Fundamental Desgin Control Central Chilled Water Plant Ip

Fundamentals of Design and Control of Central Chilled-Water Plants Steven T. Taylor, PE

I-P A Course Book for Self-Directed or Group Learning

Includes Skill Development Exercises for PDH or LU Credits

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Fundamentals of Design and Control of Central Chilled-Water Plants Steven T. Taylor, PE

A Course Book for Self-Directed or Group Learning

Atlanta

Fundamentals of Design and Control of Central Chilled-Water Plants (I-P) A Course Book for Self-Directed or Group Learning ISBN 978-1-939200-66-2 (paperback) ISBN 978-1-939200-67-9 (PDF) SDL Course Number: 00662 © 2017 ASHRAE All rights reserved.

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ASHRAE STAFF ASHRAE Learning Institute Karen Murray Manager of Professional Development Sarah Boyle Managing Editor of Professional Development Kelly Arnold Professional Development

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Dear Student, Welcome to this ASHRAE Learning Institute (ALI) self-directed or group learning course. We look forward to working with you to help you achieve maximum results from this course. You may take this course on a self-testing basis (no continuing education credits awarded) or on an ALImonitored basis with credits (PDHs or LUs) awarded. ALI staff will provide support, and you will have access to technical experts who can answer inquiries about the course material. For questions or technical assistance, contact us at 404-636-8400 or [email protected]. Skill Development Exercises at the end of each chapter will gauge your comprehension of the course material. If you take this course for credit via the ALI online-monitoring system, please complete the exercises in the workbook then submit your answers at www.ashrae.org/sdlonline (preferred method) or email copies from each chapter to [email protected]. To log in, please enter your student ID number and the SDL number. Your student ID number can be the last five digits of your Social Security number or another unique five-digit number you create when first registering online. The SDL course number is located near the top of the copyright page of this book. Please keep copies of your completed Skill Development Exercises for your records. When you finish all exercises, you will receive a Certificate of Completion indicating 35 PDHs/LUs of continuing education credit. The ALI does not award partial credit for self-directed or group learning courses. All exercises must be completed to receive full continuing education credit. You will have two years from the date of purchase to complete each course. We hope your educational experience is satisfying and successful.

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Karen M. Murray Manager of Professional Development

Continuing Education Opportunities from ASHRAE Learning Institute ASHRAE Learning Institute (ALI) provides professional development through in-depth training that is timely, practical, and targeted to engineers in consulting practices, facility management, or supplier support with instruction on applying ASHRAE standards and employing new technologies essential for advanced building performance.

HVAC Design Essentials and Applications Training—Instructor Led at Approved Locations Expand your knowledge and understanding of the fundamentals and technical aspects to design and maintain HVAC systems. Level I covers essentials. Level II instructs on use of ASHRAE Standards 55, 62.1, 90.1, and 189.1. A companion course explains improving existing building operations. www.ashrae.org/hvactraining

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Self-Directed Learning Texts—Self Study or Texts for Group Instruction For those seeking traditional book-based instruction, ASHRAE offers Learning Texts for selfstudy or group training with instructor materials. Texts cover the basics of what a practicing engineer needs for real-world HVAC&R applications. Skill Development Exercises are included to evaluate progress. Students receive a course completion certificate designating continuing education credits. www.ashrae.org/sdl

ASHRAE Learning Institute

·

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Steven T. Taylor, PE, Fellow ASHRAE, is the founding principal of Taylor Engineering, Alameda, CA. He is a registered mechanical engineer specializing in HVAC system design, control system design, indoor air quality engineering, computerized building energy analysis, and HVAC system commissioning. Mr. Taylor graduated from Stanford University with a BS in Physics and a MS in Mechanical Engineering and has 40 years of commercial HVAC system design and construction experience. He was one of the primary authors of the HVAC sections of ASHRAE Standard 90.1, Energy Standard for Buildings Except LowRise Residential Buildings and California’s Title 24 energy standards and ventilation standards. Other ASHRAE projects and technical committees Mr. Taylor has participated in include ASHRAE Standard 62.1 on indoor air quality (chair), ASHRAE Standard 55 on thermal comfort (member), Guideline 13 on specifying DDC (chair), Guideline 16 on economizer dampers (chair), Guideline 36 on advanced control sequences (founder and member), the TC 1.4 on controls (chair), and the TC 4.3 on ventilation (chair). He is past vice-chair of the U.S. Green Building Council (USGBC) Leadership in Energy and Environmental Design (LEED) Indoor Environmental Quality Technical Advisory Group, a member of the CSU Mechanical Review Board, and a 16-year member of the International Association of Plumbing and Mechanical Officials (IAPMO) Mechanical Technical Committee administering the Uniform Mechanical Code.

Contents Preface . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ix Acknowledgments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .xii Acronyms. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xv Chapter 1: Overview . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 How to Use This Course . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 Organization of Material . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 Chapter 2: Chilled-Water Plant Loads. . . . . . . . . . . . . . . . . . . . . . . . . 3 Instructions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3 Understanding Loads and Their Impact on Design . . . . . . . . . . . 3 Determining Peak Loads . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9 Determining Hourly Load Profiles . . . . . . . . . . . . . . . . . . . . . . 11 References. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13 Skill Development Exercises for Chapter 2 . . . . . . . . . . . . . . . . . . . 14 Chapter 3: Chilled-Water Plant Equipment. . . . . . . . . . . . . . . . . . . . 17 Instructions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17 Water Chillers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17 Water Chiller Components . . . . . . . . . . . . . . . . . . . . . . . . . . . 21 Heat Rejection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 40 Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 54 Variable-Frequency Drives (VFDs) . . . . . . . . . . . . . . . . . . . . . . 71 References. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 77 Skill Development Exercises for Chapter 3 . . . . . . . . . . . . . . . . . . . 79 Chapter 4: Hydronic Distribution Systems . . . . . . . . . . . . . . . . . . . . 81 Instructions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81 Chilled-Water Distribution Systems . . . . . . . . . . . . . . . . . . . . . 81 Condenser Water Systems . . . . . . . . . . . . . . . . . . . . . . . . . . 119 Plant Layout . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 133 References. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 135 Skill Development Exercises for Chapter 4 . . . . . . . . . . . . . . . . . . 138 Chapter 5: Optimizing Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 141 Instructions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 141 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 141 Design Procedure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 142 Selecting Chilled-Water Distribution System Flow Arrangement. . 143 Optimizing Piping Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 154 Optimizing Chilled-Water Design Temperatures . . . . . . . . . . . . . 159 Optimizing Condenser Water Design Temperatures . . . . . . . . . . 164

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Contents Selecting Cooling Towers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .167 Water-Side Economizers (WSEs) . . . . . . . . . . . . . . . . . . . . . . . . .173 References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .178 Skill Development Exercises for Chapter 5. . . . . . . . . . . . . . . . . . 179 Chapter 6: Chiller Procurement . . . . . . . . . . . . . . . . . . . . . . . . . . . 181 Instructions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 181 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 181 Chiller Procurement Procedures . . . . . . . . . . . . . . . . . . . . . . 181 Case Study . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 199 Simplified Procurement Procedure . . . . . . . . . . . . . . . . . . . . . 201 References. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 204 Skill Development Exercises for Chapter 6. . . . . . . . . . . . . . . . . . 205 Chapter 7: Controls. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 207 Instructions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 207 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 207 Sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 209 Control Valves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 226 Controllers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 230 Network Interfaces . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 231 Performance Monitoring . . . . . . . . . . . . . . . . . . . . . . . . . . . . 233 Control Schematics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 235 Control Sequences . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 239 Appendix A—TOPP Model Coefficients. . . . . . . . . . . . . . . . . . . . 264 Appendix B—Detailed Sequence of Operation (SOO) . . . . . . . . . 268 References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 277 Skill Development Exercises for Chapter 7. . . . . . . . . . . . . . . . . . 279 Chapter 8: Commissioning . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 283 Instructions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 283 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 283 Commissioning Overview . . . . . . . . . . . . . . . . . . . . . . . . . . . 283 Commissioning Focus . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 284 Sequence of Operation (SOO) Review . . . . . . . . . . . . . . . . . 285 Point-to-Point Checkout . . . . . . . . . . . . . . . . . . . . . . . . . . . . 286 Functional Testing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 288 Trend Review . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 290 References. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 292 Skill Development Exercises for Chapter 8. . . . . . . . . . . . . . . . . . 293 Online Supplemental Files: This SDL is accompanied by Excel spreadsheets, which can be found at ashrae.org/CHWSDL. These files include a chiller bid form, a simplified chiller bid form, and a pipe size optimization tool. If the files or information at the link are not accessible, please contact the publisher.

Preface Chilled-water plants are typically the most costly part of large building or campus HVAC systems and the largest energy user. Optimizing the design and control of chilled water plants can therefore have a large reduction in HVAC system life-cycle costs. But true optimization requires extensive analysis, for which few system designers have the time or funding. This course is intended to improve on the state of the art by providing updated design techniques based on rigorous lifecycle cost analysis that can provide near-optimum chilled-water plant life-cycle costs with little or no more engineering time than current practice. Recommended control sequences, also based on rigorous analysis, can improve plant performance with no more complexity than typical current practice. In addition to these design techniques, this course includes practical tips for laying out and piping chilled-water plants. The guidance applies to small plants serving small buildings as well as to district cooling plants. This SDL is accompanied by Microsoft® Excel® spreadsheets, which can be found at ashrae.org/CHWSDL. These files include a chiller bid form, a simplified chiller bid form, and a pipe size optimization tool. If the files or information at the link are not accessible, please contact the publisher.

Acknowledgments The author would like to thank Pacific Gas and Electric Company for allowing ASHRAE to use its 1999 CoolTools™ Chilled Water Plant Design Guide as the basis of this course. The CoolTools™ guide was co-authored by Steve Taylor (author of this course), Mark Hydeman, Paul DuPont, and Tom Hartman. Others who provided review and input to this course include: Brandon Gill, Taylor Engineering Mick Schwedler, Trane Steve Duda, Ross & Baruzzini Bryson Borzini, P2S Engineering Tony Mueller, P2S Engineering Anna Zhou, Taylor Engineering Steven T. Taylor, PE Taylor Engineering September 6, 2017

Fundamentals of Design and Control of Central Chilled-Water Plants

Acronyms

A/D AFLV AHU ATS BAS BEP CAD CBV CFC CHW CHWST COP COV CS CT CVRMSE CW CWFd CWFR CWFSP CWRT Cx CxA D/A DDC DI DO DP DV/DT DX EMI EMT EOR EPDM FCU

= = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = =

analog to digital automatic flow-limiting valve air-handling unit automatic transfer switch building automation system best efficiency point computer-aided design and drafting calibrated balancing valve chlorofluorocarbon chilled water chilled-water supply temperature coefficient of performance change of value constant speed current transformer coefficient of variation of root mean squared error condenser water design CW flow rate condenser water flow ratio condenser water flow set point condenser return temperature commissioning commissioning authority digital to analog direct digital control digital input digital output differential pressure derivative of voltage with respect to time direct expansion electromagnetic interference electrical metallic tubing engineer of record ethylene propylene diene monomer fan-coil unit

xvi

Acronyms GWP HBM HCFC HFC HFO HGBP HOA HX HXLWT I/O IPLV LCCA LOT MBE NPLV NPSHA NPSHR OAT ODP OEM PHXLWT PICCV PID PLC PLR PVC PWM RFI RMS RTD RTS SAT SOO SPLR TAB TDH TES THHN TOPP TXV UPS VAV VFD

= = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = =

global warming potential heat balance method hydrochlorofluorocarbon hydrofluorocarbon hydrofluoroolefin hot-gas bypass hand-off-auto heat exchanger heat exchanger leaving water temperature input/output integrated part-load value life-cycle cost analysis lockout temperature mean bias error nonstandard part-load value net positive suction head available net positive suction head required outdoor air temperature ozone depletion potential original equipment manufacturer predicted heat exchanger leaving water temperature pressure-independent characterized control valve proportional integral differential programmable logic controller part-load ratio polyvinyl chloride pulse width modulation radio frequency interference root mean square resistance temperature detector radiant time series supply air temperature sequence of operation staging part-load ratio testing, adjusting, and balancing total dynamic head thermal energy storage thermoplastic high heat resistant nylon coated theoretical optimum plant performance thermal expansion valve uninterruptible power source variable air volume variable-frequency drive

Fundamentals of Design and Control of Central Chilled-Water Plants I-P VS VSD WSE XHHW-2 XLPE

= = = = =

xvii

variable speed variable-speed drive water-side economizer cross-linked polyethylene high heat-resistant water-resistant cross-linked polyethylene

Overview

How to Use This Course The purpose of this course is to provide guidance to designers and operators of new and existing central chilled-water (CHW) plants ranging from small, singlechiller plants to large, district-cooling plants. While design engineers are the primary audience, the guide also provides useful information for operation and maintenance personnel, mechanical contractors, and building managers. Upon completion of this course, the student should have a thorough understanding of CHW plant fundamentals and principles that will be useful in conjunction with plant design or operation. The course is divided into chapters, each addressing a specific topic. It is important that you understand each topic before going on. At the end of each chapter there are questions that are intended to reinforce certain topics and to test your level of understanding. Your responses should be given to ASHRAE at www.ashrae.org/sdlonline in order to receive credit and to obtain the answer sheets.

Introduction Many large buildings, campuses, and other facilities have plants that make chilled water and distribute it to air-handling units (AHUs) and other cooling equipment. The design, operation, and maintenance of these CHW plants has a very large impact on building energy use and energy operating cost. The intent of this course is to provide tools and guidance to engineers so that the plants they design have a near optimum balance of first costs and future operating costs. The course can also be used by plant operators to understand and resolve operational problems and improve energy efficiency through controls optimization.

Organization of Material The course is organized in eight chapters. The first chapter is this overview. Loads. Chapter 2 discusses the nature of CHW loads and how they should be considered in the design of CHW plants. In the past, most engineers have only estimated the peak or maximum load. However, accounting for the time pattern of loads can be just as important. Methods of calculating peak loads

2

Chapter 1 Overview and hourly loads are reviewed. These include site measurements (for existing facilities), computer simulations, rules of thumb, and prototype buildings. Equipment. Chapter 3 reviews some basics on chillers, cooling towers, pumps, and other plant equipment. This chapter discusses the basic refrigeration cycle, water chillers, cooling towers, air-cooled condensers, pumps, and variable-speed drives. Distribution Systems. Chapter 4 discusses different ways of arranging CHW equipment in the system to meet loads while achieving energy efficiency and operational simplicity. The pros and cons of constant-flow and variableflow systems are discussed along with different primary-only and primary/secondary pumping systems. Optimizing Design. Chapter 5 provides procedures and analysis techniques for optimizing CHW plant design. Topics include optimizing the selection of distribution systems and optimizing the selection of CHW and condenser water design temperatures and pipe sizes. A spreadsheet for sizing piping and calculating pump head is provided at ashrae.org/CHWSDL (Pipe Size Optimization Tool spreadsheet). Recommendations were developed from in-depth life-cycle cost analysis of typical chiller plants and are provided as easy-to-use rules of thumb and procedures to simplify plant design while still achieving near-optimum life-cycle performance. Chiller Procurement. Chapter 6 discusses strategies for evaluating chiller options and selecting and procuring an energy-efficient and cost-effective chiller. Case studies of the chiller selection process are provided. Sample chiller bid forms are also provided (Chiller Bid Form and Simplified Chiller Bid Form). Controls. Chapter 7 explores the many design and performance issues related to controls and instrumentation of CHW plants. Topics include types of flow and temperature sensors, styles of and selection criteria for control valves, controller requirements and interfacing issues, performance monitoring, and recommended near-optimum control sequences for CHW plants, including allvariable-speed plants where all components have variable-speed drives. Commissioning. Chapter 8 discusses key elements of the commissioning process, addressing in detail sequence of operation review, point-to-point checkout, functional testing, and trend reviews. Supplemental Files. Supplemental material for this SDL is available at ashrae.org/CHWSDL. These files include a chiller bid form, a simplified chiller bid form, and a pipe size optimization tool.

Chilled-Water Plant Loads

Instructions Read the material in Chapter 2. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the Skill Development Exercises without referring to the text. Review those sections of the chapter as needed to complete the exercises.

Introduction This chapter discusses CHW plant peak loads and annual cooling load profiles and how they affect plant design and equipment capacity. Fundamental to the design process is a keen understanding of the chiller plant cooling loads and how they vary with time. If an existing plant is being modified or expanded, it is possible to monitor the cooling load and obtain an accurate estimate of both the peak load and the cooling load profile. A great many plants, however, are designed with only preliminary information available about the building’s design and function. Getting accurate peak load and cooling load profile information for these plants is much more difficult. This chapter discusses the uncertainties involved with predicting chiller plant loads and the impact of these uncertainties on the design process.

Understanding Loads and Their Impact on Design To provide an optimum CHW plant design, the designer must determine both a design (peak) load and a cooling load profile that describes how the load varies over time. The design load defines the overall installed plant capacity including the chillers, pumps, piping, and towers. The cooling load profile is required to design the plant to handle often widely variable loads stably and efficiently. This includes design decisions such as the unloading mechanisms of the chillers; the application of variable-frequency drives (VFDs) on the chillers, towers, and pumps; and the relative sizes of each piece of equipment. Certain key load parameters affect the cooling load profile and consequently the nature of the plant design. These parameters include the following: •

The use of outdoor air economizers and 100% outdoor air units

4

Chapter 2 Chilled-Water Plant Loads •

The climate in which the plant is located

•

Hours of building or facility operation

•

Base (24/7) loads such as computer rooms

For example, the cooling load profile of a San Francisco office building that operates five days per week was analyzed with and without economizers. As Figure 2-1 shows, the number of hours that the plant operates increases dramatically when an economizer is not used. Additionally, the shape of the profile changes dramatically. The profile influences the optimum selection of the number and capacity of the chillers as well as the full-load and part-load energy efficiency of the machines. What happens if this same CHW plant serving the office with economizers must also serve a relatively small data center without an economizer? Figure 22 shows the resulting load profile. This profile is also typical of district and campus cooling plants that serve relatively small 24/7 loads continuously, along with much larger peak summer loads. The plant clearly will need to operate efficiently at low loads. The plant must be designed very differently than the one serving the office building alone.

Figure 2-1

Cooling load profiles, five-day office in San Francisco.

Fundamentals of Design and Control of Central Chilled-Water Plants I-P

Figure 2-2

5

Cooling load profile of CHW plant serving office building plus a small data center.

Peak Loads Overview The process for estimating peak cooling loads in new construction is explained thoroughly in 2017 ASHRAE Handbook—Fundamentals (Chapter 18). The basic variables for peak load calculations include weather conditions, building envelope design, internal heat gain, ventilation, and, to a lesser extent, infiltration. Less obvious but nonetheless important is the diversity between the various load components. The diversity of loads is the probability of simultaneous occurrence of dynamic peak loads. In other words, diversity accounts for the fact that the envelope, occupancy, lighting, and plug loads will not each peak at the same time in all spaces simultaneously. Diversity is often underestimated by designers, particularly for large central plants. It is not uncommon for peak central plant loads to be less than half of the connected building design peak loads. There is also inherent uncertainty in peak load calculations. Any number of elements can make the actual load differ from the calculated load. For instance, the following may occur: •

Weather conditions can vary over a period of time as a result of increasing urbanization, climate change, and changes in land use.

6

Chapter 2 Chilled-Water Plant Loads •

• •

Building envelope elements do not always perform as expected due to issues such as thermal short-circuits of structural members and poor air barriers, among others. Changes may occur in operation (such as tenants moving into or out of the building). Internal loads (lighting, plug loads, and people) can be significantly different from those estimated in load calculations and can vary over time.

Often the characteristics of the loads served are not clear at the time of the plant design. This is often the case with district or campus systems where the designer must essentially guess at system and infrastructure capacity to support future growth. Simulation tools (discussed in the Determining Peak Loads and Determining Hourly Load Profiles sections) and budgets based on measured existing buildings’ usage can be quite helpful. A plant expansion or remodel provides the opportunity to monitor the existing plant for peak and operating loads. Most building automation systems (BASs) have the capability of supporting trend logs. Of course, the plant must also be provided with instrumentation (such as flowmeters and temperature sensors) to provide useful load information. Also, a good operator can often accurately report on the percent of full load that the plant sees during peak weather conditions. For most designers the perceived risks of understating the peak load condition (and undersizing the cooling plant) are much greater than overstating the peak load. An undersized cooling plant may not meet the owner’s expectations for comfort and may affect the owner’s ability to manufacture products or provide essential services. Oversizing the cooling plant, on the other hand, carries an incremental first-cost penalty and can have a positive or negative energy impact depending on the piece of equipment and how it is controlled. Oversized cooling towers and pipes tend to reduce the energy costs of operating the plant. Oversized pumps and chillers often run inefficiently at low loads, although the use of variable-frequency drives (VFDs) mitigates this to a great extent. Because oversizing always carries a first-cost premium, it is prudent to not oversize plants. Where actual loads, future growth, or diversity are uncertain, starting small with provisions (space and piping manifold sizes) for the addition of future pumps, towers, and chillers is advisable.

Annual Load Profiles Overview A cooling load profile is a time series of cooling plant loads along with concurrent weather data. The primary role of a cooling load profile is to facilitate the correct relative evaluation of competing design options. An accurate understanding of the cooling load profile affects the plant configuration. For example, a plant that serves a hotel complex with long periods of very low loads would be designed differently than a plant that serves widely varying loads only in mild and warm weather during the daytime, such as a plant serving an office building.


7

If the actual cooling loads are closely related to weather data, then temperature bin estimating techniques may produce satisfactory analysis results. However, in most cases the load is not strongly correlated with outdoor air temperature due to the predominance of internal and solar loads, which are not dependent on outdoor air temperature and humidity. Using bin weather data alone for optimization calculations will seldom provide the accuracy needed for a truly optimized plant. Therefore, to accurately address the impact of the expected load profile in a chiller plant’s design, it is necessary to have hourly load data for an entire “typical” year. As noted above, the designer’s ability to accurately project hourly load profiles into the future includes significant uncertainty. Designers should address uncertainty in the development of the annual load profiles. One approach is to consider multiple load profiles, representing a reasonable range of changes in operating conditions, when designing a plant.

Oversizing/Undersizing Considerations Because of the uncertainty inherent in design parameters and the risks associated with undersizing the plant, most CHW plants are larger than needed to meet maximum load conditions. Some impacts of oversizing a CHW plant are as follows: •

•

•

•

•

•

Oversized plants always cost more to build. While a plant’s cost may not vary linearly with its total capacity, larger plants have more expensive chillers, larger pumps, and possibly larger piping. When operating at part loads, an oversized fixed-speed chiller may not perform as efficiently as a smaller machine. Conversely, a variable-speed chiller at part load and reduced lift may operate more efficiently than a smaller machine at full load. Oversized chillers have larger CHW and condenser water (CW) pumps that consume more energy if the pumps are constant speed. This penalty can be significantly reduced if the pumps have VFDs or if the CHW plant consists of multiple smaller pumps. Oversized chillers can result in greater wear and tear and greater fluctuations in CHW supply temperature because chillers can only turn down so much before they must cycle off their compressors and then wait to restart them. The larger piping in an oversized plant will have less pressure drop than that of a plant whose piping is “rightsized.” Rightsized piping will reduce pumping energy if pumps have VFDs. For campuses where future loads are extremely uncertain, oversizing piping is usually a very good investment. An oversized plant’s cooling towers may save energy by allowing the tower fans to run slower if fans have VFDs. Also, they may produce lower CW temperatures for more efficient part-load operation of the chillers. Conversely, oversized cooling towers may have flow turndown problems that force the operators to use fewer cells at higher fan speeds, which can increase plant energy use.

8

Chapter 2 Chilled-Water Plant Loads The owner’s criteria may call for incorporating redundant chillers, pumps, and other equipment to reduce exposure to equipment failure. Redundant or spare equipment is a separate issue from oversizing, because it does not reduce the ability of the plant to adjust capacity to match the load. To mitigate problems with oversizing, a CHW plant must run efficiently at low loads. Chapter 5 discusses strategies for achieving optimum selection of chiller configurations. The following example from a computer simulation model helps demonstrate the issue of oversizing. In this case, an 800 ton cooling plant serves an office complex that operates on a basic five days per week schedule. Typical load profiles were scaled for peak cooling load of exactly 450 tons. The plant was modeled with the following scenarios: • • • •

A single 800 ton centrifugal chiller with inlet vane control The same 800 ton centrifugal chiller with VFDs Two 400 ton centrifugal chillers with inlet vane control The same two 400 ton centrifugal chillers each with VFDs

Figure 2-3 shows the results of this simulation. Note the dramatic reduction in annual cooling energy consumption when the VFD is added to the 800 ton machine and also when multiple machines are used. Although other scenarios may produce similar or better results, this example illustrates that the energy penalty for an oversized plant can be dramatically reduced if efficient turndown is incorporated into the design. By either adding a VFD on a single chiller or providing two smaller fixed-speed chillers, the annual energy is reduced by approximately one third. Combining these measures (two chillers with VFDs) reduces the annual energy by nearly one half.

Figure 2-3

Cooling energy usage for four design alternatives.


9

The following is some guidance with respect to loads and their impact on system design: •

• •

•

Avoid serving small 24/7 loads without economizers, such as a server room fan-coil, with a system that primarily serves large intermittent loads (such as office building air handlers with outdoor air economizers). The plant would have to run at very low loads, which is not only inefficient but can also decrease equipment life due to cycling. These auxiliary loads may be better served by separate cooling systems, such as direct-expansion (DX) units. If many hours at low loads are unavoidable, provide multiple chillers and perhaps unequally sized chillers to improve low-load performance. Where a wide range of loads can be expected (which is true of most plants, even those serving typical data centers), use VFDs on all equipment (all pumps, tower fans, and chillers), creating what is called an all-variable-speed plant. As discussed in Chapter 5, VFDs are almost always cost-effective on all equipment (with the possible exception of CW pumps), and they substantially mitigate the energy impact of oversized plants. The benefits of rightsizing or tightsizing (sizing the plant precisely for the expected loads) are often overstated, particularly for multi-chiller allvariable-speed plants, which can operate efficiently over a wide range of loads. There are also disadvantages to tightsizing, such as having to retrofit additional or larger equipment at extremely high cost if any of the many assumptions made about future loads are wrong. Owners expect and deserve flexibility to handle future loads within reason, so aggressive sizing may not be the best approach despite some first-cost and efficiency benefits.

Determining Peak Loads Calculations/Simulations ASHRAE Handbook—Fundamentals, Chapter 18, defines accepted methods and procedures for cooling load calculations. These well-known procedures include information on ventilation and infiltration, climatic design information, residential and nonresidential load calculations, fenestration, and energy estimating methods. In discussing cooling load principles, the Handbook emphasizes the importance of analyzing each variable that may affect cooling load calculations: The variables affecting cooling load calculations are numerous, often difficult to define precisely, and always intricately interrelated. Many cooling load components vary in magnitude over a wide range during a 24-h period. Because these cyclic changes in load components are often not in phase with each other, each must be analyzed to establish the resultant maximum cooling load for a building or zone. (18.1)

10

Chapter 2 Chilled-Water Plant Loads Starting in the 2001 edition, the Handbook supports only two methods of load calculation: the heat balance method (HBM) (a fundamental first-principles approach) and the radiant time series (RTS) method (an approximation of the heat balance method). For all practical purposes, both of these methods require computer simulation to analyze. Although these calculation techniques have worked very well over the years, designers must be aware of the limitations of these techniques and recognize that the methods do not all predict the same loads. Because of the uncertainties previously discussed, the design load calculations may be different than the actual chiller plant peak load. Selecting the maximum capacity of the plant is important, but it is perhaps even more important to consider the plant’s part-load performance.

Site Measurements When an existing chiller plant is being remodeled or expanded, it is possible to monitor the actual peak cooling load to obtain invaluable information. The monitoring can be short term (several months) to establish peak load and daily trends or can be long term (one year or longer) to determine annual load profiles. Successfully measuring energy and load performance of a cooling plant requires rigorous monitoring protocols (see ASHRAE Guideline 22, Instrumentation for Monitoring Central Chilled-Water Plant Efficiency [2012], for example). These monitoring protocols comprise four stages: 1. Survey of monitoring sites: Conduct a complete audit of the CHW plant. Develop a comprehensive systems diagram. 2. Monitoring plan: From the comprehensive systems diagrams prepare a plan for determining the data to be monitored, the monitoring equipment needed, and the duration of monitoring. Typical monitoring equipment includes data loggers, flow measurement devices, temperature measurement devices, and power measurement devices. Also required are concurrent measurements of weather data, including dry-bulb temperature and wet-bulb temperature. In many modern plants, the necessary instrumentation for measuring and trending load and weather is permanently installed as part of the plant’s controls system infrastructure, obviating the need for additional short-term sensors and data loggers. Weather data may also be available online from nearby government weather stations. 3. Field installation: Install instrumentation in accordance with the monitoring plan and the installation instructions. Take spot measurements to ensure that the equipment is calibrated properly and that all sensors and instruments are working correctly. Provide guidelines to operators. Have a plan for removal of instrumentation and patching of insulation, etc. 4. Data collection and analysis: Obtain data and provide validation. Perform analysis on both a basic level (for example, simple temperature logs of chiller energy usage) and a more detailed level (for example, chiller plant energy performance as a function of various elements such as time and weather). If the weather in the monitoring period does not reach the design


11

weather conditions, then it will be necessary to extrapolate the measured load as a function of weather to determine the peak load. This should be done with care as load curves regressed from available data may not be an accurate representation of real-world loads when extrapolated to temperatures well outside the monitoring range. From this procedure, the peak loads will emerge, as well as the relationships and interactions of the various components. The quality of the monitoring protocol will determine the accuracy and usefulness of the results.

Determining Hourly Load Profiles There are several methods for determining annual cooling load profiles depending on what stage the project is in and the resources available for analysis. The following are common methods for determining annual cooling load profiles: • •

Computer simulation models (customized and prototypical) Site measurements

For new construction, custom computer simulations and prototypical simulations are the two most common methods. Customized simulations have the greatest potential for accuracy but can be costly to develop and are subject to modeling error. Prototypical simulations offer quick and relatively inexpensive analysis but may not be as accurate as customized simulations. For retrofit and expansion of existing plants, site visits may be conducted to measure profiles. This technique yields the most accurate results but requires special planning, technical expertise, equipment, budget, and time. Each of these methods can be combined with statistical and mathematical techniques from a variety of sources including short-term measurements, site data, and billing data. These hybrid approaches offer the best possibility to balance accuracy and effort. The following sections discuss each technique.

Computer Simulation Models Computer simulation models customized for a specific project can take between a few hours to several person-weeks of time to develop, depending on the complexity of the building geometry and the effort spent on making the model accurate. With recent advances in simulation tool data exchange, the effort to build these models has significantly decreased. For example, building geometry can be imported from computer-aided design and drafting (CAD) programs into some load or simulation tools. Examples include EnergyPlus (which supports both IAI IFCs and GBXML), DOE2, and several commercial load and energy programs that support GBXML. Although these tools are far from “plug and play” they still dramatically reduce the time required to create models, and they reduce modeling errors.

12

Chapter 2 Chilled-Water Plant Loads For projects that are early in design and evaluations for campus systems, prototypical models are a useful tool. Many of the simulation tools now incorporate wizards that enable designers to develop a typical building for analysis in a matter of minutes. Examples include eQUEST, a free, front-end interface to DOE2.2 and DOE2.3. Computer simulation models require experienced modelers for inputting data and checking results. To assess the impact of uncertainties, the modeler should consider a range of input variations representing the best estimate, possible but likely low loads, and possible but likely high loads.

Site Measurements Site monitoring to determine peak loads was discussed earlier in this chapter. The same site monitoring protocol can be used for determining cooling load profiles based on either short-term or long-term measurements. Longterm monitoring is not common because it is costly and time-consuming to obtain the data. Long-term trend data of plant performance are sometimes available from BAS trends, but the data are often inaccurate or incomplete. Experience with long-term data indicates that due to weather and other variables, a single year’s measurement would not match the second year’s data and as a result is not deterministically exact. The utility of long-term monitoring is maximized by ensuring that the monitoring period captures the full range of anticipated weather conditions (often necessitating four to six months of data centered on a swing season, depending on climate zone) and all unique seasonal operating profiles. For example, a college central plant will have unique load profiles during in-session and out-of-session periods that must both be monitored. If these weather and schedule range criteria are met, then the data can used to create a robust regression model accounting for seasonal factors, day type (weekday, weekend, holiday, etc.), time of day, and ambient weather conditions. The resulting model can then be applied to a prototypical weather year to generate an expected annual load profile for the plant. If the weather and schedule criteria are not met, then the model runs the risk of generating invalid results when extrapolated outside of the conditions observed during the monitoring period. When long-term data are not available, or are not practical to capture, short-term data can potentially be used to define the basic shape of a typical 24-hour load profile by season or month. However, such data are climate sensitive and the associated weather/load profiles are difficult to record, especially considering the solar aspect of the load. When a few weeks of continuous short-term load and weather data are available, but are insufficient to generate a robust time- and temperaturedependent regression model as discussed above, these data may instead be used to calibrate a computer simulation model. This hybrid modeling and site measurement approach is fairly laborious. The modeler must use the weather data collected during the short-term monitoring period to create a custom simulation weather file for the site corresponding to the monitoring period. The simu-


13

lation model is then run with that weather file and calibrated to approximate the site load data over the monitoring period by adjusting schedule and load profile variables. The validity of such calibrations is measured using statistical metrics including mean bias error (MBE) and coefficient of variation of root mean squared error (CVRMSE). Refer to International Performance Measurement & Verification Protocol (EVO 2002) for further discussion of these calibration metrics.

References ASHRAE. 2012. ASHRAE Guideline 22, Instrumentation for monitoring central chilled-water plant efficiency. Atlanta: ASHRAE. ASHRAE. 2017. Chapter 18, ASHRAE Handbook—Fundamentals. Atlanta: ASHRAE. EVO. 2002. International Performance Measurement & Verification Protocol. Washington, DC: Efficiency Valuation Organization.

14


Skill Development Exercises for Chapter 2 2-1

Why is the shape of a CHW plant’s cooling load profile a critical factor in plant design? a. It dictates the conditions under which the plant must operate efficiently to minimize energy costs. b. It impacts the selection of chillers because the plant must be able to handle the full range of expected load conditions stably. c. It drives the peak capacity required of the plant. d. Both (a) and (b). e. All of the above.

2-2

Which of the following are true regarding the impact of air-side economizing on the annual load profile of a plant serving an office building? i. It reduces the total annual ton hours served by the plant. ii. It shifts the most common load percentage to a lower value. iii. It reduces the peak load of the plant. iv. It reduces the plant’s run hours. a. (i), (ii), (iv) b. (i), (ii) c. (i), (iii), (iv) d. (i), (ii), (iii), (iv)

2-3

ASHRAE Handbook—Fundamentals supports which of the following load calculations methodologies? a. RTS b. HBM c. Transfer function method d. Only (a) and (b) e. All of the above

2-4

Oversizing CHW plants a. Typically yields more efficient pumping in variable-speed applications due to lower friction losses. b. Usually leads to more efficient chiller operation. c. May cause controllability issues if chillers are not properly selected for stable low-load operation. d. Is problematic when the condenser and CHW pumps are variable speed. e. Both (a) and (c).

2-5

You are replacing oversized chillers in an existing CHW plant with modern direct digital control (DDC) controls, trending capabilities, and recently calibrated instrumentation. Which of the following is the recommended approach for determining peak load to size the new chillers? a. Develop a load model of the facility using a simulation tool and utilize the peak load estimated therefrom.


2-6

15

b. Develop a load model of the facility using a simulation tool and calibrate the model on an annual basis using utility billing data, then assess peak load with the model. c. Utilize the DDC system’s primary CHW loop flowmeter and supplyand return-temperature sensors to trend load. Use a few months of trended load and local weather data from the summer and/or swing seasons to develop a load profile and predict peak load therefrom. d. Install temporary National Institute of Standards and Technology (NIST)-calibrated instrumentation, including an ultrasonic flowmeter and supply- and return-temperature sensors to trend load. Use the same approach as option (c) to predict peak load. True or False: In early design, developing a prototypical model of the proposed building is usually too cost prohibitive to assist in plant design development. a. True b. False

Chilled-Water Plant Equipment


Introduction Design engineers seeking to maximize the performance and economic benefits of upgraded or new CHW plants need a thorough understanding of the major equipment used in these plants. This chapter provides an overview of the primary equipment, as well as essential information on how the components relate to one another, how they are controlled, and what their physical and operational limitations are. This chapter discusses the following: • • • • •

The basic vapor compression refrigeration cycle The components commonly used in commercial water chillers Methods of heat rejection, such as cooling towers and air-cooled refrigerant condensers The characteristics of different types of pumps, pump and system curves The application and efficiency of VFDs

The intent is to familiarize the reader with basic components. For additional and more in-depth information, consult with equipment manufacturers, references such as ASHRAE Handbook—HVAC Systems and Equipment (2016d), and other ASHRAE self-directed learning courses such as Fundamentals of Water System Design (2015).

Water Chillers This section presents an overview of the current water chiller technologies. Technology changes rapidly, so students are encouraged to browse manufacturers’ websites for the most current information on technologies and refrigerants.

18

Chapter 3 Chilled-Water Plant Equipment

Vapor Compression Refrigeration Cycle The vapor compression refrigeration cycle is the fundamental thermodynamic basis for removing heat from buildings and rejecting it to the outdoors. (Absorption chillers, which use a completely different technology, are discussed in the Absorption Chillers section.) The refrigeration cycle requires four basic components: • • • •

Compressor Evaporator Condenser Expansion device

The vapor compression refrigeration cycle diagram (Figure 3-1) shows the relationship of these components, as does the pressure-enthalpy chart, also known as a P-h diagram (Figure 3-2). These diagrams cover the liquid-vapor regions specific to the cycle refrigerant. The following is a description of the refrigeration cycle using the points noted on Figure 3-2: •

Figure 3-1

Starting at Point A, the refrigerant is a liquid at high pressure. As it passes through the expansion device to Point B, the pressure drops. At Point B the refrigerant is a mixture of liquid and gas. At this point the gas is called flash gas. Alternatively, the liquid could be subcooled to Point A , which is below the saturation temperature. If this is done, the liquid would pass through the expansion device, resulting in less flash gas present at Point B .

The refrigeration cycle.


Figure 3-2

19

Pressure-enthalpy chart.

•

From Point B to Point D, the liquid is converted to a gas by absorbing heat (refrigeration effect). Notice the gas leaving the evaporator at Point D has been heated to a level greater than saturation as shown by Point C. The heat from Point C to D is called superheat. Superheating in the evaporator ensures that there is no liquid in the refrigerant as it moves into the compressor.

•

From Point D the refrigerant is drawn into the suction of the compressor where the gas is compressed, as shown by Point E. At Point E, the temperature and pressure of the gas have been increased. The refrigerant is now called hot gas. Notice that this point is to the right of the saturation curve, which also represents a superheated state. The hot gas, Point E, moves into the condenser where the condensing medium (either air or water) absorbs heat and changes the refrigerant from a gas back to a liquid as shown by Point A. At Point A the liquid is at an elevated temperature and pressure. The liquid is forced through the liquid line to the throttling device and the cycle is repeated.

•

The difference between the condensing temperature and evaporating temperature is called the lift. The lift is a primary driver of the efficiency of the chiller, discussed in the following sections, including Water Chiller Components.

Refrigerants To address safety and environmental concerns, refrigerants must have low toxicity, low flammability, and a long atmospheric life. They also must have zero or minimal impact on stratospheric ozone and on global warming via greenhouse effects.

