING INSTRUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . COMMUNICATION DEVICES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HYDRONIC FLOW MEASURING DEVICES . . . . . . . . . . . . . . . . . . . . . . . . . . . .

11.1 11.1 11.9 11.12 11.16 11.22 11.23 11.24

CHAPTER 12 PRELIMINARY TAB PROCEDURES 12.1 12.2 12.3 12.4 12.5 12.6

INITIAL PLANNING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . CONTRACT DOCUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SYSTEM REVIEW AND ANALYSIS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . THE AGENDA . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PLANNING FIELD TAB PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PRELIMINARY FIELD PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

12.1 12.1 12.2 12.4 12.5 12.6

CHAPTER 13 GENERAL AIR SYSTEM TAB PROCEDURES 13.1 13.2 13.3 13.4 13.5 13.6 13.7 viii

BASIC FAN TESTING PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SYSTEM STARTUP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN TESTING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . DEFICIENCY REVIEW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . RETURN AND OUTSIDE AIR SETTINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ANALYSIS OF MEASUREMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . RECORDING DATA . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

13.1 13.1 13.1 13.2 13.2 13.3 13.3


13.8 13.9 13.10 13.11 13.12 13.13 13.14 13.15 13.16 13.17 13.18 13.19 13.20 13.21 13.22 13.23 13.24 13.25

PROPORTIONAL BALANCING (RATIO) METHOD . . . . . . . . . . . . . . . . . . . . . . . 13.3 PERCENTAGE OF DESIGN AIRFLOW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.3 SYSTEM AIRFLOW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.5 BASIC OUTLET BALANCING PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.5 STEPWISE METHOD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.5 FAN ADJUSTMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 WET COIL CONDITIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 AIRFLOW TOTALS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 EXHAUST FANS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 FAN DRIVE ADJUSTMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 DAMPER ADJUSTMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.7 DUCT TRAVERSES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.7 SYSTEM DEFICIENCIES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.7 FUME HOOD EXHAUST BALANCING PROCEDURES . . . . . . . . . . . . . . . . . . . 13.7 DUST COLLECTION AND EXHAUST BALANCING PROCEDURES . . . . . . . 13.8 AIR FLOW MEASUREMENTS ON DISCHARGE STACKS . . . . . . . . . . . . . . . . 13.11 INDUSTRIAL VENTILATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.12 SELECTION OF INSTRUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.12

CHAPTER 14 TAB PROCEDURES FOR SPECIFIC AIR SYSTEMS 14.1 14.2 14.3 14.4 14.5 14.6 14.7

INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . VARIABLE AIR VOLUME (VAV) SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . MULTI-ZONE SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . INDUCTION UNIT SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . DUAL DUCT SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SPECIAL EXHAUST AIR SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PROCESS EXHAUST AIR SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

14.1 14.1 14.13 14.14 14.14 14.16 14.17

CHAPTER 15 HYDRONIC SYSTEM TAB PROCEDURES 15.1 15.2 15.3 15.4 15.5 15.6 15.7

HYDRONIC SYSTEM MEASUREMENT METHODS . . . . . . . . . . . . . . . . . . . . . . 15.1 BASIC HYDRONIC SYSTEM PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.3 PIPING SYSTEM BALANCING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.4 BALANCING SPECIFIC SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.5 VARIABLE VOLUME FLOW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.9 PRIMARY-SECONDAR Y SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.11 SUMMER-WINTER SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.11

CHAPTER 16 TAB REPORT FORMS 16.1 16.2

PREPARING TAB REPORT FORMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . DESCRIPTION OF USE . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

APPENDIX A

16.1 16.1

DUCT DESIGN TABLES & CHARTS DUCT DESIGN TABLES AND CHARTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HVAC EQUATIONS - (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HVAC EQUATIONS - (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SI UNITS AND EQUIVALENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SOUND DESIGN EQUATIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FITTING EQUIVALENTS (WATER) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PROPERTIES OF STEAM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . STEAM PIPING (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . STEAM PIPING (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

A.1 A.31 A.35 A.39 A.41 A.43 A.44 A.45 A.49 A.54

GLOSSARY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

G.1

INDEX . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

I.1

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

ix


TABLES

5-1 6-1

Typical Fan Rating Table . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.7 Typical Ratios of Damper to System Resistance for Flow Characteristic Curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.6 6-2 Guide to Use of Various Outlets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.12 6-3 Recommended Return Air Inlet Face Velocities . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.14 6-4 Air Outlets and Diffusers Total Pressure Loss Average—in. wg (Pa) . . . . . . . . . . . 6.15 6-5 Supply Registers Total Pressure Loss Average—in. wg (Pa) . . . . . . . . . . . . . . . . . . 6.15 6-6 Return Registers Total Pressure Loss Average—in. wg (Pa) . . . . . . . . . . . . . . . . . . 6.15 8-1 Characteristics of Centrifugal Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.3 8-2 Characteristics of Common Types of Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.3 8-3 Flow vs Total Head (Cooling Tower Application) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.11 9-1 Hydronic Trouble Analysis Guide . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9.8 11-1 Airflow Measuring Instruments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.9 11-2 Instruments for Hydronic Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.11 11-3 Hydronic Measuring Instruments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.11 11-4 Rotation Measuring Instruments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.13 11-5 Instrumentation for Air & Hydronic Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.16 11-6 Instruments for Air Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.17 11-7 Temperature Measuring Instruments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.21 15-1 Load-Flow Variations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.10 A-1 Duct Material Roughness Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.3 A-2 Circulation Equivalents of Rectangular Ducts for Equal Friction and Capacity (I-P) (2) Dimensions in Inches . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.5 A-2 Circulation Equivalents of Rectangular Ducts for Equal Friction and Capacity (I-P) (2) Dimensions in Inches (continued) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.6 A-3 Circular Equivalents of Rectangular Ducts for Equal Friction and Capacity (SI) (2) Dimensions in mm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.7 A-3 Circular Equivalents of Rectangular Ducts for Equal Friction and Capacity (SI) (2) Dimensions in mm (continued) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.8 A-4 Velocities/Velocity Pressures (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.9 A-5 Velocities/Velocity Pressures (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.10 A-6 Angular Conversion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.10 A-7 Loss Coefficients for Straight-Through Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.11 A-8 Recommended Criteria for Louver Sizing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.12 A-9 Typical Design Velocities for Duct Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.13 A-10 Elbow Loss Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.14 A-1 1 Transition Loss Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.17 A-12 Rectangular Branch Connection Loss Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . A.19 A-13 Round Branch Connection Loss Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.23 A-14 Miscellaneous Fitting Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.27 HVAC Equations (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.31 A-15 Converting Pressure In Inches of Mercury to Feet of Water at Various Water Temperatures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.33 A-16 Air Density Correction Factors (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.34 HVAC Equations (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.35 A-17 Air Density Correction Factors (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.38 A-18 SI Units And Equivalents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.39 A-19 SI Equivalents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.40 A-20 Sound Design Equations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.41 A-21 Equivalent Length in Feet of Pipe for 90 Elbows . . . . . . . . . . . . . . . . . . . . . . . . . . A.43 A-22 Equivalent Length in Meters of Pipe for 90 Elbows . . . . . . . . . . . . . . . . . . . . . . . . A.43 A-23 Iron and Copper Elbow Equivalents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.43 A-24 Properties of Saturated Steam (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.44 A-25 Properties of Saturated Steam (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.44 A-26 Steam Piping (I-P) Flow Rate of Steam in Schedule 40 Pipe at Initial Saturation Pressure of 3.5 and 12 psig (Flow Rate expressed in Pounds per Hour) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.45 x

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


TABLES (continued) A-27 Comparative Capacity of Steam Lines at Various Pitches for Steam and Condensate Flowing in Opposite Directions (Pitch of Pipe in Inches per 10 Feet – Velocity in Feet per Second) . . . . . . . . . A.45 A-28 Pressure Drops In Common Use for Sizing Steam Pipe (For Corresponding Initial Steam Pressure) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.46 A-29 Length in Feet of Pipe to be Added to Actual Length of Run — Owing to Fittings — to Obtain Equivalent Length . . . . . . . . . . . . . . . . . . A.46 A-30 Steam Pipe Capacities for Low Pressure Systems (For Use on One-Pipe Systems or Two-Pipe Systems in which Condensate Flows Against the Steam Flow) . A.47 A-31 Return Main and Riser Capacities for Low-Pressure Systems—Pounds per Hour (Reference to this table will be made by column letter G through V) . . . . . . . . . A.48 A-32 Flow Rate in kg/h of Steam in Schedule 40 Pipe at Initial Saturation Pressure of 15 and 85 kPa Above Atmospheric . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.49 A-33 Comparative Capacity of Steam Lines at Various Pitches for Steam and Condensate Flowing in Opposite Directions . . . . . . . . . . . . . . . . . . . . . . . . . . A.49 A-34 Equivalent Length of Fittings to be Added to Pipe Run . . . . . . . . . . . . . . . . . . . . . . A.50 A-35 Steam Pipe Capacities for Low-Pressure Systems (For Use on One-Pipe Systems or Two-Pipe Systems in which Condensate Flows Against the Steam Flow) . A.51 A-36 Return Main and Riser Capacities for Low-Pressure Systems — kg/h . . . . . . . . A.52

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

xi


FIGURES

2-1 Heat Transfer by Conduction and Radiation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-2 Convection Heat Transfer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-3 Counterflow Airstreams . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-4 Parallel Flow Airstreams . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-5 Cross-flow Airstreams . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-6 Parallel and Counterflow Heat Transfer Curves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-7 Psychrometric Chart (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-8 Psychrometric Chart - Typical Condition Points (SI) . . . . . . . . . . . . . . . . . . . . . . . . . 2-9 Psychrometric Chart - Typical Condition Points . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-10 Sensible Heating and Cooling (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-1 1 Humidification and Dehumidification (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-12 Psychrometric Chart - Processes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-13 Cooling and Dehumidifying (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-14 Heating and Humidification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-15 Mixing of Two Airstreams (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-16 Tank Static Head . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-17 Velocity Profile . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-18 Pressure Changes During Flow in Ducts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-19 Sample Fitting Loss Coefficient Table . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-20 Pump with Static Head and Suction Head . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-21 Pump with Suction Lift . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-1 Series-Parallel Circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-2 Single-Phase AC Service . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-3 Current And Voltage-T ime Curves and Power Factor . . . . . . . . . . . . . . . . . . . . . . . . 3-4 220-Volt Three-Wire Delta Three-Phase Circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-5 220-Volt Delta Three-Phase Circuit with 110-V olt Single-Phase Supply . . . . . . . 3-6 120/208-Volt Four-Wire Wye Circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-7 Transformer with TaPped Secondary . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-8 Typical Performance of Standard Squirrel Cage Induction Motors . . . . . . . . . . . . . 3-9 Interlocked Starters with Control Transformers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-10 VFD Added to Existing Air Handling Unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-1 Valve Throttling Characteristic Comparison . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-2 ATC Valve Arrangements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-3 Typical Multiblade Dampers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-4 Desktop Computer Displaying Status of Building HVAC Systems . . . . . . . . . . . . . . 4-5 Functional Block Diagram A Centralized Computer Control System . . . . . . . . . . . 4-6 HVAC Controls Panel with Original Pneumatic Controls. . . . . . . . . . . . . . . . . . . . . . 4-7 The Same HVAC Control Panel After Upgrading to Direct Digital Control (DDC). 4-8 Portable Computer Plugged Into Electronic Wall Thermostat During System Balancing. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-1 Centrifugal Fan Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-2 Characteristic Curves for FC Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-3 Characteristic Curves for BI Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-4 Characteristic Curves for Air Foil . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-5 Axial Fan Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-6 Characteristic Curves for Propeller Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-7 Characteristic Curves for Vaneaxial Fans (High Performance) . . . . . . . . . . . . . . . . 5-8 Tubular Centrifugal Fan . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-9 Characteristic Curves for Tubular Centrifugal Fans . . . . . . . . . . . . . . . . . . . . . . . . . 5-10 Fan Class Standards (I-P) (SW BI Fans) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-1 1 Fan Class Standards (SI) (SW BI Fans) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-12 Drive Arrangements For Centrifugal Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-13 Arrangement 1 In-Line Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-14 Arrangement 4 in-line fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-15 Arrangement 9 in-Line fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-16 Centrifugal Fan Motor Locations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-17 Direction of Rotation And Discharge . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xii

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.2 2.2 2.3 2.3 2.4 2.4 2.9 2.10 2.11 2.12 2.13 2.14 2.15 2.15 2.17 2.20 2.21 2.22 2.24 2.28 2.29 3.2 3.2 3.3 3.4 3.4 3.4 3.5 3.7 3.9 3.10 4.3 4.4 4.4 4.7 4.8 4.9 4.10 4.10 5.1 5.1 5.2 5.2 5.2 5.3 5.3 5.3 5.4 5.4 5.4 5.5 5.8 5.9 5.9 5.10 5.11


FIGURES (continued) 5-18 Fan Total Pressure (TP) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-19 Fan Static Pressure (SP) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-20 Fan Velocity Pressure (VP) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-21 Tip Speed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-22 System Resistance Curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-23 Operating Point . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-24 Variations from Design Air Shortage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-25 Fan Law - RPM Change . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-26 Effect of Density Change (Constant Volume) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-27 Effect of Density Change (Constant Static Pressure) . . . . . . . . . . . . . . . . . . . . . . . . 5-28 AMCA Fan Test - Pitot Tube . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-29 Effect of Density Change (Constant Mass Flow) . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-30 Effects of System Effect . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-31 Fan Outlet Effective Duct Length . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-32 Non-Uniform Flow Conditions Into Fan Inlet . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-1 Constant Volume Fan-Powered Box . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-2 Bypass-Type Fan-Powered Box . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-3 Multiblade Volume Dampers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-4 Flow Characteristics for a Parallel Operating Damper . . . . . . . . . . . . . . . . . . . . . . . 6-5 Flow Characteristics for an Opposed Operating Damper . . . . . . . . . . . . . . . . . . . . . 6-6 Volume Dampers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-7 Surface (Coanda) Effect . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-8 Some Elements Affecting Body Heat Loss . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-9 Four Zones in Jet Expansion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-10 Typical Supply Outlets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-1 Single Duct System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-2 Typical Equipment for Single Zone Duct System . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-3 Variable Air Volume (VAV) System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-4 Terminal Reheat System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-5 Induction Reheat System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-6 Dual Duct High Velocity System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-7 Multi-Zone System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-8 System Layout (I-P Units) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-9 System Layout (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-10 Fan Duct Connections . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-1 Typical Centrifugal Pump Cross Section . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-2 Descriptions of Centrifugal Pumps Used in Hydronic Systems . . . . . . . . . . . . . . . . 8-3 Coupling Alignment with Straight Edge . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-4 Typical Required NPSH Curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-5 Pump Curve for 1750 rpm Operation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-6 Typical Design Pump Selection Point (from Abbreviated Curve) . . . . . . . . . . . . . . 8-7 System Curve Plotted on Pump Curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-8 Typical Open Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-9 Typical Cooling Tower Application . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-10 System Curve for Open Circuit False Operating Point . . . . . . . . . . . . . . . . . . . . . . . 8-1 1 System Curve for Open Circuit True Operating Point . . . . . . . . . . . . . . . . . . . . . . . . 8-12 Pump Operating Points . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-13 Multiple Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-14 Pump and System Curves for Parallel Pumping . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-15 Pump and System Curves for Series Pumping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-16 Gage Location . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-17 Relative Gage Elevations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-18 Effect of Viscosity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-1 A Series Loop System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-2 A One-Pipe System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-3 Direct Return Two-Pipe System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-4 Reverse Return Two-Pipe System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-5 Example of Primary and Secondary Pumping Circuits . . . . . . . . . . . . . . . . . . . . . . . 9-6 Return Mix System Room Unit Controls . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.12 5.12 5.13 5.13 5.14 5.14 5.15 5.15 5.16 5.17 5.18 5.18 5.19 5.20 5.20 6.2 6.3 6.4 6.5 6.6 6.7 6.8 6.10 6.11 6.12 7.3 7.3 7.4 7.5 7.6 7.7 7.8 7.11 7.12 7.14 8.1 8.2 8.4 8.6 8.7 8.8 8.8 8.9 8.9 8.9 8.10 8.10 8.11 8.11 8.12 8.12 8.12 8.13 9.2 9.2 9.3 9.3 9.4 9.5 xiii


FIGURES (continued) 9-7 Four Pipe System Room Unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-8 Boiler Piping for a Multiple-Zone, Multiple-Purpose Heating System . . . . . . . . . . 9-9 Water Cooled Condenser Connections for City Water . . . . . . . . . . . . . . . . . . . . . . . 9-10 Cooling Tower Piping System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-1 1 Basic Piping Circuits for Gravity Flow of Condensate . . . . . . . . . . . . . . . . . . . . . . . 9-12 Basic Piping Circuits for Mechanical Return Systems . . . . . . . . . . . . . . . . . . . . . . . 9-13 Typical Two-Pipe Vacuum Steam System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-14 Thermostatic Trap . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-15 Inverter Bucket Trap . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-16 Float and Thermostatic Trap . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-17 Typical Connections to Finned Tube Heating Coils . . . . . . . . . . . . . . . . . . . . . . . . . . 10-1 Refrigerant Cycle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-2 Locations of Thermal Bulbs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-1 U-T ube Manometer Equipped with Over-Pressure Traps . . . . . . . . . . . . . . . . . . . . 11-2 Inclined-V ertical Manometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-3 Electronic/Multi-meter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-4 Pitot Tube Connections . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-5 Pitot Tube . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-6 Magnehelic Gage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-7 Rotating Vane Anemometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-8 Electronic Analog Rotating Vane Anemometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-9 Deflecting Vane Anemometer Set . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-10 Thermal Anemometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-1 1 Flow Measuring Hood . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-12 Calibrated Pressure Gages . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-13 Single Gage Being Used to Measure a Differential Pressure . . . . . . . . . . . . . . . . 11-14 Single Gage Being Used to Measure a Differential Pressure . . . . . . . . . . . . . . . . 11-15 Differential Pressure Gage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-16 Chronometric Tachometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-17 Digital Optical Tachometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-18 Digital Contact Tachometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-19 Stroboscope . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-20 Multi-range, Dual Function (Optical/Contact Tachometer) . . . . . . . . . . . . . . . . . . 11-21 Glass Tube Thermometers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-22 Dial Thermometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-23 Thermocouple . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-24 Thermistor Thermometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-24 Infrared Digital Thermometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-26 Resistance Temperature Detector . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-27 Electronic Thermometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-28 Sling Psychrometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-29 Digital Psychrometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-30 Thermohygrometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-31 Clamp-on Volt Ammeter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-32 Accessing Automation System with Laptop Computer . . . . . . . . . . . . . . . . . . . . . . 11-33 Orifice as a Measuring Device . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-34 Flow Meter Types . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-35 Annular Flow Indicator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-36 Calibrated Balancing Valve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12-1 Schematic Duct System Layout . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12-2 Instruments Selected for a Specific Job . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13-1 Sample Supply Air Duct (Part) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13-2 Typical Air Diffuser CFM Measurement . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13-3 Measuring Exhaust Air Velocity on Lab Exhaust Hood with Sash Height . . . . . . 13-4 Example of Exhaust Hood Air Balance Label . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13-5 Sample Dust Collection Exhaust System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

xiv

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

9.5 9.7 9.12 9.13 9.14 9.16 9.16 9.17 9.17 9.18 9.18 10.2 10.5 11.1 11.2 11.2 11.3 11.4 11.5 11.6 11.6 11.7 11.7 11.8 11.10 11.12 11.12 11.13 11.14 11.14 11.15 11.15 11.15 11.18 11.18 11.19 11.19 11.19 11.20 11.20 11.22 11.22 11.23 11.23 11.24 11.25 11.26 11.26 11.26 12.3 12.5 13.4 13.6 13.7 13.8 13.9


FIGURES (continued) 14-1 Typical Variable Air Volume (VAV) System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-2 Open Loop Fan Volume Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-3 Closed Loop Fan Volume Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-4 Fan and System Curves, Constant Speed Fan . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-5 Fan and System Curves, Variable Speed Fan . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-6 Series Fan Powered VAV Unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-7 Parallel Fan Powered VAV Unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-8 Paper Strip at VAV Box Return Before Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . 14-9 Paper Strip at VAV Box After Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-10 Constant Fan VAV Box . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-1 1 Intermittent Fan VAV Box (Parallel) Cycle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-12 Multi-zone System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-13 Dual Duct System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-14 Induction Unit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15-1 Hydronic Flow Measurement . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15-2 External Ultrasonic Flow Sensor on Pipe with Insulation Removed . . . . . . . . . . . 15-3 Ultrasonic Flow Meter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15-4 Effects of Flow Variation on Heat Transfer 20F (11C) ∆t at 200F (93C) . . . 15-5 Percent Variation to Maintain 90% Terminal Heat Transfer . . . . . . . . . . . . . . . . . . 15-6 Chilled Water Terminal Flow Versus Heat Transfer . . . . . . . . . . . . . . . . . . . . . . . . . 15-7 Pump With Variable Speed Drive . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15-8 Example of Primary and Secondary Pumping Circuits . . . . . . . . . . . . . . . . . . . . . . 15-9 Summer-Winter Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-1 Duct Friction Loss Chart (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-2 Duct Friction Loss Chart (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-3 Duct Friction Loss Correction Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-4 Velocities/Velocity Pressures (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-5 Air Density Friction Chart Correction Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-6 Louver Velocity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-7 Elbow Equivalents of Tees at Various Flow Conditions . . . . . . . . . . . . . . . . . . . . . .

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

14.1 14.2 14.3 14.4 14.4 14.9 14.9 14.9 14.10 14.12 14.13 14.14 14.15 14.16 15.1 15.2 15.2 15.9 15.9 15.10 15.11 15.12 15.13 A.1 A.2 A.4 A.9 A.11 A.12 A.43

xv


THIS PAGE INTENTIONALLY LEFT BLANK

xvi

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


CHAPTER 1

INTRODUCTION


CHAPTER 1 1.1

INTRODUCTION TO TAB WORK

1.1.1

New Buildings

Testing, adjusting and balancing (TAB) of all HVAC systems in a new building is needed to complete the installation and to make the systems perform as the de− signer intended. Assuming that the system design and installation meets the comfort needs of the building occupants, good testing, adjusting and balancing of the HVAC system provides occupant comfort with minimum en− ergy input. This is extremely important in this era of rising energy costs. It is also important to make sure all factory equipment startup service has been completed before beginning any TAB work. Most specifications on new building construction usually require a factory representative to be present during the initial startup and adjustment of central boilers, chillers, large variable speed motor drives, and cooling towers. This initial equipment checkout is also usually required to activate the factory warranties and are not be part of the TAB contractor’s responsibility. After this initial startup service has been completed, the TAB contractor should be in− formed that the systems are operating properly, that all safety interlocks and protective devices are function− ing, and the systems are ready to be balanced. The Testing, Adjusting, and Balancing or TAB phase of any building construction or renovation is intended to verify that all HVAC water and air flows and pres− sures meet the design intent and equipment manufac− turer’s operating requirements. It is rare to find an HVAC system of any size that will perform completely satisfactorily without the benefit of TAB work. This is why it is necessary for the designer to specify that TAB work be part of the HVAC system installation. A sam− ple TAB specification can be found in the Appendix. Commissioning services for any new building construction or renovation are intended to verify all HVAC, lighting, plumbing, electrical, and security systems operate properly and meet performance crite− ria.

INTRODUCTION It should be made clear that the Testing, Adjusting, and Balancing (TAB) services may be the only HVAC sys− tem testing services contracted on most projects, but TAB work is not intended to be ?commissioning." Most commissioning services are completed by firms having technicians experienced with each of the indi− vidual building systems mentioned above. These firms will usually subcontract the services of an independent TAB contractor to verify HVAC system balancing as part of their more inclusive commission− ing contract. 1.1.2

There are few buildings in existence that have not ex− perienced changes in internal loads and wall reloca− tions since they were designed and built. These build− ings should have their HVAC systems rebalanced to achieve maximum operating efficiency and comfort. Many buildings require rebalancing twice each year with the seasonal change from heating to cooling or the reverse. Firms with a good TAB team have had a natural lead−in to service contracts and retrofit work because the TAB work identifies system operating deficiencies. 1.2

THE TAB TECHNICIAN/TEAM

1.2.1

The Technician

Throughout this publication, TAB technician will be used to designate the person in charge of the TAB work being done on the HVAC system discussed. It will be apparent after reading this publication and observing TAB procedures on a complicated HVAC system that the TAB technician must be a highly skilled and knowledgeable individual. This person must know the fundamentals of airflow, hydronic flow, refrigeration and electricity and be familiar with all types of HVAC temperature control and refrigeration systems. They must also know how to take pressure, temperature and flow measurements; and be able to perform effective trouble−shooting. The days of bal− ancing using a wet finger and cigarette smoke are long gone! 1.2.2

Commissioning also includes the testing of all build− ing controls for each mode of operation to verify all systems are being sequenced correctly with each other, and that all interlocks are functioning. The commis− sioning agent must document the results of each equip− ment test performed as it is completed.

Existing Buildings

The Team

There are TAB jobs that can be done by one person. However, many HVAC systems need a TAB team to complete the TAB work in a reasonable time period. It is equally important that the other members of the TAB team be trained and become knowledgeable in

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

1.1


the basic fundamentals and procedures of TAB work. Many of the local Joint Apprentice Training Programs have TAB courses, and the International Training In− stitute (ITI) has a Testing Adjusting and Balancing Bu− reau (TABB) training program.

1.3

GENERAL REQUIREMENTS

In addition to having the training to meet the demand− ing requirements of a TAB technician, a complete cali− brated set of balancing instruments is necessary to do TAB work on any commercial or institutional project.

1.2

The required instruments are detailed in Chapter 11CTAB Instruments. Sample test report forms may be found in Chapter 16CTAB Report Forms. These TAB report forms may be copied and used by SMACNA Contractors who fol− low the procedures and methods outlined in this manu− al. The forms are preceded by a description of their use. A sample outline specification has been included in the Appendix that can be used by the HVAC system de− signer to obtain a good, accurate TAB report based on the methods and procedures found in this manual.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


CHAPTER 2

HVAC FUNDAMENTALS


CHAPTER 2 2.1

HEAT FLOW

2.1.1

Introduction

Large HVAC systems must be designed by profession− al engineers who have received a high degree of educa− tion and practical experience in the fundamentals of heating, ventilating, and air conditioning. The TAB technician does not have to be an expert in the funda− mentals of HVAC systems, but must have had experi− ence in these systems and a basic knowledge of these fundamentals in order to perform a good balancing job and to understand what is happening. This chapter on HVAC fundamentals will include ba− sic thermodynamic and fluidic fundamentals that in− clude heat transfer, psychometrics, and fluid mechan− ics. 2.1.2

There are two fundamental laws of thermodynamics which can be stated in different, but equivalent ways: First Law

Energy can neither be created nor destroyed (the net increase in the energy content of a particular system in a given period is equal to the energy content of the ma− terial leaving the system, plus the work done on the system, plus the heat added to the system). 2.1.2.2

Second Law

It is impossible for a self−acting machine, unaided by any external agency, to convey heat from a body of lower temperature to one of higher temperature (heat flow always occurs from the higher temperature level to the lower temperature level). 2.1.3

scale of a thermometer, but the Celsius scale is used in the rest of the world. A typical temperature spectrum for the HVAC industry is where water freezes at 32F (0C) and boils at 212F (100C). The temperature at which a substance has no molecu− lar action is called absolute zero, which is −460F (−273C). The absolute temperature used in tempera− ture/pressure/volume calculations can be obtained by using the following equations: Equation 2-1 (I-P) R  °F  460°F Where: R  Absolutetemperature(Rankine) °F  Fahrenheittemperature

Thermodynamics

Heat is one of the several forms of energy which can be converted by various methods to, or from, energy in mechanical, chemical, electrical, and nuclear forms. Thermodynamics is the science of heat energy and its transformations to and from these other forms of energy.

2.1.2.1

HVAC FUNDAMENTALS

Units of Measurement

The intensity of heat of a substance traditionally has been measured in the United States on the Fahrenheit

Equation 2-1 (SI) K  °C  273°C Where: K  Absolutetemperature(Kelvin) °C  Celsiustemperature The quantity or amount of heat in a substance is mea− sured in British Thermal Units (Btu) which is the heat required to heat one pound of water one degree Fah− renheit. It is easy to realize that a swimming pool full of water at 95F, needs substantially more heat than a cup of water at 95F to increase each of them to 96F. In the metric system, the amount of heat required to heat one kilogram of water one degree Celsius is 4.18 kilojoules (kJ). 2.1.4

Heat Transfer

Heat flow is adding or removing heat at a given rate, which is measured in Btu per hour (Btuh) in the U.S. system and is measured in joules per second (J/s) or watts (W) in the metric system (1 J/s = 1 W). Figures 2−1 and 2−2 give examples of heat conduction, convection, and radiation, which are the three methods of heat transfer in environmental systems. Equation 2−2 may be used for heat flow through a ma− terial(s) that separates different air temperatures.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.1


Radiant heat

Heat Flow Heat flows from warmer body to cooler body by conduction

from flame

Rod heated by flame becomes hot as heat flows by conduction from one end to the other.

FIGURE 2-1 HEAT TRANSFER BY CONDUCTION AND RADIATION Q  A  U  Dt

Equation 2-2

Where: Q  Rateofheattransfer  Btuh(W) A  Areaofsurface  sq.ft.(m2) U  Coefficientofheattransfer Dt  Temperaturedifference  °F(°C) ?Delta" (  ), as used above, usually indicates a small change or difference; in this case, Dt is the tempera− ture difference. In Equation 2−2, neither Dt or Q represent heat. In the past, heat was thought to be a tangible quantity like a gallon of water, a pint of milk, or a bushel of wheat. Despite this carryover from the past, one cannot feel ?hot," as heat is not a tangible quantity. What one does feel is temperature. Temperature can be ?hot" when compared to some other reference point such as 98.6F (37C) body temperature, but one cannot feel how much heat is in an object. The amount of heat con− tained within the object varies with the object. The amount of heat contained within the object varies with the object’s temperature, mass, and substance. The

amount of heat in any given object at any given tem− perature can be calculated, but the HVAC industry does not find that a particularly useful function. What is useful is to know how fast heat is given up from that object, or the rate of heat transfer (Q) expressed in Btu per hour or watts. There also are other equations for calculating ?Q". Figures 2−3, 2−4, and 2−5 show the difference between counterflow, parallel flow, and cross−flow airstreams in coils or heat exchangers. The importance of the il− lustration in Figure 2−6 is that the final temperature at− tained by both of the mediums is affected by the direc− tion of the flow of the two different mediums at the heat transfer points.

Example 2.1 (I−P) A room (70F) has two separate walls which have an unheated space on the other side. The wall exposed to outdoors (30F) is 20’ × 8’ and has a ?U" factor of 0.12. The 24’  8’ wall exposed to the unoccupied space (55F) has a ?U" factor of 0.30. Which wall has the greatest heat loss?

Heating Coil

Steam or Hot Water in Pipes

Airflow

(A) Heat Flow by Natural Convection

(B) Heat Flow by Forced Convection

FIGURE 2-2 CONVECTION HEAT TRANSFER 2.2

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


B

A

B

A

B A

A

B A

B

A B

B

FIGURE 2-3 COUNTER FLOW AIRSTREAMS Solution Outside wall: Q  A  U  Dt  20  8  0.12  (70°F  30°F)  160  0.12  40°F  768Btuh Inside wall: Q  A  U  Dt  24  8  0.30  (70°F  55°F)  192  0.30  15°F  864Btuh (or the higher loss wall)

Example 2.1 (SI) A room (21C) has two separate walls which have an unheated space on the other side. The wall exposed to outdoors (0C) is 6 m  2.5 m and has a ?U" factor of 1.2. The 7 m  2.5 m wall exposed to the unoccupied space (12C) has a ?U" factor of 3.0. Which wall has the greatest heat loss?

Solution Outside wall: Q =MA  U  nt = 6 2.5 1.2 (21_C  0C) =M15  1.2  21C = 378 watts Inside wall: Q =MA × U × nt = 7 × 2.5 × 3.0 × (21_C − 12C) =M17.5 × 3 × 9_C = 473 watts

FIGURE 2-4 PARALLEL FLOW AIRSTREAMS

(or the higher loss wall) To illustrate the difference between temperature and the heat flow rate, and to show that a system can be bal− anced to either, assume that the airflow being supplied to different rooms is to be balanced so that each room has the same temperature reading. This procedure can be used in existing buildings when original engineer− ing calculations are not available and when the build− ing is experiencing large differences in temperatures between rooms. The TAB technician would then at− tempt to balance the system by adjusting the airflow so that each room had the designed space temperature. The other balancing procedure usually is required in new buildings where the designer has calculated the airflow rates that normally establish equal tempera− tures for each of the various room spaces. The TAB technician then balances the airstream flow rate to that scheduled for each room. The rooms should achieve the desired temperatures if the design calculations were correct. If a change in temperatures is required after occupancy, the additional balancing should not be done at no cost unless there was a provision for this extra work included in the TAB contract. In any event, the TAB technician balances to the flow rate of the me− dium, which actually is balancing to the heat flow rate that is being transferred by the medium. Balancing by heat flow and temperature are therefore not the same. In Equation 2−2 (Q = A U nt), ?U" is the variable affected by building insulation materials. Insulation ?values" can become quite confusing. They can be given per inch of thickness of material or for the actual thickness of the material, such as a ?six inch thick batt." Values can be given as conductance or resistan− ce. Chapter 25 of the 2001 ASHRAE Fundamentals

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.3


A1

A2

”C” AIRSTREAM

each component of the wall is the only way the total resistance (RT) can be obtained for a particular wall construction. Other building enclosure surfaces are treated in a similar manner (windows, floors, ceilings, etc.). ?R" values of air surface films must be used as indicated After the total resistance is obtained by adding the in− dividual resistances, the ?U" value is obtained by tak− ing its reciprocal as shown in Equation 2−4. Insulating materials play a large role in reducing the heat flow rate into or from buildings, ducts, pipes, etc. Obvious− ly, exterior surfaces having high resistances or low ?U" values will have less heat transfer between the outside environment and the interior of the structure, duct, or pipe.

EXCHANGER C2

“A” AIRSTREAM

Equation 2-3 RT =R1 +R2 +R3 +. . . +Rn

FIGURE 2-5 CROSS-FLOW AIRSTREAMS

Equation 2-4 Handbook entitled Thermal and Water Vapor Trans− mission Data contains thermal resistance values for different materials. Conductivity (k) indicates how much heat will pass through an inch of a material. Conductance (C) is a somewhat similar value, but is for a given thickness of material. Resistances (R), which are the reciprocals of ?k" or ?C", can be added sequentially for the heat flow through combinations of different materials. Conduc− tivity (k) and conductance (C) values can not be added in this manner. Therefore, addition of the resistance of

U  1 RT Where: U = Coefficient of heat transfer C Btuh/ ft2N⋅N°F (W/m2N⋅N°C) RT = Total of the resistances The values of U for the SI system are about 17.6 per− cent of the values in I−P Units, i.e., a ?U" of 1.0 in I−P Units is 0.176 in SI units.

TB (WARM)

TA (HOT)

TA (WARM)

TB (COLD) Distance Through Exchanger

TEMPERATURE

TEMPERATURE

TA (HOT)

TA (WARM)

TB (WARM)

TB (COLD) Distance Through Exchanger

FIGURE 2-6 PARALLEL AND COUNTERFLOW HEAT TRANSFER CURVES 2.4

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Example 2.2 An outside wall of a building has the following resist− ances:

Outside surface film

— 0.17

Masonry

— 1.60

Furring (air space)

— 0.94

Drywall

— 0.45

Inside surface film

— 0.68

Find the coefficient of heat transfer U for this wall.

Solution R T  R1  R2  R3   R n R T  0.17  1.6  0.94  0.45  0.68 R T  3.84 U  1  1  0.26 RT 3.84 2.1.5 EQUIPMENT HEAT FLOW Heat flow in HVAC equipment is normally from a fluid (or gas) through a thin wall into another fluid (or gas); or it is into a transfer substance which moves into or to another cooler fluid (or gas) to deposit its energy. Major factors in the transfer of heat by conduction are: a.

temperature difference

b.

size and shape of the transfer surface

c.

type of fluid (or gas) and flow velocity

d.

conductivity of heat transfer material

e.

conductivity of the boundary layer

There also are many other factors to consider such as film coefficients, fouling, corrosion, condensables, frost or freezing, and poor maintenance.

Going back to the ?laws of thermodynamics" stated earlier in this chapter, the concept can be restated that the same amount of heat that is given up by one me− dium is gained by the other, and that all heat can be ac− counted for (energy can neither be created nor de− stroyed). This is not quite true in this atomic era, but it can be used as a basic principle for this TAB work. Equation 2-5 (I-P) Hydronic systems: Q  500  gpm  Dt Where: Q  Heatflow(Btuh) gpm  Gallonsperminute(water) Dt  Temperaturedifference(°F) Equation 2-5 (SI) Q  4190  m3s  Dt or Q  4.19  Ls  Dt Where: Q  Heatflow(kW) m 3s  Cubicmeterspersecond(water) Ls  Literspersecond(water) Dt  Temperaturedifference(°C) Note that in Equation 2−5 the 500 (4190 or 4.19) is a ?constant" that is used specifically for water. This constant will change when the system medium is other than water, such as a glycol mixture, steam, refriger− ant, or air. In fact, the comparable equation for sensible heat flow of air is shown below. Equation 2-6 (I-P) Air systems: Q  1.08  cfm  Dt Where: Q  Sensibleheatflow(Btuh) cfm  Cubicfeetperminute(air) Dt  Temperaturedifference(°F)

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.5


Equation 2-6 (SI) Q  1.23  Ls  Dt Where: Q  Sensibleheatflow(watts) Ls  Literspersecond(air) Dt  Temperaturedifference(°C) Sensible heat is defined as the heat associated with temperature differences only as measured by a ther− mometer. This is not affected by the method of heat transfer, such as radiation, convection, and conduc− tion. Transmission heat gains are those which occur through conduction of the heat through a surface as a wall. Convection has been taken into account by film coefficients on each side of the wallCwhether inside or outside. Example 2.3 (I−P) 30 gpm of water at 200F circulates through a heating coil. If 4000 cfm of air increases in temperature from 50F to 120F, determine the leaving temperature of the water. Solution Q  1.08  cfm  Dt  1.08  4000  (120°F  50°F)  302, 400Btuh  500  gpm  Dt Q 302, 400 Dt    500  gpm 500  30  20.16°F  (T1  T2) T2%  T1MM∆tNN200F ON20.16F N179.84F Water with a temperature of 180F is leaving the coil. Example 2.3 (SI) 2 L/s of 93F water circulates through a heating coil. If 2000 L/s of air increases in temperature from 10F to 50C, determine the leaving temperature of the wa− ter. Solution Q  1.23  Ls  Dt  1.23  2000  (50°C  10°C)  98, 400watts(98.4kW)  4.19  Ls  Dt Q   98.4kW 4.19  Ls 4.19  2Ls ∆tPM11.74CMM(T1 M–MT2 ) 2.6

T2 PMT1M – ∆tMM93CM–M11.74CM=M81.26C Water with a temperature of 81.3C is leaving the coil. 2.2

PSYCHROMETRICS

2.2.1

Introduction

Psychrometrics is a study of the thermodynamic prop− erties of moist air and the application of these proper− ties to the environment and environmental systems. Thermodynamics previously has been defined as the science of heat energy and its transformation, or change, from one form of energy to another. Since air is the final environment and one of the major fluids of the systems, whatever affects air affects the systems and the environment. Whatever happens to the air and the moisture it contains, under both natural circumstances and artificial conditions imposed by the systems and the environment, is of concern to the TAB technician. The language of psychrometrics, and to be able to use psychrometric charts and tables as tools to change existing conditions to those desired or re− quired, is a requirement for a good TAB technician. 2.2.2

Properties of Air

Dry air is an unequal mixture of gases consisting prin− cipally of nitrogen, oxygen, and small amounts of neon, helium, and argon. The percentage of each gas normally will be the same from sample to sample, al− though carbon dioxide and pollutants might be present in varying quantities. Air in our atmosphere, however, is not dry but contains small amounts of moisture in the form of water vapor, and the percentage may vary from sample to sample. This air−water mixture is the moist air referred to in the subject of psychrometrics. The amount of water vapor in atmospheric air normally represents less than 1 per− cent of the weight of the moist air mixture. If the aver− age weight of air is approximately 0.075 pounds per cubic foot (1.204 kg/m3), the moisture contained therein will weight less than 0.00075 pounds per cubic foot (0.012 kg/m3). This would seem to be an insignificant amount to cause so much concern. Normally, atmospheric air contains only a portion of the water it is able to absorb (partially saturated). If the proper conditions occur, air will absorb additional moisture until it can absorb no more, and it is then saturated. Air and water vapor be− have as though the other were not present. Each will act as an independent gas and exert the same pressure as if it were alone. The barometric pressure is the sum

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


of the two pressuresCthe partial pressure of the air, plus the partial pressure of the water vapor. The dry air component of the moist air exists only as a gas under all environmental conditions. it cannot be liquified by pressure alone, and therefore acts as a per− fect gas. However, water vapor does coexist with water as a liquid at all environmental temperatures, and it can be liquified by pressure. As it is not a perfect gas, the properties of water vapor are determined experi− mentally. 2.2.3

Air-V apor Relationship

Throughout the normal ranges of atmospheric pres− sures and temperatures, the air and water vapor mix− ture behaves as a perfect gas provided no condensation or evaporation takes place. A perfect gas is one where the relationship of pressure, temperature, and volume may be defined and predicted by Equation 2−7. Equation 2-7 PV  R T

T  70°F(21°C) P  14.696psi(101.325kPa)] V  13.33ft 3lb(0.831m3kg) d  0.075lbft 3(1.204kgm3) Example 2.4 (I−P) Using Equation 2−9, any condition other than standard may be calculated if one of the final conditions is known. For example, if one pound of standard air was heated to 700F, as in a process application, find the new volume of air per pound. Solution VT V 2  1 2  13.33  460  700 T1 460  70 1160  13.33   29.2ft 3lb 530 Similar calculations may be made for any value of temperature so long as the pressure remains constant. Example 2.4 (SI) If one kilogram of standard air is heated to 370C, find the new volume of air per kilogram.

Where: P'=MAbsoluteM pressureMMlb/Mft 2M(kPa) V'=MVolumeMMft3M(m3) T'=MAbsoluteMtemperature 460M + FM(273M+MC) R'=MGas constant The actual value of R has little meaning here, but the fact that R remains constant for any given perfect gas is extremely important. It is possible to equate the P, V, T values for two different conditions for the same gas: Equation 2-8 P 1V 1 PV  2 2 T1 T2

Solution 0.831  (273°C  370°C) V 1T 2  T1 (273°C  21°C) 8  0.831  643°  1.817m 3kg 294°

V2 

Example 2.5 (I−P Units) It is possible to make similar calculations which in− clude pressure variations. An example might be to find the correct volume for an altitude of 5000 feet. Assume that the temperature is the same at both points for con− venience.

If it is assumed that atmospheric pressure remains es− sentially constant at a given elevation on earth, then: Equation 2-9 VT V2  1 2 T1 Using Standard Air, which is the fixed reference for air conditioning calculations, the following properties oc− cur:

Solution P 1V 1 PV  2 2 , T 1  T 2 and P1V 1  P 2V2 T1 T2 P orV 2  V1 1 P2 The standard atmospheric pressure at 5000 feet is 24.90 inches of mercury (see Table A−16).

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.7


V 2  V 1

P1  13.33  29.92  16.02ft 3lb P2 24.90

Referring to the sample charts in Figures 2−7 and 2−8, note that the groups or families of curves have been la− beled to indicate which part or parts of the graph are for the different properties of most air. Definitions of the properties may be found in the Glossary.

Example 2.5 (SI) 2.2.4.1 Find the correct volume for an altitude of 1500 meters (T1 = T2 ) at standard conditions.

Solution V P 0.831  101.3 V2  1 1  P2 95.1(fromTableA  17)  0.989m3kg The conclusion to be reached from these examples is that large deviations from standard air temperature and pressure values require corrections to the calculations, while relatively small variations may be ignored. In the case of pressure, a correction normally is not used in applications below 2000 feet (600 m) above sea le− vel. Above 2000 feet (600 m) it becomes necessary to make the correction, since air has a significant reduc− tion in its ability to carry heat. By a series of calcula− tions and laboratory measurements, a long list of val− ues are obtained and are listed in Air Density and Correction Factor Tables in Appendix ACEngineer− ing Data, Tables and Charts. To be meaningful, these values must have a reference, which has been stated to be standard atmospheric pressure at sea level: 29.92 inches of mercury (101.325 kPa). Since the variations caused by pressure are not serious below an altitude of 2000 feet (600 m), the values obtained by maintaining the 29.92 inches of mercury (101.325 kPa) are ade− quate for use with HVAC systems. However, as seen in Example 2−4, temperature varia− tions in process systems can cause wide variations in air density: density = 1/29.2 = 0.034 lb/ft3, or a reduc− tion of 54 percent. For this reason, many engineering catalogs use standard cfm (scfm) which is ?cfm" cor− rected to ?standard air" conditions. Similar conditions apply to the SI system. 2.2.4

Using the Psychrometric Chart

A psychrometric chart is a series of graphs or curves arranged in such a way that, by knowing a specific val− ue of each of two different properties, a point can be obtained that will determine the values for all other properties under the same conditions. Charts are avail− able for different elevations above sea level, higher or lower temperatures, and many other variations. 2.8

Basic Grid, Humidity Ratio (Specify humidity)

The horizontal parallel and equidistant grid lines (with the scale displayed along the right side of the chart) in− dicate the grains of moisture per pound of dry air or pounds of moisture per pound of dry air. The bottom line of the chart is zero humidity and represents totally dry air. The chemical industry prefers to use mol−ra− tios, and grams of moisture per kilogram of dry air is used in the SI system. 2.2.4.2

Enthalpy (Total Heat)

Enthalpy lines are slanted from the top−left to the bot− tom−right. Enthalpy is designated by the letter ?h" and, as all values on the chart, is referred to a pound (kilo− gram) of dry air. This is the only value which does not change through various processes, such as heating− cooling, compression−expansion, and humidification− dehumidification. Other constant value lines are su− perimposed upon this basic grid by plotting and are neither equidistant nor parallel, although they may ap− pear to be so. 2.2.4.3

Dry Bulb Temperature

Constant dry bulb temperature lines on the chart are nearly vertical. They diverge slightly towards the top of the chart. The scale is at the bottom, along the dry air line (zero humidity ratio) from left to right, and val− ues are given in F in Inch−Pound units, English or cus− tomary, and in C in SI Units, the International System of Units, Metric. 2.2.4.4

Saturation Line, Dew Point and Wet Bulb

The curved upper borderline of the chart, which runs from the bottom left to the mid−top, is an experimen− tally plotted saturation line. It shows the maximum amount of water vapor (in pounds or kilograms) which can be associated with a pound (kilogram) of dry air at a given dry bulb temperature. Air is said to be saturated with moisture at this point. The temperature at this point is known as the saturation temperature. It can be seen from the saturation line that at higher temperatures air can hold more moisture than at lower temperatures. Conversely, the capacity to hold mois−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Pounds of moisture per pound of dry air Enthalpy at saturation Btu per pound of dry air Grains of moisture per pound of dry air 85 90 95 100 105 .35 180 .025 170 .024 160

150

.023

.021 140

130

.45

.020 .019 .018

120

.40

.022

.50

.55

.017 .60

110

.016 .65 .015

100 .014 90

80

.013 .012

.70 .75 .80 .85 .90 .95

.011 70

60

.010

Sensible heat factor

.009 .008

50 .007 40

.006 .005

30 .004 20

80% 60%

10

40%

0 25

30

.002 .001

20% 20

.003

35

40

45

50

55

60

65

70

75

80

85

90

95

100

105

0 110

Wet-Bulb, Dewpoint or Saturation Temperature F

FIGURE 2-7 PSYCHROMETRIC CHART (I-P) ture is less at lower temperatures. Therefore, if the al− ready saturated moist air is being cooled, the excess moisture instantly begins to separate from the moist air by condensation either in the form of fog (the left side of the saturation curve is the fog area) or as dew if it condenses on a cold surface.

The point of saturation (saturation temperature) is also called the dew point. At this point the saturation tem− perature, the dry bulb temperature and the wet bulb temperature are all the same value. The saturation temperature values (also dew point or wet bulb tem− peratures) are shown in F (C) along the saturation line where it is crossed by the same value dry bulb lines.

2.2.4.5

Specific Volume

Specific volume lines run at a steep angle from top left to bottom right. The numerical values, along the bot− tom of the chart at the end of these lines, are given in cubic feet per pound of dry air in I−P Units (and in cu− bic meters per kilogram in the SI System).

The slant and spacing of the specific volume lines show that moist air becomes lighter (by expansion) with an increase in temperature and with an increase in moisture content (moist air is lighter than dry air at the same temperature). The specific volume of dry air can be found at the intersection with the bottom, zero humidity ratio line. The volume of the water vapor can be found by subtracting the volume of the dry air from the volume of the moist air. 2.2.4.6

Relative Humidity

The curves of constant relative humidity (RH) lie be− tween the zero humidity ratio line at the bottom and the curved saturation line above. The curvature decreases as curves approach the dry air line. Relative humidity expresses the proximity of the sub− ject moist air to that of saturated air at the same tem− perature. The saturation line represents 100 percent RH and the bottom line of the chart is 0 percent RH. Another term used to define proximity of the moist air to saturation is the degree of saturation which is the ra− tio of the humidity ratio of the moist air to that of the saturated moist air at the same temperature and pres−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.9


30

28 30

26

H H

24

SENSIBLE HEAT TOTAL HEAT

25

=

22

HUMIDITY RATIO (w) GRAMS MOISTURE PER KILOGRAM DRY AIR

20

20

15

18

16

14

12 0.85 0.90

10

0.95 1.0

8

10

6 5

4

2

0

20

30

40

50

DRY BULB TEMPERATURE C

FIGURE 2-8 PSYCHROMETRIC CHART - TYPICAL CONDITION POINTS (SI) sure. The difference between both meanings is small but still noticeable within the comfort conditions. While the degree of saturation of 50 percent lies on the dry bulb temperature line directly in the middle be− tween 0 percent and 100 percent, the 50 percent RH point is slightly below the mid point. Measurement of relative humidity depends on changes in humidity responsive materials. While low cost and quite accurate, these instruments require frequent cal− ibration. 2.2.4.7

Wet Bulb Temperature

The wet bulb method was developed to obtain a more practical way to measure relative humidity and de− pends on the evaporation of water around the bulb of a mercury thermometer. Since evaporation and the de− pression of the wet bulb temperature depends on the relative humidity of the moist air, it can be used to measure relative humidity by use of conversion tables.

bulb temperatures are equal at this point and are also equal to the dew point temperature (saturation temper− ature). As the wet bulb lines run so close to the enthal− py lines, most psychrometric charts use the same lines for both wet bulb temperatures and enthalpy and show correction of either enthalpy values or wet bulb values by another family of curves. 2.2.4.8

Plotting Conditions

Consider point A, in Figure 2−9, representing summer outdoor design conditions. By finding the point which represents 95F DB and 78F WB, the values for the other properties are:

Dew point temperature = 71.98F Relative humidity = 48% Enthalpy (total heat) = 41.6 Btu/lb dry air Moisture content = 118 grains/lb dry air

At the saturation point, there is no evaporation. Since the wet bulb depression is zero, the wet bulb and dry 2.10

or 0.0169 lb moisture/lb dry air.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Enthalpy at saturation Btu per pound of dry air 85

90

Grains of moisture per pound of dry air 95

100

Pounds of moisture per pound of dry air

105 .35

180 .025

170 .024

.023

160

.40

.022 150

41.6 BTU/Lb

78F WB

.021

POINT “A”

.45

140

.020

.019

.50

130 .018 .55 120

.017 .60 .016

110

118 GR/Lb

.65 .015

71.9F DP

.70

100 .014

.75

48% RH

62.7F WB

.0169 Lb/Lb

.013

90

.80 .85

.012

.90 .95

80

28.3 BTU/Lb

.011

70

.010

.009

Sensible heat factor

60

POINT “B”

65 GR/Lb

.008

50 .007

50% RH

55F DP

.006 40

.0093 Lb/Lb

.005 30 .004

.003

20 80%

Wet-Bulb Dewpoint or Saturation Temperature F

60%

.

40%

95F DB

75F DB

.002 10 .001

20% 0

Dry-Bulb F

20

25

30

35

40

45

50

55

60

65

70

75

80

85

90

95

100

105

0 110

FIGURE 2-9 PSYCHROMETRIC CHART - TYPICAL CONDITION POINTS Now consider point B in Figure 2−9 representing sum− mer indoor design conditions. By finding the point which represents 75F DB and 50 percent RH, the val− ues for the other properties are: Dew point temperature = 55F Wet bulb temperature = 62.7F Enthalpy (total heat) = 28.3 Btu/lb dry air Moisture content = 65 grains/lb dry air or 0.0093 lb moisture/lb dry air.

a change are the environmental systems. The designer, by knowing what design conditions must be satisfied, but without knowing the airflow rate, is able to plot the various changes in different portions of the system and the environment. The designer then is able to deter− mine what systems are capable of accomplishing the necessary results. In addition, the heat values are used in the design calculations to see immediately if the de− sign conditions are practical or impossible. But first, the various condition changes must be illus− trated and understood. For this purpose, skeleton psychrometric charts have been used in the related dia− grams, alternating between I−P Units and SI units. 2.2.5.1

In this way, any condition may be plotted on the chart for normal environmental systems. Other charts are available for plotting from tabular data for conditions not found here, but the need for such variations is not common. 2.2.5

Condition Changes on Psychrometric Charts

Using the chart, simple logic leads to the conclusion that there must be some way to change the properties of the air, initially at the condition of point A, to the conditions of point B. The means to accomplish such

Sensible Changes

Sensible heat, by definition, indicates only dry bulb temperatures changes. Therefore, any heating system not including humidification is a sensible heat process or system and is represented on the chart as a horizon− tal straight line. Conversely, any cooling system utiliz− ing a dry coil which does not dehumidify or where the surface of the coil does not fall to or below the air dew point is also a horizontal straight line. a.

Heating

Assume that air which is at 70F and 20 percent RH is heated to 105F. This process is indicated in Figure

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.11


2−10. Conditions other than those mentioned here have been determined from the psychrometric chart.

a change in latent heat. This process is indicated on the sample metric chart in Figure 2−11. Humidification and dehumidification are defined as the addition or subtraction of moisture from air. Since each is a change of state from liquid to gas and gas to liquid, each occurs at a constant dry bulb temperature, but a varying wet bulb temperature. Note that this is the same process as the addition or subtraction of latent heat and is also the same vertical line on the chart in Figure 2−11 at a constant dry bulb temperature. Hu− midification and dehumidification are both latent heat processes, and both are shown on the same chart.