20

Chapter 3 Chilled-Water Plant Equipment The relative ability of a refrigerant to destroy stratospheric ozone is called its ozone depletion potential (ODP). Older refrigerants—particularly chlorofluorocarbons (CFCs)—are known to destroy stratospheric ozone; CFCs have been phased out according to the 1987 Montreal Protocol (see Table 3-1). The production of CFCs in developed countries ceased in 1995, and another common refrigerant type, hydrochlorofluorocarbon (HCFCs), have been or are due to be phased out in a few years. HCFC-22, commonly used for small air conditioners and chillers, has already been phased out for use in new equipment. The most common HCFC refrigerant used in chillers is HCFC-123, which is scheduled to be phased out of production for new equipment in 2020 and production will be banned in 2030 in developed countries. Chillers installed now can be expected to be operational well past the production ban date. However, HCFC123 will likely be available well into the middle of the twenty-first century and certainly within the lifetimes of machines currently being manufactured due to stockpiling, recovery, and recycling of HCFC-123 from existing chillers as they are replaced. In response to the Montreal Protocol, several zero-ODP hydrofluorocarbon (HFC) refrigerants were developed, the most common of which is HFC-134a. The global warming potential (GWP) of refrigerants is another significant environmental issue. Gases that absorb infrared energy enhance the greenhouse effect in the atmosphere, leading to the warming of the earth. Refrigerants have been identified as greenhouse gases. A chart showing the ODP versus GWP of various refrigerants is shown in Table 3-2. Theoretically, the best refrigerants would have zero ODP and zero GWP. Unfortunately, many of the refrigerants with zero or low ODP and GWP, such as R717 (ammonia), are flammable or toxic or both. Table 3-1

Montreal Protocol

The 1987 Montreal Protocol, and subsequent revisions, established the following timeline for the phaseout of chlorofluorocarbons (CFC) and hydrochlorofluorocarbon (HCFC). Refrigerant

Year

Restrictions

CFC-11

1996

Ban on production

CFC-12

1996

Ban on production

HCFC-22

2010

Production freeze and ban on use in new equipment

2020

Ban on production

2015

Production freeze

2020

Ban on use in new equipment in developed countries

2030

Ban on production in developed countries

—

No restrictions at this point in time*

HCFC-123

HFC-134a

* As of this date there are no restrictions in North America on the use of R-134a. This could change, so the reader is advised to seek out the most recent information. HFCs have been or are proposed to be banned in many European countries.

Fundamentals of Design and Control of Central Chilled-Water Plants I-P Table 3-2

21

ODP Versus GWP for Common Refrigerants

Refrigerant

Ozone Depletion Potential (ODP)

Global Warming Potential* (GWP)

Theoretical Efficiency, kW/ton**

R-11 Trichlorofluoromethane

1

4750

6.6

R-12 Dichlorodifluoromethane

1

10,900

6.3

R-22 Chlorodifluoromethane

0.04

1810

6.2

R-123 Dichlorotrifluoroethane

0.02

77

6.5

R-1234yf

0

4

6.4

R-134a Tetrafluoroethane

0

1430

6.3

R290 Propane

0

3

6.2

R407C (23% R-32, 25% R-125, 52% R-134a)

0

1770

6.0

R-410A (50% R-32, 50% R-125)

0

2080

5.9

R717 Ammonia—NH3

0

0

6.3

* GWP values from IPCC (2007). ** Theoretical efficiencies from Calm (2005).

Table 3-2 also shows the theoretical efficiencies of each refrigerant for a typical cooling application. Refrigerant type is not the only factor that determines actual chiller efficiency; factors such as compressor type and design and the heat transfer effectiveness of the evaporator and condenser also play major roles. So, from a user’s perspective, refrigerant theoretical efficiency is not important with respect to chiller selection; the actual efficiency of the equipment is what matters. In the constant effort to simultaneously minimize GWP and ODP, hydrofluoroolefin (HFO) refrigerants have been developed, including R-1234yf (already used in automobile air conditioning in place of R-134a), R-1234ze (another R-134a replacement), and R-1233zd (a low-pressure refrigerant comparable to R-123). Some of these are slightly flammable, which has given rise to a new flammability class 2L, for which application regulations are currently under development (see ASHRAE Standard 34-2016 for more information on flammability classes).

Water Chiller Components Compressors The four most common types of compressors used in packaged water chillers are

22

Chapter 3 Chilled-Water Plant Equipment •

Reciprocating

•

Scroll

•

Screw

•

Centrifugal

The first three types are called positive displacement compressors because they compress the refrigerant by trapping a fixed amount and forcing (displacing) the trapped vapor into smaller and smaller volumes. Reciprocating compressors are almost nonexistent in modern chillers, replaced primarily by scroll and screw compressors. Hence they are not discussed here; additional information about them is available in ASHRAE Handbook—HVAC Systems and Equipment (2016d).

Scroll Scroll compressors (Figure 3-3) are the most common compressor type on for smaller chiller sizes, although there are scroll machines available up to 400 tons in capacity. They are mostly used in outdoor air-cooled chillers. Scroll compressors used in chillers typically range from 5 to 50 tons and are single speed without unloading capability; compressors are cycled to control capacity. Some advanced scroll compressors achieve variable unloading capacity by rapidly engaging and disengaging the scrolls. These compressors run at constant speed and have unloading efficiencies similar to cycling compressors but with much finer temperature control (smaller temperature swings). Variable-speed scroll compressors are also available and are beginning to be applied to chiller applications.

Figure 3-3

Scroll compressor.


Figure 3-4

23

Single screw compressor.

Screw There are two common screw compressor types: single screw and twin screw.

Single Screw The single screw (Figure 3-4) consists of a single cylindrical main rotor that works with a pair of gate rotors. The compressor is driven through the main rotor shaft, and the gate rotors, followed by direct meshing action. As the main rotor turns, the teeth of the gate rotor, the sides of the screw, and the casing trap refrigerant. As rotation continues, the groove volume decreases and compression occurs. Because there are two gate rotors, each side of the screw acts independently. Single-screw compressors are noted for long bearing life, as the bearing loads are inherently balanced. Some single-screw compressors have a centrifugal economizer built into them. This economizer has an intermediate pressure chamber that takes the flash gas (via a centrifugal separator) from the liquid and injects it into a closed groove in the compression cycle, which increases efficiency. The capacity of the single screw compressor is typically controlled from a slide valve in the compressor casing that changes the location where the refrigerant is introduced into the compression zone. This causes a reduction in groove volume, and hence the volume of gas compressed varies (variable compressor displacement). These compressors are fully modulating. The single screw has slide valves on each side that can be operated independently. This allows the machine to have a very low turndown with good part-load energy performance.

Twin Screw The twin screw (see Figure 3-5) is also known as a double helical rotary screw. The twin screw consists of two mating helically grooved rotors, one

24

Chapter 3 Chilled-Water Plant Equipment male and the other female. Either the male or female rotor can be driven. The other rotor either follows the driven rotor on a light oil film or is driven with synchronized timing gears. At the suction side of the compressor, the gas fills a void between the male and female rotors. As the rotors turn, the male and female rotors mesh and work with the casing to trap the gas. Continued rotation decreases the space between lobes, and the gas is compressed. The gas is discharged at the end of the rotors. The twin screw has a slide valve for capacity control located near the discharge side of the rotors, which bypasses a portion of the trapped gas back to the suction side of the compressor. Some manufacturers offer screw chillers with VFDs. In addition to excellent part-load and part-lift performance, these chillers offer significantly reduced noise and wear at off-design conditions. Variable-speed screw chillers, unlike centrifugal chillers, do not have surge issues (discussed below) and thus can operate at lower speeds at higher lifts.

Centrifugal Centrifugal compressors are dynamic (as opposed to positive displacement) compression devices that on a continuous basis exchange angular momentum between a rotating mechanical element and a steadily flowing fluid. Like centrifugal pumps, centrifugal chillers have an impeller that rotates at high speed. The refrigerant enters the rotating impeller in the axial direction and is discharged radially at a higher velocity. The dynamic pressure (kinetic energy) of the refrigerant obtained by the higher velocity is converted to static pressure through a diffusion process that occurs in the stationary discharge or diffuser portion of the compressor just outside the impeller. A centrifugal compressor (see Figure 3-6) can be single stage (having only one impeller) or multistage (having two or more impellers). On a multistage centrifugal compressor, the discharge gas from the first impeller is directed to

Figure 3-5

Twin screw compressor.


Figure 3-6

25

Hermetic centrifugal compressor.

the suction of the second impeller and so on for each stage provided. Like the rotary compressor, multiple stage centrifugal chillers can incorporate economizers, which take flash gas from the liquid line at intermediate pressures and feed this into the suction at various stages of compression. The result is a significant increase in energy efficiency. Centrifugal compressors can be either open or hermetic. Open centrifugal compressors have the motors located outside the casings with the shaft penetrating the casing through a seal. Hermetic centrifugal compressors have the motor and compressor fully contained within the same housing, with the motor in direct contact with the refrigerant. Because the discharge pressure developed by the compressor is a function of the velocity of the tip of the impeller, for a given pressure, smaller-diameter impellers result in faster impeller speeds. Similarly, for a given pressure, the more stages of compression there are, the smaller the impeller diameter needs to be. With these variables in mind, some manufacturers have chosen to use gear drives to increase the speed of a smaller impeller, while other manufacturers use direct drives with larger impellers and/or multiple stages. High-speed directly coupled motor-impeller compressors are also available. Recently, centrifugal chillers from some manufacturers have become available with oil-free bearings, either magnetic “frictionless” bearings or ceramic bearings. This improves efficiency by almost eliminating bearing losses, and the removal of oil from the system improves heat transfer efficiency. The elimination of oil also substantially reduces the minimum differential pressure (DP) across the condenser and evaporator (head pressure) (see the Chapter 7 section Control Schematic for a Typical Plant for more information on head pressure). Chillers requiring oil must maintain a minimum head pressure to ensure that oil can circulate through the system. This limits how much the plant controls can take advantage of mild weather to reduce condensing temperatures and chiller lift. As is discussed in more detail below, the lower the lift, the higher the efficiency.

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Chapter 3 Chilled-Water Plant Equipment One of the characteristics of the centrifugal compressor is that it can surge. Surge is a condition that occurs when the compressor is required to produce high lift at low volumetric flow. Centrifugal compressors must be controlled to prevent surge, and this is a limit on part-load performance. During a surge condition, the refrigerant alternately moves backward and forward through the impeller, creating noise, vibration, and heat. Prolonged operation of the machine in surge condition can lead to failure. Surge is relatively easy to detect in that the electrical current to the compressor will alternately increase and decrease with the changing refrigerant flow. Just before entering surge, the compressor may exhibit a property called incipient surge, in which the machine gurgles and churns. This is not harmful to the compressor but may create unwanted vibration. The electrical current does not vary during incipient surge. The capacity of centrifugal compressors may be controlled by two methods. The most common is to use inlet guide vanes or pre-rotation vanes (see Figure 3-7). The adjustable vanes are located in the compressor’s suction at the eye of the impeller and swirl the entering refrigerant in the direction of rotation. This changes the volumetric flow characteristics of the impeller, providing the basis for unloading. A second control method is to vary the speed of the impeller in conjunction with using inlet guide vanes. As with a variable-speed fan or pump, reducing the impeller speed produces extremely efficient part-load efficiency. But with fans and pumps, the required flow and pressure vary together; as the flow rate falls, the pressure required falls as well, roughly as the square of the flow rate (see subsequent discussion on pumps). But chillers must maintain a minimum speed that does not necessarily vary with refrigerant flow and chiller capacity. Rather minimum speed depends on the following: •

Figure 3-7

Minimum speed required to move the refrigerant from the low-pressure side (evaporator) to the high-pressure side (condenser): Condenser and evaporator DP can vary with chiller load somewhat, depending on the application. For an office building, the condensing temperature can be reduced in mild

Inlet guide vanes.


•

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weather. The evaporator temperature can, likewise, be raised in mild weather due to reduced ventilation and heat transfer loads on the plant. The same would not be the case for a data center, because the load is weather independent. A minimum DP also must be maintained for oil circulation unless the chillers are oil free. Minimum speed required to avoid surge: The chiller is most efficient when operating at its lowest speed just before going into surge, but efficiency falls and damage can occur if the surge line is crossed. So the controls must dynamically determine the surge line with a fairly robust strategy to stay close to but out of the surge region. Some controls use a chiller map of the surge line as a function of load and DP built into the controller. Others will lower speed until current spikes are sensed as the compressor enters surge, then respond by increasing speed.

When the impeller is at the minimum speed, further reductions in capacity are obtained by using the inlet guide vanes. Variable-speed centrifugal compressors can produce the most energy-efficient part-load performance of any compressor type. But to do so the minimum speed must be as low as possible, which in turn requires that the condenser and evaporator DP and temperature (lift) be as low as possible. Minimizing lift requires aggressive water temperature reset strategies. Without these strategies, which are discussed in Chapter 6, variable-speed centrifugal chillers can be no more efficient than fixed-speed chillers. In fact, due to the inefficiency of the drive, if the lift is not reduced, a variable-speed chiller may be less efficient than a constantspeed chiller.

Absorption Chillers The absorption process is another way to evaporate and condense refrigerants, but the process is thermal/chemical rather than mechanical. Though appearing quite complex, absorption chillers use the same refrigeration process discussed for mechanical compression except that phase change is achieved with an absorber, generator, pump, and recuperative heat exchanger (HX). The design used by almost all commercial absorption chillers uses lithium bromide as the absorbent and water as the refrigerant. See Chapter 18 of ASHRAE Handbook—Refrigeration (2014) for a description of how the absorption cycle works. A single-effect absorption process (Figure 3-8) is similar to a double-effect absorption process (Figure 3-9), except that a generator, condenser, and HX are added for the double-effect absorption process. The refrigerant vapor from the primary generator runs through a HX (secondary generator) before entering the condenser. The secondary generator with the hot vapor on one side of the HX boils some of the lithium bromide and refrigerant solution, creating the double effect. The double-effect absorption process is significantly more energy efficient than the single-effect absorption process, but it requires a higher temperature heat source.

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Figure 3-8

Single-effect absorption.

Figure 3-9

Double-effect absorption.


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Absorption machines can be direct fired or indirect fired. The direct-fired absorber has an integral combustion heat source that is used in the primary generator. An indirect-fired absorber uses steam or hot water from a remote source. Absorption machines are controlled by modulating the firing rate of a direct-fired machine or modulating the flow of steam or hot water in an indirect-fired machine. Variable-speed refrigerant and solution pumps greatly enhance the controllability of the absorption machine.

Evaporators Two types of evaporators are used in water chillers—the flooded shell and tube and the DX. DX evaporators may be shell-and-tube type or brazedflat-plate type. Flooded shell and tube HXs are typically used with large screw and centrifugal chillers, while DX evaporators are usually used with positive displacement chillers like the scroll and reciprocating machines. While water is the most common fluid cooled in the evaporator, other fluids are also used. These include a variety of antifreeze solutions, the most common of which are mixtures of ethylene glycol or propylene glycol and water. The use of antifreeze solutions significantly negatively affects the performance of the evaporator but may be needed for low-temperature applications. The fluid creates different heat transfer characteristics within the tubes and has different pressure drop characteristics. Machine performance is generally derated when using fluids other than water.

Flooded Shell and Tube The flooded shell and tube HX has the cooled fluid (chilled water) inside the tubes and the refrigerant on the shell side outside the tubes. The liquid refrigerant is uniformly distributed along the bottom of the HX over the full length. The tubes are partially submerged in the liquid. Distributors are used as a means to ensure uniform distribution of vapor along the entire tube length, and eliminators prevent the violently boiling liquid refrigerant from entering the compressor suction line. The eliminators are made from parallel plates bent into a Z shape, wire mesh screens, or both plates and screens. An expansion valve, float valve, or orifice maintains the level of the refrigerant. The tubes for the HX are usually both internally and externally enhanced (ribbed) to improve heat transfer effectiveness. Manufacturers typically limit water flow on the high end to prevent erosion of the piping and on the low end (typically around 3 ft/s with smooth tubes and much lower with enhanced tubes) to maintain Reynolds numbers above the laminar flow regime to maintain high heat transfer coefficients. It is best to check with the manufacturers for their specific flow rate limitations on each chiller. Flooded shell and tube HXs are available with multiple passes, with two being the most common for temperature differences from roughly 8°F to 18°F and three passes for 18°F to 25°F temperature differences. The greater the number of passes, the lower the minimum flow requirements.

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Figure 3-10

Direct expansion (DX).

Direct Expansion (DX) The DX shell and tube evaporator (Figure 3-10) has the refrigerant inside the tubes and the cooled fluid (chilled water) on the shell side (outside the tubes). Larger DX evaporators have two separate refrigeration circuits that help return oil to the positive displacement compressors during part load. DX coolers have internally enhanced (ribbed) tubes to improve heat transfer effectiveness. The tubes are supported on a series of polypropylene internal baffles, which are used to direct the water flow in an up-and-down motion from one end of the tubes to the other. DX evaporators often are limited to 15°F to 18°F temperature differences; where a high temperature difference is desired (see Chapter 5), chillers must be piped in series.

Condensers There are a number of different kinds of condensers manufactured for packaged water chillers. These include water-cooled, air-cooled, and evaporativecooled condensers. (Air-cooled and evaporative condensers are discussed later in this chapter with cooling towers and heat rejection devices.) Numerous types of water-cooled condensers are available including shell and tube, double pipe, brazed flat plate, and shell and coil. This discussion focuses on the condenser most commonly used on packaged water chillers—the shell and tube HX. A horizontal shell and tube condenser (Figure 3-11) has straight tubes through which water is circulated while the refrigerant surrounds the tubes on the outside. Hot gas from the compressor enters the condenser at the top where it strikes a baffle. The baffle distributes the hot gas along the entire length of the condenser. The refrigerant condenses on the surface of the tubes and falls to the bottom where it is collected and directed back to the expansion device then to the evaporator. The bottom tubes are usually the first pass (coldest) of the condenser water and are used to subcool the refrigerant. Often the condenser is used as the refrigerant receiver where the refrigerant is stored when not in use. The tubes can be enhanced (ribbed) on both the inside and outside. However, because the condenser water often comes from an open cooling tower, the inside of the condenser tubes may become fouled and require mechanical


Figure 3-11

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Polypropylene internal baffles.

cleaning. Inside enhancement—usually with straight or spiral grooves—may be problematic because the grooves will be the first areas to become fouled. Fouling can become a problem when concentrations of dissolved solids increase greatly above recommendations and when tube velocities drop into the laminar flow regime (below about 3 ft/s) for a significant amount of time. Even considering decreased performance of the enhanced condenser tube due to fouling, the heat exchange effectiveness with the enhanced tube may still be greater than a smooth bore tube. Design condenser water velocities range from about 3 to 12 fps. Lower speeds are acceptable for short-term conditions, such as for head pressure control during start-up, but many manufacturers recommend higher velocities for most run hours to reduce the risk of fouling. Water-cooled condensers are usually multiple pass, with two pass being most common. The condenser water side can be split into two separate tube bundles to accommodate a heat recovery mode or to add a level of redundancy in the event that the tubes need cleaning while the machine is still operational.

Accessories and Common Options Purge Units Centrifugal chillers that use low-pressure refrigerants such as R-123 operate below atmospheric pressure. When they leak, air and moisture are drawn into the machine. Purge units remove the noncondensable gases that collect in the condenser during normal operation and ultimately reduce the heat transfer effectiveness, causing greater refrigerant head pressures. Moisture inside the unit causes the formation of acids that break down the oil and increase internal corrosion. Purge units consist of compressors, motors, separators, and con-

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Chapter 3 Chilled-Water Plant Equipment densers that can be automatic or manual. Automatic purge units are preferred because they maintain the highest chiller efficiencies possible. Purge units that reduce refrigerant losses during operation should always be used. Discharge from purge units must be piped outdoors.

Oil Coolers Lubricants must be cooled, especially those used with screw machines. A small HX is provided for this purpose. The heat can be rejected through a city water connection or a CHW connection, or it may be air cooled or internally cooled by the refrigerant.

Hot-Gas Bypass (HGBP) Hot-gas bypass (HGBP) is a means of false-loading the chiller to reduce short-cycling compressors and oscillating water temperatures that occur once the chiller has reached its minimum unloading capacity. The minimum stable operating load ratio varies with chiller design. With most scroll compressor chillers, the minimum load is typically that of the smallest compressor. Some scroll chillers are available with variable-capacity compressors that can unload stably to very low loads. Screw and centrifugal chillers typically can unload to about 10% to 15% of design capacity. Below the minimum capacity, the compressor must be cycled off. If the chiller experiences many hours at loads below its minimum unloading capacity, the compressor can cycle excessively, which reduces the longevity of the equipment, particularly for fixed-speed chillers. To mitigate this problem, HGBP can be used to unload a machine to near-zero load by directing the hot gas from the compressor discharge back into the suction. There are no part-load energy savings with HGBP—chiller energy remains at that required for the minimum unloading capacity regardless of actual load. HGBP is a fairly inexpensive option, so it may be a good investment for screw and centrifugal chillers to prevent short cycling should loads be unexpectedly low, and it wastes no energy if loads turn out to be above the minimum. HGBP is usually not needed for scroll chillers with variable-capacity compressors because they have very low minimum loads. For scroll chillers with constant capacity compressors, HGBP should be avoided because of the energy waste; instead a storage tank should be added to the CHW loop to provide sufficient thermal mass to minimize cycling. The chiller manufacturer generally provides guidance for sizing the tank.

Heat Recovery Heat recovered from chillers can be used to heat buildings, domestic hot water, or a wide variety of low-temperature heating applications. Two types of heat recovery can be applied to chillers: a desuperheater condenser placed immediately at the discharge of the compressor and in series with the chiller’s main condenser and parallel condensers called a double bundle condenser. In some locales, condensers used to heat potable hot water have to be double wall


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so that any refrigerant leaks cannot contaminate the domestic water circuit. The economics of applying heat recovery condensers must consider the load profile of the source to be heated. Desuperheater condensers are generally applied on reciprocating chillers, particularly air-cooled machines. These condensers roughly recover 30% of a chiller’s heat rejection capacity and can often generate hot water 140°F, depending on load and condensing temperatures. Desuperheaters can slightly improve chiller efficiencies. As these condensers are located in series with the unit’s main condenser, they must be designed for a very low refrigerant pressure drop. Often desuperheaters are field retrofitted to chillers. Double bundle condensers are usually applied to centrifugal chillers and can recover the machine’s entire heat rejection capacity. A double bundle condenser is typically split into two separate tube bundles, or two separate condensers, with the heating water piped to one side and the cooling tower water piped to the other side. Heat is first rejected to the heating bundle and when the heating requirement decreases, the extra heat is rejected to the cooling tower. Double bundle condensers can be inefficient, as the condensing temperature will be elevated to achieve even the smallest amount of heat recovery and HGBP is often needed on centrifugal chillers to avoid surge at even medium loads due to the high lift. Due to the low temperatures recovered with double bundled condenser and the load matching requirements to recover heat efficiently, double bundle condensers are rarely applied.

Marine Water Boxes An accessory for shell and tube HXs is the marine water box, which is a header assembly that allows mechanical cleaning of the tubes without disassembling the connecting piping. However, they add to first costs and pressure drop and, because mechanical cleaning is so seldom required (particularly on the CHW side), the future maintenance cost savings seldom justify the added first costs and pump energy costs. The labor cost of tube pull and cleaning can be reduced by using flanged or mechanical joints on the fittings at the chiller and condenser connections for ease of temporarily removing the piping.

Performance Characteristics and Efficiency Ratings Performance Issues There are a number of variables that determine the operational characteristics and energy performance of water chillers. A chiller is selected to meet a specific maximum capacity requirement at certain design conditions, to meet this capacity at specific (maximum) power draw, and to have specific part-load operating characteristics. To design chillers that meet the performance specifications, manufacturers of packaged water chillers must consider a very wide range of variables. These variables include the following: • •

Compressor design Internal refrigerant pressure drops

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Chapter 3 Chilled-Water Plant Equipment • • • • • • •

Heat gains: motors, oil pumps, casings Over/under-compression Motor efficiency Use of refrigerant economizers Surface area of evaporators/condensers Tube heat transfer coefficients: fouling, tube enhancement, velocity of fluids Refrigerant

Each design decision has first-cost implications. Because of this complexity, products on the market have a wide variety of performance characteristics.

Chiller Efficiency Ratings The efficiency of water chillers is characterized by the coefficient of performance (COP). The COP is the ratio of the rate of heat removal to the rate of energy input in consistent units for a complete refrigerating system or some specific portion of that system under designated operating conditions. The formula for COP is Net Useful Refrigerating Effect COP = ----------------------------------------------------------------------------------------------Energy Supplied from External Sources

(3-1)

The higher the number, the more energy efficient the machine. ANSI/ ASHRAE/IES Standard 90.1-2016 and California’s Title 24 energy standards (CBSC 2016) provide minimum energy efficiency standards for water chillers. The theoretical limit of efficiency is the Carnot efficiency: TE COP = ------------------TC – TE

(3-2)

where TE is the evaporation temperature and TC is the condensing temperature, both measured in absolute degrees (°R or K). So, for example, the theoretical maximum efficiency of a chiller operating at 40°F (500°R) evaporation temperature and 100°F (560°R) condensing temperature is 500 COP = ------------------------ = 8.3 560 – 500

(3-3)

Chiller efficiencies are also characterized in terms of kW/ton, which is essentially the inverse of COP (kW/ton = 3.517/COP) and is more commonly used in the U.S. than COP. The lower the kW/ton, the more energy efficient the machine. In the example in Equation 3-3, the theoretical lowest kW/ton at these conditions is 0.42. Equation 3-2 also demonstrates how reduced lift (difference between condenser and evaporator temperatures) improves efficiency. The closer the two temperatures are to each other, the higher the COP.


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Standard chiller ratings are based on Air-Conditioning, Heating, and Refrigeration Institute (AHRI) standard conditions, which set parameters for the rating capacity of different machines. These parameters are established in AHRI Standards 550/590 (2011) (vapor-compression chillers) and AHRI Standard 560 (2000) (absorption chillers). For water chillers the AHRI rating conditions are as listed in Table 3-3. Another energy efficiency rating is the integrated part-load value (IPLV). The IPLV is a single-number figure of merit based on part-load kW/ton. Partload efficiency for equipment is based on the weighted operation at various load capacities for the equipment. The equipment kW/ton is derived for 100%, 75%, 50%, and 25% loads (with consideration for condenser water relief) and is based on a weighted percentage of operational hours (assumed) at each condition. A weighted average is determined to express a single part-load/part-lift efficiency number. The weighting factors are as follows: 1% at 100% load, 42% at 75% load, 45% at 50% load, and 12% at 25% load. For water-cooled chillers, condenser water relief assumes that the temperature of the water entering the condenser declines as a straight line from 85°F at 100% load to 65°F at 50% load and below, implying a correlation between weather and cooling load. This represents a 4°F decline for a 10% change in load. The nonstandard part-load value (NPLV) is another useful energy efficiency rating. This is used to customize the IPLV when some value in the IPLV calculation is different than standard, such as using 42°F leaving chilled water in lieu of 44°F. While IPLV and NPLV are useful energy performance indicators for individual chillers, particularly for equipment efficiency standards and regulations, the large majority of chillers are installed in multiple-chiller plants. Individual chillers operating in a multiple-chiller plant may be more heavily loaded than single chillers within single-chiller systems and operate at different condenser water temperatures than those assumed. When evaluating a multiple-chiller plant, a comprehensive analysis must be used to predict the CHW system performance. This is discussed in detail in Chapter 6. Table 3-3

AHRI 550/590-2011 and 560-2000 Rating Conditions for Water Chillers

Leaving CHW Temperature

44°F

Evaporator Water Flow Rate

2.4 gpm/ton

Entering Condenser Water Temperature

85°F

Condenser Water Flow Rate (Electric)

3.0 gpm/ton

Condenser Water Flow Rate (Absorber)

3.6 gpm/ton (single stage) 4.5 gpm/ton (two stage)

Ambient Air (for Air-Cooled)

95°F

Fouling Factors

0.00010 h·ft2·°F/Btu(evaporator) 0.00025 h·ft2·°F/Btu (condenser)

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Scroll Chillers These chillers are widely used in tonnage ranges from 50 to 230 tons, although they are available up to much larger sizes (400 tons and up). Capacity modulation is typically achieved through staging of multiple compressors that are grouped (piped in parallel) in several circuits. This creates some redundancy should a compressor fail. Some manufacturers offer variable-capacity (also called digital) scroll compressors that can unload down to 10% using a pulse width modulation (PWM) approach opening and closing scroll plates. A growing number offer variable-speed scroll compressors, which reduce both minimum turndown ratio and energy use. As positive displacement machines, they retain near-full cooling capacity even when operated at temperatures above the design conditions, and they are, therefore, very suitable for aircooled applications. For the same reason, they are also suitable for use as heat recovery machines.

Screw Chillers Rotary screw chillers are also positive displacement machines. Like scroll chillers, they are particularly suitable as air-cooled chillers but are popular in both air- and water-cooled configurations. Screw chillers tend to be most cost competitive in the 100 to 300 ton range, although they are available in a wider range of capacities. In the low capacities, they compete less successfully with scroll chillers, and, in the high capacities, centrifugal chillers tend to be more cost-effective. Most screw chillers have excellent turndown capability. Some chillers incorporate multiple compressors. This provides additional efficiency advantages during part load and allows unloading below 10%. Screw chillers are inherently more efficient than scroll compressors because they incorporate refrigerant economizers (discussed in the Performance Issues section). They have very few moving parts and have balanced forces on the main bearings. As a result, these machines are very reliable and generally have the lowest maintenance costs. Screw machines are usually controlled with a slide valve and are fully modulating, although some less expensive models use multiple discrete injection ports with stepped controls. Variable-speed control is also now being offered on singlecompressor machines and on one or more compressors on dual-compressor machines. Screw chillers tend to be noisy at design conditions due to the high speed of operation. The variable-speed-driven screws offer significant acoustical benefits at low loads and have less wear and tear on the bearings.

Centrifugal Chillers Centrifugal chillers have the highest efficiency ratings of all the chillers. They are available in sizes from 80 to 10,000 tons, but the most common factory-built sizes are from 200 to 3000 tons. Above 3000 tons, they are generally field erected. They are available in both air-cooled and water-cooled versions, but, because of very low kW/tons and very high initial cost, air-cooled centrifugal chillers are uncommon.


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Centrifugal chillers are controlled with inlet guide vanes, which allow for full modulation to as low as 10% to 15% capacity (with condenser water relief). Note that chiller efficiency drops off severely at low loads. VFDs can be added to enhance the part-load/part-lift operation characteristics and are usually cost-effective when evaluated through life-cycle cost analysis (LCCA). In addition to the energy savings, centrifugal chillers with VFDs are quieter at part load. Because of the economics of centrifugal chiller manufacturing, there are product differences among all the major manufacturers. There are countless pros and cons to the various features of these products; the following discussion presents some of the main differences.

Direct Drive Versus Gear Drive Direct-drive chillers typically operate at 3600 rpm. Gears allow impellers to rotate at speeds up to 35,000 rpm. This allows smaller impellers to be used, reducing the machine’s size and first cost. There is an efficiency loss in the gear train of 1.5% to 2%. Also, the gears have additional bearings and require regular maintenance, whereas direct-drive machines do not. The proper selection of impeller diameter and gear ratio allows the machines to be selected very near their highest performance level or sweet spot, whereas the direct-drive machines, because of limited impeller diameter choices, sometimes are selected several efficiency points away from their sweet spot. Direct-drive machines sometimes have multiple stages (more than one impeller). In this situation, economizers can be added to enhance the energy performance of the machine.

Open Drive Versus Hermetic Open-drive machines have the motor located outside the casing. Efficiency ratings do not include motor losses (4% to 5% on larger machines). The heat from an open-drive motor must be removed from the machine room, which usually requires additional mechanical cooling. Open-drive machines have seals that leak and are subject to failure. On high-pressure machines refrigerant can leak out with dire consequences, and on low-pressure machines air can leak in, causing more purge compressor time and loss of efficiency. In the event of a catastrophic motor failure, an open-drive machine can be repaired and placed back in service relatively easily, whereas a hermetic machine will require significantly more attention. Motor failures in hermetic machines are almost always catastrophic. Fortunately, motor failures are rare. Hermetic centrifugal chillers have the motor totally enclosed within the chiller casing. The motors are kept clean and are cooled by the refrigerant stream. Hermetic machines have a lower likelihood of refrigerant leakage than open-drive machines.

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Fixed Orifice Versus Variable Orifice or Float Valve When a fixed orifice is used as the thermal expansion device, a minimum DP must be held between the condenser and evaporator to ensure proper refrigerant flow. This may limit the degree of condenser water relief that can be obtained during off-peak time, with the consequence that the machine will not have as good a part-load performance as a machine with a variable orifice or float valve.

Oil Return Chillers may have an oil pump, but most require a minimum DP between the condenser and evaporator to be maintained to ensure proper refrigerant flow. This condition is often exerted by the manufacturer requiring a minimum 15°F to 25°F between the leaving CHW temperature and the leaving condenser water temperature (an indicator of refrigerant lift and often called lift). Using an oil pump can reduce the minimum lift to about 5.5°C, resulting in improved chiller efficiencies at low condenser water temperatures, particularly with variable-speed chillers. With oil-free chillers, the minimum lift need only be a few degrees and, in some chillers, may be zero or even negative.

Absorption Chillers Absorption chillers can be either single or double effect. Single-effect chillers have COPs of 0.60 to 0.70 and double-effect chillers have COPs of 0.92 to 1.20. Because the double-effect machines are 50% to 100% more efficient than the single-effect chillers, there is little doubt about which to choose if absorption is being considered. Single-effect chillers are beneficial where waste steam is available or where hot-water temperatures are not high enough to fuel a double-effect absorption chiller. Triple- and quadruple-effect machines are being developed but are not yet on the market. Absorption machines can be direct or indirect fired. Direct-fired machines have the advantage that they can also be used to heat the building and/or domestic hot water. If a direct-fired absorption machine is also used as a heater, the avoided cost of a separate boiler and boiler room (space) may help offset some of the added cost of the machine. Sizes for absorption chillers range from 100 to 1700 tons. Absorption machines typically cost two or more times that of an electric-driven chiller. Because of absorption chillers’ low, the heat rejection system must be about 50% larger than with a compression chiller plant, increasing the cost of condenser water pumps, piping, and cooling towers. Commercial absorption chillers have additional operating disadvantages that should be considered: •

They cannot produce water at temperatures as low as those of electric chillers. The minimum CHW supply temperature is typically 43°F or 44°F, which limits their use with thermal energy storage (TES) systems—certainly ice-storage systems but also CHW storage tanks where 39°F water is desired because water at that temperature has the lowest density, enhancing tank stratification and increasing storage capacity.

Fundamentals of Design and Control of Central Chilled-Water Plants I-P •

•

• •

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They take longer to start up and to shut down, thus requiring longer time between cycles than electric chillers. The slow start-up is in part due to the capacitance of their refrigerant. The cycling is due to the chemistry. As CHW flow is maintained through the chiller during start-up and shut-down periods, at lower or no produced cooling capacity, maintenance of system CHW supply temperature can be an issue. This limits their use in plants such as those for data centers, where rapid deployment is an issue. They cannot abide low flows or temperatures on the condenser water side. This limitation can hamper the performance of plants with water-side economizers (WSEs) and hybrid plants where variable-speed-driven electric centrifugal chillers might be optimized by low condenser flows and temperatures at part-load conditions. Primary/secondary condenser water pumping may be required for most efficient plant operation. They are significantly larger than electric chillers and require larger towers. They may not last as long as electric chillers and are subject to failure if not properly maintained. The absorption chiller’s chemistry is corrosive and will destroy the chiller if inhibitors are not properly maintained.

Because of these operating disadvantages, much higher first costs, and much higher operating costs, absorption chillers are seldom the best choice. However, there are a few applications where an absorption chiller may make sense: • • • •

Very high electrical costs, including demand and low natural gas cost Electrical service not available or too costly to upgrade Low-cost gas from landfill, solar, or biomass available Waste or very low-cost steam or hot water available (e.g., from a cogeneration plant or solar thermal panels)

Turbine-Driven and Engine-Driven Chillers While not a large segment of the chiller market, turbine-driven and enginedriven chillers are sometimes economically viable. Both use the same vapor compression cycle as an electric machine except they use either a reciprocating engine or a gas- or steam-driven turbine as the prime mover. For larger applications, the refrigeration component is usually an open-screw or centrifugal chiller. Because these chillers use variable-speed technology, the part-load characteristics are comparable to variable-speed electric chillers. Engines use natural gas or diesel fuel. Some hybrid units have both an engine and an electric motor so that the fuel may be switched depending on the utility rates at the time. Engines require heat rejection from the jacket water. Heat can be rejected out the cooling tower (through a HX) or smaller units can be air cooled. The jacket water heat is available for heat recovery of domestic water or other loads occurring at the same time as the engine runs. Heat recovery water temperatures at 180°F to 200°F are easily produced, availing heat recovery to a wider range of loads, which if amply available can significantly impact the economics. Engines need additional maintenance, with top-end overhauls required every 12,000 hours and complete overhauls at 35,000 hours. Reciprocating engines are

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Chapter 3 Chilled-Water Plant Equipment much louder than electric-driven or absorption machines and may require special enclosures or acoustical abatement. Natural gas and steam turbines are a very small part of the market and are used in very large plants (up to 10,000 tons). As there are limited manufacturers of these products, care is required in procuring them. A flat specification for a turbine-driven chiller on a large plant can give a single manufacturer an unfair advantage on bidding the entire plant including the turbine and electric chillers.

Heat Rejection The primary means of heat rejection in the HVAC industry are cooling towers, air-cooled refrigerant condensers, and evaporative refrigerant condensers.

Cooling Towers The conversion of liquid water to a gaseous phase requires an amount of energy called the latent heat of vaporization. Cooling towers use the internal heat from water to vaporize the water in a near-adiabatic saturation process. A cooling tower’s purpose is to expose as much water surface area to air as possible to promote the evaporation of the water. In a cooling tower, approximately 1% of the total flow is evaporated for each 12.5°F temperature change. There are two important terms used in the discussion of cooling towers: • •

Range: The temperature difference between the water entering the cooling tower and the temperature leaving the tower Approach: The temperature difference between the water leaving the cooling tower and the ambient wet-bulb temperature

The performance of a cooling tower is a function of the ambient wet-bulb temperature, entering water temperature, airflow and water flow. The dry-bulb temperature has an insignificant effect on the performance of a cooling tower. Nominal cooling tower tons are the capacity based on a 3 gpm flow, 95°F entering water temperature, 85°F leaving water temperature, and 78°F entering wet-bulb temperature. For these conditions the range is 10°F (95–85) and the approach is 7°F (85–78). Significant confusion in the industry has been caused because cooling tower tons and chiller tons use the same units (tons) but have different values; accordingly, the use of the term cooling tower tons has been waning and is no longer common. This is beneficial because the heat rejection capacity of a cooling tower varies widely depending on flow and temperatures, so the term was also misleading.

Types of Cooling Towers Cooling towers come in a variety of shapes and configurations. A direct tower is one in which the fluid being cooled is in direct contact with the air. This is also known as an open tower. An indirect tower is one in which the fluid being cooled is contained within an HX or coil and the evaporating water cascades over the outside of the tubes. This is also known as a closed-circuit cooling tower or a fluid cooler.


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The tower airflow can be driven by a fan (mechanical draft) or can be induced by a high-pressure water spray, although the spray type is rarely used. The mechanical draft units can blow the air through the tower (forced draft) or can pull the air through the tower (induced draft). The water invariably flows vertically from the top down, but the air can be moved horizontally through the water (cross flow) or can be drawn vertically upward against the flow (counterflow). Water-to-air surface area is increased by using fill. Fill can be splash type or film type. Film-type fill is most commonly used and consists of closely spaced sheets of corrugated polyvinyl chloride (PVC) arranged vertically. Splash-type fill uses bars to break up the water as it cascades through staggered rows. Typically in the HVAC industry, cooling towers are packaged towers that are factory fabricated and shipped intact to a site. Field-erected towers mostly serve very large chiller plants and industrial/utility projects. When aesthetics play a role in the selection of the type of tower, custom-designed field-erected cooling towers are sometimes used. In these towers, the splash-type fill is often made of ceramic or concrete blocks. The following is a discussion of the most common types of cooling towers encountered in HVAC CHW plants.

Forced-Draft Cooling Towers Forced-draft towers (Figure 3-12) can be of the cross-flow or counterflow type, with axial or centrifugal fans. Forward-curved centrifugal fans are commonly used in forced-draft cooling towers. The primary advantage of a centrifugal fan is that it has capability to overcome high static pressures that might be encountered if the tower were located within a building or if sound attenuators were located on the inlet and/or outlet of the tower to reduce ambient noise, as might be needed for towers located in noise-sensitive residential areas. Crossflow towers with centrifugal fans are also used where low-profile towers are needed. These towers are relatively quieter than other types of towers in the

Figure 3-12

Forced-draft cooling tower.

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Chapter 3 Chilled-Water Plant Equipment high-frequency bands. However, towers with centrifugal fans are not energy efficient. The energy to operate this tower is more than twice that required for a tower with an axial fan. Another disadvantage of the forced-draft tower is that, because of low discharge air velocities, they are more susceptible to recirculation than induced-draft towers. This is discussed in further detail in the section Induced-Draft Cooling Towers.