105F DB 8% RH 20F DP 64F WB

70F DB 20% RH 28F DP 50F WB

HEATING

COOLING

FIGURE 2-10 SENSIBLE HEATING AND COOLING (I-P)

In this example, the only constant value is the dry bulb temperature; all other properties increase for humidifi− cation and decrease for dehumidification. Note that this process is essentially an illustration and cannot normally be reproduced in environmental systems. 2.2.5.3

By considering the values, it may be seen that the dew point has not changed and the total moisture content in grains/lb has not changed. The dry bulb temperature has gone up, the heat content has gone up, the wet bulb temperature has gone up, and the ability of the air to absorb moisture has gone up as indicated by the de− crease in relative humidity. The example is theoretical because moisture from people, cooking, infiltration, etc., will make some contribution to the moisture in the air. However, this is the design diagram for heating without deliberate humidification. b.

Cooling

Now consider the process and ignore the extraneous moisture sources noted above. In ideal conditions, air gives up its heat along the same line to maintain the oc− cupied spaces at the given 70F DB and 20 percent RH. A cooling coil, selected to cool air from 105F DB and 20F DP to 70F DB would also produce condi− tions along the same line as long as the coil surface temperature was above 20F. In most systems, this would be impractical if not impossible, since the coil would immediately clog with frost. 2.2.5.2

Latent Changes

Latent heat, by definition, involves a change of state to or from a fluid; and in the case of air, this means the addition or removal of moisture. It must be remem− bered that the dry bulb temperature does not change during this addition or removal of moisture. Therefore, a vertical line on the psychrometric chart between any two points at constant dry bulb temperature represents 2.12

Combination Changes

Combination sensible−latent heat processes are the rule in most systems. The addition or subtraction of la− tent and sensible heat appears as a combination pro− cess with all changes occurring simultaneously. The result is neither a horizontal nor vertical line but a slanted one tilted in the direction dictated by process. Referring to Figure 2−12, consider the general rules be− low based on the two endpoints of the process; the first being the initial condition of the air; and the second be− ing the final condition after the process or a portion of the air treatment has been completed. On the chart, all processes have the same initial point, and the arrow point indicates each arbitrary final point: Sensible heating is a horizontal line from left to right. Sensible cooling is a horizontal line from right to left. Humidification is a vertical line upward. Dehumidification is a vertical line downward. Heating and humidification is a line sloping upward to the right. Cooling and dehumidification is a line sloping down− ward to the left. Evaporative cooling is a line sloping upward to the left. Chemical dehydration or dehumidification is a line sloping downward to the right. Now consider the specific cooling−dehumidifying ex− ample which might represent the conditions obtained

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Humidification—Addition of Moisture and Latent Heat by Evaporation of Steam Injection

Dehumidification— Removal of Moisture and Latent Heat by Condensation

41C DB 28.2C WB 24C DP 38% RH 90 kJ/kg 18.9 g/kg 41C DB 20.1C WB 6.8C DP 13% RH 58.9 kJ/kg 6.2 g/kg

FIGURE 2-11 HUMIDIFICATION AND DEHUMIDIFICATION (I-P) from a cooling coil using 100 percent recirculated air. Assume that the entering air and room conditions are 80F DB and 50 percent RH. Also assume that the cooling coil can produce leaving conditions of the air at 55F DB and 54F WB. The illustration of the pro− cess is shown in Figure 2−13. The line drawn between the initial and final conditions represents the change of air properties produced by the cooling coil and is conveniently drawn straight. In ac− tual fact, the process follows a curve, but the deviation is not usually important to the system analysis. To the coil designer, however, the curvature is critical, since it indicates the heat transfer conditions from point−to− point through the coil depth. The amount of moisture that was removed from the air (condensed on the coil) was 16 grains/lb of dry air (77T–T61T=T16). The reverse operation, heating and humidifying, could be explained by working the cooling−dehumidifying diagram backwards. However, a more practical ap− plication may be obtained by using a new diagram. As− suming that 21C DB, 40 percent RH air returns to a heating coil and that a humidifier has been added, the

combination process, assuming the required leaving conditions to be 41C DB and 38 percent RH, is illus− trated in Figure 2−14. 12.7 grams of moisture per kilo− gram of dry air was added in the process (18.9T–T6.2T=T 12.7) Equation 2−10 is used for the change in the total heat content of the air, including the moisture content. Equation 2-10 (I-P) Q(Total)  4.5  cfm  Dh Where: Q  Totalheatflow(Btuh) cfm  Airflow Dh  Enthalpydifference(Btuhlbdryair) Equation 2-10 (SI) Q(Total)  1.2  Ls  Dh Where: Q  Totalheatflow Ls  Airflow Dh  Enthalpydifference(kJkgdryair)

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.13


49

Pounds of moisture per pound of dry air

Grains of moisture per pound of dry air

48

Enthalpt at saturation Btu per pound of dry air

90

100

95

105

47

85

.35

46

180

45

.025 170

43

44

.024

.023

.40

42

80

160

41

.022 150 .45

39

40

.021

.020

.019

75

37

38

-0.1 Btu

140

.50

130

-0.2 Btu

.018 .55 120

Heating & Humidification

Evaporative Cooling

.017 .60 .016

-0.3 Btu

70

33

34

90% Relative Humidity

35

36

Humidification

110 .65 .015 .70

80% 30

100

65

28 27

26 25

55

50% -.06 Btu

23 22

Sensible Cooling

.80 .85

.012

.90

.010

Sensible heat factor

.95

80

70

.009 -.08 Btu

21

.75

90

Sensible Heating

-.02 Btu

24

60% 60

20

18

Common Initial Reference

70%

Enthalpy deviation Btu per pound of dry air

29

.014

60

40%

18

17

.008 50

16

50 .007

11

13

40

20%

35

9

12

10

Chemical Dehydration or Dehumidification

Cooling & Dehumidification Dehumidification

14

12

15

30% 45

30

.006 40

.005 30 .004

40.1 Btu

.003

8

20 40.2 Btu

.002

60%

Saturation Temperature F

40.3 Btu

40%

10 40.4 Btu

.001

20% 40.5 Btu

0 60

55

65

70

80

75

85

90

100

95

105

0 110

14.0 cu ft

50

13.5 cu ft

45

13.0 cu ft

40

35

30

12.5 cu ft

20

25

14.5 cu ft per pound of dry air

Wet-Bulb, Dewpoint or

10%

80%

7

25

FIGURE 2-12 PSYCHROMETRIC CHART - PROCESSES Psychrometric charts base all the information given in content per pound (kilogram) of dry air. The standard equations are derived from these values and give a very close approximation of the actual calculation if all the conditions would have been worked out using the basic figures on the psychrometric chart. From any two given points on a psychrometric chart, the Btuh (watts) obtained for enthalpy is always equal to or greater than the Btuh obtained for sensible heat only. The reason for this is that the moisture contained in the air has heat content.

b. Locate the final condition on the chart as Point ?B". The wet bulb temperature is 62.7F and the en− thalpy is approximately 28.3 Btu per pound of dry air. c.

The decrease in enthalpy is: h = 41.6 − 28.3 = 13.3 Btu per pound

d.

Q (Total)

= 4.5  10,000  13.3 Q (Total) e.

Solution a. Locate the initial condition on the psychrometric chart, as Point ?A" in Figure 2−9. The corresponding wet bulb temperature is 78F, and the enthalpy is approximately 41.6 Btu per pound of dry air. 2.14

= 598,500 Btuh

Tonsofrefrigeration  Btuh  Btuhton 12, 000 

Example 2.6 (I−P) Air at 95F DB and 48 percent RH enters a cooling coil at a rate of 10,000 cfm. If the air is cooled to a condi− tion of 75F DB and 50 percent RH, find the cooling load in Btuh, and in tons of refrigeration.

= 4.5  cfm  h

598, 500  49.9tons 12, 000

Example 2.6 (SI) Air at 36C DB and 50 percent RH enters a coil at a rate of 5000 L/s. If the air is cooled to a condition of 24C and 50 percent RH, find the cooling load in watts and kilowatts (use Figure 2−8). Solution a. The enthalpy for 36C DB and 50 percent RH is approximately 85.3 kJ/kg.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


AIR ENTERING 80F DB 67F WB 50% RH

AIR LEAVING 55F DB 54F WB 94% RH

FIGURE 2-13 COOLING AND DEHUMIDIFYING (I-P)

AIR LEAVING

AIR ENTERING

21C DB 40% RH 6.8C DP 13.1C WB 37.1 kJ/kg 6.2 g/kg

41C DB 38% RH 28.2C WB 24C DP 90 kJ/kg 18.9 g/kg

FIGURE 2-14 HEATING AND HUMIDIFICATION

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.15


b. The enthalpy for 24C and 50 percent RH is approximately 48.0 kJ/kg. c.

The decrease in enthalpy is: ∆h = 85.3 − 48.0 = 37.3 kJ/kg

d.

Q (Total) = 1.2 x L/s × h = 1.2 × 5000 × 37.3 Q (Total) = 223,800 (223.8 kW)

It is sometimes necessary to calculate the weight of the air. This is shown in the solution to Example 2.7 (I−P). The average person is not used to thinking of air as having weight. Air is a fluid and has weight just as wa− ter is a fluid and has weight. As a reminder, ?Standard air," at 0.075 pounds per cubic foot (1.204 kg/m3) is very much lighter than water at 62.4 pounds per cubic foot (1000 kg/m3). The specific volume of standard air is the reciprocal: 1  0.075  13.33 cubic feet per pound 1/0.075 = 13.33 cubic feet per pound (1/1.204 = 0.8305 m3/kg).

Solution 10, 000cuftmin  750.2poundsperminute 13.33cuftlb 750.2  60 minutes = 45,012 pounds per hour 45,012  13.3 (∆h) = 598,660 Btuh 598,660/12,000 = 49.9 tons The total heat content of air also can be taken from tables when the wet bulb temperature is known.

Example 2.7 (SI) Obtain the solution to Example 2.6 (SI) using the weight of the airflow volume.

Solution 5.0m 3s  6.02kgs 0.83m 3kg 6.02kgs  37.3kJkg  224.5kJs

Example 2.7 (I−P) 1W  1Js, so Obtain the solution to Example 2.6 (I−P) using the weight of the airflow volume.

2.16

224.5kJs  224.5kW

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


85.3

MIXED AIR

_ _

Q 3=10,000 l/s t 3=DB = ?3 WB3 = ?

C

_

Point B

_ _

56.3

49

_ Condition of mixture (Point C) 35_C DB

_

Point A

FIGURE 2-- 15 MIXING OF TWO AIRSTREAMS (SI)

2.2.5.4

Airstream Mixtures

Mixtures of two or more airstreams are a common requirement of the environmental system. The mixing of outside and return air on the entering side of the air cooling and heating equipment is the common method of introducing the outside ventilation air to the system. Outside air can be one of the greatest heating and cooling loads in more extreme climates.

 Assume that the airstreams are being mixed in a 50 percent ratio which causes the mixed point to be directly between points “A” and “B”.  Then assume that the damper moves to a position to take in less hot outside air (“B”). This will cause point “C” to move away from “B” (the line “B” to “C” gets longer when less air is taken in). If point “C” coincides with point “A,” 100 percent return air “A” will be used.

Figure 2-15 illustrates the mixing of an outside airstream and a return airstream, which usually is found in most air handling unit system applications.

 The mixed air temperature will be closest to the air temperature of the largest airstream.

When two airstreams are mixed and are plotted in graph form, the following steps should be used:

 Only the dry bulb air temperature can be obtained by using this method or by using Equation 2-11.

HVAC SYSTEMS Testing, Adjusting & Balancing  Third Edition

2.17


Equation 2--11 Xo T o + X r T r Tm = 100

b) From Figure 2-15:

Where:

27_C WB = 85.3 kJ/kg

Tm = Temperature of mixed air — _F (_C) Xo = Percentage of outdoor air — %

Using Equation 2-12:

To = Temperature of outdoor air—_F (_C)

X oH o + X rH r 100 20% × 85.3 + 80% × 49.0 Hm = 100 1706 + 3920 Hm = = 56.26 kJ∕kg 100 On a psychrometric chart, 56.26 kJ/kg = approximately 19.7_C wet bulb temperature.

Xr = Percentage of return air — % To = Temperature of return air—_F (_C) The wet bulb of the mixed airstream can be obtained by substituting the enthalpies of the two airstreams in Figure 2-15 in Equation 2-12 and calculating the enthalpy of the mixed airstream. From this value, the mixed airstream wet bulb temperature can be obtained from the psychrometric chart. Equation 2--12 X oH o + X rT r Hm = 100 Where: Hm = Mixed air enthalpy — Btu/lb (kJ/kg) Xo = Percentage of outdoor air — % Ho = Outdoor air enthalpy — Btu/lb (kJ/kg) Xr = Percentage of return air — % Ho = Return air enthalpy — Btu/lb (kJ/kg) Example 2.8 (SI) Calculate the dry bulb and wet bulb temperatures of the mixed airstream of Figure 2-15 using equations and by plotting in graph form on a psychrometric chart. Solution a)

Using Equation 2-11 for the dry bulb temperature: X oT o + X rT r 100 20% × 35˚C + 80% × 24˚C Tm = 100 700 + 1920 Tm = = 2620 = 26.2˚C (DB) 100 100 Tm =

2.18

17_C WB = 49.0 kJ/kg

Hm =

The psychrometric chart provides a quick method for calculating the mixed airstream conditions using only a scale as is shown in Figure 2-15. First, plot the two conditions (“A” and “B”) on the chart and draw a straight line between the two. Divide this distance in proportion to the mixed air quantities, and to scale, plot the mixed air point “C” so that it is closest to the conditions of the largest original quantity in the mixture. This is usually closest to the lowest dry bulb point since outside air quantities are usually less than 50 percent. (If they are 100 percent, no mixture determination is required.) The point “C” represents what mixture conditions should be if the air quantity proportions are correct. All of the properties of the mixture at point “C” are immediately available from the psychrometric chart (26.2_C DB and 19.7_C WB are two of them). One common error made by many novices, is the improper location of the mixed air point on the charts. Some reverse the ratio of the mixing streams, causing the mixed point shown in Figure 2-15 to occur near the top right hand point “B”. When two airstreams are mixed and are plotted in graph form, the following steps should be remembered: Assume that the air streams are being mixed a 50 percent ratio which causes the mixed point to be directly between points “A” and “B”. Then assume that a damper moves to a position to take in less hot outside air (at “B”). This will cause the point to move away from “B” (as the line “B” to “C” gets longer when less air is taken in). By the time it reaches “A”, 100 percent of the air of quality “A” will be used. It should be noted that the use of the psychrometric chart in the design to determine the properties of the

HVAC SYSTEMS Testing, Adjusting & Balancing  Third Edition


mixture does not establish the airflow quantity. This determination must be made independently. The de− signer must establish the total air quantity required from the sensible heat load and the outside air quantity from the design ventilation requirements. 2.2.5.5

Related Tables and Equations

Air (standard) density = 0.075 lb/ft3 (1.204 kg/m3 ) Water (standard) density = 62.4 lb/ft3 (1000 kg/m3) Specific volume is the reciprocal of density and is used to determine cubic feet of volume if the pounds of weight are known:

Chapter 6, Psychrometrics of the 2001 ASHRAE Fun− damentals Handbook has more detailed theory and data on this subject along with the necessary equations and psychrometric tables. Chapter 8, Thermal Comfort includes the comfort zone and effective temperature scale superimposed on a standard psychrometric chart. 2.3

FLUID MECHANICS

When the word fluid is mentioned, the average person thinks in terms of water. However, the full definition of fluid in the Glossary is ?gas, vapor, or liquid." In TAB work, the word fluid normally means air (from the atmosphere), water (or a heat transfer fluid), steam, refrigerants, and occasionally, a few other gases. This section will contain a detailed description of the behavior of air and water, as the properties of these two fluids affect most TAB work. Air and water generally have similar fluid properties except that the numerical values assigned to each property vary considerably. 2.3.1

Compressibility

For TAB purposes, water cannot be compressed. Air can be compressed and the volume of air can be pre− dicted by using Equations 2−7 to 2−9. 2.3.1.2

Water (Standard) specific volume = 0.016 ft3/lb(0.001 m3/kg) Standard Conditions for air as used above correspond to dry air to 70F (21C) and at an atmospheric pres− sure of 29.92 in. Hg. (101.325 kPa). For water, stan− dard conditions are 68F (20C) at the same baromet− ric pressure. 2.3.1.3

Weight, Density and Specific Volume Relationships

In TAB work, weight is measured in pounds (kilo− grams) and density in pounds per cubic foot (kg/m3); therefore, for standard conditions:

Specific Heat

The following specific heat values at standard condi− tions were used to develop Equations 2−5 (for water) and 2−6 (for air). Water: Specific Heat (Cp ) = 1.00 Btu/lbF (4190 J/kg C)

Fluid Properties

The basic categories of fluid properties are state, com− pressibility, viscosity, weight or density, volume or specific volume, volatility, specific heat and heat con− tent; but only those important to TAB technicians will be discussed. 2.3.1.1

Air (Standard) specific volume = 13.33 ft3/lb (0.831 m3/kg)

Air: Specific heat (Cp ) = 0.24 Btu/lbF (1000 J/kg C) Air Equation 2−6 only applies to sensible heat transfer (that which affects the dry bulb thermometer). 2.3.2

Fluid Statics

Static head is the pressure developed by the weight of the fluid at rest (not moving) in a system. The static head of air is insignificant and is ignored in TAB cal− culations, but not the static head of water. As Standard Atmospheric Pressure is measured at 14.696 psi (101.325 kPa), then Absolute Pressure is obtained by adding the 14.696 psi (101.325 kPa), atmospheric pressure to the gage pressure of the system static head.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.19


To obtain the weight of one foot of water in psi and also absolute pressure: 62.4lbsqft 1footofwater  144sqinsqft = 0.433 psi gage pressure (psig) or 1 ft of water= 0.433 + 14.696 =15.129 psia (absolute pressure) Other I−P Unit pressure/weight/height relations are: 1 inch of mercury (Hg) = 13.6 inches of water gage (in.wg) 1 foot of mercury (Hg) = 5.89 psi 1 psi = 2.31 ft wg = 2.04 in.Hg 14.696 psi = 29.92 in.Hg = 33.9 feet of water gage (ft wg) In the SI system: 1 meter (water) = 9.807 kPa

1 millimeter (mercury) = 133.32 kPa From the above relationships, it can be seen that using water instead of mercury in a U−tube manometer for a pressure of 15 psi (103.4 kPa) would be quite cumber− some. However, it would be feasible to use when mea− suring the pressure of HVAC duct systems. Figure 2−16 shows the relationship between height and pressure of an open tank of water. The hydrostatic head at point B is 30 feet (9 m) and the fluid head at point B is 30 feet of water (9 m) which is equal to 13.0 psi gage (90 kPa) and to 27.7 psia (191.3 kPa a). 2.3.3 2.3.3.1

Fluid Dynamics Velocity

The term ?dynamics" is used to describe the condi− tions of motion of a fluid or gas in a system. The ?ve− locity" of the fluid is based on the cross−sectional area of the conduit (pipe or duct) through which it is flow− ing and the volume of fluid within the conduit. When there is no turbulence, the velocity varies within the conduit as shown in the diagram in Figure 2−17, ?Velocity Profile." This phenomenon, known as the velocity profile, is caused by the friction between the conduit walls and the fluid.

OPEN TANK MAINTAINS

ATMOSPHERIC PRESSURE

CONSTANT WATER LEVEL

14.7 psi (101.3 kPa)

GAGE 10ft(3m)

HYDROSTATIC HEADS

30ft(9m) 50ft(15m)

PRESSURES

ABSOLUTE PRESSURES

4.3 psi (29.7 kPa)

19.0 psi (131.1 kPa a)

13.0 psi (89.7 kPa)

27.7 psia (191.1 kPa a)

21.7 psi (149.7 kPa)

36.4 psia (251.2 kPa a)

30.3 psi (209.1 kPa)

45.0 psia (310.5 kPa a)

70ft(21m)

FIGURE 2-16 TANK STATIC HEAD

2.20

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


V min (Zero at Wall) V avg

V max

FIGURE 2-17 VELOCITY PROFILE

The only purpose of this device is to produce a pressure sufficient to overcome the resistance of the system to the flow of the fluid. The pressure produced is indi− cated by the pressure difference from the pump or fan inlet to the pump or fan discharge. This is exactly equal to the system resistance to flow and, in the case of wa− ter, the elevation differences for the amount of fluid being pumped. A measurement of the difference be− tween the inlet and discharge pressures of a pressure device of any kind is then a measurement of the system resistance at a particular flow rate. 2.3.4

When the flow is ?turbulent," the friction rate in− creases, heat transfer through the walls of an exchang− er increases, and usually, so does the system noise that is created by the fluid flow.

Q  AV

Equation 2-13

Where: Q = Fluid flow A = Area

Air

Water

cfm (L/s)

gpm (L/s)

ft2 (m2)

ft2 (m2)

V = Velocity fpm (m/s) fpm (m/s) *Correction constants are needed so that the units in the equation are compatible. 2.3.3.2

Friction

It has been indicated that the flow of the fluid is re− sisted by a well known paradox of nature called fric− tion. Friction is the natural resistance caused by a sub− stance with which it is in contact. One substance may be stationary and the other moving or both may be moving at different velocities. If it were possible to start the fluids flowing in a system and then eliminate friction losses of the conduit and the dynamic losses of the fittings, it would be possible to eliminate all power consuming pumping equipment in closed systems. The only purpose of the pumps is to overcome these losses which result from the flow they produce. 2.3.3.3

Pressure

Pressure is the force required to overcome the friction and dynamic losses of a system. This pressure is pro− duced by a pumping device which in HVAC systems may be a circulating pump, fan, or a gaseous refriger− ant compressor.

2.3.4.1

Air System Basics Duct Pressure Changes

The pressures in air systems are simpler than those in hydronic systems because the weight of air in the sys− tems is ignored in most calculations. The resistance to airflow, imposed by a duct system, is overcome by the fan, which supplies the energy (in the form of total pressure) to overcome this resistance and maintain the necessary airflow. Figure 2−18 illustrates an example of the typical pressure changes in a duct system with the total pressure and static pressure grade lines in ref− erence to the atmospheric pressure datum line. In air conditioning and ventilating work, the pressure differences are ordinarily so small that incompressible flow is assumed. Relationships are expressed for air at a standard density of 0.075 lb/ft3 (1.204 kg/m3) and corrections are necessary for significant differences in density due to altitude or temperature. Static pressure and velocity pressure are mutually convertible and can either increase or decrease in the direction of flow. To− tal pressure, however, always decreases in the direc− tion of airflow. At any cross−section, the total pressure (TP) is the sum of the static pressure (SP) and the velocity pressure (Vp ): TP  SP  Vp

Equation 2-14

For all constant−area straight duct sections, the change in static pressure losses are equivalent to the total pres− sure losses because the change in velocity pressure (Vp ) equals zero, as the velocity is constant. These pressure losses in straight duct sections are termed friction losses. Where the straight duct sections have smaller cross−sectional areas, such as duct sections BC and FG in Figure 2−18, the pressure lines fall more rap− idly than those of the larger area ducts (pressure losses increase as the square of the velocity). When the duct cross−sectional areas are reduced, such as at converging sections B (abrupt) and F (gradual),

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.21


A

B

C

D

E

F

G

AIR FLOW

ENTRY

H EXIT Diffuser

TP

SP

TP

Vp

ATMOSPHERIC PRESSURE SP TOTAL PRESSURE (TP)

VELOCITY PRESSURE (V ) p

STATIC PRESSURE (SP)

FIGURE 2-18 PRESSURE CHANGES DURING FLOW IN DUCTS both the velocity and velocity pressure increase in the direction of airflow and the absolute value of both the total pressure and static pressure decreases. The pres− sure losses at points B and F are dynamic losses.

Dynamic losses are due to changes in direction or ve− locity of the air and occur at transitions, elbows, and duct obstructions, such as dampers, etc. Dynamic losses can be expressed as a loss coefficient (the constant which produces the dynamic pressure losses when multiplied by the velocity pressure) or by the equivalent length of straight duct which has the same loss magnitude.

Increases in duct cross−sectional areas, such as at di− verging sections C (gradual) and G (abrupt), cause a decrease in velocity and velocity pressure, a continu− ing decrease in total pressure and an increase in static pressure caused by the conversion of velocity pressure to static pressure. This increase in static pressure is commonly known as static regain and is expressed in terms of either the upstream or downstream velocity pressure. 2.22

At the exit fitting, section H, the total pressure loss co− efficient may be greater than one upstream velocity pressure, equal to one velocity pressure, or less than one velocity pressure. The static pressure just upstream of the discharge fit− ting can be calculated by subtracting the upstream ve− locity pressure from the total pressure upstream. The entrance fitting at section A in Figure 2−18 also may have total pressure loss coefficients less than 1.0 or greater than 1.0. These coefficients are references to the downstream velocity pressure. Immediately downstream of the entrance, the total pressure is sim− ply the sum of the static pressure and velocity pressure. Note that on the suction side of the fan, the static pres− sure is negative with respect to the atmospheric pres− sure. However, velocity pressure is always a positive value. It is important to distinguish between static and total pressure. Static pressure is the blow−up pressure (like a balloon) which commonly has been used as the basis for duct system design. Total pressure determines how much energy must be supplied to the system to main−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


tain airflow. Total pressure always decreases in the di− rection of airflow. But static pressure may decrease, then increase in the direction of airflow (as it does in Figure 2−18) and may go through several more in− creases and decreases in the course of the system. It can become negative (below atmospheric) on the dis− charge side of the fan, as demonstrated by points G and H. The distinction must be made between static pres− sure loss (sections BC or FG) and static pressure change as a result of conversion of velocity pressure (section C or G). The total resistance to airflow is noted by nTPsys in Figure 2−18. Since the prime mover is a vaneaxial fan, the inlet and outlet velocity pressures are equivalent; i.e. nTPsys = nSPsys . When the prime mover is a cen− trifugal fan, the inlet and outlet areas usually are not equal, thus the suction and discharge velocity pres− sures are not equal, and obviously nTPsys nSPsys . If one needs to know the static pressure requirements of a centrifugal fan, knowing the total pressure require− ments, the following relationship may be used: Equation 2-15 FanSP  TP d  TP s  Vp d (orasSP  TP  Vp) FanSP  SP d  TP s where the subscripts d and s refer to the discharge and suction sections respectively of the fan.

Above 2000 feet (600 m) altitude, below 50F (10C), or above 90F (32C) temperature, the friction loss obtained from Tables A−1 and A−2 must be corrected for the air density. Table A−12 presents the correction factors for temperature and altitude. The actual air vol− ume is used to find the friction loss from Table A−1 and A−2 and this loss is multiplied by the correction factor or factors from Tables A−3 and A−4 to obtain the actual friction loss. 2.3.4.3

HVAC duct systems usually are sized first as round ductwork. Then, if rectangular ducts are desired, duct sizes are selected to provide flow rates equivalent to those of the round ducts originally selected. Tables A−5 and A−6 give the circular equivalents of rectangu− lar ducts for equal friction and flow rate. Note that the mean velocity in a rectangular duct will be less than in its circular equivalent. Rectangular duct sizes should not be calculated direct− ly from the actual duct cross−sectional area. Tables A−5 and A−6 should be used. If this is not done, the resulting duct sizes will be smaller, with a greater velocity and friction loss for the given airflow. 2.3.4.4

2.3.4.2

Circular Equivalent for Rectangular Ducts

Dynamic Losses

Friction Losses

Pressure drop in a straight duct is caused by surface friction, and varies with the air velocity, the duct size and length, and the interior surface roughness. Friction loss is most readily determined from Air Friction Charts, Tables A−1 and A−2 in the Appendix. They are based on standard air with a density of 0.075 lb/ft3 (1.204 kg/m3 ) flowing through average, clean, round, galvanized metal ducts (with an absolute roughness factor of 0.0003 ft. The values may be used without correction for temperatures between 50F (10C) and 90F (32C) and for altitudes up to 2000 feet (600 m) and for any relative humidity. Beyond those limits, corrections should be made for other than standard air densities, and when using other duct materials or flex− ible duct. Tables A−3 and A−4 are new tables and charts that pro− vide correction factors for ducts of materials other than hot−dipped galvanized sheet metal construction. The correction factor is multiplied by the friction loss ob− tained from Table A−1 and A−2 for each straight duct section.

Where turbulent flow is present, brought about by sud− den changes in the direction or magnitude of the veloc− ity of the air flowing, a greater loss in total pressures takes place than would occur in a steady flow through a similar length of straight duct having a uniform cross−section. The amount of this loss in excess of straight−duct friction is termed dynamic loss. 2.3.4.5

Duct Fitting Loss Coefficients

The duct fitting loss coefficient ?C" is dimensionless and represents the number of velocity heads lost at the duct transition or bend. Values of the loss coefficient for elbows and other duct elements have been deter− mined experimentally or computed and can be found in tables in the SMACNA HVAC Systems 9 Duct De− sign manual or the ASHRAE Fundamentals Hand− book. Tables A−4 and A−5 which show the relation of veloc− ity to velocity pressure for standard air, can be used to find the dynamic pressure loss for any duct element whose loss coefficient ?C" is known.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.23


Fitting loss coefficient Velocity pressure—in.wg

Tee, 45 Entry, Rectangular Main and Branch Use the Vp of the upstream section Ac

Branch, Coefficient C

Qc

Vc

V / V b c

Qs Ab

Qb

Vs Ac = A s

As

Qb/ Qc 0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.2

0.91

0.4

0.81

0.6

0.77

0.72

0.70

0.8

0.78

0.73

0.69

0.66

1.0

0.78

0.98

0.85

0.79

0.74

1.2

0.90

1.11

1.16

1.23

1.03

0.86

1.4

1.19

1.22

1.26

1.29

1.54

1.25

0.92

1.6

1.35

1.42

1.55

1.59

1.63

1.50

1.31

1.8

1.44

1.50

1.75

1.74

1.72

2.24

1.63

0.79

FIGURE 2-19 SAMPLE FITTING LOSS COEFFICIENT TABLE

TP  C  Vp

Equation 2-16

Equation 2-17 (SI) Q V 1000A Where:

Where:

V = Velocity (m/s) TP = Total pressure loss C in. wg (Pa) Q = Duct airflow (L/s) C = Fitting loss coefficient Vp = Velocity pressure C in. wg (Pa) The velocity pressure (Vp) used for rectangular duct fittings must be obtained from the velocity (V) ob− tained by using the following equation:

A = Duct cross−sectional area (m2) Where different areas are involved, letters with or without subscripts are used to denote the area at which the mean velocity is to be calculated, such as A for in− let area, A1, for outlet area and Ao for orifice area, etc. The equation for obtaining the velocity pressure (Vp) from the velocity is:

Equation 2-17 (IP)

Q V A

Equation 2-18 (I-P)

Vp  V 4005

2

Equation 2-18 (SI) V p  0.602V 2

Where: V = Velocity (fpm) Q = Duct airflow (cfm)

Where: Vp = Velocity pressure C in. wg (Pa) V = Velocity C fpm (m/s)

A = Duct cross−sectional area (sq ft)

2.24

Commonly used fitting loss coefficient tables ex− tracted from the SMACNA HVAC Systems9Duct De− sign manual can be found in Tables A−17 and A−20.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Equation 2-19

Example 2.8 (I−P) Q2 rpm  rpm2 1 Q1

A main duct has an airflow of 10,000 cfm at 2000 fpm. A 45 entry tap is used for the branch duct that requires 3,000 cfm at 1,600 fpm. Find the pressure loss to the branch of the fitting.

d2 rpm  rpm2 1 d1

Using the data from Figure 2−19: Qb 3, 000   0.3 Qc 10, 000 Vb 1, 600   0.8 Vc 2, 000 Branch loss coefficient C = 0.69 2, 000

 (0.5)  0.25 V  4, 005 2

2

p

TP  C  Vp  0.69  0.25 TP  0.1725in. wg Example 2.8 (SI) A main duct has an airflow of 5,000 L/s at 10 m/s. A 45 entry tap is used for the branch duct that requires 1500 L/s at 8 m/s. Find the pressure loss to the branch of the fitting. Solution Using the data from Figure 2−19: Qb  1500  0.3 Qc 5000 Vb  8  0.8 Vc 10 BranchlosscoefficientC  0.69 V p  0.602  (10) 2  60.2Pa TP  C  Vp  0.69  60.2Pa  41.5Pa

2.3.4.6

bp 2 rpm  rpm2 1 bp 1

Solution

Fan Laws

The TAB technician can use fan curves or tables pub− lished by the fan manufacturer to determine the output of a fan under certain conditions. The following fan law equations also allow the TAB technician to calcu− late the necessary changes to be made to a system or a system component prior to the actual work.

Equation 2-20

P2 rpm  rpm2 P1 1

2

Equation 2-21 3

Equation 2-22 2

Where: QPP=MAirflow C cfm(L/s) rpmP=Fan revolutions per minute P'P=Static or total pressure – in. wg (Pa) bpPM=Fan brake power – HP (W) dPP=Density C lb/ft3 (kg/m 3) The relationship between the fan laws, fan curves, and system curves will be discussed in Chapter 5 Fans. One of the most important things to remember is that the fan brake power increases as the cube of the fan rpm increase (or of the airflow increase by combining Equations 2−17 and 2−19). So when the TAB technician attempts to increase the airflow in a system without making other changes, the fan brake power (and the fan energy consumption) in− creases dramatically. 2.3.4.7

System Pressure

The total system pressure that the system fan must han− dle then is the sum of the friction losses of each straight duct section, the dynamic losses of each duct fitting or obstruction, and the pressure loss of each duct compo− nent such as coils, filters, dampers, etc. Examples can be found in Chapter 7 Air Systems. In a given duct system with a known airflow rate, and when the position of all dampers is stable, a specific total pressure can be measured. If the airflow is in− creased without any other changes, then Equation 2−21 can be used (this equation was derived from fan law Equations 2−17 and 2−18);

P2 Q2  P1 Q1

Equation 2-23 2

Where:

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.25


P  Systempressure  in.wg(Pa) Q  SystemAirflow  cfm(Ls)

Obviously, the 5 HP motor would be inadequate and a 10 HP motor would be marginal.

Example 2.9 (I−P Units) Example 2.10 (SI) A duct system is operating at 2.0 in.wg with an airflow of 10,000 cfm. If the airflow is increased to 13,000 cfm without any other change, determine the new duct sys− tem pressure.

Solution

000 QQ  2.0 13,

10, 000 2

P 2  P 1

The same system used in Example 2.9 has a 3.8 kilo− watt motor operating at 3.6 kilowatts. Find the fan brake power and standard motor size that would be re− quired if the airflow was increased to 6500 L/s.

2

2

Solution

1

 2.0(1.3)2  3.38in.wg

 7.91kilowatts  3.6 6500 5000

Q bp 2  bp1 2 Q1

Example 2.9 (SI) A duct system is operating at 500 Pa with an airflow of 5000 L/s. If the airflow is increased to 6500 L/s without any other change, determine the new duct sys− tem pressure.

 845Pa  500 6500 5000

Q P 2  P 1 2 Q1

2

2

Example 2.10 (I−P Units) The same system used in Example 2.9 has a 5 HP mo− tor operating at 4.82 bhp. Find the bhp and standard motor size that would be required if the airflow was in− creased to 13,000 cfm. Solution

000 QQ  4.82 13,

10, 000 3

bp 2  bp1

2 1

  4.82(1.3)3  10.59

2.26

3

3

3

The 3.8 kW motor is inadequate and a 7.5 kW motor would be marginal. 2.3.5 2.3.5.1

Hydronic System Basics Hydronic Pressure Losses

In air systems, the weight of air in the system is ignored by system designers and TAB technicians. In hydronic systems, it is the velocity head that is ignored because the values usually are insignificant. Otherwise, hy− dronic systems are subject to similar friction losses in the straight runs and dynamic losses in the fittings. Manufacturers normally supply pressure loss data for equipment used in piping systems. Pressure losses for hydronic systems are given in terms of equivalent feet of pipe, in pounds per square inch (psi), or in feet of water (ft wg) in I−P Units. In SI units, meters of water and Pascals or kilopascals are used.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Equation 2−22 is the hydronic equivalent of air systems Equation 2−21: Equation 2-24

P2 Q2  P1 Q1

2

towers, are subject to greater corrosion and therefore have higher pressure losses per 100 feet of pipe. Prop− erly sized pumps used with these systems generally will deliver a higher flow rate than required by a newly installed system because the anticipated corrosion has not occurred. However, after a few years, a partially closed balancing cock at the pump will have to be opened.

Where: 2.3.5.3 PP= Pressure difference C psi(Pa or kPa) QP= Flow rate C gpm(L/s)

Example 2−11 (I−P) A piping system has a flow rate of 100 gpm at a 20 ft wg head. Calculate the flow rate if the flow resistance is reduced to a 10 ft wg head.

Solution

2

P2 Q2  ; Q2  Q1 P1 Q1 Q 2  100

PP

10  71gpm 20

2

A piping system has a flow rate of 6.3 L/s at a 6 meter head. Calculate the flow rate if the flow resistance is reduced to a 3 meter head.

2

P2 Q2  ; Q2  Q1 P1 Q1 Q 2  6.3

PP

36  4.45Ls

Static head, discussed earlier, is the pressure due to the weight of the fluid above the point of measurement. In a closed system, the selection of the pump capacity is not affected as the static head is equal on both sides of the pump. However, the pump casing must be designed to handle the highest static head. Suction head is the height of fluid from the centerline of the pump on the suction side to the level of the fluid surface as is shown in Figure 2−20. The actual static head pressure loss that is added to the piping and sys− tem pressure loss in order to size the pump is the differ− ence between A minus B.

1

Example 2−11 (SI)

Solution

Heads

2 2

Suction lift is the height of fluid that a pump must lift on the suction side of the pump from the level of the fluid surface to the pump centerline as shown in Figure 2−20. This pressure loss value is added to any other sys− tem or pump pressure losses if additional piping or equipment is involved. 2.3.5.4

Pump Laws

The following equations are similar to (or the same as) the fan law equations. Again, they allow the TAB tech− nician to calculate changes that could occur in a given hydronic system when one or more of the conditions is altered. Most equations required for TAB work also can be found in the Appendix along with the SI equiva− lents. Equation 2-25

2.3.5.2

Friction Losses

Friction loss tables for hydronic systems vary in value depending on the condition of the piping system and the type of pipe or tubing used. Closed systems, where the fluid continuously recirculates, such as hot and chilled water systems for HVAC work, stay relatively clean and free from deposits that could roughen the in− terior surfaces. Open systems, such as domestic hot water systems and condenser water systems with normally open cooling

Q2 rpm  rpm2 1 Q1 Equation 2-26 Q2 D  2 D1 Q1

H2 rpm  rpm2 H1 1

H2 D2  H1 D1

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

Equation 2-27 2

Equation 2-28 2

2.27


STATIC PRESSURE

STATIC HEAD (DIFFERENCE = A - B) TO BE ADDED TO PUMP HEAD

A

SUCTION HEAD

B

CLOF PUMP TANK

PUMP

FIGURE 2-20 PUMP WITH STATIC HEAD AND SUCTION HEAD

bp 2 rpm  rpm2 1 bp 1

bp 2 D2  D1 bp 1

2.28

Equation 2-29

Equation 2-30 3

Where:

3

Q rpm D H bp

= Fluid flowCgpm (L/s) = Revolutions per minute = Impeller diameterCinches (mm) = HeadCfeet (meters) = Brake powerCHP (kW)

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


CLOF PUMP

PUMP

L SUCTION LIFT

TANK

FIGURE 2-21 PUMP WITH SUCTION LIFT

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

2.29


THIS PAGE INTENTIONALLY LEFT BLANK

2.30

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


CHAPTER 3

ELECTRICAL EQUIPMENT AND CONTROLS


CHAPTER 3

ELECTRICAL EQUIPMENT AND CONTROLS

3.1

ELECTRICAL SYSTEMS

3.1.1

Basic Electricity

Equation 3-3 1R t  1R1  1R2  1R3    1R n (Parallel Circuits)

The electrical industry concerns itself with a broad range of subjects, many of which are not necessarily directly related to the HVAC industry. However, it is necessary to know basic electricity as it applies to that part of building construction which is related to me− chanical and electrical systems. Electric motors drive or power almost all HVAC equipment, including fans and pumps. Understanding the operation of, and the differences between, the many types of motors, the related controls, and the control circuits, is a necessity for the TAB technician. The TAB technician must be able to determine the brake horsepower that is being applied to air handling equipment and ensure that the motor is properly con− nected and protected. A few simple equations should be kept in mind when dealing with electricity. The first is a derivative of Ohm’s law: The current in a circuit is equal to the elec− tromotive−force activity in the circuit divided by the resistance in the circuit.

Where: R t  TotalSystemResistance R n  IndividualResistances

Equation 3−3 states mathematically that the parallel current flow will work similarly to hydronic flow, with the circuit with the highest resistance receiving the lowest flow. 3.1.2.2

Series Circuits

Resistances are added together for electrical circuits in series as in hydronic circuits. As more resistances are added, the flow becomes less and less. Equation 3-4 R t  R1  R2  R3    R n (Series Circuits) Where: R t  TotalResistance R n  IndividualResistances

Equation 3-1 I  E (or)E  IR R P  IE

Equation 3-2

Where: I  Amps(A) E  Volts(V) R  Ohms(W) P  Watts(W) 3.1.2 Electrical Resistances 3.1.2.1

Parallel Circuits

Parallel electrical circuits resemble HVAC terminal units piped in parallel circuits. Using a simple circuit with two units and one pump, it is known that the water flow will split in accordance with the resistance across each unit. If both resistances are the same, the flow will split 50−50.

The electrical diagram in Figure 3−1 is similar to a pip− ing circuit with a pump at ?E", two chillers in series at ?R1" and ?R2", seven terminal units piped in parallel, and a strainer, valve, etc., piped in series in the pump suction piping. This shows the similarity of electrical calculations to those for hydronic and air, where resist− ances in series are added, and those in parallel are com− bined. 3.2

ELECTRICAL SERVICES

3.2.1

Single Phase Circuit Voltages

A measured voltage may not be exactly one of the val− ues of voltages indicated in Figure 3−2. Voltages can vary, and in normal situations a variation of ±10 per− cent will not adversely affect equipment operation. The basic 115 volt two−wire circuit shown in part ?A" of Figure 3−2 is very common. There is a potential or pressure of 115 volts between the hot wire and the neu− tral or ground. Normal household use, such as a lamp, is representative of such a circuit. The 115 volt poten− tial in the hot wire will exist between the hot wire and the neutral, or between the hot wire and any other

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

3.1


Parallel

IT

R1

R2

(Etc.)

Series E Series

Series-Parallel

FIGURE 3-1 SERIES-P ARALLEL CIRCUIT

FUSE MAIN SWITCH

TO LOAD

115 V

GROUND

TO EQUIPMENT

GROUND LINE GROUNDS

A. TWO-WIRE CIRCUIT

GROUND LINE GROUND TO EQUIPMENT

MAIN FUSE HOT

MAIN SWITCH 115 V

115 V LOAD

115 V

115 V LOAD

NEUTRAL MAIN FUSE HOT CIRCUIT FUSES GROUND LINE

B. THREE-WIRE CIRCUIT

FIGURE 3-2 SINGLE-PHASE AC SERVICE

3.2

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

230 VOLT LOAD


+

Time

+

Time

+

Time

Current I

Voltage E

Voltage E

Voltage E

Current I

Current I

--

-Difference in Peaks Causing Power Loss Inductance Load Current Voltage

-Difference in Peaks Causing Power Loss Capacitance Load Current Voltage

Peaks Concurrent Providing Maximum Useful Power Inductance -- Capacitance Current Voltage (Power Factor Corrected)

FIGURE 3--3 CURRENT AND VOLTAGE-- TIME CURVES AND POWER FACTOR ground, such as a pipe or a person which might contact the wire.

The neutral or ground wire is another matter. The neutral normally has no voltage potential. Theoretically, if the neutral contacts a pipe or a person, nothing will happen. The neutral is connected to the ground. The term theoretically is used, because in actual field conditions, stray currents can find their way into the neutral and it then can be dangerous. A neutral should be treated with the same respect as a known hot wire.

Part “B” of Figure 3-2 shows another common single phase circuit. It is also a household circuit which serves items requiring greater power such as ranges, clothes driers and air conditioners. This circuit represents the type of three-wire service normally entering residences. Two of the three wires are hot wires and one is neutral. The voltage potential between either of the hot wires and the neutral is 115 volts. There are actually three circuits; two separate 115 volt circuits (from each of the hot wires to the neutral) and a 230 volt circuit (between the two hot wires). The neutral in a 230 volt circuit found in some appliance connections serves as a ground for safety, but it is not used as part of the power circuit. It is connected to the frame of the machine to carry off any stray currents or any short circuits resulting from failures. Ground or neutral wires are never switched or fused. The main advantage of the 230 volt, two hot wire circuit is that it allows each of the hot wires to carry half of the current flow. Therefore, twice the current will be handled by the same size wires.

3.2.2

Three Phase Circuit Voltages

The three-phase (3∞) concept is somewhat more difficult to understand. Figure 3-3 shows alternating current pulses of voltage and current changing with time. In the case of the single-phase, three-wire circuit, two different pulses are being sent down two different hot wires. After one starts, the second starts 1/120th of a second later. These pulses continue indefinitely at the same frequency and have the same phase relationship between the 2 wires. This can be thought of as +115 volts and -115 volts between the hot wires and the neutral wire. The pulses would average 115 volts between the hot wire and the neutral wire, would change to -115 volts and back to +115 volts and so on 120 times per second. The pulses going down the other wire would do the same, but because their timing is out of phase, an instantaneous look at the two 115 volt wires would show that the voltage in one wire would be +115 volts, and the other wire (because of its delay in phase) would be -115 volts. When voltage readings are taken with a voltmeter, there is no apparent way to tell the difference between 220 volt single phase circuits and 220 volt three-phase circuits. When measurements are taken, it is found that voltages do vary somewhat; that three-phase circuits are usually 220 volt, and that single-phase circuits are usually 230 volt. However, phasing cannot be determined by just voltage readings. Four-wire, three-phase circuits, as illustrated in Figure 3-6, produce 208 volts between phase wires. This arrangement is commonly used in commercial buildings, as the 120 volt loads can be divided equal between the three hot (phase) wires.

HVAC SYSTEMS Testing, Adjusting & Balancing  Third Edition

3.3


Three-Phase Motor (220 V)

L1 Main Switch Fuse

220 V

220 V

L2

220 V L3

FIGURE 3-4 220-VOL T THREE-WIRE DELTA THREE-PHASE CIRCUIT

Center Tap

Approx. 177 V

Fuse

Three-Phase Motor (220 V)

L1 Transformer

Main Switch L2

110 V 220 V L2

Fuses

Single-Phase Loads (110 V) Neutral

FIGURE 3-5 220-VOL T DELTA THREE-PHASE CIRCUIT WITH 110-VOL T SINGLE-PHASE SUPPLY

Fuse

Three-Phase Motor (208 V)

L1 Main Switch

208 V

208 V

L2

208 V L2

Fuses

Single-Phase Loads (120 V) Neutral

FIGURE 3-6 120/208-VOL T FOUR-WIRE WYE CIRCUIT

3.4

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


3.3

TRANSFORMERS

For the voltage reduction indicated for the transformer illustrated in Figure 3−7, the number of turns shown on the primary side should be twice the number of turns shown on the secondary side. Voltage is transformed by the transformer stepping down the voltage to one half the original voltage. By swapping the primary and secondary connections, this same transformer could step up the voltage from 440 volts to 880 volts. The ballasts in fluorescent lights in buildings step up the voltages from 115, 220, or 227 volts to voltages near 2500 volts, the required voltage to produce light in the tubes. The function of the center tap of a transformer also is illustrated in Figure 3−7. If a 220 volt difference exists between the legs of the secondary side, it is logical that a 110 volt difference would exist between one leg and a center tap. Most single−phase residential transform− ers have high voltages on the primary side, but the sec− ondary voltages use a ?center tap" (the ground) to fur− nish two 110 volt circuits along with the 220 volt power (Figure 3−2). This size transformer, which looks like a large can, usually is attached to a pole near the residence. It can supply power to several residences or buildings, or just to a single building. Larger transformers are ground mounted, and if out− side, are usually in a metal housing. These transform− ers can be either single−phase or three−phase. The ?tap" (or neutral) cannot be used with hot line ?L2" for 110 volt single−phase loads (Figure 3−5) from a three− wire, 220 volt circuit. However, all three hot line or phase wires may be used for 120 volt loads in four− wire, 208 volt circuits (Figure 3−6).

110 V Secondary 440 V Primary

(Center Tap) 220 V Secondary 110 V Secondary

FIGURE 3-7 TRANSFORMER WITH TAPPED SECONDARY 3.4

MOTORS

3.4.1

Types of Motors

Most motors used on HVAC system equipment are de− signed for alternating current. Small motors will use single−phase current, while the larger motors will use three−phase current. However, some rural areas must use larger motors on single−phase current, as three−

phase current is not available. There are many differ− ent motor speeds, but 1800 rpm and 3600 rpm are the most common. The actual speed of the motor will vary with the load imposed. Split phase, capacitor start, synchronous, induction, shaded poleCall are part of the many different types of motors that the TAB team will need to recognize. The characteristics of each is important for troubleshooting, as the wrong type of motor may be used. 3.4.2

Rotation

Motors rotating in the wrong direction are a common occurrence when a new system is started. The normal TAB procedures deal with this situation, as correct motor rotation is vital to the performance of the unit. The direction of the motor usually is changed in three− phase motors by switching any two of the three−phase power wires. In single−phase motors, the change of di− rection is accomplished by switching two of the inter− nal motor leads that connect to the motor line terminal lugs. CAUTION: Certain fans and most pumps will develop measurable pressures and some fluid flow when the rotation is incorrect. Rotation arrows can be found on many types of equipment. Correct rotation is obvious on some units. Flow and amperage readings also can be used to determine whether something is amiss. Whenever a piece of equipment does not perform as specified and the current flow is much lower than de− sign, rotation is one item to be checked. 3.4.3

Nameplate Data

Except for some small motors, an attached nameplate will supply the basic information that the TAB techni− cians needs: full load amps, rpm, horsepower or watts capacity, voltages, line phase, and cycles. Many motor nameplates contain starting load amps that are quite large compared to full load amps. Although starting load amps are not as important (and are not recorded), the amperage value must be used by the system design− er for electrical circuits, circuit breaker panels, and motor starting equipment. Information on special motors might have to be ob− tained from specification sheets or from the HVAC unit nameplate. Since voltage and amperage measure− ments are seldom the same as the nameplate values, the actual motor horsepower being produced can be es− timated. However, the no load amps reading is difficult to obtain from close−coupled equipment where the load cannot easily be detached from the motor. The motor must be running with all drives disconnected for this no load amps reading.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

3.5


3.4.4

Operating Temperatures

Every electric motor generates heat as well as power. The more inefficient the motor, the greater the amount of heat produced. Unless this heat is dissipated, the temperature within the motor will rise until the insula− tion is destroyed. The amount of heat produced in the motor also de− pends upon the load. Therefore, most motors are rated on the basis of a certain temperature rise when running fully loaded. Open motors are generally designed to run at a temperature rise of 72F (40C) above the sur− rounding (ambient) air temperature. In other words, an open motor when running at full nameplate condi− tions, could be running at a temperature of 142F (61C) if the surrounding air is 70F (21C). Totally enclosed motors generally are rated at a 99F (55C) temperature rise. Under the latter conditions, a totally enclosed motor would run at 169F (76C). 3.4.5

Motor Performance

Figure 3−8 indicates ways in which various motor fac− tors are interrelated. The point at which the speed and the amperes cross corresponds to 55 percent of the maximum amps and over 60 percent of the maximum horsepower. At this point, the speed is between 97 and 98 percent of the maximum synchronous speed (which is not a great change) and the efficiency curve stays fairly flat close to 90 percent. The power factor also stays between 80 and 90 percent. Single−Phase Circuits: Equation 3-5 (I-P) I  E  P.F.  Eff. bhp  746 Equation 3-5 (SI) I  E  P.F.  Eff. kW  1000 Three−Phase Circuits: Equation 3-6 (I-P) I  E  P.F.  Eff.  1.73 bhp  746

3.6

Equation 3-6 (SI) kW 

I  E  P.F.  Eff.  1.73 1000

Where: bhp'= Brake horsepower (I−P) kW' = Kilowatts (brake power) (SI) I' = Amps E' = Volts P.F.'= Power factor Eff.'= Efficiency In Equation 3−5 and 3−6, the power factor and efficien− cy values must be used to obtain the actual motor brake power. As these values usually are difficult to obtain, a reasonable estimate can be used. Referring to Figure 3−8, the normal range of both curves is between 80 and 90 percent. Therefore, 80 percent might be used for one value and 90 percent for the other value to obtain a brake power estimate. Brake power is calculated to verify that the proper size motor has been installed, i.e., that the installed motor is not overloaded and is operating within its service factor. It also is used to determine that the pump or fan is operating with the required efficiencies. The system designer usually has specified the total amount of pow− er or energy that may be consumed to perform a specif− ic function. For example, a pump is selected to circulate 120 gpm (7.6 L/s) of water at a 40 foot (12 m) head and consume not more than two horsepower (1.5 kW). Designers in the past may have used a three horsepower (2.3 kW) motor on the pump in order to have a safety factor, or to have extra capacity for future loads that may be planned for the system. With energy conservation con− siderations, the installation of a three horsepower (2.3 kW) motor just to have a safety factor should not be used unless a future load is planned. When the designer does not want to use more than the rated two horse− power (1.5 kW), it is essential to calculate the actual power consumption unless the amperage reading is well within the full load amperage rating of the motor. There are two equations that must give a less theoreti− cal, but more practical approach to the calculation of motor full load (F.L.) amperage and brake horsepower (kW).

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Equation 3-7 ActualF.L.amps 

Solution

F.L.amps *  voltage * Actualvoltage

a.

Using Equation 3−7: 8.16  220V 210V  8.55amps

ActualF.L.amps 

Equation 3-8 bhp(kW)  HP * (kW *)  (Motoroperatingamps)  (Noloadamps  0.5)

b.

Using Equation 3−8:

(ActualF.L.amps)  (Noloadamps  0.5) bph (kW)'= 3 HP (2.3 kW)

Use Equations 3−7 and 3−8 to obtain an accurate (but not exact) brake power by measuring motor amper− ages and voltages under no load and full load condi− tions.



(6.2)  (4.7  0.5) (8.55)  (4.7  0.5)

bph(kW)  3HP(2.3kW)  *Nameplate ratings that supply the basic information.

(6.2)  (2.35) (8.55)  2.35)

bph(kW)  approx.1.86HP(1.43kW)

Example 3.1 Example 3.2 A fan has a 3 HP (2.3 kW), 220 volt, 3 phase motor that actually draws 6.2 amps at 210 volts. The full load am− perage shown on the nameplate is 8.16 amps and the ?no load" measurement is 4.7 amps. Determine the approximate fan brake power.

The following loads are paralleled across a 220 volt, single−phase, 60 Hz (hertz) source: a.

A 10,000 watt electric heater

90 100

80

Efficiency 90

% Power Factor

100

70 100 100

60

99

50

98

40

97

30 96 20

75

50

25

% Current (Amperes)

70

% Synchronous Speed

% Efficiency

80

95

10

0 0

10

20

30

40

50

60

70

80

90

100

% Motor Horsepower

FIGURE 3-8 TYPICAL PERFORMANCE OF STANDARD SQUIRREL CAGE INDUCTION MOTORS HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

3.7


b.

A 5 horsepower (3.8 kW) electric motor oper− ating at a P.F. of one, with a 90 percent effi− ciency and pulling full load

The reason for the variance in answers between the I−P and SI examples is that the 3.8 kW motor size is a nom− inal size, but 3.73 kW is the exact equivalent.

c.

A 220 to 110 volt transformer (assume 100 percent efficient) with fifty 100 watt incan− descent light bulbs in parallel with the secon− dary.

3.5

MOTOR CONTROLS

3.5.1

Safety Switches

Heater

Transformer 110 V

220 V

10,000 Watt

1&

Motor 5HP 50 @ 100 watts each

Determine the total current requirements on the 220 volt source. Solution a.

b.

P

= EI, I = P/E = 10,000 W/220 V

I

= 45.5 amps

Using Equation 3−5 (I−P):

I  E  P.F.  Eff. 746 P  P.F.  Eff. bhp  746 bhp  746 5  746  4144watts P  P.F.  Eff. 1  0.9 bhp 

Using Equation 3 − 5 (SI): I  E  P.F.  Eff. 1000 P  P.F.  Eff. kW  1000 kW  1000 3.8  1000  4222watts P  P.F.  Eff. 1  0.9 I  P  4144  18.8amps) E 220 I  4222 220 = 18.8 amps (I−P) kW 

c.

For transformers with 100 percent efficiency, P(in) = P(out) P  50  100watts  5000watts I  P  5000  22.7amps E 220

I (Total) = 45.5 + 18.8 + 22.7 = 87.0 amps (I−P) I (Total) = 45.5 + 19.2 + 22.7 = 87.4 amps (SI) 3.8

A simple on−off toggle switch, a safety switch, or an individual circuit breaker in an electrical power panel is not an overload protection device for a motor. Many ordinary looking toggle switches do contain overload protection for smaller single−phase motors. Many small motors do have built−in overload protec− tion, and do not need additional protection. The circuit breaker only provides overload protection for the wir− ing circuit, but not any connected motor(s). The electric current to a motor must be switched off and on to stop and start the motor (manually or auto− matically). The switching device is commonly called a motor starter. This is not to be confused with a safety switch, which is a device that must be placed in the off position before any work is done on a motor or electri− cal equipment. This prevents the motor from acciden− tally starting from remote control devices. 3.5.2

Motor Starters

There are a large number of different types of starters, each with various advantages and limitations. In most cases a specific type of starter is required by a particu− lar type of motor. For example, a full voltage magnetic starter usually is used with an induction motor. Re− duced voltage or reduced current starters, while more expensive than a magnetic starter, often must be used with larger horsepower motors to prevent disruption (by producing large drops in line voltage) of marginal− ly adequate power services. Many electrical utility companies have mandatory requirements for these starters above a certain horsepower (this varies with the type of equipment and voltage). The motor starter or the safety switch is the main source of access to motor terminal leads for measure− ment of voltage and amperage. The starter also can contain holding coils, auxiliary contacts, control trans− formers, and a push−button station or a hand−off−auto− matic selector switch. This last item is useful in trou− bleshooting. If the switch is turned to the hand position, the motor should run if each phase line is hot, unless there is trouble in the motor. This is because the hand portion of the switch bypasses the various con− trols in the circuit. The automatic portion of the switch is connected to the circuit containing auxiliary devices

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


such as thermostats, safety lockouts and other external switches used to control or turn off the motor. If the motor runs on hand, but not on automatic, contacts in one of the control or safety interlocks should be open.

The TAB technician can find many different voltages around starters and starter combinations (see Figure 3−9). If a remote room thermostat is added in series with the push button station of the ?A" unit, the 110 volt control circuit then may be required to be 24 volts.

It is not good practice to use line voltages (110 volt or higher) for control circuits, but to reduce costs, 240 volt control circuits are not uncommon.

The motor starter overload protection devices or heat− er coils should be sized from the actual motor name− plate data rather than from data from motor or starter manufacturer’s catalogs. Heater coils never should be oversized under any condition.

3.5.3

Push Button Stations

There are two basic types of push button stationsCthe maintained contact station and the momentary contact station. The important point to remember is that after an interruption of the current with the momentary con− tact station, the motor will not restart until the start but− ton is pushed. 3.6

VARIABLE FREQUENCY DRIVES

Variable speed or variable frequency drives (VFD) are becoming commonplace in new commercial and insti− tutional construction. On renovation projects, this electrical component is ?spliced" into the existing wir− ing between the electrical source and the fan or pump motor disconnect. Figure 3–10 shows an existing H & V unit that was retrofitted with a VFD wired into the motor disconnect. Drive manufacturers achieve variable motor speed op− eration by modulating the frequency of the electrical power being supplied to the motor during normal op− erations, as well as voltage adjustment during startup

440/3/60

MAINTAINED CONTACT PUSH BUTTON STATION

220/3/60 MAIN CONTACTS

L1

L2

L3

L1

L2

L3

T1

T2

T3

STOP AUXILIARY CONTACT

START T1

T2

HOLDING COILS

T3

HEATER COILS

110 V

440/110 V CONTROL TRANSFORMER

M A. CONTROLLING STARTER-MOT OR

110 V 220/110 V CONTROL TRANSFORMER

M B. CONTROLLED STARTER-MOT OR

FIGURE 3-9 INTERLOCKED STARTERS WITH CONTROL TRANSFORMERS HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

3.9


to approximately 40% of their full design speed with no noticeable effect on motor performance or life. However, unless specifically designed for variable fre− quency operation, most motors will develop a notice− able increase in noise level and temperature below this point. Continuous operation below 25 percent of full load speed is not recommended and this lower range should only be used to reduce high current surges and belt stress during fan startup or shutdown. Just as lowering the frequency below 60 cycles per sec− ond will lower the motor speed below nameplate RPM, you can also increase motor speed above name− plate rating by increasing the frequency above 60 cycles per second. All initial air balancing should be carried out with the VFD set for approximately 55 cycles per second (Hz), even if drive pulleys and belts must be changed to achieve the design peak air flow. This will provide added ?headroom" for fan capacity when the system experiences duct and coil static pres− sure loses from dirt buildup. 3.6.1 FIGURE 3-10 VFD ADDED TO EXISTING AIR HANDLING UNIT and shutdown. Although most 3−phase motors will op− erate satisfactorily on a VFD, low cost motors and some energy efficient motors may experience higher noise or heat levels, especially at very low speed op− eration. Many manufacturers now offer modified mo− tor designs specifically for VFD operation. Varying the speed of an AC motor is much more com− plex than DC motor speed control. Most traditional DC motors have a commutator and brushes to provide electrical power to the rotating armature coil. Varying the voltage to this coil using a resistor or a rheostat changes the motor’s speed fairly effectively. AC motors do not have a commutator or brushes since the constantly alternating electricity induces an oppos− ing electrical field in the armature coil much like the primary coil of a transformer inducing a voltage in the secondary coil. Each AC motor is designed for a spe− cific voltage and reducing the supply voltage below this value will cause the motor to quickly lose its load capacity and stall, causing overheating and eventual motor burnout. Keeping the supply voltage to an AC motor at its de− sign point while varying the frequency of this voltage results in a corresponding change of motor speed. Most AC motors can be operated continuously down 3.10

VFD Operation During TAB Work

Most VFD devices can be programmed with a maxi− mum and minimum motor speed, and ?ramp up" and ?ramp down" rates to reduce wear on drive belts and bearings. The days of a TAB technician replacing mo− tor pulleys to adjust fan speed is drawing to a close, and today’s TAB technician needs to become familiar with variable frequency motor drives and their setpoint pro− gramming. In most cases, drive and motor manufacturers do not recommend operating these systems for extended peri− ods of time below 40 percent motor speed, and all VFD programmed setpoints should be verified and recorded during the balancing process. 3.6.2

VFD Bypass

Almost all VFD devices include a hand−off−auto switch and a manual speed control dial that can be used in manual operation. If the VFD device is placed in manual mode during TAB work, be sure it is returned to the auto mode before completing this work. Also note that in manual mode it is possible to operate the fan or pump at full speed. Since a remote duct or piping pressure sensor may be connected to the VFD device to allow maintaining a fixed duct or water sup− ply pressure, manually operating the fan or pump while down stream dampers or valves may be closed could cause very high pressures to develop in ducts or piping.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


CHAPTER 4

TEMPERATURE CONTROL


CHAPTER 4 4.1

AUTOMATIC TEMPERATURE CONTROL SYSTEMS

4.1.1

Introduction

Automatic temperature control of HVAC includes the control of temperature, humidity, and sometimes sys− tem or building pressures. The automatic temperature control (ATC) system constantly adjusts the HVAC systems to maintain design conditions within the occu− pied space. Because TAB work is concerned with the operation of the HVAC system, and because the control system constantly adjusts the HVAC system, it is imperative that the TAB technician understands the function and use of automatic temperature control systems. It must be remembered that motor controls covered in Chapter 3, are not automatic temperature controls, but the automatic temperature control system may moni− tor or operate motors and motor controllers. The use of the word control for both systems and devices some− times causes this basic difference to be overlooked re− lated to responsibility. 4.1.2

Types of ATC Systems

There are four basic types of controls for HVAC or en− vironmental systems:

   

electric pneumatic electronic self−contained

There are also combinations of the above types. 4.1.2.1

Electric Controls

Electric controls are those which are line voltage or less (generally 110 volts maximum). Reduced volt− ages are obtained from transformers, either locally or centrally situated. Many of these controls are simply on−off devices such as a high limit thermostat control− ling an exhaust fan. Generally, complex systems re− quire more sophistication than these controls can pro− duce. 4.1.2.2

Pneumatic Controls

Until the advent of micro−electronics, all HVAC con− trols of air handling systems, chillers, and boilers were pneumatic controls. These systems used compressed air to operate diaphragms and mechanical relays to

TEMPERATURE CONTROL position dampers and valves. Although fairly easy to visually observe the operation of each control device, changing the sequence of operation for any HVAC sys− tem was in many cases a plumbing ?nightmare," since everything was interconnected by air tubes. At the close of the 1990’s, most pneumatic logic controls have been replaced by direct digital controls (DDC), with the exception of very large dampers and valves which may still utilize pneumatic damper motors as large electronic motors are still relatively expensive. On most of today’s commercial and institutional pro− jects, the TAB technician may find that all HVAC con− trols are now based on micro−electronics. However, we are still including a review of pneumatic control basics in this chapter as these systems still exist and may be encountered during HVAC system renovation and expansion. 4.1.2.3

Electronic Controls

The term ?electronic controls" was first used to de− scribe newer control technology being installed to re− place some of the functions of the older pneumatic control devices. Most of these first generation elec− tronic controls did little more than monitor HVAC sys− tem operation and provide on/off control of fans, valves, and dampers. The control ?logic" was special− ized software operating on a large main frame central computer, communicating with field interface panels connected to temperature sensors and relays. Any pneumatic controlled dampers or valves were still op− erated by their original pneumatic controls, with E/P relays switched between fixed setpoints. These earlier electronic control systems were of little interest to the TAB technician other than requesting an unseen con− trol operator to start or stop a fan. During the 1990’s this older technology was replaced by DDC which no longer required the pneumatic con− trol devices to carry out the positioning of dampers, valves, and setpoint controllers. In addition, with the advent of micro−electronics, the operating program software is now contained within the remote field pan− els and the large central computer is no longer re− quired. Unlike the earlier electronic control systems, this new DDC technology does have a significant impact on the balancing contractor on all but the smallest projects. HVAC system manufacturers are finding that it is less expensive to use these easily programmed ?black box" control devices. There are no longer pneumatic control devices that must be constantly calibrated and ad− justed to maintain system reliability. The days of a TAB technician adjusting a pneumatic controller to reposition the damper on a VAV box may

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

4.1


be coming to an end, and today’s TAB technician may find a hand held computer indispensable on a balanc− ing job. What may have required above ceiling linkage adjust− ments while standing on a shaky ladder, can now be ac− complished by plugging a hand held device into the nearest electronic wall thermostat. The TAB techni− cian can instantly verify high and low air flows, adjust operating setpoints, and monitor room conditions as the HVAC system responds to these control input changes. It is very important for any SMACNA contractor en− tering today’s TAB field to have a good understanding of DDC basics. 4.1.2.4

Self-Contained Controls

Self−contained controls differ from the above types in that they do not use an external source of power, but develop their own power. Often used in automatic valves, a bellows or other sensing element has enough strength to move the valve. Because of strength and large mass involved in its construction, it is not capa− ble of providing as close control as other types of sys− tems. Applications include:

   

condenser water regulating valves on refrig− eration compressor units (city water). thermostatic expansion valves. steam control valves for heating domestic hot water. self−contained radiator control valves.

Other combinations are electro−hydraulic, commonly applied to valve operators, and electro−pneumatic sys− tems using electronic devices to sense temperature and pressure, and pneumatic devices to operate valves and dampers. This dual system combines advantages of electronic systems (sensitivity, wide range of adjust− ability) with simplicity of pneumatic operators. 4.1.3

Control Categories

Control systems also are divided into two categories: operating controls and safety or limit controls. 4.1.3.1

Operating Controls

Operating controls are used for the control of room conditions and system setpoints. The most common example of an operating control is a room thermostat. 4.2

4.1.3.2

Safety Controls

Safety or limit controls are used to provide safe equip− ment operation. Safety or limit controls must be set properly to avoid unsafe conditions such as pressures or temperatures that are too high or too low, and imple− ment emergency equipment shut off. Safety or limit controls may interrupt the operating controls at any given time to ensure safe system opera− tion. Examples of safety controls are freeze stats, fire stats, flow switches, smoke detectors, and refrigera− tion high−low pressure cutouts. 4.2

CONTROL LOOPS

No matter which type of control system is used, all control applications must involve a fundamental con− trol loop. A control loop consists of three components:

  

a controller (thermostat) a controlled device (valve, damper) a sensing device (transmitter, bi−metal strip)

For example, a sensing device (remote bulb) monitors the temperature of a supply air duct and sends a signal to the controller. The controller monitors the signal as sent by the sens− ing device, and reacts by either opening or closing a controlled device (valve or damper). As a result of the resulting change in system output, (such as hot water in a heating coil), the action of the controlled device creates a change in the sensing device which provide feedback to the controller that the system change took place. The operation of any control loop is continuous during normal operation of the HVAC system. 4.2.1

Controllers

Controllers (such as thermostats and humidistats) have two possible sets of control actions:

  4.2.1.1

Modulating or two position Direct or reverse acting Modulating/Two Position

Modulating control (also called proportional control) is obtained when the control signal sent by the control− ler to the controlled device is constantly changing in small increments to gradually increase or decrease the capacity of a system component to suit the load condi− tions. Two position control, which also can be on−off,

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


only assumes two positions; fully open or fully closed. Two position controllers are normally used with equip− ment that only operates in an on−off application. Equipment such as gas valves or small air conditioning compressors would fall under this category. 4.2.1.2

Two−way valves are normally used for water or steam service. Three−way mixing and diverting valves are used only in hydronic piping (Figure 4−2). The valve constant or flow coefficient is used to calculate the flow and pressure drop of ATC valves in the wide open position.

Direct/Reverse Acting Equation 4-1 (I-P)

Modulating system control devices may increase or decrease the output (branch) control signal with changes in space conditions monitored by the sensing element. A direct acting controller will increase the output (branch) control signal as the controlled vari− able (temperature, humidity, pressure) increases. A re− verse acting controller increases the control signal as the controlled variable decreases. The action of the controller must be properly matched with the control device or the control loop will produce unexpected re− sults or those opposite of that desired.

DP 

CQ

2

v

Where: DP  Pressuredifferential(psi) Q  Flowthroughvalve(gpm) C v  FlowCoefficient

Equation 4-1 (SI)

DP 

KQ

2

v

The position of a controlled device when de−energized is considered the normal position. Control devices such as valves or dampers are either normally open (N.O.) or normally closed (N.C.). Some electric de− vices also contain switches that are normally open or normally closed until moved to the opposite position by a controller.

Where: DP  Pressuredifferential(kPa) Q  Flowthroughvalve(Ls) K v  FlowCoefficient

4.2.2

A control valve must be selected to control a flow of 20 gpm at a maximum 4 psi pressure drop. Calculate the Cv of the valve.

Controlled Devices

Controlled devices that affect the TAB technician the most are automatic control dampers and automatic control valves. Both affect flow and both can be two position or modulating.

Example 4−1 (I−P)

Solution 4.2.2.1

CQ ; DP  CQ 2

ATC Valves

DP 

v

Figure 4−1 illustrates the throttling characteristics of the different types of modulating ATC valves. Two position valves such as those used for automatic shut− off in seasonal change−over piping need no specific throttling characteristic, the major concern being tight shutoff. Gate

Butterfly

100%

v

Q C v    20  10 DP 4 Example 4.1 (SI) A control valve must be selected to control a flow of 1.3 L/s at a maximum 28 kPa pressure drop. Calculate the Kv of the valve.

Ideal Straight Line Characteristic Throttling Plug Travel 50%

Globe

Solution

KQ ; DP  KQ 2

50%

100%

DP 

% Flow

FIGURE 4-1 VALVE THROTTLING CHARACTERISTIC COMPARISON

K v 

v

v

Q   1.3   0.25 D P 28

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

4.3


IN

OUT

OUT

IN

OUT

IN 3 Way Mixing

3 Way Diverting

FIGURE 4-2 ATC VALVE ARRANGEMENTS For proper control action, it is desirable for an ATC valve to be sized so the pressure drop cross the wide open valve at design flow rate will give an appreciable pressure drop. For example, in the case of a steam valve, it is considered good practice for the pressure drop at design flow to be approximately 50 percent of the absolute steam pressure available at the valve. 4.2.2.2

closed, produces a control characteristic that is unsat− isfactory. Opposed blade dampers do not eliminate the above problems, but they improve the control ability by clos− ing blades toward each other so that throttling begins sooner. Close and accurate control is improved but still limited. The linear operating characteristic is not as

ATC Dampers

Dampers used for automatic temperature control have either parallel or opposed blades as shown in Figure 4−3. Quality, tight fitting dampers with long lasting blade edge seals or the equivalent are necessary for ATC work. Parallel blade dampers are almost always used for two position or open−closed control. Opposed blade damp− ers are used for modulating control of airflow. Damp− ers present throttling problems similar to valves which is difficult to correct. Parallel blade dampers often have a throttling characteristic which is worse than gate valves. This deficiency is complicated by the pro− cedure of selecting damper sizes based on low air ve− locities across the dampers. The play in the dampers and damper linkages caused by flexibility and distor− tion, and the difficulty of seating the blades when 4.4

PARALLEL OPERATION

OPPOSED OPERATION

FIGURE 4-3 TYPICAL MULTIBLADE DAMPERS

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


nearly achieved as in the case of valves, but may be more closely approximated by sizing the dampers on higher velocities and by providing more sections and more rigidly constructed.

the sequence of control and the ATC diagrams would indicate which devices to inspect.

4.2.2.3

Most control system sensors and controllers are linear (which means straight line). Figure 4−1 indicates the error induced by non−linear control devices such as dampers and valves. Linear control in pneumatic sys− tems may translate to one degree of temperature change from one psi of air pressure change. One psi (kPa) or fluid pressure change on the discharge side of a pump also could result in a two or three psi (kPa) of control pressure change. These systems are linear as long as each increment of controlled variable produces the same increment of signal. The system would be non−linear if different amounts of signals emanate from a fixed increment of the controlled variable. For example, a system is non−linear if at 70F (21C), a one degree change produces a one psi (kPa) control signal; but at 90F (32C), a one degree change pro− duces a two psi (kPa) control signal.

Valve and Damper Linkages

The operation of automatic valves can often be re− versed in the field to suit the action of the controllers, although in some cases it may be necessary to change the operator. The action of automatic dampers can usu− ally be reversed by resetting the damper arm or other parts of the linkage. Sometimes it is necessary to limit the travel of dampers or valves in order to provide proper control. Most pneumatic and electric devices have operator stops that can be adjusted so that the valve or damper operator is permitted to complete a portion of its stroke. Electric operators have limit switches which can be positioned to electrically stop the operator at a desired position. The correct setting of stops and/or limit switches is important for the suc− cessful operation of a system, and personnel must un− derstand such adjustments, or equipment could be damaged. Although adjustments of ATC valves and dampers on larger systems are normally made by the Temperature Control Contractor, the TAB technician needs to un− derstand the factors required for proper valve and damper adjustment. 4.3

4.4

CONTROL RELATIONSHIPS

An actuator is considered linear if it has a signal range of ten psi (kPa) from fully open to fully closed. So a five volt or five psi (kPa) signal will cause a 50 percent travel. However, if the actuator device is used on a valve or damper, an actuator change of 50 percent will seldom change the fluid flow by the same 50 percent. From Figure 4−1, one can see that a 50 percent stem travel of a gate valve from wide open will have little effect on the fluid flow.

CONTROL DIAGRAMS Equation 4-2 (I-P)

In a typical job specification, there are general descrip− tions of various types of control applications , called the sequence of controls, which the automatic Temper− ature Control Contractor must translate into a set of drawings called control diagrams. These control dia− grams and related written sequence of controls for each HVAC system, are used by the ATC contractor for control system installation in coordination with the HVAC system contractor. The data found in these dia− grams is extremely important to the TAB technician and these diagrams frequently are the only description of how a complicated HVAC system will operate. Control system diagrams also can be used to assist the TAB technician in troubleshooting. For example, when the hand−off−automatic switch of a fan motor starter is in the automatic position, it is found that the fan will not run. But the fan will run when the switch is in the hand position. This indicates that some type of automatic temperature control device or safety switch is preventing the fan from running. A review of

C v  Q

(2.3)½   Q (H)½

2.3 H

Equation 4-2 (SI) Q K v  DP

Equation 4-3 Q21 H1 DP 1   2  H2 Q2 DP 2 Where: Cv Kv Q P H

= = = = =

Flow coefficient or valve constant (I−P) Flow coefficient or valve constant (SI) Fluid flow rateCgpm (L/s) Pressure differenceCpsi (kPa) Head loss or pressure dropCfeet (meters)

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

4.5


Equations 4−2 and 4−3 indicate that the pressure drop across the valve is proportional to the square of the fluid flow rate. This relationship is indicated by the general curve shown in Figure 4−1, which can apply to most systems, although the numbers may vary. The non−linearity of the controlling device is apparent with this curve. In order to minimize the resulting control inaccuracies, the controller and the controlled device must be carefully matched so that an average linearity is achieved. This cannot be done across the entire range of the device, therefore, the devices are matched for a normal operating range, which is a matter of judg− ment of the system designer or the ATC Contractor.

trol system is required. Outside air and exhaust air dampers usually are interlocked with the supply fan to open to a fixed minimum outside air position when the fan is started. A mixed air temperature sensor could then control the outside air, return air, and exhaust air dampers to maintain a set mixed air temperature. At a pre−set temperature or a high outside air humidity, the outside air damper often will be returned to a mini− mum position to decrease the cooling load of the out− side air. In a case of power or control system failure, the outside air damper usually closes automatically. A freeze stat also can stop the fan and close the outside air dampers.

4.5

4.6.2

ATC SYSTEM ADJUSTMENT

After completion of the physical installation, the ATC system components must be adjusted and calibrated so that they may operate individually and collectively to provide the specified environmental system control. The amount of adjustment and calibration will depend on the complexities of the ATC system. All calibration of ATC system instruments should have been done by the ATC installer prior to system balan− cing. However, there are some specific adjustments which should be done in conjunction with TAB per− sonnel during system adjustment and balancing. Fail− ure to provide this coordination may lead to the inabil− ity of the HVAC system to perform satisfactorily under load. Setting automatic dampers for proper air quantities, positioning hot and cold deck dampers, and maintain− ing valves open or closed to maintain design operating conditions are among the multitude of factors which affect systems operation and TAB work. After the installation has been completed, accepted, and the building occupied, problems can arise which may or may not be attributable to the TAB work. It is not unusual for accidental maladjustment of controls to produce symptoms which seem to point to improper HVAC system balancing. Outside air dampers that have slipped, or a reheat coil thermostat that malfunc− tions, are excellent examples. The ability to recognize the real source of the problem not only saves time but vindicates the TAB work. 4.6

TAB/ATC RELATIONSHIP

4.6.1

Related Problems

To properly balance and adjust any HVAC system, a thorough knowledge of the installed temperature con− 4.6

Controllers

A thermostat in the duct system often will control a heating or cooling coil valve, face and by−pass damp− ers or mixing dampers. A room thermostat can control a hot water, steam, chilled water or electric booster coil, and/or hot and cold mixing dampers. A humidis− tat can control a humidifier or a cooling coil for dehu− midification. Controls can be direct acting, reverse acting, modulating or two position, stepped, master, sub−master, series, or parallel. Controls can actuate dampers, valves, and relays; start, modulate or stop motors, fans, and other equipment; and be controlled by time clocks, time delay relays, static pressure con− trollers, air switches, flow switches, level controllers, fire and smoke detectors. They can be connected to alarm systems, and be controlled, readjusted, and be indicated from or at remote control panels. Finally, controls can be very simple or very complex. 4.6.3

Ventilation Air

Probably the most important effect on TAB work is the setting of the outside air, return air, and exhaust air dampers. After the fan speed (rpm) and airflow capac− ity have been checked out, set the outside air dampers for minimum outside air. Use thermometers or a ther− mocouple to measure outside air, return air and mixed air temperatures. Use the mixed air temperature Equa− tion 4−4 to determine the amount of outside air. Work− ing with the temperature control contractor, set the minimum outside air conditions. Mark the dampers for the minimum position and recheck the air flows. If an economizer cycle is used, next check using 100 percent outside air and again check air flows. Follow the procedure for 25, 50, 75 percent outside air. The mixing of the return air with outside air to give a known mixed air temperature can be determined by the following:

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


X oT o  X rT r T m  100

Equation 4-4

Tm = Temperature of the mixture of return air− and outdoor air Xo = Percentage of outdoor air Xr = Percentage of return air To = Temperature of outdoor air, F (C) Tr = Temperature of return air, F (C) Being familiar with the interactions and functions of these control systems will go a long way in reducing on site system balancing time. The following equations are used for determining per− centages of outside air. For this work, more convenient forms of expressing Equation 4−4 are given in Equa− tion 4−5 and 4−6.

(Tr  Tm) X o  100 (T r  T o)

(Tm  To) X r  100 (Tr  To)

Equation 4-5

Example 4.2 (SI) 24C return air is mixed with −4 outside air and the mixed air temperature is 13C. Find the percentage of outside air.

Solution (Tr  Tm) (T r  T o) (24°C  13°C)  100 [24°C  ( 4°C)]  100  11°C  39.3% 28°C X o  100

4.7

CENTRALIZED CONTROL SYSTEMS

The concept of centralized controls or energy manage− ment systems (EMCS) briefly addressed in the begin− ning of this Chapter is applied to many buildings being constructed or modernized today.

Equation 4-6

Example 4−2 (I−P) 75F return air is mixed with 25F outside air and the mixed air temperature is 55F. Find the percentage of outside air.

Solution (Tr  Tm) (T r  T o) (75°  55°)  100 (75°  25°)  100  20°  40% 50° X o  100

FIGURE 4-4 DESKTOP COMPUTER DISPLAYING STATUS OF BUILDING HVAC SYSTEMS

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

4.7


Figure 4−4 shows a standard desktop computer being used to monitor and control all HVAC systems in a hos− pital. These software programs are becoming very easy to use and can display photos and diagrams of a given system, with all ?live" temperature and setpoint data displayed next to each system component. Almost all new buildings will utilize a computerized HVAC control system. Even the most basic dial time clock for start/stop control have been replaced by less expensive electronic time clocks that can adjust for seasonal length of daylight and changing outdoor tem− peratures. Large HVAC systems now have one or more field control panels that include programmable com− puterized memories containing all of the sequence of control logics and control algorithms for the systems and devices being controlled. These field panels include four types of control inputs and outputs, plus the ability to communicate local con− ditions and setpoints to other field panels or remote monitoring locations.

Digital Input Digital Output Analog Input Analog Output

4.7.1

4.7.2

In addition to these input and output field devices, a typical centralized computer control system consists of one or more field interface devices, and one or more programmable stand alone controllers as shown in Fig− ure 4−5. Each programmable stand alone controller contains a microcomputer and battery backed up memory con− taining all of the programming for all field devices and

SENSORS

Many field interface panels include a manual/auto switch for each control output. This allows local by− pass of the control system during system testing and balancing, but bypassing any controls should be autho− rized by the system operator and all switches returned to ?auto" mode when the TAB work is complete. Digital Input

A digital input is an on/off, open/closed, hi/low, or oth− er two position feedback signal input to the computer− ized control system from the HVAC equipment being monitored.

Four basic EMCS signals:

   

HVAC systems it controls. After this software has been created on a desktop computer, it is ?downloaded" to each field controller. Since the actual sequence of con− trols and operating schedules reside in the field stand alone controller, the central control computer is only needed when operating schedules or system setpoints need to be changed by the operator, or to display sys− tem alarms sent from the field controllers. Since the programmable stand alone controllers are a micro− computer device, a field interface device provides re− lays and analog to digital transducers which allows the tiny electronic circuits to control larger current field devices like motor starters and damper operators.

Digital Output

A digital output is an on/off, open/close, hi/low, or oth− er two position control signal command output from the computerized control system to the HVAC system being controlled. 4.7.3

Analog Input

An analog input is a variable feedback signal to the computerized control system from the HVAC system being monitored. It could indicate the position of a

DESKTOP COMPUTER FIELD INTERFACE DEVICES & RELAYS

PROGRAMMABLE STAND ALONE CONTROLLER(S)

ACTUATORS

FIGURE 4-5 FUNCTIONAL BLOCK DIAGRAM A CENTRALIZED COMPUTER CONTROL SYSTEM 4.8

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


valve or damper, speed of a fan or pump, or a room temperature or humidity. 4.7.4

Analog Output

An analog output is a variable command signal output from the computerized control system to the HVAC system being controlled. This signal could be adjust− ing a 0 to 20 psi pressure to a valve or damper motor, or a 0 to 10 volt signal to a variable speed fan motor drive, or other variable signal to match the device be− ing controlled. 4.7.5

EMCS Communications

In addition to the above control inputs and outputs, al− most all computerized building control systems have the ability to communicate with other field panels or central monitoring displays and alarm printers.

vision of the Centralized Control installer or ATC con− tractor. Also, readings obtained from centralized sys− tems can be used by the TAB technician to balance the HVAC system being controlled. 4.7.7

How an EMCS Helps TAB Work

A person entering the TAB field may be wondering what all this has to do with the testing and balancing field. Prior to micro−electronics and DDC, all HVAC sys− tems were controlled by pneumatic devices. Figure 4−6 shows a typical pneumatic control cabinet for a large air handling unit. Notice the high concentration of ?spaghetti" tubing which are used to interconnect all of the pneumatic control devices including pneu− matic relays, receiver/controllers, and various pneu− matic logic controls.

Originally, this communication was by direct hard wire or dial up phone connection, but the trend today is towards less manufacturer specific and more open communication ?protocols," allowing any computer on the same computer ?network" having the proper software and access codes to view and/or change any HVAC system on the same network. This system of in− terconnect also reduces the need for multiple sensors. For example, one outside air sensor can now have its present temperature reading accessed by all air han− dling units and their controls for outside temperature reset instead of a separate sensor for each. 4.7.6

EMCS Points List

In addition to the control contractor providing a writ− ten sequence of controls and related control diagrams, the control documentation for a computerized automa− tion control system should also include a ?points list." This list is actually a table or chart, which at a glance indicates each physical piece of equipment being con− trolled down the ?y" axis, and the different types of software programs utilized across the top ?x" axis. The type of points, analog input (AI) analog output (AO), digital input (DI), and digital output (DO) is also indi− cated. By placing an ?x" or ?dot" where each column and row intersect, it is easy to see the interaction of the physical world with the software programming. Reviewing this points list is helpful to the TAB technician to under− stand how a system is being operated. The TAB technicians should not attempt to adjust or change control settings except when under the super−

FIGURE 4-6 HVAC CONTROLS PANEL WITH ORIGINAL PNEUMATIC CONTROLS.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

4.9


Unfortunately, many times this calibration has not been completed prior to the TAB work. For this reason, if you have access to the automation system display during the balancing work, be sure to record air and water flows being displayed at the time of your own measurements and advise the automation system con− troller. This is especially critical on variable air volume sys− tems since out of calibration automation flow stations can cause an air handling unit that was just balanced to produce much higher or lower flow rates than in− tended. Figure 4−8 shows a TAB technician using a small por− table laptop computer to adjust the minimum and max− imum air flows on an above ceiling VAV box. Note how the computer is ?plugged" directly into a jack that is provided under each electronic wall thermostat con− nected to a DDC system. The computer screen is displaying actual VAV box dis− charge air cfm, discharge air temperature, percent damper position, reheat coil discharge temperature, branch duct supply temperature, duct static pressure, maximum discharge cfm setpoint, and minimum dis− charge cfm setpoint. All of these values can be easily read and adjusted if necessary during system testing

FIGURE 4-7 THE SAME HVAC CONTROL PANEL AFTER UPGRADING TO DIRECT DIGITAL CONTROL (DDC).

Figure 4−7 shows the same control cabinet after all of the pneumatic controls and control tubing were re− placed with a DDC system. Now changing discharge air temperature from the unit or adjusting the outside air damper is as simple as moving the ?mouse" across the control screen in the building manager’s office. Many building automation systems include air flow and water flow measuring stations in addition to the many temperature and humidity sensors. Before being tempted to use these easy to read values, keep in mind that these control input devices require calibration in the software program which may not have been com− pleted. On new construction projects, many control contractors will install all of these field devices using factory ?default" settings. This allows faster system startup and the default valves are acceptable for initial startup operation. 4.10

FIGURE 4-8 PORTABLE COMPUTER PLUGGED INTO ELECTRONIC WALL THERMOSTAT DURING SYSTEM BALANCING.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


and balancing without the need to climb a ladder or re− move ceiling tiles and access covers. Since each control manufacturer may have their own custom software and plug−in thermostat to computer cables, most TAB technicians have developed a good working relationship with the control system installers

who may be able to provide these programs and cables at little or no cost. The ability of a TAB technician to use these portable control devices can reduce the time the control contractor needs to be on site during the TAB work, and in return, the TAB technician does not need to schedule his site times to meet the availability of the control contractor.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

4.11


THIS PAGE INTENTIONALLY LEFT BLANK

4.12

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


CHAPTER 5

FANS


CHAPTER 5 5.1

FAN CHARACTERISTICS

5.1.1

Introduction

This chapter will apply the basic airflow fundamentals discussed in Chapter 2 HVAC Fundamentals to fans. As stated earlier, each type of system needs a pump to overcome the friction and dynamic losses of the sys− tem. This device can be either a centrifugal pump, a fan, a compressor, a turbine, or some other sophisti− cated device. Therefore, each device must be studied by the TAB technician so that not only all of its unique characteristics are known, but that the device has been applied properly within the system, and that the system has been designed to circulate fluid in the most eco− nomical manner and to provide maximum comfort. With an in−depth understanding of HVAC fans and their relationship to HVAC systems, it becomes easy for the TAB technician to apply proper balancing pro− cedures in the correct sequence when on the job. Centrifugal Fans

Three basic types of fans are used in HVAC systems, the centrifugal fan or blower, the axial flow fan, and special designs using fans or blowers in different hou− sings. The airflow within the centrifugal fan is sub− stantially radial through the wheel, while the airflow through the axial flow fan is parallel to the fan shaft. The components of centrifugal fans are identified in Figure 5−1. The three variations of the centrifugal fan used in HVAC work are forward curved, backward in− clined, and airfoil. 5.1.2.1

*SCROLL SIDE SCROLL PIECE SIDE SHEET SIDE PLATE

*OUTLET DISCHARGE

*BACKPLATE HUB DISK HUBPLATE *BLADES FINS INLET CONE INLET RING INLET BELL INLET FLARE INLET NOZZLE VENTURI

*SCROLL CASING HOUSING *IMPELLER WHEEL SCROLL HOUSING *RIM VOLUTE MOTOR SHROUD WHEEL RING WHEEL CONE *SUPPORTS RETAINING RING STIFFENERS INLET RIM *INLET COLLARWHEEL RIM INLET SLEEVEFLANGE INLET BAND INLET PLATE * PREFERRED NOMENCLATURE

PEDESTAL

FIGURE 5-1 CENTRIFUGAL FAN COMPONENTS shape of its performance curve, which allows the pos− sibility of overloading the motor, if system static pres− sure decreases. It also is not suitable for material han− dling because it has an inherently weak structure. Therefore, FC fans are generally not capable of the high speeds necessary for developing higher static pressures. STATIC PRESSURE CURVE STATIC EFFICIENCY CURVE BHP CURVE

100

Forward Curved (FC) Fans

The FC centrifugal fan turns at a relatively slow speed and generally is used for producing high airflow vol− umes at low static pressures. The FC fan will surge, but the magnitude is less than for other types. The static pressure proportion of the total pressure discharge is 20 percent while the velocity pressure is 80 percent. Typical operating range of this type of fan is from 30 percent to 80 percent wide open volume (see Figure 5−2). The maximum static efficiency of 60−80 percent generally occurs slightly to the right of peak static pressure. The horsepower curve has an increasing slope and therefore is referred to as an overloading type fan.

70 SE. SP AND BHP

5.1.2

FANS

0

30

80

100

FIGURE 5-2 CHARACTERISTIC CURVES FOR FC FANS 5.1.2.2

Advantages of the FC fan are its low cost, and the slow speed which minimizes shaft and bearing size, and its wide operating range. Disadvantages include the

0

Backward Inclined (BI) Fans

Backward inclined fans travel at about twice the speed of the FC fan. The normal selection range of the BI fan

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.1


is approximately 40−85 percent of wide open airflow volume (see Figure 5−3). Maximum static efficiency of about 80 percent generally occurs close to the edge of its normal operating range. Generally, using a larger fan will allow greater efficiency for a given selection. The magnitude of surge for a BI fan is greater than for the FC fan.

STATIC PRESSURE CURVE STATIC EFFICIENCY CURVE BHP CURVE

100

STATIC PRESSURE CURVE STATIC EFFICIENCY CURVE BHP CURVE

SE. SP AND BHP

86

SE., SP AND BHP

100

80 0

50

CFM

85

100

FIGURE 5-4 CHARACTERISTIC CURVES FOR AIR FOIL are shown in Figure 5−4. For a specific application, the airfoil fan has the highest rpm of the three centrifugal fans. 5.1.3 0

0

40

CFM

85

100

FIGURE 5-3 CHARACTERISTIC CURVES FOR BI FANS Advantages of the BI fan include its higher efficiency and non−overloading horsepower curve. The horse− power curve generally reaches a maximum in the middle of the normal operating range, thus overload− ing is normally not a problem. Inherently stronger de− sign makes it suitable for the higher static pressure op− eration of 70 percent of the total pressure measured at the fan discharge. This leaves the measured velocity pressure at only 30 percent. Disadvantages include the higher speed, which re− quires larger shaft and bearing sizes and places more importance on proper wheel balance; and unstable op− eration, which occurs as block−tight static pressure is approached. 5.1.2.3

Airfoil Fans

A refinement of the flat bladed BI fan is a fan that uses airfoil shaped blades. This improves the static effi− ciency to about 86 percent and reduces noise level slightly. The magnitude of surge also increases with the airfoil blades. Characteristic curves for airfoil fans 5.2

0

Axial Fans

Components of axial fans are illustrated in Figure 5−5. HVAC axial fans may be divided into three groups, propeller, tubeaxial, and vaneaxial.

GUIDE VANE

INLET CONE OR INLET BELL WHEEL ROTOR IMPELLER

MOTOR

BLADE

HOUSING CASING

HUB

NOSE COVER PLATE SPINNER

FIGURE 5-5 AXIAL FAN COMPONENTS 5.1.3.1

Propeller Fans

HVAC propeller fans normally are not connected to duct systems. They are well suited for handling high volumes of air at very low or no static pressures and low efficiencies (see Figure 5−6). 5.1.3.2

Tubeaxial and Vaneaxial Fans

Tubeaxial and vaneaxial fans are simply propeller fans mounted in a cylinder and are similar except for vane−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


STATIC PRESSURE CURVE STATIC EFFICIENCY CURVE BHP CURVE

STATIC PRESSURE CURVE STATIC EFFICIENCY CURVE BHP CURVE

100

SE.SP ANDBHP

SE. SP AND BHP

100

50

80

0 0

0

CFM

65

0

CFM

100

FIGURE 5-6 CHARACTERISTIC CURVES FOR PROPELLER FANS

Tubeaxial fans and vaneaxial fans generally are used for handling large volumes of air at low static pres− sures.

90 100

FIGURE 5-7 CHARACTERISTIC CURVES FOR VANEAXIAL FANS (HIGH PERFORMANCE) 5.1.4.1

type straighteners on the vaneaxial. These vanes re− move much of the swirl from the air and improve the efficiency. A vaneaxial fan is more efficient than a tu− beaxial fan and can reach higher pressures. Note that with axial fans the brake horsepower (BHP) is maxi− mum at the blocktight static pressure (see Figure 5−7).

65

Tubular Centrifugal Fans

Tubular centrifugal fans, illustrated in Figure 5−8, gen− erally consist of a single width airfoil wheel arranged in a cylinder to discharge air radially against the inside of the cylinder. Air is then deflected parallel with the fan shaft to provide straight−through flow. Vanes are used to recover static pressure and to straighten air flow.

SW CENTRIFUGAL FAN WHEEL

Advantages of tubeaxial fans and vaneaxial fans in− clude the reduced size and weight, and the straight− through airflow which frequently eliminates elbows in the ductwork. The maximum static efficiency of an in− dustrial vaneaxial fan is approximately 65 percent. The operating range for axial fans is from 65 percent to 90 percent. Disadvantages of axial fans include high noise levels and efficiencies lower than those of centrifugal fans. 5.1.4

Special Designs

There are variations of both centrifugal and axial fans that are designated special design fans. These include tubular centrifugal fans and power roof ventilators.

STREAMLINE INLET

AIR OUT AIR IN

STRAIGHTENING VANES

FIGURE 5-8 TUBULAR CENTRIFUGAL FAN Characteristic curves are shown in Figure 5−9. The selection range is generally about the same as the scroll type BI or airfoil bladed wheelC50 to 85 per− cent of wide open volume. However, because there is

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.3


limitations of the wheels, bearings, and housing of fans. Under the most recent class standards, there are three classifications, as shown in Figures 5−10 and 5−11.

STATIC PRESSURE CURVE STATIC EFFICIENCY CURVE BHP CURVE

Note the line of demarcation between Class I and Class II construction.

100

SE. SP AND BHP

Example 5.1 (I−P) A fan operates at 9056 cfm, 1478 rpm, requiring 5.08 BHP at 1.0 in. wg SP. The airflow must be increased to 10,188 cfm to handle an additional load. Find the new SP and BHP.

70

15 14 13 1/2” @ 3780 13 RATINGS MAY BE PUBLISHED IN THIS UPPER RANGE

12

0

50

85

100

CFM

FIGURE 5-9 CHARACTERISTIC CURVES FOR TUBULAR CENTRIFUGAL FANS no housing of the turbulent air flow path through the fan, static efficiency is reduced to a maximum of about 72 percent and noise level is increased.

(SP)INCHES OF WATER

0

10

STATIC PRESSURE

11

6

TYPICAL CLASS II CHARACTERISTIC CURVE MINIMUM PERFORMANCE CLASS III

9 8 1/2” @ 3000 8

CLASS III SELECTION ZONE

7 6 3/4” @ 5260

5” @ 2300 5 4 3

CLASS I SELECTION ZONE

1000

2000

5000

6000

7000

3750 3500 3375 Pa @ 18.9 3250 RATINGS MAY BE PUBLISHED IN THIS UPPER RANGE

3000 2750

STATIC PRESSURE (SP)-P ASCALS (Pa)

Fan Classes

4000

FIGURE 5-10 FAN CLASS STANDARDS (I-P) (SW BI FANS)

TYPICAL CLASS II CHARACTERISTIC CURVE

2500

5.2.1

3000

OUTLET VELOCITY (OV) FEET PER MINUTE

Power roof ventilators allow the air to discharge in a full circle from the impeller, which may be either cen− trifugal or axial with similar characteristics. A large advantage is that they provide positive exhaust ven− tilation over gravity ventilators.

FAN CONSTRUCTION

RATINGS MAY BE PUBLISHED IN THIS LOWER RANGE

1

Power Roof Ventilators

5.2

4 1/2” @ 4175

2 1/2” @ 3200

Frequently, the straight−through flow results in signifi− cant space savings. This is the main advantage of tubu− lar centrifugal fans.

Disadvantages include lower available static pressures than centrifugal fans and loss of the discharge velocity pressure component that is recovered.

CLASS II SELECTION ZONE

MINIMUM PERFORMANCE CLASS I

2

5.1.4.2

MINIMUM PERFORMANCE CLASS II

MINIMUM PERFORMANCE CLASS III

2250 2125 Pa @ 15.0 2000

CLASS III SELECTION ZONE

1750 1563 Pa @ 26.3 1500 MINIMUM PERFORMANCE CLASS II

1250 Pa @ 11.5 1250 1000

MINIMUM PERFORMANCE CLASS I

750

1063 Pa @ 20.9

CLASS II SELECTION ZONE

625 Pa @ 16.0

When using a fan rating table published by the fan manufacturer, if fan speeds and static pressures in− crease above certain given conditions, the class of the fan changes. Class again refers to an Air Movement and Control Association, Inc. (AMCA) standard which has been developed to regulate actual structural 5.4

500

CLASS I SELECTION ZONE

RATINGS MAY BE PUBLISHED IN THIS LOWER RANGE

250

5

10

15

20

25

30

OUTLET VELOCITY-METRES PER SECOND (m/s)

FIGURE 5-11 FAN CLASS STANDARDS (SI) (SW BI FANS)

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

35


SW—Single Width SI—Single Inlet

DW—Double Width DI—Double Inlet

Arrangements 1, 3, 7 and 8 are also available with bearings mounted on pedestals or base set independent of the fan housing. For designation of rotation and discharge (see Figure 5–17) For motor position, belt or chain drive (see Figure 5–16)

ARR. 2 SWSI For belt drive or direct connection. Impeller overhung Bearings in bracket supported by fan housing.

ARR. 3 SWSI For belt drive or direct connection. One bearing on each side and supported by fan housing.

ARR. 1 SWSI For belt drive or direct connection. Impeller overhung. Two bearings on base.

ARR. 3 DWDI For belt drive or direct connection. One bearing on each side and supported by fan housing.

ARR. 4 SWSI For belt drive. Impeller overhung on prime mover shaft. No bearings on fan. Prime mover base mounted or integrally directly connected.

ARR. 7 SWSI For belt drive or direct connection. Arrangement 3 plus base for prime mover.

ARR. 7 DWDI For belt drive or direct connection. Arrangement 3 plus base for prime mover.

ARR. 8 SWSI For belt drive or direct connection. Arrangement 1 plus extended base for prime mover.

ARR. 9 SWSI For belt drive. Impeller overhung, two bearings with prime mover outside base.

ARR. 10 SWSI For belt drive. Impeller overhung, two bearings with prime mover inside base.

FIGURE 5-12 DRIVE ARRANGEMENTS FOR CENTRIFUGAL FANS HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.5


THIS PAGE INTENTIONALLY LEFT BLANK

5.6

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


VEL

OUT VEL

Press

CFM

FPM

H20

2264 2547

800 900

2830 3113

VOL

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

0.125

SP

0.250

SP

0.375

SP

0.500

SP

0.625

SP

0.750

SP

0.875

SP

1.000

SP

1.250

SP

1.500

SP

1.750

SP

2.000

SP

RPM

BHP

RPM

BHP

RPM

BHP

RPM

RPM

RPM

BHP

RPM

BHP

RPM

BHP

RPM

BHP

RPM

BHP

BHP

BHP

RPM

BHP

RPM

BHP

0.04 0.05

398 434

0.10 0.13

456 487

0.15 0.19

507 536

0.21 0.25

567 578

0.26 0.30

608 624

0.32 0.37

656 669

0.40 0.44

703 712

0.47 0.51

747 755

0.55 0.60

835

0.78

1000 1100

0.06 0.08

472 510

0.17 0.21

519 552

0.23 0.27

565 595

0.29 0.34

608 636

0.36 0.42

645 675

0.42 0.49

686 708

0.49 0.56

727 745

0.57 0.63

767 782

0.65 0.71

843 855

0.83 0.89

916 924

1.03 1.10

991

1.31

3396

1200

0.09

549

0.26

587

0.33

627

0.40

666

0.48

702

0.56

738

0.64

768

0.71

802

0.79

870

0.97

936

1.17

999

1.39

1062

1.63

3679

1300

0.11

589

0.32

624

0.39

661

0.47

697

0.55

731

0.64

765

0.73

798

0.81

825

0.89

888

1.07

950

1.26

1012

1.48

1070

1.71

3962

1400

0.12

629

0.39

662

0.46

695

0.54

729

0.63

762

0.72

794

0.81

826

0.91

856

1.01

909

1.17

967

1.37

1026

1.58

1083

1.82

4245

1500

0.14

668

0.46

700

0.54

730

0.62

762

0.72

794

0.81

825

0.91

854

1.01

884

1.12

936

1.30

989

1.50

1043

1.71

1097

1.94

4528

1600

0.16

709

0.55

739

0.63

767

0.72

796

0.81

827

0.91

856

1.02

884

1.13

912

1.23

967

1.46

1013

1.64

1063

1.86

1114

2.09

4811

1700

0.18

749

0.65

778

0.74

805

0.83

832

0.92

860

1.03

888

1.14

915

1.25

942

1.36

994

1.59

1044

1.82

1087

2.01

1134

2.25

5094

1800

0.20

790

0.75

818

0.85

843

0.95

868

1.05

894

1.15

921

1.26

948

1.38

973

1.50

1023

1.74

1073

1.99

1115

2.21

1157

2.43

5377

1900

0.23

830

0.88

857

0.98

882

1.08

906

1.19

930

1.29

955

1.40

980

1.53

1005

1.65

1053

1.90

1100

2.16

1146

2.42

1185

2.64

5660

2000

0.25

872

1.01

897

1.12

921

1.23

944

1.33

966

1.44

989

1.56

1014

1.68

1038

1.81

1084

2.08

1129

2.34

1173

2.61

1217

2.89

5943

2100

0.27

913

1.16

937

1.27

960

1.39

982

1.50

1004

1.61

1025

1.73

1048

1.85

1071

1.99

1116

2.26

1160

2.54

1202

2.82

1245

3.12

6226

2200

0.30

954

1.32

977

1.44

999

1.56

1021

1.68

1042

1.80

1062

1.91

1083

2.04

1104

2.17

1148

2.46

1191

2.75

1231

3.04

1272

3.34

6509

2300

0.33

995

1.50

1017

1.62

1039

1.75

1059

1.87

1080

1.99

1100

2.12

1119

2.24

1139

2.38

1181

2.57

1222

2.97

1262

3.28

1301

3.58

6792

2400

0.36

1037

1.70

1067

1.82

1079

1.95

1099

2.08

1118

2.21

1137

2.34

1156

2.47

1175

2.60

1215

2.90

1255

3.20

1293

3.52

1331

3.84

7358

2600

0.42

1120

2.13

1139

2.26

1159

2.40

1178

2.55

1196

2.68

1214

2.82

1231

2.97

1248

3.10

1284

3.40

1321

3.72

1358

4.06

1393

4.40

7924 8490

2800 3000

0.49 0.56

1204 1287

2.64 3.23

1221 1303

2.78 3.38

1239 1320

2.93 3.53

1257 1337

3.08 3.70

1274 1353

3.23 3.86

1291 1370

3.38 4.02

1308 1385

3.53 4.18

1324 1401

3.69 4.34

1356 1431

3.99 4.67

1389 1461

4.32 5.00

1424 1492

4.67 5.35

1458 1525

5.03 5.73

9056 9622 10754

3200 3400 3600 3800

0.64 0.72 0.81 0.90

1371 1455 1539 1623

3.90 4.66 5.51 6.46

1386 1469 1552 1636

4.05 4.82 5.68 6.64

1401 1483 1566 1648

4.21 4.99 5.86 6.82

1417 1498 1579 1661

4.39 5.16 6.04 7.01

1433 1513 1594 1674

4.56 5.35 6.24 7.21

1448 1528 1608 1688

4.74 5.54 6.43 7.42

1464 1542 1621 1701

4.91 5.72 6.63 7.63

1478 1556 1636 1714

5.08 5.91 6.82 7.84

1507 1583 1661 1740

5.43 6.27 7.20 8.25

1535 1611 1687 1764

5.77 6.64 7.59 8.65

1663 1637 1713 1768

6.13 7.00 7.99 9.06

1593 1664 1737 1813

6.51 7.39 8.37 9.48

11320

4000

1.00

1707

7.52

1719

7.70

1731

7.89

1743

8.09

1755

8.29

1769

8.52

1781

8.74

1794

8.95

1818

9.39

1841

9.80

1865

10.24

1888

10.68

10188

IN

Pressure class limits:

Class I II

Maximum RPM 1550 2140

Table 5-1 Typical Fan Rating Table 5.7


MOTOR LEFT

VIEW FACING DISCHARGE

FIGURE 5-13 ARRANGEMENT 1 IN-LINE FANS for motors too large for fan casing. Arrangement 4 (Figure 5−14)Cdirect drive with wheel overhung on motor shaft.