Induced-Draft Cooling Towers The induced-draft tower is by far the most widely used cooling tower available in the HVAC industry. These towers can be cross flow or counterflow and use axial fans (Figure 3-13). Most field-erected cooling towers are the induced-draft type. Because the air discharges at a high velocity, they are not as susceptible to recirculation as forced-draft towers. The large blades of

Figure 3-13

Induced-draft cooling towers.


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an axial fan can create noise that is difficult to attenuate and, depending on the location on the property, could cause problems. Many manufacturers offer low-sound blades that reduce noise levels but often reduce airflow rates and efficiency as well. Low-sound propeller fan towers can be as quiet as centrifugal fan towers without sound attenuators.

Closed-Circuit Fluid Coolers With a closed-circuit fluid cooler, the fluid is located within a coil (rows of tubes) rather than being open to the environment. A pump draws water from a sump and delivers it to a header where the water is sprayed over the coil. With proper initial chemical treatment, the fluid does not foul the condenser tubes, so chiller maintenance is reduced and energy efficiency is always at peak. Because of the additional heat exchange process, for the same capacity as an open tower, a closed-circuit fluid cooler is physically much larger and significantly more expensive than conventional open towers.

Cooling-Tower Performance Given a fan selection, flow rate, range, entering wet-bulb temperature, and fill volume, cooling towers have a wide range of performance characteristics. Typical performance curves (see Figure 3-14) show the relationship between these variables at different operating conditions. In reviewing the typical performance curve, one feature not well understood is that for a given range, as the entering wet-bulb temperature decreases, the approach increases. As entering wet-bulb temperature drops, it is likely that the load (range) will also decrease for the same flow rate. Yet even at this condition, the approach usually increases over design condition. This is particularly important when considering the selection of cooling towers for use with WSEs. To obtain the maximum effectiveness at low wet-bulb temperatures, a cooling tower used in a WSE system should often be larger (selected for lower approach) than a tower selected just for maximum peak duty. Tower efficiency is defined in ANSI/ASHRAE/IES Standard 90.1 as the maximum flow rate (gpm) the tower can cool from 95°F to 85°F at 75°F entering wet-bulb temperature, divided by the motor horsepower (2016b). Typical efficiencies range from 20 to 50 gpm/hp for centrifugal fan towers and from 40 to 120 gpm/hp for propeller fan towers. Higher-efficiency towers usually are physically larger (more fill) with smaller fan motors operating at lower speeds. Cooling towers are relatively inexpensive when compared to the total cost of a chiller plant and incremental increases in tower efficiency can be purchased at a relatively low cost. More efficient towers also tend to be quieter due to lower fan speeds. Optimum tower efficiency for various applications is discussed further in Chapter 5.

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Figure 3-14

Typical tower performance at constant heat rejection load.

Application Issues Siting and Recirculation When the saturated air leaving the cooling tower is drawn back into the intake of the tower, the recirculation that occurs degrades the performance of the tower. Wind forces create a low-pressure zone on the downwind (leeward) side of the tower that causes this phenomenon. Wind forces on the leeward side of the building can also create downward air movement. When cooling towers are located in such a way that the discharge from one tower is directed into the intake of an adjacent tower, recirculation can also occur. Recirculation is a greater problem when cooling towers are confined within pits or have screen walls surrounding them, typically to hide them for architectural reasons. If the


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tower is sited in a pit or well, it is essential to follow the tower manufacturer’s guidelines to determine the proper location of the outlet and minimum clearances for the air intake. Typically the manufacturer will require that the tower outlet be no less than flush with the height of the walls of the enclosure. Crossflow side discharge towers should never be used in pits or wells, because recirculation is almost assured. The Cooling Technology Institute (CTI) recommends that recirculation effects be accounted for in the selection of the tower. Their tests show that as much as 8% of the discharge air could be recirculated back into the intake and that the worst conditions occur with winds of 8 to 10 mph. Where recirculation is a concern, a rule of thumb is that the entering wet-bulb temperature used to select the tower should be increased by 1°F above the ambient temperature to account for recirculation effects. Tower height is often a consideration in selection due to the architectural impact. The lowest profile towers are usually blow-through centrifugal type, but they are more expensive and less efficient than other options and thus should be used only in extreme cases. Double inlet cross-flow induced-draft towers also have low profiles, but they have a large footprint and thus require more plan space. Counterflow-induced-draft towers have the smallest footprint but tend to require the most height. However, sometimes the height of these towers is an advantage when the tower is located in a well and the height of the well is determined, for example, by other tall penthouse elements, such as a traction elevator machine room. Because the tower discharge should be at least flush with the walls of the well, tall cooling towers may avoid the need for and cost of high support pedestals (which also make maintenance access more difficult) or fan discharge duct extensions.

Capacity Control Like most air-conditioning equipment, cooling towers are selected to deliver peak capacity at design weather conditions, but most of the time they operate at well less than peak capacity. There are a number of methods used to control the temperature of the water leaving the cooling tower, including the following: •

•

On/off: Cycling fans is a viable method but leads to increased wear on belts and gear drives (if used) and can lead to premature motor failure. It is also the least energy-efficient control option and can result in large variations in condenser water temperature, which can cause unstable chiller operation. Cycling is therefore the least favorable method of controlling temperature. Two-speed motors: Multiple wound motors or reduced-voltage starters can be used to change the speed of the fan for capacity control. This method is cost-effective and well proven. Because of basic fan laws, there are significant energy savings when the fans are run at low speed. One pitfall with two-speed fans is that when switching from high to low speed, the fan rpm must reduce to below low speed before energizing the low-speed step to avoid motor overload trips.

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Chapter 3 Chilled-Water Plant Equipment •

•

•

Pony motors: This is another version of the two-speed approach. A second, smaller motor is belted to a common fan shaft. For low-speed operation, the larger motor is de-energized and the smaller motor energized for a lower speed. Again, when going from high speed to low speed, the fan must slow down sufficiently before energizing the low-speed motor. This option has fallen out of favor as VFD costs have fallen, making it more expensive than alternative options. It is also typically an option only available on centrifugal fan blow-through towers. Modulating discharge dampers: Used exclusively with centrifugal fans, discharge dampers built into the fan scroll can be modulated for capacity control. Although it does save energy by riding the fan curve, other methods of capacity control provide better energy savings results at lower costs. Hence, this option is seldom used in modern towers. Variable-frequency drive (VFD): Adjustable-frequency VFDs can be added to the motors for speed control. This method provides the tightest temperature control performance and is the most energy-efficient method. One pitfall to avoid with VFDs is to not run the fans at critical speeds, which are speeds that result in resonant frequency vibrations and can severely damage the fans. Critical speeds are typically determined empirically postinstallation. Minimum fan speeds are discussed in the sections Belt Versus Gear Drive and VFD Accessories and Application Considerations: Minimum Speed Setpoint. Selecting the best control option is discussed in Chapter 5.

Chemical Treatment and Cleaning Cooling towers are notorious for having high maintenance costs. Unfortunately, cooling towers are very good air scrubbers. A 200 ton open cooling tower can remove 600 lb of particulate matter in 100 hours of operation. Because tower water is open to the atmosphere, the water is oxygen saturated, which can cause corrosion in the tower and associated piping. Towers evaporate water, leaving behind dissolved solids such as calcium carbonate that can precipitate out on piping and condenser tubes and decrease heat transfer and energy efficiency. To avoid these problems, towers must have water treatment systems and should be inspected and cleaned regularly. It is best to contract with a cooling tower water treatment specialist to assist in determining the appropriate water treatment program and to provide regular monitoring. The following are some of the strategies to consider in a good chemical treatment program: •

Blowdown: To control dissolved solids, a portion of the flow of the tower must be bled from the system. Depending on the quality of the water (e.g., silica and other dissolved solid content) and water treatment approach (chemical versus nonchemical), the cycles of concentration of dissolved solids (ratio of blowdown to incoming water concentration) can vary from


•

•

•

47

2 to 10, equal to a blowdown rate of about 0.06 to 0.03gpm/ton, respectively. This is typically controlled by sensing water conductivity (mhos) and bleeding water to maintain a certain conductivity set point. This set point should be determined by a water treatment specialist. Corrosion control: Corrosion can be caused by high oxygen content, carbon dioxide (carbonic acid), low pH, or high dissolved solids. Regular injection of corrosion inhibitor chemicals is a common solution. Biological growth: Slime and algae are handled with shock treatments of chlorine or chlorine compounds. It is best to alternate between two different compounds so that organisms do not develop a tolerance to the chemicals. At least one of the biocides should be effective against Legionella, the organism responsible for Legionnaires’ disease. Scale prevention: Control of the pH (acid levels) is extremely important in areas with very hard water. Usually acids, inorganic phosphates, or similar compounds are used to control pH. Blowdown is usually effective in areas with neutral or soft water.

Technologies that purport to eliminate the need for inhibitors and/or biocides have come and gone over the years, usually with mixed or poor results. One promising technology employs pulsed electromagnetism to remove dissolved solids and inhibit biological growth. The appropriate use of these systems depends on the local makeup water quality and other local conditions. A local water treatment specialist should be consulted. To get an unbiased recommendation, the specialist should represent or operate both chemical and nonchemical water treatment systems. Another water treatment option is particulate filtration. The filter can be mounted in-line with the primary condenser water flow but more commonly it is mounted in a small sidestream configuration with its own pump to reduce the energy penalty of the filter pressure drop. Sidestream filters generally circulate about 10% of the system flow. Common filters are centrifugal separators or sand filters; the latter remove finer particles but require higher pump energy and more frequent backwash and associated water use. A common accessory is a basin eductor distribution system: the sidestream filter pump draws water from the basin, pumps it through the filter, and then discharges it through an array of eductors mounted on the tower basin designed to stir up settling particles so they can be effectively removed by the filter. The advantage of this design is that it reduces tower maintenance costs by reducing how frequently the basin must be isolated and cleaned of dirt that precipitates out in the basin. But the first costs and energy costs are high—in fact it is not unusual for the basin pump to use more energy than the cooling tower fan. Claims that these filters reduce fouling of condenser water tubes and thus improve energy efficiency have been made by filter manufacturers but to date have never been demonstrated with unbiased research. Sidestream filters may be desirable where the tower is located adjacent to ambient air with high particle concentrations, such as near farming, and possibly where nonchemical water treatment systems are used, because they are designed to have particles coagulate and

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Chapter 3 Chilled-Water Plant Equipment precipitate out in the tower basin. To limit energy use, filters should be operated on a time schedule for a few hours per day on days when the towers have operated, preferably during off-peak energy rate periods; the exact time period required must be determined empirically by the plant operator.

Flow Limits For a given design of a cooling tower, the manufacturer will specify maximum and minimum flow rates through the tower. The maximum flow is usually based on the capacity of the water distribution system within the tower to adequately distribute the water over the fill. Too much flow will overflow the tower distribution pans and create a situation where the tower does not get adequate mixing of air and water to perform properly. Below the minimum flow rate, the water may not distribute evenly across the entire fill face. This creates voids where there is no water in the fill. At the boundary where the wet and dry portions of the fill meet, dissolved solids can drop out of solution and plate out on the fill. Prolonged operation below the minimum water flow can thus cause significant scaling to occur. However, this may not be a problem in areas with excellent water quality. ANSI/ASHRAE/IES Standard 90.1 (2016b) and California’s Title 24 (CBSC 2016) require that in plants with multiple condenser water pumps, the tower minimum flow rate must be low enough to handle flow from the smallest pump down to 50% of the total design flow rate. This is to allow more cells to be active even when the plant is at part load; as discussed in Chapters 4, 5, and 6, energy efficiency is maximized by running as many tower cells as possible. Low minimum flow is generally achieved with weir dams in the distribution and/or adjusting distribution nozzle type and size.

Cooling Tower Accessories and Options The following is a list of accessories and options that should be considered when purchasing a cooling tower: •

Vibration switch: This stops the fan if vibration exceeds a certain limit. It could prevent catastrophic failure of the fan. Codes in some areas require the installation of a vibration switch.

•

Side inlet and internal distribution: Cross-flow towers are often supplied with field-erected overhead piping to the gravity basins. But this distribution can be factory installed within the tower to reduce overall height and substantially reduce installed costs. This is because field-erected overhead piping cannot be supported off of the towers and thus must have expensive field-installed support frames connected to adjacent structures. It is also not uncommon for overhead piping to be self-vented and drain when pumps are stopped, sometimes causing cold-water basin overflow. Air locks can also form that starve one cell while overflowing the hot-water basins of the other(s). This is a highly recommended option.


•

•

•

•

49

Ladders and railings: Ladders to the tower supply basins and associated safety railings may be a convenience to maintenance personnel, although they add to the cost and sometimes to the height of the tower. Using the side inlet and internal distribution piping option eliminates the need for accessing supply shutoff valves on the top of the tower, reducing the need for this option for cross-flow towers. For tall counterflow towers with external motors, ladders and railings are recommended for safe access to motor and drive maintenance. Stainless steel or protective coatings: In most applications, using stainless steel or coated galvanized steel hot- and cold-water basins is recommended. The entire tower can also be built from these materials but at very high cost, usually justified only where future tower replacement will be very expensive. Basin heaters: In freezing climates where the tower cannot easily be winterized (drained) during the cold-weather season, thermostatically controlled electric basin heaters can be provided. Typically in these instances piping must also be protected with heat tracing and insulation. Electronic fill controls: The standard water level controller is a float valve, much like a toilet fill valve. Electronic fill controls using an electronic level sensor controlling an electric motorized fill valve will improve reliability. (An even better approach is to use a separate level sensor mounted in an equalizer standpipe controlling a central makeup water valve located inside the chiller room. See Chapter 4 for design details.) Calibrated balancing valves (CBVs): Balancing valves are almost never needed on cooling towers. Most plants are near self-balancing simply because their compact size does not result in large differences in pressure drop across each tower circuit. Cooling tower performance is also very forgiving to flow imbalance: the cells with excess flow will create warmer water and the ones with low flow will create colder water, but when they are mixed, the resulting temperature is almost exactly the same as it would be if the cells had equal flow.

Choosing the Type of Cooling Tower When choosing which cooling tower is most appropriate for a particular application, the following factors should be considered.

Packaged Versus Field-Erected Cooling Towers The type of cooling tower selected will be determined to a great extent by the required capacity. Packaged cooling towers are manufactured to be cost-effective and to ship on standard-size carriers. Typically, a single cell of a packaged tower will handle a maximum cooling capacity of 650 to 1000 tons at nominal conditions. Larger plants will require multiple cells. If the CHW plant is very large, field-erected cooling tower, cells may be more cost-effective than a packaged cooling tower. Field-erected cooling towers also offer a greater degree of energy efficiency because the design flexibility makes it

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Chapter 3 Chilled-Water Plant Equipment possible to match lower-horsepower fans with larger fill volume. Although fielderected cooling towers offer greater flexibility when the site has physical constraints, they may have longer procurement times than packaged cooling towers and they can be much more expensive than multiple packaged towers.

Open Versus Closed Circuit Open-circuit cooling towers are the most prevalent type of cooling tower used in the HVAC industry. Closed-circuit cooling towers (also called fluid coolers) are at a disadvantage due to their higher first cost, additional energy cost, and larger physical size. Closed-circuit cooling towers are therefore only used where • • • •

the condenser water pumps must be located remotely from the tower, the cooling tower is located below the condensers, it is necessary to keep the condenser water free from contamination with dirt or impurities due to poor local water quality, or the condenser water is mixed with fluids from other closed systems (like the chilled water or hot water) such as on hydronic heat pump systems.

Forced Versus Induced Draft Most axial fan towers are induced draft for the same reason that most air handlers are draw through: the air distribution through the fill is inherently more uniform when air is drawn through it rather than blown through it. Blowthrough designs benefit from the fact that motors and fans are not in the wet, more corrosive atmosphere of the tower effluent, but the cost benefits of the induced-draft design generally outweigh these considerations. Induced-draft towers also are less likely to recirculate discharge air due to the higher velocity.

Centrifugal Versus Axial Fans Centrifugal fans are significantly less energy efficient in cooling tower applications than axial fans and generally more expensive. The use of centrifugal fans in cooling towers should be limited to situations where • •

low-profile towers are absolutely required architecturally or sound attenuators or other acoustical considerations make a centrifugal fan the only choice.

Cross Flow Versus Counterflow Cross-flow and counterflow induced-draft towers are the most common cooling towers used in CHW plants. Advantages and disadvantages include the following: •

Minimum flow. Cross-flow towers can typically achieve lower minimum flow rates than counterflow towers because they use gravity-fed distribution pans that can be readily modified to provide minimum flow rates as low as


•

•

•

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30% to 50% using weir dams and flow nozzle extensions. Counterflow towers use pressurized nozzles that typically have much more limited turndown; to meet the 50% turndown limit required by California’s Title 24 (CBSC 2016) and ANSI/ASHRAE/IES Standard 90.1, special low-turndown nozzles must be specified. (Note that new nozzle designs available from some counterflow tower manufacturers can provide performance comparable to counterflow towers.) The lower minimum flow rate can improve efficiency and eliminate the need for expensive cell automatic isolation valves. This is discussed in more detail in Chapter 5. Maintenance access. Cross-flow towers generally are easier to maintain. The distribution pans are on the top where they can be easily accessed and cleaned, whereas access to counterflow spray nozzles requires removal of mist eliminators. Efficiency and cost. The counterflow design is usually slightly more efficient for the same cost, or said another way, slightly less expensive for towers with similar efficiency. Height and footprint. As discussed above under Siting and Recirculation, counterflow towers tend to be tall but with a small footprint while crossflow towers are the opposite. Either can be an advantage, depending on the architectural constraints. This factor alone can often drive the selection between the two types.

Belt Versus Gear Drive Induced-draft axial fans can have either a belt drive or a direct-drive connection to the motor. Direct-drive fans use gear reducers to maintain the low speeds of the fan. Gear drives cost more but reduce maintenance frequency and may reduce lifetime maintenance costs compared to belt drives. But the increasing popularity of VFDs has also increased the popularity of belt drives; the belts last longer due to the soft start feature of VFDs, and belt drives allow near-zero minimum speeds while many gear drives require on the order of 20% minimum speed to ensure adequate lubrication. Lower minimum speed reduces fan cycling, which reduces wear and tear, reduces noise levels and abrupt changes in noise levels, reduces condenser water temperature fluctuations, and (slightly) improves energy efficiency.

Air-Cooled Refrigerant Condensers Types Another method of heat rejection commonly used in chiller plants is the aircooled refrigerant condenser. This can be coupled with the compressor and evaporator in a packaged air-cooled chiller (Figure 3-15) or can be located remotely. Remote air-cooled condensers are usually located outdoors and have propeller fans and finned refrigerant coils housed in a weatherproof casing. Some remote aircooled condensers have centrifugal fans and finned refrigerant coils and are installed indoors in what amounts to an AHU. Indoor condensers are only used on

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Figure 3-15

Packaged air-cooled chiller.

small chillers and will not be discussed further here. Air-cooled condensers, whether remote or packaged within an air-cooled chiller, normally operate with a temperature difference between the refrigerant and the ambient air of 10°F to 30°F, with fan power consumption of less than 0.08 hp/ton cooling. Maximum size for remote air-cooled refrigerant condensers is about 500 tons, with a more common maximum of 250 ton. Air-cooled chillers are available up to about 400 tons. There are a number of reasons air-cooled chillers are used. These include • • • •

Water shortages or quality problems Lower cost than water-cooled equipment With packaged air-cooled chillers, no need for machine rooms with safety monitoring, venting, etc. Less maintenance required than cooling towers

Air-cooled chillers are not as energy efficient as water-cooled chillers in most applications. When comparing the energy efficiency of air-cooled to water-cooled chillers, care must be taken to include in the water-cooled chiller the energy consumed by the condenser water pump and cooling tower. Aircooled chillers can have very good part-load performance if properly controlled; as the outdoor air temperature drops, the kW/ton improves significantly if condensing pressure is allowed to float downward. Remote air-cooled refrigerant condensers in CHW plants are very seldom used because of the physical size for the larger capacity machines. Air-cooled chillers are more often used in smaller chiller plants, generally below 300 tons, as space, water treatment, and the additional maintenance cost associated with cooling towers or evaporative condensers outweighs the energy benefit. California’s Title 24 limits the size of air-cooled chiller plants to 300 tons (CBSC 2016). ANSI/ASHRAE/IES Standard 90.1 has no current limit on the use of air-cooled chillers. Air-cooled chillers require little maintenance but they do need to have coils cleaned regularly, they require standard lubrication, and the refrigerant charge needs to be periodically checked. If leaves from trees or other debris become a problem, permanent air filters are available to protect the coils. However, air filters slightly degrade system performance and require additional maintenance.


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As with siting cooling towers, air-cooled chillers can potentially recirculate the warm discharge air, especially when multiple condensers are located adjacent to one another or condensers are located within a pit or screen wall. Consult the manufacturer’s location guidelines for multiple machines or pit locations.

Controls When air-cooled condensers operate, typically the fan runs continually in conjunction with the compressor. When the outdoor temperature falls, it is possible to decrease the liquid refrigerant pressure too much to adequately overcome the thermal expansion valve (TXV) pressure drop. In this case, controls are required to limit the heat rejection. These controls include the following: •

• • •

Flooded coil: Control valves back up liquid refrigerant into the condenser to limit the heat transfer surface. This requires a receiver and a large refrigerant charge. Fan cycling: Usually need multiple fans with one or more cycling on and off to maintain minimum head pressure. Dampers: Discharge dampers on condenser fan restrict airflow. This option has become less popular as VFD costs have fallen. Variable-speed fans: Fan speed modulates airflow.

For systems not intended to run at cold temperatures (less than 40°F), fan cycling is usually the most appropriate choice for control. For systems intended to run at temperatures down to 0°F, fan speed control is the most common and most efficient.

Evaporative Condensers Evaporative condensers (Figure 3-16) are similar to closed-circuit fluid coolers but with refrigerant rather than closed-circuit water in the HX tubes. A pump draws water from a sump and sprays it on the outside of a coil. Air is blown (drawn)

Figure 3-16

Evaporative condenser.

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Chapter 3 Chilled-Water Plant Equipment across the coil and some of the water evaporates, causing heat transfer. The hot gas from the compressor condenses inside the tubes. Evaporative condensers are primarily used with DX air-conditioning units, although a few manufacturers produce small packaged water chillers with evaporative condensers. The effectiveness of the evaporation of the water and the refrigerant in the heat transfer process means that for a given load, evaporative condensers can have the smallest footprint of any heat rejection method. The evaporative condenser causes lower condensing temperatures and, as a result, is far more efficient than air-cooled condensing. Maintenance and control of evaporative condensers is similar to that of a closed-circuit fluid cooler. As with cooling towers, the style of the tower can significantly impact fan power.

Pumps In the HVAC industry, most pumps are single-stage (one impeller) centrifugal pumps that have either a single inlet (end suction) or a double inlet (double suction). Vertical turbine pumps are sometimes used in a cooling tower sump application. Most pumps in the HVAC industry are bronze fitted, meaning they have a bronze impeller and wear rings, a bronze or stainless steel shaft sleeve, a stainless steel shaft, and a cast iron casing. Centrifugal pumps come with mechanical seals (most common) or packing gland seals. Packing gland seals are sometimes (but infrequently) used in condenser water systems, where an accumulation of dirt can damage mechanical seals.

Pump Types The following is a brief discussion of the various types of pumps used in a CHW plant.

Single (End) Suction Single- or end-suction pumps can be either flexible coupled (also known as base mounted [Figure 3-17]) or close coupled (Figure 3-18). Close-coupled pumps use a special motor that has an extended shaft to which the pump impeller is directly connected. The motor and pump cannot be misaligned, and they take up

Figure 3-17

Base-mounted end-suction pump.


Figure 3-18

Close-coupled end-suction pump.

Figure 3-19

In-line pump.

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less floor space than flexible coupled pumps. However, replacement motors can have a long lead time and can be difficult to obtain after a breakdown, especially for larger sizes (25 hp). Flexible coupled pumps allow the motor or pump to be removed without disturbing the other. The flexible coupling requires very careful alignment and a coupling guard. Standard EPDM couplings have been known to fail when used with VFDs at low speeds; flexible polyurethane couplings are recommended where there will be many hours of operation below about 50% speed. The flexible coupled pump is usually less expensive than the close-coupled pump for pumps with motors 20 hp and larger. Usually end-suction pumps are the most cost-effective for use up to 1500 gpm to 2000 gpm but are available from some manufacturers up to 4000 gpm. In-line pumps (Figure 3-19) are end-suction pumps with a suction fitting designed so that they can be inserted directly into a pipe. They can also be mounted on a base like other pumps. In the past, these pumps were used almost exclusively for small pumps, but now they are available in the full range of sizes, with larger pumps generally mounted on a base for ease of maintenance. Inline pumps are not quite as efficient as end-suction pumps due to the inlet fitting. These pumps can save considerable space when mounted in-line with the piping, but maintenance access is worse when the pump and motor are well above the floor.

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Figure 3-20

Double-suction pump.

Figure 3-21

Vertical turbine pump.

Double Suction In the double-suction pump (Figure 3-20), water is introduced on each side of the impeller and the pump is flexibly connected to the motor. These pumps are preferred for larger flow systems (typically greater than 1500 gpm), because they are very efficient and can be opened, inspected, and serviced without disturbing the motor, impeller, or the piping connections. Typically, the pumps are mounted horizontally but can be mounted vertically. The pump case can be split axially (parallel to shaft) or vertically for servicing. This pump takes more floor space than end-suction pumps, particularly with the traditional horizontal inlet and discharge (as shown in Figure 3-20). However, some manufacturers offer a style with vertical inlet and/or discharge to reduce space requirements.

Vertical Turbine Vertical turbine pumps (Figure 3-21) are axial-type pumps that are used almost exclusively for cooling tower site-built sump applications. These pumps can be purchased with enclosures or “cans” around the bowls when not sump mounted.


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Pump Performance Curves For a given impeller size and rotational speed, the performance of a pump can be represented on a head-capacity curve of total head (in ft of water) versus flow (in gallons per minute). Total dynamic head (TDH) is the difference between suction and discharge pressure and includes the difference between the velocity head at the suction and discharge connection. Starting from zero flow, as the pump delivers more water, the mechanical efficiency of the pump increases until a best efficiency point (BEP) is reached. Increasing the flow further decreases the efficiency until a point where the manufacturer no longer publishes the performance (end of curve). Pump performance data are generally shown as a family of curves for different size impellers (Figure 3-22). Notice as the impellers get smaller, the pump efficiency decreases. The power (hp) requirements are also shown on the performance curve; notice that the power lines cross the pump curve until one value does not cross. This value is called the non-overloading horsepower, because operation at any point on the published pump curve will not overload the motor. Finally, information on the net positive suction head required (NPSHR) is shown on the pump curve. This is discussed in greater detail in the Pump Inlet Limitations section.

Figure 3-22

Pump head capacity curve.

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Chapter 3 Chilled-Water Plant Equipment Pump curves are also characterized as “steep” or “flat.” A flat-curve pump is defined as a pump where the pressure from shutoff head (head at zero flow) to the pressure at the BEP does not vary more than a factor of 1.1 to 1.2.

Parallel and Series Pumping When two or more pumps are operated in parallel (Figure 3-23), a combined parallel pump curve can be drawn that holds the head constant and adds the flow. Similarly, a series pump curve can be drawn that holds the flow constant and adds the head. Pumps are rarely placed in series in HVAC applications.

Variable-Speed Pumping For a given impeller size, a family of curves can be drawn to represent the variable-speed performance of a pump (see Figure 3-24). Notice that the BEP

Figure 3-23

Parallel pump curve.

Figure 3-24

Variable-speed performance curve.


59

follows a parabolic curve just like an ideal system curve (which is discussed in greater detail in the System Curves section). Also notice that the NPSHR lines follow fairly closely with the published end-of-curve lines for the various speeds. The power lines decrease rapidly as the speed decreases, which graphically demonstrates the potential power savings of variable-speed operation in variable-flow systems. For a more detailed example of variable-flow applications, refer to Chapters 4 and 6. Variable-speed pumps should, in general, not be piped in parallel with constant-speed pumps if both are expected to operate together. The constantspeed pump will overpower the variable-speed pump, keeping its check valve closed until the variable speed is increased sufficiently high to meet the pressure created by the constant-speed pump. Thus the VFD provides much lower energy savings when both pumps are operating. This can be seen by the example in Figure 3-25, where two variable-speed pumps in parallel will use 7.6 bhp while a variable-speed pump in parallel with a constant-speed pump will use 8.5 bhp. ANSI/ASHRAE/IES Standard 90.1-2016 and California Title 24 (CBSC 2016) requires variable-flow design for all CHW systems with more than three control valves. The standards also require VFDs on all variable-flow CHW systems with pump motors greater than 5 hp.

Figure 3-25

Performance of two VFD pumps versus one VFD and one constant-speed pump.

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Figure 3-26

Variable-speed performance at varying DP set points.

Source: Taylor 2012.

ANSI/ASHRAE/IES Standard 90.1-2016 also requires that the DP sensor used to control pump speed be located at the most remote coil or HX. This is because the lower the DP set point, the lower pump energy will be, as shown in Figure 3-26. If the DP sensor were located at the pump discharge, the set point would have to be set for the design pump head downstream of the pump to ensure adequate flow at design conditions. The result will be that the pump runs at near full speed most of the time and pump energy is not much better than if the pump had no VFD and simply rode its curve as flow reduced.

Selecting Pumps Pump Type Table 3-4 summarizes the advantages and disadvantages of the most commonly used pumps for CHW plants. For plants with variable-speed pumps expected to operate at low speeds, flexible-coupled pumps should be avoided unless special couplings are provided, as discussed previously. From a cost and space perspective, typical CHW plant pumps with motors 15 hp and less (20 hp and less with VFDs) should be close-coupled end-suction type. Larger pumps should be flexible-coupled end-suction type until they reach about 1500 to 2000 gpm, where double-suction pumps become cost competitive and their lower maintenance costs make them the best option.

Fundamentals of Design and Control of Central Chilled-Water Plants I-P Table 3-4 Type • • Double suction

• • • •

Pump Selection Summary

Advantages High efficiency Hydraulically balanced (less bearing wear) Longest life Only pump type available >2500 gpm or so Low cost Uses standard motors

• • • • • • •

End suction Base mounted (flexible coupled)

• End suction Close coupled

Lowest cost for pumps for 2 to 15 hp or so

• • • • •

• In-line Close coupled or • Flexible coupled •

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Lowest cost for small pumps <2 hp or so Lowest cost to install when mounted in-line Least space required

• •

Disadvantages Highest cost for pump (<2000 gpm) Requires most space Highest cost for inertia base Alignment required Possible coupling failure Higher cost than close-coupled <20 hp More space required than close-coupled Higher cost for inertia base Alignment required Possible coupling failure Motors above 25 hp are hard to find and costly Unusual motors of all sizes are often harder to find Often less efficient (due to inlet volute) Maintenance is more difficult if installed in-line

Pump Selection Pumps should be selected between 65% and 115% of the flow rate at the BEP for the impeller required by the design flow and head (see Figure 3-27) (not the BEP of the largest impeller). Selecting a pump too close to shutoff head or too near the end of the curve presents problems with radial thrust and potential cavitation. This is discussed further in the section Radial Thrust. Constant-speed CHW pumps serving terminals with two-way valves (variable flow) generally should be flat-curved pumps to reduce the additional head generated as the pump rides up its curve. If variable speed, efficiency is improved at low loads if the pump is selected to the right of the BEP. For multiple constant-speed pumps in parallel, motor sizes should be selected so that the power curve does not cross the pump curve at any point (non-overloading). Selecting for non-overloading is not required for singlepump constant-flow systems, because the system can be balanced with valves or the impeller can be trimmed to prevent overloading. Selecting for non-overloading is optional for pumps with VFDs, because the VFD can be configured to limit speed to prevent overloading, but a non-overloading motor will allow for greater pump flow when one pump is operating alone, which may be an advantage in case one pump fails.

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Figure 3-27

Pump selection recommendations.

Pumps with VFDs can also be selected with larger impellers to operate at less than 60 Hz at design conditions. As seen in Figure 3-22, pump efficiency generally increases with larger impellers, depending on the design flow relative to the BEP flow. Simply put, a pump with a VFD should be selected with the largest impeller available that does not increase motor and VFD sizes at full speed. For instance, assume the pump in Figure 3-25 was selected for 280 gpm at 50 ft of head. It requires about 5.1 bhp, requiring a 7.5 hp motor. Instead of trimming the impeller to the design duty (~8 in. as seen from the pump curve), it could be selected to max out the 7.5 hp motor. Using pump laws, D2 3 HP2 = HP1  -------  D1 HP2 D2 = D1  ---------- HP1 7.5 D2 = 8  -------  5.1

1/3

1/3

= 9.1 in.

So, a 9.1 in. impeller should be selected. The VFD would then automatically slow the pump to 8/9.1 · 60 = 53 Hz to achieve the desired duty. By oversizing the impeller in this way, pump efficiency increases from about 70% to about 75% at no additional cost. If the application allowed the pump to ride out its curve, the VFD could be configured to limit pump speed to prevent overloading. A common error is to take this one step further: if a 9.1 in. impeller is better than 8 in., why not select the maximum impeller for the pump, 9.5 in. in this case, and still use the 7.5 hp motor, again relying on the VFD to limit speed to


63

prevent overloading? Unfortunately, this will not work. With a 9.5 in.impeller, the pump would have to slow down to 8/9.5 · 60 = 50 Hz and at that speed the delivered hp would drop to 7.5 · (50/60)3 = 4.5 bhp (assuming the VFD is configured to a quadratic V/Hz relationship typically used for pump applications), less than the required 5.1 bhp. So, the impeller can only be maxed out within the capability of the selected motor size. The motor and VFD could have been increased to 10 hp with the 9.5 in. impeller, but overall efficiency usually gets worse with such an increase, because at low loads, motor and VFD efficiencies fall quickly, as discussed in the section Drive System Efficiency.

System Curves The system curve describes how the pressure drop in the system varies with flow. For a given system geometry, the head can be calculated using Equation 3-4: H = Q

x

(3-4)

where the exponent x varies from 2 for turbulent flow to 1 for laminar flow. The pressure drop of fittings, such as valves and elbows, can be modeled well with a constant exponent of 2. Straight piping and coil tubing seldom have velocities high enough for fully developed turbulent flow and usually are in the mixed region between turbulent and laminar. An exponent of about 1.8 is often used to describe the average relationship of flow versus pressure drop in common HVAC systems. In open systems (cooling towers), the static head of the tower is a constant, as is the DP set point in a variable-flow system where pump speed is controlled to maintain DP in the distribution system. These constant pressures are represented by raising the starting point of the curve at the zero flow line to the pressure that remains constant (Figure 3-28).

Figure 3-28

System curve with fixed head.

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Chapter 3 Chilled-Water Plant Equipment The system will operate where the pump curve and the system curve intersect (Figure 3-29). This point (gpm and head) is usually different from the design gpm and calculated head used to select the pump, due to the inaccuracies of pump head calculations and/or field changes (e.g., greater or fewer elbows installed than assumed). For constant-speed pumps, if the actual head is lower than the calculated head, one of two actions can be taken: the most efficient option is to trim the impeller to match the actual requirements. The system curve can be used to size the new impeller diameter (Figure 3-30). The second option is to partially

Figure 3-29

System curve and pump curve.

Figure 3-30

System curve (for sizing new impeller).


65

close balancing valves, which simply raises the actual system curve back up to the design system curve with obvious increases in pump energy usage. Trimming the impeller is required by ANSI/ASHRAE/IES Standard 90.1 with some exceptions that allow valve throttling (2016). For pumps with VFDs, neither impeller trimming nor valve throttling is required because the pump speed will automatically adjust to the reduced head. If the calculated head is below the actual head for fixed-speed pumps, the impeller must be replaced with a larger impeller. This may require a larger motor as well. For pumps with VFDs, the VFD can be operated at higher than 60 Hz to increase flow until the motor and VFD are at maximum current, in which case they will need to be replaced with a larger motor and drive. There is no value to increasing the pump impeller diameter in this case; speeding up the pump with the VFD is less expensive and the result is basically the same. The system curve modeled with a fixed exponent is reasonably accurate as long as nothing in the system changes (the geometry is fixed). When two-way valves are incorporated into the system, the variable pressure drop created changes the system curve dynamically. So, in a variable-flow system with multiple two-way valves, there are an infinite number of system curves that may represent the operating condition at any one time. Also, as the flow decreases, the velocities of the water in piping and coils can fall closer to the laminar flow regime (exponent = 1), reducing the average exponent and making the system curve more linear. So a two-way valve system can best be described as a range of system curves (Figure 3-31). However, for simplicity in energy modeling, the system curve is generally modeled as having a fixed exponent as if it were a constant-geometry system. It is not known how accurate this assumption is in representing real two-way valve systems.

Pump Inlet Limitations The boiling temperature of water is a function of the absolute pressure surrounding the water. In a pump, the pressure at the eye of the impeller can be the lowest in the system and, depending on the temperature, the water can boil (vaporize). As the liquid moves through the impeller and gains pressure, the water vapor collapses back into liquid. This process is called cavitation and can be very harmful to the impeller. In the field, the pump sounds like it is pumping gravel. The pressure at the inlet of the pump must therefore be high enough to prevent the water from boiling. Manufacturers publish NPSHR curves with pump curves (e.g., Figure 3-25). Notice that the NPSHR increases when the flow gets higher. The designer must ensure that the system will have enough net positive suction head available (NPSHA) to prevent cavitation. NPSHA (in ft) is calculated using Equation 3-5: 2 2 0.102 NPSHA = -------------  P a – P vp  +  V a – V s   193 +  z a – z s  – H f ,a  s

s

(3-5)

where s is the specific gravity of the fluid (= 1 for water), P is absolute pressure (psia), V is velocity (ft/min), and Hf is head loss (ft) due to friction. These are defined at two points: point a, which is a reference point in the system where

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Chapter 3 Chilled-Water Plant Equipment conditions are known, and point s, which is the pump suction. For closed systems, point a is typically the expansion tank connection; for open systems, point a is the top of the tower cold-water basin. Pvp is the vapor pressure of the fluid at the pump suction, which is a function of fluid temperature. For water in chiller plants, it ranges from 0.25 psia at 60°F (typical of a CHW pump suction) to 0.5 psia at 80°F (typical of a condenser water pump suction). Maintaining a high NPSHA is easily done in closed systems by adjusting the precharge pressure of the expansion tank (Pa). Because the vapor pressure is so low in CHW systems, all that is required to ensure adequate NPSHA is to maintain a positive gage pressure at the pump suction. In open systems, such as open cooling tower condenser water systems, many engineers get concerned about maintaining NPSHA and consequently insist on elevating the tower basin well above the pump suction. However, this is, in fact, not necessary because atmospheric pressure (Pa) is so high (14.7 psia at sea level). For instance, assuming the suction line pressure drop Hf is 2 ft and the pump suction elevation is the same as the basin elevation, the equation above indicates 2

 0 – 12 2  2.31 NPSHA = ----------  14.7 – 0.5  + ------------------------ +  0  – 2 61.3 1 = 28.5 ft

(3-6)

This is well above NPSHR for common pumps. So, cavitation is seldom an issue as long as the pressure drop between the basin and the pump suction Hf is reasonable. See additional discussion in the Pump Installation section that follows. What is often mistaken as cavitation is another phenomenon that occurs when the pump is not elevated well below the basin: air that is dissolved in the water in the tower basin comes out of solution as the pressure drops when water flows into the pump suction. These air bubbles can make a gurgling sound that is similar to (but not as loud as) the sound of cavitation. It is generally harmless to the pump, although the pump efficiency can be slightly reduced. To completely prevent this from occurring, the pump must be below the tower basin by an elevation equal to or larger than the NPSHR plus the pressure drop of the inlet piping Hf. This is almost never practical when the tower and pumps are located on the same floor of the building (e.g., the roof). Because the air bubbles do no harm to the pump, it does not make sense to spend a lot of money raising the tower elevation to ensure the bubbles do not occur. Another pump inlet problem to be avoided is vortexing. Any time water is drawn from an open tank or sump, there is a potential that a vortex will occur. Vortexing causes air to enter the pump suction line and decreases the effectiveness of the pump. Vortexing occurs even with very deep sumps. Any time water is drawn from a sump or open tank, anti-vortexing devices should be installed. These are standard devices supplied with cooling towers.


Figure 3-31

67

Range of system curves for two-way valve system.

When using vertical turbine pumps, care must be taken to ensure that manufacturer recommendations are followed to maintain a minimum submergence distance above the inlet bell. Also, adequate clearance must be maintained from the tank’s bottom to the pump’s inlet.