Solution New static pressure

 1.0  10188 9056 2

cfm2 P 2  P1  cfm1

2

 1.27in.8 8 wgSP New brake horsepower:

cfm

 5.08  10188 cfm 9056 3

BP 2  BP1 

3

2 1

BP 2  7.23BHP A review of Table 5−1 not only confirms the calcula− tions, but also indicates that the change in capacity moved the fan into a different pressure classification which could result in a failure of the fan wheel and/or bearings. CAUTION)Always check with the published rat− ings of fan equipment to make sure that revised op− erating conditions do not require a different class fan. Often this type of change also could change the pressure classification of part or all of a duct system to higher duct construction and sealing require− ments. 5.2.2

Fan Nomenclature

5.2.2.1

Drive Arrangements

AMCA has developed standard fan drive arrange− ments (shown in Figure 5−12) for various bearing and drive locations. Axial or in−line fans are designated in much the same way as standard centrifugal fans. Stan− dard arrangements for in−line are: Arrangement 1 (Figure 5−13)Cbelt drive with motor mounted independent of fan casingCtypically used 5.8

Arrangement 9 (Figure 5−15)Cbelt drive with motor located on periphery of casing in one of eight standard locations designated by the letters beginning with A at the top and proceeding clockwise at eight equal inter− vals through the letter H when viewing the fan from the discharge. Vertical units are designated as either upblast or down− blast and generally are available only in Arrangements 4 and 9. 5.2.2.2

Motor Arrangement Locations

Motor location is specified at W, X, Y, and Z as shown in Figure 5−16. This motor location always is deter− mined by facing the fan drive sheave. It is independent of the discharge or rotation. 5.2.2.3

Rotation

Rotation is determined by the direction the fan wheel will be turning for proper operation as viewed from the drive side of the fan. Rotation is designated as clock− wise (CW) or counter clockwise (CCW). 5.2.2.4

Non-Sparking Construction

For applications where sparks generated in the air stream could be dangerous, AMCA provides three non−sparking construction classifications based on the degree of assurance desired. For all classes, bearings must be out of the air stream, the fan must be grounded, and non−sparking belts are required. The three classes are: 1.

AMCA A. Requires all components in the airstream be made of nonferrous material.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


VIEW FACING DISCHARGE

FIGURE 5-14 ARRANGEMENT 4 IN-LINE FANS

A B

H

G

C

D

F MOTOR SHOWN IN POSITION A

E VIEW FACING DISCHARGE

FIGURE 5-15 ARRANGEMENT 9 IN-LINE FANS 2.

AMCA B. This requires all components in the airstream be made of nonferrous material. Housing can be steel.

3.

AMCA C. Nonferrous wear ring is required on the inlet cone so that, if the impeller shifts, it will rub the nonferrous material.

Generally, AMCA A is the most expensive and AMCA C is the least expensive. 5.2.3

Fan Motors and Drives

5.2.3.1

General

Most fans are driven at constant speed by constant speed motors, and they generally deliver a constant air quantity. The motors range from small single phase fractional horsepower motors to large polyphase mo− tors. Motors generally are connected to the driven fan by means of a V belt drive which not only transmits power but allows the synchronous speed of the motor such as 1200, 1800, 3600 rpm, to be converted to the lower fan speed. Some small fans have motors directly connected. Some V belt drives have an adjustable speed range by providing a variable drive sheave on which the pitch diameter can be manually adjusted to allow for minor speed variations. Variable air volume (VAV) systems are now commonly used, and some reduce system air− flow by using variable speed motors or drives.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.9


d.

Ratios should not exceed 8:1.

e.

Belt speed, preferably should not exceed 5,000 fpm (25 m/s), or be less than 1,000 fpm (5 m/s). Best practice is about 4,000 fpm (20 m/s).

f.

Sheaves should be dynamically balanced when used for speeds in excess of 5,000 fpm (25 m/s) rim speed.

FAN MOTOR

Z

W Y

X DRIVE

Equation 5-1 rpm(fan) Pitchdiam.motorpulley   rpm(motor) Pitchdiam.fanpulley

FIGURE 5-16 CENTRIFUGAL FAN MOTOR LOCATIONS

5.2.3.3

Drive Installations

When installing or reviewing fan drives, these points should be particularly watched: Of major importance to the TAB technician, is that the V belt drives must be properly aligned before testing, and the belt tension adjusted properly. Too little belt tension results in belt slippage and excessive belt wear. Too much belt tension can cause excessive bearing loading, causing motor bearing or fan bearing failure. One further caution to the TAB technician is that the motor must have sufficient starting torque to over− come the inertia of the fan wheel and drive package. Most HVAC supply air systems do not have this pro− blem. However, in return air or exhaust air systems where design airflow volumes may be high and fan to− tal pressures low, check to assure that the installed mo− tor has sufficient starting torque to accelerate the fan to its design speed. 5.2.3.2

b.

c.

5.10

Be sure that shafts are parallel and sheaves are in proper alignment. Check again after a few hours of operation.

b.

Do not drive sheaves on or off shafts. Wipe shaft, key, and bore clean with oil. Tighten screws carefully. Recheck and retighten after a few hours of operation.

c.

Belts should never be forced over sheaves.

d.

In mounting belts be sure the slack in each belt is on the same side of the drive. This should be the slack side of the drive.

e.

Belt tension should be reasonable. When in operation, the tight side of the belts should be in a straight line from sheave to sheave and with a slight bow on the slack side. All drives should be inspected periodically to be sure belts are under proper tension and are not slipping.

f.

When making replacements of multiple belts on a drive, be sure to replace the entire set with a new set of matched belts.

Drive Design

Regardless of whether drives consist of stock or spe− cial items, there are certain primary conditions to con− sider with respect to the design of satisfactory drives. The conditions most commonly encountered are: a.

a.

Drives should be installed with provisions for center distance adjustment. This is essential, as all belts stretch.

5.3

FAN AIRFLOW AND PRESSURES

5.3.1

Fan Air Volume

Centers should not exceed 2½ to 3 times the sum of the sheave diameters nor be less than the diameter of the larger sheave.

The airflow volume (cfm or L/s) produced by a fan in a given system is independent of the air density.

Arc of contact on the smaller sheave should not be less than 120.

cfm (L/s)Ccubic feet per minute (liters per second) of air handled by a fan at any air density.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Counter Clockwise Top Horizontal

Clockwise Top Horizontal

Clockwise Bottom Horizontal

Counter Clockwise Bottom Horizontal

Clockwise

Counter Clockwise

Counter Clockwise

Clockwise

Up Blast

Up Blast

Down Blast

Down Blast

Counter Clockwise

Clockwise

Clockwise

Counter Clockwise

Top Angular Down

Top Angular Down

Bottom Angular Up

Bottom Angular Up

Clockwise

Counter Clockwise

Counter Clockwise

Clockwise

Bottom Angular Down

Bottom Angular Down

Top Angular Up

Top Angular Up

FIGURE 5-17 DIRECTION OF ROTATION AND DISCHARGE

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.11


scfm (sL/s)Ccubic feet per minute (liters per second) of standard air (0.075 lb/ft3 or 1.2041 kg/m3 density) handled by a fan. 5.3.2

Where it is possible to take field measurements, care must be taken to measure fan total pressure at the fan inlet duct rather than fan static pressure.

Fan Total Pressure (TP)

Fan total pressure is the difference between the total pressure at the fan outlet and the total pressure at the fan inlet. The fan total pressure is a measure of the total mechanical energy added to the air or gas by the fan. This generally can be measured accurately only (as il− lustrated in Figure 5−18) in a test laboratory.

IMPACT TUBE

FAN

FAN

STATIC TUBE

IMPACT TUBE

AIR FLOW

AIR FLOW

SP

FIGURE 5-19 FAN STATIC PRESSURE (SP)

5.3.4 IMPACT TUBE

TP

Fan velocity pressure (Figure 5−20) is the pressure cor− responding to the fan outlet velocity pressure. It is the kinetic energy per unit volume of flowing air.

FIGURE 5-18 FAN TOTAL PRESSURE (TP)

5.3.3

Fan Static Pressure (SP)

Fan static pressure (Figure 5−19) is the fan total pres− sure less the fan velocity pressure. It can be calculated by subtracting the total pressure at the fan inlet from the static pressure at the fan outlet. This is a source of some confusion within the industry, but, by definition: Fan SP = Fan TP (outlet)  TP (inlet)  Vp (out− let) Also, TP (outlet)  SP (outlet) = Vp (outlet)

5.3.5

Fan SP = SP (outlet)  TP (inlet)

Fan Outlet Velocity

Fan outlet velocity is the theoretical velocity of the air as it leaves the fan outlet, and is calculated by dividing the air volume in cfm (L/s) by the fan outlet area in square feet (m2). However, all fans have a non−uni− form outlet velocity; that is, the velocity varies over the cross−section of the fan outlet. Therefore fan outlet velocity as calculated above is only a theoretical value that could occur at a point downstream from the fan. All velocity (velocity pressure) readings, including to− tal pressure and static pressure should be taken down− stream in a straight duct connected to the fan discharge where the flow is more uniform. A large portion of the discharge airflow occurs at the side of the fan outlet farthest from the fan shaft. Veloc− ity readings taken at the side of the duct nearest the shaft, may indicate air appearing to flow from the duct back into the fan. 5.3.6

and substituting:

5.12

Fan Velocity Pressure (Vp)

Fan Brake Power

Fan brake power is the actual power required to drive the fan. It is greater than a theoretical air power be−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Example 5.2 (I−P) Find the tip speed of a 30 inch diameter fan wheel ro− tating at 954 rpm.

PITOT TUBE TOTAL PRESSURE

VELOCITY PRESSURE

Solution TipSpeed  p  D  RPM  p  30  954 12 12 Tip Speed = 7493 (fpm)

STATIC PRESSURE

Example 5.2 (SI) VELOCITY PRESSURE  TOTAL PRESSURE = STATIC PRESSURE

Find the tip speed of a 750 mm diameter fan wheel ro− tating at 954 rpm.

FIGURE 5-20 FAN VELOCITY PRESSURE (VP) Solution cause it includes loss due to turbulence and other inef− ficiencies in the fan, plus bearing losses. Fan brake power is an important value to the TAB technician be− cause it is the power furnished by the fan motor. 5.3.7

Tip8 8 Speed  p  D  RPM  p  0.75  954 60 60 Tip Speed = 37.46 (m/s)

TIP SPEED

Also called peripheral velocity, tip speed equals the circumference of the fan wheel times the rpm of the fan and is expressed in feet per minute (meters per second) (Figure 5−21)

RPM

D

Equation 5-2 (I-P) TipSpeed  p  d  RPM 12in.ft. Where: TipSpeed  Feetperminute D  Wheeldiameter  inches RPM  Revolutionsperminute Equation 5-2 (SI) p  d  RPM TipSpeed  60secmin Where: TipSpeed  Metersperminute D  Wheeldiameter  meters RPM  Revolutionsperminute

FIGURE 5-21 TIP SPEED

5.4

FAN/SYSTEM CURVE RELATIONSHIP

5.4.1

System Curve

Duct system resistance is the sum of all pressure losses through filters, coils, dampers, and ductwork. The sys− tem curve or system resistance curve (Figure 5−22) is a plot of the pressure that is required to move air through the system. For fixed systems, that is, with no changes in damper settings, etc., system resistance

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.13


varies as the square of the airflow. The system curve for any system is represented by a single curve. For exam− ple, consider a system handling 1,000 cfm (500 L/s) with a static pressure (SP) resistance of 1 in. wg (250 Pa).

System Curve BHP Curve (W) Fan Curve

Operating Point

3

(750)

AND

(1000)

SP

4

(500)

BHP

Static Pressure - In wg (Pa)

5

2 1 (250) 0

1000 (500) AIRFLOW - CFM (l/s)

2000 (1000)

CFM (L/S)

FIGURE 5-22 SYSTEM RESISTANCE CURVE

FIGURE 5-23 OPERATING POINT If the airflow is doubled, the SP resistance will in− crease by that ratio squared (4) to 4 in. wg (1000 Pa). This system curve changes, however, as filters load with dirt, coils start condensing moisture, or when bal− ancing dampers are moved to a new position.

reading across the fan and concluding that if it is at or above design requirements, the airflow is also at or above design requirements. 5.4.3

5.4.2

The system operating point (Figure 5−23), a point at which the fan and system will simultaneously perform, is determined by the intersection of the system curve with the fan performance curve for each designated speed (rpm). Every fan operates only along its perfor− mance curve. If the designed system SP resistance is not the same as the SP resistance in the installed sys− tem, the operating point will move along the fan curve and the SP and volume delivered will not be as calcu− lated. In Figure 5−24 the actual duct system has more pres− sure drop then predicted by the system designer. Thus, airflow is reduced because the SP increased. The shape of the horsepower curve typically would result in a re− duction in fan power. Typically, the fan rpm would then be increased, and more fan power would be need− ed to achieve the desired airflow. In many cases, when there is a difference between actual and calculated fan output, the difference is due to a change in system re− sistance rather than to any shortcomings of the fan or motor. Frequently, the mistake is made of taking the SP 5.14

Fan Law Relationships

System Operating Point Fan law equations 2−17 and 2−18 from Chapter 2 ap− plying a change only in fan rpm (with the system re− maining unchanged), are graphically shown in Figure 5−25. Use Equation 2−17 to obtain rpm 2: rpm 2  rpm1 

Q2 Q1

Then use Equation 2−18 to obtain the new static pres− sure:

rpm P 2  P1 rpm 2 1

2

5.4.4

Density

5.4.4.1

When Volume is Constant

The resistance of an HVAC duct system is dependent on the density of the air flowing through the system. Air density at standard conditions of 0.075 pounds per cubic feet (1.2041 kilograms per cubic meter) is used for rating fans in the HVAC industry. A fan is a constant volume machine and will produce the same

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


volume of airflow regardless of the air density being handled (see Figure 5−26). The fan SP and fan power, however, will vary directly as the air density increases or decreases.

SP. The fan is drawing 9.22 Bhp from a 10 HP motor. If the fan is run with the oven off (70F ambient), cal− culate the new SP and Bhp. (Air density at 250F = 0.0563 lb/ft3 ).

Equation 5-3 SP 2 d   2 SP 1 d1

Solution Equation 5-4

FP 2 d   2 FP 1 d1

Using Equations 5−3 and 5−4: d2  2.6  0.075 d1 0.0563 SP 2  3.46in.w.g. d FP 2  FP1  2  9.22  0.075 0.0563 d1 FP 2  12.28Bhp With only a 10 HP motor, a 23 percent motor overload occurs. SP 2  SP1 

Where (airflow1 = airflow2 ): SP  Staticpressure  in.wg(Pa) 3

d  Density  lbft (kgm3) FP  Fanpower  Bhp(W) In other words, the heavier or more dense the air, the greater the fan power or SP will be.

SP @ RPM

2

System Curve

Fan Curve Actual System Curve Design System Curve

SP

2

New Operation Point Change

SP

SP Increase

SP @ RPM 1 SP

1

AIirflow Reduction CFM (L/S)

FIGURE 5-24 VARIATIONS FROM DESIGN AIR SHORTAGE

Example 5.3 (I−P) A 15,000 cfm fan is delivering 250F air from an oven through an air−to−air heat exchanger against 2.6 in. wg

Airflow 1

Airflow 2

FIGURE 5-25 FAN LAW - RPM CHANGE

Example 5.3 (SI) A 7500 L/s fan is delivering 125C air from an oven through an air−to−air heat exchanger against 650 Pa SP. The fan is drawing 6.88 kW from a 7.5 kW motor. If the fan is run with the oven off (20C ambient) calcu− late the new SP and kW. (Air density at 125C = 0.891 kg/m 3).

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.15


Where: (SP1 = SP2 ):

Solution

Q  Airflow  cfm(Ls) d  Density  lbft3(kgm3) RPM  FanSpeed FP  Fanpower  Bhp(W)

Using Equations 6−2 and 6−3: d SP 2  SP1  2  650  1.2041 0.891 d1 SP 2  878Pa d FP 2  FP1  2  6.88  1.2041 d1 0.891 FP 2  9.30kW With only a 7.5 kW motor, a 24 percent motor overload occurs. 5.4.4.2

When Static Pressure is Constant

SP @ d 1

SP1

If the system SP remains constant, the airflow volume, fan speed and fan power will vary inversely as the square root of the density (see Figure 5−27).

Q1   Q2

SP @ d

Chg. 2

Equation 5-5

SP 2

d2 d1

SYSTEM d 1

RPM 1   RPM 2

Equation 5-6 SYSTEM d 2

d2 d1

Airflow1= Airflow 2

FIGURE 5-26 EFFECT OF DENSITY CHANGE (CONSTANT VOLUME) FP 1   FP 2

d2 d1

Equation 5-7 5.4.4.3

Constant Mass Flow

With a constant mass flow rate in a system that remains constant without any changes and using the same fan with a variable drive, the airflow rate, RPM and SP will vary inversely with the air density. The fan brake power will vary inversely with the square of the densi− ty (see Figure 5−28). Equation 5-8 Q1 d   2 Q2 d1 Equation 5-9 RPM 1 d   2 RPM 2 d1 Equation 5-10 SP 1 d   2 SP 2 d1

FP 1 d   2 FP 2 d1

5.16

Equation 5-11 2

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Static Pressure

Example 5−4 SI SP@d1

SP@d 2

A fan is required to handle 7500 L/s at 500 Pa, 75C, 950 rpm, 750 meters altitude and 4.2 kW. Find the fan airflow, SP, RPM and fan brake power that must be se− lected from the fan table.

Solution

Chg.

SP1 -SP2

Using an air density correction factor table found in Appendix A, the correction factor to standard condi− tions is 0.78 or: d 2(Actual)   0.78 d 1(Standard) Airflow (d1< d 2)

Using Equations 5−8 to 5−11:

FIGURE 5-27 EFFECT OF DENSITY CHANGE (CONSTANT STATIC PRESSURE)

d2 d1 Q 1  7500  0.78  5850Ls SP 1(Std)  500  0.78  390Pa RPM 1(Std.) RPM 2(Act.)  d 2d 1 RPM 2  9500.78  1218rpm FP1(Std.) FP 2(Act.)  (d2d 1)2 FP 2  4.2(0.78)2  6.90kW FAN CAPACITY RATINGS Q 1(Std)  Q2(Act.)

Where:  Airflow  cfm(Ls) Q d  Density  lbft 3(kgm3) RPM  Fanspeed SP  Staticpressure  in.w.g.(Pa) FP  Fanpower  Bhp(W) 5.5 Example 5.4 (I−P) 5.5.1 A fan is required to handle 15,000 cfm at 2 in. wg, 150F, 950 RPM, 2000 feet altitude and 5.6 Bhp. Find the fan cfm, SP, Bhp and RPM that must be selected from the fan table.

Solution Using an air density correction factor table found in Appendix A, the correction factor to Standard condi− tions is 0.81 or: d 2(Actual)   0.81 d 1(Standard)

Using Equations 5−8 to 5−11:

Fan Testing

Most fan manufacturers rate the performance of their products from tests made in accordance with ANSI/ AMCA Standard 210, Laboratory Methods of Testing Fans for Rating. The purpose of Standard 210 is to es− tablish uniform methods for laboratory testing of fans and other air moving devices to determine perfor− mance in terms of flow rate, pressure, power, air densi− ty, speed of rotation and efficiency, for rating or guar− antee purposes. Two basic methods of measuring airflow are included, the Pitot tube and the long radius flow nozzle. These are incorporated into a number of different setups or figures. In general, a fan is tested on the setup which most closely simulates the way in which it will be installed in an HVAC system. Centrif− ugal, tubeaxial and vaneaxial fans are usually tested with only an outlet duct. Figure 5−29 is a reproduction of a test setup from AMCA Standard 210. Note that this particular setup includes a long straight duct connected to the outlet of the fan. A straightener is located upstream of the Pitot

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.17


PL.1

PL.2

PL.3 L 2,3 10 D3 MIN. +0.25

8.5 D 3

-0.00

D3 MIN. D3 5 D3

+0.25 -0.00

D3 t d3

D3

FAN

PITOT TUBE STRAIGHTENER

TRAVERSE

TRANSFORMATION PIECE

THROTTLING DEVICE

FIGURE 5-28 AMCA FAN TEST - PITOT TUBE tube traverse to remove swirl and rotational compo− nents from the airflow and to ensure that the flow at the plane of measurement is as near to uniform as possible.

A manufacturer may test a fan with or without an outlet duct or inlet duct. Catalog ratings should state whether ducts were used during the rating tests. If the fans are not to be applied with similar duct configurations as used in the test setup, an allowance should be made for the difference in the resulting performance.

Static Pressure

SYSTEM @ d 1

SP2 @ d AND RPM 2

5.5.2

SP2 SP @ d 1 AND RPM 1

Chg.

Airflow1

System Effect

For years, many HVAC system designers, system in− stallers, fan company sales engineers and testing, ad− justing, and balancing (TAB) contractors have found that system total pressure measurements and airflow capacities were considerably less than the fan horse− power and rpm curves indicated.

SP1

Airflow2

FIGURE 5-29 EFFECT OF DENSITY CHANGE (CONSTANT MASS FLOW)

The angle of the transition between the test duct and the fan outlet is limited to ensure that uniform flow will be maintained. A steep transition, or abrupt change of cross−section would cause turbulence and eddies, and lead to non−uniform flow. 5.18

Uniform flow conditions ensure consistency and re− producibility of test results and permits the fan to de− velop its maximum performance. In any installation where uniform flow conditions do not exist, the fan’s performance will be reduced.

This derating of duct system fans is called system ef− fect, and it is very important that this phenomenon be taken into account by all concerned with HVAC sys− tems if they are to operate as designed. System effect diminishes a fan’s performance because of the interaction of the fan and the connected duct sys− tem; and system effect factors are used to compensate for the fan’s decreased performance. In general, sys− tem effect factors are approximations obtained from many research studies. Some studies have been pub− lished previously by individual fan manufacturers, and

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


many represent the consensus of engineers with con− siderable experience in fan applications. Fans of different types (and even fans of the same type from different manufacturers) will not necessarily react with a duct system in exactly the same way. Therefore, it is necessary to use judgment, based on ac− tual experience, in applying the system effect factors. 5.5.2.1

Fan Selection

Figure 5−30 illustrates deficient fan/system perfor− mance caused by system effect. HVAC system pres− sure losses have been determined and the fan selected to operate at Point 1 (system curve A). However, no al− lowance has been made for the effect of poor duct con− nections to the fan. To compensate, a system effect fac− tor must be added to the calculated system pressure losses to determine a new system curve that is then used to select the fan. The point of intersection between the fan performance curve and this new ?phantom" system curve B is Point 4. Therefore, the actual system flow volume is defi− cient by the difference from Point 1 to Point 4. To achieve the design airflow volume, a system effect factor equal to the pressure difference between Point 1 and Point 2 should be added to the calculated system pressure losses. The fan should be selected to operate at Point 2 where the new corrected rpm curve crosses phantom system curve B. A higher fan brake horse− power will also be required.

corrected fan, the airflow volume and static pressure will be established as point 1, because that is where the system actually is operating. The system is not operat− ing on the phantom system curve, which was used only to select the derated capacity fan. System effect cannot be measured in the field, but only calculated after a visual inspection is made of the fan/duct system con− nections. Because system effect is velocity related, the differ− ence between Points 1 and 2 is greater than the differ− ence between Points 3 and 4. The system effect factor includes only the effect of the system configuration on the fan’s performance. All duct fitting pressure losses are calculated as part of the HVAC system pressure losses and are part of system curve A. 5.5.2.2

Figure 5−31 shows the changes in velocity profiles from the fan outlet to where a uniform velocity profile has developed in the duct. The distance of this point from the fan is called the effective duct length. To ob− tain 100 percent of the energy recovery or static regain, duct fittings or abrupt changes in duct configuration should not be used within that space. In other words, any changes to the discharge duct con− figuration within the effective duct length (which dif− fers from the duct configuration used when the fan was tested and rated) may cause the fan to perform less effi− ciently. 5.5.2.3

However, when a testing and balancing technician measures the actual HVAC system conditions with the

PHANTOM CURVE B WITH SYSTEM EFFECT CURVE A CALCULATED DUCT SYSTEM WITH NO ALLOWANCE FOR SYSTEM EFFECT 2

SYSTEM EFFECT LOSS AT DESIGN VOLUME

4

DESIGN PRESSURE

1 3

SYSTEM EFFECT AT ACTUAL FLOW VOLUME

SELECTED FAN CURVE FAN CATALOG PRESSURE-VOLUME CURVE

DEFICIENT PERFORMANCE DESIGN VOLUME

FIGURE 5-30 EFFECTS OF SYSTEM EFFECT

Fan Outlets

Fan Inlets

Power roof exhausters are tested and rated when mounted on a roof curb through which the exhaust air duct passes, so system effect is not a problem. The problem occurs with HVAC centrifugal and axial flow fans that are tested without any inlet obstructions or in− let duct connections. For rated performance, the air must enter the fan uni− formly over the inlet area in an axial direction without pre−rotation. Non−uniform flow into the inlet is the most common cause of reduced fan performance. Such inlet conditions are not equivalent to a simple in− crease in the system resistance, so they cannot be treated as a percentage decrease in the fan airflow and pressure output. A poor inlet condition results in an en− tirely new fan performance. Many system effect curves for various round and rec− tangular elbows may be found in AMCA Publication 201−90, Fans and Systems or the SMACNA HVAC

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.19


BLAST AREA CENTRIFUGAL CUTOFF FAN

OUTLET AREA

DISCHARGE DUCT

LENGTH OF DUCT R

100% EFFECTIVE DUCT LENGTH

a. Round Elbow

FIGURE 5-31 FAN OUTLET EFFECTIVE DUCT LENGTH

System9Duct Design manual. When a suitable (often sizeable) length of duct is used between the fan inlet and return air duct elbow, system effect may be avoi− ded. These improvements help maintain uniform flow into the fan inlet and thereby approach the flow condi− tions of the laboratory test setup. Most often where space is at a premium, the inlet duct will be mounted directly to the fan inlet, as shown in Figure 5−32 b. The reduction in capacity and pressure for this type of inlet condition is impossible to tabulate. The many possible variations in width and depth of the duct influence the reduction in performance to varying degrees. Therefore, this inlet should be avoided. Fans and Systems and the HVAC Systems9Duct De− sign manual states that capacity losses as high as 45 percent have been observed in poorly designed inlets, such as those shown in Figure 5−32 b. Field fabricated or factory designed inlet boxes (see Figure 5−32 c) may often eliminate or substantially reduce system effect at fan inlets. Inlet elbows at axial fans may cause an instability in fan operation in addition to system effect that could re− sult in serious damage to the fan. It is strongly advised that inlet elbows be installed at least three duct diame− ters away from any axial fan inlet.

5.20

b. Rectangular Duct

c. Inlet Box

FIGURE 5-32 NON-UNIFORM FLOW CONDITIONS INTO FAN INLET 5.5.2.4

Field Measurements

Recent research has determined that accurate duct ve− locity measurements cannot be made until a near uni− form velocity profile has developed. This point may vary from 3 to 20 duct diameters downstream from the object causing the turbulence. So, any accurate mea− surements on the discharge side of a fan must be well away from the point where system effect occurs. On the fan inlet, non−uniform airflow, spin in the air− flow or a duct condition that produces a vortex create the problem. When one observes the various poor duct fan inlet conditions normally installed, accurate mea− surements are impossible. However, on the fan inlet side, the system effect loss usually occurs within the entry to the fan wheel, so it is not field−measurable. Finally, system effect is a real and often occurring pro− blem. It may be avoided by using better fan/system duct connections where space permits. it also may be avoided if the installing contractor would order HVAC fans and equipment with the proper inlet and discharge connection configurations so that elbows would not have to immediately change airflow direction. Just re− member that: fan capacity reductions due to system ef− fect cannot be measured in the field by TAB techni− cians, system effect losses are approximate, and that system effect factors must not be confused with duct fitting loss coefficients.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


CHAPTER 6

AIR DISTRIBUTION AND DEVICES


CHAPTER 6

AIR DISTRIBUTION AND DEVICES

6.1

AIR TERMINAL BOXES

6.1.2.1

6.1.1

Introduction

Constant flow rate controllers may be of the pneumatic or electric volume regulator type. They typically re− quire internal differential pressure sensing, selector devices, and pneumatic or electric motors for opera− tion.

An air terminal box or terminal unit is a device that controls the volume of conditioned air introduced into a space or zone from the HVAC air duct system. The air terminal box manually or automatically fulfills one or more of the following functions. 6.1.1.1

Pressure

The air terminal box may control the pressure of the discharge airflow. 6.1.1.2

Airflow Rate

The air terminal box may control the rate and velocity of the discharge airflow. 6.1.1.3

Temperature

The air terminal box may mix airstreams of different temperatures or humidities, or include a coil to add additional heating or cooling capacity.

6.1.2.2

Air Blending

Variable Air Volume

Variable air volume (VAV) controllers incorporate a means to reset the constant volume regulation auto− matically to a different control point within the range of the control device in response to an outside signal, such as from a thermostat. Boxes with this feature are pressure independent and may be used with reheat components. Variable flow rate may also be obtained by using a modulating damper ahead of a constant vol− ume regulator. This arrangement typically allows for variations in flow between high and low limits or be− tween a high limit and shutoff. These boxes are pres− sure dependent and volume limiting in function. Pneu− matic variable volume may be either pressure independent, volume limiting, or pressure dependent, according to the equipment selected. 6.1.3

6.1.1.4

Constant Airflow

Box Power Sources

A terminal box commonly integrates a sound chamber to reduce noise generated by the manual damper or flow controller reducing the inlet air velocity or pres− sure. The sound attenuation chamber is typically lined with thermal and sound insulating material and is equipped with baffles. Special sound attenuation in the air discharge ducts usually is not required in smaller boxes.

Terminal boxes can be further categorized as being system powered, wherein the assembly derives all of the energy necessary for its operation from the supply air within the distribution system, or as externally powered, wherein the assembly derives part or all of the energy necessary for its operation from a pneumat− ic or electric outside source. In addition, assemblies are self−contained (when they are furnished with all necessary controls for their operation, including actua− tors, regulators, motors, and thermostats), as opposed to non−self−contained assemblies (where part or all of the necessary controls for operation are furnished by someone other than the terminal box manufacturer). In this latter case, the controls may be mounted on the as− sembly by the assembly manufacturer or may be mounted by others after delivery of the equipment.

6.1.2

6.1.4

Types of Air Terminal Boxes

6.1.4.1

Reheat Terminal Boxes

The air terminal box may mix primary air at high ve− locity and/or secondary air from the conditioned space. 6.1.1.5

Sound Attenuation

Categories

Terminal boxes are typically categorized according to the function of their airflow volume controllers, which are generally either constant or variable air volume (VAV) devices. They are further categorized as being pressure dependent, where the airflow rate through the assembly varies in response to changes in system pres− sure, or as pressure independent where the airflow rate through the device does not vary in response to changes in system pressure.

Reheat terminal boxes add sensible heat to the supply air. They cannot be used in many applications unless the added heat is recovered heat due to energy con− servation codes. Water or steam coils or electric resist− ance heaters can be located within or attached directly to the air discharge end of the box. These boxes typi− cally are single duct, and operation can be either

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.1


constant or variable volume. However, if they are VAV, they must maintain some minimum airflow to accomplish the reheat function. 6.1.4.2

Dual Duct Terminal Boxes

Duct terminal boxes receive warm and cold air from separate air supply ducts in accordance with space re− quirements. Pneumatic and electric volume regulated boxes often have individual modulating dampers and operators to regulate the amount of warm and cool air. When a single modulating damper operator regulates the amount of both warm and cold air, a separate pres− sure reducing damper or volume controller (either pneumatic or mechanical) is needed in the box to re− duce pressure and limit airflow. Specially designed baffles may be required within the unit or at the box discharge to mix varying amounts of warm and cold air and/or to provide uniform flow downstream. Dual duct boxes can be equipped with constant flow rate or vari− able flow rate devices to be either pressure indepen− dent or pressure dependent to provide a number of vol− ume and temperature control functions. 6.1.4.3

Ceiling Induction Boxes

The ceiling induction box provides either primary air or a mixture of primary air and relatively warm air to the conditioned space. It accomplishes this function by permitting the primary air to induce air from the ceil− ing plenum or via inducted return air from conditioned space. A single duct supplies primary air at a tempera− ture cool enough to satisfy all zone cooling loads. The ceiling return air inducted into the primary air is at a higher temperature than the room because heat from recessed lighting fixtures enters the plenum directly. The induction box contains damper assemblies con− trolled by an actuator in response to a thermostat to control the amount of cool primary air and warm in− duced air. As reduction in cooling is required, the pri− mary air flow rate is gradually reduced and the induced air rate is generally increased. Reheat coils can be used in the primary air supply and/or in the induction open− ing to meet occasional interior and perimeter load re− quirements. 6.1.4.4

Fan Powered Boxes

Fan powered boxes differ from the above induction boxes in that they are equipped with a blower. This blower, generally driven by a fractional horsepower motor, draws air from the conditioned space (secon− dary air) to be mixed with the cool primary air. The ad− vantage of fan assisted boxes over basic VAV boxes is 6.2

that for a small energy expenditure to the terminal fan, constant air circulation can be maintained in the space. Fan assisted boxes operate at a lower primary air static pressure than air induction boxes, and perimeter zones can be heated without operating the primary fan during unoccupied periods. Warm air from the ceiling return can be used for low to medium heating loads depend− ing on construction of the building envelope. As the load increases, heating coils in the perimeter boxes can be activated to heat the recirculated plenum air to the necessary level. Fan assisted boxes can be divided into two categories: constant volume and bypass−type units. Constant volume, fan assisted boxes (Figure 6−1) are used when constant air circulation is desired in the space. The unit has two inletsCone for cool primary air from the central fan system and one for the secon− dary air. All air delivered to the space passes through the blower. The blower operates continuously when− ever the primary air fan is on and can be cycled to de− liver heat, as required, when the primary fan is off. As the cooling load decreases, a damper throttles the amount of primary air delivered to the blower. The blower makes up for this reduction of primary air by drawing air in the space or ceiling plenum through the return air opening.

DISCHARGE AIR FAN

(CONSTANT VOLUME)

PRIMARY AIR

DAMPER

RETURN AIR (PLENUM)

FIGURE 6-1 CONSTANT VOLUME FAN-POWERED BOX

In the bypass−type fan assisted box (Figure 6−2), the cool primary air bypasses the blower portion of the unit and is delivered directly to the space. The blower section draws in plenum air only and is mounted in par− allel with the primary air damper. A back draft damper prevents primary air from flowing in to the blower sec− tion when the blower is not energized. The blower in these units generally is energized after the damper in the primary air is partially or completely throttled. Some electronically controlled units gradually in− crease the fan speed as the primary air damper is throttled to maintain constant airflow, while permit− ting the fan to shut off when it is in the full cooling mode.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


6.2

VARIABLE AIR VOLUME (VAV) TERMINAL BOXES

6.2.1

Introduction

A VAV system controls the dry bulb temperature with− in a space by varying the volume of supply air rather than the supply air temperature. VAV systems can be applied to interior or perimeter zones, with common or separate fan systems, common or separate air tempera− ture control, and with or without auxiliary heating de− vices. As the VAV boxes on a given supply duct system begin to reduce their air flow, duct pressure controls sense this duct pressure increase and the supply fan air flow is reduced accordingly.

PRIMARY AIR

DAMPER DISCHARGE AIR

6.2.3

Combination Pressure Dependent-Independent Boxes

These combination boxes regulate maximum volume, but the airflow volume below the maximum flow rate varies with the inlet pressure variation. Generally, air− flow will oscillate when pressure varies. They are less expensive than pressure independent units and can be used where pressure independence is required only at maximum volume, where system pressure variations are relatively minor, and where some degree of hunt− ing is tolerable. 6.2.4

Pressure Dependent Boxes

Pressure dependent boxes do not regulate the airflow volume, but position the volume regulating device in response to the thermostat. They are the least expen− sive and should only be used where there is no need for limit control and the system pressure is stable. 6.2.5

Bypass (Dumping) Boxes

(VARIABLE VOLUME) RETURN AIR (PLENUM)

FAN

FIGURE 6-2 BYPASS-TYPE FANPOWERED BOX

A space thermostat can control flow by varying a damper, or a volume regulating device in the duct, or a pressure reducing terminal box. Depending on the complexity of the air distribution system, and consid− erations, VAV may or may not be combined with fan or system static pressure controls. The fan system is designed to handle the largest simul− taneous block load, not the sum of the individual peaks. As each zone peaks at a different time of day, it borrows the extra air from off−peak zones. This transfer of air from low−load to high−load zones occurs only in a true VAV system. 6.2.2

Pressure Independent Boxes

Pressure independent boxes regulate the airflow vol− ume in response to the thermostat’s call for heating or cooling. The required airflow is maintained regardless of fluctuation of the VAV unit inlet or system pressure. These units can be field or factory adjusted for maxi− mum and minimum cfm (L/s) settings. They will oper− ate at inlet static pressures as low as 0.2 in. of water (50 Pa) at maximum system design volumes.

VAV room supply is accomplished in constant volume systems by returning excess supply air into the return ceiling plenum or return air duct, thus bypassing the room. However, this reduction of system volume is not energy efficient. Use generally is restricted to small systems where a simple method of temperature control is desired, initial cost is modest, and energy conserva− tion is deemed unimportant. 6.3

OTHER AIRFLOW DEVICES

6.3.1

Pressure Reducing Valves

Pressure reducing valves or air valves each consist of a series of gang operated vane sections mounted within a rigid casing and gasketed to reduce as much air leak− age as possible between the valve and duct. They usu− ally are installed between a high pressure trunk duct and a lower pressure branch duct. Pressure is reduced by partially closing the valve, which results in a high pressure drop through the valve. This action generates noise, which must be attenuated in the low pressure discharge duct. The length and type of duct lining depend on the amount and frequency of noise to be attenuated. Volume control is obtained by adjustment of the valve manually, mechanically, or automatically. Automatic adjustment is achieved by a pneumatic or electric con− trol motor actuated by a pressure regulator or a thermo− stat. Pressure reducing valves are generally equal in size to the low pressure branch duct connected to the valve

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.3


discharge. This arrangement provides minimum pressure drop with valves opened fully. 6.3.2

thermostatically controlled. The opening varies in approximate proportion to the air volume to maintain discharge throw pattern stability, even with low air quantities. Since these units are pressure dependent, constant pressure regulators are usually required in the duct system. Noise is a particular concern when selecting outlets.

Supply Outlet Throttling Units

The area of the throat or the discharge opening of these supply outlets, which are usually linear diffusers, is

FRAME: 2’ or 1 1_w" ¢ 1 1_w" ¢ 1_i" STRUCTURAL OR FORMED CHANNEL

FRAME 3_i" OR 1_w" DIA.

SHAFTS

18 GA MIN. BLADES

ANGLE STOP 1_w" ¢ 1_w" BAR OPTIONAL

SHAFT EXTENSION

SECTION

FIG. A OPPOSED ACTION

CHANNEL FRAME

PIN & BRONZE BUSHING

CONNECTING BAR FRAME

NOTICE 48" MAX. WIDTH FRAME

STOP SHAFT EXTENSION FIG. B PARALLEL ACTION

SEE TEXT ON VOLUME DAMPERS

SECTION FIG. C

FIGURE 6--3 MULTIBLADE VOLUME DAMPERS

6.4

HVAC SYSTEMS Testing, Adjusting & Balancing  Third Edition


Volume Dampers

6.3.3.1

Introduction

Volume dampers are primary elements in the duct sys− tem. They are used for controlling airflow rates by introducing a resistance to airflow in the system. In higher pressure systems, the damper is referred to as a pressure reducing valve. Volume control or balancing dampers should be installed in each branch of zone duct. Single leaf dampers which are part of a manufactured air grille are not acceptable for system balancing. Opposed blade dampers which are part of a manufactured air grille can be used if there is not enough room for a regular damper and if sufficient space is provided behind the grille face for proper operation of the damper. Other− wise, a balancing damper should be installed in the branch register termination at a location where it is ac− cessible from the grille or diffuser opening, or a quad− rant damper should be used. Volume dampers installed in branch ducts where the total estimated static pressure is less than 0.5 in. wg (125 Pa) can be single leaf type. Volume dampers installed in ductwork where the total estimated system static pressure exceeds 0.5 in. wg (125 Pa) should be manufactured in accordance with Figure 6−3. 6.3.3.2

Multiblade Dampers

Figure 6−3 shows two types of multiple blade dampers: parallel blade and opposed blade. The terms parallel and opposed refer to the movement of the adjacent bla− des. In the parallel blade damper, all of the blades move in parallel. The opposed blade damper has a linkage which causes the adjacent blades to move in opposite directions. Partial closing of a damper increases the resistance of the duct system to airflow. The reduction in airflow with closure of the damper may or may not be propor− tional to the amount of adjustment of the damper. That is, closing the damper half way does not necessarily mean that the air volume will be reduced to fifty per− cent of that volume which flows through the damper when it is wide open. The relation between the position of the damper and the percent of air that flows through the damper with respect to the airflow through the wide open damper is termed the flow characteristic. Typical flow characteristic curves for parallel blade and opposed blade dampers are shown in Figures 6−4 and 6−5. In Figure 6−4 the flow characteristic curves for

the parallel blade damper show that as the damper is closing, the flow reduction may be proportional to the closing of the damper as is shown by curve J, or partial closure of the damper may have little effect on the flow as is shown by curve A. 100 90 PERCENT OF MAXIMUM FLOW

6.3.3

80 A B

70

C D

60

E F G

50

H J

40

K

30 20 10 0 10

20

30 40 50 60 70 DAMPER POSITION, DEGREES OPEN

80

90

FIGURE 6-4 FLOW CHARACTERISTICS FOR A PARALLEL OPERATING DAMPER The manner in which the damper performs in any duct system is determined by how complicated the system is. If the system is very simple and the damper makes up a major part of the resistance, then any movement of the damper will change the resistance of the entire system and good control of the airflow will result. If the damper resistance is very small in relation to that of the entire system, a poor flow characteristic such as curve ?A" in Figures 6−4 and 6−5 will result. Typical ratios of damper to system resistance are shown in Table 6−1 for each flow characteristic curve. The set of curves for the opposed blade damper (Figure 6−5) shows that for a given ratio of damper to system resistance, a better flow characteristic usually results than with the parallel blade damper (Figure 6−4). As the opposed blade damper is closed, it introduces more resistance to airflow for a given position than a parallel blade damper. When balancing systems, it should be realized that the flow characteristics of a damper are not constant and will vary from one system to another. The actual effect of closing the damper can only be determined by mea− surements in the particular system unless the system designer has taken into account the damper flow char− acteristics in his system design. It is important that the TAB technician understand the airflow patterns of multiblade dampers. The parallel blade damper has a tendency to throw the air toward one side of the duct. This uneven pattern may adverse− ly affect coil or fan performance, or airflow into

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.5


branch ducts if the damper is located just upstream of any system component. 100 PERCENT OF MAXIMUM FLOW

90 80 70

Where dampers should have tight shutoff when closed, the linkage between blades must be properly adjusted. Damper motor linkage also must be properly adjusted. Cold deck and hot deck dampers, as used in multi−zone units, must close tightly, as must face and by−pass dampers used in some air handling units.

A

60

B C

50

D E F

40

G H

30 20 10 0 10

20

30 40 50 60 70 DAMPER POSITION, DEGREES OPEN

80

90

FIGURE 6-5 FLOW CHARACTERISTICS FOR AN OPPOSED OPERATING DAMPER

These flow patterns should be noted when it is neces− sary to measure airflow in a duct near a damper. Where possible, make any measurements upstream rather than downstream of a damper. 6.3.3.3

Parallel−leaf dampers

0.5− 1.0 1.0− 1.5 1.5− 2.5 2.5− 3.5 3.5− 5.5 5.5− 9.0 9.0−15.0 15.0−20.0

6.4

AIR DISTRIBUTION BASICS

6.4.1

Introduction

Air distribution criteria will vary considerably in com− mercial and institutional buildings as well as zone tem− perature and humidity levels. People sitting with little activity require closer tolerances than those actively moving about. Spillover from open refrigerated dis− play equipment in super markets causes frequent com− plaints from customers. An understanding of the prin− ciples of room air distribution helps in the selection, design, control, and operation of HVAC duct systems. The real evaluation of air distribution in a space, how− ever, is if most occupants are comfortable.

Quadrants and Linkages

When dampers are located within ducts and are manu− ally controlled, they are usually secured in place with locking linkage or quadrant such as those shown in Figure 6−6. Varying in strength and locking ability, they should be of suitable size for the damper with which they are used. When adjusting a damper, the regulator or quadrant must be tightened securely to en− sure that the damper remains as set.

Open damper resistance, percent of system resistance

Do not always accept the position of the regulator pointer as indicating the actual position of the damper blade. When in doubt, inspect the end of the damper rod at the face of the regulator. A groove, usually cut by a hacksaw, will indicate that the damper blade runs in the same direction as the cut.

Opposed−leaf dampers

Flow character− istic curve

Open damper resistance, percent of system resistance

Flow character− istic curve

A B C D E F G H

0.3− 0.5 0.5− 0.8 0.8− 1.5 1.5− 2.5 2.5− 5.5 5.5−13.5 13.5−25.5 25.5−37.5

A B C D E F G H

The object of good air distribution in HVAC systems is to create the proper combination of temperature, hu− midity, and air motion, in the occupied zone of the con− ditioned room from the floor to 6 feet (2 m) above floor level. To obtain comfort conditions within this zone, standard limits have been established as acceptable ef− fective draft temperature. This term includes air tem− perature, air motion, relative humidity, and the physio− logical effects on the human body. Any variation from accepted standards of one of these elements causes dis− comfort to occupants. Lack of uniform conditions within the space or excessive fluctuation of conditions in the same part of the space may produce less than ac− ceptable conditions. Although the percentage of room occupants who ob− ject to certain conditions may change over the years, more recent research has shown that a person tolerates higher air flow velocities and lower temperatures at ankle level than at neck level. Because of this, condi− tions in the zone extending from approximately 30 to 60 inches (0.75 to 1.5 m) above the floor are more criti− cal than conditions nearer the floor.

6.4.2 Table 6-1 Typical Ratios of Damper to System Resistance for Flow Characteristic Curve 6.6

Air Velocity And Air Entrainment

For comfortable air distribution, room air velocities within the occupied zone (floor to 6 feet [2 m] above

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


ROD. WASHER, LOCK NUT OR HINGE

MIN.-12 " FLOW

SET SCREW 1 ROD TO 24" DEPTH 2 RODS 25" TO 60" 3 RODS 61" & OVER

SPLITTER DAMPER

BEARING OPTION

ROD CONTINUOUS ON 2" WG CLASS AND ON ALL DAMPERS OVER 12" DIA.

NUT

ARM FIG D ELEVATION TWO BLADE ARRANGEMENT

FIG C ROUND DAMPER

DUCT

CIRCULAR DUCTS

3 8

DUCT

" QUADRANT

DUCT

QUADRANT 1" 2

" PIN

ROD-PIN

12" MAX

3 8

1" 2

22 Ga. BLADE. 1" 8

16 Ga. BLADE 1" 8

CLEARANCE ALL AROUND UP TO 18" FIG A

CLEARANCE ALL AROUND 19" TO 48" FIG B

NOTE: OVER 12" HIGH USE MULTIPLE BLADES. SEE FIG 14-3

D

DUCT DEPTH

STIFFEN AS REQUIRED

D

1" 2

FIG A OR B SIDE ELEVATION

RECTANGULAR DUCTS

FIGURE 6-6 VOLUME DAMPERS HVAC SYSTEMS Testing, Adjusting & Balancing â&#x20AC;¢ Third Edition

6.7


floor level) should be in a range of 20 fpm to 70 fpm (0.1 to 0.35 m/s) with 50 fpm (0.25 m/s) normally be− ing used. Stagnant air areas should be avoided as tem− perature in these areas may not be acceptable to the oc− cupants. Room air velocities less than 50 fpm (0.25 m/s) are ac− ceptable; however, even higher velocities may be ac− ceptable to some occupants. ASHRAE Standard 55−1981 recommends elevated air speeds at elevated air temperatures. No minimum air speeds are recom− mended for comfort, although air speeds below 20 fpm (0.1 m/s) are usually imperceptible. The velocity of the air (primary air) emerging from the supply outlet induces air movement within the room area (secondary air). This process of entrainment or capturing of secondary air into the primary air is an es− sential part of air distribution to create total air move− ment within the room, thereby eliminating stagnant air areas and reducing temperature differences to accept− able levels before the air enters the occupied zone. Air entrainment will also tend to overcome natural con− vection and radiation effects within the room. 6.4.3

Surface Effect

Air entrainment takes place only along one surface of the outlet discharge jet when the outlet discharges air directly parallel and adjacent to a wall or ceiling. The

surface effect (Coanda effect) is illustrated in Figure 6−7. Since turbulent jet airflow from a grille or diffuser is dynamically unstable, it may veer rapidly back and forth. When the jet airflow veers close to a parallel and adjacent wall or ceiling, the surface interrupts the flow path on that side as shown in Figure 6−7B. The result is that no more secondary air is flowing on that side to replace the air being entrained with the jet airflow. This causes a lowering of the pressure on that side of the outlet device, creating a low pressure bubble that causes the jet airflow to become stable and remain at− tached to the adjacent surface throughout the length of the throw. The surface effect counteracts the drop of horizontally projected cool airstreams. Ceiling diffusers exhibit surface effect to a high degree because a circular air pattern blankets the entire ceil− ing area surrounding each outlet. Slot diffusers, which discharge the airstream across the ceiling, exhibit sur− face effect only if they are long enough to blanket the ceiling area. Grilles exhibit varying degrees of surface effect, depending on the spread of the particular air pattern. In many installations, the outlets must be mounted on an exposed duct and discharge the airstream into free space. In this type of installation, the airstream en− trains air on both its upper and lower surfaces; as a re− sult, a higher rate of entrainment is obtained and the throw is shortened by approximately 33 percent. Air− flow per unit area for these types of outlets can, there−

SEPARATION BUBBLE CEILING

CEILING JET FLOW

WALL

JET FLOW

ENTRAINED AIRFLOW (SECONDARY AIR)

FLOOR (A)

ENTRAINED AIRFLOW (SECONDARY AIR)

FLOOR (B)

FIGURE 6-7 SURFACE (COANDA) EFFECT

6.8

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


fore, be increased. Because there is no surface effect from ceiling diffusers installed on the bottom of ex− posed ducts, the air drops rapidly to the floor. There− fore, temperature differentials in air conditioning sys− tems must be restricted to a range of 15F to 20F (8C to 1C). Airstreams from slot diffusers and grilles show a marked tendency to drop because of the lack of surface effect. 6.4.4

Smudging

Smudging may be a problem with ceiling and slot dif− fusers. Dirt particles held in suspension in the secon− dary (room) air are subjected to turbulence at the outlet face. This turbulence, along with surface effect, is pri− marily responsible for smudging. Smudging can be ex− pected in areas of high pedestrian traffic (lobbies, stores, etc.) When ceiling diffusers are installed on smooth ceilings (such as plaster or metal pan), smudg− ing is usually in the form of a narrow band of discolor− ation around the diffuser. Anti−smudge rings may re− duce this type of smudging. On highly textured ceiling surfaces (such as rough plaster and sprayed−on com− position), smudging often occurs over a more exten− sive area. 6.4.5

Sound Levels

The sound level of an outlet is a function of the air dis− charge velocity and the transmission of HVAC equip− ment noise, which is a function of the size of the outlet. Higher frequency sounds can be the result of excessive outlet velocity, but may also be generated in the duct by the moving airstream. Lower pitched sounds are generally the result of mechanical equipment noise transmitted through the duct system and outlet. The cause of higher frequency sounds can be pin− pointed as outlet or equipment sounds by removing the outlet during operation. A reduction in sound level in− dicates that the outlet is causing noise. If the sound lev− el remains essentially unchanged, the system is at fault. Chapter 46 Sound and Vibration Control in the 1999 ASHRAE HVAC Applications Handbook has more information on design criteria, acoustic treat− ment, and selection procedures. 6.4.6

Effect Of Blades

Blades affect grille performance if their depth is at least equal to the distance between the blades. If the blade ratio is less than one, effective control of the air− stream discharged from the grille by means of the blades is impossible. Increasing the blade ratio above

two has little or no effect, so blade ratios should be be− tween one and two. A grille discharging air uniformly forward (blades in straight position) has a spread of 14 to 24, depend− ing on the type of outlet, duct approach, and discharge velocity. Turning the blades influences the direction and throw of the discharged airstream. A grille with diverging blades (vertical blades with uniformly increasing angular deflection from the cent− erline to a maximum at each end of 45) has a spread of about 60, and reduces the throw considerably. With increasing divergence, the quantity of air dis− charged by a grille for a given upstream total pressure decreases. A grille with converging blades (vertical blades with uniformly decreasing angular deflection from the centerline) has a slightly higher throw than a grille with straight blades, but the spread is approximately the same for both settings. The airstream converges slightly for a short distance in front of the outlet and then spreads more rapidly than air discharged from a grille with straight blades. In addition to vertical blades that normally spread the air horizontally, horizontal blades may spread the air vertically. However, spreading the air vertically risks hitting beams or other obstructions or blowing primary air at excessive velocities into the occupied zone. On the other hand, vertical deflection may increase adher− ence to the ceiling and reduce the drop. In spaces with exposed beams, the outlets should be lo− cated below the bottom of the lowest beam level, pre− ferably low enough to employ an upward or arched air path. The air path should be arched sufficiently to miss the beams and prevent the primary or induced air− stream from striking furniture and obstacles and pro− ducing objectionable drafts. 6.5

ROOM AIR DISTRIBUTION

6.5.1

Natural Airflow

The natural air convection currents flowing down the glass during heating, and up the glass during cooling as shown in Figure 6−8, are a major influence on the air distribution in the perimeter zones of a building. During heating, these currents carry cool air down to the floor level causing a stratification of air in layers of increasing temperatures from the floor to the cei− ling. The severity of the temperature gradient depends

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.9


0

75F

95F

75F 80F

HEATING

COOLING

FIGURE 6-8 SOME ELEMENTS AFFECTING BODY HEAT LOSS on outdoor temperature, construction, and air distribu− tion. It is easily understood that warm supply air introduced at the base of the wall would tend to coun− teract these currents and reduce or eliminate stratifica− tion. Optimum air distribution in perimeter zones re− quires perimeter introduction of air or supplementary radiation at the perimeter. During cooling, currents carry warm air up the wall to ceiling level. Stratification then forms from the ceiling down. To eliminate stratification, cool air should be projected into this region near the ceiling. To do this most effectively, supply air outlets should be located high in the wall or in the ceiling. 6.5.2

Supply Air Outlet Performance

6.5.2.1

Outlet Throw

Extensive studies of supply outlet performance have shown that air discharge throw from free round open− ings, grilles, perforated panels and ceiling diffusers are related to the average velocity at the face of the supply outlet or opening. An air jet discharged from a free opening has four zones of expansion and the centerline velocity of the jet in any zone is related to the initial velocity as shown in Figure 6−9. Regardless of the type of outlet, the air stream will tend to assume a circular shape in free space. The important point is that the performance of any supply outlet is related to the initial velocity and initial area as shown in Figure 6−9.

the initial volume of the jet at any distance from the point of origin depends mainly on the ratio of the initial velocity (Vo) to the terminal velocity (Vx). For exam− ple, doubling the initial velocity for the same terminal velocity doubles the induction ratio and also the throw. In zone 4 where the terminal velocity is relatively low and specifically for terminal velocities of 50 fpm (0.25 m/s), the throw should be reduced 20 percent. The buoyant forces with non−isothermal jets cause the air jet to rise during heating and drop during cooling. These conditions result in shorter throws when the throw is reduced to a terminal velocity less than 150 fpm (0.75 m/s) The discussion of throw and drop has been limited to free space applications. If the air discharge jet is pro− jected parallel to and within a few inches of a surface, the jet performance will be affected by the surface, which limits the induction on the surface side of the jet. This creates a low pressure region between the jet and the surface which draws the jet toward the surface. In fact, this effect will prevail if the angle of discharge be− tween the jet and the surface is less than 40. Surface effect will draw the jet from a sidewall outlet to the ceiling if the outlet is within one foot (0.3 m) of the ceiling. The jet from a floor outlet will be drawn to the wall and the jet from a ceiling outlet will be drawn to the ceiling. Surface effect increases the throw for all types of outlets and decreases the drop for horizontally projected air streams. 6.5.2.2

Beyond the second zone, the jet is a mixture of supply and room air. The air jet expands because of induction of room air. The ratio of the total volume of the jet, to 6.10

VAV Applications

Air distribution is very important in VAV applications. Consideration must be given to distribution and to

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


1

3

2

4

X

PRIMARY

HIGH VELOCITY USUALLY ON OR NEAR ROOM SURFACES

PRIMARY AND INDUCED ROOM AIR

AIR

HIGH TEMP-HEATING LOW TEMP-COOLING

ROOM AIR GENTLE MOVEMENT

GREATEST POSSIBLE SOURCE OF DRAFTS FREE JET—ROOM AIR INDUCED ON ALL SIDES JET NEAR A SURFACE—HIGH VELOCITY AIR HUGS SURFACE AND INDUCES AIR ON ONLY ROOM SIDE OF JET

ZONE

SUPPLY VELOCITY,VO JET CENTERLINE VELOCITY,VX

1

V

2

V

3

V

4

V

X

X

=

V

X

»

V

O

X

»

V

O

O

/

X

/

X

APPROACHES ROOM VELOCITY

FIGURE 6-9 FOUR ZONES IN JET EXPANSION sound levels at maximum and minimum airflow. If the combined sound level of the terminal unit and diffuser at maximum flow is at least 3 dB below the room ambi− ent sound level, variations will not be noticed. In gen− eral, several important considerations are listed for variable volume system air distribution using outlets for horizontal discharge patterns as follows: a.

b.

An outlet with a low throw coefficient should be used. A small throw coefficient gives a smaller absolute change in the throw values with variation in volume and thus tends to minimize the change in air motion within the occupied space due to change in airstream pattern. Outlets should be chosen for small quantities of air. In this manner, absolute values of throw will vary a minimum with the variation in flow rate for the outlet. If the system ap− plication requires modular outlet arrange− ments for occupancy flexibility, as with dif− fusers in combination with light troffers or with ceiling suspensions, no increase in the

number of outlets is necessary to satisfy this requirement. For under window air distribution, vertical throw out− lets with nonspreading pattern should be used. To pre− vent cool air dropping back into the occupied space at minimum flow conditions, the outlet discharge veloc− ity should be 500 fpm (2.5 m/s) minimum. The throw coefficient should be higher to project air up to the cei− ling. With these exceptions, the preceding items also apply to under window distribution. 6.5.3

Supply Outlets

Outlets are selected for each specific room, based on air quantity required, distance available for throw or radius of diffusion, structural characteristics and ar− chitectural concepts being considered. Table 6−2 is based on experience and typical ratings of various out− lets. It may be used as a guide to the outlets applicable for use with various room air loadings. Special condi− tions, such as ceiling heights greater than the normal 8 to 12 feet (2.4 to 3.5 m) and exposed duct mounting,

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.11


FIGURE 6-10 TYPICAL SUPPLY OUTLETS

as well as product modifications and unusual condi− tions of room occupancy, can modify this table.

6.5.3.1

The adjustable blade grille is the most common type of grille used as a supply outlet. The single deflection blades install behind and at right angles to the face bla− des. This grille controls the airstream in both the hori− zontal and vertical planes.

Grille Slot Perforated Panel Ceiling Diffuser Perforated Ceiling

Air Loading, cfm/ft2 (L/s per m2) of Floor Space

Approx. Max. Air Changes @Hour for 10’ (3 m) Ceiling

0.6 to 1.2 (3 to 6) 0.8 to 2.0 (4 to 10) 0.9 to 3.0 (5 to 15) 0.9 to 5.0 (5 to 25) 1.0 to 10.0 (5 to 50)

7 12 18 30 60

Table 6-2 Guide to Use of Various Outlets 6.12

Fixed Blade Grilles

The fixed blade grille is similar to the single deflection grille, except that the blades are not adjustable; the blades may be straight or set at an angle. The angle at which the air is discharged from this grille depends on the type of deflection blades.

Adjustable Blade Grilles

Type of Outlet

6.5.3.2

6.5.3.3

Stamped Grilles

The stamped grille is stamped from a single sheet of metal to form a pattern of small openings through which air can pass. Various designs are used, varying from square or rectangular holes to intricate ornamen− tal designs. 6.5.3.4

Variable Area Grilles

The variable area grille is similar to the adjustable double deflection grille but can vary the discharge area to achieve an air volume change (variable air volume outlet) at constant pressure, so that the variation in throw is minimized for a given change in supply air volume.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


6.5.3.5

Slot Diffusers

6.5.3.9

Perforated Face Ceiling Diffusers

A slot diffuser is an elongated outlet consisting of a single or multiple number of slots. It is usually installed in long continuous lengths. Outlets with di− mensional aspect ratios of 25 to 1 or greater and a max− imum height of approximately 3 inches (75 mm) gen− erally meet the performance criteria for slot diffusers.

Perforated metal diffusers meet architectural demands for air outlets that blend into perforated ceilings. Each has a perforated metal face with an open area of 10 to 50 percent which determines its capacity. Units are usually equipped with a deflection device to attain multi−pattern horizontal air discharge.

6.5.3.6

6.5.3.10 Variable Area Ceiling Diffusers

Air-Light Diffusers

Air−light slot diffusers have a single slot discharge in nominal 2, 3, and 4 foot (0.6, 0.9, and 1.2 m) lengths and are available for use in conjunction with recessed fluorescent light troffers. A diffuser mates with a light fixture and is entirely concealed from the room. It dis− charges air through suitable openings in the fixture and is available with fixed or adjustable air discharge pat− terns, air distribution plenum, inlet dampers for bal− ancing, and inlet collars suitable for flexible duct con− nections. Light fixtures adapted for slot diffusers are available in styles to fit common ceiling constructions. Various slot diffuser and light fixture manufacturers may furnish products compatible with one another’s equipment. 6.5.3.7

Multi-Passage Ceiling Diffusers

Multi−passage ceiling diffusers consist of a series of flaring rings or louvers, which form a series of concen− tric air passages. They may be round, square, or rectan− gular. For easy installation, these diffusers are usually made in two parts; an outer shell with duct collar and an easily removable inner assembly. 6.5.3.8

Flush And Stepped-Down Diffusers

Flush and stepped−down diffusers also are available. In the flush unit, all rings or louvers project a plane sur− face, whereas in the stepped−down unit, they project beyond the surface of the outer shell. Common variations of this diffuser type are the adjust− able pattern diffuser and the multi−pattern diffuser. In the adjustable pattern diffuser, the air discharge pat− tern may be changed from a horizontal to a vertical or downblow pattern. Special construction of the diffuser or separate deflection devices allow adjustment. Mul− tipattern diffusers are square or rectangular and have special louvers to discharge the air in one or more di− rections. Other outlets available as standard equipment are half round diffusers, supply and return diffusers, and light fixture air diffuser combinations.

Variable area ceiling diffusers may be round, square, or linear and have parallel or concentric passages or a perforated face. In addition, they feature a means of ef− fectively varying the discharge area to achieve an air volume change (VAV outlet) at constant pressure, so that the variation in throw is minimized for a given change in supply air volume. 6.5.4

Under Floor Distribution

Raised floor plenums for air distribution are primarily used in computer rooms or research facilities having high concentrations of heat generating electronic equipment requiring very clean and cool supply air. Distributing conditioned air under a raised floor al− lows supply diffusers to be located directly under elec− tronic equipment cabinets having high sensible heat generation. To meet these high equipment cooling needs with ceiling supply diffusers would require dis− charge air flows that would be very uncomfortable for most room occupants, and may still not provide ade− quate cooling air flow into these equipment cabinets. In most applications, these spaces are served by a dedi− cated packaged air conditioning unit also located in the space, which discharges supply air directly down and into the under floor cable access plenum. Since the primary purpose of this system is to provide adequate cooling of very expensive electronic equip− ment, it is important for the TAB technician to obtain direction from those responsible for this equipment before balancing these floor registers. In some cases, equipment nameplate data may indicate a design air flow temperature and flow rate, if this information is not provided by the design drawings. It should also be understood that most raised floors are constructed of manufactured removable metal panel sections to allow easy access to the high concentration of wiring and easy addition or removal of equipment. For this reason, the air distribution must remain flex− ible and easy to modify.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.13


6.5.5

Return Air Inlet Performance

6.5.5.1

Location

Return air and exhaust air inlets should be located to suit architectural design requirements including ap− pearance and compatibility with supply outlets and ductwork. Generally, inlets are installed to return room air of the greatest temperature differential that collects in the stagnant areas. The location of return and ex− haust inlets does not significantly affect air motion. The location of return and exhaust inlets will not com− pensate for ineffective supply air distribution. A return air inlet should not be located directly in the primary airstream from supply outlets. To do so will short circuit the supply air back into the return air sys− tem without allowing it to mix with the room air. The TAB technician should remember that the supply air maintains the conditions within a space by mixing and dilution. Any removal of excess warm or cold air which is allowed to stratify before mixing within the space will permit lower temperature differentials or lower airflow rates. Removal of excess warm or cold air is accomplished with hoods in certain industrial processes. It can also be done by selecting the supply outlet performance to promote the formation of the stagnant zone directly from the local heat gain or loss. The return intake or exhaust would then be located in the stagnant zone. 6.5.5.2

Noise

In addition to the location of the return intake as dis− cussed above, the intake should be sized to return the proper amount of air to the HVAC unit with minimum static pressure requirements and noise levels. In gener− al, most commercial return grilles have a free area of between 45 and 55 percent, because they are designed so that one cannot see through them. With this type of grille, the velocity should not exceed approximately 500 fpm (2.5 m/s) to have reasonable pressure drop re− quirements and a reasonable sound level (see Table 6−3). In general, return air inlets should be sized on available pressure requirements and sound data, rather than relying on indicated free area values. The problem of return inlet noise is the same as that for supply outlets. In computing resultant room noise lev− els from operation of an air conditioning system, the return inlet must be included as a part of the total grille area. The major difference between supply outlets and return inlets is the frequent installation of the later at ear level. When they are so located, the return inlet ve− 6.14

locity should not exceed 75 percent of maximum per− missible outlet velocity.

Inlet Location Above occupied zone Within occupied zone Not near seats Within occupied zone Near seats Door or wall louvers Undercut doors

Velocity Over Gross Inlet Area−fpm (m/s) 800 Up (4.0 Up) 600−800 (3.0−4.0)

400−600 (2.0−3.0) 200−300 (1.0−1.5) 200−300 (1.0−1.5)

Table 6-3 Recommended Return Air Inlet Face Velocities 6.5.6

Return Air Inlets

6.5.6.1

Adjustable Blade Grilles

The same adjustable blade grilles used for air supply are used to match the deflection setting of the blades with that of the supply outlets. 6.5.6.2

Fixed Blade Grilles

The same fixed blade grilles described in the supply air section are used. This grille is the most common return air inlet. Blades are straight or set at a certain angle, the latter being preferred when appearance is important. 6.5.6.3

V-Blade Grille

The V−blade grille is made with blades in the shape of inverted v’s stacked within the grille frame, this grille has the advantage of being sight proof; it can be viewed from any angle without detracting from appea− rance. Door grilles are usually v−blade grilles. The air− flow capacity of the grille decreases as visibility through the grille decreases. 6.5.6.4

Light Proof Grille

The light proof grille is used to transfer air to or from darkrooms. The blades of this type of grille form a lab− yrinth and are painted black. The blades may take the form of several sets of v−blades or be of some special interlocking louver design to provide the required lab− yrinth. 6.5.6.5

Stamped Grilles

Stamped grilles frequently are used as return air and exhaust air inlets, particularly in more demanding areas like rest rooms and utility areas. 6.5.6.6

Ceiling And Slot Diffusers

Supply air ceiling diffusers also may be used as return air and exhaust air inlets.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Neck VelocityCfpm (m/s)

400 (2.0)

500 (2.5)

600 (3.0)

700 (3.5)

800 (4.0)

1000 (5.0)

Round Diffuser Half Round Diffuser Half Round Diffuser, Flush Square Diffuser Square Diffuser, Adjustable Rectangular Diffuser Curved Blade Diffuser Perforated Diffuser High Capacity Diffuser Slimline Diffuser, 2 Way* Extruded Fineline Diffuser 0.25 in.(6.4 mm) Bar Spacing* Linear Slot Diffuser*

0.024 (6.0) 0.035 (8.7) 0.046 (11.5) 0.021 (5.2) 0.036 (9.0) 0.043 (10.7) 0.056 (13.9) 0.037 (9.2) — 0.010 (2.5)

0.039 (9.7) 0.054 (13.4) 0.074 (18.4) 0.033 (8.2) 0.057 (14.2) 0.066 (16.4) 0.090 (22.4) 0.058 (14.4) C 0.015 (3.7)

0.056 (13.9) 0.080 (19.9) 0.106 (26.4) 0.048 (12.0) 0.080 (19.9) 0.096 (23.9) 0.131 (32.6) 0.083 (20.7) C 0.022 (5.5)

0.075 (18.7) 0.107 (26.6) 0.143 (35.6) 0.064 (15.9) 0.112 (27.9) 0.131 (32.6) 0.175 (43.6) C 0.050 (12.5) 0.028 (7.0)

0.096 (23.9) 0.141 (35.1) 0.184 (45.8) 0.083 (20.7) 0.144 (35.9) 0.170 (42.3) 0.225 (56.0) 0.148 (36.9) 0.060 (14.9) 0.040 (10.0)

0.152 (37.8) 0.219 (54.5) 0.290 (72.2) 0.130 (32.4) 0.226 (56.3) C 0.355 (88.4) 0.230 (57.3) 0.100 (24.9) 0.063 (15.7)

0.011 (2.7) 0.051 (12.7)

0.015 (3.7) 0.079 (19.7)

0.024 (6.0) 0.110 (27.4)

0.030 (7.5) 0.150 (37.4)

0.044 (11.0) 0.200 (49.8)

0.069 (17.2) C

Extractor

0.004 (1.0)

0.006 (1.5)

0.010 (2.5)

0.013 (3.2)

0.017 (4.2)

0.023 (5.7)

*Velocity Through Face Open Area

Table 6-4 Air Outlets and Diffusers Total Pressure Loss Average—in. wg (Pa)

VelocityCfpm (m/s)

300 (1.5)

400 (2.0)

500 (2.5)

600 (3.0)

800 (4.0)

1000 (5.0)

0 Deflection

0.010 (2.5)

0.017 (4.2)

0.028 (7.0)

0.038 (9.5)

0.069 (17.2)

0.107 (26.6)

22½ Deflection

0.011 (2.7)

0.019 (4.7)

0.031 (7.7)

0.043 (10.7)

0.078 (19.4)

0.120 (29.9)

45 Deflection

0.016 (4.0)

0.029 (7.2)

0.047 (11.7)

0.064 (15.9)

0.117 (29.1)

0.181 (45.1)

Table 6-5 Supply Registers Total Pressure Loss Average—in. wg (Pa)

Velocity—fpm (m/s)

300 (1.5)

400 (2.0)

500 (3.0)

600 (3.0)

800 (4.0)

900 (4.5)

0.033 (8.2)

0.060 (14.9)

0.092 (22.9)

0.068 (16.9)

0.122 (30.4)

0.187 (46.6)

0.134 (33.4)

0.238 (59.3)

0.302 (75.2)

0.272 (67.7)

0.483 (120.7)

0.055 (13.7)

0.098 (24.4)

0.614 (152.9)

0.152 (37.8)

0.222 (55.3)

0.390 (97.1)

0.496 (123.5)

0.025 (6.2)

0.060 (14.9)

0.080 (19.9)

0.100 (24.9)

0.180 (44.8)

0.230 (52.3)

0.012 (3.0)

0.020 (5.0)

0.032 (8.0)

0.046 (11.5)

0.080 (19.9)

0.102 (25.4)

0.033 (8.2)

0.055 (13.7)

0.088 (21.9)

0.126 (31.4)

0.220 (54.8)

0.275 (68.5)

Register, 0Deflection

0.054 (13.4)

0.090 (22.4)

0.144 (35.9)

0.207 (51.5)

0.360 (89.6)

C

Register, 30 Deflection

0.042 (10.5)

0.070 (17.4)

0.112 (27.9)

0.161 (40.1)

0.280 (69.7)

0.350 (87.2)

Rectangular Diffuser 12  24 in. (300  300 mm) 24  24 in. (600  600 mm) 12  21 in. (300  525 mm) Perforated Return Diffuser (Neck Velocity)

Register, 45 Deflection Register, Perforated Face

Table 6-6 Return Registers Total Pressure Loss Average—in. wg (Pa) HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.15


THIS PAGE INTENTIONALLY LEFT BLANK

6.16

HVAC SYSTEMS Testing, Adjusting & Balancing â&#x20AC;¢ Third Edition


CHAPTER 7

AIR SYSTEMS


CHAPTER 7

AIR SYSTEMS

7.1

INTRODUCTION

7.1.2

7.1.1

Categories

In general, air systems offer the following advantages:

This chapter covers design and application of air sys− tems used in single and multiple zoning applications. An air system is defined as a system that provides total sensible and latent cooling in the cold air supplied by the system. No additional cooling is required at the ter− minal units. Heating may be accomplished by the same airstream, either from the central system or at the terminal devices. In some applications, heating is ac− complished by a separate air, water, steam, or electric heating system. The term zone implies the provision or the need for separate thermostatic control, while the term room implies a partitioned area which may or may not require separate control. Air systems may be classified into two basic catego− ries: single−path systems and dual−path systems. 7.1.1.1

Single-Path Systems

Single−path systems are those which contain the main heating and cooling coils in a series flow air path, using common duct distribution system at a common air temperature to feed all terminal apparatus.

7.1.2.1

Air System Advantages

Consolidation

Air systems permit centralized location of major equipment, they consolidate operation and mainte− nance in unoccupied areas, and permit maximum choice of filtration systems, odor and noise control, and high quality, durable equipment. There is com− plete absence of drain piping, electrical equipment power wiring, and filters in the conditioned space. 7.1.2.2

Outdoor Air Cooling

The greatest advantage of air systems is the number of free cooling season hours that may be had when out− door air can be used for cooling in lieu of mechanical refrigeration. Economizer control systems usually are more trouble−free than enthalpy control systems. 7.1.2.3

Flexibility

Air systems allow a wide choice of zonability, flexibil− ity, and humidity control under all operating condi− tions, with simultaneous availability of heating and cooling during off season periods. 7.1.2.4

Heat Recovery

Single−path systems may be: a.

single duct, single zone, constant volume,

b.

single duct, variable air volume (VAV),

c.

single duct, VAV induction, and

d.

single duct zoned reheat.

7.1.1.2

Dual-Path Systems

Dual−path systems are those which contain the main heating and cooling coils in a parallel flow, or series− parallel flow air path, using either: (1) a separate cold and warm air duct distribution system, which is blended at terminal (dual duct systems); or (2) a sepa− rate supply duct to each zone, with blending of warm and cold air at the main supply fan. Dual−path systems may be: a.

dual duct (including dual duct, VAV), and

b.

multi−zone.

Air systems are readily adapted to heat recovery de− vices. 7.1.2.5

Design Freedom

Air systems allow full design freedom for optimum air distribution in air motion, draft control, and extenuat− ing local requirements. 7.1.2.6

Makeup Air

Air systems are best suited to applications requiring abnormal exhaust makeup. 7.1.2.7

Adaptable

Air systems are easily adaptable to automatic seasonal changeover and winter humidification. 7.1.3

Air System Disadvantages

Air systems have the following disadvantages: 7.1.3.1

Duct Space Requirements

Additional duct clearance requirements can penalize floor space for duct risers and fan rooms, and building

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

7.1


height for ceiling clearances. (Particularly true of low velocity systems). 7.1.3.2

Longer Fan Hours

In those systems which use air (not radiation) for pe− rimeter heating, longer fan−operating hours are re− quired to take care of unoccupied period heating (in low temperature locales). 7.1.3.3

System Balancing

In those systems which have no built−in zone self−bal− ancing devices, air balancing is difficult and may have to be done several times when a common air system serves areas which are not rented simultaneously. 7.1.3.4

Temporary Heat

Air heating perimeter systems may not be available for use during building construction as rapidly as perime− ter hydronic systems. 7.1.3.5

Terminal Devices

Accessibility to terminal devices demands close coop− eration between architectural, mechanical, and struc− tural designers.

ment stores, small individual shops in a shopping cen− ter, individual classrooms of schools, computer rooms, etc. A rooftop unit, for example, complete with refrig− eration system serving an individual space, would be considered a single zone system. The refrigeration sys− tem, however, could be remote and serve several single zone units in a larger installation. A schematic of a more sophisticated single zone cen− tral system is shown in Figure 7−2. The return air fan may be used if 100 percent outdoor air is used for cool− ing purposes, and may be eliminated if air is relieved from the space with very little pressure loss through the relief system. However, objectionable pressuriza− tion of conditioned spaces should be avoided to allow entrance doors to open or close normally. Control of the single zone system can be affected by varying the quantity of cooling medium, providing re− heat, face and bypass dampers or a combination of these. Single duct systems with reheat satisfy varia− tions in load by providing independent sources of heat− ing and cooling. When a humidifier is used, humidity control may be completely responsive to space needs. Single duct systems without reheat offer cooling flexi− bility, but cannot control summer humidity indepen− dent of temperature requirements. 7.2.2

7.1.3.6

Reheat Prohibition

Energy inefficiency of reheat type systems may pro− hibit use. 7.2

TYPES OF AIR SYSTEMS

7.2.1

Single Zone Systems

The simplest form of the air system is a single condi− tioner serving a single temperature control zone (see Figure 7−1). The unit may be installed within or remote from the space it serves and may operate with, or with− out distributing ductwork. Ideally, this can provide a system which is completely responsive to the needs of the space. Well designed systems can maintain tem− perature and humidity closely and efficiently. They can be shut down when desired, without affecting the operation of adjacent areas. A single zone system responds to only one set of space conditions. Its use is limited to situations where varia− tions occur almost uniformly throughout the zone or where the load is stable; but when multiple units are installed, they can handle a variety of conditions effi− ciently. Single zone systems are used in small depart− 7.2

Variable Air Volume (VAV) Systems

Control of dry bulb temperatures within a space re− quires that a balance be established between the space load and the air supplied to offset the load. The design− er may choose between varying the supply air temper− ature (constant volume) or varying the airflow volume (variable air volume) as the space load changes. To control part load volume reduction, supply air temper− atures and air volumes may be controlled simulta− neously. VAV systems (Figure 7−3), may be applied to interior or perimeter zones, with common or separate fan sys− tems, common or separate air temperature control, and with or without auxiliary heating devices. The VAV concept may apply to volume variation in the main system total airstream and to the zones of control. Variation of flow under control of a space thermostat may be accomplished by positioning a simple damper or a volume regulating device in a duct, in a VAV ter− minal box, or at a diffuser or grille. Depending on the complexity of the air distribution system, first cost considerations, the lowest throttling ratio expected at part load, and the complexibility of the initial and part load balancing problems, VAV may or may not be

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


POSSIBLE PREHEAT COIL

OUTDOOR AIR INTAKE

FILTERS

OUTDOOR DAMPER

COOLING COIL

HEATING COIL

SUPPLY AIR FAN RETURN AIR DAMPER

RELIEF AIR LOUVER

USE OF BAROMETRIC RELIEF AIR LOUVERS NOT RECOMMENDED

SUPPLY AIR SYSTEM

RELIEF AIR DAMPER RETURN AIR SYSTEM

FIGURE 7-1 SINGLE DUCT SYSTEM

EXHAUST AIR OR RELIEF AIR LOUVER & DAMPER RETURN AIR FAN

RETURN AIR DUCT REHEAT COIL IF BYPASS IS USED

RETURN AIR DAMPER BYPASS DAMPER SPRAYS

MIN. O.A. DAMPER MAX. O.A. DAMPER

SUPPLY AIR DUCT

FACE AND BYPASS DAMPER

SUPPLY AIR FAN OUTDOOR AIR LOUVER

MIXED AIR PLENUM

FILTERS PREHEAT COIL

COOLING COIL

REHEAT COIL IF BYPASS IS NOT USED

SPRAY PUMP

FIGURE 7-2 TYPICAL EQUIPMENT FOR SINGLE ZONE DUCT SYSTEM

HVAC SYSTEMS Testing, Adjusting & Balancing â&#x20AC;¢ Third Edition

7.3


OUTDOOR AIR INTAKE

POSSIBLE PREHEAT COIL

FILTERS

HEATING COIL

COOLING COIL

PRIMARY AIR DUCT S.P. CONTROLLER

SUPPLY FAN WITH RETURN S.P. CONTROL AIR DAMPER

OUTDOOR DAMPER

VAV TERMINAL UNITS

EXHAUST AIR LOUVER

EXHAUST AIR DAMPER

OPTIONALRETURN AIR FAN RETURN AIR SYSTEM

T

T

FIGURE 7-3 VARIABLE AIR VOLUME (VAV) SYSTEM

combined with fan volume or system static pressure controls. It is possible to permit system airflow volume varia− tions without fan volume variation by using a simple fan bypass. It is possible to vary zone air volume only, while keeping fan and system volume substantially constant, by dumping excess air into a return air ceil− ing plenum or directly into the return air duct system. These methods of system control do not provide the fan horsepower savings usually associated with VAV systems. 7.2.3

Terminal Reheat Systems

The terminal reheat system (Figure 7−4) is a modifica− tion of the single zone system. It permits zone or space control for areas of unequal loading, provides heating or cooling of perimeter areas with different exposures, and promotes process or comfort applications where close control of space conditions is desired. As the word reheat implies, the application of heat is a secondary process being applied to either precondi− tioned primary air or recirculated room air. Under present energy codes, the use of Na reheat system is dis− couraged or prohibited unless recovered heat is used. A single low pressure reheat system is produced when a heating coil is inserted into the duct system down− stream of the cooling coil(s). The more sophisticated 7.4

systems use higher pressure duct designs and pressure reduction devices to permit system balancing at the re− heat zone. The medium for heating may be hot water, steam, or electricity. A big advantage of the reheat sys− tem is that it has the capability of maintaining very close control of space humidity. The system is generally applied to hospitals, laborato− ries, or spaces where wide load variations are expec− ted. Terminal units are designed to permit heating of primary air, or secondary air inducted from the condi− tioned space, located either under the window or in the duct system overhead. Conditioned air is supplied from a central unit at a fixed cold air temperature de− signed to offset the maximum cooling load in the space(s). The control thermostat simply calls for heat as the cooling load in the space drops below maxi− mum. 7.2.4

Induction Reheat Systems

The induction reheat system is shown schematically in Figure 7−5. Full cooling capacity is provided in the pri− mary higher pressure airstream and supplied by the central equipment to the terminal. Zone control is ac− complished by heating the secondary or induced air− stream. This type of terminal is used when it is desir− able to introduce supply air to the space at a higher temperature, or permit higher space air movement without increasing the quantity of primary air over the amount of air required for cooling.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


POSSIBLE OUTDOOR PREHEAT AIR INTAKE COIL

COOLING COIL

RETURN AIR DAMPER

EXHAUST AIR EXHAUST AIR DAMPER LOUVER

POSSIBLE RETURN AIR FAN

SUPPLY AIR SYSTEM

SUPPLY AIR FAN REHEAT COIL NO. 1

RETURN AIR SYSTEM

REHEAT COIL NO. 2

T

T SUPPLY DUCT TO ZONE 1

SUPPLY DUCT TO ZONE 2

FIGURE 7-4 TERMINAL REHEAT SYSTEM The primary airN is discharged from nozzles arranged to induce room air into the induction unit approximate− ly four times the volume of the primary air. The in− duced air is cooled or heated by a secondary water coil. The water coil may be supplied by a 2−pipe system where either chilled water or heated water is available, but not simultaneously; by a 3−pipe system where sep− arate supplies of hot or chilled water are continuously available and, after passing through the unit, are mixed into a common return; or by a 4−pipe system, where a supply and return of hot water and chilled water are both continuously available. Induction units generally are located under the win− dow to offset winter downdrafts. Overhead installa− tions are limited, since ductwork connections carrying induced air have limited static pressure available, thereby decreasing induction air volume and unit ca− pacity. When installed under the window, this unit has the advantage of providing gravity heating during off− hour operation, permitting shutdown of the air system. The primary supply air fan operates at high pressures requiring high horsepower input. When balancing, at− tempt to reduce the primary air volume and pressure to the minimum required to operate the induction ter− minal units under full load conditions. Induction unit nozzles may be worn through many years of cleaning and operation, resulting in increased primary air quantity at lower air velocities with lower induced air volumes. Check each induction unit and either repair the nozzles or replace them before at− tempting any balancing work on the system.

7.2.5

Variable Air Volume (VAV) Reheat Systems

The VAV concept, when applied to reheat systems, permits flow reduction as a first step in control, there− by suspending the application of heat until flow condi− tions reach a predetermined minimum. By proper application of VAV, the reheat system may be designed to permit initial and operating cost sav− ings. With air volume selected for maximum instanta− neous peak loads rather than the sum of all peaks, the total system air volume is reduced. Also, any addition− al system diversity, such as areas with intermittent loads (conference rooms, office equipment rooms, etc.) may be included in the total volume reduction. When air volume reduction is used as a first step in control, reheat is not applied until the minimum vol− ume is reached. This procedure reduces system operat− ing costs appreciably for summer and intermediate weather. 7.2.6

Dual Duct Systems

7.2.6.1

Low Velocity Systems

The low velocity dual duct system distributes condi− tioned air through two parallel ducts. One duct carries cold air and the other warm air, allowing heating or cooling at all times. In each conditioned space or zone, automatic control dampers, responsive to a room ther− mostat, mix the warm and cold air in proper propor− tions to satisfy the prevailing heat load of the space. The return air fan shown in Figure 7−6 may be elimi− nated on small installations if provisions are made to

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

7.5


OUTDOOR AIR INTAKE

POSSIBLE PREHEAT COIL

FILTERS

HEATING COIL

COOLING COIL

HIGH VELOCITY PRIMARY AIR SYSTEM

T

RETURN SUPPLY FAN AIR DAMPER

OUTDOOR AIR DAMPER

T

INDUCED AIR POSSIBLE EXHAUST AIR EXHAUST AIR DAMPER RETURN LOUVER AIR FAN

SECONDARY WATER

INDUCTION

FIGURE 7-5 INDUCTION REHEAT SYSTEM relieve excess outdoor air from the conditioned spaces. They generally are required for economizer cooling cycles and in systems with substantial return air ductwork. 7.2.6.2

High Velocity Systems

High velocity dual duct systems operate in the same manner as the low velocity systems except that the supply fan runs at a much higher pressure and each zone requires a mixing box with sound attenuation. A large amount of energy is required to operate the fan at high pressure. When balancing, a close analysis of the pressure drops within the duct system should be made and the fan pressure reduced to the minimum re− quired to operate the mixing boxes. 7.2.6.3

Energy Savings Ideas

In conditions when there is no cooling load, install controls to close off the cold air duct; de−energize chillers and cold water pumps and operate as a single duct system, rescheduling the warmer air duct temper− ature according to heating loads only. Under conditions where there is no heating load, install controls to close off the warm air ducts; shut off hot water, steam, or electricity to the warm duct and operate the system with the cold duct air only; resched− uling supply air temperature according to cooling loads. 7.6

Replace obsolete or defective mixing boxes to elimi− nate leakage of hot or cold air when the respective damper is closed. Provide volume control for the supply air fan and re− duce capacity preferably by speed reduction when both the hot deck and cold deck air quantities can be reduced to meet peak loads. Reducing the heat loss and heat gain provides an opportunity to reduce the amount of air circulated. When there is more than one air handling unit in a dual air system, modify duct work, if possible, so that each unit supplies a separate zone to provide an opportunity to reduce hot and cold duct temperatures according to shifting loads. Change dual duct systems to VAV systems when ener− gy analysis is favorable and the payback in energy saved is sufficiently attractive by adding VAV boxes and fan control. 7.2.7

Multi-Zone Systems

The multi−zone system (Figure 7−7) is applicable for serving a relatively small number of zones from a single, central air handling unit. The requirements of the different zones are met by mixing cold and warm air through zone dampers at the central air handler in response to zone thermostats. The mixed conditioned air is distributed throughout the building by a system of single zone ducts. Either packaged HVAC units complete with all components or field fabricated HVAC components may be used. Return air is usually handled in a conventional manner.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


EXHAUST AIR LOUVER

EXHAUST AIR DAMPER

RETURN AIR FAN (OPTIONAL)

SUPPLY AIR

RETURN AIR DAMPER

HEATING COIL

FAN

Ô Ô Ô

OUTDOOR AIR INTAKE

FILTERS

OUTDOOR AIR DAMPER

COOLING COIL RETURN AIR SYSTEM

POSSIBLE PREHEAT COIL

T

ZONE 2

T

ZONE 1

MIXING BOXES SUPPLY DUCT

SUPPLY DUCT

SYSTEM (COLD)

SYSTEM (HOT)

FIGURE 7-6 DUAL DUCT HIGH VELOCITY SYSTEM Multi−zone systems are somewhat similar to dual duct systems. They can provide a smaller building with some of the advantages of dual duct systems at a lower first cost with a wide variety of packaged HVAC units, but are limited to handling smaller projects by multi− ple runs of single zone ducts. Most packaged HVAC units lack the control sophistication for comfort and operating economy that can be built into dual duct sys− tems.

Multi−zone systems may handle more than one room with a single duct. Multizone packaged HVAC equip− ment is usually limited to about 12 zones, while built− up systems may have as many as can be physically in− corporated into the layout.

7.2.7.1

VAV Terminal Devices

VAV may be applied to multi−zone systems with pack− aged or built−up systems having the necessary zone volume regulation and fan controls. However, it is sel− dom applied in this manner for entire distribution sys− tems except for TV studios and other critical noise lev− el applications. More often, a few select rooms in a zone may incorporate VAV terminal devices, when off−peak requirements permit this approach and air balancing considerations indicate there will be no problems from omission of fan control or static pres− sure regulation. 7.2.7.2

Zone Coils

Some multi−zone units have individual heating and cooling coils for each zone supply duct. These units

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

7.7


use less energy then units with common coils as the supply air is heated or cooled only to the degree re− quired to meet the zone load. 7.2.7.3

Install controls or adjust existing controls to give the minimum hot deck temperature and maximum cold deck temperature consistent with the loads of critical zones.

TAB Arrange the controls so that when all of the hot duct dampers are partially closed, the hot deck tempera− tures will progressively reduce until one or more zone dampers are partially closed; the cold duct tempera− ture will progressively increase until one or more of the zone dampers are fully opened.

Before starting testing and balancing work, analyze multi−zone systems carefully and treat each zone as a single zone system. Adjust air volumes and tempera− ture accordingly. 7.2.7.4

Energy Savings Ideas Install controls to shut off the fan and all heating con− trol valves during unoccupied periods in the cooling season, and shut off the cooling valve during unoccu− pied periods in the heating season.

Hot and cold deck dampers are often of poor quality and allow considerable air leakage even where fully closed. Modify these dampers to avoid leakage.

EXHAUST AIR LOUVER

EXHAUST AIR DAMPER

POSSIBLE RETURN AIR FAN RETURN AIR SYSTEM ZONE 4

SUPPLY AIR FAN

OUTDOOR AIR INTAKE OUTDOOR AIR DAMPER

FILTERS COOLING COIL

HEATING COIL ZONE MIXING DAMPERS

RETURN AIR DAMPER

ZONE 3 ZONE 2 ZONE 1 SUPPLY AIR SYSTEMS

POSSIBLE PRE-HEAT COIL

FIGURE 7-7 MULTI-ZONE SYSTEM

7.8

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


7.3

AIR SYSTEM DESIGN

7.3.1

Introduction

Knowing the basics of good air system duct design will allow the TAB technician to determine if HVAC sys− tems can be balanced properly, in addition to helping solve routine problems while balancing. This section contains highlights only of duct design basics found in the SMACNA HVAC Systems9Duct Design manual. The tables and charts found in the Appendix of this manual are not as extensive as those found in the duct design manual, but they should be adequate for all TAB work. Air systems may be designed at higher or lower pres− sures. Higher friction rates and system pressures are required with higher velocity systems to reduce duct sizes and save space. For some lower velocity systems, higher pressures may be desirable for ease of balanc− ing and for flow control regulators that have a substan− tial pressure drop. On any of the variable air flow systems, there will be an infinite set of operating conditions which will create rates of airflow in the ducts entirely different from those used in the design. Every effort should be made when designing the ductwork to reduce the total fan power needed. This will assure a quieter system, reduced duct leakages, and in most cases maximum operating economy for the system owner. High velocity duct systems have been used mainly be− cause of space limitations created by architectural and structural practices. On the other hand where space is not at a premium, the use of high velocities and pres− sures are not economical. Some installations have areas of great space restriction where duct velocities must be higher. As soon as these points are passed and space becomes less critical, ve− locity rates should be sharply dropped, then gradually reduced toward the end of the duct system. 7.3.2

Equal Friction Design Method

The equal friction method of duct sizing probably has been the most universally used means of sizing low pressure supply air, return air and exhaust air duct sys− tems. It also is being adapted by many HVAC system designers for use in medium pressure systems. It nor− mally has not been used for sizing high pressure sys− tem. This design method automatically reduces air ve− locities in the direction of the airflow, so that by using a reasonable initial airflow velocity the chances of

introducing airflow generated noise are reduced or eliminated. When noise is an important consideration, the system velocity may be readily checked at any point during the design. Then there is the opportunity to reduce ve− locity created noise by increasing duct size or adding sound attenuation materials (such as duct lining). The major disadvantages of the equal friction method are there are no natural provisions for equalizing pres− sure drops in the branches (except in the few cases of a symmetrical layout); and there are no means of pro− viding the same static pressure behind each supply or return terminal device. Consequently, balancing can be difficult, even with a considerable amount of damp− ering in short duct runs. However, the equal friction method can be modified by designing portions of the longest run with different friction rates from those used for the shorter runs (or branches from the long run). Duct static regain (or loss) due to airflow velocity changes is included in the duct fitting pressure losses calculated using the duct fitting loss coefficient tables found in the Appendix. Otherwise, the omission of sys− tem static regain, when using older tables, could cause the calculated system fan static pressure to be greater than actual field conditions, particularly in larger, more complicated systems. Equal friction does not mean that total friction remains constant throughout the system. It means that a specific friction loss or stat− ic pressure loss per 100 feet (per meter) of duct is se− lected before the ductwork is laid out, and that this loss per 100 feet (per meter) is used constantly throughout the design. The SMACNA Duct Design System Calcu− lator makes this design method easy to use. 7.3.3

Supply Air Duct Sizing Procedures

To size the main supply air duct leaving the fan, the usual procedure is to select an initial velocity from Figures A−1 and A−2 found in the Appendix of this Manual. This velocity could be selected above the low velocity shaded section of the Duct Friction Loss Chart (I−P) A−1, and the Duct Friction Loss Chart (SI) A−2 if higher sound levels and energy conservation are not limiting factors. The charts on Appendix A.1 and A.2 are used to determine the friction loss by using the de− sign air quantity (cfm or L/s) and the selected velocity (fpm or m/s). A friction loss value commonly used for low pressure duct sizing is 0.1 in. wg per 100 feet of ductwork (0.8 Pa/m), although other values found in the low velocity shaded area, both lower and higher, are used by some designers as their standard or for spe−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

7.9


cial applications. This same friction loss value gener− ally is maintained throughout the design, and the re− spective round duct diameters are obtained from the charts in Figures A−1 and A−2. The round duct diame− ters are used to select the equivalent rectangular duct sizes from chart A−1, unless round ductwork is to be used. Round ductwork is generally preferred on the higher pressure duct systems. The friction losses of each duct section may be cor− rected for otherN materials and construction methods by use of Table A−1 and Figure A−3. The correction factor from the Duct Friction Loss Correction Factors Figure A−3 is applied to the duct friction loss as only for the straight sections of the duct. The airflow rate used in the next section (and subsequent sections) of the main supply duct after each branch takeoff, is the original airflow rate (cfm or L/s) of the preceding sec− tion reduced by the amount of airflow into the branch. Using charts in Figures A−1 and A−2, the new airflow rate value (using the recommended friction rate of 0.1 in. wg per 100 ft (or 0.8 Pa/m) will determine the duct velocity and diameter for that section. The equivalent rectangular size of that duct section again is obtained from the Circulation Equivalents of Rectangular Ducts for Equal Friction and Capacity (I−P) (2) Dimen− sions in Inches Table A−2 and the (SI) version Table A−3 (if needed). All additional sections of the main supply air duct and all branch duct sections can be sized using charts in Figures A−1 and A−2 and approxi− mately the same friction loss rates. The pressure drop at each terminal device or air outlet (or inlet) of a small duct system, or of branch ducts of a larger system, should not differ more than 0.05 in. wg (12 Pa). If the pressure difference between the termi− nals exceeds that amount, dampering would be re− quired that could create objectionable air noise levels, and balancing may become more complicated. Example 7.1 (IP) A 70 foot section of N36 × 24 in. galvanized sheetN met− al duct is handling 10,000 cfm of air. What is the actual pressure loss and velocity of this duct section, and is it in the ?low velocity" category?

Solution Using Table A−2, a 36 × 24 in. duct has a circular equivalent of 32.0 inches. From Figure A−1, the 32.0 in. equivalent diameter duct has a velocity of 1800 fpm at a 10,000 cfm airflow rate and a pressure loss rate of 7.10

0.12 in. wg per 100 ft. It is in the low velocity shaded area. From Table A−1 the duct roughness category is medium smooth and from Figure A−3, a correction fac− tor is not needed. 0.12in.wg  70ft. Sectionpressureloss  100ft. 8 8  0.084in.wg NOTE: The pressure loss of any duct% fittings or ac− cessories contained in this 70 foot section of duct would be added to the above duct friction pressure loss.

Example 7.1 (SI) A 21 meter section of 900 × 600 mm galvanized sheet metal duct is handling 5000 L/s of air. What is the actu− al pressure loss and velocity of this duct section, and is it in the ?low velocity" category?

Solution Using Table A−3, a 900 × 600 mm duct has a circular equivalent of 799 mm. From Figure A−2, the 799 mm equivalent diameter duct has a velocity of 10 m/s at a 5000 L/s airflow rate, and a pressure loss rate of 1.1 Pa/ m. It is in the low velocity shaded area. From Table A−1, the duct roughness category is medium smooth, and from Figure A−3, a correction factor is not needed. Section pressure loss = 1.1 Pa/m × 21 m = 23.1 Pa 7.3.4

Modified Design Method

The modified equal friction method is used for sizing duct systems that are not symmetrical or that have both long and short runs. Instead of depending upon volume dampers to artificially increase the pressure drop of short branch runs, the branch ducts are sized (as nearly as possible) to dissipate (bleed off) the available pres− sure by using higher duct friction loss values. Only the main duct, which is usually the longest run, is sized by the original duct friction loss rate. Care should be exer− cised to prevent excessive high velocities in the short branches (with the higher friction rates). If calculated velocities are found to be too high, then duct sizes must be recalculated to yield lower velocities, and opposed blade volume dampers or static pressure plates must be installed in the branch duct at or near the main duct to dissipate the excess pressure. Regardless, it is good de− sign practice to always include balancing dampers in HVAC duct systems to balance the airflow to each branch.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Return Air Duct Systems

32"

× 16"

A DAMPER

× 16"

Return air ducts should be sized using the equal fric− tion method at lower velocities. One scheme to simpli− fy return air systems is to use the space above hung ceilings or corridors for return air plenums and collect all return airflows at central points on each floor. In some localities there are code restrictions to using this method.

30’

45’

32"

7.3.5

2000 cfm

Terminal Unit Ductwork

7.4

DUCT SIZING EXAMPLES

7.4.1

Sizing In I-P Units

Example 7.2 (I−P Units) What is the total pressure loss of the portion of the fi− brous glass HVAC duct system shown in Figure 7−8.

× 16"

Low pressure ducts leaving terminal units are sized as any other conventional low pressure ductwork. Expe− rience indicates that the least expensive and the safest way to assure a quiet installation is to have some length of lined ductwork on the leaving side of terminal units. Lined ductwork, especially when it contains one or two elbows, can be a very effective sound attenuator. Lined ductwork helps if noise regeneration should oc− cur in the air distributing system because of poorly constructed ducts, fittings, and taps.

B 35’

22"

7.3.6

C

2000 cfm

25’

2000 cfm

D

14"

× 16"

×

28" 14" Grille PD=0.12 in. wg

FIGURE 7-8 SYSTEM LAYOUT (I-P UNITS)

Figure A−3, Correction factor = 1.43 Table A−14D, Opposed Blade Damper (set wide open), C = 0.52 Table A−10F , Sq. Elbow, 4.5 in. ?R", single thickness vanes, C = 0.23

Solution Using the figures, tables, and charts from the Appen− dix:

Velocity = 6000 cfm/32 × 16/144 = 1688 fpm Table A−4, Vp = 0.18 in. wg

Duct AB: Duct ABC32 × 16 inches, 6000 cfm Table A−2, Circ. Equiv. = 24.4 in. diameter

DuctC45 ft + 30 ft = 75 ft/100 ft × 0.17 × 1.43 = 0.182 in. wg DamperCVp × C = 0.18 × 0.52 = 0.094 in. wg

Figure A−1, Friction loss = 0.17 in. wg/100 ft, velocity = 1850 fpm

ElbowCVp × C = 0.18 × 0.23′ = 0.041 in. wg

Table A−1, category = medium rough

Duct Section <AB" Total = 0.317 in. wg

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

7.11


Duct BC:

10 m

Duct BCC22 × 16 in., 4000 cfm

× 400

A

× 400

DAMPER

15 m

550

Table A− 2, Circ. Equiv. = 20.4 inches diameter

800

Figure A−1, Friction loss = 0.19 in. wg/100 ft, velocity = 1790 fpm

1000 L/s

B

× 400

Figure A−3, Correction factor for medium rough = 1.45

550

Table A−11C, 32 × 16 inches to 22 × 16 inches

12 m

60

1000 L/s

C

contraction, A1/A = 1.45

8m

C = 0.06 Velocity = 4000 cfm/22 × 16/144 = 1636 fpm,

1000 L/s

D

350

Table A−4, Vp = 0.17 in. wg DuctC35 ft/100 ft × 0.1 9 × 1.45 = 0.096 in. wg

× 400

×

700 350 Grille PD= 30 Pa

FIGURE 7-9 SYSTEM LAYOUT (SI)

TransitionCVp × C = 0.17 × 0.06 = 0.010 in. wg

TransitionCVp × C = 0.11 × 0.06 = 0.007 in. wg

Duct Section <BC" Total = 0.106 in. wg Duct CD:

ElbowCVp × C = 0.11 × 1.2 = 0.132 in. wg

Duct CDC14 × 16 inches, 2000 cfm

Grille (given) = 0.120 in. wg

Table A−2, Circ. Equiv. = 16.3 in. diameter Figure A−1, Friction loss = 0.16 in. wg/100 ft

Duct Section <CD" Total = 0.316 in. wg Duct−Run ABCD:

Figure A−3, Correction factor for medium rough = 1.43 (velocity = 1380 fpm)

ABC0.317 in. wg BCC0.106 in. wg

Table A−11C, 22 × 16 in. to 14 × 16 in. 60 contraction, A1/A = 1.57, C = 0.06 Table A−10D, 90 mitered elbow, H/W =6/14 = 1.14, C = 1.2

CDC0.316 in. wg Total Pressure Loss)0.739 in. wg 7.4.2

Sizing In SI Units

Velocity = 2000 cfm/14 × 16/144 = 1286 fpm Table A−4, Vp = 0.11 DuctC25 ft/100 ft × 0.16 × 1.43 = 0.057 in. wg 7.12

Example 7−2 (SI) Find the total pressure loss of the portion of the fibrous glass HVAC duct system shown in Figure 7−9.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Solution Using the figures, tables, and charts from the Appen− dix:

C = 0.06 Velocity = 2.0 m3/s/(0.55 × 0.4) = 9.1 m/s Table A−5, Vp = 49.9 Pa

Duct AB:

DuctC12 m × 1.9 Pa/m × 1.46 = 33 Pa

Duct ABC800 × 400 mm, 3000 L/s: Table A− 3, Circ. Equiv. = 609 mm diameter, Figure A−2, Friction loss = 1.6 Pa/m,

TransitionCVp × C = 49.9 × 0 .06 = 3 Pa Duct Section <BC" Total = 36 Pa Duct CD:

velocity = 10 m/s

Duct CDC350 × 400 mm, 1000 L/s:

Table A−1, category = medium rough,

Table A−3, Circ. Equiv. = 409 mm diameter,

Figure A−3, Correction factor = 1.45,

Figure A−2, Friction loss = 1.5 Pa/m

Table A−14D, Opposed Blade Damper (set wide open), C = 0.52,

Velocity = 7.7 m/s

Table A−10F, Sq. Elbow, 114 mm ?R", single thickness vanes, C = 0.23, Velocity = 3.0 m3/s/0.8 × 0.4 = 9.4 m/s (1000 L/s = 1.0 m3/s) (1000 mm = 1.0 m)

Figure A−3, Correction factor for medium rough = 1.46 Table A−11C, 550 × 400 mm to 350 × 400 mm, 60 contraction, A1/A = 1.57, C = 0.06

Table A−5, Vp = 53.2 Pa DuctC10m + 15m = 25m × 1.6 Pa/m × 45 = 58 Pa DamperCV p × C = 53.2 × 0.52 = 28 Pa ElbowCV p × C = 53.2 × 0.23 = 12 Pa Duct Section <AB" Total = 98 Pa Duct BC:

Table A−10D, 90 mitered elbow, H/W = 1.14, C = 1.2 Velocity = 1.0 m3/s/(0.35 × 0.4) = 7.1 m/s Table A−5, Vp = 30.3 Pa DuctC8 m × 1.5 Pa/m × 1.46 = 18 Pa TransitionCVp × C = 30.3 × 0.06 = 2 Pa ElbowCVp × C =30.3 × 1.2 = 36 Pa

Duct BCC550 × 400 mm, 2000 L/s: Table A−3, Circ. Equiv. = 511 mm diameter, Figure A−2, Friction loss = 1.9 Pa/m,

Grille (given) = 30 Pa Duct Section <CD" Total = 86 Pa Duct−Run ABCD:

velocity = 10 m/s

ABC98 Pa

Figure A−3, Correction factor for medium rough = 1.46

BCC36 Pa

Table A−11C, 800 × 400 mm to 550 × 440mm, 60 contraction, A1/A = 1.45,

CDC86 Pa Total Pressure LossC220 Pa

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

7.13


7.5

SUMMARY

if it is, and calculations are painstakingly made in ac− cordance with all tables and charts, the actual total sys− tem losses can vary from the design losses. Generally, balancing a system adds additional losses to the system total or static pressure losses when compared to a sys− tem with all dampers wide open.

Step by step procedures for sizing a low pressure sup− ply and return system and a medium pressure system can be found in the SMACNA HVAC Systems9Duct Design manual. On new construction work, and with a good background in duct system design, TAB techni− cians will be able to recognize where problems exist and what can be done to correct them before construc− tion has progressed to the point where changes are dif− ficult to make.

Testing and balancing personnel also must know about diffusers and air patterns in order to establish a truly comfortable system with even temperatures through− out the entire space. The system designer and the TAB team have the same objectivesCa properly operating HVAC system. Understanding system design is the key to it all.

The TAB technician must know how duct systems are designed in order to troubleshoot problem jobs. The ductwork is seldom installed exactly as shown. Even

POOR POOR

BEST

IMPROVED

POOR

IMPROVED

POOR

IMPROVED

IMPROVED BEST

POOR

IMPROVED

IMPROVED

IMPROVED

POOR

IMPROVED

POOR

FIGURE 7-10 FAN DUCT CONNECTIONS

7.14

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

IMPROVED


CHAPTER 8

HYDRONIC EQUIPMENT


CHAPTER 8 8.1

PUMPS

8.1.1

Pump Laws

HYDRONIC EQUIPMENT

Centrifugal pumps are used in heating and air condi− tioning systems to produce fluid flow in piping. This section provides general information on centrifugal pumps and their application to heating and air condi− tioning or hydronic systems. Other pumps, such as re− ciprocating or rotary pumps, play a minor role in the HVAC industry. Pumps interact with a hydronic system in almost the same manner as fans do in an air system, and pump laws are similar to fan laws. Pumps usually are direct driven by being coupled to the shaft of the motor, and their speed is not changed unless a variable speed drive or motor is used. If the pump speed can be changed, the volume of liquid that is pumped will vary directly as the speed. The pressure or head imposed within the piping system will vary as the square of the rpm. The power required to run the pump will vary as the cube of the rpm.

Cutoff Discharge Nozzle Impeller



The pressure or head within the system varies directly as the square of the diameter of im− peller,



The horsepower or power required to drive the pump varies directly as the cube of the di− ameter of the impeller.

It is more efficient to change the pump impeller than to throttle a pump (using a discharge valve) to ?short circuit" part of output flow. On hydronic systems having larger pumps, or for spe− cialized systems requiring the ability to adjust water flow based on system loads, a variable frequency drive (VFD) may be used. Although the pump motor ramp up and ramp down set− points are usually programmed differently than for a motor driving a fan, the same VFD can be used. Since motor horsepower input is a cubic function of pump rpm, even a small 10 percent reduction in pump flow can reduce motor horsepower and corresponding electrical consumption over 30 percent. Most HVAC systems are sized for maximum calcu− lated loads, which may only occur for several hours during an entire year. The ability to reduce system wa− ter flow rates to meet reduced radiation or air handling unit coil loads for the remainder of the year can result in a significant utility cost savings.