Radial Thrust When pumps operate at points on the pump curve other than BEP, nonuniform pressures can develop on the impeller. This is called radial thrust and can cause shaft deflection, excessive wear on pump bearings, and even shaft failure if high radial thrust is maintained for long periods of time. Radial thrust is greatest when pumps are operated at or near the shutoff pressures or near the end of the curve (Figure 3-32). This is one reason pumps should be selected close to the BEP (Figure 3-27). The use of VFDs on variable-flow systems also reduces radial thrust, because they keep the pump operating near the BEP; when constant-speed pumps ride up their curves, radial thrust increases as shown in Figure 3-32.

Pump Installation Locating Condenser Water Pumps

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Figure 3-32

Radial thrust.

When using an open-circuit cooling tower, the water falls by gravity into the collection basin and sump. Typically, on a packaged cooling tower, the collection basin and sump are an integral part of the tower. At the outlet to the pump suction there should be a screen for collecting larger debris and a device for breaking the vortex created by the suction. These are standard components provided with packaged towers. From the cooling tower outlet, the water flows into the suction of the pump with these application considerations: •

•

•

•

The pump must be at the same level as or below the tower basin water level. As discussed in the previous section, the pump need not be located well below the tower basin to avoid cavitation, because atmospheric pressure is well above net positive suction head requirements. Piping from the basin to the pump inlet must have a low pressure drop if the pump and tower elevation are nearly the same. Cavitation can only occur if the pressure drop from the tower to the pump inlet has a very high-pressure drop. That is why no strainer is shown in the piping at the pump inlet in Figure 3-34. The strainer can get clogged and generate a high enough pressure drop that cavitation can occur and the pump could be damaged. So, if the pump is located near the same elevation as the tower basin, strainers should be located on the discharge side of the pump, typically at the inlets to each chiller condenser. The pump impellers will be adequately protected by the screen provided with the tower at the basin outlet. That will keep any large debris from impacting the impeller; small particles will pass through the pump without causing any damage. Piping from the basin to the pump inlet should be at or below the basin water level elevation. If the piping rises above the water level, any air leakage into the piping can cause it to vent and drain when the pump shuts off. The pump prime is thus lost and when the pump restarts it will operate dry and be damaged. The pump must include a check valve if the pump discharge piping rises above the basin elevation. This will prevent backflow and basin overflow


•

69

when the pump shuts off. The piping is vented from the supply nozzles on the tower supply distribution side. To avoid trapping air bubbles, reducers should be eccentric with the top flat.

Pump Piping Considerations The following are installation and operation guidelines that help ensure proper operation of pumps: •

•

•

•

•

•

•

The minimum flow through a pump should be sufficient to remove the heat of compression (motor input power) with a limited temperature rise to avoid damage to seals or excessive warming of supply water. This is never an issue with primary CHW and condenser water pumps because minimum evaporator and condenser flows are so high. It can be an issue on secondary CHW pumps serving only a few coils, but not for those with VFDs with low minimum speed (see Variable-Frequency Drives [VFDs] section below) due to low pump power. A minimum of 4 to 6 diameters of straight pipe should be located upstream of the pump suction, or a suction diffuser should be installed. Suction diffusers are essentially turning vanes and can have strainers installed in them and thus provide double duty. Typically condenser water pumps follow the straight pipe recommendation because, as noted above, they should not have strainers at their inlets unless their elevation is well below the tower basin. CHW pumps typically are fitted with suction diffusers to reduce costs and space requirements. Variable-flow pumps should never have balance valves installed on the discharge, as flow balancing is accomplished automatically by varying the speed of the pump. When using a combination shutoff, balance, and check valve (also called a triple duty valve), install an additional shutoff valve downstream so that the check valve can be maintained. This is also advisable because most triple duty valves make poor shutoff valves because they require hand wrenches and multiple turns to close. Alternatively, simply do not use combination valves, per Chapter 5. A single pump gage should be used and piped with shutoff cocks to the taps that are provided with the pump at the suction and discharge. This allows the gage reading to be more accurately compared to the pump curves to determine pump operating point, because any error in the gage reading is canceled out when taking the difference in readings. Temperature gages and test plugs at each pump connection are of little value because pumps generate almost no temperature rise across the pump (only that due to the inefficiency of the pump impeller). Vibration isolation must be considered for pumps located above or below noise-sensitive occupied spaces. Twin-sphere EPDM-style isolators provide vibration dampening from the pump as well as the water flowing through the pump—stiff, braided-type flex connections are ineffective at dampening

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Chapter 3 Chilled-Water Plant Equipment vibrations due to impeller-generated water pulsations. Pumps in remote, dedicated plant buildings typically do not need vibration isolation. Figure 3-33 and Figure 3-34 show typical CHW and condenser water pump piping details following these guidelines.

Figure 3-33

Typical CHW pump piping detail.

Figure 3-34

Typical condenser water pump piping detail.


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Variable-Frequency Drives (VFDs) One of the greatest improvements in the efficiency of chiller plants is the result of VSDs, also known as VFDs. (The more commonly used acronym, VFD, is used herein.) The advent of a cost-effective means to vary the speed of chiller impellers, tower fans, and pump impellers has meant that greater operating efficiencies are now possible. Systems are also quieter, provide more stable temperature control, and have lower maintenance costs due to reduced or eliminated cycling and reduced radial thrust. The adjustable-frequency VFD is an electronic device that works by converting a fixed three-phase voltage and 60 Hz frequency source into a variable voltage and frequency source. Frequency of the source to the motor controls the motor speed. In order to keep the required torque of the motor, the voltage and frequency relationship must be maintained. This is called the volts to hertz relationship. Most HVAC applications require a variable torque because as the speed of the motor decreases, the load (torque) also decreases. For these applications, the drive is most efficient if it has a quadratic volts to hertz relationship, although the linear relationship still works. The large majority of the drives currently in HVAC systems use PWM technology. The PWM drive has a fixed diode rectifier that converts the AC input voltage to fixed DC voltage. The DC voltage is filtered and sent to the inverter section that changes the fixed DC voltage to variable AC voltage and changes fixed frequency to variable frequency. The inverter uses power transistors to chop the DC voltage to create the variable output. The transistors are turned on and off at a variable rate (carrier frequency) to create the variable output voltage and frequency. PWM VFDs have very high efficiency with little motor heating, have a constant input power factor, can run at low speeds, and have reduced audible motor noise (because of high carrier frequency). VFDs have many benefits: •

• • •

•

•

Energy savings: For CHW and condenser water pumps, the savings is multiplied because the reduced pump heat also reduces the load on the chiller or tower. It is also possible to select larger pump impellers to improve pump efficiency, as noted previously. Soft start: This reduces bearing wear and belt wear on cooling tower fans. Reduced radial thrust: Reduces bearing wear. Reduced noise from fans and impellers: VFDs can create a whiny noise from the motor windings, but this can be minimized by increasing the carrier frequency of the drive. See the Motor Noise section. Better temperature control: Due to reduced cycling on cooling tower fans and due to better controllability of control valves on variable-flow pumping systems. Pump power and status available to DDC system via network cards: Most VFDs include network cards (e.g., BACnet® MS/TP [ASHRAE 2016c]) that allow them to be integrated into the DDC system for remote monitor-

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Chapter 3 Chilled-Water Plant Equipment ing and control without hardwiring input/output (I/O) points. One benefit is that VFDs include a built-in power meter so pump power is known without having to install a separate meter. VFDs do have some disadvantages, including a possible negative impact on power quality, motor noise, electromagnetic interference (EMI), radio frequency interference (RFI), potential motor bearing pitting and failure, high-frequency motor noise (whining from vibrations in the motor windings), and nuisance tripping. The advantages far outweigh the disadvantages in most applications.

Drive System Efficiency The efficiency of the drive system takes into account electrical losses from the VFD, the connected motor, and the combination of the two devices. VFDs have losses in the form of thermal power from the inverter (60%) and rectifier (30%) and leaking current and power lost in the cooling equipment (10%). The inverter and rectifier losses are proportional to the motor shaft speed, and the other losses are fairly constant. Drive losses are 2% to 3% of the nominal power of the drive (i.e., VFD efficiency is 97% to 98%). The motor losses are from rotating losses, including friction and iron losses, and resistance losses caused by the resistance in the internal wiring in both the stator and the rotor. Motor efficiencies can be obtained from motor manufacturers or simply assumed to be the minimum efficiencies prescribed by the National Electrical Manufacturers Association (NEMA) for the motor type. The actual efficiency of the drive and motor operating together can be calculated if both the motor and drive losses are known over the entire speed range. The typical performance of the drive and motor, plus a combined curve, are shown in Figure 3-35. For this figure, data were provided by two popular drive and motor manufacturers and averaged. The y-axis is percent of full-load efficiency. So, the combined efficiency of a premium-efficiency 25 hp motor (93.6% efficiency at 25 bhp) with a VFD (~97% efficiency at 25 bhp) at 20% load (5 bhp) is about 0.936·0.97·0.8 (from curve) = ~73%. Fan power as measured by the VFD is shown in Figure 3-36, along with predicted power, assuming power varies as the cube of the flow adjusted by the combined drive and motor efficiency curves shown in Figure 3-35. The plot shows that the motor/drive efficiency curves are conservative at low motor loads but reasonably accurate above about 40% motor loads. The primary reason for the discrepancy is that the power measured by the VFD’s internal power meter is outgoing power, not incoming power, so it does not include VFD inefficiency. As shown in Figure 3-35, VFD efficiency is high at higher loads but falls at lower loads. Figure 3-36 show that there is little energy saved below about 25% speed; the cube law savings is offset by the reduced efficiency of the VFD and motor. However, the low minimum speed does improve temperature control and limit cycling, so it is beneficial to set the minimum speed set point as low as possible anyway. See the Minimum Speed Set Point section.


Figure 3-35

Typical VFD and motor efficiency at part load.

Figure 3-36

Fan performance—measured versus model.

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VFD Accessories and Application Considerations Bypass Starter A VFD can be equipped with a manual or automatic bypass switch and starter. Normally this is provided to allow the operation of the motor across the line (with no speed control) if the VFD fails or while it is being serviced. This requires two sets of contacts, one on the power side and one on the load side to isolate the drive for servicing. Care must be taken when operating the motor across the line. Larger motors (60 hp and above) are usually not started across the line due to limitations on inrush current. When operating in the bypass

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Chapter 3 Chilled-Water Plant Equipment mode, the system may be overpressurized, causing control valves to control more erratically or to be pushed open if their shutoff capability is less than the pump head. Bypasses cost about half as much as the VFD itself, and because VFDs have become very reliable, the cost/benefit of bypasses is questionable in noncritical plants. Even in more critical applications, bypasses might not add much value if the equipment served by the VFD (e.g., pump, tower cell fan) is itself redundant; modern VFDs are probably no more likely to fail than the equipment they serve.

Motors Motors powered through VFDs must be able to withstand spike and transient voltages induced by VFDs. For common pump and fan applications, motors called VFD ready or inverter ready that comply with NEMA MG-1 (2016), part 31, are sufficient. It is not necessary to use more expensive inverter duty motors, which are designed for constant torque applications. It is no longer necessary to specify premium-efficiency motors in the United States; they are mandated by federal law. Care must be taken when applying a new VFD to an existing motor. In some older motors, Class B insulation windings may not be sufficient to handle the voltage surges and additional overheating caused by the VFD. In these cases, the motors should be replaced with inverter duty motors or the existing motors can have a Megger test that will verify the condition of the insulation on the windings. Contact the VFD supplier; some maintain a database of existing motors and the expected failure rate when applying a VFD.

Shaft Grounding VFDs have been shown to create voltage differences between the rotating shaft and the casing of the motor. Electrical current passes from the shaft through the bearings into the casing. This electrical current can cause damage in the form of pitting of the bearing cases. A pattern known as fluting can form, which will eventually cause premature failure of the bearings. Experience suggests that the effects of bearing currents from common mode voltages are seldom appreciable, and, therefore, shaft grounding protection is not needed provided the following are true: • • •

• •

Motors are inverter ready (i.e., fully compliant with NEMA MG-1 [2016], part 31). Motors are 75 hp and smaller. The installation adheres strictly to the motor and drive manufacturers’ recommendations regarding the cabling and grounding, the latter being the key. The PWM switching (also called carrier) frequency is below 8 kHz. VFD speed is not fixed at or near one frequency.


75

Where these criteria do not apply, shaft ground kits can be installed on the shaft to neutralize the shaft voltage differences. Installing ceramic bearings in the non-drive end of motors is further recommended in applications requiring motors greater than 100 hp in order to mitigate high-frequency circulating currents.

Minimum Speed Set Point There are many misconceptions about minimum speed set points on VFDs. Some designers are concerned about maintaining a minimum speed for motor cooling. But because pump and fan power falls roughly with the cube of flow, the motor heat generated at very low loads is small, and the temperature ratings of motors are very high (about 390°F). So, for variable-torque applications like pumps and fans, there simply is no minimum speed required for motor cooling. The minimum speed should be set in the field as that required to cause the pump or fan to barely move: simply manually raise the VFD speed output until the pump or fan can be seen to move. This is typically about 5% to 10% speed. If concerns remain about having minimum speeds this low, consider using 10% as the minimum, because almost all major motor manufacturers advertise 10:1 turndown for VFD-ready motors used in variable-torque applications. (As noted previously, for gear-drive cooling towers, a higher minimum speed may be required to maintain crankcase oil flow.) For typical CHW and condenser water pumps, the minimum speed is never reached because minimum flow requirements through evaporators and condensers are so high. But for pumps used to control flow through cooling coils (see Chapter 4) and for cooling tower fans, more stable control is provided if minimum speed is set as low as possible, because it minimizes the need to cycle the motor.

VFD Location and Enclosure Type VFDs should always be located indoors wherever possible. Those that must be located outdoors require NEMA 3R enclosures. But even with these enclosures, VFD service life will not be as long if located outdoors. Capacity and efficiency can also degrade due to very high temperatures that can occur on a sunny rooftop location. Concerns about the length of wiring between the drive and motor are seldom valid arguments for locating VFDs outdoors; the maximum length recommended by most manufacturers is 300 ft or more, based on manufacturer.

Motor Noise PWM VFDs cause motors to hum due to vibrations in the motor windings. This can be mitigated by increasing the PWM switching (also called carrier) frequency. Most VFD manufacturers offer multiple switching frequency set points. But high switching frequency has several negative impacts: it increases the likelihood of shaft grounding issues and motor insulation failures, maximum current is reduced, maximum conductor length between the VFD and the

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Chapter 3 Chilled-Water Plant Equipment motor is reduced, and motor/VFD efficiency is reduced. So, raising carrier frequency should be the last resort for solving motor noise problems. Here is a suggested procedure for setting switching frequency: 1. Set to the standard frequency recommended by the manufacturer (typically 4 kHz), then check for motor noise in nearby occupiable spaces. 2. If motor noise is audible in occupied spaces, enable the noise smoothing feature offered by some VFD manufacturers. This feature randomly changes the switching frequency, resulting in more of a white noise effect rather than a continuous tonal hum. 3. If noise is still a problem, raise switching frequency to 8 kHz. 4. Raising the switching frequency above 8 kHz should be done only as a last resort and only if motor shaft grounding is provided.

Input Line Reactors Most modern VFDs include line reactors or DC chokes to reduce harmonics to the power line and as protection from AC line transients. They are adequate at controlling harmonics in almost all applications. The exception may be dedicated chiller plant buildings housing all-variable-speed plants. In this case, almost every motor has a VFD so harmonic noise may be an issue. In this case, the primary VFD supplier or the electrical engineer should perform an IEEE 519 (2014) analysis. The study should consider not only the utility point of common coupling but also other portions of the electrical distribution system and the effects on any alternative power sources (e.g., diesel generators). In addition to the nonlinear load characteristics of VFDs, other nonlinear load types (e.g., uninterruptible power supplies [UPSs]) should be taken into consideration. If it is determined that line reactors are required, they should be included in VFD specifications.

Output Load Reactors and Derivative of Voltage with Respect to Time (DV/DT) Filters The maximum conductor length between the VFD and the motor recommended by the VFD manufacturer should not be exceeded to prevent excessive electrical conditions such as voltage reflection and or capacitive coupling. Exceeding the maximum cable length value can cause damage to the VFD, motor, and cables. The maximum cable length varies depending on the VFD manufacturer/model and the selected switching frequency. If the length must exceed the VFD manufacturer’s recommended maximum length, discuss the application with the VFD manufacturer and consider a DV/DT filter in lieu of an output load reactor, because output load reactors may reduce available motor torque at full load.


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Output Power Wiring For most applications, wiring between the VFD and the motor should be separate conductors with XHHW-2 insulation field installed in a single metallic raceway (e.g., electrical metallic tubing [EMT]), including an internal equipment ground conductor. Standard PVC-insulated conductors (e.g., THHN) should not be used in new VFD installations due to inconsistencies in the manufacturing process (e.g., air voids) or nicks/cuts in the insulation during the conductor pulling/installation process, which may lead to corona and insulation degradation. XHHW-2 insulation is a high-temperature XLPE insulation that handles the high-voltage spikes from VFD outputs well, provides superior corona resistance, and will last longer than THHN PVC insulation.

Generators An IEEE 519 (2014) study should be performed when generators are supporting VFD loads. As a rule of thumb, VFD loads on a generator must be less than approximately 50% of generator capacity to limit total harmonic distortion to less than 15%. In addition to the IEEE 519 study, a generator sizing study should be performed to confirm generator compatibility with the VFD nonlinear load and other loads that the generator may be supporting (e.g., UPSs, medical imaging, elevators with regenerative loads). The generator sizing study should address voltage and frequency transient response, dips, overshoot, recovery time, leading power factor, and other sizing criteria at various loading conditions (not just full load) per the application requirements for the specific installation. Many of the generator manufacturers can assist with generator sizing studies with custom software for their specific generators. Automatic transfer switches (ATSs) that switch loads from utility power to generator power and back again should be selected with a programmable transition delay feature such that the ATS can be programmed to be placed in a neutral position (not connected to either source) for a minimum of 7 to 10 seconds (or as recommended by the VFD manufacturer) when transferring between two sources. This will ensure the VFD is not damaged when power is switched while VFDs are under power. This is recommended even when VFDs have a “flying start” feature that allow them to start into a rotating load without tripping.

References AHRI. 2011. AHRI Standards 550/590, 2011 Standard for performance rating of water-chilling and heat pump water-heating packages using the vapor compression cycle. Arlington, VA: Air-Conditioning, Heating, and Refrigeration Institute.

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Chapter 3 Chilled-Water Plant Equipment AHRI. 2000. AHRI Standard 560, 2000 Standard for Absorption Water Chilling and Water Heating Packages. Arlington, VA: Air-Conditioning, Heating, and Refrigeration Institute. ASHRAE. 2014. Chapter 18, ASHRAE Handbook—Refrigeration. Atlanta: ASHRAE. ASHRAE. 2015. Fundamentals of water system design. Atlanta: ASHRAE. ASHRAE. 2016a. ANSI/ASHRAE Standard 34, Designation and safety classification of refrigerants. Atlanta: ASHRAE. ASHRAE. 2016b. ANSI/ASHRAE/IES Standard 90.1-2016, Energy standard for buildings except low-rise residential buildings. Atlanta: ASHRAE. ASHRAE. 2016c. ANSI/ASHRAE Standard 135-2012, BACnet—A data communication protocol for building automation and control networks. Atlanta: ASHRAE. ASHRAE. 2016d. ASHRAE Handbook—HVAC systems and equipment. Atlanta: ASHRAE. IEEE. 2014. IEEE Standard 519-2014, IEEE recommended practice and requirements for harmonic control in electric power systems. Piscataway, NJ: Institute of Electrical and Electronics Engineers. IPCC. 2007. Table 2.14 (Errata). IPCC Fourth Assessment Report: Climate Change 2007. https://www.ipcc.ch/publications_and_data/ar4/wg1/en/ errataserrata-errata.html#table214. Calm, James M. 2005. Comparative efficiencies and implications for greenhouse gas emissions of chiller refrigerants. Non-CO2 Greenhouse Gases (NCGG-4). Rotterdam: Millpress. http://jamesmcalm.com/pubs/Calm%20JM,%20 2005.%20Comparative%20Efficiencies%20and%20Implications%20for %20GHG%20Emissions%20of%20Chiller%20Refrigerants.%20NCGG4.pdf. CBSC. 2016. 2016 California Building Standards Code. California Code of Regulations, Title 24. Sacramento, CA: California Building Standards Commission. NEMA. 2016. NEMA MG-1 2016. Washington, DC: National Electrical Manufacturers Association. Taylor, S. 2012. Optimizing design & control of chilled water plants: Part 5: Optimized control sequences. ASHRAE Journal 6:56–74.


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In the vapor compression refrigeration cycle, refrigerant flows from the compressor discharge, then a. through the condenser, evaporator, expansion valve, and back to the compressor. b. through the condenser, expansion valve, evaporator, and back to the compressor. c. through the expansion valve, condenser, evaporator, and back to the compressor. d. through the evaporator, expansion valve, condenser, and back to the compressor.

3-2

Which of the following are valid reasons for a building owner to prefer selecting a chiller utilizing R-134a instead of R-123 refrigerant? i. ii. iii. iv. a. b. c. d.

ODP Purge unit requirements GWP Future refrigerant availability

(i), (ii) (i), (ii), (iv) (i), (ii), (iii), (iv) (iii), (iv)

3-3

You are selecting a 300 ton water-cooled chiller. The available compressor types in this size range include which of the following? a. Scroll, screw, and centrifugal b. Scroll and screw c. Screw and centrifugal d. Centrifugal

3-4

Manufacturers typically limit evaporator flow rates on the low end to avoid _________ and on the high end to avoid _________. a. Surge; noise b. Deposit formation; erosion c. Laminar flow; erosion d. Deposit formation; noise

3-5

Benefits of gear-driven centrifugal chillers relative to direct-drive centrifugal chillers include the following: a. Smaller physical footprint b. The more common incorporation of multiple stages, and thus refrigerant economizers c. Greater flexibility in optimizing compressor efficiency

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Chapter 3 Chilled-Water Plant Equipment d. All of the above e. (b) and (c) f. (a) and (c) 3-6

You are designing a 300 ton primary-only CHW plant with a 20°F CHW T. The pump type with the best combination of efficiency and cost for this application is most likely to be a. Close coupled, end suction b. Base mounted, double suction c. Base mounted, end suction d. Close coupled, in-line

3-7

A customer expresses concern that the condenser water pumps in a plant located in San Diego, CA make a gurgling noise. The pumps are installed 3 ft below the basin water level. The strainers are located downstream of the pumps and the towers have vortex shedders. What is the most likely cause of the issue? a. Cavitation b. Dissolved air coming out of solution c. Vortices forming at the tower suction d. None of the above

3-8

You are designing a 700 ton central plant for a 20-story high-rise building. The cooling towers are to be located in a well on the roof where the primary constraint is roof area. Acoustics are not a concern. The type of tower best suited to this application is a. Field erected, induced draft, counterflow b. Packaged, induced draft, cross flow c. Packaged, induced draft, counterflow d. Packaged, forced draft, counterflow

3-9

A cooling tower has been selected for a 15°F range with a 6°F approach when the ambient wet-bulb temperature is 72°F. At design flow rate and 15°F range, but with an ambient wet-bulb temperature of 62°F, the tower is capable of providing a. A better approach b. A worse approach c. The same approach

3-10

Pump and tower VFD minimum speed set points of 6 Hz or less a. Result in significant energy savings relative to a minimum speed set point of 12 Hz. b. Are not feasible, because they cause motors to overheat. c. Improve controllability under low-load situations. d. None of the above.

Hydronic Distribution Systems Instructions Read the material in Chapter 4. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the Skill Development Exercises without referring to the text. Review those sections of the chapter as needed to complete the exercises.

Introduction This chapter addresses piping layouts and design issues related to CHW distribution systems and condenser water systems. The first section (ChilledWater Distribution Systems) addresses the CHW (evaporator) side of the chillers, CHW pumps, and cooling coils. The second section (Condenser Water Systems) addresses the condenser water side of the chillers, including cooling towers, condenser water pumps, WSEs, and other design options. The third section addresses optimal plant layout approaches. This chapter also introduces distribution system fundamentals. Chapter 5 addresses how these systems are applied to various applications and how components are selected for minimum life-cycle cost.

Chilled-Water Distribution Systems The chilled-water distribution system consists of chillers, pumps, piping, cooling coils, controls and other components on the evaporator side of the chillers. Understanding how hydronic distribution systems react to varying loads and how their components interact is essential for designing a reliable, energy-efficient, and cost-effective CHW plant. This section introduces fundamental distribution system types followed by a discussion of system design considerations. A brief outline follows: •

Constant-flow CHW systems ° Single or multiple chillers serving a single cooling load ° Single chiller serving multiple cooling loads ° Multiple chillers (in parallel or series) serving multiple cooling loads

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Chapter 4 Hydronic Distribution Systems •

Variable-flow CHW systems ° ° ° ° °

•

Concerns about variable flow in evaporators Primary/secondary variable-flow design Primary-only variable-flow design Distributed systems in larger plants Coil pumping strategies

CHW system design considerations ° ° ° ° °

Primary pump configuration Balancing variable-flow systems Degrading T syndrome Connecting heat recovery chillers Connecting Thermal Energy Storage (TES)

Constant-Flow Chilled-Water Systems Single Chiller Serving a Single Cooling Load With one or more chillers serving a single cooling coil (Figure 4-1), the simplest design strategy is to eliminate the traditional three-way control valve at the coil and to use a constant-volume pump to circulate water between the evaporator and the coil. Supply air temperature control is provided by resetting the temperature of the chilled water leaving the chiller. While constant CHW flow results in constant pump energy, chiller performance is improved when the leaving CHW temperature is reset to be as high as possible. A VFD could also be added to the pump to make the system partially variable flow (minimum flow is limited by the chiller minimum flow rate), but that adds cost and complexity and

Figure 4-1

Constant-flow system, single chiller, single coil.

Source: Taylor 2011a.


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may not be cost-effective, depending on the chiller’s minimum flow rate and system head requirements. Typically in a single-chiller, single-coil plant, pump head is small because the chiller and coil are close-coupled (physically close together). Many engineers have concerns about causing high space humidity with this design because CHW temperatures must be aggressively reset to maintain supply air temperature at set point under low-load conditions. However, in fact, that is never a concern, because the leaving supply air condition is about the same regardless of CHW temperature. For example, if the supply air temperature leaving the coil is 55°F, the air leaving the coil is close to saturated whether the CHW supply temperature is 42°F or 50°F. It is the supply air temperature set point that determines supply air humidity conditions, not the CHW temperature. Chillers must have a sufficient volume of water in the piping system to prevent unstable temperature swings. This may be an issue with single-chiller, single-coil systems. Often a small storage tank is required if the chiller is closely coupled to the coil. The minimum water volume varies with the chiller unloading capabilities and should be verified with the chiller manufacturer. Figure 4-1 shows a single chiller, but any number of chillers can be used. When two chillers are used, this is a good application for piping chillers in series rather than in parallel. Series chiller piping is discussed further below in the section “Multiple Series Chillers Serving Multiple Cooling Loads.”

Single Chiller Serving Multiple Cooling Loads In a constant-flow system serving multiple coils (Figure 4-2), three-way valves are used at the cooling coils to modulate the load at each air handler.

Figure 4-2

Constant-flow system, single chiller, multiple coils.

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Chapter 4 Hydronic Distribution Systems Despite the use of the term constant flow with three-way valve systems, threeway valves do not actually provide constant flow. This is demonstrated in Figure 4-3. The first four rows of the table show the system pressure drop by component at a fixed flow rate of 100 gpm to the branch circuit. The resulting pressure drop at the point of connection to the branch circuit varies from 20 ft of head at both the full (100%) flow and no (0%) flow conditions. The bypass balance valve is provided in the coil bypass for this reason. At the 50% flow condition, the pressure drops to 11.5 ft of head because there are now two paths for the water to take and the pressure drop varies roughly as the square of the flow. In the bottom row of the table, we see what the flow is through the branch if the system pressure is held at 20 ft of head at the branch circuit point of connection. At both the 100% flow and 0% flow conditions through the coil, the flow is 100 gpm (design). But at the 50% valve position, the flow increases to 132 gpm, 32% over design flow. Thus, at part-load conditions, systems with three-way valves can experience starved coils at the most remote parts of the system. Using automatic flow-limiting

Item

Pressure Drop (ft) @ 100 gpm 100% to Coil

50% to Coil

0% to Coil

Pipe/valves

2

2

2

Coil and/or bypass

8

2

6

Globe control valve

10

7.5

12

Total

20

11.5

20

gpm @ 20 ft P*

100

132

100

Actual P available may change.

Figure 4-3

Flow variation as a function of valve position.


85

valves (AFLVs) at each coil is one way to maintain system balance. However, a better solution is to provide a few two-way valves in the system, as discussed in the Variable-Flow Chilled-Water Systems section.

Multiple Parallel Chillers Serving Multiple Cooling Loads Figure 4-4 is the same as Figure 4-2, except there are now multiple chillers. When the system operates near full load, performance is satisfactory because all chillers and pumps are operating. However, constant-flow systems can have problems during part-load or off-peak conditions depending on how coil loads vary. To demonstrate the point, consider a system like that shown in Figure 4-4, with two equally sized chillers that serve two equally sized coils with each coil in turn serving a hotel meeting room. If there are functions in both rooms and both rooms are at or near full-load, the system operates well: both chillers with their associated pumps are running and each function space is receiving its design flow. But a problem can occur if the loads are significantly out of balance. For instance, what happens when only one of the two function spaces is occupied (say Coil A) and the other (Coil B) is vacant. The system as a whole is at half load, so in theory only one chiller and pump could satisfy the load. However, Coil B will also take its design flow, although it will merely bypass

Figure 4-4

Constant-flow system, multiple parallel chillers, multiple coils.

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Chapter 4 Hydronic Distribution Systems this flow from the supply to the return. If the plant operates with only one chiller and pump, it has sufficient load capacity but it cannot meet the flow demands. One half of the water will flow through Coil A, which is less than it needs to meet the load, and the other half of the water will flow through Coil B. To avoid starving Coil A, both pumps will have to operate and both chillers will have to operate at 50% load. This problem is one of the reasons designers migrated to variable-flow designs, as discussed in the section Variable-Flow Chilled-Water Systems. But this system can work well as long as all coil loads tend to vary in the same proportion, as they might if all coils served similar occupancies (e.g., all serve offices on the same schedule). For instance, if the coils served are below half load and only one chiller and pump are operating, all coils will be capable of meeting their loads. The system is thus a quasi-variable-flow system in that pumps and chillers can be staged. Also, because the loads vary similarly, CHW temperature may be reset aggressively, which allows the plant to be about as efficient as one of the true variable-flow systems discussed in the section Variable-Flow ChilledWater Systems. So, this system is a reasonable choice for small applications with only a few coils serving similar loads; it is simple and inexpensive and avoids all of the complexities of variable-flow systems. Also, small systems like this are typically close-coupled so there is not much pump energy to save. Note that ANSI/ ASHRAE/IES Standard 90.1 only allows this approach for systems with three coils or fewer or a total CHW pump system power of 10 hp and less (2016).

Multiple Series Chillers Serving Multiple Cooling Loads One solution to the staging problems associated with constant-flow systems with multiple chillers is to put the chillers in series where the entire flow passes through each machine (Figure 4-5). This design was popular in the 1960s and 1970s because it simplified controls. The design eventually lost favor to the parallel arrangement (Figure 4-4), which has lower first costs—the series arrangement requires additional piping and valves to allow one chiller to be offline for service while the other is operational, and parallel flow allows pipe sizes to be reduced at each chiller connection. Series chillers are now coming back into favor as an energy conservation measure, in part because of the movement toward high CHW temperature differences (see Chapter 5). With the series arrangement, the upstream chiller reduces the water temperature to roughly half of the overall range (e.g., from 62°F to 52°F) and the downstream chiller does the rest (e.g., 52°F to 42°F). Full-load chiller power is primarily a function of the temperature leaving the chiller; the entering temperature has almost no impact on efficiency. Hence, the upstream chiller will be more efficient than the downstream chiller, and so the overall chiller power will be less than with two parallel chillers. However, this savings is at least partly offset by the increased pump energy because the evaporators are in series. The pump head through the evaporator can be minimized by judiciously selecting the chillers and/or reducing the number of passes through the evaporators from two (or three) to one (or two), with some reduction in the downstream chiller efficiency.


Figure 4-5

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Multiple series chillers, multiple coils.

Variable-Flow Chilled-Water Systems Introduction Variable flow has many advantages in large CHW systems with multiple chillers and multiple loads or coils. Varying the flow allows the load and flow to better track, making chiller staging more efficient. Significant pumping energy can also be saved. This section discusses many important aspects of variable-flow CHW systems, including the following: • • • •

The effect of variable flow through the evaporator How primary/secondary distribution systems work The emergence of primary-only variable-flow pumping Energy efficiency opportunities with distributed pumping techniques

Variable Flow in the Evaporator of a Chiller Flow in the evaporator can be dynamically varied but within several constraints. If a chiller is operating in a stable condition and flow in the evaporator is reduced, the leaving CHW temperature will drop. If the flow reduction

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Chapter 4 Hydronic Distribution Systems occurs slowly, the controls will have adequate time to respond and the system will remain stable. But a rapid change in flow will cause the leaving water temperature to drop quickly. If the controls react too slowly, the chiller may shut down on low-temperature safety. This is a significant nuisance because the operator must manually reset the safety control and the chiller must remain off for a minimum period of time before restarting. Fortunately, modern digital controls are much more robust than older controllers and can accommodate changes in evaporator flow as long as they are not too abrupt. Another issue is avoiding laminar flow in the evaporator tube. A fluid velocity resulting in ~3000 Reynolds number and above (roughly equal to about 3 ft per second and above with smooth tubes) is recommended to maintain good heat transfer. Tubes enhanced with rifling (grooves) on the inside can operate at even lower velocities before laminar flow sets in. Consult the manufacturer’s literature for chiller minimum flow rates. The distribution system design must maintain flow rates through the chiller greater than the minimum. Because of these issues, designers traditionally chose constant-flow rates through the evaporators, leading to primary/secondary distribution systems. However, due to cost and energy advantages (discussed in Chapter 5) coupled with the success of robust new chiller controllers, variable flow through the evaporator is rapidly becoming standard practice.

Primary/Secondary Variable-Flow Design Primary/secondary refers to two hydraulic loops, each with its own pump, that share a common piping element called the common pipe or common leg. If the pressure drop in the common leg is very low (Figure 4-6), the two loops are hydraulically independent, meaning that flow through one loop has little or no impact on flow in the other loop. This concept was ideal for early CHW systems. To accommodate wide-ranging coil loads and reduce pumping costs, two-way valve variable-flow distribution is desirable. However, chillers required stable and (at the time) near-constant flow.

Figure 4-6

Primary/secondary hydraulic independence.


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Primary/secondary pumping solved that problem: the primary loop could provide constant flow through the chillers while the secondary loop could provide variable flow through the coils (Figure 4-7). Until recently, primary/secondary was the standard design for central CHW plants with multiple chillers and multiple cooling loads, and it still has many applications, as discussed in Chapter 5. Primary pumps are typically constant-volume, low-head pumps intended to provide a constant flow through the chiller’s evaporator. The secondary pumps deliver the chilled water from the common leg to coils and then back to the common leg. Secondary pumps are typically fitted with VFDs. VFDs are typically costeffective except on very small systems. Note that ANSI/ASHRAE/IES Standard 90.1 requires VFDs on CHW pumps exceeding 5 hp (2016). The VFDs are controlled by a DP sensor located near the most remote coil so that the DP set point can be as low as possible; this is also a requirement of ANSI/ ASHRAE/IES Standard 90.1. Locating the sensor near the pump requires a high DP set point and eliminates most of the energy savings from the VFD. See additional discussion on pump controls in Chapter 7.

Figure 4-7

Primary/secondary variable-flow piping, multiple parallel chillers, and multiple coils.

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Chapter 4 Hydronic Distribution Systems It is essential to have the flow rate in the primary loop equal to or greater than the flow rate in the secondary loop. When the secondary flow exceeds the primary, return water from the system flows back through the common pipe and mixes with the supply water from the chillers. This increases the temperature of the supply water to the secondary system. The warmer supply temperature in turn causes the control valves at each cooling coil to open even more (more water is required to meet the coil load when it is warmer), creating an ever-increasing demand for secondary system flow. Eventually, the secondary system will be at full flow and coils will be starved. This phenomenon (often referred to as the death spiral) is well documented in literature, as it has plagued many CHW systems that suffer from degrading T syndrome. See the discussion under Degrading T Syndrome for mitigation measures. The arrangement of the primary/secondary common leg piping connections is considered important by some engineers. The concern is the migration of the secondary system return water back to the supply piping, increasing supply water temperature. When the flow enters into a bullhead tee, the velocity pressure of the water can cause the flow to be forced into the supply. As shown in Figure 4-8, the tee used in the secondary piping system return should be arranged so that flow from the bypass is induced into the return system. If this is not possible, separate the return connection bullhead tee from the secondary supply connection by three or four pipe diameters. However, this issue is only important when the primary and secondary flow rates are nearly equal, which is not a common situation, so in practice the configuration of the piping seldom has a significant impact on plant performance. Another issue in primary/secondary piping is the size of the common leg. In a multiple-chiller system that is properly controlled, the maximum flow in the bypass should not exceed 110% to 115% of the flow from one chiller. However, in many cases the bypass line may be the same size as the supply and return headers for the following reasons:

Figure 4-8

Water flow through tee in common piping.

Fundamentals of Design and Control of Central Chilled-Water Plants I-P • •

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The first cost of saving material by decreasing the size of the bypass line can be more than offset by the labor for the extra fittings required. Lower pressure drop in the common leg makes the two circuits more hydraulically independent.

In very large systems (e.g., pipes larger than 20 in.), particularly if the common leg is long, the cost savings achieved by downsizing the common pipe should be considered. Another issue is where to locate the common leg. Figure 4-9 shows the common leg in the traditional front-loaded position (top) and back-loaded position (bottom). The former always results in the two chillers being equally loaded (assuming they each are controlled to the same leaving water temperature) because they have the same entering water temperature. But the back-loaded location causes the (a)

(b)

Figure 4-9

(a) Front-loaded versus (b) back-loaded common leg.

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Chapter 4 Hydronic Distribution Systems chillers to be unequally loaded; the chiller closest to the return will be fully loaded, while the one closest to the common leg will handle the rest of the load, often a very low load at which chillers operate very inefficiently. So the back-loaded common leg results in less efficient operation for most chiller plant applications. Primary/secondary systems can also have chillers piped in series. This could be done using the approach shown in Figure 4-5, but a more flexible and more energy-efficient design is shown in Figure 4-10. In this case each chiller has its own common leg. Pumping energy is lower in this scheme than with the design used in Figure 4-5 at low loads when only one chiller is on, because the inactive chiller is bypassed, creating no added pressure drop effectively. This design also allows chillers to be unequally sized, which can improve performance at very low loads, common to plants serving AHUs with outdoor air economizers. One application for piping chillers in series is coupling an absorption or engine-driven chiller with an electric chiller. This arrangement allows the operator flexibility to choose which machine to load based on utility rates or other criteria. In addition, a series configuration allows the absorption machine to operate at higher outlet temperatures, which increases its efficiency and allows it to operate in a cold-water chiller plant, often used in large distribution sys-

Figure 4-10

Primary/secondary variable-flow piping, multiple series chillers, and multiple coils.


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tems to increase T and reduce piping costs. Absorption chillers generally cannot produce colder than 40°F chilled water and thus are incompatible with a plant designed for 38°F–40°F chilled water when piped in a parallel configuration. The series configuration also allows the chillers to be unequally loaded (if, for example, you wanted to preferentially load the thermal machine during on-peak times). This is done by adjusting CHW set points; the lower the set point, the more the chiller is loaded. With this control, either the upstream or downstream chiller can be preferentially loaded.

Primary-Only Variable-Flow Design The simplest variable-flow distribution system is shown in Figure 4-11. Twoway valves are installed at most coils with just enough three-way valves installed to maintain the minimum flow required by the chiller. This minimum rate, which can be obtained from the manufacturer, varies with design CHW flow rate and the chiller type, size, and manufacturer but is typically 25% to 50% of the design flow. Locating the three-way valves near the chiller minimizes pump energy if the pump has a VFD. Locating them remotely increases flow to the extremes of the system, which increases the pump pressure and power required. There is usually no benefit to locating three-way valves remotely to keep the system

Figure 4-11

Primary-only variable-flow system with three-way valve minimum flow, single chiller, multiple coils.