Casing or Volute

Vanes

Impeller “Eye” Shaft

FIGURE 8-1 TYPICAL CENTRIFUGAL PUMP CROSS SECTION

In small HVAC hydronic systems having pumps under 5 hp, the design flow rate is usually reduced by the pump supplier by reducing or ?trimming" the diameter of the pump’s impeller. Changing the diameter of the impeller has the same affect as changing the pump speed. Pump laws found in subsection 2.3.5 of Chapter 2 and in the Appendix can be restated in a different way:



The pump volume (gpm) varies directly with the impeller diameter,

As recommended in section 3.6 dealing with VFD con− trol of HVAC fans, understanding how to program VFD motor controls can be very helpful for any TAB controller. 8.1.2

Pump Types

Types of centrifugal pumps used in the heating and air conditioning industry can be defined by the type of im− peller, number of impellers, type of casing, method of connection to driver, and mounting position. Two types of pump impellers are used in these pumps, single suction and double suction. The single suction impeller has one suction or intake, while the double suction impeller has two suctions or intakes. Most cen− trifugal pumps for heating and air conditioning are single suction; the significant example of double suc− tion impeller is the single stage, horizontal split−case pump. Pumps also may have single or multiple impel− lers. When they have multiple impellers, they are called multistage pumps. As with impellers, there are two basic types of casings for these pumps, volute and diffuser. The volute types

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

8.1


include all pumps that collect water from the impeller and discharge it perpendicularly to the pump shaft. Diffuser−type casings collect water from the impeller and discharge it parallel to the pump shaft. All pumps described here are the volute type, except the vertical turbine pump, which is a diffuser type.

sure and temperature limitations vary depending on the liquid being pumped and the style of seal. The seal material and style are supplied by the manufacturer af− ter being informed of the kind of liquid to be pumped and the temperature and pressure limitations. 8.1.3.4

Pumps can be classified by their method of connection to an electric motor, close coupled or flexible coupled. The close coupled pump has the impeller mounted di− rectly on a motor shaft extension, while the flexible coupled pump has an impeller shaft supported by a frame or bracket, connected to the electric motor through a flexible coupling. Pumps also are labeled by their mounting position; either horizontal or vertical. Seven significant types of pumps are used in heating and air conditioning or hydronic systems, and are shown in Figure 8−2. Many variations of these pumps are offered by manufacturers for particular applica− tions. 8.1.3

Packing Glands

Pumps with packing glands are extensively used, par− ticularly where abrasive substances included in the water are not detrimental to system operation. Some leakage at the packing gland is needed to lubricate and cool the area between the packing material and the shaft. 8.1.3.5

Shaft Sleeves

Shaft sleeves protect the motor or pump shaft, espe− cially with packing.

Pump Construction Features FORM

Important construction features of centrifugal pumps follow. 8.1.3.1

Material Types

Centrifugal pumps are generally offered in bronze fitted, all bronze, or iron fitted construction. In bronze fitted construction, the impeller, shaft sleeve (if used), and wear rings are bronze, and the casing is cast iron. These construction materials refer to the liquid end of the pump (those parts of the pump that contact the liq− uid being pumped). 8.1.3.2

Stuffing Box

The stuffing box is that portion of the pump where the rotating shaft enters the pump casing. To seal undesir− able leakage at this point, a mechanical seal or packing is used in the stuffing box.

TYPICAL APPLICATIONS  Residential Hydronic System  Domestic Hot Water  Recirculation  Multizone Circulation

CIRCULATOR

CLOSE-COUPLED END SUCTION

FRAME-MOUNTED END SUCTION

BASE-MOUNTED HORIZONTAL SPLIT CASE SINGLE-ST AGE DOUBLE-SUCTION BASE-MOUNTED MULTISTAGE HORIZONTAL SPLIT CASE

 Cooling Tower  Condenser Water  Chilled Water  Primary & Secondary  Hot Water  Boiler Feed  Condensate Return

VERTICAL INLINE VERTICAL TURBINE SINGLE-ST AGE OR MULTISTAGE

8.1.3.3

Mechanical Seals

Pumps with mechanical seals are used successfully in a wide variety of applications. Like pumps, many styles and types of seals are available. There are unbal− anced and balanced (for higher pressures) seals. Inside seals operate inside the stuffing box while outside seals have their rotating element outside the box. Pres−

8.2

SUMP MOUNTED

CAN PUMP

FIGURE 8-2 DESCRIPTIONS OF CENTRIFUGAL PUMPS USED IN HYDRONIC SYSTEMS

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Impeller Type

Type

Number of Impellers

Motor Mounting Position

Motor Connection

Casing

Circulator

Single suction

One

Volute

Flexible coupled

Horizontal

Close coupled, end suction

Single suction

One or two

Volute

Close coupled

Horizontal

Frame mounted, end suction

Single suction

One or two

Volute

Flexible coupled

Horizontal

Double suction, horizontal split case

Double suction One

Volute

Flexible

Horizontal

Horizontal split split case, multistage

Single suction

Two to five

Volute

Flexible coupled

Horizontal

Vertical inline

Single suction

One

Volute

Flexible or close

Vertical

Vertical turbine

Single Suction

One to twenty Diffuser

Flexible coupled

Vertical

Table 8-1 Characteristics of Centrifugal Pumps

Positive Displacement Pumps

Centrifugal Pumps

Characteristics Rotary

Piston

Radial

Mixed Flow

Axial Flow

Flow

Even

Pulsating

Even

Even

Even

Effect of increasing head: on flow on bhp

Negligible decrease Increase

Increase

Decrease Decrease

Decrease Large increase

Effect of decreasing head: on flow on bhp

Decrease Small decrease to large increase

Negligible increase Decrease

Decrease

Increase Increase

Increase Slight increase to decease

Increase Decrease

Up to 30% increase Decrease 50%-60%

Considerable increase 10% decrease to 80% increase

Large increase Increase 80%-150%

Effect of closing discharge valve: on pressure

Can destruct unless relief valve is used Increase to destruction

on bhp

Table 8-2 Characteristics of Common Types of Pumps 8.1.3.6

Bearings

Ball bearings are most frequently used in larger pumps. Circulators use sleeve type bearings for motor and pump bearings.

8.1.3.7

Wearing Rings

Wearing rings are for the impeller and/or casing. They are replaceable and prevent wear to the impeller or casing.

HVAC SYSTEMS Testing, Adjusting & Balancing â&#x20AC;˘ Third Edition

8.3


8.1.3.8

Balance Rings

The balance ring is placed on the back side of a single inlet, enclosed impeller to reduce the axial load. Double inlet impellers are inherently balanced axially.

Straight Edge C L Pump and Water P

M Aligned Gap

8.1.3.9

Rotation of any pump is fixed by the configuration and type of vanes and the suction and discharge connec− tions. An arrow to indicate proper direction is often cast directly into the casing metal. In addition to prop− er position of the pump in the piping, rotation is also dependent upon the motor or driven rotation. Rotation of motor and pump must be tested prior to operation. However, pumps with mechanical seals must not be run dry, even for ?bumping," to determine rotation.

C L Pump M

A pump may be driven by any appropriate means. For the most part, environmental system pumps are motor driven, and the shafts of the motor and pump are con− nected end to end by some type of coupling. Some pumps, usually those employed in pumping fuel oil, are belt driven. In this case the belts may be used as a protective device, either slipping or breaking before damage to the pump can occur in the event of overload. The couplings between motor shaft and pump shaft are made in two pieces so that the two coupling halves may be disconnected for removal of the pump without disturbing the motor, for running the motor indepen− dently of the pump, or for removal of the motor with− out disturbing the pump. The coupling also serves as a means of adjustment of the pump and shaft align− ment. The ideal alignment condition is that both shafts are in a straight line and concentric under all condi− tions of operation and shut down. Because of tempera− ture changes, an unequal expansion of parts causes a change of alignment during operation. Base mounted pumps, especially in larger sizes, re− quire at least an alignment check in the field. This may be done in a superficial but often satisfactory way with a straight edge since the outside perimeters of the cou− pling halves are machined to the same diameter and are perpendicular to each shaft. Center lines and cou− pling faces must be true as shown in Figure 8−3. 8.1.3.11 Operating Speeds Operating speeds of motors may be selected in the range between 600 and 3500 rpm, with 1800 rpm being the most common speed. Pumps operating at higher speeds are generally less expensive, but for quieter

P C L Motor Coupling (Typ.) Misaligned C L Pump

C L Motor M Gap and Angle Straight Edge Misaligned

8.1.3.10 Pump Drives

8.4

Straight Edge

Pump Rotation

FIGURE 8-3 COUPLING ALIGNMENT WITH STRAIGHT EDGE performance or lesser NPSH requirements, lower speeds are preferred. 8.1.4

Pump Pressure or Heads

The purpose of a pump for HVAC work is to establish fluid flow and produce sufficient pressure to overcome the resistance of a system and the system components at the design flow rate. 8.1.4.1

Pump “Head” Definitions

When working with pumps, the word head often will be used to define pressure. Definition of these and oth− er common head terms are noted here, even though some may be defined again under other discussions: Friction head is the pressure in psi or feet (pascals or meters) of the liquid pumped which represents system resistance that must be overcome. Velocity head is the pressure needed to accelerate the liquid being pumped. (For practical purposes, the ve− locity head is insignificant and usually can be ignored in HVAC system calculations.) Static suction lift is the distance in feet (meters) be− tween the pump centerline and the source of liquid be− low the pump centerline. Suction lift is the combination of static suction lift and friction head in the suction piping when the source of liquid is below the pump centerline.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Suction head is the positive pressure on the pump inlet when the source of liquid supply is above the pump centerline. Static suction head is the positive vertical height in feet (meters) from the pump centerline to the top of the level of the liquid source. Dynamic suction lift is the sum of suction lift and ve− locity head at the pump suction when the source is be− low the pump centerline. Dynamic suction head is positive static suction head minus friction head and minus velocity head. Dynamic discharge head is static discharge head plus friction head plus velocity head. Total dynamic head is dynamic discharge head (static discharge head, plus friction head, plus velocity head) plus dynamic suction lift, or dynamic discharge head minus dynamic suction head. 8.1.4.2

Pressure Relationships

For the pressure relationship to the inlet and suction side of the pump, the discharge pressure is higher. In the process of establishing this head, the impeller pro− duces a lower or relatively negative pressure on the suction side. It is important to note the deliberate use of the term rel− ative in a discussion of the pressures which pumps pro− duce. The system elevation static pressure is of major consequence in the liquid system. During the time when the pump is running, there is a redistribution of pressures in the system because of the combination of the elevation pressures and the pressures produced by the pump. However, when the pump is shut down, the system pressures return to the same values of elevation static as before the pump was started. The pressures produced by the pump merely add to or subtract from the initial shutdown pressures. Since the operating discharge pressure produced by the pump is an increase, this value is added to the shut− down pressure in the discharge piping. Since the oper− ating suction pressure produced by the pump is a de− crease, this value is subtracted from the shutdown pressure in the suction piping. Assuming that the sys− tem piping and equipment losses were properly calcu− lated and that the pump was properly selected to over− come those losses and to withstand the system static and dynamic pressures, it would be expected that the pump would produce the required fluid flow. Howev−

er, this may not be the case because of the pump’s sen− sitivity to the pressure conditions on its inlet (the dis− charge conditions of the pump do not generally present a problem). 8.1.4.3

Frequent Pumping Problems

The fluid being pumped, usually water, generally con− tains some entrained air which has been absorbed as a result of atmospheric pressure when the fluid was ex− posed to the atmosphere prior to being introduced into the system. This air is released because of an increase in fluid temperature, a decrease in fluid pressure, or because of the fluid vapor pressure. Air driven out of the water when it is heated must be vented from the piping, often several times at the be− ginning of the season, and throughout the heating op− eration if fresh water must be continuously added to re− place that amount dripped through pump packing. Most of the air released in the process of heating of the water should be removed at or near the heat exchanger or hot water generator. Therefore, there may be a point in the environmental fluid pumping system where air may be released from the fluid being pumped if the pressure is low enough and/or the liquid may change to a gas. Should these conditions occur, the pump, which has been designed to move liquid, is generally unable to cope, and the flow of liquid is either greatly reduced or stopped com− pletely. However, at some point within the pump where the impeller produces sufficient pressure, the bubbles of gaseous liquid will be re−liquified and the bubbles of air will be reabsorbed. This transition oc− curs suddenly and is accompanied by crackling or ex− plosive noise. The phenomenon is called cavitation and may cause destructive pitting and wearing of the impeller and casing as well as noise and vibration. Any one or all of these conditions will reduce pump perfor− mance and life. 8.1.4.4

Net Positive Suction Head (NPSH)

To eliminate the cavitation problem, it is necessary to maintain a minimum suction pressure at the inlet side of the pump. The actual value in psi or ft. wg (pascals or meters) of internal pump losses depends on the pump size and de− sign, and the volume of water being pumped. This must be determined by the pump manufacturer, and is given by numerical values of NPSH. Required NPSH, sometimes designed NPSHR, can be considered to be the amount of pressure, in excess of the vapor pressure, required to overcome internal pump losses and so keep

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

8.5


water flowing into the pump. For a given pump, the re− quired NPSH increases as capacity increases. Each system, as a result of design and physical limitations, will produce an available NPSH sometimes desig− nated NPSHA. When the available NPSH is greater than the required NPSH, the problems of air release, vaporization, and cavitation will not arise. The required NPSH for a specific pump is available from the manufacturer, either in catalogue data or upon request. Although usually given as a single num− ber, the value varies with flow and head. For any pump, the full range of values for each impeller size and oper− ating speed is expressed as a curve (see Figure 8−4). The TAB technician must remember, however, that for satisfactory pump operation, NPSHA must always ex− ceed the NPSHR; if it does not, bubbles and pockets of vapor will form in the pump. The results will be reduc− tion in capacity, loss of efficiency, noise, vibration, and cavitation. The available NPSH in a specific system may be ex− pressed by the following equation: Equation 8-1 NPSHA  P a  Ps  V   P vp 2g 2

Where: NPSHA = Net positive suction head available − ft wg (m wg) Pa = Atmospheric pressure at elevation of installationCuse 34 ft wg (10.32 m wg) at sea level

strainer on the suction side of the pump should become clogged. 8.1.4.5

Vortex

Vortex is not a pressure, but a term used to describe whirling or spinning of the liquid in a piping system. The condition is similar to a weather cyclone or torna− do and may occur anywhere in the piping system where conditions cause or allow the vortex to be pro− duced. On the discharge or relatively positive pressure side of a pump the worst effect of a vortex is normally noise, and the configuration of the piping is usually the cause. Several elbows, always turning the same way, may produce a vortex with a low pressure center in the pipe, and noise bubbles of air can temporarily be released in the same way as a pump suction with inadequate NPSH. On the suction, or relatively negative pressure side of the pump, the same condition may also occur. Howev− er, the suction vortex problem is more commonly caused by placing the suction pipe termination too close to the surface of the system liquid, as might occur in a cooling tower pan. The low pressure produced at the pipe entrance produces a vortex or whirlpool simi− lar to that produced when a stopper is removed from a sink full of water. The result of this condition is to introduce air directly into the eye of the pump impel− ler, impairing the efficiency of the pump and produc− ing undesirable noise. This condition may occur even though calculations indicate adequate NPSH.

Ps = Pressure at pump centerlineCft wg (m. 2 wg) V = Absolute vapor pressure at 2g pumping temperatureCft wg (m wg) g = Gravity accelerationC32.2 ft/s2 (9.81 m/s2) NPSH is normally not a consideration in closed sys− tems, especially where the pump is at the bottom of a riser. It is also not ordinarily a factor in most open sys− tems unless pumping hot water, or if there is a consid− erable suction lift, or if there is considerable friction in the pump suction pipe. In unusual considerations of excessive suction line friction, there could be insuffi− cient NPSHA. Such a condition could exist because of an undersized pipe, or too many fittings, or if a valve in the suction line was throttled, or if a fine mesh

8.6

Total Head - Ft

Pump Performance Curve

Required NPSH Flow Rate - gpm

FIGURE 8-4 TYPICAL REQUIRED NPSH CURVE

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


Some vortex producing conditions can be eliminated by proper piping. In cooling tower sumps, a plate often is installed to prevent a vortex from forming. 8.2

PUMP / SYSTEM CURVE RELATIONSHIP

8.2.1

Pump Curves

Pump performance characteristics are presented in graphical curves or tabular form in the same manner as fans. However, pump curves usually are available, and performance tables are not; which is contrary to the situation with fan data. The result is a relatively quick, visual pump selection to determine the conse− quences of system pressure changes. Typical pump selection curves are illustrated in Figure 8−5. The curves presented might illustrate the ratings of a single pump of a family similar to the fan curves pre− viously discussed. In addition, the same pump may have two sets of curves. The one illustrated represents operation at 1750 rpm, while another set of conditions is available for the same pump operating at 3500 rpm (or 3450 rpm). The group or sub−family of curves on each graph is produced by different impeller sizes in the same ca− sing. Those impeller sizes shown are for standard im− peller diameters, usually in inch or half inch incre− ments in U.S. units and similar increments in metric units. Within the minimum and maximum impeller sizes indicated, any impeller diameter may be made to accomplish specific requirements simply by shaving a standard size on a lathe.

TOTAL HEAD IN FEET

The position of the pump capacity selection point is best located in or slightly to the left of the highest effi−

45% 55% 60%

60 50

7” 6 1_w”

68% 65% 60% 55%

40 30

5 1_w”

8.2.2

Closed System Curve

The system curve is simply a plot of the change in en− ergy head resulting from a fluid flow change in a fixed piping circuit. System curve construction methods dif− fer between open and closed piping circuits. From the pipe size and design flow rate, a calculated energy head pressure drop is determined. It should be particularly noted that system static height is of no im− portance in determining energy head pressure drop. This is because the static heights of the supply and re− turn legs are in balance; the energy head required to raise water to the top of the supply riser is balanced by the energy head regain as water flows down the return riser.

Example 8.1 (I−P)

45%

5”

3HP

20 10 0

The motor size must be made so that the pump will not overload at the design conditions and if possible at any curve condition along the selected impeller curve.

A design flow rate of 200 gpm (12.0 L/s) establishes 30 foot (9 m) pressure drop in a typical system. This particular point can be plotted on a foot head versus gpm pump curve as shown in Figure 8−7. What pres− sure drop would occur if the flow were changed to 163 gpm (9.8 L/s) through the piping circuit?

65% 68%

ciency area. Not only is efficiency high, but should it be a factor, NPSH is low. Just as important is the matter of quiet operation. Therefore, the selection is not nor− mally made at or near either the maximum or mini− mum impeller sizes. When the impeller diameter is chosen in the mid−range, it may be replaced in the field, if required, with either a larger or smaller size. Furthermore, the slope of the pump curve requires se− rious consideration. Too flat a curve results in large changes in flow rate for small changes in system head. Too steep a curve will often dip at the left, which will produce surging or unstable operation should the pump need to operate in this area.

2HP 1 1/2 HP

20

3 HP 1HP 1_w HP 4 40 60 80 100 120 140 160 CAPACITY IN U.S. GALLONS PER MINUTE

180

FIGURE 8-5 PUMP CURVE FOR 1750 RPM OPERATION

Solution Using Equations 2−24 and 2−26 (combined) 2 Q2 H 2  H 1 Q1

2

H 2  30 163   19.9feet 200

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

8.7


CUT-OFF FOR SELECTED IMPELLER DIAMETER PUMP HEAD-CAPACITY CURVE FOR FIXED IMPELLER SIZE MOTOR POWER (STANDARD SIZES)

TOTAL HEAD

MOTOR WILL OVERLOAD IF OPERATING POINT SHIFTS TO THE RIGHT OF THIS INTERSECTION, SELECT MOTOR B FOR NONOVERLOADING

HEAD - CAPACITY CURVE FOR DESIGN CONDITIONS AND IMPELLER DIAMETER

DESIGN HEAD

DESIGN OPERATING POINT

MOTOR B MOTOR A DESIGN FLOW FLOW

EFFICIENCY CURVES - SELECTION AT OR TO LEFT OF MAXIMUM MAINTAINS HIGH EFFICIENCY IF ACTUAL OPERATING POINT OCCURS TO RIGHT OF DESIGN OPERATING POINT

FIGURE 8-6 TYPICAL DESIGN PUMP SELECTION POINT (FROM ABBREVIATED CURVE) The same procedure carried out for a 116 gpm (7.0 L/s) flow rate would result in a 10.1 (3 m) pressure drop. These points may be plotted on a foot head versus gpm chart as shown in Figure 8−7. Connection of these three points, along with other condition combinations, de− scribes a system curve. The system curve is a statement of the change in pipe friction drop with water flow change for a fixed piping circuit. This is a most impor− tant working tool for pump application.

Equation 8-2

H2 Q2   H1 Q1

2

Where: H = HeadCft wg (m wg)

TOTAL HEAD - FEET(m)

Q = Fluid flowCgpm (L/s or m3/s)

50 (15) 40 (12) 30 (9) 20 (6) 10 (3) 0

The operation of the pump in Figure 8−7 installed in the piping circuit described by the system curve must be at the intersection of the pump curve with the system curve because of the First Law of Thermodynamics.

PUMP CURVE

POINT OF OPERATION

8.2.3

SYSTEM CURVE 50 (3)

100 (6)

150 (9)

200 (12)

250 (15)

CAPACITY - US GALLONS PER MINUTE (LITERS PER SECOND)

FIGURE 8-7 SYSTEM CURVE PLOTTED ON PUMP CURVE 8.8

Open System Curve

In plotting the system curve for an open system, the statics of the system must be analyzed in addition to the friction loss. The different static conditions are il− lustrated in Figure 8−8. A typical cooling tower application is illustrated in Figure 8−9. In this system, the pump is drawing water from the tower sump and discharging it through the condenser to the tower nozzles, at a 10 foot (3 m) high− er elevation than the sump level.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


TOTAL STATIC HEAD

TOTAL STATIC HEAD STATIC SUCTION HEAD

STATIC STATIC DISCHARGE DISCHARGE HEAD HEAD

TOTAL STATIC HEAD

STATIC SUCTION HEAD

STATIC DISCHARGE HEAD

STATIC SUCTION LIFT

STATIC SUCTION HEAD LESS THAN STATIC DISCHARGE HEAD

STATIC SUCTION LIFT PLUS STATIC DISCHARGE HEAD

STATIC SUCTION HEAD GREATER THAN STATIC DISCHARGE HEAD

FIGURE 8-8 TYPICAL OPEN SYSTEMS

This system curve cannot be applied directly to the pump curve and the intersection taken as the accurate pumping point for the open system. A false evaluation using this criteria, but without evaluating the static height of the tower, is shown in Figure 8−10. The illustration is false because the pump must also provide the necessary energy to raise water from the tower sump to the spray nozzles. In this case, the pump NOZZLES

10 FT(3 m) TOTAL STATIC HEAD

COOLING TOWER SUMP

must raise each pound of water 10 feet (3 m) in height, or it must provide 10 feet (3 m) of energy head due to the static difference in height between the water levels. The static difference of 10 feet (3 m) must be added to the piping pressure drop to provide total required head for each of the gpm points previously noted. The re− vised fluid flow versus total required head is shown in Table 8−3.

60 (18)

TOTAL HEAD - FEET(m)

Total friction loss (suction and discharge piping, con− denser, nozzles, etc.) is 30 foot (9 m) at a design flow rate of 200 gpm (12 L/s), the change in piping pressure drop for a change in water flow rates is determined and plotted to develop a system curve.

PUMP No. 1 PERFORMANCE CURVE

50 (15) 40 (12)

FALSE OPERATING POINT

30 (9) 20 (6)

SYSTEM CURVE

10 (3)

STATIC DISCHARGE HEAD STATIC SUCTION HEAD

0

75 (4.5)

150 (9.0)

225 (13.5)

300 (18.0)

375 (22.5)

CAPACITY - US GALLONS PER MINUTE (LITERS PER SECOND) CONDENSER

FIGURE 8-9 TYPICAL COOLING TOWER APPLICATION

FIGURE 8-10 SYSTEM CURVE FOR OPEN CIRCUIT FALSE OPERATING POINT

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

8.9


PUMP HEAD CAPACITY CURVE PUMP NO.2 PERFORMANCE CURVE

POINT 4

POINT 2

50 (15)

POINT 1 DESIGN HEAD

TRUE OPERATING POINT PUMP NO.2

40 (12)

DESIGN SYSTEM HEAD CURVE

PUMP NO.1 PERFORMANCE CURVE

30 (9)

SYSTEM CURVE

20 (6)

TRUE OPERATING POINT PUMP NO.1 10 (3) 10’(3m) 0 75 (4.5)

150 (9.0)

225 (13.5)

300 (18.0)

375 (22.5)

CAPACITY - US GALLONS PER MINUTE (LITERS PER SECOND)

SYSTEM AND PUMP HEAD

TOTAL HEAD - FEET(m)

60 (18)

OVERPRESSURE WITH CONSTANT SPEED PUMP POINT 3

ACTUAL SYSTEM HEAD CURVE

POINT 5

FIGURE 8-11 SYSTEM CURVE FOR OPEN CIRCUIT TRUE OPERATING POINT

50% DESIGN 100% DESIGN FLOW

FLOW

SYSTEM FLOW

FIGURE 8-12 PUMP OPERATING POINTS

The correct procedure for plotting a system curve for the circuit shown in Figure 8−9 is illustrated in Figure 8−11. 8.2.4

Pump Operating Points

In Figure 8−12, if the system is of the free flowing type without control valves, with an actual system head curve as shown, the pump will operate at Point 2, not Point 1; the pump will produce a higher flow rate than design flow rate. If the system is of the controlled flow type with two way valves on all heating or cooling coils, at design flow, the pump will operate at Point 1 and will create an over−pressure on the coils and con− trol valves equal to the head difference between Points 1 and 3. If the system flow is reduced to 50 percent of design on such a system, the over pressure will in− crease to the amount between Points 4 and 5. Pump op− eration will be as follows.

head and system head being converted into over−pres− sure, consumed by the control valves. Recognizing that over−pressure can occur in con− trolled flow systems where coils are equipped with two way control valves, the selection of pumps for these systems must include methods of limiting over−pres− sure to an economic minimum. These methods in− clude: a.

multiple pumps operating in parallel

b.

multiple pumps operating in series

NON-CONTROLLED FLOW

c.

multi−speed pumps

On a system without control valves, the pump will al− ways operate at the point of intersection of the pump head capacity curve and the system head curve.

d.

variable speed pumps

8.2.4.1

8.2.4.2

CONTROLLED FLOW

On controlled flow systems, the pump will follow its head capacity curve, the difference between pump

8.10

The actual method used on a specific hydronic system depends on the economics of that system. The effects of these methods on over−pressuring a particular sys− tem can be determined by developing the system head curve and plotting pump head capacity curves on the same graph with the system head curve.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


0 (0) 0 (0)

116 (7.0) 10 (3)

Feet (Meters) of Head Design 163 185 200 (9.8) (11.1) (12.0) 20 (6) 25 (7.5) 30 (9)

Static Head

10 (3)

10 (3)

10 (3)

10 (3)

10 (3)

10 (3)

10 (3)

Total Head

10 (3)

20 (6)

30 (9)

35 (10.5)

40 (12)

45 (13.5)

50 (15)

GPM (L/s) Piping System

215 (12.9) 35 (10.5)

230 (13.8) 40 (12)

Table 8-3 Flow vs Total Head (Cooling Tower Application) 8.2.5

Multiple Pumps

PUMP HEAD CAPACITY CURVE ONE-PUMP OPERATION

SYSTEM DESIGN HEAD

SYSTEM AND PUMP HEAD

Multiple pumps in parallel is the most common meth− od of eliminating over−pressure. Figure 8−13A de− scribes two pumps piped in parallel, while Figure 8−14 includes a system head curve as well as the head capac− ity curves for single pump and two pump operation. It is obvious from Figure 8−14 that one pump at 50 per− cent system flow will reduce the over−pressure caused by two−pump operation or one pump designed to han− dle maximum design flow and head.

Figure 8−13B illustrates two pumps piped in series with bypasses for single pump operation. Figure 8−15 indicates the use of series pumping on a hydronic sys− tem with a system head curve consisting of a large amount of system friction. For such systems, series pumping can greatly reduce the overpressure on a con− trolled flow system. Series pumping should not be used on hydronic systems with flat system head curves similar to the one shown in Figure 8−14.

For such a system, one pump operation with series connection would result in the pump running at shutoff head and producing no flow in the system.

CHECK VALVE

CHECK VALVE

BYPASSES FOR SINGLE PUMP OPERATION

A. PARALLEL PUMPING

B. SERIES PUMPING

FIGURE 8-13 MULTIPLE PUMPS

TWO-PUMP OPERATION

TWO PUMPS ONE PUMP MAXIMUM POINTS OF OPERATION

SYSTEM HEAD CURVE

INDEPENDENT HEAD

50% DESIGN FLOW

100% DESIGN FLOW

SYSTEM FLOW

FIGURE 8-14 PUMP AND SYSTEM CURVES FOR PARALLEL PUMPING

8.3

PUMP INSTALLATION CRITERIA

8.3.1

Pressure Gage Location

To eliminate the effect of pipe friction, fittings, valves, and other obstructions, the most desirable gage loca− tion for accuracy would be at the pump flanges. How− ever, this is not usually practical. Gages should be lo− cated as close to the flanges as possible as shown in Figure 8−16. To eliminate an elevation static head correction, the gages on suction and discharges should be at the same height with respect to the pump centerline. If this pre− caution is not taken, the difference in gage elevation, even though usually of small numerical value, must be accounted for in the gage differential.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

8.11


versed or if one or both gages were located below the horizontal pipe or pump centerline. Gate Valve

Unacceptable Locations

Gate Valve Check Valve

Strainer

Pump

Preferable Location on Each Side

Acceptable Location Each Side

FIGURE 8-16 GAGE LOCATION

In Figure 8−17 there is a physical difference in height of 2 feet (0.6 m). If the gage pressure, when converted, measured 50 feet (15 m) on the discharge and 30 feet (9 m) on the suction, subtraction alone would indicate a differential of 20 feet (6 m). However, with respect to the discharge gage which is two feet (0.6 m) lower in the piping, the suction gage reads two feet (0.6 m) of head too little, and at the same elevation as the dis− charge gage would read 32 feet (9.6 m). The correct differential is then 50 − 32 = 18 feet (15 − 9.6 = 5.4 m). A similar analysis would apply if the positions were re−

TWO-PUMP OPERATION

Fluid Viscosity

It should be noted that as long as the head − fluid flow curve is based on feet (meters) of head, no correction need be made for temperature or density since feet (meters) of head and gallons per minute (liters per sec− ond) account for these factors. However, density does increase the pump power requirements. The pump wa− ter power curves are developed at near maximum den− sity at approximately 85F (29C). Since density de− creases as temperature rises, pump water power will decrease, but the change usually is ignored. Viscosity can change the pump impeller head capacity curve provided the change in viscosity is greater than the change of water viscosity between 40F and 400F (4C and 204C). The effect on the curve is illustrated in Figure 8−18.

2 ft (0.6m)

Pump PUMP HEAD CAPACITY CURVES TWO PUMPS

100% DESIGN HEAD

SYSTEM AND PUMP HEAD

8.3.2

Difference in Gage Readings is Not Pump Differential

ONE-PUMP

FIGURE 8-17 RELATIVE GAGE ELEVATIONS

MAXIMUM POINTS

ONE-PUMP OPERATION

OF OPERATION

8.3.3

Installation Criteria

SYSTEM HEAD CURVE 50% DESIGN HEAD

Some of the important points for the TAB technician to consider in installing a pump are:

INDEPENDENT HEAD

100% DESIGN FLOW

a.

suction piping should be air tight and free of air traps

b.

piping should provide a smooth flow into the suction without unnecessary elbows

c.

suction pipe should be one or two sizes larger than pump inlet (eccentric reducer or reduc− ing elbow to connect inlet to piping)

d.

reduce or eliminate restrictions at pump suc− tion

SYSTEM FLOW

FIGURE 8-15 PUMP AND SYSTEM CURVES FOR SERIES PUMPING

8.12

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


e.

piping supported independently of pump cas− ing

f.

use of vertical silent check valve in pump dis− charge in multi−pump installations

g.

manual air vent in pump casing and piping

h.

pressure gages on suction and discharge at same elevation

i.

when vibration isolation is used, isolate pip− ing and pump as a system (preferable), or pro− vide pump isolators and piping flexible con− nectors

j.

recheck pump alignment after installation even if guaranteed by manufacturer

k.

lubricate prior to start up

l.

check rotation, but do not run mechanical seals dry

Total Head

Water Impeller Curve

Increased Viscosity Curve for Same Impeller

Flow

Boiler heating surface is the area of fluid backed sur− face exposed to the products of combustion, or the fire side surface. Various codes and standards define al− lowable heat transfer rates in terms of heating surface. Boiler design provides for connections to a piping sys− tem which delivers heated fluid to the place of use and returns the cooled fluid to the boiler. 8.4.2

Heat exchangers or converters are used as heat sources for many hot water heating systems. Heat exchangers may be of three general types: a) steam−to−water, b) water−to−water, or c) water−to−steam (generators). Steam−to−water heat exchangers usually take the form of shell and tube units. Steam is admitted to the shell, and water is heated as it circulates through the tubes. Steam−to−water converters are useful where an addi− tion is to be made to an existing steam system and where hot water heating is desired. They are also wide− ly used in areas where district steam is available and individual buildings are to be heated with a hot water system. High rise buildings can be zoned vertically by using steam distribution and installing converters at various levels to serve several floors, thus limiting maximum operating pressures in the zone. Water−to−water heat exchangers (generally shell and tube units) are used in high temperature water (HTW) systems to produce lower temperature water for cer− tain zones or in process water or domestic water ser− vices. Water−to−steam heat exchangers generally consist of U−tube bundle installed in a tank or pressure vessel to provide space for the release of steam. They are used in HTW systems to provide process steam where re− quired. 8.4.3

FIGURE 8-18 EFFECT OF VISCOSITY

Heat Exchangers

Water Chillers

8.4

HYDRONIC HEATING AND COOLING SOURCES

The source of cooling in a chilled water or a dual tem− perature system is a water chiller. There are three gen− eral types of water chillers: (1) reciprocating, (2) cen− trifugal, and (3) absorption. For further information, see the 2000 ASHRAE Systems and Equipment Hand− book.

8.4.1

Boilers

8.4.4

A boiler is a cast iron or steel pressure vessel heat ex− changer, designed with and for fuel burning devices and other equipment to burn fossil fuels (or use electric current) and transfer the released heat to water (in wa− ter boilers) or to water and steam (in steam boilers).

Heat Pumps

A heat pump may serve as a source for both hot water and chilled water in a dual temperature system. Heat pumps are described in the 2000 ASHRAE Systems and Equipment Handbook. Water temperatures avail− able are generally low in winter (about 90F to 130F

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

8.13


or 32C to 54C) and terminal heat transfer must be designed for operation under these conditions. In some cases, a supplementary heat source is used to raise temperature levels. 8.5

TERMINAL HEATING AND COOLING UNITS

8.5.1

General

Many types of terminal units are available for central water systems. Some are suited to only one type of sys− tem and others may be used in all types of systems. Terminal units may be classified in several ways. 8.5.1.1

Natural Convection

Natural convection units include cast iron radiators, cabinet convectors, baseboard and finned tube radi− ation are used in heating systems. 8.5.1.2

Forced Convection

Forced convection units include unit heaters, unit ven− tilators, fan coil units, induction units, air handling units, heating and cooling coils in central station units, and most process heat exchangers. Fan coil units, unit ventilators, and central station units can be used for heating, ventilating and cooling. 8.5.1.3

Radiation

Radiation units include panel systems, unit radiant panels, and certain special types of cast iron radiation. All transfer some heat by convection. Such units are generally used for heating in low temperature water (LTW) systems. However, special designs of overhead radiant surfaces, both tubular and panel, are being used in medium and high temperature water systems to take advantage of the lowered surface requirements achieved through the use of high surface temperatures. Panel cooling is applied in conjunction with control of space humidity to maintain the space dew point below the panel surface temperature. 8.5.1.4

Mixing Different Types Of Units

In any single circuit having similar loads and a single control point, the terminal units should be of similar response types. Cast iron radiation should not be installed in the same controlled circuit as baseboard or fin tube type units. Caution should be exercised when including fan operated units with natural convection units on the same pumping circuit. 8.14

8.5.2

Radiators And Convectors

Cast iron radiation and cabinet convectors have been widely used in LTW systems. Ceiling hung radiators frequently were used where floor space was not avail− able for other types of units. Convectors are used ex− tensively in areas where high output is needed and lim− ited space is available, and where linear heat distribution is not desired. Typical areas heated with radiators or convectors include corridors, entries, toi− let rooms, storage areas, work rooms, and kitchens. 8.5.3

Baseboard And Fin Tube Radiation

Baseboard and fin tube radiation permits the blanket− ing of exposed surfaces for maximum comfort. Base− board and fin tube elements are generally rated at vari− ous average water temperatures and at one or more water velocities. Velocity corrections may be applied. Many designers feel that these units are thus limited to systems designed to a 20F (11C) temperature drop. However, careful selection can result in successful ap− plication with temperature drops much higher than 20F (11C). 8.5.4

Unit Ventilators

Unit ventilators, originally developed for specific ap− plication in school classrooms, are being used today in a much wider range of applications. Unit ventilators consist of a forced convection heating or cooling unit with dampers permitting introduction of controlled amounts of outdoor air to provide a complete cycle of heating, ventilating, ventilation cooling, or mechani− cal cooling as required. Condensation may be a prob− lem during summer operation unless chilled water flow is stopped when fans are not operating. Conden− sate drains are necessary. Comparatively low supply temperature and rise may be required. 8.5.5

Fan Coil And Induction Units

Fan coil units are generally used, with or without out− door air, in dual temperature water systems. The same coil is often used for both heating and cooling. Individ− ual control is usually achieved by the use of valves, or by using intermittent or multi−speed fan operation. Hot water ratings are usually based on flow rates or tem− perature drops at various entering water and air tem− peratures. Temperature drops of 40F to 60F (22C to 33C) frequently are used. Induction units are simi− lar to fan coil units except that air circulation is pro− vided by a central air system which handles part of the load, instead of a blower in each cabinet. 8.5.6

Unit Heaters

Unit heaters are available in several types: a) horizon− tal propeller fan, b) downblow or c) cabinet. They are

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


used where high output in a small space is required, and where no cooling is to be added. Cabinet units are frequently applied in corridors and at entrances to

blanket doors which are frequently opened. Normally, unit heaters do not provide ventilation air.

HVAC SYSTEMS Testing, Adjusting & Balancing â&#x20AC;˘ Third Edition

8.15


THIS PAGE INTENTIONALLY LEFT BLANK

8.16

HVAC SYSTEMS Testing, Adjusting & Balancing â&#x20AC;¢ Third Edition


CHAPTER 9

HYDRONIC SYSTEMS


CHAPTER 9 9.1

HYDRONIC SYSTEMS

9.1.1

General

HYDRONIC SYSTEMS 9.1.2.2

Medium Temperature Water Systems (MTW)

A hydronic or all water system is one in which hot or chilled water is used to convey heat to or from a condi− tioned space or process through piping connecting a boiler, water heater, or chiller with suitable terminal heat transfer units located at the space or process.

Medium temperature systems are hot water heating systems that operate at temperatures of 350F (177C) or less, with pressures not exceeding 150 psi (1035 kPa). The usual design supply temperature is approxi− mately 250F to 325F (121C to 163C), with a usu− al pressure rating for boilers and equipment of 150 psi (1035 kPa).

All water systems may be classified by:

9.1.2.3

a.

temperature

b.

generation of flow

c.

pressurization

d.

piping arrangement

e.

pumping arrangement

In terms of flow generation, hot water heating systems are of two types: (1) the gravity system, in which cir− culation of the water is due to the difference in weight between the supply and return water columns of any circuit or system; and (2) the forced system in which a pump, usually driven by an electric motor, maintains the necessary flow. Water systems can be either once− through or recirculating systems.

High temperature systems are hot water heating sys− tems that operate at temperatures over 350F (177C) and pressures of about 300 psi (2070 kPa). The maxi− mum design supply water temperature is 400F to 450F (204C to 232C), with a pressure rating for boilers and equipment of about 300 psi (2070 kPa). It is necessary that the pressure temperature rating of each component be checked against the design charac− teristics of the particular system. 9.1.2.4

Temperature Classifications

Water systems may be classified according to operat− ing temperature as follows. 9.1.2.1

Low Temperature Water Systems (LTW)

Low temperature systems are hot water heating sys− tems that operate within the pressure and temperature limits of the ASME boiler construction code for low pressure heating boilers. The maximum allowable working pressure for low pressure heating boilers is 160 psi (1104 kPa) with a maximum temperature limi− tation of 250F (121C). The usual maximum work− ing pressure for boilers for LTW systems is 30 psi (207 kPa), although boilers specifically designed, tested, and stamped for higher pressures may frequently be used with working pressures to 160 psi (1104 kPa). Steam−to−water or water−to−water heat exchangers also are used.

Chilled Water System (CW)

A chilled water system operates with a normal design supply water temperature of 40F to 55F (4C to 13C) and a pressure range of 125 psi (62 kPa). Anti− freeze or brine solutions may be used for systems (usu− ally process applications) that require temperatures below 40F (4C). Well water systems may use supply temperatures of 60F (16C) or higher. 9.1.2.5

9.1.2

High Temperature Water Systems (HTW)

Dual Temperature Water System (DTW)

A dual temperature system is a combination hot water heating and chilled water cooling system that circu− lates hot and/or chilled water to provide heating or cooling using common piping and terminal heat trans− fer apparatus. They are operated within the pressure and temperature limits of LTW systems, with winter design supply water temperatures of 100F to 150F (3C to 66C) and summer supply water temperatures of 40F to 55F (4C to 13C). 9.1.3

Piping Classifications

Generally, the most economical hydronic distribution system layout has pipe mains that are run by the short− est and most convenient route to the terminal equip− ment that has the largest flow rate requirements. Branch or secondary circuits then are connected to these mains. Hydronic distribution mains are most frequently lo− cated in corridor ceilings, above hung ceilings, along

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

9.1


a perimeter wall, in pipe trenches, in crawl spaces and in basements. System piping need not be run at a defi− nite level or pitch, but may change up or down as re− quired by architectural or structural needs. Hydronic system piping may be divided into two arbitrary size classificationsCsmall systems and large systems. 9.1.3.1

Small Systems

Piping circuits suitable for complete small systems or for terminal or branch circuits on large systems consist of: a.

series loops

b.

one−pipe systems

c.

two−pipe reverse return systems

d.

two−pipe direct return systems

9.1.3.2

Large Systems

Main distribution piping used to convey water to and from terminal units or circuits in large systems consist of: a.

two−pipe direct return systems

b.

two−pipe reverse return systems

c.

three−pipe systems

d.

four−pipe systems

9.1.3.3

Pump

Series Loop System

A series loop system is a continuous run of pipe or tube from a supply connection to a return connection. Ter− minal units are a part of the loop. Figure 9−1 shows a system of two series loops on a supply and return main (split series loop). One or more series loops may be used in a complete system. Loops may connect to mains, or all loops may run directly to and from the boiler. The water temperature drops progressively as each ra− diator transfers heat to the air, the amount of drop de− pending on radiator output and the water flow rate.

Boiler Adjusting Cock

FIGURE 9-1 A SERIES LOOP SYSTEM considered one radiator, and all units can be sized at the AWT of the loop. One floor of a small dwelling with open interior doorways is such an interconnecting space. If individual units on a loop are in separate en− closed spaces, each unit must be sized at the actual AWT at that point. A decrease in loop water flow rate increases the tem− perature drop in each unit and in the entire loop. The average water temperature shifts downward progres− sively from the first to the last unit in the series. Conse− quently, comfort may not be able to be maintained in separate spaces heated with a single series loop if wa− ter flow rate is varied. Control of output from individu− al terminal units on a series loop is impractical except by control of heated airflow. Manual dampers can be used on natural convection units; automatic fan or face and bypass damper control can be used on forced air units.

One Special Return Fitting (Upfeed) Boiler

Pump Downfeed (Two Special Fittings)

The true system operating water temperature and flow rate must be known to calculate the average water temperature (AWT) for each unit in the loop. If all ter− minal units are in series on one loop in one zone of in− terconnecting air space, the entire set of units can be 9.2

FIGURE 9-2 A ONE-PIPE SYSTEM

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


9.1.3.4

One Pipe System (Diverting Fitting)

One−pipe circuits form a single loop to and from the boiler. For each terminal unit, a supply and a return tee are installed on the same pipe main. One of the two tees is a special diverting tee which creates a pressure drop in the fluid flow to divert a portion of the flow to the unit. One (return) diverting tee usually is sufficient for upfeed (units above main) systems. Two special fit− tings (supply and return tees) are usually required for downfeed units to overcome the thermal head. Special tees are proprietary, so consult the manufacturer’s lit− erature for flow rates and pressure drop data. One−pipe circuits allow manual or automatic control of flow to individual connected heating units. On−off control rather than flow modulation control is advis− able because of the relatively low pressure drop and low diverted flow. Piping length and load imposed on a one−pipe circuit are usually small because of the lim− itations listed.

ed balancing fitting pressure drops at the same flow rate. 9.1.3.6

Combination Piping Systems

The basic piping circuits exist only to describe func− tion as one type can grade into another. A piping sys− tem can contain one or more types and thus cannot be described as a particular type. Figure 9−5 illustrates a primary circuit and two secondary pumping circuits. As pipe lengths and number of units vary, and as circuit types are combined, basic names for piping circuits be− come meaningless; flow, temperature, and head must be determined for each circuit and for the complete system.

T S

9.1.3.5

Two-Pipe Systems R

Two−pipe circuits may be direct return where the re− turn main flow direction is opposite the supply main flow and the return water from each unit takes the shortest path back to the boiler; or reverse return where the return main flow is in the same direction as the sup− ply flow. After the last unit is supplied, the return main returns all water to the boiler. The direct return system is popular because less piping is required; however, circuit balancing valves usually are required on all units and/or sub−circuits. Since water flow distance from and to the boiler is virtually the same through any unit on a reverse−return system, balancing valves are adjusted less. Operating (pumping) costs are likely to be higher with direct return piping because of the add−

Z Y X Terminal Units Pump

Boiler or Chiller

FIGURE 9-3 DIRECT RETURN TWO-PIPE SYSTEM

Pump

Terminal Units Boiler or Chiller

FIGURE 9-4 REVERSE RETURN TWO-PIPE SYSTEM

9.1.3.7

Three-Pipe Systems

The three−pipe system satisfies variations in load by providing independent sources of heating and cooling to a terminal unit in the form of constant temperature, primary and secondary chilled, and warm water. The terminal unit contains a single secondary water coil. A modulating three−way valve at the inlet of the coil admits water from either the warm water or cold water supply, as required. The water leaving the coil is carried in a common return pipe to either the secon− dary cooling or heating equipment. The usual room control for three−pipe systems is a special three−way modulating valve that modulates either the warm or the cold water in sequence, but does not mix the streams. The cold primary air is furnished at the same temperature all year. During the period between seasons, if both hot and cold secondary water is available, any unit can be op− erated within a wide capacity range from maximum

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

9.3


Control Valves

Terminal Unit 3-W ay control Valve for secondary circuit

Common Flow

Secondary Pump

Common Flow Secondary Pump B

C

D

Balance Cock

Boiler or Chiller

E

F A

Primary Pump

FIGURE 9-5 EXAMPLE OF PRIMARY AND SECONDARY PUMPING CIRCUITS

cooling to maximum heating within the limits set by the temperature of the secondary chilled or warm wa− ter. Any unit in the system can be operated through its full range of capacity without regard to the operation of any other unit in the system, recognizing the operat− ing cost penalty that will result from simultaneous heating and cooling loads. All units are selected on the basis of their peak capacity requirements.

The three−way control valves used (Figure 9−6) are a special design in which the hot port gradually moves from open to fully closed, and the cold port gradually moves from fully closed to open. The valves are constructed so that at mid−range there is an interval in which both ports are completely closed. Room control action is the same during all seasons. 9.4

9.1.3.8

Four-Pipe Systems

Four−pipe systems used for induction, fan coil, or ra− diant panel systems derive their name from the four pipes connected to each terminal unit. These pipes connect to the cold water supply, cold water return, warm water supply, and warm water return. The four− pipe system satisfies variation in cooling and heating to the room unit in the form of constant temperature primary air, secondary chilled water, and secondary warm water. The terminal unit usually is provided with two inde− pendent secondary water coils, one served by warm water, the other by cold water. The primary air is cold and remains at the same temperature year−round. Dur− ing peak cooling and heating, the four−pipe system performs in a manner similar to the two−pipe system,

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


with essentially the same operating characteristics. During the period between seasons, any unit can be op− erated at any capacity level from maximum cooling to maximum heating, if both cold water and warm water are being circulated. Any unit can be operated at or be− tween these extremes without regard to the operation of any other unit.

T UNIT THERMOSTAT

HOT WATER RETURN

R

HOT WATER SUPPLY

HOT COIL CONTROL VALVE COLD COIL

Since the primary air is supplied at a constant cool tem− perature at all times, it is sometimes feasible for fan coil or radiant panel systems to extend the interior sys− tem supply to the perimeter spaces, eliminating the need for a separate primary air system. The type of ter− minal unit and the characteristics of the interior system will be determining factors.

T

CHECK VALVE

UNIT THERMOSTAT

COLD WATER RETURN A - SEPARATE COILS

COLD WATER SUPPLY

T UNIT THERMOSTAT R HOT WATER RETURN

HOT WATER SUPPLY

COMMON SECONDARY WATER COIL

COLD WATER RETURN

2 - POSITION DIVERTING VALVE

SEQUENCE VALVE

COMMON SECONDARY WATER COIL COMMON RETURN

FIGURE 9-6 RETURN MIX SYSTEM ROOM UNIT CONTROLS

The four−pipe terminal unit is usually provided with two completely separated secondary water coils, one receiving hot water and the second receiving cold wa− ter. The coils are operated in sequence by the same thermostat. The coils are never operated simulta− neously, and the unit receives either hot water or cold water in varying amounts, or else no flow is present. This is shown in Figure 9−7A. Figure 9−7B illustrates another unit and control config− uration which sometimes is used. A single secondary water coil is provided at the unit, and three−way valves located at the inlet and leaving side of the coil admit the water from either the warm or cold water supply, as required, and divert it to the appropriate return pipe. This arrangement requires a special three−way modu− lating valve, originally developed for one form of the three−pipe system, described earlier, which controls the warm or cold water selectively and proportionally but does not mix the streams. The valve at the coil out−

COLD WATER SUPPLY

FIGURE 9-7 FOUR PIPE SYSTEM ROOM UNIT

SEQUENCE VALVE COLD WATER SUPPLY

HOT WATER SUPPLY

let is a two−position valve that opens to either the warm or cold water return. 9.1.4

Hydronic Piping Devices

9.1.4.1

Air Control And Venting

If air and other gases are not eliminated from hydronic flow circuits, they may cause air binding in the termi− nal heat transfer elements and noise in the piping cir− cuits. High points in piping systems and terminal units should be vented with manual or automatic air vents. As automatic air vents may malfunction, valves should be provided at each vent to permit service or re− placement without draining the system. If a common tank is used for an expansion tank, free air should be routed from the piping circuit into the expansion tank by a boiler dip tube or air separation device. If a dia− phragm−type tank is used, all air should be vented from the system at all high points. 9.1.4.2

Drains And Shut-Offs

All low points should be equipped with drains. Provi− sions should be made for separate shutoff and drain of individual equipment and circuits so that the entire system does not have to be drained for service of a par− ticular item.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

9.5


9.1.4.3

Balance Fittings

Balance fittings should be applied as needed to permit balancing of individual terminal units and major sub− circuits. Such fittings should be placed at the circuit re− turn when possible. 9.1.4.4

Pitch

9.1.4.6

Thermometers or thermometer wells assist the system operator and the TAB technician, and may be used for troubleshooting. Permanently installed thermometers with separable sockets should be used at all points where temperature readings are regularly needed. Thermometer wells only may be installed where read− ings will be needed during start−up and balancing. 9.1.4.7

Hydronic piping need not be pitched, but may be run level, provided that flow velocities in excess of 1.5 feet per second (0.45 m/s) are maintained. 9.1.4.5

Strainers

Strainers should be used where necessary to protect the elements of a system. Strainers at the pump suction need to be large enough to avoid cavitation. Large sep− arating chambers are available, which serve as both main air venting points and dirt strainers. Automatic control valves require protection from particles that may readily pass through the pump and its larger mesh strainer. Individual fine mesh strainers may therefore be required ahead of each control valve. If a cooling tower is used, the strainer provided in the tower basin usually is adequate.

9.6

Thermometers

Flexible Connections

Flexible connectors are installed at pumps and ma− chinery to reduce vibration into the piping circuit. However, vibrations may be transmitted through the water across flexible connections, reducing the effec− tiveness of the connectors. Flexible connectors, how− ever, do prevent damage caused by slight misalign− ment of equipment connections to the piping. 9.1.4.8

Gages

Gage cocks should be installed at points where pres− sure readings will be required. At a minimum, these shall be located immediately entering and leaving each devise to be measured. There must be no fittings or transistors between port location and device. Note that gages permanently installed in the system may de− teriorate due to vibration and pulsation, and may not be reliable when needed.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


SYMBOLS GATE VALVE FLOW OR WEIGHTED CHECK VALVE AIR CONNECTION

ADJUSTING COCK GLOBE VALVE

EXPANSION TANK

GAGE GLASS

AUTO MIXING VALVE CIRCULATING PUMP

WATER FEEDER

DRAIN

A

ALTITUDE GAGE

ZONE SUPPLIES

THERMOMETER

MANUAL AIR VENT

A

A

A BOILER

BOILER

AIR SEPARATOR

DRAIN

DRAIN

DRAIN ZONE RETURN

FIGURE 9-8 BOILER PIPING FOR A MULTIPLE-ZONE, MULTIPLE-PURPOSE HEATING SYSTEM

HVAC SYSTEMS Testing, Adjusting & Balancing â&#x20AC;¢ Third Edition

9.7


COMPLAINT POSSIBLE CAUSE

RECOMMENDED ACTION

Shaft misalignment

 Check and realign

Worn coupling

 Replace and realign

Worn pump/motor bearings

 Replace, check manufacturer’s lubrica− tion recommendation  Check and realign shafts

Improper founda− tion or installation

Pump or system noise

 Check foundation bolting or proper grout− ing.  Check possible shift− ing due to piping expan− sion/contraction  Realign shafts.

COMPLAINT POSSIBLE CAUSE

RECOMMENDED ACTION

Pump running backwards (3-phase)

 Reverse any two-motor leads.

Broken pump coupling

 Replace and realign

Improper motor speed

 Check motor nameplate wiring and voltage.

Pump (or impeller diameter) too small

 Check pump selection (impeller diameter) against specified system requirements.

Clogged strainer(s)

 Inspect and clean screen.

Inadequate or no circulation

Pipe vibration and/ or strain caused by pipe expansion/ contraction

 Inspect, alter or add hangers and expansion provision to eliminate strain on pump(s)

System not completely filled

 Check setting of PRV fill valve.  Vent terminal units and piping high points

Water velocity

 Check actual pump performance against spe− cified and reduce impel− ler diameter as required.  Check for excessive throttling by balance valves or control valves.

Balance valves or isolating valves improperly set

 Check settings and adjust as required.

Air-bound system

 Vent piping and terminal units.  Check location of expansion tank connection line relative to pump suction.  Review provision for air elimination.

Pump Operating close to or beyond end point of perfor− mance curve

 Check actual pump performance against spe− cified and reduce impel− ler diameter as required.