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Chapter 4 Hydronic Distribution Systems cold so that chilled water is instantly available to coils; it typically takes just seconds or perhaps minutes for water to travel from the chiller to the most remote coil, and the load will not be lost in that short time. Nevertheless, the best location for the three-valves is usually at the most remote coils because this engages the water volume (heat capacity) of the system, which minimizes chiller short cycling. The design also will retain the self-balancing nature of two-way valve systems (see Balancing Variable-Flow Systems). These operational considerations are more important than the small pump energy penalty. The energy efficiency of a single-chiller primary-only variable-flow system could be enhanced by providing two-way valves at each coil and maintaining minimum flow using a flowmeter and bypass valve as shown in Figure 4-12. Flow and, therefore, pump energy are lower with this design, but with a significant increase in control complexity of the bypass (discussed further below in this section and in Chapter 5). Figure 4-13 shows a primary-only variable-flow system with multiple chillers in parallel and a minimum flow bypass. Unlike with a single chiller, the use of three-way valves for minimum chiller flow control is not a good choice with multiple chillers because of likely chiller staging problems discussed previously for constant-flow multiple-chiller systems. With multiple chillers, performance is enhanced when flow and load track each other, so all coil valves should be two-way.

Figure 4-12

Primary-only variable-flow piping with bypass, single chiller, multiple coils.


Figure 4-13

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Primary-only variable-flow piping, multiple parallel chillers, multiple coils.

The type and size of the minimum flow bypass valve is critical to the stability of the system. A common mistake is to oversize the bypass as if it were the common leg of a primary/secondary system. This results in unstable, fluctuating flow through the bypass that can cause fluctuating water temperatures to the chillers, making their controllers unstable and possibly causing chiller short cycling. For systems with constant-speed pumps, the valve should be sized for the high pump head that will occur as the pump rides up its curve at minimum flow. For variable-speed pumps, the head available at the bypass may be about as low as the DP set point used to control the pump due to the low flow out to coils. In both cases, the valve should usually be sized for the minimum flow rate of the lead chiller; no bypass is likely to be needed when multiple chillers are operating, except perhaps if they are very unequally sized. The valve should be a ball, globe, or pressure-independent valve, not a butterfly valve (which does not have the desired flow characteristic). Flow rate is usually mea-

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Chapter 4 Hydronic Distribution Systems sured using a flowmeter in the primary line serving the chillers. It is also possible to deduce flow using a DP sensor across the chillers and correlating pressure drop to flow using chiller manufacturers’ curves. This is usually less expensive, but a flowmeter is more common because flow is typically desirable for determining plant load used for chiller and pump staging and plant performance calculations, as discussed in Chapter 7. If the bypass valve is controlled only to maintain the minimum flow rate recommended by the chiller manufacturer, which typically is 25% to 50% of the design flow rate, pump energy will be lower than for primary/secondary systems with constant-speed primary pumps. See more discussion of relative pump energy in Chapter 5. The bypass valve can be located either adjacent to the chiller as shown in Figure 4-13 or far out in the system. Locating it close to the chillers provides the best energy performance because it reduces flow in the distribution piping, hence reducing pump energy. It also allows the chiller flowmeter and the bypass valve control to be from the same control panel, ensuring that DDC system network delays or failures do not affect control of the bypass valve. On the other hand, a remote location results in more consistent pressure drop across the valve, which makes control more stable, and any instabilities are less likely to cause fluctuating chiller entering water temperature, which can cause chiller cycling. For campus situations, a more remote location also ensures that the water in the distribution system is kept cold. For a large campus loop, the time required to cool down the mass of water in the system can be substantial. However, keeping the loop cold can be accomplished by other means (e.g., a threeway valve at the end of the system) and is generally not necessary in single building systems where chilled water can reach a remote coil in seconds or (at worst) a few minutes. Because of the energy and control system cost advantages of the close location, it is usually preferred; however, proper valve sizing and loop tuning are essential.

Large Plant or Building Distribution Systems Distributed Pumping System See Figure 4-14. The primary/secondary pumping arrangements described above have the secondary pumps located near the common piping (within the central plant) and serving all of the secondary system. While this strategy is reasonable for most systems, it uses more pump energy than necessary on systems serving coils with widely differing head requirements, such as those serving large hospitals, airports, and university campuses. For these applications, distributed secondary pumps as shown in Figure 4-14 can reduce pump energy. In the traditional primary/secondary arrangement, pressure created by the secondary pumps must be sufficient to deliver the chilled water to the most remote load or coil. As a result, the coils located closest to the secondary pumps have excess DP that is simply throttled by the control valve. This overpressurization not only represents energy waste, but it strains the ability of the control valve to accurately modulate water flow and may even cause valves to lift off their seats.


Figure 4-14

97

Primary/secondary variable-flow piping, distributed pumping.

A distributed pumping system moves the secondary pumps from within the plant and locates them remotely nearer to the loads they serve. In a single large building these pumps could be located either at the coils or on distribution branches that serve a group of smaller loads. In a large system, the distributed secondary pumps would be located in each building served by the plant. In either case, the distributed secondary pumps are sized for the pressure drop needed to move the water from the common pipe at the plant to their most remote coil within the building and back to the common pipe.

Tertiary Pumping System See Figure 4-15. A variation of the standard primary/secondary pumping strategy is to provide standard secondary pumps within the central plant and locate tertiary pumps remotely within the building, at the loads. Traditionally, a crossover bridge with a two-way valve would be used at the connection point between the main distribution piping connections and the tertiary pump (Figure 4-15, Building A). The crossover bridge hydraulically isolates the tertiary pumping system in the building from the main secondary system. The two-way valve ensures that the secondary flow through the crossover bridge is equal to or less than the building flow. This valve is typically controlled to provide building chilled-water supply temperatures (CHWSTs) 1°F warmer than the secondary loop supply temperature (to ensure that all of the secondary flow goes to the building).

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Chapter 4 Hydronic Distribution Systems

Figure 4-15

Primary/secondary variable-flow piping, tertiary pumping.

With modern variable-speed technology and control, however, the crossover bridge is not necessary with the tertiary pump piped in series with the secondary pumps (Figure 4-15, Building B). Without the crossover bridge, the closest buildings can use available DP, with the tertiary pumps simply supplementing this when needed. When using a tertiary pump without the crossover bridge, a bypass line with check valve can be added around the tertiary pump so that during periods of high DP in the secondary piping, the tertiary pump can be de-energized. Series-piped tertiary pumps must be controlled with VFDs with speed controlled to maintain the building coil pressure. This will


99

ensure that total pressure from the pumps in series does not overpressurize the building, and it ensures that the tertiary pump does not disrupt the flow in the tertiary loops of adjacent buildings.

Coil Pumping Strategies Four coil pumping strategies are discussed here and shown in Figure 4-16. Figure 4-16 (Coil Pumping Scheme A) shows a variable-speed coil pump used in place of a two-way control valve; rather than controlling flow by modulating a valve, pump speed is used. This reduces pump energy and can reduce first costs when applied to large AHUs in a distributed design approach. This design is discussed further in Chapter 5. Figure 4-16 (Coil Pumping Scheme B) shows a primary/secondary coil pump. This design is used to maintain a constant flow through the coil, which can be an effective means of coil freeze protection if the coil is exposed to subfreezing outdoor air. In this design, the coil pump and plant distribution pumps

Figure 4-16

Coil pumping strategies.

100

Chapter 4 Hydronic Distribution Systems are hydraulically independent—flow through the coil is unaffected by the operation of the distribution pumps or the two-way control valve. The distribution pumps must have sufficient head to deliver water through the common leg and back to the plant. Figure 4-16 (Coil Pumping Scheme C) is an option when the distribution pumps do not have sufficient head to deliver water to the coil. When the threeway valve is open to the distribution system, the coil pump is in series with the distribution pumps and can pull water to the coil. When the valve is closed to the distribution system, the pump maintains constant flow through the coil (e.g., for freeze protection). Caution must be applied to this approach because the pump may reduce or even reverse the DP across the distribution supply and return lines, possibly starving nonpumped coils. In Figure 4-16 (Coil Pumping Scheme D), the coil pump is in parallel with the main distribution pumps. Again, this design is primarily used for coil freeze protection as in Figure 4-16 (Coil Pumping Scheme B), but it has the advantage that either the main distribution pumps or the coil pump can maintain flow through the coil should either of them fail. The pump is usually controlled to operate when the two-way valve has closed below 50% or so then turn off at higher primary flow rates. Coil Pumping Schemes B, C, and D have been used also to ensure the coil flow rate never falls into the laminar region. The theory was that coil performance (T) would degrade as the flow through the coil becomes laminar, reducing plant T. In fact, with fully circuited coils, T improves at lower flows and the T never drops below design even in the laminar flow region (see Figure 4-20). Furthermore, simulations have shown coil pumps use more energy than they could possibly save even if the T is assumed to degrade drastically at low flows. So, other than perhaps to provide freeze protection (not typically needed with CHW coils), these configurations should be avoided; they increase both installed costs and operating costs.

Chilled-Water System Design Considerations Primary Pump Arrangements There are three common options for piping primary pumps on a CHW system: • • •

Option A: Dedicate a pump for each evaporator. See Figure 4-17a. Option B: Provide a common header at the pump discharge and two-way automatic isolation valves for each evaporator. See Figure 4-17b. Option C: Provide a common header with normally closed (NC) manual isolation valves in the header between pumps. See Figure 4-17c.

The advantages of dedicated pumps for each chiller (Option A) include the following:


(a)

(b)

(c)

Figure 4-17

Piping options for primary pumps: (a) Option A: dedicated pumps, (b) Option B: headered pumps with automatic-isolation valves, and (c) Option C: headered pumps with manual isolation valves.

101

102


•

•

The pump can be custom-selected for the chiller it serves. Pump selection can then take into account variations in evaporator pressure drop and flow rates when chillers are not identical. This can reduce pump energy compared to Option B, where the head of each pump must be the same and sized for the evaporator with the highest pressure drop; balance valves at the other evaporators must be throttled to generate this same pressure drop. Controls are a bit simpler because the pump can be controlled using the contact provided with the chiller controller. This ensures that the pump starts and stops when the chiller wants it to. With Option B, the control of the isolation valves and pumps is by the DDC system and must be coordinated with the needs of the chiller controller to avoid nuisance trips. For instance, the pumps generally must run for several minutes after the command for the chiller to stop so that the chiller can pump down the refrigerant. Pump failures do not cause multiple-chiller trips. With dedicated pumps, if a pump fails, only the chiller it serves will see a flow disruption and trip. With Option B, all operating chillers will see a flow reduction when a pump fails, possibly causing more than one chiller to trip due to low flow or a sudden drop in evaporator temperature. This has become less of an issue with modern, robust chiller controllers.

The advantages of headered (manifolded) pumps (Option B) include the following: •

•

•

Redundancy is improved. With Option A, if a pump fails and a chiller other than the one it serves also fails (albeit, this is a rare event), then two chillers will be inoperative. With Option B, any pump can serve any chiller and under many conditions one pump can provide enough flow for two chillers to operate near full capacity. Including a standby pump is much simpler. Adding a standby pump to Option A is cumbersome and expensive because it requires extensive piping and manual or automatic isolation valves. If standby pumps are desired, Option B is the best option. Isolation valves can be modulating or slow-acting to reduce sudden flow variations when starting and stopping pumps and chillers. With dedicated pumps, when a chiller and pump are enabled, there is no flow through the pump or chiller until the pump speed matches that of the other operating primary pumps—otherwise, the pump check valve is closed. So, flow changes can be sudden, which can cause chiller trips with primary-only systems, as discussed in Chapter 5.

Headered pumps with manual isolation valves (Option C) can have the advantages of Option A (although this option works best with identical chillers). It also mitigates the redundancy disadvantage of Option A, although accommodating a pump failure requires manual manipulation of valves versus the automatic response in Option B. Including a standby pump is possible with


103

Option C, but it only works (depending on which pump fails) with the header isolation valves open, and chillers must be staged by manually opening and closing their isolation valves. First costs are usually lowest with Option A if the chiller and pump pairs are close coupled (physically close together) and the manual isolation valves between the two are eliminated (each chiller-pump pair is isolated for service as a pair). Costs can be higher with Option A if pumps are grouped together and not adjacent to the chillers due to long piping runs from the pumps to their dedicated chillers. Option C is usually less expensive than Option B, but Option B is usually the best choice on primary-only variable-flow systems and where standby pumps are required.

Balancing Variable-Flow Systems Balance valves are often provided to deliver design flow rates at each coil or other heat exchange device in hydronic systems. There is no debate about the need to balance flow in constant-flow (three-way valve) systems. However, the need to balance variable-flow (two-way valve) hydronic systems is controversial and the subject of many articles and papers. Balancing of variable-flow systems is intended to do the following: • •

Ensure adequate flow available at all coils to meet loads. Note that “adequate” flow may be less than the design flow depending on coil loads. Ensure DP across control valves is not so high as to cause erratic control such as two-positioning, where the valve cracks open, provides too much flow due to the high DP, and then immediately closes after overshooting set point.

The discussion that follows is based on a detailed analysis (Taylor and Stein 2002) of a real design of variable-flow chilled- and hot-water systems using a hydronic analysis program to evaluate the following design options: 1. No balancing (relying on two-way control valves to automatically provide balancing) 2. Manual balancing, most commonly using CBVs to measure and adjust flow 3. AFLVs 4. Reverse return 5. Oversized main piping 6. Undersized branch piping 7. Undersized control valves At the time the referenced study was made, pressure-independent control valves were only available from one manufacturer, so they were not included in the study. They are now made by several manufacturers and have become more cost competitive. Pressure-independent valves generally have two components:

104

Chapter 4 Hydronic Distribution Systems a standard ball- or globe-type control valve controlled by the digital control system and a self-powered pressure control section that maintains a constant 4 to 5 psi DP across the control valve section regardless of the available DP. The latter component also typically provides flow-limiting duty, preventing the valve from using more than design flow. While not included in the detailed study, pressure-independent valves are included in the discussion and recommendations below. Table 4-1 summarizes the design valve pressure drops and transient flow rates of the balancing methods studied in the article. The columns with maximum pressure drop data are for the condition of all coils at design flow. The data indicate how much DP the worst-case control valve has to shed to get design flow. This is an indirect indication of valve controllability (the higher the pressure, the more closed the valve is under design conditions). Standard electric/electronic control valves typically have 50:1 to 100:1 turndown ratio. This allows the valve to control reasonably stably with a DP on the order of 30 psi (70 ft), based on the authors’ experience. Still, lower DPs are desirable. The results show that all options result in sufficiently low DPs to provide reasonable controllability. Reverse return (Option 4) provides the lowest DP while Option 1 (No balancing) results in the highest DP. Option 3 (AFLVs) Table 4-1

Pressure Drop and Flow Extremes of Balancing Options

Balancing Method

Maximum Pressure Drop of Control Valve Required for Design Flow, feet

Percent of Design Flow (Percent of Design Coil Sensible Capacity) with All Control Valves 100% Open Maximum Flow through Closest Coil

Minimum Flow through Most Remote Coil

CHW

HW

CHW

HW

CHW

HW

1

No balancing

20.5

44.4

143%

212%

73%

75%

2

Manual balance using calibrated balancing valves

0

0

100%

100%

100%

100%

3

Automatic flow limiting valves

Note 6

Note 6

100%

100%

100%

100%

4

Reverse-return

1.2

10.4

103%

150%

99%

85%

5

Oversized main piping

7.0

20.9

122%

173%

94%

82%

6 Undersized branch piping

19.5

NA

142%

NA

73%

NA

7 Undersized control valves

8.0

NA

120%

NA

86%

NA


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also results in the same high DPs as Option 1. This is because these valves only act to prevent flow rates above design; when the flow is below design, as it always is, other than during rare design load conditions, the valves do nothing. The last four columns indicate the condition of a transient warm-up or cooldown scenario where all the control valves are completely open. Proponents of balancing point to this condition as the justification for balancing: without it, parts of the building may warm up or cool down much more slowly than others due to flow imbalances. The numbers in each column indicate the percentage design flow. The numbers in parentheses indicate the percentage sensible coil capacity at this flow. Coils are very nonlinear so percent flow and percent capacity do not track each other. For the CHW coil with the lowest circuit pressure drop, in Option 1 (No balancing), the flow was 212% of the design flow, but this represented only 119% of design coil capacity. The worstcase cooling coil had only 73% of the design flow but 89% of its design capacity. The resulting capacities are so close to design (arguably within the accuracy of load calculation programs) that it is clear that system warm-up and or cooldown time will not be significantly affected by these flow imbalances. Table 4-2 summarizes the pump head, annual pump energy costs, and incremental first costs for each of the options. The first costs are relative to Option 1 (No balancing). Table 4-3 ranks each option with respect to controllability, pump energy costs, and first costs from best (1) to worst (8). Table 4-2

Balancing Method

Energy and Installed Cost of Balancing Options

Pump head, ft

Annual Pump Energy, $/yr

Incremental First Costs vs. Option 1 $

$ per design gpm

CHW

HW

CHW

HW

CHW

HW

CHW

HW

1

No balancing

58.5

82.7

$1910

$3930

–

–

–

–

2


60.3

83.6

$1970

$3970

$7960

$47,530

$6.60

$88.00

3


66.6

90.8

$2170

$4310

$11,420

$50,750

$9.50

$94.00

4

Reverse-return

55.3

80.0

$1810

$3800

$28,460

$17,290

$23.70

$32.00

5


45.0

59.3

$1470

$2820

$12,900

$7040

$10.80

$13.00

6

Undersized branch piping

58.5

NA

$1910

NA

($250)

NA

($0.20)

NA

7

Undersized control valves

58.5

NA

$1910

NA

($2340)

NA

($2.00)

NA

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Chapter 4 Hydronic Distribution Systems Table 4-3

Summary of Balancing Alternatives

Balancing Method

Controllability (All Conditions)

Pump Energy Costs

First Costs

No balancing

7

3

3


4

6

6


7

7

7

Reverse-return

2

2

5


3

1

4

Undersized branch piping

6

4

2

Undersized control valves

5

4

1

Pressure independent control valve

1

7

8

Recommendations for balancing of variable-flow systems are as follows: •

•

•

•

•

For other than very large distribution systems with high pump head, Option 1 (No balancing) appears to be the best option. It has a very low first costs, excellent energy performance, and minimal or insignificant operational problems. Using AFLVs is one of the most expensive options, yet it offers only the benefit of reducing imbalances during transients which, as shown in Table 4-1, has almost no functional impact on performance. Accordingly, this is generally the worst balancing option for variable-flow systems. Manual balancing with calibrated balance valves also adds to first costs with little benefit. They can also introduce coil performance problems under certain operating conditions (refer to Taylor and Stein [2002] for a detailed discussion). However, calibrated balance valves can be very handy for diagnosing flow problems because flow can be readily measured. If they are provided for this purpose, they should not be used for balancing (i.e., all valves should be wide open). For systems with long hours of operation, the added cost of oversized mains may be cost-effective, based on pump energy savings, but also offers considerable flexibility for future changes because future loads can be added anywhere to the system. This is an excellent option for campuses where future additions can never be well predicted. Reverse return is effective at reducing DP across control valves and also can reduce pump head, but the cost can be high and this option is unlikely to be cost-effective. However, reverse return is sometimes almost free (e.g., for floors with a loop distribution, reverse return can be achieved by looping the CHW supply in one direction, for example, clockwise, while looping the CHW return in the other, for example, counterclockwise). In these cases, reverse return should be used.


Figure 4-18

107

•

Undersizing piping and valves on the noncritical runs can reduce first costs but requires significant additional engineering time. If the branches are reduced too far, they could become the index runs and cause increased energy usage. The index run also may vary with varying time schedules (e.g., a coil may become the index run when other more remote coils are scheduled off).

•

For coils that will experience high and varying DPs (e.g., coils nearest to high-head distribution pumps), consider using pressure-independent control valves. They will provide more stable control and should reduce the time required to tune control loops. By providing more stable control, CHW flow rate can be decreased for the same sensible load (i.e., the same supply air temperature set point). Theoretically, there is exactly one flow rate (and one T) that will meet the desired set point given the coil entering air and water conditions at any moment in time. So, if standard valves could be tuned to provide stable control, pressure-independent valves would provide no improvement in flow and T. But because coils are nonlinear, fluctuating control that results in an average supply air temperature around set point will result in an average flow rate that is higher than the rate resulting from stable control, and thus average T will decrease. This can be seen in Figure 4-18, which shows pressure-independent characterized control valve (PICCV) performance versus a standard pressure-dependent globe valve. Notice on the right side how the average T of the globe valve is lower than that of the more stable PICCV. (This figure also shows improved T on the left side where control for both valve types is stable; this is not supported by physics and probably reflects some bias by the PICCV manufacturer who performed these tests.) Figure 4-18 shows a T improvement of about 10%. Coil modeling shows that to achieve that savings, supply air temperature of a typical cooling coil would have to vary by ±5°F, which is very unstable control. So, only those valves that are hard to tune due to high and varying DP greater

Pressure-independent valve performance (as rated by manufacturers).

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Chapter 4 Hydronic Distribution Systems than approximately 35 psi (80 ft)—as a rule of thumb—will benefit from pressure-independent control. The first-cost premium for these valves, currently very high but falling, is hard to justify in other applications.

Degrading T Syndrome The load is directly proportional to flow rate and T (the difference between return and supply CHW temperatures): Q = m· c p T

(4-1)

In most variable-flow CHW plants, it is assumed that T will remain relatively constant. If T is constant, it follows that the flow rate must vary proportionally with the load. However, in almost every real-world chiller plant, T often falls well short of design levels, particularly at low loads. Figure 4-19 shows data from a large central plant during cool weather (January through March); T is well below the 10°F design T. As T falls, flow rate must increase and no longer proportionally tracks the load. This phenomenon, often called degrading (or low) T syndrome, will impact plant energy performance in several ways: • •

Figure 4-19

Pump energy increases as flow increases. Chillers may have to be prematurely staged on in primary/secondary systems. As noted previously, primary flow must exceed secondary flow to

Example of degrading T syndrome.


•

•

109

avoid the death spiral that results in starved coils. If the secondary flow is high due to low T, primary pumps and chillers must be staged on prior to being fully loaded. If the primary pumps and chillers do not have VFDs, this will increase energy usage. The plant may be unable to meet coil demands despite adequate chiller capacity; it may run out of flow capability before running out of chiller capacity. For plants with CHW TES, TES storage capacity is reduced, possibly causing chillers to run in partial or peak demand periods (full storage) or increasing their load in these periods (partial storage). Causes of degrading T can be broken into three categories:

• • •

Causes that can be avoided by proper design or operation of the CHW system Causes that can be mitigated but through measures that may not result in overall energy savings Causes that are inevitable and simply cannot be avoided

These are summarized in the following subsections. More details can be found in “Degrading Chilled Water Plant Delta-T: Causes and Mitigation” (Taylor 2002).

Degrading T—Causes that Can Be Eliminated by Design/Operation • •

•

•

•

Improper set points or calibration: For example, dropping coil supply air temperature set point by only 2°F can double the flow rate and halve the T. Use of three-way valves: Three-way valves should pretty much never be used in variable-flow CHW systems. Some designers use them to ensure that water is “instantly” available at coils, but there is seldom a need: cooling loads change slowly, and, in a typical building, chilled water will reach a coil when its valve is opened after seconds or perhaps minutes, but definitely fast enough to avoid overheating. Three-way valves simply increase flow unnecessarily and reduce plant T. They also cost more than two-way valves. No control valve interlock: It is essential that control valves and associated control loops be disabled when the associated AHU is shut off. Otherwise the valve will naturally open as the supply air or room temperature cannot meet set point. This is often overlooked in DDC system programming. Coils piped backwards: CHW coils must be piped in an overall counterflow arrangement entering at the air discharge side of the coil. If piped backwards, the coil effectiveness is significantly reduced and the supply air temperature set point can almost never be reached. Uncontrolled process loads: Plants sometimes provide cooling to process loads such as medical and biological laboratory freezers. These devices must have automatic valves that shut off flow when the equipment does not need it.

110


•

•

Incorrectly selected control valves: Valves must have sufficient actuator power to shut off against the DP created by the pumps. This is less of a problem now with electric actuators and the common use of ball valves instead of globe valves—small pneumatic globe valves typically have only about 30 psi of shutoff head capability versus 200 ft for an electric ball valve. Oversized valves can also be an issue because they result in hunting; as shown in Figure 4-17, oscillating flow can increase average T. Again, modern control valves mitigate this problem; they have a rangeability of 50:1 up to 100:1, compared to 10:1 up to 15:1 for large pneumatic valves. Two-position (on/off) control valves, such as those used to control small fan-coil units (FCUs), are often blamed for low T problems. If these valves are not equipped with flow-limiting valves, or piped in a reverse return arrangement, they may consume more water flow when open than the design calls for. With full flow through the coil at partial loads, the T will invariably be lower than design. However, because the air temperature entering the FCU is fairly constant and is usually not subject to outdoor air conditions, the T will not degrade significantly. Incorrectly selected coils: When connecting to a central plant, it is essential for the designer to select coils that meet the minimum design T for which the plant was designed. It is not uncommon for engineers to select coils for a 10°F T simply because it is the basis of chiller AHRI standard ratings, instead of the 20°F or more assumed in the plant design. On campuses where there may be many engineering firms designing buildings over the years, it can be hard to police their coil selections. Consequently, undersized coils are often the most common source of degrading T in campus CHW plants. Improper bridge connection and control: Tertiary pump bridge connections (Figure 4-15) should be controlled off of the CHW supply temperature to the building at a set point that is a few degrees above the temperature of the water supplied by the central plant. If the set point is at or lower than the plant temperature, the bridge control valve will be wide open and water will recirculate directly to the return, substantially reducing T. Nonetheless, a proper set point will not help improve T if the coils downstream are not maintaining a high T. To resolve this problem, some designers move the temperature sensor controlling the two-way valve from the supply line to the return water line. The control valve is then modulated to maintain the return water temperature at design levels. If return water is too cold, it is recirculated back into the building in an attempt to make it absorb more heat. While instinctively it may make sense that recirculating water will increase the return water temperature, the return water temperature is driven primarily by the entering air temperature and the coil effectiveness, not the entering water temperature. Table 4-4 (based on an eight row, 96 fpf coil designed for 77°F entering dry-bulb, 62°F entering wet-bulb, and 55°F leaving dry-bulb temperatures) shows that for a given load, increasing entering CHW temperature results in a lower, not higher, leaving return water temperature. So, recirculating water not only increases the flow in


111

Coil Performance with Increasing CHWST

Entering CHW Temperature, °F

Flow Rate, gpm

T, °F

Leaving CHW Temperature, °F

42

30

16.7

58.7

44

34.5

14.7

58.7

46

41

12.3

58.3

48

53

9.5

57.5

the building tertiary loop, it also slightly increases flow in the secondary loop. Furthermore, the same death spiral that occurs with primary/secondary systems can occur here: as water is recirculated the T gets worse so more water is recirculated until the valve is closed and the system is fully recirculating water. Eventually, in this mode, the return water temperature will rise but only because coils are starved. Once return water temperature does rise above set point, the CHW valve will open, but it will soon close as the return water temperature once again drops. Clearly, this is not a good control strategy.

Low T—Measures that Improve T but with an Energy Trade-Off It must be remembered that degrading T is not a problem in and of itself—it is a problem because it increases plant energy usage. Therefore, solutions that improve T but do not result in an overall reduction in plant energy seldom make any sense. Here are two examples: •

•

Reducing CHW temperature: Reducing CHW temperature will definitely improve T; typically a 1°F drop in supply water temperature will increase T by 1.5°F to 2°F. That will reduce pump energy and perhaps improve efficiency due to improved chiller staging, but it will increase chiller energy because chillers are more efficient at higher leaving water temperatures. Simulations have shown that resetting CHW temperature downwards always results in a net increase in plant energy, even when pump head is high (greater than 100 ft). See Chapter 7 for more discussion. So, it is seldom a reasonable strategy to reduce CHW temperature to increase T in an existing plant. (With a new plant, there are substantial first-cost benefits to designing for large Ts and subsequent lower CHW temperatures; see Chapter 5.) Coil pumps: At design conditions, flow through cooling coils is typically in the fully turbulent flow regime. But as load falls and flow rate is reduced, flow will quickly fall into the mixed region and down into the laminar flow region. Theory suggests that in the laminar flow region, heat transfer rates will fall dramatically, resulting in reduced coil effectiveness

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Chapter 4 Hydronic Distribution Systems and reduced T. In fact, most coils do not behave that way. First, laminar flow is interrupted due to tube bends at the end of each row. Second, there is another factor occurring at the same time that more than offsets this rise in heat transfer resistance: at low flow rates, the coil is effectively oversized (i.e., the ratio of flow rate to coil surface area of is low). The water stays in the coil longer, and more heat transfer occurs, which causes temperature rise to increase rather than decrease. Figure 4-20 shows T as a function of space sensible load in a variable-air-volume (VAV) system for three coil types. The effect of increasing coil effectiveness due to the reduced flow rates overcomes the increase in film resistance so that coil T remains above design T except at the transition point into the laminar flow regime. Contrary to conventional thinking, T below the onset of laminar flow increases rather than decreases. But even if laminar flow did negatively affect coil performance causing degrading T, adding coil pumps would not be a good mitigation: Figure 4-21 shows pumping energy for a three-chiller plant first using primary/secondary pumping assuming T degraded proportionally with load, and second with coil pumps added assuming a constant T. The added energy of the constant-speed coil pumps more than offsets the reduced secondary pump energy resulting from improved T. So, coil pumps should never be used for the sole purpose of improving T.

Figure 4-20

Cooling coil T at part load.


Figure 4-21

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CHW pump energy—three-chiller plant.

Degrading T—Causes that Cannot be Eliminated Despite the best design, most plants experience some T degradation, particularly in mild weather. Two reasons are as follows: •

•

Coil heat transfer degradation: Coil heat transfer effectiveness is reduced by water-side fouling (e.g., slime, scale, or corrosion on the inside of coil tubes), air-side fouling (e.g., dirt buildup on coil fins), air-side deterioration (e.g., deteriorating fins), nonuniform air distribution across the cooling coil, and coil bypass air. Any reduction in coil effectiveness increases the flow rate of water required to deliver the desired leaving water temperature, thus reducing T. Coil cleaning, particularly on the air side, should be performed regularly and when fouling is visible. Air economizers and 100% outdoor air systems: Air-handling systems designed for high T (above about 14°F) that have integrated outdoor air-side economizers or supply 100% outdoor air will experience degrading T when the weather is cool but not cold enough to provide 100% of the system’s cooling load. Under these conditions, the air temperature entering the coil is low, causing correspondingly low return water temperatures. For instance, a coil might be designed for 80°F entering air temperature with a CHW return temperature of 60°F. When the outdoor air temperature is 60°F, it is clearly impossible to maintain a 60°F return water temperature. A coil on a VAV system designed for 44°F chilled water and an 18°F T would only be able to achieve a 11°F to 15°F T at 55°F to 65°F outdoor air temperatures.

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Designing Chiller Plants to Accommodate Low T The previous discussion shows that degrading T syndrome is caused by many conditions, most of which can be avoided by careful design and maintenance practices. There are, however, situations where degrading T syndrome is inevitable. Therefore, the plant must be designed to accommodate some T degradation while still operating efficiently. Coil effectiveness degradation can be mitigated by selecting chillers so that they can operate at ~1°F lower water temperature than that used to size the coils. For instance, if coils are selected for 43°F (see Chapter 5 for how this is determined), the chiller could be selected for 42°F supply temperature at the same T. This will allow the chiller to provide colder water on peak days without surging or tripping on high lift. The increase in pump energy caused by reduced T cannot be avoided. So, the issue is: how can the plant be designed to avoid the energy impact of premature staging of chillers and primary pumps? Common solutions include the following: • • •

Put a check valve in the common leg of primary/secondary systems. Use variable-speed chillers and pumps. Use primary-only distribution systems. Each of these is discussed in more detail in the following subsections.

Mitigating Degrading T—Check Valve in Common Leg When a chiller plant has multiple machines piped in parallel and the T is lower than design, the result is that one or more additional chillers must be activated, not because of load but because of flow. This causes an increase in primary CHW pump energy and condenser water pump energy as pumps are added to serve the newly activated chiller, and it results in more chillers operating at lower load, which reduces chiller efficiency for fixed-speed chillers. To mitigate this phenomenon, flow through the evaporators on the primary side must be increased to match the secondary flow. One way to do this is to insert a check valve in the common leg (Figure 4-22). If a low T situation occurs and the secondary flow increases above the primary flow of one chiller, the check valve will close and cause the secondary pump to drive the water through the primary pump (i.e., the primary and secondary pumps will be in series). The increase in additional flow through the primary pump depends on where the primary pump was operating on its pump curve, the steepness of the pump curve, and the excess horsepower of the secondary pumps. A primary flow-rate increase of 25% to 40% is usually possible. The pressure drop in the primary circuit increases roughly as the square of the flow, and pump power increases roughly with the cube of the flow, so flow increases are generally limited by available secondary pump horsepower. Adding a check valve is an excellent retrofit opportunity for existing plants with fixed-speed chillers that suffer from degrading T syndrome. Using a check valve in this manner is controversial. Some concerns are largely unsubstantiated, such as the following:


Figure 4-22

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Check valve in common pipe.

•

•

•

The check valve makes the primary and secondary circuits hydraulically dependent when the secondary flow rate exceeds the primary flow rate. Some purists find this heresy, but there is no intrinsic value to pressure independence in modern variable-speed, variable-flow systems. Evaporator flow may exceed the maximum limit published by chiller manufacturers. This is very unlikely to be true in practice, particularly if chillers are selected for a high T as recommended in Chapter 5. The amount of flow increase is limited by the cube law increase in pump power, as noted previously. But even if the maximum is exceeded, it should be noted that the maximum flow rate established by chiller manufacturers is largely an arbitrary limit based on rules of thumb to limit erosion; the chiller will not suddenly fail when that limit is exceeded. Because the hours that the system will be operating above the maximum will be few, erosion should not be an issue. When the primary and secondary pumps are in series, the additional pressure will force control valves open, overcooling spaces and wasting energy. This is simply not true when the secondary pumps have VFDs, as they always should. The pump controller will simply back off on the pump speed to maintain the desired DP in the system.

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Primary pumps may ride out beyond the end of their pump curves as the flow increases through them, causing cavitation. Again, due to the limitation in flow increase, this is seldom a problem and is one that can be avoided by picking pumps with lots of room on the right side of their curves so that they do not fall off the end even at 40% excess flow. However, there are some real issues with the check valve:

•

•

•

Deadheading secondary pumps: When primary CHW pumps are piped in a headered arrangement (Figure 4-17, Option B), if primary pumps are off and chiller isolation valves are closed, the check valve will stop flow through the secondary pumps and they will be deadheaded. Without flow, an operating pump will soon cause water within the pump to get very hot, damaging seals. The solution is to logically interlock secondary pumps to primary pumps in DDC system software; if primary pumps are off for more than 5 minutes when secondary pumps are on, shut the secondary pumps off. “Ghost” flow through inactive chillers: When primary CHW pumps are piped in a dedicated arrangement (Figure 4-17, Option A or C), the check valve can cause flow to be pushed through inactive primary pumps and chillers because the primary pump check valves allow flow in that direction. This water bypasses the operating chillers, raising the overall leaving water temperature, which can exacerbate the degrading T problem. The solution is to always use headered pumps (Figure 4-17, Option B) when using a check valve in the common leg. “Ghost” flow through inactive coils with coil pumps: When coils are controlled by variable-speed coil pumps in lieu of control valves (Figure 4-16 [Coil Pumping Scheme A]), the pressure drop across the check valve can induce a positive DP across the secondary loop, pushing water through inactive coil pumps and coils. This can add a load to the CHW system if these coils are part of air handlers operating in air economizer mode. The solution is to use low pressure drop swing check valves in the common leg, not spring-loaded “silent” check valves, which have a higher pressure drop.

Mitigating Degrading T—Variable-Speed Chillers and Pumps As noted in the previous subsection, degrading T can cause an increase in primary pump energy, condenser water pump energy, and chiller energy if chillers are staged on prematurely due to high secondary flow rather than load. But this assumes constant-speed pumps and chillers. If primary CHW pumps and condenser water pumps have VFDs, there will be very little increase in pump power when a new chiller and associated pumps are staged on if these pumps are controlled by flow, as recommended in Chapter 7. The speed of the primary pumps is varied to match the flow of the secondary system so that the flow in the common pipe is near zero. And as also explained in Chapter 7, variable-speed chillers are more efficient at low loads (as long as lift is reduced) than at high loads, thus staging chillers on prior to their being at full load is deliberate, not premature. Hence, efficiently accommodating degrading T is inherent in an all variable-speed chiller plants controlled using the sequences recommended in Chapter 7.


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Mitigating Degrading T—Primary-Only Distribution With primary-only distribution systems, the whole issue of backwards flow through the common leg disappears because there is no common leg. The chillers are generally staged on load alone, not flow. If degrading T is severe, it is possible the chillers will have to be staged on prematurely because the very high-pressure drop in the operating chillers (remember pressure drop increases as the square of the flow) can reduce the DP available in the distribution system, potentially starving coils. However, this should not be the case if the system is well designed (the “inevitable” causes of degrading T should not result in such severe degradation), and the problem is resolved by using variable-speed chillers and possibly variablespeed condenser water pumps, as recommended in Chapter 5.

Connecting Heat Recovery Chillers The heat rejected from the condenser of a chiller can be used for many purposes, including domestic water preheating, process heating, and building heating. Heat recovery chillers are usually sized for a small portion of the total cooling load because of the need to have a simultaneous mechanical cooling load and heating load and because of the lower cooling efficiency of heat recovery chillers. An unsatisfactory strategy for incorporating a heat recovery machine into a chiller plant is to pipe the chiller in parallel with the other chillers in the primary loop. The problem with this approach is that constant-flow primary chillers will almost always have a percentage of the cold supply water bypassed into the return, thereby decreasing the temperature of the water entering the chiller. This decreased entering temperature can diminish the heat recovery potential (cooling load) of the machine. In primary-only variable-flow systems where the flow through the evaporator is allowed to vary, this is not as much of a concern. Another method of dealing with a heat recovery chiller is to pipe it for preferential loading. Figure 4-23 shows a heat recovery machine piped in parallel with other chillers, but the location of the heat recovery primary pump suction pipe is such that it receives only the warmest return water from the system. Any bypass flow will go to the other chillers unless the heat recovery machine is the only one on. A problem with this approach is that there is no way to effectively unload the heat recovery machine during times when the heating load is low. The preferentially loaded machine will be required to cool its full volume of warm return water and, because the need for recovered heat is low, most of the heat will be rejected out the cooling tower. Mitigation includes providing a variable-speed primary CHW pump, which will allow the heat recovery chiller to be partly unloaded down to the minimum flow allowed by the chiller. Figure 4-24 shows a configuration for a heat recovery machine that puts it in series with the remaining chillers in the plant. Warm return water is pumped to the chiller, then back into the return, thereby precooling the inlet water to the other chillers. The heat recovery machine can remove as much or as little heat as is needed for the heating load.

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Figure 4-23

Preferentially loaded heat recovery chiller.

Figure 4-24

Heat recovery chiller in series.


Figure 4-25

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Thermal energy storage (TES).

Connecting Thermal Energy Storage (TES) Chilled water can be generated during utility off-peak periods and stored for use during on-peak periods to reduce energy costs. This is called chilledwater TES. Figure 4-25 shows an elegant way to include a stratified TES tank into a system: locate the tank in the common leg. In this position, the tank can be charged and discharged nonsimultaneously (full storage) or simultaneously (partial storage) without any opening and closing automatic valves.

Condenser Water Systems Introduction In designing energy-efficient central CHW plants, it is important to select the proper condenser water system. The efficiency of the chillers is affected not only by the operation of the cooling towers and associated pumps but also by the temperature and quality of the condenser water. In the next sections the following aspects of condenser water systems are discussed:

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Chapter 4 Hydronic Distribution Systems • • •

Options for piping multiple chillers and cooling towers (Piping Multiple Chillers and Cooling Towers) Integration of WSEs (Piping for Water-Side Economizers [WSEs]) Piping heat recovery chillers (Piping Heat Recovery Options)

Piping Multiple Chillers and Cooling Towers Condenser Water Pump Arrangements The common piping options for condenser water pumps are the same as for CHW pumps. See Options A, B, and C in Figure 4-17. The advantages and disadvantages are also similar. For ease of use, duplicate advantages and disadvantages are repeated here. The advantages of dedicated pumps for each condenser (Option A) include the following: •

•

•

The pump can be custom-selected for the condenser it serves. Pump selection can then take into account variations in condenser pressure drop and flow rates when chillers are not identical. This can reduce pump energy compared to Option B, where the head of each pump must be the same and sized for the condenser with the highest pressure drop; balance valves at the other condensers must be throttled to generate this same pressure drop. Controls are a bit simpler because the pump can be controlled using the contact provided with the chiller controller. This ensures that the pump starts and stops when the chiller wants it to. With Option B, the control of the isolation valves and pumps is by the DDC system and must be coordinated with the needs of the chiller controller to avoid nuisance trips. For instance, the pumps generally must run for several minutes after the command for the chiller to stop so that the chiller can pump down the refrigerant. Pump failures do not cause multiple-chiller trips. With dedicated pumps, if a pump fails, only the chiller it serves will see a flow disruption and trip. With Option B, all operating chillers will see a flow reduction when a pump fails, possibly causing more than one chiller to trip due to low flow or high refrigerant head. However, if there is a lag or standby pump with Option B that can be started quickly, trips can usually be avoided because it takes some time for refrigerant head to rise.

The advantages of headered (manifolded) pumps (Option B) include the following: •

•

Redundancy is improved. With Option A, if a pump fails and a chiller other than the one it serves also fails (albeit this is a rare event), then two chillers will be inoperative. With Option B, any pump can serve any chiller and under many conditions one pump can provide enough flow for two chillers to operate near full capacity. Including a standby pump is much simpler. Adding a standby pump to Option A is cumbersome and expensive because it requires extensive piping and manual or automatic isolation valves. If standby pumps are desired, Option B is the best option.


•

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Isolation valves can double as head pressure control valves. See the discussion on head pressure control below in the section Refrigerant Head Pressure Control. For Option A, head pressure control would require the addition of VFDs on condenser water pumps or tower bypass valves. It is easier to integrate a WSE. See the discussion on WSEs below in the section Piping for Water-Side Economizers (WSEs). Because WSEs are only operational in cold weather when loads are generally low, the condenser water side can use one (or more) of the condenser water pumps serving chillers rather than a dedicated pump. This reduces first costs.

Headered pumps with manual isolation valves (Option C) can have the advantages of Option A (although this option works best with identical chillers). It also mitigates the redundancy disadvantage of Option A, although accommodating a pump failure requires manual manipulation of valves versus the automatic response in Option B. Including a standby pump is possible with Option C, but it only works (depending on which pump fails) with the header isolation valves open, and chillers must be staged by manually opening and closing their isolation valves. First costs are usually lowest with Option A if the chiller and pump pairs are close coupled and the manual isolation valves between the two are eliminated (each chiller/pump pair is isolated for service as a pair). Costs can be higher with Option A if pumps are grouped together and not adjacent to the chillers due to long piping runs from the pumps to their dedicated chillers. Option C is usually less expensive than Option B, but Option B is usually the best choice where head pressure control and standby pumps are required. If pumps are constant speed with Option B, some engineers have concerns about overpumping condensers as pumps ride out their curves due to reduced losses in common piping. This is almost never a consideration because the maximum condenser flow rate is seldom reached, particularly if pumps are selected for a high T as recommended in Chapter 5, and the small increase in pump energy as the pump rides out its curve is offset by the improved chiller efficiency due to lower condenser-leaving water temperatures. There is, therefore, no need to limit condenser water flow at chiller condensers, and use of flow-limiting valves is strongly discouraged because they can easily become clogged in this open-circuit environment.

Equalizer Piping and Maintaining Sump Levels When piping multiple cooling towers, the water flow rate drawn from the sump is never exactly equal to the amount distributed into the inlet. This can lead to an overflowing collection basin with makeup water supplied to other basins and possibly to air entrainment into the suction piping if water levels get extremely low. It is virtually impossible to balance flow accurately enough to prevent over/underflow from occurring—even a few gpm imbalance is enough over time. To prevent this problem an equalizer must be provided between sumps, typically one of these two designs:

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•

Equalizer flume gates installed between tower basins: This requires that the tower basins be physically attached to each other. A removable cover is usually provided to allow one basin to be separated from the others for maintenance (e.g., cleaning the basin). Because the cover requires that the operator climb into the tower basin, this option is not as convenient for maintenance. Equalizer piping installed between tower basins: The equalizer line allows flow by gravity from one basin to the next. Because the force that moves the water through the equalizer line is the difference in water level between the sumps (which is sometimes just several inches), it is essential that the equalizer line be sized for a very low pressure drop. The equalizer line must be independent of the suction piping due to the various pressure differentials in the suction piping.

Cooling Tower Makeup and Level Control Typically, makeup water to cooling towers is provided by domestic or industrial water piped to a float valve in each tower basin. Float valves, which are similar to the valves in tank-type toilets, are not very reliable—they tend to get stuck open and leak. An alternative is to install an electronic level switch in each tower basin wired to an electric slow-closing solenoid makeup water valve. Another novel makeup water system design is summarized as follows: • •

Figure 4-26

No makeup water connection or assembly is provided with the cooling towers. The cooling tower equalizers are fitted with shutoff valves at each cell and a standpipe is installed in the common equalizer pipe as shown in Figure 4-26. The standpipe is open at the top and is fitted with an analog level sensor;

Cooling tower equalizer standpipe.


•

•

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capacitance type sensors are typically the most cost-effective. This is tied to the BAS. The level of water in the standpipe is the same as that in the cooling tower basins for any basin open to the equalizer. A basin that is valved off and drained (e.g., for cleaning) has no impact on the level sensor reading. Makeup water is piped through a slow-closing valve to the condenser water system in either the supply or return piping to the tower; it does not matter which, provided the makeup water pressure is higher than that in the piping and there are no automatic valves or check valves between the connection point and the towers. Typically, the connection will be indoors near the chillers. The valve is wired to the BAS. The level sensor is programmed in the BAS to open the makeup water valve as required to maintain the tower basin water level within the tower manufacturer’s recommended range whenever the condenser water pumps are on. It can also be programmed to provide high- and low-level alarms. The advantages of this design include the following:

•

•

• •

Makeup water piping need not be piped to each cell and generally need not be piped outdoors. This is a particular advantage in freezing climates because it obviates the need for heat tracing and insulating the makeup water piping. The analog sensor provides control as well as high- and low-level alarms with set points easily adjusted in software. False alarms are virtually eliminated; when level switches are provided at each cell for alarming, a false alarm is generated if the cell is drained for cleaning. (California Title 24 energy standards [CBSC 2016] require a high-level alarm to reduce water waste; this design meets that requirement.) Float valves are eliminated. Costs are lower than electronic makeup water controls, especially if they include high- and low-level alarms to the BAS. Costs are also typically lower than float valve makeup water when the system has more than two cells or is located in a freezing climate.

Maximum and Minimum Flow Rates When water enters the cooling tower, it is distributed uniformly across the fill by means of spray nozzles or a gravity distribution basin with properly sized nozzles. Each cell of a cooling tower has a maximum and a minimum flow rate. The maximum flow rate is that required to prevent overflow of gravity distribution systems or excessive spray through nozzles. The minimum flow rate is that required to ensure that tower fill is fully wetted. See Chapter 3 for more details. In plants with multiple cooling towers and chillers, it is desirable to stage condenser water pumps with the chillers (Chapter 7) so there will be times when one condenser water pump will operate alone. This will reduce the flow rate through cooling towers. Options for maintaining minimum flow rates (Figure 4-27) include the following:

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(a)

(b)

(c)

Figure 4-27

Cooling tower cell isolation options: (a) Option A, weir dams and/or low-flow nozzles; (b) Option B, automatic-isolation valves on supply only; and (c) Option C, automatic-isolation valves on supply and suction.

Source: Taylor 2011b.

•

Option A: Select tower weir dams and/or nozzles to allow one pump to serve all towers. For systems with two or three tower cells this can eliminate the need for isolation valves, which cost much more than the weir dams and nozzles. This option is also the most efficient; tower energy usage is minimized by operating as many cells as possible, particularly when tower fans are controlled by VFDs. This is because fan speed is


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reduced (reducing fan power by almost the cube of the speed) and cooling is achieved through tower cells even when fans are off. With most manufacturers and tower types, nozzles and dams are available to reduce flow by 50%, and many can go down to 33% or even 25%, depending on the selection and design flow rate. Because of low cost and high efficiency, this option should always be the first choice. When a plant has many tower cells and automatic isolation valves are unavoidable, the dams and nozzles should still be selected to allow as many cells to operate as possible. •

Option B: Install automatic isolation valves on supply lines only. This option uses the equalizer to keep basin levels between overflow and fill lines and will require that equalizers be oversized from that required for normal duty. For example, assume there are three tower cells and only one is active; supply flow to the others is shut off. However, water is drawn out of all three cell basins because the suction lines have no automatic isolation valves. The water level in the basin of the cell that is supplied will rise while the other two basin levels will fall. The difference in the two elevations must provide enough head for water to transfer from the supplied cell to the others through the equalizer. If the equalizer is undersized, water will overflow in the supplied cell and the others will be drawn so low that makeup water valves open, wasting water and water treatment chemicals. There are only a few inches of elevation difference between the overflow and fill lines, so it is imperative that the equalizer be properly sized for this option to work. Another option is to eliminate the basins at each tower and use a common sump, often located indoors in cold climates. This avoids the need for equalizer lines entirely but is much more expensive.

•

Option C: Install automatic isolation valves on both supply and suction lines. This is usually the most expensive option because automatic valves are expensive relative to an incremental increase in equalizer size. This design also increases exposure to a valve failure—an oversized equalizer line has no failure modes. It also increases the risk of freezing (or increases the energy used by basin heaters) in the basins of inactive cells in systems that must operate in cold weather. However, this is often the best option when there are many tower cells that are not located close together (i.e., when equalizer lines would be too long for Option B).

Start-Up Conditions After a condenser water system has been shut down, water will drain by gravity from the inlet piping into the cooling tower basin. Depending on the size of the lines, this volume could be enough to overflow the sump, thereby wasting valuable treated water. Conversely, when the pump starts, there needs to be enough water in the sump to fill the empty piping without drawing the volume so low that air is entrained into the suction piping. The amount of water moved back and forth during start-up and shutdown must be minimized by one or more of the following options:

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Chapter 4 Hydronic Distribution Systems • •

•

•

Locate the tower at the highest elevation of the system. This prevents water from draining back to the tower when pumps shut off. Install inverted traps at the tower return inlet. This is recommended if there is a significant amount of piping at or just above the elevation of tower return. The return connection is typically self-venting, so any water in piping at that level will drain into the tower when the pump stops. An inverted trap with the top above the highest piping elevation prevents any water upstream of the trap from draining back to the tower. Use automatic valves that shut off when the pumps shut off. Valves will be necessary if there is significant piping and equipment located at a higher elevation than the towers. Properly size the sump volume. This is a practical option when the sump is field built. It is not an option on factory-built towers.

Refrigerant Head Pressure Control All chillers require a minimum refrigerant head (lift) between the evaporator and condenser. This can be quite high for most screw chillers and some hermetic centrifugal chillers and very low for magnetic bearing or ceramic bearing chillers, which have no oil return considerations. There are two common reasons low refrigerant head pressure can occur: •

•

Cold water in the cooling tower basins at start-up: Some chillers can operate for a short period of time with low start-up head while others will trip on low head pressure safeties almost immediately. To determine if head pressure control is required for cold starts, consult with the chiller manufacturer. Integrated WSE operation (discussed below in Piping for Water-Side Economizers [WSEs]): Head pressure control is almost always mandatory because cooling tower water temperatures are deliberately kept very cold for long periods. Options to avoid low head pressure problems include the following:

•

Tower three-way bypass valves: The bypass water is diverted around the tower fill into the cooling tower sump or into the suction piping, thus avoiding the natural cooling that occurs across the tower fill even when tower fans are off. Piping the bypass to the suction line also avoids the mass of water in the basin for an even faster warm-up, but the design can be problematic: unless the bypass line is balanced to create a pressure drop equal to the height of the cooling tower, air will be drawn into the system backwards from the spray nozzles because piping above the basin will fall below atmospheric pressure. For staged or variable condenser water flow systems, the bypass must be balanced at the lowest expected flow rate. This creates a high-pressure drop and reduced flow if more pumps operate. However, reduced flow is acceptable when the intent of the bypass is to raise head


•

•

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pressure. The bypass valve is controlled by supply water temperature typically with a low limit set point well below the normal set point used to control tower fan on/off and speed. Tower bypass is most commonly used where towers must operate in very cold weather to avoid freezing in the fill itself. The other options below are less expensive and therefore preferred in other applications. For systems with dedicated condenser water pumps (Option A or C, Figure 4-17), VFDs on the pumps can be used to reduce water flow to the chiller. Head pressure can be maintained even with very cold supply water as long as the flow rate can be reduced so that the condenser refrigerant pressure can be high enough (head pressure depends on the condenser water temperature leaving the chiller, not entering the chiller). Pump speed can be controlled by the temperature leaving the condenser at a set point that corresponds to minimum condenser pressure or (preferably) by a signal from the chiller controller indicating head pressure needs—most chiller controllers have an analog output dedicated for this purpose. For systems with headered pumps (Option B, Figure 4-17), the isolation valves can double as head pressure control valves by converting them from two position to modulating. Valve position is typically controlled by the chiller controller head pressure control analog output, either directly or through the digital control system. This signal will close the valve when the chiller shuts off.

The second two options above reduce flow through the condenser. Many engineers have concerns about low condenser water flow contributing to fouling of the condenser tubes, but there is little definitive evidence to support the concept that high velocity keeps tubes clean; strainers and sidestream filters that prevent particles from entering the condenser in the first place are preferred. However, even if this is an issue, for most head pressure control applications there are few hours at reduced flow—only during cold starts—so the impact on tube fouling should not be significant. Low flow through the cooling tower may also be an issue (see discussion in Chapter 3) but, again, it should not be given the short duration.

Piping for Water-Side Economizers (WSEs) WSEs are an alternative to air-side economizers. Air-side economizers are usually more energy efficient, but they are not always practical and can be much more expensive. Applications where WSEs are often preferred include floor-by-floor air handlers in a high-rise office building or computer room air handlers serving a large data center. A WSE uses cold water generated at the cooling tower to produce chilled water without, or with reduced, mechanical refrigeration. This is accomplished by running the cooling towers to produce water temperatures typically 45°F and less during periods of low ambient wetbulb temperatures. The cold water is pumped through a high-effectiveness water-to-water HX, usually a plate and frame type, to produce chilled water at temperatures of 50°F or less. The HX protects the CHW system from the corrosion, dirt, and debris typical of open-circuit condenser water.

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Chapter 4 Hydronic Distribution Systems For detailed design guidance on sizing WSE HXs and flow rates, see Chapter 5. Figure 4-28 shows a non-integrated WSE where the economizer HX is piped in parallel with the chiller evaporators on the CHW side. This design allows the economizer to operate only if the chillers are not operating and vice versa—they cannot operate together. This design was the most common when WSEs first became popular in the 1980s, but it is not very efficient and is no longer allowed to be used by energy standards such as ANSI/ASHRAE/IES Standard 90.1. Instead, WSEs must use an integrated piping arrangement shown in Figure 4-29 for a primary/secondary system and in Figure 4-30 for a primary-only system. Integrated systems, which cost only slightly more than non-integrated systems, allow simultaneous operation of the chillers and the economizer because the HX is piped in series with the chiller evaporators on the CHW side. The economizer can provide some precooling of the return CHW temperature even if it cannot provide all of the cooling. This substantially extends the number of hours the economizer can be operational. Controls are also much simpler for integrated economizers.

Figure 4-28

Water-side economizer (WSE), non-integrated.



Figure 4-29

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WSE, integrated, primary/secondary.


Figure 4-29 shows two options for how to provide flow through the HX. The least expensive option is to place a two-position valve in the CHW return line. The valve closes when the economizer is enabled and is open otherwise. This option requires that secondary pumps have VFDs so that they can slow down when the HX is out of the circuit and vice versa. The secondary pumps generally do not need to be sized for the added head of the HX because the HX will be in the loop only when the economizer is active and cooling loads (and flows) are low. If secondary pumps are constant speed (rarely true in modern plants) or if the design flow rate through the HX is much lower than the expected CHW flow during economizer operation, a sidestream pump should be used instead of the two-position valve. This sidestream pump is sized with enough head to draw water out of the secondary return, pump it through the HX, then back to the return. In both the integrated and non-integrated designs, the HX is generally not provided with its own condenser water pumps. Because the load will be low when the weather is cold enough for the towers to deliver cold water, it should

130


Figure 4-30

WSE, integrated, primary only.


not be necessary to run all chillers, so one of the chiller condenser water pumps can serve the HX. The HX should be selected so that its pressure drop is similar to the pressure drop across chiller condensers. Figure 4-30 includes a bypass WSE-only valve. It is opened when the WSE is able to provide cold enough water that none of the chillers need to operate, thus reducing pressure drop through the system. When using WSEs, refrigerant head pressure control is required (except for with some magnetic bearing chillers) because of the cold water coming off the cooling tower. See the discussion in the section Refrigerant Head Pressure Control regarding head pressure control options.

Piping Heat Recovery Options Heat rejected from chillers can be used in numerous ways, including preheating domestic hot water and—with the use of double-bundle heat recovery chillers—heating buildings. In the case of preheating domestic hot water, the


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condenser water is routed through a double-wall HX that is either an integral part of a storage tank or is remotely located with a circulation pump to the storage tank (see Figure 4-31). For double-bundle heat recovery chillers, the piping of the cooling tower side of the double-bundle condenser involves using a three-way valve that controls the water temperature leaving the heat recovery side of the double-bundle condenser (see Figure 4-32). In a double-bundle heat recovery condenser, the hot gas from the compressor first enters the heat recovery side of the condenser where the building’s heating system removes the heat at a suitable temperature (105°F to 130°F). If all of the heat from the chiller is not rejected in the heat recovery bundle, the leaving heating water temperature (and refrigerant pressure) will rise above set temperature. This will cause the temperature controller to modulate the three-way valve on the cooling tower side of the condenser to maintain set temperature. A balance valve must be provided on the bypass line that goes back to the pump suction. Centrifugal heat recovery chillers have limited unloading capability when in the heat recovery mode due to the high condensing temperature level. If the cooling load is small relative to the design chiller capacity, HGBP must be used to prevent surge, severely increasing energy usage. Therefore, where heat recovery

Figure 4-31

Preheat of domestic hot water.

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Figure 4-32

Piping for double-bundle heat recovery chiller.

chillers are used, it is important to install one or more non-heat-recovery chillers. The more efficient non-heat-recovery chillers can be operated at low loads when there is insufficient load to keep the heat recovery chiller on without HGBP. They can also be run when the cooling load exceeds the capacity of the heat recovery chiller. Using chiller heat recovery for space heating and using economizers (air or water) are generally mutually exclusive because the economizers will keep the chillers from operating in cold weather, so there is no condenser heat to recover. During cold weather when the heating load is equal to or greater than the amount of heat rejected from the chillers, it can be shown that using heat recovery chillers is more energy efficient than air-side economizers and gas-fired heating systems. In most commercial buildings, the cooling load will be small in cold weather because only the interior zones need cooling and their loads are usually relatively small. In this case, heat recovery will be more efficient than economizers. But as the weather gets milder, heating loads get smaller and cooling loads can get larger as some sunny perimeter zones switch from requiring heating to requiring cooling. At this point, economizers begin to outperform heat recovery systems. On an annual basis, economizer systems tend to be more energy efficient in mild climates because of the following: • •

The heating season is relatively short and mild. Integrated economizers reduce energy usage even when heating is not required. For instance, in mild weather (55°F to 65°F), integrated economizers reduce the cooling load, which can be substantial because both interior and perimeter zones require cooling in this case, while heat recovery systems will do little because heating loads are very small or nonexistent. In California’s mild climate, for instance, a substantial number of building operating hours fall into this temperature range. (In very mild climates like San Francisco’s, more than 85% of the operating hours fall into this range.)


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As noted previously, heat recovery chillers have limited unloading capability when in the heat recovery mode, often requiring the use of HGBP. When using HGBP, the chiller is essentially acting like a costly electric resistance heater. Heating systems using recovered heat must be designed for low temperatures (e.g., 110°F to 130°F) and low-temperature differences because of the limits of the heat recovery condenser. This will increase hot-water flow rates and may require larger heating coils, increasing both air- and water-side pressure drops and thus increasing fan and pump energy. If there is a large constant heating load, such as that for domestic hot water in a hotel, heat recovery will probably outperform economizer systems. A detailed computer analysis would be required to evaluate the two design options in this application. It is important to include maintenance costs in the analysis because heat recovery systems require the chiller to operate all day long, all year long, increasing the maintenance costs and reducing the service life of this expensive machine. It is possible to combine economizers and heat recovery to maximize energy savings. The heat recovery mode is used, with economizers locked out, when the heating load is large enough (e.g., when outdoor air temperature [OAT] < 45°F) and the economizer mode is used during milder weather. However, the first costs of this design can be prohibitive.

Plant Layout Figure 4-33 shows a CHW plant with primary-only pumping. Conceptually the design is fine, but the designer of this plant unnecessarily increased first costs. First, reverse-return piping was provided for both condenser water and CHW systems. Reverse return, where the first device supplied is the last returned, is a piping scheme intended to self-balance hydronic systems but, more importantly, to keep the pressure drop across modulating two-way control valves relatively low and equal among all valves as they open and close with changing loads (see the Balancing Variable-Flow Systems section). However, in this case, the valves at each chiller are essentially two-position and the difference in pressure across each valve is very small even if the system is direct return (first supplied, first returned) and was not manually balanced. A balanced direct-return piping system will result in the same flow rates across each chiller as the reverse-return design, regardless of how many chillers are enabled. So, there is no value in performance to reverse return in this application, yet it substantially increases first costs. Another often expensive design concept is to group pumps together and pipe all the pumps first to a larger common pipe before distributing the supply water to the chillers. In this example, the common pipe size is 12 in. on the CHW side and 14 in. on the condenser water side. It is not uncommon to see all the pumps squished into a corner of the chiller room. There is little synergy to grouping pumps next to each other, and doing so can increase first costs and reduce space around the pumps for maintenance.

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Figure 4-33

Expensive CHW plant.

Figure 4-34 shows the same plant as Figure 4-33 but instead of reverse-return and grouped pumps, the pumps are aligned with chillers and are piped into a common header on the discharge side of the pump and also on the discharge side of the condensers. (The discharge side of the evaporators is the same as that in Figure 4-33.) This design still allows any pump to serve any chiller, but it shortens piping runs and it reduces pipe sizes because there is no common pipe that serves the total system flow. On the condenser water side, all pipes are 10 in.; the 14 in. common pipe is eliminated. Similarly, the 12 in. CHW piping at the discharge of the pump to the evaporators is eliminated; all pipes are 8 in. First costs are substantially reduced with no impact on performance. This layout also usually reduces the footprint of the plant, reducing the floor area required for the chiller room. In fact, locating the equipment as close together as possible is key to the first-cost savings. Figure 4-35 shows a


Figure 4-34

135

Less expensive CHW plant.

floor plan of the plant shown schematically in Figure 4-34. The CHW and condenser water pumps align with the chillers but are offset to provide tube pull space and piping connections between the pumps. This is a very compact layout, but it still provides adequate maintenance access for all equipment.

References ASHRAE. 2016. ANSI/ASHRAE/IES Standard 90.1-2013, Energy standard for buildings except low-rise residential buildings. Atlanta: ASHRAE. CBSC. 2016. 2016 California building standards code. California Code of Regulations, Title 24. Sacramento, CA: California Building Standards Commission.

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Figure 4-35 CHW plant floor plan.


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Taylor, S. 2002. Degrading chilled water plant Delta-T: Causes and mitigation. ASHRAE Transactions 108(1). Taylor, S., and J. Stein. 2002. Balancing variable-flow hydronic systems. ASHRAE Journal 10. Taylor, S. 2011a. Optimizing design & control of chilled water plants: Part 1: Chilled water distribution system selection. ASHRAE Journal 6:14–25. Taylor, S. 2011b. Optimizing design & control of chilled water plants: Part 2: Condenser water system design. ASHRAE Journal 9:14–25.

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Aggressive CHW temperature reset a. Causes issues with thermal comfort because higher supply water temperature yields significantly higher supply air moisture content for a given set of entering air conditions and thus higher space humidity. b. Typically increases pump energy usage significantly enough to offset the benefit of the decrease in chiller energy use. c. Has little to no impact on space humidity control. d. Both (a) and (b).

4-2

A plant consists of two identical fixed-speed centrifugal chillers, each with a dedicated constant-speed primary CHW pump. The chillers supply chilled water to one large built-up air handler that primarily serves daytime commercial office space loads and another large air handler that serves an auditorium space most frequently occupied in the evening. Both air handlers have threeway CHW control valves and are of approximately equal size. Which of the following are true? i. The design will require two chillers to operate when the auditorium air handler is at full load, even if the office air handler is off. ii. The plant will operate least efficiently in the rare instances that both the office air handler and auditorium air handler are at near design load. iii. The plant will operate most efficiently in the rare instances that both the office air handler and auditorium air handler are at near design load. iv. Headering the primary pumps would increase the controls complexity with no benefit in system redundancy. a. b. c. d.

4-3

(i), (iii), (iv) (i), (iii) (ii), (iv) (i), (ii)

For three-way valve systems a. The flow rate through the branch serving the coil is constant, irrespective of valve position. b. The flow rate through the branch serving the coil peaks when the valve is fully open to the coil. c. The flow rate through the branch serving the coil peaks when the valve is 50% open. d. Balancing of the bypass leg is never necessary.

Fundamentals of Design and Control of Central Chilled-Water Plants I-P 4-4

139

The back-loaded position of the common leg shown below a. Causes one chiller to be almost fully loaded with the remainder of the load handled by the other chiller. b. Causes unbalanced flow through the two chillers. c. Results in the same energy performance as a system with a common leg located in the normal position just upstream of the secondary pumps. d. Is a reasonable location for most plants if it is less expensive due to the physical layout of the plant.

4-5

Construction of a variable primary CHW plant is just about complete when it is discovered that the design includes only two-way valves and no means to maintain minimum flow. Which of the following last-minute design change options will resolve the problem at minimum cost? a. Install a CHW bypass locally at the CHW plant. Measure DP across the chillers to indirectly measure flow. b. Install a CHW bypass locally at the CHW plant. Install a flowmeter in the main return line at the plant to measure flow. c. Install a CHW bypass at the end of the line. Install a flowmeter in the main return line at the plant to measure flow. d. Install enough three-way valves at end-of-line coils to maintain minimum flow.

4-6

True or false? Both headered and dedicated pump per chiller configurations are equally appropriate for plants requiring a standby pump. a. True b. False

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Which balancing method is most appropriate for all but very large distribution systems? a. Manual balancing using CBVs to measure and adjust flow. b. No balancing. c. AFLVs at all coils. d. Reverse-return piping.

4-8

True or False: Air-side economizing systems can contribute to low CHW T issues in systems with high T designs. a. True b. False

4-9

What factors constrain the number of cooling towers that can be operated with a given number of constant-speed condenser water pumps enabled? a. Minimum per-tower flow requirements. b. Maximum per-tower flow requirements. c. Neither (a) nor (b). d. Both (a) and (b).

4-10

Isolating cooling towers by means of isolation valves on the tower supply piping only a. Requires that the equalizer be oversized. b. Does not require the equalizer to be oversized. c. Requires that all towers be operated whenever the plant is enabled. The isolation valves are only installed to prevent tower overflow upon plant shutdown. d. Is not a viable design option.

4-11

True or false? Most modern centrifugal chillers can operate with an integrated WSE without any means of head pressure control. a. True b. False

Optimizing Design


Introduction Previous chapters discussed the basic principles behind central CHW plant components and distribution system design. This chapter provides procedures and analysis techniques for optimizing the design to minimize first costs and operating costs (in particular, energy costs) over the plant’s life cycle. This chapter primarily applies to new CHW plants, but many of the techniques can be used for retrofit projects as well. To rigorously optimize a central plant design would be a Herculean task due to the almost infinite number of design decisions that affect energy costs and first costs. For instance, energy costs are determined by the • • • • • • •

full-load and part-load/part-lift efficiency of each piece of plant equipment (e.g., chillers, towers, pumps), quantity and staging of each type of equipment, design of the distribution system (e.g., variable flow versus constant flow, primary only versus primary/secondary), control sequences, pipe and valve sizing, flow rate sizing, and more!

First costs can be even more complex to account for during initial design. There are many reasons for this, including the fact that • • • •

costs are not a continuous function of capacity, capacity for some equipment and materials is only available in discrete sizes, costs vary by manufacturer and by market conditions, and costs vary widely depending on the physical layout of the plant and other design details.

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Chapter 5 Optimizing Design Rather than trying to account for every design variable, this chapter suggests a chiller plant design approach that combines detailed analysis and ruleof-thumb recommendations. This approach provides much better results in terms of plant performance and cost compared to traditional design procedures with little or no more effort.

Design Procedure Rigorous optimization of a CHW system would require a mathematical model of each possible system component and design option, accurately describing and defining its operating performance and first costs. Unfortunately, this approach is not practical; there are simply too many options and their cost and performance cannot always be described by continuous mathematical functions. Nevertheless, it is possible to partially optimize the chiller plant with a reasonable engineering effort. The key is to break the chiller plant into subsystems and then optimize within those subsystems. The plant will not be completely optimized due to the complex interactions between subsystems, but the result should be close to optimum. Before beginning the detailed design process, however, it is necessary to first develop plant load profiles as described in Chapter 2. It is also essential that the designer be very knowledgeable about chiller plant equipment (see Chapter 3) and hydronic distribution systems (see Chapter 4). For most chiller plants, near-optimum plant design can be achieved by the following step-by-step procedure: 1. Select CHW distribution system. 2. Select CHW temperatures, flow rate, and primary pipe sizes. 3. Select condenser water distribution system. 4. Select condenser water temperatures, flow rate, and primary pipe sizes. 5. Select cooling tower type, speed control option, efficiency, and approach temperature, and make cooling tower selection. 6. Select chillers. 7. Finalize piping system design, calculate pump head, and select pumps. 8. Develop and optimize control sequences. Steps 1 to 5 are discussed in this chapter. Chiller selection, Step 6, is discussed in Chapter 6. Step 7 is discussed in this chapter and Chapter 4. Control system design, Step 8, is discussed in Chapter 7.


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Selecting Chilled-Water Distribution System Flow Arrangement Simplified Distribution System Selection Table 5-1 lists recommendations for “life-cycle cost optimum” distribution systems based on the size and number of loads served and the number of chillers. “Life-cycle cost optimum” is in quotation marks because these recommendations are generalizations that should apply to the majority of typical HVAC applications but may not prove to be optimum for every application and have not been rigorously proven as the best choice. The recommendations are based on the author’s design and commissioning experience, analysis that was done in conjunction with development of this manual, work done on an earlier CHW plant design manual (EDR 1999), and the prescriptive requirements of ANSI/ ASHRAE/IES Standard 90.1 (2016b). The intent is to allow designers to select the system that is most often the best choice from a life-cycle cost perspective for a given application without having to perform any lengthy analyses. Figures of CHW distribution systems are repeated from Chapter 4 for convenience. In each figure, multiple chillers are shown in parallel. For most applications with two—or a multiple of two—equally sized chillers, the chillers could alternatively be piped in series. This results in lower chiller energy usage, partly Table 5-1

Chilled-Water Distribution System

Application Number

Number of Coils/Loads Served

Number of Chillers

Size of Coils/ Loads Served

Control Valves

Recommended Distribution Type

1

One

Any

Any

None

Primary only/single coil (Figure 5-1)

2

More than one

One

Small (≲100 gpm)

Two way and three way

Primary only/single chiller (Figure 5-2)

3

Few coils serving similar loads

More than one

Three way

Primary only/ multiple chillers/few coils with similar loads (Figure 5-3)

Small (≲100 gpm)

Two way

Primary-only variable flow (Figure 5-4) or Primary/secondary (Figure 5-5)

Small (≲100 gpm)

4

Many coils serving similar More than loads or one any serving dissimilar loads

5

More than one

Any

Large campus

Two way

Primary/distributed secondary (Figure 5-6)

6

More than one

Any

Large coils (≳100 gpm)

None

Primary/coil secondary (Figure 5-7)

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Chapter 5 Optimizing Design offset by higher pump energy usage. However, first costs are usually higher with series piping due to larger piping and pumps and bypass piping typically provided to allow one chiller to operate while the other is down for maintenance. Because of limited funding, the life-cycle costs of series piping were not evaluated, and thus series piping is not included in the recommendations in Table 5-1. This option will be evaluated for cost-effectiveness in future versions of this textbook. Note that air-cooled chillers using DX evaporators are not able to achieve the high Ts recommended below in the section Optimizing Chilled-Water Design Temperatures. For those applications, series chillers may have lower first costs than chillers piped in parallel due to lower piping costs and, thus, are likely to have lower life-cycle costs than a system using parallel piping.

Primary Only/Single Coil With one or more chillers serving a single cooling coil (Figure 5-1), the simplest design strategy is to not use any control valves at the coil. Instead, a constantvolume pump circulates water between the chiller and the coil, and supply air temperature is controlled by resetting the temperature of the chilled water leaving the chiller. While constant CHW flow results in constant pump energy, chiller performance is improved when the leaving CHW temperature is reset to be as high as possible. A VFD could also be added to the pump to make the system variable flow, but that adds cost and complexity. VFDs are seldom cost-effective because pump power is typically small in a single-coil plant because the chiller and coil are usually close coupled and it is more efficient to increase CHW temperature than to reduce pump speed and pump energy. (Chapter 7 further discusses the trade-off between resetting CHW temperature and pump energy.) Figure 5-1 shows a single chiller, but any number of chillers can be used. When two chillers are used, this is a good application for piping chillers in series rather than in parallel.

Figure 5-1

Primary only/single coil.


145

Primary Only/Single Chiller Small CHW plants commonly have a single chiller, typically air cooled. Single-chiller plants do not have to deal with flow and staging problems common to multiple-chiller plants and thus can have a simple distribution and control system. The recommended design is shown in Figure 5-2. It is the simplest variable-flow primary-only system. Two-way valves are installed at most coils with just enough three-way valves installed to maintain the minimum flow required by the chiller. This minimum rate, which can be obtained from the manufacturer, will vary with design CHW flow rate and the chiller type, size, and manufacturer but is typically 25% to 50% of the design flow. A VFD is shown in Figure 5-2; VFDs are typically cost-effective, except on very small systems. Note that ANSI/ASHRAE/IES Standard 90.1 requires VFDs on CHW pumps exceeding 5 hp (2016b). The VFD is controlled by a DP sensor located near the most remote coil so that the DP set point can be as low as possible; this is also a requirement of ANSI/ASHRAE/IES Standard 90.1. Locating the sensor near the pump requires a high DP set point and eliminates most of the energy savings from the VFD. See Chapter 4 for a discussion of considerations for locating the three-way and two-way valves.

Figure 5-2

Primary only/single chiller.

146

Chapter 5 Optimizing Design

Primary Only/Multiple Chillers/Few Coils with Similar Loads When systems have multiple chillers, chiller staging can be a problem when flow and load do not track, and they generally do not when three-way valves are used. But three-way valves (Figure 5-3) can work well with this type of system so long as all coil loads tend to vary in the same proportion, as they might if all coils serve similar occupancies (e.g., all serve offices on the same schedule). For instance, if the coils served are below half load and only one chiller and pump are operating, all coils will be capable of meeting their loads. The system is thus a quasi-variable-flow system in that pumps and chillers can be staged. Also, because the loads vary similarly, CHW temperature may be reset aggressively, which allows the plant to be about as efficient as one of the true variable-flow systems discussed below in the next section Primary-Only Variable-Flow and Primary/Secondary Systems. So, this system is a reasonable choice for small applications with only a few coils serving similar loads—it is simple and inexpensive and avoids all of the complexities of variable-flow systems. Also, small systems like this are typically close coupled, so there is not much pump energy to save. Note that ANSI/ASHRAE/IES Standard 90.1 only allows this approach for systems with three coils or fewer or a total CHW pump system power of 10 hp and less (2016b).

Figure 5-3

Primary only/multiple chillers/few coils with similar loads.



147

Primary-Only Variable-Flow and Primary/Secondary Systems Application 4 in Table 5-1 is probably the most common. It applies to systems serving many small coils (or a few coils with dissimilar loads) and more than one chiller. In this case, either of two systems is recommended: primaryonly variable flow (Figure 5-4) or primary/secondary (Figure 5-5). Both systems have positives and negatives, as summarized in Table 5-2. Primary-only systems always cost less and take up less space than primary/ secondary systems, and, with VFDs, primary-only systems also always use less pump energy than traditional primary/secondary systems. The pump energy savings are due to the following: •

Figure 5-4

Reduced system head as a result of the elimination of the extra set of pumps and related piping and devices (shutoff valves, strainers, suction diffusers, check valves, etc.)

Primary-only variable flow.


148


Figure 5-5

Primary/secondary.


Table 5-2

Advantages and Disadvantages of Primary-Only Versus Primary/Secondary Systems

Advantages of Primary Only

Disadvantages of Primary Only

Lower first costs Less plant space required Reduced pump peak power Lower pump annual energy usage

Complexity of bypass control Complexity of staging chillers

•

•

More efficient pumps. The primary pumps in the primary/secondary system will be inherently less efficient due to their high flow and low head. This can be partially mitigated by using larger pumps running at lower speed but with an increase in first costs. Variable flow through the evaporator, which allows flow to drop below design flow to some minimum flow rate prescribed by the chiller manufacturer. VFDs can be added to the primary pumps of a primary/secondary


149

system and controlled to track secondary flow down to the chiller minimum flow rate but with an increase in first costs and control complexity. (This is, however, a cost-effective way to improve the performance of existing primary/secondary systems and less expensive than converting the system to primary only.) The lower energy costs and lower first costs of the primary-only system often make it an easy choice versus primary/secondary, but the system does have two significant disadvantages, as discussed in the following subsections.

Bypass Control Complexities A bypass valve (shown in Figure 5-4)is required to ensure that minimum flow rates are maintained through operating chillers. The valve must be automatically controlled by flow, typically using a flowmeter in the primary circuit (as shown in Figure 5-4) or DP sensors across chillers correlated to flow. The flowmeter is more costly but is more easily adapted into plant load calculations, which are necessary for optimum chiller and CHW pump staging (discussed in Chapter 7). Selecting the bypass control valve and tuning the control loop is sometimes difficult because of the widely ranging DPs across the valve caused by its location near the pumps. The valve must be large enough to bypass the minimum chiller flow through it with a pressure drop as low as the DP set point used to control CHW pump VFDs. This is necessary because if only a few valves are open in the system, the pressure at the DP sensor location will be nearly equal to the DP at the plant because there is little pressure drop between these two points due to the low flow rate. This constraint makes the valve oversized for other flow scenarios that can occur, so tuning can be difficult. If the control loop is unstable, cold CHW supply can be fed back into the return intermittently and cause chillers to cycle off due to low load or cold supply water temperatures. However, if the loop is too slow, it may not respond quickly enough to sudden changes in flow (e.g., when a large number of AHUs shut off at the same time), resulting in insufficient flow through the chillers, thereby causing them to trip on low flow or low temperature. Complex control systems are prone to failure, so at some point in the life of the plant one can expect the bypass control to fail. A failure of the bypass system can cause nuisance chiller trips, which generally require a manual reset. If an operator is not present to reset the chiller, the plant can be out of service for some time.

Staging Control Complexities When one or more chillers are operating and another chiller is started by abruptly opening its isolation valve (or starting its pump for dedicated pump configurations), flow through the operating chillers will abruptly drop. The reason for this is simple: flow is determined by the demand of the CHW coils as controlled by their control valves. Starting another chiller will not create an increase in required flow, so flow will be split among the active machines. If this occurs suddenly, the drop in flow will cause operating chillers to trip.

150

Chapter 5 Optimizing Design To stage the chillers without a trip, active chillers must first be temporarily unloaded (demand limited or set point raised), then flow must be slowly increased through the new chiller by slowly opening its isolation valve. Then all chillers can be allowed to ramp up to the required load together. During the staging sequence, CHW temperatures will rise somewhat. This is seldom a problem in comfort applications but may be an issue for some industrial applications. Given these considerations, primary-only systems are most appropriate for the following situations: •

•

Plants with many chillers (more than three) and with fairly high base loads, as might be expected in an industrial or data center application. For these plants, the need for bypass is minimal or nil due to the high base loads, and flow fluctuations during staging are small due to the large number of chillers. Plants where design engineers and future on-site operators understand the complexity of the controls and the need to maintain them.