Table 9-1 Hydronic Trouble Analysis Guide

9.1.4.9

Pump Location

Pump location varies with the size and type of system. Figures 9−3, 9−4 and 9−8 illustrate pumps in the supply main from the boiler or chiller, while Figures 9−2 and 9−5 have the pumps in the return piping. A pump (cir− culator) in the boiler return is acceptable for small sys− tems when pump head is at 12 foot (3.6 m) head or less, the compression tank is on the boiler (or a nearby main), and the highest piping and radiation is main− tained at a static pressure greater than full pump head. These conditions apply to most residential systems. When the pump head is equal to or greater than the dif− ference between the boiler fill and relief valve dis− charge pressures, or when the highest piping or radi− ation can be at a static pressure less than the total pump head, the pump(s) must be located on the supply side of the boiler with the compression tank at the pumps inlet, as illustrated in Figure 9−8. This assures that pump cycling will not cause excessive pressure varia− tions in the boiler that will create subatmospheric pres− sure at topmost system points causing air leakage into 9.8

the system. Pump cavitation is prevented by locating a properly sized compression tank near the pump inlet. 9.1.4.10 Trouble Shooting Table 9−1 is a handy chart for the TAB technician to use when balancing a hydronic system and/or when trouble is encountered. This list may be expanded as experience dictates. 9.2

HYDRONIC SYSTEM DESIGN

9.2.1

Closed Loop Systems

9.2.1.1

General

Pipe sizes for heating and chilled water systems should be based on clean pipe pressure drop charts. This is be− cause the system is closed, eliminating most normal corrosion to the piping. The use of pressure drop charts based on age−corroded pipe leads to oversized piping and pumps, which together will result in unwarranted system flow rates in excess of design. The excess flow and pressure heads, in turn, may cause control prob−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


lems, system noise, pumping troubles, and excessive pumping power usage. All hydronic system sizing charts may be found in the Appendix, Engineering Data, Tables and Charts. 9.2.1.2

General Design Range

The general range of pipe friction loss used for design of hydronic systems occurs between 1 to 4 ft/100 ft (0.1 to 0.4 kPa/m). A value of 2.5 ft/100 ft (0.25 kPa/m) represents the mean to which many systems are desig− ned. Wider ranges may be used in specific designs, if the precautions described below are considered. 9.2.1.3

Piping Noise

Closed loop hydronic system piping generally is sized below certain arbitrary upper limits, a velocity limit of 4 fps (1.2 m/s) for 1 inch pipe and under, and a pressure drop of 4 ft/100 ft (0.4 kPa/m) for piping over 1 inch diameter. Velocities in excess of 4 fps (1.2 m/s) may be used in piping of larger size, although they normally are not used for smaller pipe sizing. This limitation is generally accepted, although it is based on relatively inconclusive experience with noise in piping. It seems apparent that water velocity noise is caused, not by wa− ter, but by free air, sharp pressure drops, turbulence, or a combination of these, which in turn cause cavitation, or flashing of the water into steam. Therefore, higher velocities may be used if proper pre− cautions are taken to eliminate air and turbulence. Be− cause piping noise can be caused by free air, hydronic systems must be equipped with sufficient air separa− tion devices to minimize entrained air in the piping cir− cuits. 9.2.1.4

9.2.1.5

Valve And Fitting Pressure Drop

Valve and fitting pressure drop usually is listed in el− bow equivalents. The elbow equivalent relates the pressure drop through a valve or fitting to an equiva− lent length of pipe. The pressure drop of one elbow is approximately the same as that of a length of straight pipe 25 times the pipe diameter. Tables may be found in the Appendix under Engineering Data, Tables and Charts.

Example 9.1 (I−P) Determine the equivalent feet of pipe for a 4 inch open gate valve at a flow velocity of 4 fps (use Tables and Figures in the Appendix).

Solution From Table A−21, at 4 fps, each equivalent elbow is equal to 10.5 feet of 4 inch pipe. From Table A−23, a gate valve is equal to 0.5 elbows. The actual equivalent pipe length (added to measured circuit length for pres− sure drop determination) will be 10.6  0.5 = 5.3 equivalent feet of 4 inch pipe.

Example 9.1 (SI) Determine the equivalent meters of pipe for a 100 mm open gate valve at a flow velocity of 1.33 m/s (use Tables and Charts in Appendix A).

Air Purge Velocities

Air will be entrained in and carried along with the wa− ter flow at velocities of 1.5 to 2 fps (0.45 to 0.6 m/s) or more in pipe sizes 2 inches and under. Minimum ve− locities of 2 fps (0.6 m/s) are therefore recommended. For pipe sizes 2 inches and over, minimum velocities corresponding to a head loss of 0.75 ft/100 ft (0.075 kPa/m) normally are used. Attention to maintenance of minimum velocities should be observed in the upper floors of higher buildings because air may come out of solution because of reduced pressures. Higher veloci− ties should be used for downfeed return mains at air separation units.

Solution From Table A−22, at 1.33 m/s, each equivalent elbow is equal to 3.2 meters of 100 mm pipe. From Table A−23, a gate valve is equal to 0.5 elbows. The actual equivalent pipe length will be 3.2  0.5 = 1.6 equiva− lent meters of 100 mm pipe. Pressure drop through pipe tees varies with flow through the branch. Pressure drops are illustrated in Figure A−7 for tees of equal inlet and outlet sizes, and for the flow patterns illustrated.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

9.9


Example 9.2 (I−P) Determine the main equivalent pipe length for a 4 inch tee (size of all openings) flowing 25% to a side branch, 75% through the tee at an entering velocity of 3 fps.

Example 9.3 (I−P) Determine the side branch equivalent length for a 100 mm tee (all openings) with the same flow ratios as in Example 9.2.

Solution Solution From Table A−23, the number of equivalent elbows as applied to the main flow in the tee is equal to 0.15 el− bows. A 4 inch elbow is equivalent to 10.2 feet of 4 inch pipe (Table A−21), so the equivalent length of 4 inch pipe is 0.15  10.2 = 1.5 feet.

From Table A−21, the side branch elbow equivalent is about 13. Thus, the side branch equivalent length would be 13  3.1 = 40.3 meters of 100 mm pipe. Most tees are sized to a reduced side branch flow, in which case the curve in Figure A−7 does not apply ex− cept through the main. The side branch pressure drop, in any case, will not exceed 2 elbow equivalents for the branch pipe size in question.

Example 9.2 (SI) Determine the main equivalent pipe length for a 100 mm tee (size of all openings) flowing 25 percent to a side branch, 75 percent through the tee at an entering velocity of 1.0 m/s.

Example 9.4 (I−P) Determine the side branch equivalent length for a 4  4  2 inch tee with 100 gpm entering the tee and 50 gpm flowing out the side.

Solution

Solution

From Table A−23, the number of equivalent elbows as applied to the main flow is equal to 0.15 elbows. A 100 mm elbow is equivalent to 3.1 meters of 100 mm pipe (Table A−22), so the equivalent length of 100 mm pipe is 0.15  3.1 = 0.47 meters.

The side branch loss is approximately that of 2 equiva− lent elbows. For 2 inch pipe, each equivalent elbow is equal to about 5 feet. Thus, the side branch equivalent length is 2  5 = 10 equivalent feet of 2 inch pipe.

Example 9.4 (SI) Example 9.3 (I−P) Determine the side branch equivalent length for a 4 inch tee (all openings) with the same flow ratios as in Example 9.2.

Determine the side branch equivalent for a 100  100  50 mm tee with 6 L/s entering the tee and 3 L/s flow− ing out the side.

Solution Solution From Table A−23, the side branch elbow equivalent is about 13. Thus, the side branch equivalent length would be 13  10.2 = 133 feet of 4 inch pipe.

The side branch loss is approximately that of 2 equiva− lent elbows. For 50 mm pipe, each equivalent elbow is equal to about 1.5 meters. Thus the side branch equivalent length is 2  1.5 = 3 equivalent meters of 50 mm pipe. 9.2.1.6

It should be noted that the actual pipe friction loss through a 4 inch side branch pipe could actually be lower because of reduced side branch flow rates. 9.10

Water Flow-Pressure Drop

The energy head required to force water through pip− ing, valves and fittings, or heat transfer elements var−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


ies as the square of the change in water flow rate, as previously stated in Chapter 8. With conditions of flow and energy head (pressure drop) known at one point, one can determine pressure drop for any other flow rate by using Equation 8−2 (from Chapter 8).

H2 Q2   H1 Q1

be used to compute pressure drop through the valve at any flow rate. If the pressure drop and flow are known, the valve selection can be made in terms of a specific Cv rating.

2

DP 

CQ

Equation 9-1 2

v

Where:

Where:

H= Head – ft water (m water) Q = Flow – gpm (L/s or m3/s)

DP  Pressuredrop  psi(kPa) Q  Flow  gpm(m3sorLs) C v  Valveconstant(dimensionless) Example 9.6 (I−P)

Example 9.5 (I−P) Determine the pressure drop at 1 gpm for a fan coil unit that has a catalog rating of a 4 foot pressure drop at a flow rate of 2 gpm.

Calculate the Cv rating for a control valve to be se− lected for a pressure drop of 5 psi at a flow rate of 20 gpm.

Solution Using Equation 9−1: Solution H 2  H 1

QQ

2

2 1

2

H 2  4 1   1ftwater 2

Determine the pressure drop at 0.06 L/s for a fan coil unit that has a catalog rating of a 1.2 meter pressure drop at a flow rate of 0.12 L/s.

Q H 2  H 1 2 Q1

CQ

2

v

Q C v    20  8.95 DP 5

A valve is selected as closely as possible to the Cv rat− ing of 9.

Example 9.5 (SI)

Solution

DP 

Calculate the Cv rating for a control valve to be se− lected for a pressure drop at 34 kPa at a flow rate of 1.25 L/s.

2

Example 9.6 (SI)

2

H 2  1.2 0.06   0.3mwg 0.12 9.2.1.7 Valve Coefficients

Solution DP 

CQ

2

v

Control valve manufacturers state their pressure drop in terms of a coefficient (Cv). The coefficient is numer− ically equal to the flow rate through the valve at which a known pressure drop is obtained. When the Cv is known, the water flow pressure drop relationship can

Q C v    1.25  0.21 DP 34 Check valves, strainers, etc., may also be rated by Cv. When so rated, pressure drop at any given flow may be determined as in Example 9.7.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

9.11


Example 9.7 (I−P ) Calculate the pressure drop in ft wg at 750 gpm for a check valve that has a Cv of 500 (a pressure drop of 1 psi = 2.3 ft).

UNIONS FOR HEAD REMOVAL

OPEN SIGHT DRAIN

THERMOMETERS

T T

Solution DP 

CQ

2 STRAINER

v

Example 9.7 (SI) Calculate the pressure drop in kPa and meters water gage at 50 L/s for a check valve that has a Cv of 12.7 (1 kPa = 0.102 m water).

CQ

DRAIN

2

DP  750   2.25psi 500 DP  2.25psi  2.3  5.18ftwg

DP 

CONDENSER

CONTROL VALVE

2

v

50   15.5kPa 12.7

FIGURE 9- 9 WATER COOLED CONDENSER CONNECTIONS FOR CITY WATER corrosion on a continuing basis. For this reason, the piping will have an increased pressure drop with time, and fouling factors must be included in the condenser design. Provisions also must be made to treat the water for bacteria as well as for scale and corrosion.

2

DP 

DP  15.5kPa  0.102  1.6mwg 9.2.2

Open Systems

9.2.2.1

Condenser Water Systems

Condenser water systems for refrigerant compressors may be classified either as cooling tower systems, or as other systems of the once−through type, such as city water, well water, or pond or lake water systems. They are all open systems, in which air is continuously in contact with the water, and require a somewhat differ− ent approach to pump selection and pipe sizing than do closed heating and cooling systems. Some heat con− versation systems rely on a split condenser heating system which includes a two section condenser. Heat from one section of the condenser is used for heating in closed circuit systems, occasionally interconnected with chilled water systems. The other section of the condenser serves as a heat reject circuit, an open sys− tem connected to a cooling tower. In selecting a pump for a condenser water system, con− sideration must be given to the static head as well as to the system friction loss in sizing the pump. Proper provision must be made to assure an adequate net posi− tive suction head at the pump inlet. In addition, continuous contact with air in an open sys− tem introduces impurities which can result in scale and 9.12

Cooling tower water is available at a temperature sev− eral degrees above the design wet bulb temperature, depending on tower performance. For city, lake, river, or well water systems, the maximum water tempera− ture occurring during the operating season must be used for equipment selection. From manufacturers’ performance data with known condenser water temperature, the required flow rate may be determined for any condensing temperature and capacity. A condensing temperature and corre− sponding flow rate may then be selected to produce the required capacity with a minimum of energy input and purchased water within the load capacity of the driver. 9.2.2.2

Once-Through Systems

Figure 9−9 shows a water cooled condenser using city, well, or river water. The return is run higher than the condenser so that the condenser is always full of water. Water flow through the condenser is modulated by a control valve in the supply line, usually actuated from condenser head pressure to maintain a constant con− densing temperature with variations in load. City wa− ter systems usually require check valves and open sight drains, as shown. When more than one condenser is used on the same circuit, individual control valves are used to avoid balance problems. Piping should be sized in accordance with the prin− ciples outlined earlier in this chapter and in Chapter 2, with velocities of 5 to 10 fps (1.5 to 3.0 m/s) for the

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


flow rates. Where city water is used, a pump is not re− quired. For well or river water, pumps may be neces− sary, in which case the procedure for pump sizing gen− erally would be that indicated below for cooling tower systems. 9.2.2.3

Cooling Tower Systems

Figure 9−10 illustrates a typical cooling tower system for a refrigerant condenser. Water flows to the pump from the tower basin and is discharged under pressure to the condenser and back to the tower. Since it is usu− ally desirable to maintain condenser water tempera− ture above a predetermined minimum, water is di− verted through a control valve to maintain minimum sump temperature. Piping from the tower sump to the pump requires some cautions. Sump levels should be above the top of the pump casing to provide positive prime. Piping pressure drop should be minimized. All piping must pitch up either to the tower or the pump suction to eliminate air pockets. Suction strainers should be equipped with inlet and outlet gages to indi− cate when cleaning is required. Vortexing in the tower basin is prevented by piping connections at the tower in accordance with manufac− turer’s specification and by limiting flow to the maxi− mum allowed by the sump design. A straight section of suction pipe five times the diameter in length con− tributes to achieving expected pump performance.

If multiple cooling towers are to be connected, the pip− ing should be designed so that the pressure loss from the tower to the pump suction is approximately equal for each tower. Large equalizing lines or a common reservoir are used to maintain the same water level in each tower. 9.3

HYDRONIC DESIGN PROCEDURES

9.3.1

Determination Of Flows

Regardless of the method used to determine the flow through each item of terminal equipment, the desired result should be in terms of gpm (L/s). In an equipment schedule or on the plans, starting from the most remote terminal and working back towards the pump, progres− sively list the cumulative flow in each of the mains and branch circuits in the entire hydronic distribution sys− tem. 9.3.2

Preliminary Pipe Sizing

For each portion of the piping circuit, a tentative pipe size is selected from flow charts, in the Appendix, us− ing a value of pipe friction loss ranging from 0.75 to 4 ft per 100 ft (0.075 to 0.4 kPa/m). Residential piping size is often based on pump prese− lection, using pipe sizing tables, which are available

COOLING TOWER

The elements of required pump head are illustrated in Figure 9−10. Since there is an equal head of water be− tween the level in the tower sump and the pump on both the suction and discharge sides, these heads can− cel each other and may be disregarded. The elements of pump head are: static head from tower sump to the tower header; friction loss in suction and discharge piping; pressure loss in condenser; control valves; and strainer and tower nozzles, if used. These elements added in feet (meters) of water determine the required pump total dynamic head.

FLOAT VALVE

ROOF 3-W AY DIVERTING VALVE

MAKE-UP WATER

ALTERNATE BYPASS

T

Normally, piping is sized to yield water velocities be− tween 5 and 12 fps (1.5 to 3.6 m/s). Friction factors for 15 year old pipe commonly used are in the range of 1.5 to 1.75. The pressure drops for condenser, cooling tow− er, control valves, and strainers are obtained from manufacturers’ data. If condensers are installed in parallel, only the one with the highest pressure drop is counted. Combina− tion flow measuring and balancing valves can be used to equalize pressure drops.

GAGES

THERMOMETERS CONDENSER T

CONTROLLER DRAIN

FIGURE 9-10 COOLING TOWER PIPING SYSTEM

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

9.13


from the Hydronics Institute or from pump manufac− turers. 9.3.3

summarized for several of the longest piping circuits, to determine the precise head against which the pump must operate at design flow.

Preliminary Pressure Drop 9.3.6

Using the preliminary pipe sizing indicated above, cal− culate the pressure drop through each portion of the piping. Find the total pressure drop in several of the longest circuits to determine what maximum pressure drop through the piping, including the terminals and control valves, must be available in the form of pump head. 9.3.4

Preliminary Pump Selection

The preliminary selection should be based on the abili− ty of the pump to fulfill the capacity requirements as determined. It should be selected at a point left of cen− ter on the pump curve, and should not overload the mo− tor. Because of the squared relationship between flow and head, it will generally be found that flow capacity variation between the next closest stock selection and an exact point selection will be relatively minor. 9.3.5

Final Piping Layout

An overall examination of the piping layout should be made at this point, to determine if readjustments in the sizes of piping in some areas may be necessary. It is de− sirable to have the pressure drop in a number of the principal circuits be about equal, so that excessive heads are not required to serve a small portion of the building. In determining the final system friction loss, both the first cost of the piping system and the pump, and the operating costs of the pump should be considered. In general, lower heads and larger piping become more economical when longer amortization periods are con− sidered, especially in larger systems. On the other hand, in small systems, e.g., residences, it may be most economical to select the pump first and design the pip− ing system to meet the available head. In any event, ad− justments should be made in the piping system design and in the pump selection, until such time as the opti− mum design has been achieved. When the final piping layout has been established, the friction loss for each section of the piping system can be determined by reading directly from the pressure drop charts. After the friction loss at design flow for all sections of the piping system, all fittings, terminal units, and con− trol valves have been calculated, they should then be 9.14

Final Pump Selection

After the final pressure drop calculations have been completed, a final selection of the pump is made, using the procedure of plotting a system flow and pump curve, and selecting the pump that operates closest to the actual calculated design point. 9.4

STEAM SYSTEMS

9.4.1

General

A steam system does not need to be balanced nor can it be balanced manually in the true sense that air and hydronic systems need to be tested, adjusted and ba− lanced. However, the TAB technician needs to have a working knowledge of steam systems and their rela− tionships to air and hydronic systems. A steam heating system uses the vapor phase of water to supply heat to a conditioned space or a process, con− necting a source of steam, through piping, with suit− able terminal heat transfer units located at the space or process (water heater, fan coil units, or the heating coils of an absorption refrigerating machine). Steam heating systems are referred to as vacuum, return line vacuum, vapor, low, medium, and high pressure sys− tems. They are also referred to as central or district heating systems. The temperature and heating effects of steam systems vary over wide ranges to control space temperatures or heating processes. The steam supply temperature may be controlled according to outdoor temperature in much the same manner as a hot water system. They op− erate at 2188F (1038C) or higher during winter de− sign conditions and as low as 1258F (528C) in mild weather, so that the average temperature of the radi− ation may vary in direct relation to heating demands. 9.4.2

Properties Of Steam

Steam has particular characteristics related to specific volume, temperature, pressure, and heat content. Tables in the Appendix are a condensed version of these properties. Note that the conditions indicated are for saturated steam. This means that the steam at any given temperature and pressure will begin to condense if the temperature is even slightly lowered and will be− come superheated if the temperature is slightly increa− sed. Except when ensuring dry steam at the point of use, there is small advantage to superheating in envi−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


ronmental systems because the increase in heat con− tent is relatively small for the limited amount of super− heat which might normally be provided.

1000lb.hr. 8.33lb.gal.8 8 60minhr. =M2.0 gpm condensate In a similar manner, the pounds per hour of steam re− quired by the heat exchanger may also be calculated.

Also note that there are three enthalpy figures given. The enthalpy of saturated liquid is the heat content of the water just before evaporation and the enthalpy of saturated vapor is the heat content of the gas just after evaporation. The enthalpy of evaporation is the differ− ence between the two, or the amount of heat in Btu/lb (kJ/kg) required to change from liquid to gas at satura− tion. At standard conditions of 212F (100C) and 14.696 psia (101.325 kPa), this value is 970.3 Btu/lb (2256 kJ/kg), quite a difference from 1 Btu/lb.CF (4.19 kJ/kg.C) for liquid water. By use of the ap− propriate value, the pounds of condensate to be re− turned and pumped may be determined. For example:

970, 300Btuhr.   1000lb.hr.condensate 970.3Btulb.

Similar calculations may be made in metric units. 284, 300WorJs   0.126kgscondensate 2, 256, 00Jkg 0.126 kg/s  60 s/minP=M7.56 kg/min or L/min Once the condensate is formed, it will normally cool below the condensing temperature before returning to the boiler. The reduction in temperature is referred to as sub−cooling. 9.4.3

Types Of Steam Systems

9.4.3.1

Piping Designations

A steam heating system is known as a one−pipe system when a single main serves the dual purpose of supply−

WET RETURN (BELOW W.L.)

DRY RETURN

DRY RETURN

VENT TRAP

DRIP CONNECTION

STEAM MAIN

DRY RETURN

STEAM MAIN

WET RETURN (BELOW W.L.)

A. One-Pipe System

B. Two-Pipe System (Piping Typical of Atmospheric and Vapor System, etc.)

FIGURE 9-11 BASIC PIPING CIRCUITS FOR GRAVITY FLOW OF CONDENSATE

ing steam to the heating unit and conveying conden− sate from it. Each transferring device usually has only one connection which must serve as both the supply and the return, although separate supply and return connections may be used. A steam heating system is known as a two−pipe system when each transferring device is provided with two piping connections, and when steam and condensate flow in separate mains and branches.

9.4.3.2

Steam Flow

Heating systems may also be described as upfeed or downfeed, depending on the direction of steam flow in the risers; and as a dry return or a wet return, depending or whether the condensate mains are above or below the water line of the boiler or condensate receiver.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

9.15


TRAP

DRY RETURN

DRY RETURN

PUMP

PUMP A. Two-Pipe Direct-Return System Return-line vacuum, variable vacuum, subatmospheric, condensation on return, etc.

DRIP TRAP

TRAP

STEAM MAIN

DRIP TRAP

STEAM MAIN

B. Two-Pipe Reversed-Return System

FIGURE 9-12 BASIC PIPING CIRCUITS FOR MECHANICAL RETURN SYSTEMS 9.4.3.3

Return Systems

In gravity systems, the condensate returns to the boiler solely by gravity, but the steam flows by the effect of the steam pressure. When the condensate cannot be re− turned to the boiler by the action of gravity, either traps or pumps must be employed. These systems are known as mechanical return systems, and may be either open return systems or vacuum (closed) systems. In these systems, the condensate flows to the mechanical con− densate returning device by gravity, and the gradient for steam circulation is provided by steam pressure. However, in vacuum systems, the partial vacuum pro− duced by the pump provides a part or all of the gradi− ent, depending on the operating steam pressure. Vacu− um and condensate pumps both return condensate to the boiler. 9.4.3.4

Steam Piping Systems

The functions of the piping system are the distribution of low pressure steam, the return of the condensate, and, in systems where no local air vents are provided, the removal of the air. The distribution of the steam should be rapid, uniform, and noiseless. The release of air should be at a rate equal to or greater than the in− tended steam distribution. An air bound system will not heat readily or properly. In designing the piping ar− 9.16

TRAP

VALVE

STEAM MAIN

TRAP EQUALIZER LINE AIR DISCHARGE

BOILER WATERLINE

PUMP CONTROL

VACUUM HEATING PUMP

DRIP TRAP

RETURN

FIGURE 9-13 TYPICAL TWO-PIPE VACUUM STEAM SYSTEM

Pressures

Steam heating systems may also be classified as high pressure, intermediate (medium) pressure, low pres− sure, vapor, and vacuum systems, depending on the pressure conditions under which the system is de− signed to operate. 9.4.4

rangement, it is desirable to maintain equivalent re− sistances through the supply piping to, and the return piping from, each terminal unit or heat exchanger.

The condensate collecting in steam piping and termi− nal units must be drained to prevent interference with the ready flow of steam and air. It is important that steam piping systems distribute steam not only at full design load, but also during ex− cess and partial loads. Usually, the average winter steam demand is less than half the demand at the de− sign outdoor temperature. Moreover, when rapidly warming up a system even in moderate weather, the load on the steam main and returns momentarily may exceed the maximum operating load for severe weath− er, due to the necessity of raising the temperature of metal in the system to the steam temperature and the building to design indoor temperature.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


The basic tables for checking the sizes of low pressure system steam and condensate return piping may be found in the Appendix of this manual. 9.4.5

THERMOSTATIC DISC ELEMENT

Heating Unit Piping Connections INLET

It is important that the following good practices be fol− lowed when making piping connections to steam heat− ing units.

VALVE AND ORIFICE UNION OUTLET

a.

Condensate from steam main drip traps should not be piped into heating units.

b.

If it is necessary to keep the heater in service at all times, a bypass with globe or plug valve should be installed around the automatic tem− perature control valve.

c.

A strainer should be provided on the steam supply side of a control valve.

d.

The sizing of control valves should be based on the steam load and not on the heater supply connection.

e.

Each heater or bank of heaters installed in se− ries should have a separate trap.

FIGURE 9-14 THERMOSTATIC TRAP In general, steam traps consist of, a) an inlet connec− tion that opens into a chamber or passage into which condensate and non−condensable gases flow, b) an ori− fice through which the condensate and fixed gases are discharged, c) a valve which regulates or throttles the flow through the orifice port, and d) an outlet connec− tion. OPTIONAL INLET

OUTLET INLET

f.

Return piping from heater to trap should be of the same size as the heater outlet connection.

g.

Return piping should not be run to a main which is above the discharge connection of the heater or trap, or into mains under pres− sure regulated by control valves (modulating or on−off), but rather the heater condensate trap should discharge to a pump and receiver, or lifting trap, which then discharges to the overhead main or return main under pressure.

h.

9.4.6

Steam piping and heater sections should be supported independently. Steam Traps & Strainers

Steam traps are important components of most steam heating systems. These devices enable such systems to properly distribute the heating medium, and operate to perform two different functions: (1) to hold steam in the radiation and supply piping until its latent heat has been given up, and (2) while at the same time releasing non−condensables and condensate. Steam traps are usually regarded as drainage devices which release liquids and gases from a higher to a lower pressure.

VALVE AND ORIFICE

AIR VENT

FLOAT

DRAIN

FIGURE 9-15 INVERTER BUCKET TRAP

Without a trap as the means of confining the steam to heat transfer equipment, proper distribution and heat transfer related to the load could not take place since the pressure obtained in each unit is virtually unaf− fected by the pressure in the return piping. All steam traps, except the small thermostatic type and all steam control valves, should have a strainer installed immediately before each unit to prevent pipe scale and other debris from entering and damaging or clogging it. Whenever a TAB technician encounters

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

9.17


9.4.7

THERMOSTATIC DISC ELEMENT

INLET

OUTLET

Steam Systems—Medium And High Pressure

Medium and high pressure systems are used when heat transfer must be accomplished at higher temperatures, either to satisfy process conditions or for economy in equipment and piping costs. Higher pressure steam supply sources may be used to provide thermal capac− ity or to reduce carry over effects. To satisfy actual op− erating conditions, steam generated at higher pres− sures is reduced to various lower pressure levels by one or more pressure regulating stations.

FLOAT VALVE AND ORIFICE

FIGURE 9-16 FLOAT AND THERMOSTATIC TRAP

steam heat transfer equipment that is not operating properly, usually a clogged strainer or defective trap will be the problem if the automatic control valve is operating correctly. Leaking steam traps are a common maintenance prob− lem on all steam systems. Live steam in the return lines can cause overheating by some nearby units and create live steam problems with condensate or vacuum pumps. A critical problem in environmental systems is the re− moval of condensate from coils, especially those for 100 percent outside air. Proper operation dictates even steam flow and distribution in the coil. Condensate trapped in the coil prevents even distribution and should the temperature fall low enough, freeze−up might occur. Air must be allowed to enter the atmo− spheric system coil to prevent tubing collapse during the fall of pressure during initial heating, but it must then be immediately vented to prevent holding con− densate in the coil. Further, the coil must be rapidly drained of condensate so chance freezing can be redu− ced. Even under the best operating conditions, a ?non− freeze" coil (tube−in−tube) can freeze. It is imperative that improper venting and condensate removal is not allowed to add to the possibility.

9.18

Medium and high pressure systems may be classified as follows: 1) medium pressure, 10 to 44 psig (69 to 380 kPa) and 250 to 305F (121 to 152C); at high pressure, 55 to 125 psig (475 to 860 kPa) and 305 to 350F (152 to 177C).

GATE VALVE STRAINER FLOAT AND THERMOSTATIC TRAP

STEAM MAIN

REGULATING VALVE VACUUM EQUALIZER

STRAINER

GATE VALVES

HEATING COILS

DIRT POCKET

DIRT POCKET STRAINER RETURN MAIN

GATE VALVE

FLOAT AND THERMOSTATIC TRAP

FIGURE 9-17 TYPICAL CONNECTIONS TO FINNED TUBE HEATING COILS

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


CHAPTER 10

REFRIGERATION SYSTEMS


CHAPTER 10 10.1

REFRIGERATION SYSTEMS

10.1.1

General Checklist

There is no testing and balancing work for refrigera− tion systems, but TAB technicians must be familiar with them since they affect the work that must be done. The lack of ?cooling" can affect TAB work, so the fol− lowing is a checklist of common problems, assuming that all equipment has been turned on. a.

Check to see if the outdoor air dampers have been properly set as this can place an exces− sive load on the cooling coils.

b.

With direct expansion coils, check to see the air quantity crossing the cooling coil is ade− quate, as frost could be the result of insuffi− cient air.

c.

All automatic temperature control devices should be operating satisfactorily and chilled water should be flowing at the correct tem− perature.

REFRIGERATION SYSTEMS i.

A low water temperature differential across the chiller may also indicate fouling of the tubes.

j.

A low water flow across either condenser or evaporator will produce a high water temper− ature differential.

k.

A malfunctioning safety device such as a flow switch will prevent machine start up even with water flow.

l.

A malfunctioning operating device such as a refrigerant solenoid will prevent proper op− eration and perhaps shut down the machine.

m. Improper operation of centrifugal machine suction damper controller may cause surg− ing of the machine. These and the many other conditions which may occur do not necessarily cause immediate loss of cooling. However, if any are apparent during TAB work, nui− sance calls may be avoided by bringing the conditions to the attention of the responsible parties. 10.1.2

d.

e.

A frosted DX coil can be the result of insuffi− cient system refrigerant and a reduced evapo− rator temperature caused by the compressor trying to meet the load. A cooling tower cell fan shut down or high ambient wet bulb may cause high condenser pressures. The machine may operate, but the condensing temperature could rise above the maximum design. This condition is referred to as high head.

f.

A compressor shutdown may be the result of operation of a safety device such as high con− densing pressure (high head) cutout or low evaporator temperature (freeze protection or low temperature) cutoff.

g.

A water valve closed to condenser and/or chilled water heat exchanger will shut the machine down on a safety switch.

h.

A high water temperature differential across the condenser, even though not sufficient to shut down the machine, may indicate accu− mulation of fouling solids on the water side of the tubes which would require cleaning by rodding or acid.

Refrigeration Cycle

There are four basic components of the compression refrigeration cycle: the pump (compressor), the heat rejector (condenser), the metering device (capillary tube, thermal expansion valve, float valve), and the heat absorber (evaporator, chiller, cooler, direct ex− pansion coil). Figure 10−1 partially defines these com− ponents in terms of function in the cycle, and shows common names used for them. The refrigerant flow from the pump through the heat rejector to the port of the metering device is called the high side of the refrigeration circuit. From the meter− ing device through the heat absorber to the pump is called the low side of the circuit. These two terms refer to both pressure and temperature. The nature of the re− frigerants, which will be discussed further in later paragraphs, produces this effect. Beginning at the evaporator, the liquid refrigerant is boiled off at reduced pressure into a gas as it absorbs heat from the heat exchanger, and water, air or other environmental fluids are cooled by the heat removal. The gaseous refrigerant moves through the suction pipe or line to the compressor, which has reduced the inlet pressure to cause fluid flow. The compressor pumps the gas to the condenser through the hot gas (pipe) line. In the pump, the gas is compressed to a smaller volume, and heat (heat of compression) is add−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

10.1


Low Side

High Side

Liquid Line Solenoid Valve Sight Glass Expansion Valve (metering device)

Bulb Evaporator (Heat Absorber)

Hot Gas Piping

Suction Piping

Condenser (Heat Rejector)

Compressor (Pump) Low Side

Liquid Receiver

High Side

FIGURE 10-1 REFRIGERANT CYCLE

ed to the low pressure, low temperature gas to produce the high pressure, high temperature conditions found in the high side.

In the condenser heat exchanger, water from a cooling source such as a tower, or air moving across the con− denser coil, converts the high pressure hot gas to a high pressure, warm liquid. This warm liquid moves through the liquid line to the metering device, which allows the proper amount of liquid to flow into the evaporator, where the cycle starts over. Without the metering device, excess liquid would flow, causing loss of evaporator control and likely allowing liquid to enter the compressor, where damage could occur, since the compressor is only designed to pump gas. 10.2

10.2

REFRIGERATION TERMS AND COMPONENTS

10.2.1

Evaporator

The heat exchanger in which the medium being cooled, usually air or water, gives up heat to the refrig− erant through the exchanger transfer surface. The liq− uid refrigerant boils into a gas in the process of heat ab− sorption. Part of the evaporator contains liquid refrigerant, part contains a liquid−gas mixture, and part contains all gas. The amount of each will be deter− mined by the load and the control provided by the me− tering device.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


10.2.2

Thermal Expansion Valve

The metering device or flow control that regulates the amount of liquid refrigerant which is allowed to enter the evaporator. Flow of refrigerant is automatically regulated by the valve reaction to the pressure varia− tions in the sensing bulb being transmitted through the capillary tube to the thermal valve bellows. Since liquid flow to the compressor is undesirable and can be damaging, the thermal expansion valve is usu− ally adjusted to produce approximately 10F (5.6C) superheat in the gas leaving the evaporator to assure a dry condition of this gas entering the compressor. Slugging liquid into the compressor may cause dama− ge. The expansion valve connection to the evaporator may be a single pipe or multiple small pipes to individ− ual circuits. 10.2.3

Capillary Tube

The capillary tube is a metering device made from a thin tube approximately 2 to 20 feet (0.6 to 6 m) long and from 0.025 to 0.090 inches (0.6 to 2.3 mm) in di− ameter which feeds liquid directly to the evaporator. Usually limited to systems of 1 ton or less, it performs most of the functions of a thermal expansion valve when properly sized. 10.2.4

Suction Piping

Suction piping is the piping that returns gaseous refrig− erant to the compressor. This may be the most critical piping in the system design. Sizes must be large enough to maintain minimum friction to prevent re− duced compressor and system capacity, but must be small enough to produce adequate velocity to return oil to the compressor. Oil separates from the refriger− ant in the evaporator and must be entrained in the gas in small particles by gas flow velocity, or excess oil may collect in the piping and evaporator. 10.2.5

10.2.7

Discharge Stop Valve

The manual service valve on the discharge side or at the leaving connection of the compressor. Similar to the suction stop valve except for size and location. 10.2.8

Hot Gas Piping

The compressor discharge piping that carries the hot refrigerant gas from the compressor to the condenser. Sizing may be as critical as suction piping, as veloci− ties must be high enough to carry entrained oil. 10.2.9

Condenser

The heat exchanger in which the heat absorbed by the evaporator and some of the heat of compression introduced by the compressor are removed from the system. The gaseous refrigerant changes to a liquid, again taking advantage of the relatively large heat transfer by the change of state in the condensing pro− cess. Part of the condenser contains all gas, part con− tains a gas−liquid mixture, and part contains solid liq− uid refrigerant, in the reverse manner as the evaporator. 10.2.10 Receiver The receiver is an auxiliary storage receptacle for re− frigerant when the system is pumped down (shut down). When completely isolated by valving, this equipment provides a storage place to contain refriger− ant even when the system, including the condenser, may be opened for servicing. Receivers are optional except in systems where condenser storage capacity is inadequate, especially when the system design re− quires pump down. 10.2.11

Filter-Drier

The filter−drier is a combination device used as a strainer and moisture remover. Normally used with a three−valve bypass to allow removal of the cartridge during system operation.

Compressor 10.2.12 Liquid Solenoid Valve

The compressor is the pump that provides the pressure differential to cause fluid to flow, and in the pumping process, increases pressure of the refrigerant to the high side condition. The compressor is the separation between the low side and the high side. 10.2.6

Suction Stop Valve

The manual service valve on the inlet side of the com− pressor.

The electrically operated, automatic shut off valve in the liquid piping that closes on system shut down. It also closes off the receiver discharge when used in a pump down cycle, which prevents refrigerant migra− tion into the system. 10.2.13 Liquid Sight Glass The glass ported fitting in the liquid line used to indi− cate adequate refrigerant charge. When bubbles ap−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

10.3


pear in the glass, there is insufficient refrigerant in the system. 10.2.14 Hot Gas Bypass And Valve The piping and manual, but more often automatic, valve used to introduce compressor discharge gas di− rectly into the evaporator. This type of arrangement will maintain compressor operation at light loads by falsely loading the evaporator and compressor. 10.2.15 Relief Devices Codes require excess pressure relief devices which may be reseating relief valves, ?one shot" rupture discs, or both. Either should or must be piped to atmo− sphere. 10.3

SAFETY CONTROLS

The two most common safety controls are the high pressure cutout and the low temperature cutout. The high pressure cutout is a pressure actuated switch with its sensing connection in the compressor head. To pro− tect the compressor from pressures often caused by high condenser temperatures and pressure due to foul− ing and lack of water or air, this switch shuts the com− pressor down when the pressure setting is reached. A distinctive, increasing pitch sound is emitted from the compressor before shut down. The low pressure cutout is either a pressure or temper− ature actuated device with the sensing element in the evaporator, which will shut the system down at its con− trol setting to prevent freezing chilled water or to pre− vent coil frosting. Direct expansion equipment may not use thus device. 10.4

OPERATING CONTROLS

Many variations of operating controls are available. Some are: a.

cycling compressor with a thermostat,

b.

unloading compressor cylinders with step controllers operated from multi−stage ther− mostats,

c.

cycling fan and compressor in an air system, and

d.

reducing capacity by using compressor cylin− der unloaders actuated by refrigerant pres− sure changes.

10.4

10.5

REFRIGERANTS

Equipment manufacturers select refrigerants that will change state in the cycle at the temperatures required by the system. Consequently, the refrigerants are se− lected that will change from liquid to gas in the evapo− rator while absorbing heat, and will change back to a liquid from a gas in the condenser while rejecting heat. The ASHRAE Refrigeration Handbook lists the cur− rently used refrigerants found in HVAC work, as re− frigerant types are changing due to environmental problems. 10.6

THERMAL BULBS AND SUPERHEAT

If warmer air is passed over an evaporator coil, the re− frigerant will evaporate more quickly and the last drop of refrigerant will evaporate at a point at or before the coil outlet. This will cause the suction header to in− crease in temperature. The expansion valve bulb will then sense the increase in superheat and cause the ex− pansion valve to open further. This action will allow more refrigerant to flow through the expansion valve into the coil to overcome the higher rate of evaporation (an increase in superheat), thereby again moving the point where the last drop of refrigerant evaporated to the location just ahead of the bulb. It is in this manner that the expansion valve and its sensing bulb act to closely control the cooling load of the system by mea− suring the effects of the load at the outlet of the evapo− rator coil. Example ?C" in Figure 10−2 is an undesirable location because liquid can be trapped at the location of the bulb, giving a false temperature reading. The location of ?B" is incorrect because the bulb can never sense the refrigeration gas or liquid properly in the lower portion of the coil. Location ?A" is good from an accuracy of sensing standpoint, but liquid cannot be allowed to drain directly into the compressor or slugging could occur. The solution, therefore, is to use location?D" and to allow superheat to occur. Superheat is the temperature increase in the refrigera− tion gas after evaporation has been completed. All re− frigerant liquid evaporation should occur far enough ahead of location ?A" in Figure 10−2 to allow a temper− ature rise of 5F to 15F (2.8C to 8.3C) (or the tem− perature recommended by the manufacturer) to occur before the sensing bulb. This superheat establishes the point, where the last drop of refrigerant evaporated, deeper into the evaporator coil, thereby preventing liq− uid refrigerant from flowing into the suction line and then into the compressor. In other words, ?insurance" is being added so that all refrigerant is evaporated by

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


the time it reaches the bulb. This is also the reason that loops are used in suction line piping to collect liquid refrigerant in the event of an expansion valve malfunc− tion. To see superheat in a slightly different way, if at a certain rate of heat transfer or load condition, the

bulb was set to maintain a fully opened expansion valve, then lowering superheat would slightly close the valve to ensure that all liquid had evaporated be− fore reaching the compressor.

Evaporator Suction Header

Bulb

B. Wrong

A. Good

(Alternate direction)

or

C. Wrong

D. Good

FIGURE 10-2 LOCATIONS OF THERMAL BULBS HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

10.5


10.7

COMPRESSOR SHORT CYCLING

When the load on a coil requires an expansion valve to be half open, a compressor running fully loaded would quickly reduce the suction pressure from the evapora− tor coil and the low pressure control would stop the compressor. Since the expansion valve would still be controlled to remain at the half open point by the bulb, pressure would again build up in the suction piping, re− starting the compressor. Rapid stopping and starting of a compressor is called short cycling. This can quickly damage a compressor. Many compressors have unloading devices to prevent this short cycling (or a hot gas line bypass might be used). These devices selectively allow one or more of the cylinders of the compressor to cease operation al−

10.6

lowing the compressor to adjust to the changing load condition. To continue with the half opened expansion valve example, if the compressor has 4 cylinders, 2 of the cylinders would be unloaded or ineffective. The compressor could now run without being able to re− duce the suction pressure to a point sufficient to shut itself off. It is in this manner that the compressor is ad− justed to match the load. When the compressor is unloaded to its minimum ca− pacity during light load conditions, the compressor can once again short cycle. To prevent this, an anti− short cycle timer is used. This timer is put in the start− ing circuit of the compressor so that when the com− pressor stops, 5 minutes (or some other set time interval) is required to elapse before the compressor can again come on.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


CHAPTER 11

TAB INSTRUMENTS


CHAPTER 11 11.1

TAB INSTRUMENTS

INTRODUCTION

Instruments for the measurement of airflow, water flow, rotation, temperature and electricity are the tools of the trade for the TAB technician. Many instruments are available to accomplish TAB tasks and gather TAB data and information. For example, the instruments covered in this section are proven by experience to be reliable and accurate. In addition, new electronic mul− tipurpose instruments that are capable of providing more than one measurement, such as airflow and tem− perature, as well as other electronic types, are avail− able. 11.2

AIRFLOW MEASURING INSTRUMENTS

11.2.1

Manometer, U-Tube

cause rapid deterioration of any copper it touches in the system. c.

U−tube manometers should not be used for readings under 1.0 in. wg (250 Pa). OVER-PRESSURE TRAPS, WITH SHUT-OFF COCKS

11.2.1.1 Description The manometer (Figure 11−1) is a simple and useful means of measuring partial vacuum and pressure, both for air and hydronic systems. It is so universally used that both the inch (millimeter) of water and the inch (millimeter) of mercury have become accepted units of pressure measurements. In its simplest form, a ma− nometer consists of a U−shaped glass tube partially filled with a liquid such as tinted water, oil or mercury. The difference in height of the two fluid columns de− notes the pressure differential. U−tube manometers are made in different sizes and can be used for measuring pressure drops above 1.0 in. wg (250 Pascals) across filters, coils, fans, terminal devices, and sections of ductwork; and are not recommended for readings of less than 1.0 in. wg (250 Pa). 11.2.1.2

Recommended Uses

Air and gas (with water or oil instrument): a.

Measuring pressure drops above 1.0 in. wg (250 Pa) across filters, coils, eliminators, fans, grilles and duct sections.

b.

Measuring low manifold gas pressures.

FIGURE 11-1 U-TUBE MANOMETER EQUIPPED WITH OVER-PRESSURE TRAPS

11.2.2 11.2.1.3

Manometer, Inclined/Vertical

Limitations 11.2.2.1 Description

a.

b.

Manometer tubes should be chemically clean to be accurate and filled with the correct fluid. Use collecting safety reservoirs on each side of a mercury manometer to prevent blowing out mercury into the water system, which can

The inclined−vertical manometer (Figure 11–2) for airflow pressure readings is usually constructed from a solid transparent block of plastic. It has an inclined scale that provides accurate air pressure readings be− low 1.0 in.wg (250 Pa) and a vertical scale for reading greater pressures. Note that all air pressures are given

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.1


in ?inches of water (millimeters or pascals)". For ex− ample, ?3.0 inches of water" (75 mm of water or 750 Pa) means that the air pressure on one end of a U− shaped tube is enough to force the water 3 inches (75 mm) higher in the other leg of the tube. Instead of wa− ter, this instrument uses colored oil which is lighter than water. This means that although the scale reads in inches of water (mm), it is longer than a standard rule measurement. Whenever a manometer is used, the oil must be at normal room temperatures or the reading will not be correct. The manometer must be set level and mounted so it does not vibrate. Note the leveling screw and the magnetic clips. Some manometers have two scales C one indicating some pressure in inches (mm) of water and the other indicating velocity in feet per minute (meters per sec− ond). The manometer (or inclined draft gage) is the standard in the industry. It can be read accurately down to approximately 0.03 in. wg (7 Pa) and contains no me− chanical linkage. It is simple to adjust by setting the piston at the bottom until the meniscus of the oil is on the zero line. This instrument can be used with a Pitot tube or static probe to determine pressures or air veloc− ity in ducts across coils and fans.

11.2.2.2 Recommended Uses Use with Pitot tube or static probe. 11.2.2.3 Limitations When air velocities are extremely low, a micro−ma− nometer (hook gage) or some other more sensitive in− strument should be used for acceptable accuracy. 11.2.3

Electronic (Digital) Manometer

11.2.3.1 Description The electronic manometer (Figure 11−3) is designed to provide accurate readings at very low differential pres− sures. Some manometers measure an extremely wide range of pressures from 0.0001 to 60.00 in. wg (0.025 to 15,000 Pa). Airflow and velocity are automatically corrected for the density effect of barometric pressure and temperature. Readings can be stored and recalled with ?average" and ?total" functions. A specially de− signed grid enables the reading of face velocities at fil− ter outlets, coil face velocities and exhaust hood open− ings. Some millimeters provide additional functions such as temperature measurement.

FIGURE 11-3 ELECTRONIC/ MULTI-METER FIGURE 11-2 INCLINED-VERTICAL MANOMETER 11.2

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


11.2.3.2 Recommended Uses DUCT

a.

Use with Pitot tube or static pressure probe.

AIR FLOW SP

b.

With the velocity grid the instrument can be used for velocity measurements at HEPA fil− ter outlets, at hood openings and at coil faces.

PITOT TUBE

V TP

11.2.3.3 Limitations

Description

TP

A) PITOT TUBE CONNECTIONS FOR SUPPLY AIRSTREAM

Because the meter utilizes a time weighted average for each reading, it is often difficult to measure and identi− fy the pulsations in pressure. For this reason it may be difficult to repeat single point readings. 11.2.4

SP

DUCT AIR FLOW SP PITOT TUBE

SP V

TP

The standard Pitot tube (Figure 11−5), which is used in conjunction with a suitable manometer, provides a simple method of determining the air velocity in a duct. The Pitot tube is of double concentric tube construction, consisting of an 1_i inch (3.2 mm) O.D. inner tube which is concentrically located inside of a 5_qy inch (8.0 mm) O.D. outer tube. The outer ?static" tube has 8 equally spaced, 0.04 inch (1 mm) diameter holes around the circumference of the outer tube, lo− cated 2−1_r inches (57 mm) back from the nose or open end of the Pitot tube tip. At the base end, or tube con− nection end, the inner tube is open ended as at the head, and the outer tube has a side outlet tube connector per− pendicular to the outer tube and directly parallel with and in the same direction as the head end of the Pitot tube. Both tubes have a 90 degree radius bend in them lo− cated near the measuring end to allow the open end of the inner ?impact" tube to be positioned so that it faces directly into the airstream when the main shaft of the Pitot tube is perpendicular to the duct and the side out− let static pressure tube outlet connector is pointed in a parallel direction with airflow facing upstream. 11.2.4.1 Recommended Uses a.

Measurement of airstream ?total pressure" by connecting the inner tube outlet connector to one side of a manometer or gage (see ?TP" connections in Figure 11−4).

b.

Measurement of airstream ?static pressure" by connecting the outer tube side outlet con− nector to one side of a manometer or gage (see ?SP" connections in Figure 11−4).

TP B) PITOT TUBE CONNECTIONS IF AIRSTREAM IS EXHAUSTED FROM DUCT & TP IS POSITIVE DUCT AIR FLOW SP PITOT TUBE

SP V

TP TP C) PITOT TUBE CONNECTIONS IF AIRSTREAM IS EXHAUSTED FROM DUCT & TP IS NEGATIVE

FIGURE 11-4 PITOT TUBE CONNECTIONS

c.

Measurement of airstream velocity pressure by connecting both the inner and the outer tube connectors to opposite sides of a ma− nometer or gage (see ?Vp" connections in Figure 11−4).

d.

This instrument when used with a gage, ma− nometer or micro−manometer is a most reli− able and rugged instrument and its use is pre− ferred over any other method for the field measurement of air velocity, system total air, outdoor air, return air, fan static pressure, fan total pressure and fan outlet velocity pres− sures.

e.

The following are some of the instruments that may be used with the Pitot tube:

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.3


2 1_w" in. (63 mm)=8D

5 in. (125 mm)=16D 1_r”

(6.4 mm)

0.125 in. (3.2 mm) DIAM.

A 0.313 in.(8.0 mm)=1D A 0.938 in. (23.8 mm.) RAD

0.156 in. (4.0 mm) RAD 8 HOLES - 0.04 in (1mm) DIAM. NOSE SHALL BE FREE EQUALLY SPACED FROM NICKS AND BURRS. FREE FROM BURRS

90  1

SECTION A-A

INNER TUBING - APPROX 0.125 in. (3.2 mm) O.D.  21B & S GA STATIC PRESSURE OUTER TUBING 0.313 in. (8 mm) O.D.  APPROX. 18 B & S GA.

NOTE: Other sizes of pitot tubes when required, may be built using the same geometric proportions with the exception that the static orifices on sizes larger than standard may not exceed .04 in. (1 mm) in diameter. The minimum pitot tube stem diameter recognized under this code shall be 0.10 in. (2.5 mm). In no case shall the stem diameter exceed 1/30 of the test duct diameter.

TOTAL PRESSURE

FIGURE 11-5 PITOT TUBE

11.2.4.2



Micro−manometer (analog or digital), very low pressure differential; 0 to 6 inch (150 mm or 1500 Pa) range;



Inclined or digital manometer, moder− ate pressure differential; 0 to 10 inch (250 mm or 2500 Pa) range;



U−tube manometer, medium pressure differential; 1.0 to 100 inch (25 mm to 2500 mm or 250 Pa to 25 kPa) range;



Magnehelic gage, 0 to 0.5 inch (0 to 12 mm), 0 to 1.0 inch (0 to 25 mm), and 0 to 5 inch (0 to 125 mm) ranges. Limitations

The accuracy depends on uniformity of flow and com− pleteness of traverse. Several shapes and sizes of Pitot tubes are available for different applications. A rea− sonably large space is required adjacent to the duct penetration for maneuvering the instrument. Care must be taken to avoid pinching instrument tubing. Al− low a few seconds for duct pressures and probe tem− perature to stabilize after inserting probe. 11.4

If static pressure, velocity pressure, and total pressure are to be measured simultaneously, three draft gages are connected depending on the specific application. In any case, however, the three values measured will then fulfill the equation: TP = SP + Vp. In conducting tests, it frequently is sufficient to measure only two of these three pressures, since the third one can be ob− tained by simple addition or subtraction. Care must be taken, however, so that the signs of the pressures moni− tored are correct. If the airstream is exhausted from the duct, the static pressure is negative and the hose connections will de− pend on whether the velocity pressure is greater or smaller than the numerical value of the static pressure. If it is greater, the total pressure will be positive; if it is smaller, the total pressure will be negative. The various connections between the Pitot tube and gage are frequently made with rubber hose. Precaution must be taken so that all passages and connections are dry, clean and free of leaks, sharp bends and other ob− structions. The branching out of the rubber hose can be accomplished by the use of a T−fitting or by the use of a 2−stem nipple adapter which can be purchased as an accessory.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


11.2.5

Pressure Gage (Magnehelic)

11.2.5.1 Description A dry type diaphragm operated differential pressure gage that employs a calibrated spring loaded horse− shoe magnet lever operated from the differential pres− sure on the diaphragm. This causes rotation of a highly magnetic permeable helix that positions a pointer on the pressure scale. The Magnehelic pressure gage is operated by magnetic field linkage only, which makes it extremely sensitive and accurate. However, the construction of the gage makes it resistant to shock and vibration. The helix rotates on anti−shock mounted sapphire bearings. A zero calibration screw is located on the plastic cover. Common ranges are: 0 to 0.5 in. wg (125 Pa); 0 to 1.0 in. wg (250 Pa); and 0 to 5.0 in. wg (1250 Pa). There are approximately 30 available pressure ranges in this instrument.

b.

Should not be mounted on a vibrating sur− face.

c.

Should be held in same position as ?zeroed".

d.

Some should be used in the vertical position only.

11.2.6

Anemometer, Rotating Vane

11.2.6.1 Description The basic propeller or rotating vane anemometer (Fig− ure 11−7) consists of a lightweight, wind−driven wheel connected through a gear train to a set of recording dials that read the linear feet of air passing through the wheel in a measured length of time. The instrument is made in various sizes; 3 inch (75 mm), 4 inch (100 mm), and 6 inch (150 mm) sizes being the most com− mon. At low velocities, the friction drag of the mechanism is considerable. In order to compensate for this, a gear train that overspeeds is commonly used. For this rea− son, the correction is often additive at the lower range and subtractive at the upper range, with the least correction in the middle of the 200 to 2000 fpm (1 to 10 m/s) range. Most older instruments are not sensitive enough for use below 200 fpm (1 m/s). Newer instru− ments can read velocities as low as 30 fpm (0.15 m/s).

FIGURE 11-6 MAGNEHELIC GAGE

11.2.5.2

Recommended Uses

a.

Use with Pitot tube or static pressure probe as outlined under subsection 11.2.4.

b.

Use with specially constructed induction unit primary air total pressure measuring tip for primary air distribution balancing on high pressure induction systems and heat of light systems.

11.2.5.3 a.