The primary/secondary system may be a better choice for buildings where fail-safe operation is essential or on-site operating staff is unsophisticated or nonexistent.

Primary/Distributed Secondary For plants serving groups of large loads such as buildings on a college campus or terminals in an airport, the primary/distributed secondary system (Figure 5-6) is usually the best option. Starting from a typical primary/secondary system as discussed above, the secondary pumps at the central plant are eliminated and variable-speed pumps are added at each building. The building pumps are controlled by DP sensors at the most remote coil in each building. Building pump heads are sized for the pressure drop of the loop from the plant, to the building, through the building’s coils, then back to the plant through the common leg. Thus, each pump has a different head customized for the building. The advantages of this design compared to conventional primary/secondary and primary/secondary/tertiary systems include the following: •

•

•

Overall pump horsepower is reduced. With a conventional primary/secondary system, secondary pump head must be sized for the most remote building (say 100 ft), whereas with a distributed secondary system, the building pumps close to the central plant can have much smaller heads (say 50 ft). The system is self-balancing via the speed controls on the secondary pumps. There is no need to throttle pressure at close buildings and flow self-adjusts over time as additional buildings are connected to the system. Overpressurization of control valves located near the central plant is eliminated. With large, high-head secondary systems, these valves must operate against excess DP, which can reduce controllability and may even force flow through the valve if it does not have sufficient shutoff head.


Figure 5-6

151

Primary/distributed secondary.

•

•

•

Pump energy is reduced because of the custom pump heads and the more precise control of the VFDs. With a conventional primary/secondary system, the secondary pumps are typically controlled to maintain DP at the entry to the most remote building. Thus, the DP set point must be higher than that for the distributed pump system, which is controlled by the DP at the most remote coil in each building. At part load, the pumps can therefore operate at slower speeds. With primary/secondary/tertiary systems, the tertiary pumps are generally piped with a bridge and a two-way control valve. Control of the bridge is always difficult and, if done incorrectly, is often the cause of degrading T (Taylor 2002). With this distributed pumping system, bridge connections are eliminated. The system will be less expensive, more energy efficient, and have lower maintenance costs than a primary/secondary/tertiary system. Disadvantages include the following:

•

•

Expansion tank pressurization may have to be increased to maintain positive suction pressure at building pumps if the pumps are located at the top of campus buildings. This has only a minor cost impact on the expansion tank. Primary/distributed secondary systems usually cost more than conventional primary/secondary systems because there are more pumps and space is required to house the pumps in each building.

152


Primary/Coil Secondary For plants serving large individual air-handling systems, using distributed variable-speed-driven coil secondary pumps (Figure 5-7) is usually the best option. The advantages of this design compared to a conventional primary/secondary system include the following: •

•

•

Figure 5-7

Connected pump motor horsepower is reduced. This is due in part because of the customized heads for each pump but also because the control valve is eliminated. Two-way control valves are typically selected for a wide-open pressure drop of 27 to 34 kPa, about 10 ft. This is a substantial savings. The system is self balancing. There is no need for balancing valves of any kind nor are there any advantages to self-balancing designs such as reversereturn arrangements. Pump energy is significantly lower with this design. This is due mostly to the reduced pump heads but also because there is no need to maintain a minimum DP in the system as there is with conventional secondary pumps. Because of this minimum DP and because of the throttling caused by partially closed con-

Primary/coil secondary.


153

trol valves, conventional secondary pumps will not follow the theoretical parabolic system curve. Hence, pump efficiency will generally get worse, particularly at low load. With the variable-speed coil pump design, there are no control valves or minimum DP, so pump efficiency will be nearly constant. •

Control of large control valves is inherently slow due to their size and slow responsiveness. With the coil pump design, flow can be controlled almost instantaneously with the VFD, so control is precise. There is also no fear of overpressurizing control valves with the coil pump design. Such overpressurization can lead to reduced controllability and may negatively impact plant T.

•

Because of the eliminated control valves and lower pump horsepower, this system generally has lower costs than a conventional primary/secondary system. It is usually a little more expensive than a primary-only system.

Control valves can be thought of as brakes on a car while pumps are the car engines; from an energy perspective, it never makes sense to press both the brake and the accelerator pedals at the same time, but that is effectively what systems with control valves do. The primary/coil secondary system is therefore ideal from a pumping perspective: it has no brakes. Unfortunately, there are a few disadvantages with this system: •

All coils must have a pump. If a coil were connected to the secondary circuit without a pump, flow through the coil will be backwards from the return to the supply. For a building that has a mixture of small coils and large coils, pumps for the small coils will most likely have to be expensive multistage pumps.

•

Exposure to equipment failure is increased. A control valve is extremely reliable—the pump and VFD in this design are more likely to fail. Duplex pumps could be used to improve redundancy, but the cost is prohibitive in most situations. A good design philosophy is to provide the same level of redundancy as the rest of the system served. For instance, if the air handler has only a single fan, then it makes sense to provide only a single pump. For more critical applications, redundant pumps or an alternative distribution system design should be considered.

•

For this design to be energy efficient, coils must be large due to the inherent inefficiency of small pumps, particularly low-flow/high-head pumps. For instance, a typical pump at 60 ft of head will have an efficiency on the order of 20% at 10 gpm, 40% at 20 gpm, 50% at 50 gpm, 60% at 100 gpm, 70% at 200 gpm, and 75% to 85% at higher flow rates. That is why this system is recommended only for coils with flows greater than 100 gpm in Table 5-1. This flow limit is obviously a rough rule of thumb because efficiency will vary over a range, not drop abruptly below 100 gpm.

If a project includes both small and large coils, a hybrid system of both distributed coil pumps and conventional secondary pumps to serve small coils is possible. See Figure 5-8 for an example hybrid plant.

154


Figure 5-8

Hybrid primary/coil secondary and primary/secondary system.

Optimizing Piping Design Pipe Sizing Selecting the optimum pipe size for a given design flow rate is a function of • • •

• • • •

the location of the pipe in the system (whether or not it is in the critical circuit, i.e., the circuit that determines pump head); the first costs of installed piping; the pump energy costs, which in turn depend on pump and motor efficiency, distribution system type (constant or variable flow), annual flow profile through the system as well as the pipe in question, type of pump control (variable speed or riding pump curve), etc.; erosion considerations (high velocities can contribute to hastening of pipe wall deterioration); noise considerations, such as velocity limits to minimize noise caused by turbulence and the proximity of the pipe to noise-sensitive areas; physical constraints; and budget constraints.


155

Traditionally, most designers size piping using rules of thumb, such as limiting friction rate (e.g., 4 ft per 100 ft of pipe), water velocity (e.g., 10 ft per second), or a combination of the two. These methods are expedient and reproducible, but they seldom result in an optimum design from either a first-cost or a life-cyclecost perspective. A better approach is to size piping based on LCCA. This method is discussed in a 2008 ASHRAE Journal article (Taylor and McGuire 2008) and can easily be performed using the pipe size optimization spreadsheet provided free with this textbook (available at ashrae.org/CHWSDL). The spreadsheet is very easy to use, but it is primarily intended to analyze piping systems that are completely laid out with all valves, fittings, and other appurtenances fully identified. It is therefore not as handy to use during the early design phase when these details are not yet known. An easier tool to use in early design is a simple lookup table showing maximum flow rates for each pipe size, such as Table 5-3. This table, which was extracted from ANSI/ASHRAE/IES Standard 90.1 (2016b), was developed from the LCCA spreadsheet (Pipe Size Optimization Tool v2.0.6 (available at ashrae.org/CHWSDL) assuming a typical distribution system and ANSI/ASHRAE/IES Standard 90.1 life-cycle cost parameters (see Taylor and McGuire [2008] for details). The flow rates listed are the maximum allowed by ANSI/ASHRAE/IES Standard 90.1 for each pipe size using the prescriptive compliance approach. Tables 5-4 and 5-5 are similar tables that the author uses for preTable 5-3

Piping System Design Maximum Flow Rate in gpm

(Table 6.5.4.6 from ANSI/ASHRAE/IES Standard 90.1-2016)

Operating h/yr

2000 h/yr

>2000 and 4400 h/yr

>4400 and 8760 h/yr

Other

Variable Flow/ Variable Speed

Other


Other


2 1/2

120

180

85

130

68

110

3

180

270

140

210

110

170

4

350

530

260

400

210

320

5

410

620

310

470

250

370

6

740

1100

570

860

440

680

8

1200

1800

900

1400

700

1100

10

1800

2700

1300

2000

1000

1600

12

2500

3800

1900

2900

1500

2300

Maximum velocity for pipes over 12 in. size

8.5 fps

13.0 fps

6.5 fps

9.5 fps

5.0 fps

7.5 fps

Nominal Pipe Size, in.

156

Chapter 5 Optimizing Design liminary design of variable-flow, variable-speed and constant-flow, and constantspeed pumping systems, respectively. These tables were also developed from the LCCA spreadsheet assuming California utility rates and fairly aggressive lifecycle cost parameters for discount rates, energy rate escalation, etc., which the author believes are appropriate for green buildings ($0.15/kWh, 1% energy escalation rate, 5% discount rate, 30 year life). The maximum flow rates are a bit lower than those in Table 5-3 accordingly. Tables 5-4 and 5-5 also include a set of flow limits for piping located in acoustically sensitive areas—again, see Taylor and McGuire (2008) for details. Table 5-4 and Table 5-5 are useful for selecting pipe sizes in the early design phase; once the design is more complete, the LCCA spreadsheet can be used to select final pipe sizes.

Table 5-4 Piping System Design Maximum Flow Rate in gpm for Variable-Flow, Variable-Speed Pumping Systems (Developed from LCCA spreadsheet assuming green life-cycle cost parameters) Pipe Diameter

Non-noise Sensitive

Noise Sensitive

2000

4400

8760

2000

4400

8760

1/2

7.8

5.9

4.6

1.8

1.8

1.8

3/4

18

14

11

4.6

4.6

4.6

1

29

22

17

8.9

8.9

8.9

1 1/4

51

39

30

15

15

15

1 1/2

88

67

52

24

24

24

2

120

84

67

51

51

51

2 1/2

160

120

91

81

81

81

3

270

210

160

140

140

140

4

480

360

290

280

280

280

5

670

510

390

490

490

390

6

1100

800

630

770

770

630

8

1800

1400

1100

1500

1400

1100

10

2900

2200

1800

2700

2200

1800

12

4400

3300

2600

4200

3300

2600

14

6000

4600

3600

5400

4600

3600

16

7400

5700

4500

7200

5700

4500

18

10,000

8000

6300

9200

8000

6300

20

11,000

8800

7000

11,000

8800

7000

24

17,000

13,000

11,000

17,000

13,000

11,000

26

21,000

16,000

13,000

20,000

16,000

13,000


157

Table 5-5 Piping System Design Maximum Flow Rate in gpm for Constant-Flow, Constant-Speed Pumping Systems (Developed from LCCA spreadsheet assuming green life-cycle cost parameters) Pipe Diameter

Non-Noise Sensitive

Noise Sensitive

2000

4400

8760

2000

4400

8760

1/2

5.0

3.9

3.0

1.8

1.8

1.8

3/4

12

9.0

7.0

4.6

4.6

4.6

1

19

14

11

8.9

8.9

8.9

1 1/4

34

26

20

15

15

15

1 1/2

57

43

34

24

24

24

2

73

55

44

51

51

44

2 1/2

100

77

60

81

77

60

3

180

140

110

140

140

110

4

320

240

190

280

240

190

5

430

330

260

430

330

260

6

700

530

420

700

530

420

8

1200

900

720

1200

900

720

10

1900

1500

1200

1900

1500

1200

12

2900

2200

1700

2900

2200

1700

14

4000

3000

2400

4000

3000

2400

16

4900

3800

3000

4900

3800

3000

18

7000

5300

4200

7000

5300

4200

20

7700

5800

4600

7700

5800

4600

24

12,000

8900

7100

12,000

8900

7100

26

14,000

11,000

8500

14,000

11,000

8500

Selecting Valves and Fittings To minimize the energy use of piping systems, piping system pressure drop must be minimized. The procedures described in the Pipe Sizing section should result in near-optimum pipe size for straight pipe and fittings. However, accessories such as valves and strainers are also major contributors to system pressure and must be optimized by proper selection. The following recommendations should be considered.

Valves at Chillers and Towers CHW plants in general should have only two types of valves for flow isolation and balance: butterfly valves for large piping (typically 3 in. and larger) and ball valves for smaller piping. These are not only the least expensive types

158

Chapter 5 Optimizing Design of valves, but they also are the easiest to use because they require only a quarter turn to open and close (unlike globe and gate valves) and they may be used for balancing (unlike gate valves). They are also physically smaller than other valve types. Standard two-piece ball valves come in two types: full port (the opening in the ball is the same as the pipe size) and standard-port (the hole is smaller than the pipe size). Full-port ball valves have a lower pressure drop but cost more than standard-port ball valves. It is cost-effective to use the full-port valve for valves located in the critical circuit (the circuit that has the highest pressure drop and that determines the pump head). Otherwise, standard-port ball valves are the most cost-effective. Valves to avoid and/or that are not needed include the following: • •

•

•

Globe and gate valves: They provide no advantages over butterfly and ball valves, and they cost more and require more space. CBVs: Most plants are almost self-balancing simply because their compact size does not result in large differences in pressure drop across each evaporator or condenser. Even if no balancing is done, the plant generally will work well—a small difference in CHW or condenser water flow among chillers and towers has only a small impact on plant performance. What little balancing is needed can be accomplished by modulating the condenser/evaporator isolation butterfly valves to cause the pressure drop across the condenser/evaporator to be the same among the chillers (or proportionally the same if chillers are not all the same). Cooling tower balancing is almost never an issue, and there is no need to create equal flow to each cell; the cells with excess flow will create warmer water and the ones with low flow will create colder water, but when they are mixed, the resulting temperature is almost exactly the same as it would be if the cells had equal flow. Flow-limiting valves (also called flow control valves): These valves should never be used in a CHW plant. There are times when exceeding design flow is actually desired. For example, a chiller plant with degrading CHW T can be more efficient if higher flow is allowed to be forced through operating chillers rather than start another. Exceeding the chiller manufacturer’s maximum flow rate is rarely an issue because pressure drop increases with the square of the flow and pumps seldom have the capability to force this much water through the evaporator or condenser. Multipurpose valves (also called triple duty): These valves, typically located at pump discharges, provide three functions in one: shutoff valve, check valve, and flow measurement/balance. However, they are hard to use as shutoff valves because they are multi-turn rather than quarter turn and a wrench is required; they generally do not include handles. They can also be hard to repair because they cannot isolate themselves. When sized the same as the pipe size, as is typical, flow measurement is often not possible; they must be undersized to provide accurate flow readings, which adds to pressure drop. Flow balancing is also not needed at the


159

pump; it can and should be done at the HXs, and throttling flow for balance is not required at all if the pump has a VFD. Finally, cost savings versus a check and butterfly valve are minor or nonexistent. Therefore, triple-duty valves are unnecessary in almost all applications.

Balancing Devices at Coils As discussed in Chapter 4, two-way valve variable-flow systems do not require balancing. AFLVs should never be used with variable-flow CHW systems. Manual CBVs are also not needed, but they can be handy for diagnosing problems because they can be used to measure how much flow is going through a coil. Where used, CBVs that use ball valves are preferred because they have a lower pressure drop than those that use globe valves.

Strainers at Coils Strainers are often installed at coils to protect control valves from fouling. But strainers located at pumps usually eliminate sufficient debris from the system so that problems seldom occur at coils even without strainers. Fine debris can pass through coils and control valves without causing damage. In fact, strainers at coils may cause flow problems because they often are not readily accessible for maintenance and therefore get clogged (particularly during startup) and restrict flow.

Optimizing Chilled-Water Design Temperatures Table 5-6 shows the typical range of CHW temperature difference (commonly referred to as delta-T or T) and the general impact on energy usage and first costs. The table shows that there are significant benefits to increasing T from a first-cost standpoint, and there may be energy cost savings as well, depending on the relative size of the fan energy increase (due to increased air-side pressure drop with deeper CHW coils) versus pump energy decrease as T increases. Chiller energy usage is largely unaffected by T Table 5-6

Impact on First Costs and Energy Costs of Chilled-Water Temperature Difference (Assuming Constant Chilled-Water Supply Temperature) T Low

High

Typical range

8°F

25°F

First-cost impact

Smaller coil

Smaller pipe Smaller pump Smaller pump motor

Energy-cost impact

Lower fan energy

Lower pump energy

160

Chapter 5 Optimizing Design for a given CHW supply temperature. The leaving CHW temperature drives the evaporator temperature, which in turn drives chiller efficiency; entering water temperature has almost no impact on efficiency. Intuitively, one might think that fan energy would dominate in the energy balance between fan and CHW pump because fan energy is so much larger than pump energy annually, and the fan sees the coil pressure drop under all conditions, while the CHW pump typically only runs in warmer weather (assuming the system has an air-side economizer). However, detailed analysis has shown that not to be the case: the impact on the air side of the system is seldom significant. Table 5-7 shows a typical cooling coil’s performance over a range of CHW Ts. While the example in the table will not be true of all applications, it does suggest that air-side pressure will not increase very much as CHW T rises, while water-side pressure drop falls significantly. For VAV systems, the impact on annual fan energy is even less significant because any fullload air-side pressure drop penalty will fall rapidly as airflow decreases. Figure 5-9 shows the impact of CHW T on energy usage for a typical Oakland, CA, office building served by a VAV air-distribution system with VFD and an air-side economizer. Fan energy rises only slightly as T increases. If pipe size is left unchanged as T increases, CHW pump energy will fall substantially due to reduced flow and reduced piping losses. In real applications, pipe sizes are generally reduced to decrease first costs, but pump energy will still fall due to reduced flow rates and reduced coil and evaporator pressure drops, although not as dramatically as in Figure 5-9. Reducing CHW temperature can eliminate the fan energy penalty. Figure 5-10 shows the same system as Figure 5-9, but instead of holding CHW temperature constant and increasing coil size to increase T, CHW temperature is lowered and coil size is held constant to keep air-side pressure drop (and therefore fan energy) constant as T increases. Dropping CHW temperature increases chiller energy, but pump energy savings more than make up the difference. As with Figure 5-9, the pump energy shown in Figure 5-10 assumes that pipe sizes remain constant, which is not always the case.

Table 5-7

Typical Coil Performance Versus Chilled-Water Temperature Difference

Chilled-Water T, °F

5.5

7.2

8.9

10.6

12.2

14

Coil water pressure drop, ft H2O

7.2

4.2

2.8

2.5

2.0

1.4

Coil air-side pressure drop, in. H2O

12.2

12.7

13.2

15.2

16.0

19.8

Rows

6

6

6

8

8

8

Fins per in. (fpi)

2.9

3.3

3.7

3.0

3.4

4.6

Cooling coil pressure air- and water-side drops were determined from a manufacturer’s AHRI-certified selection program assuming 500 fpm coil face velocity, smooth tubes, maximum 12 fpi fin spacing, 43°F CHW supply temperature, 78°F/63°F entering air temperature, and 53°F leaving air temperature.


Figure 5-9

161

Typical annual energy usage versus CHW T with a constant CHW supply temperature and constant pipe sizes.


Figure 5-10

Typical annual energy usage with coils selected for constant air-side pressure drop.


Table 5-8 compares three cooling coils with 4, 6, and 8 rows that result in about 10°F, 18°F, and 25°F T, respectively. Pipe sizes were selected from Table 5-4 assuming an acoustically sensitive location with about 2000 h/yr of operation. Coil first costs were obtained from the manufacturer’s representative

162

Chapter 5 Optimizing Design Table 5-8 Cooling Coil and Associated Piping Costs (For 20,000 cfm coil sized at 500 fpm, 42°F CHW supply temperature, 78°F entering dry-bulb temperature, 62°F entering wet-bulb temperature, and 53°F leaving dry-bulb temperature) Coil

Piping

Air Fluid Pressure Pressure Pipe Size, Fins per Fluid T, Fluid Coil Cost Rows Drop, in. in. in. Flow,gpm Drop, °F H2O ft H2O

Coil Connection

Total Cost

10

4

17.8

10.1

118.7

9.1

$3598

3

$4551

$8149

11

6

16.51

18.2

66.0

7.6

$4845

2.5

$3581

$8426

10

8

20.32

24.9

47.0

5.7

$5956

2

$2101

$8057

and piping costs (including typical valve train and 20 ft of branch piping) were obtained from the LCCA spreadsheet piping cost data. Table 5-8 shows that the added cost of the deeper coil is more than offset by the savings in the cost of piping the coil. And there are additional first-cost savings from the reduced piping main, pump, pump motor, and VFD sizes. So, increasing T reduces both first costs and energy costs. Clearly lifecycle costs will be lower the higher the T. We were unable in our analysis to find a point where the negative impact on fan energy or coil costs caused lifecycle costs to start to rise with increasing T; within the range of our analysis (up to 25°F T), higher T was always better. Energy savings from high T are even greater with systems that have WSEs or CHW TES. To reiterate: our analysis suggests that it never makes sense to use the traditional 10°F or 12°F Ts that are commonly used in standard practice. As T is increased, eventually the ever-deepening coil will run into the ASHRAE Standard 62.1 coil pressure drop limit (2016a). ASHRAE Standard 62.1 uses dry coil pressure as a surrogate for the cleanability of the coil. Section 5.11.12 of the standard requires that dry coil pressure drop at 500 fpm face velocity must not exceed 0.75 in. This is roughly the pressure drop of an 8 row, 12 fpi coil (depending on the details of the fin and tube design). So, the design procedure for selecting CHW coils is simple: rather than arbitrarily selecting CHW temperatures and then selecting coils that deliver those temperatures, reverse the logic: always use the “biggest” (highest effectiveness) coil available without exceeding the ASHRAE Standard 62.1 pressure drop limits and let the CHW T be determined by the coil and design air conditions. Based on this logic, the recommended procedure for sizing CHW coils and selecting CHW design temperatures is as follows. The intent of this procedure is to achieve all of the piping first cost savings resulting from a high T but with as warm a CHW temperature as possible to improve chiller efficiency.


163

1. Calculate the CHW flow rate for all coils assuming a 25°F T.1 2. Pick primary pipe sizes (at pumps, headers, main risers, and main branch lines) in the critical circuit (that which determines pump head) using Table 5-4 or the LCCA spreadsheet. 3. With pipe sizes selected, use Table 5-4 or work backwards using the LCCA spreadsheet to find the maximum flow for each pipe size, and then recalculate the T in each pipe using these flow rates. This is the minimum average T for this leg of the circuit. 4. Use the coil manufacturer’s selection program to find the maximum coil size that complies with the ASHRAE Standard 62.1 cleanability limit, typically 8 row/12 fpi.2 Use this for all coils so that T is maximized. (For some smaller fan-coils, 8 row coils may not be an option. If so, use the largest coil available but no less than 6 rows. If that is not an option with the selected manufacturer, find another manufacturer.) 5. With the coil manufacturer’s selection program, iterate on coil selections to determine the CHW supply temperature that results in the selected T on average for each leg of the critical circuit, starting with the coil at the end of the circuit and working back to the plant. It is not necessary that all Ts be the same (and in fact they definitely will not be the same with this approach), just that the flow through the circuit equals the maximum flow determined in Step 3. The recommended minimum CHW supply temperature is 42°F.2 If this minimum is reached in any leg and the flow exceeds the maximum, then the process must be started over with a smaller T assumption in Step 1. 6. The lowest required CHW supply temperature for any coil in the circuit is then the design temperature. 7. Determine actual flow and T in other coils in other circuits using the coil selection program with this design CHW supply temperature, again maximizing coil size within ASHRAE Standard 62.1 limits (e.g., 8 rows, 12 fpi) and letting the program determine return water temperature. 8. The plant flow is that calculated using the diversified (concurrent) load and the gpm-weighted average return water temperature of all coils. 1. Some engineers may be concerned that a 25°F T is nonconservative and reduces future flexibility for load changes. Designing around large Ts results in large coils and small pipes and pumps. Designing around small Ts results in the opposite, small coils and large pipes and pumps. Both are equally forgiving with respect to possible coil load changes—one is no more conservative than the other. If excess capacity is desired for future flexibility, it should be explicitly built into the design rather than relying on accidental flexibility from design parameters. 2. In our analyses, the lowest CHW supply temperature resulting from this technique was about 42°F; we do not know if lower CHW temperatures will start to affect the lifecycle cost due to reduced chiller efficiency. So, unless the designer performs additional LCCA, we suggest limiting the design CHW supply temperature to no colder than 42°F. For most applications, this low temperature will not be required to achieve the target 25°F T. Limiting the supply temperature to 42°F also provides some conservatism in the design; should there be a miscalculation in loads or unexpectedly high loads at a certain coil, CHW temperature can be lowered below 42°F to increase coil capacity, although with a resultant loss in overall chiller capacity and efficiency.

164

Chapter 5 Optimizing Design If this procedure seems too long and complicated, here is a shortcut procedure: skip Steps 1 to 5 and just assume a CHW supply temperature of 42°F in Step 6. This will provide basically the same result except that the design CHW temperature may be lower than it needed to be to achieve the pipe size savings from high T, so the chiller design efficiency may be worse than it needed to be. The energy impact will be minimal, however, because CHW temperature should be aggressively reset, as discussed in Chapter 7. This simplified approach also results in a somewhat lower CHW flow rate, so pump size and power will be reduced.

Optimizing Condenser Water Design Temperatures Selecting optimum condenser water temperatures is more complex than selecting CHW temperatures due to the complex interactions between cooling towers and chillers. As with chilled water, there can be significant first-cost savings using high condenser water Ts (also known as cooling tower range). But with chilled water, the supply fan energy impact was small, so increasing T was found to always reduce total system energy costs. With condenser water, the energy impact on the chiller of increasing T and return condenser water temperature is not small (in fact it is very large), and T also significantly affects the energy used by the cooling tower. So, optimum condenser water temperatures are not as easily determined as those for chilled water. Table 5-9 shows the first-cost and energy impacts of condenser water temperature difference within the ranges commonly used in practice. Higher Ts reduce first costs (because pipes, pumps, and cooling towers are smaller), but the net energy-cost impact may be higher or lower depending on the specific design of the chillers and towers. Figure 5-11 shows chiller, tower, and condenser water pump energy usage for the example Oakland office building introduced in Figures 5-9 and 5-10. The condenser water temperature and T were selected so that the cooling tower

Table 5-9 Impact on First Costs and Energy Costs of Condenser Water Temperature Difference Assuming Constant Condenser Water Supply Temperature T Low

High

8F

18F

First-cost impact

Smaller condenser

Smaller pipe Smaller pump Smaller pump motor Smaller cooling tower Smaller cooling tower motor

Energy-cost impact

Lower chiller energy

Lower pump energy Lower cooling tower energy

Typical range


Figure 5-11

165

Oakland office building annual energy usage versus with condenser water supply temperature and T selected for constant tower size and constant pipe sizes.


size and fan power do not change. As the T decreases, the temperature of the water returning to the cooling tower decreases and the tower becomes less efficient. This requires the condenser water temperature leaving the tower to rise (or the tower size must be increased). The most energy-efficient combination in this case was a 14°F T. But this assumes pipe sizing is constant; however, the pipe sizes could have been reduced for the larger T designs, reducing first costs but increasing pump energy costs. Figure 5-12 shows life-cycle costs for a large office building CHW plant that was analyzed as part of the development of this SDL. Utility costs and life-cycle cost assumptions are those used in the evaluation of energy conservation measures for ANSI/ASHRAE/IES Standard 90.1 ($0.094/kWh average electricity costs and 14 scalar ratio—the scalar ratio is essentially the simple payback period). The plant was modeled in great detail (including real cooling tower and piping costs) for three climates: Oakland, CA; Albuquerque, NM; and Chicago, IL. Figure 5-12 shows results for Chicago but the trend was the same in all three climate zones: life-cycle costs were minimized at the largest of the three Ts analyzed, about 15°F. This was true for both office buildings and data centers and for both single-stage centrifugal chillers and two-stage centrifugal chillers. It was also true for low-, medium-, and high-approach cooling towers (high-approach tower data are shown Figure 5-12). (The optimum approach temperature is discussed under the section Selecting Cooling Towers, but it had no impact on the optimum T). In all cases, pipe, pump, pump motor, and pump VFD sizes reduced as T increased, and these cost differences were the primary driver in life-cycle cost differences, as shown in Figure 5-12. The differences in energy

166


Figure 5-12

Life-cycle costs of 1000 ton chilled plant in Chicago as a function of condenser water T.


use among the options is not as significant because savings in pump and tower energy largely (though not completely) offset the increase in chiller energy use. Other studies have also found that 15°F condenser water T is optimum and can even reduce annual energy costs (Trane 2005, 2011). The plant analyzed had a relatively short distance between the towers and chillers; high Ts would have an even larger life-cycle cost advantage for plants that have a large distance between them, such as a plant with chillers in the basement and towers on the roof. Based on this analysis, the following procedure is suggested to choose the condenser water T (cooling tower range): • •

•

•

Calculate the condenser water flow rate for all pipe sections assuming a range of 15°F. Pick primary pipe sizes (at pumps, headers, main risers, main branch lines) in the critical circuit (that which determines pump head) using the LCCA spreadsheet. With pipe sizes selected, work backwards using the LCCA spreadsheet to find the maximum flow for each pipe size and then recalculate the T in each pipe using these flow rates. The largest calculated T in any pipe segment is the design plant T. Recalculate all flow rates using this T.

This procedure attempts to minimize cost by reducing pipe size as much as possible. However, it then takes full advantage of the resulting pipe size to minimize T to reduce chiller energy. Pump energy will be a bit higher than if a 15°F T were simply used, but pump energy is small relative to the impact of high T on chiller energy use.

Fundamentals of Design and Control of Central Chilled-Water Plants I-P Table 5-10 1

2

Piping Section

Application

167

Example Condenser Water Pipe Sizing 3 gpm at 15°F T

4

5

Pipe Size Maximum T at in. per gpm per Maximum Table 5-5 Table 5-5 gpm

6 gpm at Maximum T

Common Constant flow/constant speed, pipe 2000 h, non-noise sensitive

1850

10

1900

14.6

1900

To each Constant flow/constant speed, equipment 2000 h, non-noise sensitive

925

8

1200

11.6

950

Example Plant Take, for example, a 1000 ton plant serving an office building in Oakland. Each pump, chiller, and tower is sized for half the load. •

At the initial chiller selection COP of 6.28, the rejected heat is about 13.9 million Btu/h. So at a15°F T, the total condenser water flow rate is about 1850 gpm and the flow rate to each individual piece of equipment is 925 gpm (Table 5-10, Column 2).

•

For an Oakland office served by a system with an air-side economizer, the chiller plant will operate for about 2000 h/yr. Assuming constant flow/constant speed, we can use Table 5-5 for pipe sizing. Because the chiller room is not noise sensitive, the pipe sizes on the left side of the table are used. The selected pipe sizes for each of the two piping sections are shown in Table 5-10, Column 3.

•

Next, the selected pipe sizes are maxed out using the maximum flow rates for each from Table 5-5 (Table 5-10, Column 4) and the T for each piping section is recalculated using this flow rate.

•

The highest T in Column 5 is selected (14.6°F) and flow rates are recalculated using this T (Table 5-10, Column 6). These flow rates would be used to select chillers, towers, and pumps.

Selecting Cooling Towers Cooling tower selection and the strategies used to control the tower fans have a significant impact on CHW plant performance. Sizing towers for a close approach to ambient wet-bulb temperature will improve chiller efficiency, but it increases tower fan energy and first costs. Other factors to consider are the efficiency of the tower itself and the tower fan speed control options.

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Tower Fan Speed Control Tower fan control is required to maintain tower leaving water temperature under all load and weather conditions. The common options (listed roughly in ascending order of first-cost premium) are as follows: • • • • •

One-speed motor (cycling) Two-speed motor where low speed is half of full speed (1800 rpm/900 rpm) Two -speed motor where low speed is two-thirds of full speed (1800 rpm/ 1200 rpm) Dual-motor drives (the smaller pony motor is typically designed for twothirds of full speed) VFDs

Figure 5-13 shows the energy performance of these options. The performance of the pony motor option will typically be close to the 100%/67% (1800/1200 rpm) two-speed motor option. The performance of the one-speed fan is not linear because of the free cooling that occurs in the off cycle due to the stack-effect-induced natural draft through the tower. The optimum fan control option depends primarily on the number of hours the tower operates under various load conditions. This, in turn, is a function of the application (e.g., office building, data center), climate, and the tower control set point strategy (see Chapter 7).

Figure 5-13

Cooling tower fan model at part load: percent power as a function of percent load.


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Despite these complex interactions, the following generalizations can be drawn from our studies of typical plants: •

One-speed control is almost never the optimum strategy for the typical chiller plant, regardless of building occupancy type, climate, or tower control set point reset strategy. Energy savings from at least one of the twospeed options result in very short payback periods. In addition, the two- or variable-speed options reduce noise and wear and tear on belts.

•

The 100%/50% (1800 rpm/900 rpm) two-speed motor is somewhat more energy efficient than the 100%/67% option, particularly when controlled as recommended in Chapter 7. The 1800/900 rpm motor is available in a onewinding motor, which is less expensive than the two-winding motor required for 1800/1200 rpm operation, although this savings is partly offset by the higher cost of the starter. At low speed, towers driven by 1800/900 motors are also quieter than those driven by 1800/1200 motors.

•

Pony motors are usually more expensive than standard two-speed motors, but the additional motor reduces exposure to motor failure. Note that this option is not available on all towers.

•

For plants with multiple towers or multiple cells, two- or variable-speed control must be provided on all cells, not just the lead cells. The towers are most efficient when all cells are running at low speed rather than some at full speed and some off. For instance, two cells operating at half speed will use about 25% of full power compared to 50% of full power when one cell is on and the other is off.

•

Because VFD efficiency is not that much better than two-speed motor efficiency (see Figure 5-13), VFDs were not found to be cost-effective when we first performed a LCCA in the late 1990s. Since that time, however, VFD costs have plummeted, and they are now the clear best option. In addition to energy savings, VFDs offer these other benefits: °

Similar or lower first cost to two-speed motors

°

Tighter temperature control than two-speed motors

°

Reduced noise at low loads and less abrupt changes in noise generation

°

Reduced belt maintenance costs for belt-driven fans

°

Operating information available from the VFD, such as motor power

Tower Efficiency Tower efficiency (as defined in California Title 24 [CBSC 2016] and ANSI/ASHRAE/IES Standard 90.1 [2016b]) is the ratio of the maximum tower flow rate (gpm) to the motor horsepower () at standard Cooling Tower Institute (CTI) rating conditions (95F to 85F at 75F wet bulb). The optimum tower efficiency depends on the number of hours the tower operates under various load and weather conditions.

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Chapter 5 Optimizing Design Two primary factors affect tower efficiency: •

•

Fan type: Towers are commonly available with either propeller fans or centrifugal blowers. The latter require about twice the fan power and are no less expensive. From an energy and first-cost perspective, propeller fan towers (whether draw through or blow through) should always be used. There are a few advantages to centrifugal fan towers: they are generally quieter than propeller fan towers and they can operate against a larger external static pressure drop, such as that caused by louvers when towers are located indoors. However, the design can usually be modified to accommodate propeller fan towers, such as by oversizing the tower to reduce fan speed and noise, using ultraquiet fan blades available from most tower manufacturers, or oversizing intake louvers to reduce pressure drop. Because of the severe energy penalty associated with centrifugal blowers, every effort should be made to accommodate a propeller fan tower before considering a centrifugal fan tower. Fan pressure drop through the fill: Pressure drop is primarily a function of the fill’s size and design but can also be affected by fan inlet and discharge configuration. Most engineers use the least expensive tower selections available in manufacturers’ selection software. However, the tower and fill can be oversized to reduce pressure drop, thereby allowing the fan to be slowed down, which reduces motor power. Whether this is cost-effective depends on the application, climate, and the added cost to oversize the tower and to accommodate the larger tower footprint and weight.

To determine optimum efficiency and approach temperature, a large office building CHW plant was analyzed using utility costs and life-cycle cost assumptions as described above. The plant was modeled in great detail for three climates: Oakland, CA; Albuquerque, NM; and Chicago, IL. Details included real equipment and piping costs. Additional analyses for optimum approach temperature were made for Miami, FL; Las Vegas, NV; and Atlanta, GA. The condenser water system was designed, cost estimated, and modeled at all permutations of the following design parameters: •

•

Three ranges of tower efficiencies: low was the least efficient available for the cross-flow propeller fan tower series analyzed, with efficiencies ranging from 45 to 60 gpm/hp; medium, with efficiencies ranging from 65 to 75 gpm/hp; and high, with efficiencies ranging from 80 to 100 gpm/hp. Note that even the low-efficiency towers are significantly more efficient than the ANSI/ ASHRAE/IES Standard 90.1 minimum of 38.2 gpm/hp (2016b). Tower approach temperatures ranging from 2.5°F to 11°F based on actual tower selections for a cross-flow propeller fan tower series.

Tower costs were based on manufacturer’s price plus sales tax and contractor markup, plus a 50% premium that is intended to estimate the secondary installed cost impact of larger towers. Tower size (both footprint and height)


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and weight increase with increasing efficiency and decreasing approach. Both can impact tower installed costs depending on tower location. The actual premium can vary from close to nothing for a tower located on grade to a significant premium if the tower is located on the roof and requires architectural screening to hide the tower from view, because larger towers will require additional structural work and larger screens. The 50% premium is probably conservative—in most cases the premium will be lower. Figure 5-14 shows life-cycle costs for the three ranges of cooling tower efficiencies for an office building in Miami with a tower range of 15°F; lifecycle costs were minimized with the high-efficiency towers. This was true in all climate zones, ranges, and approach temperatures analyzed. The analyses were based on a fairly aggressive scalar ratio (maximum simple payback period) of 14, but high-efficiency towers were found to be cost-effective down to a scalar ratio of about 5 (i.e., they will have a simple payback of five years compared to the next best tower option, even in the mildest climates). The reason is that the net cost premium for increasing tower efficiency is relatively small; the tower physical size and fill area increase but motor and VFD size and cost decrease, partially offsetting the tower cost increase. For example, the net installed first cost add for the high-efficiency tower versus the low-efficiency tower for the 1000 ton plant in Miami was only about $9000, a 6% increase, while annual energy savings were about $5500 and life-cycle energy savings were $77,000. The magnitude of the savings is smaller in milder climates, but the high-efficiency towers were found to be cost-effective in all climates analyzed.

Figure 5-14

Life-cycle costs of 1000 ton chiller plant serving a Miami office building as a function of tower efficiency range.


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Chapter 5 Optimizing Design Accordingly, cooling towers should be selected for an efficiency of at least 80 gpm/hp. Those serving data centers and other 24/7 applications should be selected for 100 gpm/hp. Plants with WSEs should have even higher efficiencies as discussed in the Water-Side Economizers (WSEs) section.

Tower Approach Temperature While increasing tower efficiency from low to high is relatively inexpensive (about 6% to 12% of tower costs in this study), the same cannot be said for reducing tower approach. As shown in Figure 5-15, the cost of a tower that provides a low approach of 2°F to 3°Fcan be 60% higher than a tower providing a 10°F to 12°F approach. A reasonable correlation was found between the sum of the life-cycle cost optimum tower approach (TA) (again using ANSI/ASHRAE/IES Standard 90.1 energy costs and scalar ratio) and tower range (also called condenser water temperature difference, TCW), and cooling degree-days base 50 (CDD50). This is shown in Figure 5-16 for each of the climate zones tested. The straight line curve fit is approximately T A + T CW = 15 – 0.0006CDD 50

Figure 5-15

(5-1)

Cooling tower installed costs for a 1000 ton chiller plant as a function of tower approach.



Figure 5-16

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Life-cycle cost optimum approach plus range as a function of cooling degree days base 50°F.


Solving for approach, T A = 15 – T CW – 0.0006CDD 50

(5-2)

Based on the limited data, the approach determined from this equation should be limited to no more than 9.5°Fand no less than 2.5°F. In summary, based on this analysis the following design criteria are recommended for selecting cooling towers for office buildings and buildings with similar load profiles: • •

Tower efficiency should be 80 gpm/hp or greater. Tower approach should be selected using Equation 5-2, with a minimum of 2.5°F and a maximum of 9.5°F.