The instrument reads in feet (m), and so a timing in− strument must be used to determine velocity. Readings are usually timed for one minute, in which case the anemometer reading (when corrected according to a calibration curve) will give the result in feet per minute or meters per minute (divide by 60 for m/s). For mod− erate velocities, it may be satisfactory to use a one−half minute timed interval, repeated as a check. A stop watch should be used to measure the timed interval, al− though a wristwatch with a sweep second hand may give satisfactory results for rough field checks. In the case of coils or filters, an uneven airflow is fre− quently found because of entrance or exit conditions. This variation is taken into account by moving the in− strument in a fixed pattern to cover the entire amount of time over all parts of the area being measured so that the varying velocities can be averaged. 11.2.6.2 a.

Measurement of supply air, return air, and ex− haust air quantities at registers and grilles.

b.

Measurement of air quantities at the faces of maximum return air dampers or openings, to− tal air across the filter or coil face areas, etc.

Limitations Readings should be made in midrange of scale.

Recommended Uses

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.5


c.

d.

If the opening is covered with a grille, the in− strument should touch the grille face but should not be pushed in between the bars. For a free opening without a grille, the anemome− ter should be held in the plane of the entrance edges of the opening. The anemometer must always be held in such a manner that the air− flow through the instrument is in the same di− rection as was used for calibration (usually from the back toward the dial face). The manufacturer’s recommendations must be followed very carefully when using this in− strument. A quality stopwatch shall be used to time the readings.

d.

Not very accurate on coils without using spe− cific correction factors calculated for each coil. See Appendix A.

11.2.7

Electronic Rotating Vane Anemometer

11.2.7.1

Description

A battery operated, direct digital or analog readout anemometer with interchangeable remote rotating vane heads. The digital readout of the velocity is auto− matically averaged for a fixed time period depending on the measured velocity and the type of instrument. Analog instruments are direct readout with a choice of velocity scales,Figure 11−8. 11.2.7.2

Recommended Uses

For measuring airflow velocities at grilles, coils, lami− nar flow cabinets and other terminal devices. 11.2.7.3

Limitations

a.

Battery operated.

b.

Total inlet area of rotating vane head must be in measured air flow.

FIGURE 11-7 ROTATING VANE ANEMOMETER

11.2.6.3

Limitations FIGURE 11-8 ELECTRONIC ANALOG ROTATING VANE ANEMOMETER

a.

Each reading from this instrument must be corrected by its calibration chart.

b.

The air terminal manufacturer’s specified ?k factor" (effective area) for this instrument must be used in computing air quantities.

11.2.8

Anemometer, Deflection Vane

11.2.8.1

Description

Total inlet area of instrument must be in mea− sured airflow.

The deflecting vane anemometer , Figure 11−9, oper− ates by having pressure exerted on a vane that causes

c.

11.6

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


a pointer to indicate the measured value. It is not de− pendent on air denisty because of the sensing of pres− sure differential to indicate velocities. The instrument is provided and always use with a dual−hose connec− tion between the meter and the probes, except as noted below.

11.2.8.3

Limitations

Instruments should not be used in extremely hot, cold, or contaminated air. 11.2.9 11.2.9.1

Thermal Anemometer Description

The operation of a thermal type anemometer (Figure 11−10) depends on the fact that the resistance of a heated wire will change with its temperature. The probe of this instrument is provided with a special type of wire element which is supplied with current from batteries contained in the instrument case. As air flows over the element in the probe the temperature of the element is changed from that which exists in still air, and the resistance change is indicated as a velocity on the indicating scale of the instrument. FIGURE 11-9 DEFLECTING VANE ANEMOMETER SET A deflecting vane anemometer set meets the needs of TAB work as most major air distribution device manufacturers have set up arear factors based on its use. The set consists of the meter, measuring probes, range selectors, and connecting hoses. Meters are scaled through the following velocity ranges: 0–300 fpm (0–1.5 m/s); 0–1250 fpm (0–6.25 m/s); 0–2500 fpm (0–12.5 m/s); 0–5000 fpm (0–25.0 m/s); 0–10,000 fpm (0–50 m/s). Three velocity probes are providedNCNthe lo−flow probe, the diffuser probe, and the Pitot tube. The lo− flow probe is used in conjunction with the 0–300 fpm (0–1.5 m/s) scale for measuring terminal air velocities in rooms or open spaces, and to measure face veloci− ties at ventilating hoods, spray booths, fume hoods, and the like. The lo−flow probe is directly mounted to the meter without the use of hoses. The Pitot tube is used to measure airstream velocities in ducts. 11.2.8.2 a.

b.

Recommended Uses This instrument may be used for measure− ments of air velocity through both supply and return air terminals using the proper jet and the proper air terminal ?k" factor (effective area) for the airflow calculation. Some instruments may also be used for mea− suring some lower velocities wher the instru− ment case itself is placed in the airstream.

FIGURE 11-10 11.2.9.2

THERMAL ANEMOMETER

Recommended Uses

a.

Used to measure very low air velocities such as a filter face velocity, room velocity and the velocity of hood openings.

b.

Can be used for velocity traverse in ducts to determine total airflow.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.7


11.2.9.3

Limitations

a.

The probe that is used with this instrument is quite directional and must be located at the proper point on the diffuser or grille as indi− cated by the manufacturer.

b.

Probes subject to fouling by dust and corro− sive air.

c.

Should not be used in flammable or explosive atmosphere.

d.

Corrections must be made for the tempera− ture of air being measured.

11.2.10

Flow Measuring Hood

11.2.10.1

Description FIGURE 11-1 1 FLOW MEASURING HOOD

The flow measuring hood ,Figure 11−11, is a device that covers the terminal air outlet device to facilitate taking air velocity or air flow. The conical or pyramid shaped hood can be used to collect all of the air dis− charged from an air terminal and guide it over flow measuring instrumentation. A velocity measuring grid and calibrated manometer in the hood will read the air− flow in cfm (L/s). The balancing hood should be tailored for the particu− lar job. The large end of the hood should be sized to fit over the complete diffuser and should have a gasket around the perimeter to prevent leakage.

11.2.11

Smoke Devices

WARNING! Before using any smoke devices, the TAB Technician must warn all people within the area so that they are aware of the use. 11.2.11.1 Description These devices generally are used for the study of air− flow and for the detection of leaks.

11.2.10.2 Recommended Uses To measure air outlet devices direction in cfm (L/s). Some digital instruments have memory, averaging, and printing capabilities.

When testing for leaks, sufficient smoke should be used to fill a volume 15 to 20 times larger than the duct or enclosure volume to be tested.

11.2.10.3 Limitations a.

b.

11.8

Smoke bombs come in various sizes with different lengths of burning time from which highly visible, non−toxic smoke readily mixes with air simplifying the observation of flow patterns.

Flow measuring hoods should not be used where the discharge velocities of the air out− lets exceed manufacturer recommendations. The hood redirects the normal pattern of air discharge which creates a slight, artificially imposed, pressure drop in the ductwork branch which can be corrected by using manufacturers backpressure compensation.

Smoke sticks and candles are convenient in that they come in different sizes and they provide an indicating stream of smoke. Some are like the puff from a ciga− rette and others smoke continuously for a few minutes to a maximum of 10 minutes. Smoke guns are valuable in tracing air currents, deter− mining the direction and velocity of airflow, and the

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


general behavior of either warm or cold air in condi− tioned rooms.

studies, hoods, filters, etc. Air motion rates below 10 fpm (0.05 m/s) can be measured with a stopwatch and distance determina− tions.

11.2.11.2 Recommended Uses a.

b.

For determining the direction and observing the velocity and pattern of airflow in room

INSTRUMENT

RECOMMENDED USES

Discharge patterns from exhaust systems, driers, hoods and stacks can be made.

LIMITATIONS

U-TUBE MANOMETER

Measuring pressure of air and gas above 1.0 in.wg (250 Pa) Measuring low manifold gas pressures

Manometer should be clean and used with correct fluid. Should not be used for readings under one inch of differential pressure.

VERTICAL INCLINED MANOMETER

Measuring pressure of air and gas above 0.02 in.wg (5 Pa) Normally used with pitot tube or static probe for determination of static, total, and velocity pressures in duct systems.

Field calibration and leveling is required before each use. For extremely low pressures, a micro manometer or some other sensitive instrument should be used for maximum accuracy.

MICROMANOMETER (ELECTRONIC)

Measuring very low pressures or velocities. Used for calibration of other instrumentation.

Because some instruments utilize a time weighted average for each reading, it is difficult to measure pressures with pulsations.

PITOT TUBE

Used with manometer for determination of total, static and velocity pressures.

Accuracy depends on uniformity of flow and completeness of duct traverse. Pitot tube and tubing must be dry, clean and free of leaks and sharp bends or obstructions.

PRESSURE GAGE (MAGNEHELIC)

Used with static probes for determination of static pressure or static pressure differential.

Readings should be made in midrange of scale. Should be “zeroed” and held in same position. Should be checked against known pressure source with each use.

ANEMOMETER ROTATING VANE (MECHANICAL AND ELECTRONIC)

Measurement of velocities at air terminals, air inlets, and filter or coil banks.

Total inlet area of rotating vane must be in measured airflow. Correction factors may apply, refer to manufacturer data.

ANEMOMETER DEFLECTING VANE

Measurement of velocities at air terminals and air inlets.

Instruments should not be used in extreme temperature or contaminated conditions.

ANEMOMETER THERMAL

Measurement of low velocities such as room air currents and airflow at hoods, troffers, and other low velocity apparatus.

Care should be taken for proper use of instrument probe. Probes are subject to fouling by dust and corrosive air. Should not be used in flammable or explosive atmosphere. Temperature corrections may apply.

FLOW MEASURING HOOD

Measurement of air distribution devices directly in CFM (L/s)

Flow measuring hoods should not be used where the discharge velocities of the terminal devices are excessive. Flow measuring hoods redirect the normal pattern of air diffusion which creates a slight, artificially imposed, pressure drop in the duct branch. Capture hood used should provide a uniform velocity profile at sensing grid or device.

Table 11-1 Airflow Measuring Instruments 11.3

PRESSURE GAGE, CALIBRATED

11.3.1

Description

The calibrated ?test gage" (Figure 11−12) shall be of a minimum?GRADE A" quality, have a Bourdon tube assembly made of stainless steel, alloy steel, monel or

bronze, and a non−reflecting white face with black let− ter graduations conforming to ANSI Specification U.S.A.S. B40−1. Test gages are usually 3−1_w inch to 6 inch (90 mm to 150 mm) diameter with bottom or back connections. Many dials are available with pressure, vacuum or compound ranges.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.9


11.3.1.1

Recommended Uses

Dial gages are used primarily for checking pump pres− sures; coil, chiller, and condenser pressure drops; and pressure drops across orifice plates, valves, and other flow calibrated devices. 11.3.1.2

Limitations

Some precautions in the use of Bourdon tube gages are: a.

Pressure ranges should be selected so the pressures to be measured fall in the middle two−thirds of the scale range.

b.

The gage should not be exposed to pressures greater than the maximum dial reading. Simi− larly, a compound gage should be used where exposed to vacuum.

c.

d.

11.3.2 11.3.2.1

Reduce or eliminate pressure pulsations by installing a needle valve between the gage and the system equipment or piping. Under extreme pulsating conditions install a pulsa− tion dampener or snubber (available from gage manufacturers). In using a gage, apply pressure slowly by gradually opening the gage cock or valve, to avoid severe strain and possible loss of accu− racy that sudden opening of the gage cock or valve can cause. Likewise, when removing pressure, slowly close the gage cock or valve, to avoid a sudden release of pressure.

pressure drop across a piece of equipment, a balancing device, or a flow measuring device. Normally, this re− quires two pressure measurements, one on the high pressure side and one on the low pressure side. The dif− ferential pressure, or pressure drop, is then the differ− ence between the two pressure readings.

A differential pressure gage is a dual inlet, ?Grade A" dual Bourdon tube pressure gage with a single indicat− ing pointer on the dial face which indicates the pres− sure differential existing between the two measured pressures. It can be calibrated in psi, inches wg or inch− es mercury (Pa, kPa, mm wg or mm Hg). The Differen− tial Pressure Gage will automatically read the differ− ence between two pressures.

Pressure Gage, Differential Description

In practically all cases of flow measurement, it will be necessary to measure a pressure differential, that is, a

11.10

FIGURE 11-12 CALIBRATED PRESSURE GAGES

Using a single gage, the gage is alternately valved to the high pressure side and the low pressure side to de− termine the pressure differential. Such an arrangement eliminates any problem concerning gage elevations, and virtually eliminates errors due to gage calibration.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


FUNCTION

RANGE

MINIMUM ACCURACY

CALIBRATION

TEMPERATURE MEASURING INSTRUMENT (CONTACT)

Minimum Range 0 to 240 F (-20  to 120C)

± 1% of full scale

12 months

HYDRONIC PRESSURE MEASURING INSTRUMENTS

0 to 30 SPI (0 to 200 kPa) 0 to 60 PSA (0to 400 kPa) 0 to 200 PSI 30 in. Hg to 30 PSI (-760 mm Hg to 200 kPa) 30 in. Hg to 60 PSI (-760 mm Hg to 400kPa)

± 1% of full scale ± 1% of full scale ± 1% of full scale ± 1% of full scale ± 1% of full scale ± 1% of full scale ± 1% of full scale ± 1% of full scale ± 1% of full scale

12 months 12 months 12 months 12 months 12 months 12 months 12 months 12 months 12 months

HYDRONIC DIFFERENTIAL PRESSURE INSTRUMENT

Minimum Range 0 to 36 in. wg (0 to 9 kPa)

± 1% of full scale

12 months *

* If used, mechanical/electronic instrument requires compliance with calibration dates noted.

Table 11-2 Instruments for Hydronic Balancing

INSTRUMENT PRESSURE GAGE (CALIBRATED)

RECOMMENDED USES

LIMITATIONS

Static pressure measurements of system equipment and/or piping.

Pressure gages should be selected so the pressures to be measured fall in the middle two-thirds of the scale range. Gage should not be exposed to pressures greater than or less than dial range.

PRESSURE GAGE (DIFFERENTIAL) FLOW MEASURING DEVICES

Pressures should be applied slowly to prevent severe strain and possible loss of accuracy of gage. Same as pressure gage.

Differential pressure measurements of system equipment and/or piping. Used to obtain highly accurate measurement of volume flow rates in fluid systems.

Must be used in accordance with recommendations of equipment manufacturer.

Table 11-3 Hydronic Measuring Instruments

Figure 11−13 illustrates the application of one type of gage modification that uses a single standard gage and eliminates the need for subtraction to determine differ− ential. The gage glass is calibrated to ft wg (kPa) at its outer periphery. During operation, the gage glass is left loose so it can be rotated. To measure a pressure differ− ential, the high pressure is applied to the gage by oper− ating the valve to the high pressure side, and the gage glass is then rotated so that its ?zero" is even with the gage pointer. Next, the high pressure valve is closed and the valve to the low pressure side is opened. The gage pointer will now indicate a pressure that is direct− ly equal to the pressure differential in ft wg (kPa). If the gage is of large diameter, such as 8 inches (200 mm) diameter, differential pressures can be read accurately to the order of 0.25 ft wg (750 Pa).

11.3.2.2

Recommended Uses

a.

This instrument when furnished in one of the lower differential pressure ranges, calibrated in inches of mercury (mm Hg), or inches of water (Pa), can be used with water hose flex− ible connectors for water distribution balanc− ing.

b.

This instrument, when furnished in one of the higher differential pressure ranges can be used in lieu of the two combination type high pressure gages.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.11


Snubber

High pressure

Low pressure (A)

Loose glass rotated to “zero” setting

Diff.pressure Ft Hd. 0

0

High pressure

Valve open

(B)

Low

High

pressure

pressure

Valve

Valve

closed

closed

Low pressure

(C)

Valve open

FIGURE 11-13 SINGLE GAGE BEING USED TO MEASURE A DIFFERENTIAL PRESSURE 11.3.2.3

Limitations

Some applications require use of a pulsation suppres− sor or needle valve. 11.4

ROTATION MEASURING INSTRUMENTS

A tachometer is an instrument used to measure the speed at which a shaft or wheel is turning. The speed is usually determined in revolutions per minute (rpm), but some have many other ranges such as rev/sec, rev/ hr, ft/sec, in./sec, cm/sec, m/min, rad/sec, and rad/min. The several types of tachometers described below vary in cost, in dependability, and in accuracy of results ob− tainable. One basic difference between the different types of tachometers is that many have digital readouts directly in revolutions per minute (rpm), while older types are primarily revolution counters that must be used with a timing device such as an accurate stop watch. 11.4.1

Tachometer, Chronometric

FIGURE 11-14 SINGLE GAGE BEING USED TO MEASURE A DIFFERENTIAL PRESSURE ter spindle will then be turning with the shaft but the instrument will not be indicating. To take a reading, the push button is pressed and then quickly released. This sets the meter hand to zero, winds the stop watch movement, and then simultaneously starts both the revolution counter and the stopwatch. After a fixed time interval, usually six seconds, the counting mech− anism is automatically uncoupled so that it no longer accumulates revolutions even though the instrument tip is still in contact with the rotating shaft. After the meter hands have stopped, the tachometer may be re− moved from the shaft and read. The meter face has two pointers and two dials, the smaller one indicating one graduation for each complete revolution of the larger pointer, and the reading will be directly in rpm (rps). Some instrument spindles must be rotating in order to be reset without damage.

11.4.1.1 Description The chronometric tachometer (Figure 11−16) com− bines a revolution counter and a stopwatch in one instrument. In using this type of tachometer, its tip is placed in contact with the rotating shaft. The tachome− 11.12

Since the timing is automatically synchronized with operation of the revolution counter, the human error that can occur when a revolution counter and separate stop watch are used, is eliminated. In general, the chro− nometric tachometer is the preferred type of instru−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


ment when the shaft end is accessible and has a coun− tersunk hole. There are new hand tachometers capable of producing instantaneous rpm measurement readings on a dial face (Eddy−current type) or, solid state instruments with digital readout. 11.4.1.2

Recommended Uses

For determining the speed of any shaft having a coun− tersunk end. 11.4.1.3

Limitations

The shaft end must be accessible and countersunk.

FIGURE 11-15 DIFFERENTIAL PRESSURE GAGE INSTRUMENT REVOLUTION COUNTER CHRONOMETRIC TACHOMETER CONTACT TACHOMETER ELECTRONIC TACHOMETER (STROBOSCOPE) OPTICAL TACHOMETER DUAL FUNCTION TACHOMETER

RECOMMENDED USES

LIMITATIONS

Contact measurement of rotating equipment speed.

Requires direct contact of rotating shaft. Must be used in conjunction with accurate timing device.

Contact measurement of rotating equipment speed.

Requires direct contact of rotating shaft.

Contact measurement of rotating and linear speeds. Non-contact measurement of rotating equipment. Non-contact measurement of rotating equipment. Contact or non-contact measurement of rotating equipment and linear speeds.

Requires direct contact of rotating shaft or device to be measured. Readings must be started at lower end of scale to avoid reading multiples (or harmonics) of the actual rpm. Must be held close to object and at correct angle. Rotating device must use reflective markings. Same as optical tachometer.

Table 11-4 Rotation Measuring Instruments

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.13


11.4.2 11.4.2.1

Contact Tachometer (Digital) Description

Contact tachometers (Figure 11−17) are available in ei− ther LCD or LED displays in multi−ranges. Some have a ?memory" button to recall the last reading as well as maximum and minimum readings. In addition, most have a measuring wheel for linear speeds. 11.4.2.2

Recommended Uses

For measuring rotational speeds and linear speeds of shafts. 11.4.2.3

FIGURE 11-16 CHRONOMETRIC TACHOMETER

Limitations

Battery operated; shaft must be accessible. 11.4.3

Optical Tachometer (Photo Tachometer)

11.4.3.1

Description

The optical tachometer or photo tachometer , Figure 11−18 uses a photocell, or eye which counts the pulses as the object rotates. Then by use of a transistorized computer circuit, it produces a direct rpm (rps) reading on the instrument dial that is either digital or analog. Several features make it adaptable for use in measur− ing fan speeds. It is completely portable and is equipped with long−life batteries for its light and pow− er source. It has good accuracy and any error can be re− duced by using more than one reflective marker on the rotating device. Its calibration can be continually checked on most jobs by directing its beam to a fluo− rescent light and comparing the indicated reading against 7200 on the rpm scale. 11.4.3.2

Recommended Uses

The optical tachometer does not have to be in contact with the rotating device. It indicates instantaneous speeds, not average speedCwhether constant or changingCthereby reading the speed as it is. It is easy to useCto read rpm, one need only place a contrasting mark on the rotating device by using chalk or reflec− tive tape. It is a good instrument to use on in−line fans and other such equipment where shaft ends are not ac− cessible. It also has good application for use on equip− ment rotating at a high rate of speed. 11.14

FIGURE 11-17 DIGITAL CONTACT TACHOMETER

11.4.3.3

Limitations

Battery operated; must be held close to object and at correct angle; mark on rotating device must reflect properly. 11.4.4

Electronic Tachometer (Stroboscope)

11.4.4.1

Description

The Stroboscope is an electronic tachometer that uses an electrically flashing light. The frequency of the

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


FIGURE 11-19 STROBOSCOPE FIGURE 11-18 DIGITAL OPTICAL TACHOMETER

flashing light is electronically controlled and adjusta− ble. When the frequency of the flashing light is ad− justed to equal the frequency of the rotating machine, the machine will appear to stand still. The Stroboscope shown in Figure 11−19 does not need to make contact with the machine being checked, but need only be pointed toward the machine so that a moving part will be illuminated by the Stroboscope light and can be viewed by the operator. The light flashes are of extremely short duration, and their fre− quency is adjustable by turning a knob on the Strobos− cope. When the frequency of the light flashes is exact− ly the same as the speed of the moving part being viewed, the part will be seen distinctly only once each cycle, and the moving part will appear to stand still. The corresponding frequency, or rpm, can be read from an analog or digital scale on the instrument. 11.4.4.2

Recommended Uses

For measurement of rotation speeds when instrument contact with the rotating equipment is not feasible. 11.4.4.3

Limitations

FIGURE 11-20 MULTI-RANGE, DUAL FUNCTION (OPTICAL/CONTACT TACHOMETER)

Care must be taken to avoid reading multiples (or har− monics) of the actual rpm (rps). Readings must be started at the lower end of the scale. HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.15


11.4.5

11.4.5.3

Dual Function Tachometer

Limitations

11.4.5.1 Description

Battery operated.

This dual−function tachometer (Figure 11−20) pro− vides both optical and contact measurements of rota− tion and linear motions. Many allow a choice of 19 ranges depending on the application. A digital display always indicates the unit of measurement to identify the operating range. The ?memory" button may be used to recall the last, maximum, minimum, and aver− age readings. Compact size and light weight make for easy one−handed operation.

11.5

TEMPERATURE FUNCTION TACHOMETER MEASURING INSTRUMENTS

11.5.1

Thermometers, Glass Tube

11.5.1.1

Description

Mercury−filled glass thermometers (Figure 11−21) have a useful temperature range of from minus 40F to over 220F (−40C to 105C). They are available in a variety of standard temperature ranges, scale gradua− tions, and lengths.

11.4.5.2 Recommended Uses For Measurement of rotation speeds by direct contact or by counting the speed of a reflective mark. FUNCTION

RANGE

RECOMMENDED ACCURACY

RECOMMENDED CALIBRATION

ROTATION MEASURING INSTRUMENT

0 to 5000 RPM

± 2%

24 months

TEMPERATURE MEASURING (IMMERSION)

-40  to -120F (-40  to 50C)

Within ½ of Scale Division

12 months*

TEMPERATURE MEASURING (IMMERSION)

0 to 220F (-20  to 105C)

Within ½ of Scale Division

12 months*

TEMPERATURE MEASURING (AIR)

-40  to 120F (-40  to 50C)

Within ½ of Scale Division

12 months*

TEMPERATURE MEASURING (AIR)

0E to 220E F (-20E to 105E C)

Within ½ of Scale Division

12 months*

ELECTRICAL MEASURING INSTRUMENTS

0 to 600 VAC 0 to 100 Amperes 0 to 30 VDC

3% of Full Scale 3% of Full Scale 3% of Full Scale

12 months* 12 months* 12 months*

*If used, mechanical/electronic instrumentation requires compliance with calibration dates noted.

Table 11-5 Instrumentation for Air & Hydronic Balancing

11.16

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


FUNCTION AIR PRESSURE MEASURING INSTRUMENTATION

RANGE

MINIMUM ACCURACY

CALIBRATION

0 to 0.5 in. wg

± - 0.01 in. wg (2.5 Pa)

12 months *

(0 to 125 Pa)

± - 0.02 in. wg (5 Pa)

12 months *

0 to 1 in. wg

± - 0.20 in. wg (50 Pa)

12 months *

0 to 5 in. wg

± - 0.5

12 months *

in. wg (125 Pa)

(0 to 1250 Pa) 0 to 18 in. wg (0 to 4500 Pa) PITOT TUBE PITOT TUBE

18 in. (450 mm) 36 in. (900 mm)

N/A N/A

N/A N/A

AIR VELOCITY MEASURING INSTRUMENT

Minimum Range: 100 to 3000 FPM (0.5 to 15 M/S)

± 10% when used in accordance with Mfg. recommendations

12 months

HUMIDITY MEASURING INSTRUMENT

10-90% RH

2% RH, Range: 10-90% RH

12 months *

AIR VOLUME MEASURING INSTRUMENT (DIRECT READING)

Minimum Range: 0 to 1400 CFM (0 to 700 l/s)

± 5% when used in accordance with Mfg. recommendations

12 months

* If used, mechanical/electronic instrument requires compliance with calibration dates noted.

Table 11-6 Instruments for Air Balancing 11.5.1.2 a.

b.

11.5.1.3 a.

b.

Recommended Uses The complete stem immersion calibrated thermometer, as the name implies, must be used with the stem completely immersed in the fluid in which the temperature is to be measured. If complete immersion of the ther− mometer stem is not possible or practical, then a correction must be made for the amount of emergent liquid column. The thermometers calibrated for partial stem immersion are more commonly used. They are used in conjunction with thermometer test wells designed to receive them. No emer− gent stem correction is required for the partial stem immersion type. Limitations Radiation effectsCwhen the temperatures of the surrounding surfaces are substantially different from the measured fluid, there is considerable radiation effect upon the ther− mometer reading, if left unshielded or other− wise unprotected from these radiation ef− fects. Proper shielding or aspiration of the thermometer bulb and stem can minimize these radiation effects. Thermometer test wellsCare used to house the test thermometer at the desired location

and permits removal and insertion of a ther− mometer without requiring removal or loss of the fluid in the system. c.

Nuclear work and many clean rooms prohibit the use of instruments containing mercury

11.5.2

Dial Thermometers

11.5.2.1

Description

Dial thermometers are of two general types: stem type (Figure 11−22) and flexible capillary type. They are constructed with various size dial heads, 1−3_r" to 5" (45 mm to 125 mm), with stainless steel encapsulated temperature sensing element. Hermetically sealed, they are rust, dust and leak proof and are actuated by sensitive bimetallic helix coils. Some can be field cali− brated. Sensing elements range in length from 2−1_w" to 24" (60 mm to 600 mm) and are available in many temperature ranges, with and without thermometer wells. The advantage of dial thermometers is that they are more rugged and more easily read than glass−stem thermometers, and they are fairly inexpensive. Small dial thermometers of this type usually use a bimetallic temperature sensing element in the stem. Temperature changes cause a change in the bend or twist of the ele− ment, and this movement is transmitted to the pointer by a mechanical linkage.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.17


11.5.2.2

Recommended Uses

Useable for checking both air and water temperature in ducts and pipe thermometer wells. 11.5.2.3

Limitations

Time lag is relatively long. 11.5.3

Thermocouple Thermometers

11.5.3.1

Description

Digital thermocouple thermometers (Figure 11−23) uses a thermocouple as a sensing device and a milli− voltmeter (or potentiometer) with a scale calibrated for reading temperatures directly.

FIGURE 11-21 GLASS TUBE THERMOMETERS

The flexible capillary type dial thermometer has a rather large temperature sensing bulb which is con− nected to the instrument with a capillary tube. The in− strument contains a Bourdon tube, the same as in pres− sure gages. The temperature sensing system, consisting of the bulb, capillary tube, and Bourdon tube, is charged with either liquid or a gas. Tempera− ture changes at the bulb cause the contained liquid or gas to expand or contract, resulting in changes in the pressure exerted within the Bourdon tube. This causes the pointer to move over a graduated scale as in a pres− sure gage, except that the thermometer dial is graduat− ed in degrees. The advantage of this type thermometer is that it can be used to read the temperature in a remote location.

In using a dial thermometer, the stem or bulb must be immersed a sufficient distance to allow this part of the thermometer to reach the temperature being measu− red. Dial thermometers have a relatively long lag time, so enough time must be allowed for the thermometer to reach the temperature and the pointer to come to rest. 11.18

Electronic type thermometers have an instrument case containing items such as batteries, various switches, knobs to adjust variable resistances, and a sensitive meter. Thermocouple temperature sensing elements are remote from the instrument case, and connected to it by means of wire or cables. Electronic type ther− mometers shown in Figure 11−24 have advantages of remote−reading, good precision, and flexibility as to temperature range. Additionally, some electronic type thermometers have multiple connection points on the instrument case, and a selector switch, enabling the use of a number of temperature sensors which can be placed in different locations, and read one at a time by use of the selector switch. 11.5.3.2 a.

Recommended Uses In balancing water circuits thermallyCwhen balancing by flow measurement is not practi− cal.

FIGURE 11-22 DIAL THERMOMETER

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


FIGURE 11-24 THERMISTOR THERMOMETER

FIGURE 11-23 THERMOCOUPLE

b.

11.5.3.3

For evaluation of certain types of boilers, fur− naces, ovens, etc. Limitations

In piping applications, it should be remembered that the surface temperature of the conduit is not equal to the fluid temperature and that a relative comparison is more reliable than an absolute reliance on readings at a single circuit or terminal unit. 11.5.4 11.5.4.1

FIGURE 11-25 INFRARED DIGITAL THERMOMETER

Electronic Thermometers Description

There are many types of rugged, light weight, battery powered digital electronic thermometers that have precision accuracy with interchangeable probes and/ or sensors. Types include: resistance temperature de−

tectors (RTD), thermistors, thermocouples, and diode sensors, with either liquid crystal or LED displays. Re− sponse time and ease of use will vary from model to model, and type to type. Resistance type electronic thermometers are shown in Figure 11−23 and in Figure 11−24.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.19


11.5.5

PSYCHROMETER

11.5.5.1

Description

The sling psychrometer (Figure 11−28) consists of two mercury filled thermometers, one of which has a cloth wick or sock around its bulb. The two thermometers are mounted side by side on a frame fitted with a han− dle by which the device can be whirled with a steady motion through the surrounding air. The whirling mo− tion is periodically stopped to take readings of the wet and dry bulb thermometers (in that order) until such time as consecutive readings become steady. Due to evaporation, the wet bulb thermometer will indicate a lower temperature than the dry bulb thermometer, and the difference is known as the wet bulb depression.

FIGURE 11-26 RESISTANCE TEMPERATURE DETECTOR

The newest type of electric thermometers is the in− frared scanner as shown in Figure 11−25.

Accurate wet bulb readings require an air velocity of between 1000 to 1500 fpm (5 to 7.5 m/s) across the wick, or a correction must be made; therefore, an in− strument with an 18 inch (450 mm) radius should be whirled at a rate of two revolutions per second. Signifi− cant errors will result if the wick becomes dirty or dry, so a constant supply of distilled water should be used. Digital battery powered versions are available that blow the ambient air over the wetted wick. These in−

When using an infrared temperature scanner, be sure to calibrate the meter for the type of surface being measured. Shiny surfaces like polished metal will have a different energy reflectivity or emissivity than a rusted or painted surface. This will cause the meter readings to be ?off set" unless the emissivity setpoint matches the material surface conditions being scanned. 11.5.4.2

Recommended Uses

Remote probe electronic thermometers may be used for checking air or liquid temperatures either im− mersed in the fluid steam or from surfaces. Infrared scanners are excellent for uninsulated overhead pip− ing, steam traps, and very hot or cold surfaces that are difficult to access. 11.5.4.3

Limitations

Resistance type have longer response times than ther− mocouple type. Infrared scanners ?see" a larger area at a distance which may introduce errors when the item being scanned is small.

11.20

FIGURE 11-27 ELECTRONIC THERMOMETER

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


struments are accurate and they can be placed into con− fined areas where there is insufficient room to whirl a sling psychrometer. 11.5.5.2

11.5.5.3 a.

Accurate wet bulb readings require an air ve− locity of between 1000 to 1500 fpm (5 to 7.5 m/s) across the wick, or a correction must be made.

b.

Significant errors will result if the wick be− comes dirty or dry.

c.

For an 18 inch (450 mm) radius unit, the in− strument should be whirled at a rate of two revolutions per second.

Recommended Uses

The sling psychrometer can be used in determining the psychrometric properties of the conditioned spaces, return air, outdoor air, mixed air and conditioned sup− ply air. The readings taken from the sling psychrome− ter can be spotted on a standard psychrometric chart from which all other psychrometric properties of the air so measured can be determined. INSTRUMENT GLASS TUBE THERMOMETERS

RECOMMENDED USES Measurement of temperatures of air and fluids

Limitations

LIMITATIONS Ambient conditions may impact measurement of fluid temperature. Glass tube thermometers require immersion in fluid or adequate test wells. Some applications prohibit use of instruments containing mercury within the work area.

DIAL THERMOMETERS

Measurement of temperatures of air and fluids.

Ambient conditions may impact measurement of fluid temperature. Stem or bulb must be immersed a sufficient distance in fluid to record accurate measurement. Time lag of measurement is relatively long.

THERMOCOUPLE THERMOMETERS

Measurement of surface temperatures of pipes and ducts.

ELECTRONICTHERMOM- Measurement of temperatures of air and ETERS fluids. Measurement of surface temperatures of pipes and ducts. PSYCHROMETERS

Measurement of dry and wet bulb air temperatures.

Surface temperatures of piping and duct may not equal fluid temperature within due to thermal conductivity of material. Use instrument within recommended range. Use thermal probes in accordance with recommendations of manufacturer. Accurate wet bulb measurements require an air velocity between 1000 and 1500 fpm (5 to 7.5 m/s) across the wick, or a correction must be made. Dirty or dry wicks will result in significant error.

ELECTRONIC THERMO-HYGRO METER

Measurement of dry and wet bulb air temperatures and direct reading of relative humidity.

Accuracy of measurement above 90% R.H. is decreased due to swelling of the sensing element.

INFRARED

Measurement of surface temperatures of distant objects

Condition and meter must be adjusted for the finish of surface being measured

Table 11-7 Temperature Measuring Instruments 11.5.6

Electronic Thermohygrometers

11.5.6.1

Description

Unlike the psychrometer, the thermohygrometer (Fig− ure 11−30) does not utilize the cooling effect of the wet bulb to determine the moisture content in the air. A thin film capacitance sensor is used as a sensing element in many instruments. As the moisture content and tem− perature change, the resistance in the sensor changes proportionally. Read out is normally in percent rela− tive humidity. Because the instruments do not rely

upon evaporation for measurement, the need for air− flow across the wetted wick or sock is eliminated. The sensing element needs only to be held in the sampled air. Typical measuring rate is 10% to 98% RH, 32F to 140F (0C to 60C). 11.5.6.2

Recommended Uses

The thermohygrometer can be used to determine the psychrometric properties of air in much the same way as the sling psychrometer. The reading can be spotted on a standard psychrometric chart from which all other

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.21


psychrometric properties of the air so measured can be determined. It can be used for measuring and monitoring of areas sensitive to change in relative humidity such as clean rooms, hospitals, museums and paper storage. Contin− uous monitoring of conditions in areas sensitive to hu− midity is possible with a greater accuracy and ease of measurement. 11.5.6.3

Limitations

At relative humidities above 90%, the accuracy of the sensor is decreased due to swelling of the sensing ele− ment.

FIGURE 11-28

SLING PSYCHROMETER

11.6

ELECTRICAL MEASURING INSTRUMENTS

11.6.1

Volt-Ammeter

11.6.1.1

Description

The testing, adjusting, and balancing of mechanical systems requires the measurement of voltages and electrical currents as a routine matter. The clamp−on− type volt−ammeter with digital readout (Figure 11−31) is one of the types used for taking field electrical mea− surements. The clamp−on type volt−ammeter shown has trigger operated, clamp−on transformer jaws which permit current readings without interrupting electrical service. Most meters have several scale ranges in amperes and volts. Two voltage test leads are furnished which may be quick−connected into the bot− tom of the volt−ammeter. 11.6.1.2

Safety & Use

When using the volt−ammeter, the proper range must be selected. When in doubt, begin with the highest range for both voltage and amperage scales. Before using, be aware of the following safety precau− tions: First C be careful not to contact an open electrical cir− cuit. Hands should never be put into the electrical box− es. Do not attempt to pry wires over into position. Do not force the instrument jaws into position. These pre− cautions reduce the risk of causing a short circuit which could injure both equipment and personnel. FIGURE 11-29 DIGITAL PSYCHROMETER

11.22

Second C when taking amperage readings do not at− tach the instrument and then start the motor. Position the instrument and read it after the motor is running at full speed. The inrush current required to start a motor

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


is from three to five times higher than the load rated full nameplate current. Therefore, starting the motor with the instrument attached could damage the instru− ment. Readings may be taken at the motor leads or from the load terminals of the starter. To determine the amper− ages of single phase motors, place the clamp about one wire. When involved with three phase current, take readings on each of three wires and average the results. To measure voltage with portable test instruments, set the meter to the most suitable range, and connect the test lead probes firmly against the terminals or other surfaces of the line under test, and read the meter, mak− ing certain to read the correct scale if the meter has more than one scale. When reading single phase volt− age the leads should be applied to the two load termi− nals. The resulting single reading is the voltage of the current being applied to the motor. When reading three phase current it is necessary to ap− ply the voltmeter terminals to Pole No. 1 and Pole No. 2; then to Pole No. 2 and Pole No. 3; and finally to Pole No. 1 and Pole No. 3. This will result in three readings, each of which will likely be a little different, but which should be close to each other. If the average voltage delivered to the motor varies by more than a few volts from the nameplate rating of the motor, several things can occur. A rise in voltage may damage the motor and will cause a drop in the amper− age reading. A drop in the voltage will cause a rise in the amperage and can cause the overload protectors on the starter to ?kick out." In either case, it is advisable to promptly report high or low voltage situations. 11.6.1.3

FIGURE 11-30

THERMOHYGROMETER

digital communication technology to access and ad− just setpoints or ?variables" of a system. Digital controllers utilize programmed algorithms to perform desired functions related to a sequence of op− eration. The algorithm(s) may call for the proportional changes, such as modulation of a variable air volume

Limitations

a.

The proper range must be selected. When in doubt begin with the highest range for both voltage and amperage scales.

b.

Depending on the conditions at the point of measurement, and the size of the volt−amme− ter, access for measurements may be restricti− ve. Caution is required, particularly when taking measurements within starters.

11.7

COMMUNICATION DEVICES

11.7.1

Introduction

The advent of microprocessor based controls has created the need for individuals working within the TAB industry to become familiar with computer and

FIGURE 11-31 CLAMP-ON VOLT AMMETER

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.23


(PDA). To speed system trouble shooting, many brands of electronic wall thermostats also include this communication jack or data port. This is extremely helpful when balancing VAV boxes serving a specific room, since the VAV box minimum and maximum or flow rates, heating and cooling temperature setpoints, and thermostat calibrations can be checked from the same point. Unfortunately, until all automation control system manufacturers standardize their communication pro− tocols, you will need either a specific handheld device for a given system, or a copy of the manufacturer’s software installed on your own laptop computer. This software may be easy to obtain since most automation system installers realize that giving you this ability re− duces their onsite setup workload. Figure 11−32 shows a TAB technician adjusting VAV box setpoints with a laptop computer ?plugged" into the data port on a modern electronic wall thermostat.

FIGURE 11-32 ACCESSING AUTOMATION SYSTEM WITH LAPTOP COMPUTER

(VAV) terminal unit, or may perform Boolean func− tions such as starting and stopping a fan. The setpoints or ?variables" of an algorithm may be predetermined and programmed. It should be noted that the TAB tech− nician may establish setpoints and communicate them to the automation system installer, which normally is required to facilitate satisfactory system operation.

11.8

HYDRONIC FLOW MEASURING DEVICES

11.8.1

Metric Measurements

Although gallons per minute (gpm) is the common hy− dronic measurement value in U.S. units, many coun− tries have not accepted S.I. units in the metric system. Many organizations use litres per second (L/s), al− though cubic meters per second (m3/s) and cubic me− ters per hour (m3/h) are used. 11.8.2

Venturi Tube and Orifice Plate

Dedicated Communication Terminals

The venturi tube or orifice plate (Figures 11−33 and 11−34) is a specific, fixed area reduction in the path of fluid flow, installed to produce a flow restriction and a pressure drop. The pressure differential (the up− stream pressure minus the downstream pressure) is re− lated to the velocity of the fluid. The pressure differen− tial also is equated to the flow in gpm (L/s) but the pressure drop is not equal to velocity pressure drop. By accurate measurement of the pressure drop with a ma− nometer at flow rates from zero fluid velocity to a max− imum fluid velocity established by a maximum practi− cal pressure drop, a calibrated flow range may be established. The flow range may then be plotted on a graph which reads pressure drop versus flow rate (gpm or L/s) or the manometer scale may be graduated di− rectly in the flow rate values.

Most microprocessor based programmable field pan− els include a jack or ?port" that allows connecting a laptop computer or hand held personal data assistant

The diagrams in Figure 11−34 illustrate the difference between the venturi tube and an orifice plate. The ven− turi tube, because of the streamlining effect of the en−

11.7.2

Computer Terminals

Today’s building temperature control systems are mi− croprocessor based, using one or more programmable field panels to provide all of the sequence of control functions and setpoints for the HVAC systems. Al− though there is still a lack of industry standards for in− terconnect cabling and communication protocols, most automation system installations include one or more desktop computers located remotely, that can monitor each field panel and adjust setpoints and oper− ating hours for each HVAC system controlled. 11.7.3

11.24

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


ORIFICE DIAMETER

ORIFICE SIZE IDENTIFICATON

PRESSURE TAPPINGS FOR INSTRUMENT CONNECTIONS

AIR VENT HOLE; LOCATE AT TOP OF HORIZONTAL PIPE IF CARRYING WATER

DRAIN HOLE: LOCATE AT BOTTOM OF PIPE IF ORIFICE IS USED IN STEAM PIPE

ORIFICE PLATE

FLOW

ORIFICE

(A) ORIFICE PLATE

(B) ORIFICE PLATE INSTALLED BETWEEN SPECIAL FLOWMETER FLANGES FLOWMETER FLANGES

FIGURE 11-33 ORIFICE AS A MEASURING DEVICE trance and the recovery cone, produces a lower pres− sure loss for the same flow rate. The full venturi tube can be extremely accurate with no appreciable system pressure loss, but it must then be extremely long. Unless such accuracy is required, a modified version with a shortened entrance and re− covery cones may be employed. The modified tube generally provides adequate accuracy with acceptable system pressure losses (still less than the orifice plate for the same accuracy) for environmental systems.

valves. They are similar to ordinary balancing valves, but the manufacturer has provided pressure taps into the inlet and outlet; and has calibrated the device by setting up known flow quantities while measuring the resistance which results from the different valve posi− tions. These positions usually are graduated on the valve body (as a dial) and the handle has a pointer to indicate the reading. The manufacturer then publishes a chart or graph which illustrates the percentage open to the valve (the dial settings), the pressure drop and the resulting flow .

11.8.3

11.8.5

Annual Flow Indicator

The Annular Flow Indicator (Figure 11−35) is a flow sensing and indicating system that is an adaptation of the principle of the Pitot tube. The upstream sensing tube has a number of holes which face the flow and so are subjected to impact pressure (velocity pressure plus static pressure). The holes are spaced so as to be representative of equal annular areas of the pipe, in the manner of selecting Pitot tube traverse points. An equalizing tube arrangement within the upstream tube averages the pressures sensed at the various holes, and this pressure is transmitted to a pressure gage. The downstream tube is similar to a reversed impact tube, and senses a pressure equal to static pressure minus ve− locity pressure at this point; this pressure is also trans− mitted to a gage. The difference between the two pres− sures, when referred to appropriate calibration data, will indicate flow in gpm (L/s). A differential pressure gage is used to directly read the pressure differential. 11.8.4

Calibrated Balancing Valves

Calibrated balancing valves (Figure 11−36) perform dual duty as flow measuring devices and as balancing

Location Of Flow Devices

Flow measuring devices including the orifice, venturi, and other types described above, give accurate and re− liable readings only when fluid flow in the line is quite uniform and free of turbulence. Pipe fittings such as el− bows, valves, etc., create turbulence and non−unifor− mity of flow. Therefore, an essential rule is that flow measuring elements must be installed far enough away from elbows, valves and other sources of flow distur− bance to permit turbulence to subside and for flow to regain uniformity. This applies particularly to condi− tions upstream of the measuring element, and it also applies downstream except to a lesser extent. The manufacturers of flow measuring devices usually specify the lengths of straight pipe required upstream and downstream of the measuring element. Lengths are specified in numbers of pipe diameters, so that the actual required lengths will depend on the size of the pipe. Requirements will vary with the type of element and the types of fittings at the ends of the straight pipe runs, ranging from about 5 to 25 pipe diameters up− stream and 2 to 5 pipe diameters downstream, or as recommended by the manufacturer.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

11.25


MODIFIED TUBE ORIFICE PLATE

FULL VENTURI TUBE

PIPE PIPE TURBULENCE

THROAT VENA CONTRACTA

ENTRANCE CONE

RECOVERY CONE

VENTURI TUBE

ORIFICE PLATE

FIGURE 11-34 FLOW METER TYPES

FIGURE 11-35 ANNULAR FLOW INDICATOR

11.26

FIGURE 11-36 CALIBRATED BALANCING VALVE

HVAC SYSTEMS Testing, Adjusting & Balancing â&#x20AC;¢ Third Edition


CHAPTER 12

PRELIMINARY TAB PROCEDURES


CHAPTER 12 12.1

INITIAL PLANNING

12.1.1 Organization Since testing, adjusting and balancing (TAB) of HVAC systems can best be accomplished by following sys− tematic procedures, the entire TAB process should be thoroughly organized and planned. All activities, in− cluding the organization, procurement of required test instrumentation and the actual system balancing should be scheduled as soon as practical after the con− tract has been consummated. Building space loads often vary with each change of season and space tem− perature levels are a significant factor in TAB work. This needs to be considered when scheduling the TAB work for any project. 12.1.2 Initial Reviews Preparatory work includes the planning and schedul− ing of all TAB procedures, collecting the necessary data, reviewing the data collected, studying the sys− tems to be balanced, making schematic system lay− outs, recording the published data on the test report forms, and finally, making preliminary field checks of the HVAC equipment and systems. If the initial study of the HVAC system plans and specifications by the TAB technician indicates that the systems may not be able to be balanced properly, the HVAC system de− signer should be sent a written notification containing suggestions for changes or the addition of balancing equipment (dampers or valves) that would allow cor− rective action before starting the balancing proce− dures. Occasionally, a system cannot be balanced or made to perform in accordance with the contract docu− ments regardless of the number of balancing dampers or valves that can be installed. 12.2

CONTRACT DOCUMENTS

Secure the latest contract drawings for the HVAC sys− tems making sure that they are complete including; floor plans, sections, schedules, riser diagrams, sche− matic flow diagrams, ATC schedules and interlocks, and any other detail drawing that is normally produced by the preparing engineer or agency. Obtain a complete set of specifications including all sections that pertain to the HVAC equipment, auto− matic temperature controls, motors, air outlets, sheet metal, VAV boxes, pumps, piping, valves, and any oth− er appurtenances that will be installed on the project.

PRELIMINARY TAB PROCEDURES Check to see that you have all change orders, bulletins or any other document that could have an impact on the installed project. Obtain the latest ?as built" shop drawing for the sheet metal and piping. Obtain system leakage rates for ductwork and secure leak test data reports if field testing was in fact required and performed. 12.2.1 HVAC Equipment Performance Data Performance data, including fan and pump curves, should be obtained for all HVAC equipment. Fan per− formance can only be verified by measuring the total static pressure. External static pressure values from manufacturer tables are helpful for initial system duct designs, but these table values do not include the ef− fects from connected ductwork. Fan performance data must relate to the actual job re− quirements and include items such as inlet vanes and altitude and temperature effects. Many times data is general and is not adjusted for these conditions. Fan performance also can be affected by improper de− sign of ductwork near the fan inlet and/or discharge. This phenomenon is called ?system effect", and it can− not be measured in the field. Performance of lower pressure fans can be substantially reduced by system effect. Fan curves or prototype curves can be obtained upon request. If none of these can be obtained, limited curves may be developed from tabulated catalog data. Use caution and determine if the pressures given are internal or external to the equipment and are total pres− sures or static pressure. Pump curves can be handled in a similar manner. Take particular note of equipment substitutions that might affect air and water pressure drops of heat ex− changers, cooling towers, coils and condensers. These include changes in coil size, fin spacing, fin configura− tion, number of rows, cooling coil wet or dry ratings, or number of tubes in the coil face. Data should include air and hydronic pressure differences, the direction of air and water flow (so that proper field installation checks can be made), temperature differentials, capac− ities, operating temperatures and pressures, and limit or safety temperatures and pressures. 12.2.2 Manufacturer’s Catalogs Obtain manufacturers catalogs for all HVAC equip− ment including pumps, air moving and air terminal de−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

12.1


vices to supplement the shop drawings and submitted data where possible.

perature setpoints, operating schedules, and interlock relationships.

12.2.3 Electrical Data

Before starting any TAB work on a system controlled by an automation system, obtain copies of the written sequence of controls, and the latest printout of all equipment setpoints and operating schedules.

Look for changes of horsepower ratings as a result of equipment substitutions. Note voltage variations such as 230 volt indicated on a nameplate instead of the 208 volt specified. Note phase substitutions, especially on packaged equipment, such as single phase on the nameplate instead of three phase specified. Often these changes can be discovered from the shop draw− ings or submittals. Motor data not available from shop drawings should be obtained in the field. Motor starters, sizes, locations, and thermal overload protection ratings should be checked against horse− power, phase and voltage for substitutions. Coordinate this information with the electrical drawings and with the electrical contractor to verify that the specified electrical service is being installed to each piece of HVAC equipment. 12.2.4 Air Distribution Devices Manufacturer’s recommendations on device testing is available in most cases. Effective Areas (K factors) usually can be obtained for all air grilles, registers, and diffuser devices for the velocity measuring instrument recommended. In addition to air pattern adjustment and sound data. Manufacturer’s Data and Test Proce− dures should be obtained for all other rated or adjust− able air handling devices, such as variable air volume boxes, constant volume regulators, static pressure con− trol dampers and all other similar equipment. Air pressure drop data across louvers, filter banks, sound traps, remote coils and other devices in the air distribution system should be obtained. Note if louvers are provided with screening; filter pressure drop data is for clean, partially dirty, or dirty filters; if sound trap pressure drops are certified and can be confirmed, and if all pressure drops for substituted equipment are within design limitations. 12.2.5 Automatic Temperature Control (ATC) And Energy Management System (EMS) Diagrams With the phase out of pneumatic control systems by microprocessor based programmable controls, it is no longer possible to determine the sequence of controls or control setpoints by visually looking into a control cabinet. Computer chips in microprocessor controls now contain all of the control logic, equipment tem− 12.2

12.2.6 Maintenance, Operating And Start-Up Obtain operating and maintenance manuals for all equipment if available for review. 12.3

SYSTEM REVIEW AND ANALYSIS

After all preliminary data has been collected, a study of each HVAC system may be performed. Two basic reasons for the importance of system review and anal− ysis are (1) to isolate any discrepancies in the data or drawings that may prevent the proper balancing and performance of a system, and (2) to establish the best approach for testing and balancing. The design engi− neer, architect, and owner may want to review what procedures will be employed during the actual testing and balancing of their systems. This also offers an ex− cellent opportunity to present any discrepancies found during the system review and analysis. 12.3.1 System Components And Types Review all available plans, specifications and equip− ment data noting such things as the types and locations of the areas served, types of system, types of compo− nents used such as fans, pumps, boilers, chillers, coils, and VAV boxes. Note such things as primary and sec− ondary systems, interlocked or interconnected sys− tems, possible tie−ins to existing systems and the loca− tion of motor control centers, breakers, and other electrical equipment. Review the equipment sched− ules, the temperature control drawings and carefully note the operating sequence of control for each system. Review the most current set of plans, notes, sections, and details and check for possible additional fans or other equipment that may not be listed in the equip− ment schedules. 12.3.2 System Schematic Drawings It is recommended that drawings or a schematic layout of each HVAC duct system be prepared as shown in Figure 12−1. You can color code an extra set of the original design drawings if you do not normally gener− ate your own plans. A similar drawing should be made for all complex piping systems. All dampers, regulat−

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition


-400 CFM EA. (200 l/s) 1200 CFM (600 l/s) 800 fpm (4 m/s)

2

1

-200 CFM EA. (100 l/s)

3

4

20"  12" (500  300 mm) FD

10

6

1800 CFM (900 l/s) VD

VD

5

FD

VD

20"  12" (500  300 mm)

1200 CFM (600 l/s) 800 fpm (4 m/s)

7

8

9

1800 CFM (900 l/s) 1200 fpm (6 m/s) 18"  12" (450  300 mm)

-200 CFM EA.(100 l/s)

3rd FL. -400 CFM EA.(200 l/s) 2400 CFM (1200 l/s) 1200 CFM (600 l/s) 11 12 13 1200 fpm (6 m/s) 800 fpm (4 m/s)

-200 CFM EA.(100 l/s)

28"  12" (700  300 mm)

14

20"  12" (500  300 mm) VD PT FD

15

FD 20

16

1800 CFM (900 l/s) VD

VD PT 1200 CFM (600 l/s) 20”  12” (500  300 mm) 800 fpm (4 m/s)

17

19

18

3600 CFM (1800 l/s) 1200 fpm (6 m/s)

-200 CFM EA.(100 l/s)

2nd FL. -400 CFM EA.(200 l/s) 4800 CFM (2400 l/s) 1200 CFM (600 l/s) 21 800 fpm (4 m/s) 1425 fpm (7.2 m/s)

22

36"  13" (900  325 mm) 23

-200 CFM EA.(100 l/s) 24

30"  18" 20"  12" (750  250 mm) (500  300 mm) VD FD VD

25

26

28

29

1800 CFM (900 l/s)

VD

20"  12" 1200 CFM (600 l/s) PT(500  200 mm) 800 fpm (4 m/s)

27

-200 CFM EA.(100 l/s)

1st FL. 7200 CFM (3600 l/s) 1600 fpm (8 m/s) 30"  26" (750  650 mm)

FD

S. FAN NO.1

FILTER

COOLING COIL

PT

35"  20" (875  500 mm) OUTDOOR AIR LOUVER

VD ATC ATC

EXHAUST LOUVER