Water-Side Economizers (WSEs) As discussed in Chapter 4, WSEs should always be piped in an integrated precooling position, as shown in Figure 5-17. As shown in Figure 5-17, the same cooling towers and condenser water pumps can be used to serve both the economizer HX and the chiller condensers. Some designers provide separate towers and pumps, at considerable expense, because of concerns about low chiller head pressure due to low condenser water supply temperatures. However, head pressure control is easily and

174


Figure 5-17

Integrated economizer.

inexpensively addressed by making the condenser isolation valves modulating and controlling them off the head pressure control signal output that is standard on most chiller controllers. The valves throttle flow through the condenser as needed to maintain chiller minimum lift regardless of how cold the condenser water supply temperature is in economizer mode. (See Chapter 4 for more details on head pressure control.) The capacity and quantity of cooling towers and condenser water pumps remain the same as they would without the economizer. For office building applications, this is intuitively clear: we know that when the economizer is on, weather will be cold so loads will be well below design loads; hence, only one of the two chillers (in Figure 5-17) will be needed, freeing the other condenser water pump to supply tower water to the economizer HX. This is also true for data centers where the load may require all chillers to run even in cold weather. The reason is that the load on the towers is actually reduced by the economizer because compressor heat is reduced, and condenser water flow to the chiller condensers may be less than design because the water tem-


175

perature is colder than design (and in fact may be reduced by the throttling of the head pressure control valves discussed in Chapter 4 in the section Refrigerant Head Pressure Control), thus making water available to the economizer HX without the need to add pumps. While cooling tower capacity is not affected by the economizer, it may be necessary to reduce the design approach temperature in order to meet ANSI/ ASHRAE/IES Standard 90.1 Section 6.5.1.2.1 (2016b) WSE requirements, particularly for plants with high loads in cold weather. This is discussed further in the example below in Example Design Procedure. It is critical that cooling towers be very efficient because they will be running at full speed many hours of the year. A minimum of 90 gpm/hp is recommended for office-type applications and 110 gpm/hp for 24/7 applications, such as data centers. These efficiencies are 10% above those shown to be costeffective for non-economizer applications. Cooling towers should be selected so that as many tower cells as possible can be enabled when the economizer is enabled to maximize efficiency and capacity while maintaining the minimum flow rates required by the tower manufacturer to prevent scaling. Low minimum flow rates can be achieved using weir dams and special nozzles in the hot-water distribution pans. CHW pump head increases due to the pressure drop of the HX when in economizer mode. However, in applications like offices, where the loads are low when the economizer is on, pump head may not need to increase above design head when the economizer is off; excess head may be available for the HX when the economizer is active due to the reduced CHW flow to coils. The HX should be a plate-and-frame type and selected for an approach of about 3°F (i.e., the temperature of the chilled water leaving the HX is equal to 3°F above the temperature of the condenser water entering the HX) in office applications. HX cost increases exponentially with approach temperature, so very close approaches should be tested for cost-effectiveness. Plate-and-frame HX cost typically increases dramatically with an increase in frame size, whereas adding additional plates to a smaller frame usually results in a linear increase in cost. As such, one effective design approach in applications such as data centers, where a close approach may prove cost-effective, is to maximize the plate capacity of a given frame. For instance, if the design condition is for a 3°F approach, but the selection leaves 20% additional frame capacity, it may prove cost-effective to max out the frame. The HX pressure drop on the condenser water side should be similar to that of the condensers so the flow rate will be similar when serving either the condensers or the HX. On the CHW side, pressure drop is typically limited to about 5 or 6 psi to minimize the CHW pump energy impact. The HX performance must be certified per AHRI 400 (2001), as required by ANSI/ASHRAE/IES Standard 90.1. To maximize economizer performance, and also the performance of the system even when not economizing, the CHW system must be designed for a very high temperature rise (T) using the procedure described above.

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Example Design Procedure 1. Calculate the CHW load at 50°F dry-bulb temperature and 45°F wet-bulb temperature. This is the performance test condition prescribed by ANSI/ ASHRAE/IES Standard 90.1 for buildings other than data centers (2013). This can be done using standard load calculation software by selecting a spring or fall month and overriding design outdoor air temperatures. All other load assumptions remain the same. Loads should be reduced from design loads due to reduced conduction and outdoor air conditioning loads. 2. Use coil selection software or other coil models to determine the warmest CHW supply temperature that can meet 100% of the CHW load determined above for all coils. To do this, first determine the warmest supply water temperature that can meet the load at the design flow for each coil, then select the coldest of these and use it to determine the CHW flow through all the other coils. The coil software will also determine the CHW return temperature from each coil. Typically the CHW supply temperature can be reset 5°F or more above the design CHW supply temperature. This can be true even for data centers, despite the consistently high load, by taking advantage of redundant air handlers to effectively increase available coil area. 3. Select the condenser water flow rate equal to a multiple of design condenser water pump flow rates as required to closely match the CHW flow rate determined above. 4. Determine the HX condenser water supply temperature equal to the required reset CHW supply temperature determined above less the 3°F approach temperature. 5. Calculate the condenser water return temperature to match the CHW load based on the condenser water flow and supply temperature determined above. 6. Use cooling tower selection software to verify that the cooling towers can provide the required condenser water supply temperature at 45°F wet-bulb temperature. If not, then cooling towers (and/or HXs) will need to be reselected for closer approach temperatures. For instance, assume the plant in Figure 5-17 serves an office building with floor-by-floor air handlers. The design conditions of the two chillers at the design cooling peak are shown in Table 5-11. The pump flow rates match the chiller design rates and cooling towers are selected at 68°F wet-bulb temperature for a 7°F approach. The cooling loads are then recalculated at 50°F/45°F outdoor- air temperature conditions; the load drops from 600 to 350 tons. Coil selection programs are then used to determine the CHW conditions required at the air handlers, and condenser water conditions are determined using the steps above. The resulting HX design conditions are shown in Table 5-12.


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Chiller Design Conditions (Each) CHW Side

Condenser Water Side

Capacity, tons

300

345

Flow, gpm

290

550

Entering water, °F

69.0

75

Leaving water, °F

44.0

90

P, psi

6.5

6.7

Table 5-12

WSE HX Design Conditions

CHW (Hot) Side

Condenser Water (Cold) Side

Total load, tons

350

350

Flow, gpm

560

550

Entering water, °F

65.0

47.0

Leaving water, °F

50.0

62.3

P, psi

5

4.8

Cooling tower selection software is then used to see if the selected cooling towers could cool 550 gpm (275 gpm across each tower) from 62.3°F to 47°F, a 2°F approach to the 45°F wet-bulb temperature. However, the software indicates that the towers are only able to deliver 48°F. So, the towers are reselected for a 73°F leaving water temperature (5°F approach to design wet-bulb temperature) in order to achieve a 2°F approach at 45°F ambient wet-bulb temperature and the load conditions prescribed in Table 5-12. Note that HX approach also could be reduced to deliver the desired CHW temperature, but it is usually more cost-effective to invest in larger cooling towers because they also improve efficiency when the economizer is off.

Selecting Type, Number, and Size of Chillers Design and selection procedures for chillers are discussed in Chapter 6. The procedures are included in a separate chapter to make them easier for engineers to use for retrofits and other projects that do not include the design of the CHW and condenser water systems addressed in this chapter.

Thermal Storage Thermal storage systems can offer significant energy cost savings where energy costs are based on time-of-day or real-time pricing. The design of these systems is beyond the scope of this SDL. For more information, refer to ASHRAE’s Design Guide for Cool Thermal Storage (Dorgan and Elleson 1993).

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References AHRI. 2001. AHRI 400-2001, Liquid-to-liquid heat exchangers. Arlington, VA: Air Conditioning, Heating and Refrigeration Institute. ASHRAE. 2016a. ANSI/ASHRAE Standard 62.1-2016, Ventilation for acceptable indoor air quality. Atlanta: ASHRAE. ASHRAE. 2016b. ANSI/ASHRAE/IES Standard 90.1-2016, Energy standard for buildings except low-rise residential buildings. Atlanta: ASHRAE. CBSC. 2016. 2016 California building standards code. California Code of Regulations, Title 24. Sacramento, CA: California Building Standards Commission. Dorgan, Charles E., and James S. Elleson. 1993. Design guide for cool thermal storage. Atlanta: ASHRAE. EDR. 1999. CoolTools chilled water plant design guide. Sonoma, CA: Energy Design Resources. McBride, M. 1995. Development of economic scalar ratios for Standard 90.1. Proceedings of Thermal Performance of the Exterior Envelopes of Buildings VI. Clearwater Beach, FL. Taylor, S. 2002. Degrading chilled water plant Delta-T: Causes and mitigation. ASHRAE Transactions 108(1). Taylor, S., and M. McGuire. 2008. Sizing pipe using life-cycle costs. ASHRAE Journal 50(10). Taylor, S. 2011a. Optimizing design & control of chilled water plants: Part 1: Chilled water distribution system selection. ASHRAE Journal 6:14–25. Taylor, S. 2011b. Optimizing design & control of chilled water plants: Part 3: Pipe sizing and optimizing T. ASHRAE Journal 12:22–34. Taylor, S. 2012. Optimizing design & control of chilled water plants: Part 4: Chiller & cooling tower selection. ASHRAE Journal 3:60–70. Trane. 2005. Trane Engineers Newsletter 34(1). ADM-APN014-EN. Trane. 2011. Chiller system design and control: Applications manual. Dublin: Trane.


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Skill Development Exercises for Chapter 5 Complete these questions by writing your answers on the worksheets at the back of this book. 5-1

You are responsible for developing the master plan for a new college campus that will include multiple buildings. Which of the following CHW distribution approaches will be the most energy efficient? a. Variable primary flow b. Primary, variable secondary c. Primary, variable secondary with tertiary bridge connected pumps at the buildings d. Primary, distributed secondary where the secondary pumps are sized for the pressure drop from and back to the primary loop

5-2

Your customer is building a sprawling six-story, 600,000 ft2 building with three wings, each of which is served by a large AHU. The building also has multiple computer rooms served by FCUs. All of the high density rooms are localized in the same area of the building. What is the likely to be the most energy-efficient CHW distribution design approach? a. Variable primary with a plant bypass leg and two-way valves at all AHU and FCU coils b. Constant primary, variable secondary with two-way valves at all AHU and FCU coils c. Variable primary with distributed variable secondary coil pumps for all AHUs and a separate variable secondary pump serving all FCUs with two-way valves d. Variable primary with distributed variable secondary pumps at all AHUs and all FCUs

5-3

When sizing piping to minimize life-cycle costs, which of the following are critical considerations? a. Expected system operating hours b. Variable-speed versus constant-speed operation c. Utility rates d. All of the above

5-4

Which of the following are benefits resulting from selecting condenser water loop T for a larger range? i.Reduced chiller energy use due to lower chiller lift ii.Reduced pumping energy use iii.Increased cooling tower capacity for a given tower selection iv.Reduced cooling tower scaling potential a. (ii.) and (iii.) b. (i.), (ii.) and (iii.) c. (ii.), (iii.), and (iv.) d. (i.), (ii.), and (iv.)

180


Which of the following are true when selecting cooling tower fan controls? a. Constant-speed towers often prove life-cycle cost-effective because of their low upfront costs. b. VFD fan speed controls are prohibitively expensive relative to the small energy benefit gained relative to two-speed motor controls. c. To achieve the energy benefits provided by variable-speed tower control, it is only necessary to install a VFD on the lead cooling tower. d. VFDs on all tower cells generally result in lowest life-cycle costs.

5-6

You are retrofitting an existing central plant serving a large office building with a WSE in a climate where freeze conditions are not a concern. The plant currently maintains minimum chiller head pressure in cool dry weather using a cooling tower bypass. How does this arrangement need to be modified to accommodate the WSE? a. No changes are needed. b. A separate cooling tower without condenser water bypass needs to be added to accommodate the WSE HX. c. Modulating actuators should be added to the condenser isolation valves on each chiller and the cooling tower bypass should be permanently shut. d. The WSE HX should be piped in series with the cooling towers and the cooling tower bypass should be shut permanently.

Chiller Procurement


Introduction This chapter discusses a novel approach to selecting chillers using lifecycle cost techniques. There are many chiller options that apply to a given project, and, in the case of centrifugal chillers, there are hundreds of options because these custom-built products include so many options (condenser and evaporator sizes and quantity of tubes, compressor and impeller sizes, motor drive and bearing options, etc.) The typical selection process is largely subjective, which seldom results in the optimum selection. A case study of the chiller selection process is provided for a new office building. Also provided are two sample chiller bid forms in spreadsheet format, available online at ashrae.org/ CHWSDL.

Chiller Procurement Procedures Table 6-1 compares typical versus recommended approaches for procuring chillers. The typical approach, shown in the table’s left column, is expedient, but it seldom leads to an optimum selection. There are simply too many selection options on large chillers to make a good choice by inspection. Not only are there often major design options such as centrifugal versus screw compressors, refrigerants R-123 versus R-134a (or, now, next-generation low-GWP refrigerants), constant versus variable speed, and conventional versus magnetic or ceramic bearings, there can be hundreds of combinations of evaporator, condenser, and compressor options for a given capacity that can radically affect chiller price and performance. The recommended approach to chiller selection is outlined in the right column of Table 6-1 and is discussed in more detail in the subsequent sections of this chapter. This method ensures selection is based on objective performance rather than subjective issues.

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Chapter 6 Chiller Procurement Table 6-1

Chiller Procurement Approaches

Typical Approach

Recommended Approach

1. Calculate or estimate the required plant total capacity. 2. Pick number of chillers, usually arbitrarily or as limited by program or space constraints. 3. Take plant load and divide by number of chillers to get chiller size (all equal). 4. Pick favorite vendor, often based on who is most attentive to the engineering firm. 5. Have vendor suggest one or two chiller options. 6. Select one option based on minimal or no analysis. 7. Bid the chillers along with the rest of the job and let market forces determine which chillers are actually installed.

1. Calculate the required plant total capacity and design temperatures and flow rates. 2. Develop a bid list of chiller vendors based on past experience, local representation, etc. 3. Request chiller bids based on a performance specification. Multiple options encouraged. 4. Adjust bids for other first-cost impacts. 5. Estimate utility costs of options based on detailed computer model simulations of the building/plant. 6. Estimate maintenance cost differences between options. 7. Calculate life-cycle costs. 8. Select the chiller option with the lowest lifecycle cost. 9. Hard-spec the selected chiller (no substitutions) and include contractor price in specifications.

The recommended performance bid approach should take place once plant capacity is well defined. This is typically at the end of the design development phase. The recommended approach has many advantages: • •

•

•

•

The owner generally benefits from lower life-cycle costs. Arbitrary selection of chiller vendor and model is eliminated, potentially lowering chiller costs due to a more competitive bid process. In traditional design/bid/build projects, the specified chiller is commonly provided in contractor’s bids even if it is not the lowest cost because of the uncertainty of the secondary cost impacts that might occur from a substitution. The procedure generally results in a more energy-efficient chiller selection. The traditional approach often leads to a low-cost mentality where the least expensive, and often least efficient, chiller is selected. Chiller vendors can make proposals that take advantage of their systems’ strengths or “sweet spots” both for cost and efficiency. The conventional approach, where size and efficiency are more arbitrarily selected, usually favors one vendor (inadvertently or intentionally) who happens to have a sweet spot for the selected piece of equipment. Chiller selections are finalized in the design stage. Selection at this time allows the designer to customize the design of the plant, including physical layout and chiller-specific design parameters such as minimum flow (which affects minimum flow bypass line and valve size), knowing that the chiller selection will not be changed at bid time. The traditional approach


183

can result in substitutions at bid time, which can lead to coordination issues (and costs) due to changes in size, weight, peak power, minimum flow rates, etc. There are also disadvantages to this approach: • •

•

•

It takes more time, both for the engineer and the equipment vendors.1 Assumptions made in the energy calculations may prove to be incorrect. For instance, utility rates, internal loads, and occupancy patterns assumed in the analysis may not be correct or may change over time. Energy escalation and inflation rates used in the analysis also have considerable uncertainty. The computer model used to calculate energy usage is imperfect in the way it models building and system loads, the chiller plant and pumping systems, and plant control strategies. Optimum plant control strategies are often complex and can vary from one chiller option to another. These strategies cannot always be modeled accurately by existing energy simulation tools. The benefits of long-term product reliability and vendor support are seldom included in the life-cycle cost calculations because their cost benefits are difficult to estimate, although they may be significant.

Despite these disadvantages, the recommended chiller procurement approach represents an improvement over conventional approaches and has been successfully used on several projects. The case study at the end of this chapter is based on an actual project that followed this recommended approach. Note that this approach has been used successfully on projects that require competitive bidding, such as most state and General Services Administration projects. This is because most of the statutes mandate only that competitive bids occur, not that they occur at the traditional bid time when the design is complete. Most allow metrics such as life-cycle cost (not first costs) to be the basis of selection provided the selection criteria are well defined.

Procurement Step #1: Calculate Plant Capacity and Design Conditions See Chapter 2 for an in-depth overview of available methods of calculating or estimating the required plant total capacity (tonnage). The chiller procurement procedures described below assume that chiller plant design parameters, such as condenser and CHW entering and leaving temperatures and flow rates, have been determined. For retrofit applications these design parameters are often set by the constraints of the existing system design. Chapter 5 describes procedures for selecting these design parameters for new chiller plants. 1. Note: ASHRAE is developing a new standard, ASHRAE Standard 205, Standard Representation of Performance Simulation Data for HVAC&R and Other Facility Equipment, that will define standard data models and formats for use in creating accurate equipment energy models. This will substantially reduce or eliminate the time burden on equipment vendors—the performance data will be generated by chiller factory engineers—and it will improve the accuracy of chiller models. The standard is expected to be published in 2017.

184

Chapter 6 Chiller Procurement

Procurement Step #2: Develop Vendor Bid List Pick a list of chiller vendors based on past experience, local representation, and other factors. Not all of the major vendors need to be on the bid list; however, the more vendors on the bid list, the better the competition will be, and, therefore, the lower the chiller energy costs and first costs are likely to be. However, low first costs and energy costs are not the only selection criteria: low maintenance costs, system reliability, vendor support, and other issues are also considerations. This approach also works when the owner has a specific chiller manufacturer they want to use based on past experience or to maintain consistency with chillers already installed at their building or campus. Optimizing chiller selection with a single vendor is still valuable because within their product line there are many options, such as larger heat transfer areas in condensers and evaporators and VFDs. It is not uncommon for centrifugal chiller selection software to generate hundreds of selections for a single set of capacity and design conditions.

Procurement Step #3: Obtain Chiller Bids Performance Specification First, develop a performance bid specification that includes all design constraints for the chillers. In the conventional procurement approach, a chiller is typically specified by capacity, efficiency, construction details, etc. These details, however, are not usually appropriate for a performance specification. Vendors should be encouraged to propose as many chiller options as possible without constraints—the LCCA will determine which option is the best, so constraints are not necessary and may inadvertently eliminate better choices. As a rule, do not include the following in the performance specification, instead leaving them up to the bidders: •

•

• • •

Number of chillers and capacity of each chiller. (For some projects, the engineer may decide to fix chiller quantity and size based on other factors or for simplicity.) Chiller efficiency metrics (kW/ton, IPLV, etc.), although the specification should state that all selections must meet all full-load and part-load energy code requirements. Variable speed or fixed speed. Conventional or magnetic or ceramic bearings. Refrigerant type.

Encourage multiple options from each vendor. They may, for instance, propose one design that uses several equally sized chillers and another that includes unequally sized chillers or a pony chiller. Another option may include VFDs on one or more of the chillers. Encourage the vendors to be imaginative.


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Also eliminate unnecessary boilerplate requirements. Conventional specifications often contain a long list of construction details that are sometimes proprietary and usually not very important from a quality and reliability perspective. If a chiller vendor has been approved for the bid list, then that should mean that the vendor’s standard construction details are acceptable. Including endless boilerplate requirements in the specifications is usually a waste of time and can be counterproductive if they mask truly important specification details. The following details should be included in the performance specification: •

Total plant capacity. In general, specify only total plant capacity, not the capacity of each chiller. See Chapter 2 for an overview of methods of determining plant total tonnage.

•

The minimum number of chillers or required level of redundancy. For redundancy and reliability, the specification may require, for instance, that there be at least two chillers in the plant, or possibly at least two compressors to allow for dual-compressor options. Alternatively, for critical plants, such as those serving an industrial process or data center, the specification may require N+1 redundancy, meaning that the chiller plant must be able to handle the load even with the failure of the largest chiller.

•

Design conditions. Design condenser and evaporator flow rates and entering and leaving water temperatures of the plant as a whole must be specified. For retrofit applications these factors are usually fixed by the constraints of the existing design. Chapter 5 describes how to determine these parameters for new buildings.

•

Energy sources available, such as electricity or gas. This procurement procedure is not limited to electric chillers, although mixing gas and electric chillers will require more cost adjustments to account for the larger heat rejection required by absorption chillers and for differences in gas and electric utility service sizes and distribution systems.

•

Typical cooling load profile. In order to offer intelligent proposals, the chiller vendors need to know the cooling load profile, typically developed from a computer model of the systems served by the plant (see Chapter 2). For instance, the plant with the load profile shown in Figure 2-1 for a typical office building will require much different chiller sizing for best efficiency than the plant with the load profile shown in Figure 2-2, which has a relatively small data center that provides a constant base load. The plant serving Figure 2-2 loads will most likely require a small pony chiller or unequally sized chillers for best performance; so, it is important that the vendors bidding the chillers be aware of this need.

•

Limited mechanical room space. If this is a factor, include a plan of the mechanical room in the specification.

•

Available voltage and, for retrofit projects, available capacity of the electrical service.

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Chapter 6 Chiller Procurement •

Type of allowed refrigerants. For instance, the owner may disallow R-123 as an option because it is to be phased out of production in the near future in developed countries, or the owner may prefer R-123 because it has a lower GWP. (See Chapter 3 for additional comments on refrigerant issues.)

•

Sound power limits. If noise is a design constraint because the chiller is located adjacent to noise-sensitive spaces, an acoustical engineer should back-calculate the maximum chiller sound power levels needed to meet noise criteria levels in occupied spaces. These limits should then be included in the performance specification. Either the proposed chillers have to meet these sound power limits inherently or the vendor has to include a chiller sound enclosure in their proposal.

Performance Verification The following are optional requirements in the specification to ensure that the performance data from the chiller manufacturer are accurate.

Factory Tests Factory testing is an option offered by almost all manufacturers. It can be witnessed by the owner’s representative (often the design engineer) or nonwitnessed. The latter is usually much less expensive because testing can occur right when the chiller comes off the assembly line; witnessed tests must be arranged well in advance and can cause production inefficiencies, in addition to the cost of travel for the witness. It is highly recommended that factory tests be required, despite their cost, for these reasons: •

Factory tests make it possible to reject the chiller if it does not meet performance specifications. Once a chiller is installed in the field, rejecting the chiller is no longer a practical option.

•

Testing is more accurate in the factory than in the field, resulting in less controversy about whether and to what extent chillers fail to meet performance requirements. The factory test stand is fitted with calibrated laboratory-grade flowmeters and temperature sensors. Even if accurate instruments were available in the field, their accuracy is easily questioned by the chiller manufacturer should the tests indicate chiller underperformance.

•

Field tests cannot supply the necessary load and ambient conditions to test the chillers at all the specified conditions.

•

Factory tests obviate the need for field tests for commissioning or performance verification purposes. Tests are more easily and less expensively done in the factory than in the field, and often field tests are not practical because of incomplete instrumentation (e.g., lack of a condenser water flowmeter).


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Liquidated Damages If performance tests are required, it is recommended that monetary penalties apply if a chiller fails to meet performance requirements. Without “teeth” in the specification, vendors will be more likely to exaggerate chiller performance in their proposals. An example liquidated damages clause follows: Chiller manufacturer shall repair or replace equipment, at no cost to the Buyer, until equipment is certified by AHRI 550/590 test procedures to meet the performance indicated in the manufacturer’s proposal. In the event that these revisions do not achieve submitted performance or require longer than 2 weeks, the following penalties will be imposed. 1. Capacity penalty: For each ton below the proposed design capacity, one thousand dollars per ton will be deducted from the contract price. 2. Power consumption penalty: The power consumption penalty for each tested load point shall be = [Measured kW – (Measured Tons × Proposed kW/Ton)] × $2000/kW. 3. Total penalty: The total performance penalty will be the sum of the capacity penalty and power consumption penalty at each tested load point for each chiller tested.

Zero AHRI Tolerance AHRI Standard 550/590 (2015) allows a tolerance in the measured versus actual chiller capacity and efficiency. The magnitude of the tolerance varies as a function of load, increasing as load decreases, as shown in Figure 6-1. The tolerance curve, established over 30 years ago, was intended to account for variations in manufacturing, but over time manufacturers have found that their manufacturing tolerances are smaller than AHRI standard tolerances. Market forces then caused manufacturers to take credit for the tolerances in the chiller performance reported by their rating programs. In other words, the power reported by manufacturers’ rating programs for a given chiller has been decreased from the actual power required by an amount equal, or almost equal, to the allowed AHRI tolerance. Obviously, to get an accurate picture of predicted chiller performance in the LCCA, actual chiller performance data must be obtained from chiller vendors. Figure 6-2 shows predicted chiller performance for the same chiller with and without AHRI tolerance assuming standard AHRI condenser water relief. The AHRI tolerance data were from the chiller manufacturer’s selection program without adjustment. The zero tolerance data were from the same manufacturer for the same chiller but with the stipulation that the chiller be factory tested and must meet predicted performance claims with zero tolerance. It is clear that low-load chiller kW/ton is significantly underpredicted when AHRI tolerance is allowed. This can completely skew chiller selection. For instance, the manufacturers who take the most advantage of the AHRI tolerances (not all push the limit) will

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Figure 6-1

AHRI Standard 550/590 tolerance.

Figure 6-2

Chiller-predicted performance with and without AHRI Standard 550/590 tolerance.


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have a significant advantage. If the analysis was in part to determine how many chillers or what size they should be for plants with unequally sized chillers, the AHRI tolerance can result in an incorrect answer because it makes it appear that large chillers are efficient at low loads. It is strongly recommended, therefore, that the performance claims by manufacturers be required to have zero AHRI tolerance. The bid specification should state that no tolerance, such as that normally allowed by AHRI Standard 550/590, will be accepted on proposed efficiency and capacity data. Note that obtaining zero tolerance data from vendors will often elicit complaints because it can be difficult due to selection software constraints, but it is absolutely essential for accurate analysis and proper chiller selection.

Other Bid Inclusions The project construction documents will ultimately include contractor pricing for the chillers in the bid specifications. For pricing to be complete, costs normally included in vendor proposals must be clearly identified and included in the performance bid. This includes cost of freight to the job site. Including local sales taxes is also a good idea because some part of the chiller bid may not be taxed, such as freight and some start-up services.

Performance Bid Forms Bid forms are used to collect both chiller pricing and chiller performance data. It should be made clear to chiller vendors that chiller pricing and selection using this procedure are final; there will be no opportunity to reprice chillers when the project is bid at the 100% construction documents phase. To create chiller simulation models using common simulation programs such as DOE-2.1E, DOE-2.2, DOE-2.3, and EnergyPlus, chiller performance data over a wide range of operating conditions must be collected. From these data, regression models of the chillers can be made so that the chiller’s performance can be accurately modeled. A sample performance spreadsheet (Chiller Bid Form) is included as supplemental material for this SDL (available at ashrae.org/CHWSDL). The form includes pricing and also generates full-load and part-load performance data tabs based on design conditions that the chiller vendor must complete. The performance data tabs are designed to ensure that all operating conditions expected for the particular project are within the bounds of the collected data so that regression models do not need to extrapolate. For instance, part-load data are collected at the chiller’s minimum and maximum lift conditions. The spreadsheet includes a macro that generates the regression coefficients for DOE-2.1E, DOE-2.2, DOE-2.3, and EnergyPlus chiller models once the data fields have been completed. The regression coefficients can then be entered directly into the programs to create accurate models of the proposed chillers.

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Procurement Step #4: Adjust for Other First-Cost Impacts After obtaining chiller bids, adjustments must be made to account for other installation cost impacts. Chiller prices obtained through the bidding process are not necessarily directly comparable. Here are some reasons prices may not be apples to apples: •

•

•

• •

One option may have more chillers than another (for instance, three versus two). The added cost of rigging, piping, wiring, and controls for the additional chiller must be accounted for. Three-pass evaporators can increase piping costs compared to two-pass evaporators. Sometimes, depending on the plant layout, the opposite may be true. Open-drive compressors require more chiller room cooling than hermetic motor options. In many climates, cooling is not needed at all for chiller rooms housing hermetic chillers because the room has very few heat sources (only those due to the inefficiency of pump motors) and substantial heat sinks (e.g., heat flow into CHW piping and evaporators). But heat gains from open-drive chiller motors are substantial, requiring mechanical cooling of the chiller room in almost all climates. (The energy used by this mechanical cooling also must be addressed in the energy calculations.) One option may have less efficient chillers, which can affect electrical service costs. One proposal may include a control system that can be directly connected to the BAS network, while another may require a field-installed gateway.

For the first pass of the LCCA, these secondary first-cost impacts can be roughly estimated. Then, if the chillers for which rough estimates were made end up with the lowest life-cycle costs, a more accurate estimate of the secondary costs should be made. This is best handled by a contractor. If a contractor is not part of the design team, a local mechanical contractor who is likely to bid the project should be consulted. Many will provide rough estimating services at no cost to the project—they write off the costs internally as a marketing expense, and they also have the advantage of gaining project familiarity prior to bid.

Procurement Step #5: Estimate Utility Costs One of the most important elements of the CHW plant procurement process is accurately estimating the utility costs of each chiller option. Utility costs include electricity, gas (if applicable), and water (especially if both waterand air-cooled chillers are being considered). The level of accuracy and detail necessary for energy calculations depends on the project’s size and engineering budget. The primary engineering cost of the computer model is not creating the chiller plant model itself but creating the model of the building and systems generating the plant load.


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Many modern projects include energy modeling in order to meet energy code requirements (e.g., the energy cost budget approach to ANSI/ASHRAE/ IES Standard 90.1 compliance [2016]), utility incentive programs, or green building rating programs such as U.S. Green Building Council’s (USGBC’s) Leadership for Energy and Environmental Design (LEED). In this case, the added cost to perform chiller life-cycle costing is relatively low, particularly when the design team becomes more adept at the process over time. For projects that are not using energy modeling for other reasons, the cost of creating the plant profile can be substantially reduced by using prototypical profiles developed from generic models of buildings with similar occupancies and HVAC systems. For instance, the eQUEST program (a free front-end interface to DOE-2.2) has a wizard that can create a prototypical building in less than an hour. For existing projects, measured performance data may be used. See Chapter 2 for a more detailed discussion of load profiles and energy simulation programs. Once performance bids from chiller vendors have been obtained and a computer model of the building and HVAC systems is complete, create regression model coefficients of each of the proposed chillers using the macro within the performance spreadsheet (Chiller Bid Form) provided with this SDL (available at ashrae.org/CHWSDL). Then enter the regression coefficients into the simulation program. The following factors should be considered in the simulation.

Utility Rates Utility rates vary over time and are difficult to predict, particularly given the current transition from regulated to unregulated utility markets. There are a number of energy forecasts that project electricity and natural gas costs up to 30 years in the future (e.g., those of the Department of Energy/Energy Information Administration [DOE/EIA], American Gas Association [AGA], Gas Technology Institute [GTI]) that attempt to account for the impact of deregulation as well as world oil and gas reserves. However, these reports are typically generalized for the nation or the world as a whole and may not reflect local utility rate trends. Given this uncertainty and for the sake of simplicity, it is typically assumed that current rate structures will be in effect during the chiller plant’s life cycle. The LCCA (discussed below [Procurement Step #7: Calculate Life-Cycle Costs]) includes utility escalation rates that can roughly address future rate changes. Virtually all utilities charge both for energy consumption as well as for power demand. Because chillers are one of the largest energy users in typical buildings, it is essential that demand charges be properly taken into account. This is particularly true when demand charges are ratcheted, meaning the owner pays some percentage of the maximum peak demand over the year, regardless of actual monthly demand. Most hourly simulation programs allow rate schedules to be entered explicitly—this is usually much more accurate than using effective annual rates.

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Control Logic and Set Points The optimum control logic will vary for different chillers. It is therefore essential to iterate on control logic and set points in order to simulate the chillers as they will perform (presuming of course that sequences are customized to the chiller as recommended in Chapter 7). Unfortunately, most simulation programs are not able to precisely model the sequences recommended in Chapter 7. But they can get close. The following are some control variables and logic that should be optimized in the model.

Tower Fan Control (Condenser Water Set Point Reset) As noted previously, condenser water temperature/tower fan control options in energy programs are limited. It is recommended that various tower control strategies that may be available in the simulation tool be modeled to see which option is best. The actual logic recommended in Chapter 7 will (according to theoretical optimum plant performance [TOPP] models) be better, but optimizing within the available program model options should be close enough for an apples-to-apples comparison of the chiller options. For instance, one or more of the following control models for condenser supply temperature set point should be available in the simulation program; all of them should be tried to see which results in the lowest chiller energy usage. •

•

•

•

Constant set point: Iterate on the set point from the minimum allowed by the chiller manufacturer up to the design set point. The optimum is likely to be closer to the minimum for most centrifugal chillers, particularly those with VFDs. A set point equal to the design wet-bulb temperature has proven to be a surprisingly effective set point for constant-speed chillers. Wet-bulb reset: This is usually modeled as a constant approach to outdoor air wet-bulb temperature within limits. The offset and limits should be adjusted to minimize energy usage. As noted in Chapter 7, this strategy requires that a humidity or wet-bulb sensor be installed in the control system, which increases maintenance costs and reduces reliability. Load reset: This is usually modeled as a linear reset from a given temperature at 100% load to some lower set point at a lower load percentage. The various set points should be adjusted to find best performance. This logic is the closest to the sequences recommended in Chapter 7. Optimum tower set point: Some energy analysis tools dynamically determine the tower set point to optimize the sum of chiller plus tower energy use.

Staging Fixed-speed chillers are almost always optimally controlled by running the fewest chillers necessary to meet the load. This is easily modeled by most plant simulation tools and is usually the default control strategy.


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Variable-speed chillers, however, use less energy when they are operating at low-lift conditions, except at very low loads. Therefore, a plant with multiple variable-speed chillers will be more efficient when more chillers are operating at part load than if fewer are operating near full load, despite the additional energy used by CHW and condenser water pumps. (See the Control Sequences section in Chapter 7 for an additional discussion of optimizing CHW plant design.) For estimating energy usage of chiller plants with multiple variablespeed chillers, the staging on/off point should be tested to find the optimum.

Chilled-Water Reset As noted in Chapter 7, CHW temperature reset is almost always the best strategy versus a fixed set point, even for plants with variable-flow variablespeed CHW pumping systems. Reset strategies available to most simulation tools are limited. Here are a few common models: • •

Reset by load (valve position): This is the most effective reset logic, as noted in Chapter 7. Reset by outdoor air temperature: This model inversely resets CHW temperature proportionally to outdoor air temperature.

The simulation program’s ability to accurately simulate reset is a significant issue. The program must be able to model the effect reset has on coolingcoil effectiveness; it requires an accurate coil model and most programs (including DOE-2.1 and 2.2) have poor models that do not properly account for variations in coil performance with flow and entering water temperature. If all the chillers being evaluated are of a similar type (for example, all are centrifugal chillers), it may be best not to model reset to avoid skewing the results due to coil modeling errors.

Variable-Flow Pumping Systems Most plant computer models are limited in their ability to accurately model variable-flow pumping systems, particularly primary-only systems. It is often necessary to “fake” modeling the actual system in the program by changing default. performance curves or adjusting pump heads. This can require considerable judgment on the part of the engineer doing the modeling. Fortunately, in most cases, the pumping scheme is the same for all the chiller options being considered, so errors in the model tend to cancel out when options are compared. The important differences between chiller options that must be accounted for are variations in condenser and evaporator pressure drops and, for variable-flow primary systems, the minimum and maximum evaporator flow rates. Pressure drop differences can usually be accurately accounted for by adjusting pump heads in the computer model and, in any case, have a small impact on overall energy. Differences in minimum and maximum flow rates with variable-flow primary systems cannot be modeled well using current simulation programs. However, the impact of varying minimum flow rates is usually small unless there are

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Chapter 6 Chiller Procurement large differences between options and many hours when the plant operates at low loads. The impact of varying maximum flow rates is seldom an issue because the maximum rates seldom occur unless the system experiences very bad T degradation. Variable-flow condenser water systems are also not modeled well with some programs (including DOE-2.1) because they use condenser water supply temperature in chiller regressions. This is accurate for constant condenser water flow but not for variable flow because it is condenser leaving water temperature that correlates to condenser temperature and pressure (which is the drive for chiller efficiency), and leaving condenser water temperature rises as condenser water flow rate falls. DOE-2.3 and EnergyPlus include improved models that account for condenser water flow variation.

Procurement Step #6: Estimate Maintenance Costs Maintenance costs are more difficult to estimate accurately than energy costs. There are few data available indicating the relative maintenance costs of various chiller types (for example, screw, centrifugal, absorption, enginedriven) and among the various manufacturers of each chiller type. Manufacturers make claims about their products’ advantages, but they seldom have hard, independently collected data to support those claims. To further complicate the issue, annual maintenance costs are not constant. The costs are low for the first few years, jump during years when a complete overhaul is required, and increase gradually as equipment wears. The length of time that different pieces of equipment last before they must be replaced also varies, although usually this is not an issue in chiller selection unless very long life cycles are analyzed. Because maintenance costs are difficult to estimate, they are often ignored in the chiller selection life-cycle costing and considered only as a “soft” issue when making the final chiller selection. This is probably a reasonable approach when the number and types of chillers are the same. However, when the number of chillers in each option varies, both air- and water-cooled options are being considered, or chillers of different types (e.g., electric and gas-engine driven) are being considered, maintenance costs need to be included explicitly in the life-cycle cost calculation for best results. One method to estimate long-term maintenance costs is to request a five- or ten-year full maintenance contract bid, perhaps including refrigerant replacement, from each manufacturer. This gives some measure of the manufacturers’ costs for maintenance. This approach should be used with care; sometimes quoted costs and actual maintenance costs may not jibe, depending on the willingness of the vendors (who are not insurance or actuarial professionals) to take on risk. Factors that affect maintenance costs include the following: •

Gear versus direct drive: Gear-drive machines have slightly higher maintenance costs than direct-drive machines.


•

•

• •

•

•

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Open-drive versus hermetic motors: Open-drive machines have slightly higher maintenance costs than hermetic machines, although they are less costly to repair should there be a motor burnout. (In chillers with hermetic motors, motor burnout typically contaminates the refrigerant.) Variable-frequency drives (VFDs): VFDs introduce another component subject to failure into the system. Few data are available on their reliability because the latest generation of VFDs is fairly new. It is also likely that VFDs will need to be replaced one or more times in the chiller’s lifetime. Accounting for VFD replacement cost should be considered if the years of analysis exceeds 15 years, the typical expected service life of VFDs. Varying manufacturer quality: There are many claims and anecdotes but few data on who makes the best chiller. For the purpose of maintenance cost analysis, it is generally assumed that if the chiller manufacturer is on the bid list, they make a reliable product. Compressor type: Screw chillers usually require less maintenance than centrifugal chillers, although this varies from manufacturer to manufacturer. Number of chillers (for options where the number of chillers differs): The more chillers in the system, the higher the maintenance costs will be, even for the same total plant capacity. The costs for maintaining a chiller are not strongly dependent on the chiller’s size. If costs are not available, a reasonable estimate of annual maintenance cost per water-cooled chiller is about $2000 to $4000 for the first few years, with an additional $1000 per year in repairs after the first five years or so. There is a slight increase in the permachine maintenance cost at about 1000 tons, and there may be another incremental jump at 2000 tons. Air versus water cooled: Whether the system is air or water cooled has by far the most significant impact on maintenance costs among the issues listed here. Water-cooled systems always cost more to maintain due to the constant water treatment requirements and the need for regular tube cleaning. Water-cooled chillers also generally last longer, particularly in harsh environments such as near oceans where salt in the air can significantly shorten the life of air-cooled condensers. To estimate the differences in maintenance costs between air- and water-cooled systems, request input from local HVAC service companies. ASHRAE Handbooks (particularly ASHRAE Handbook—HVAC Applications, Chapter 37 [2015]) also provide some maintenance cost information. Absent other information, a reasonable estimate of annual maintenance cost savings is $1000 to $2000 per year per chiller for the first five to ten years (plus the cost of cooling tower maintenance and chemicals). Gas-engine-driven chillers: Gas-engine-driven chillers are known to have much higher maintenance costs than either electric or absorption chillers. These chillers have frequent engine-related maintenance requirements (e.g., for spark plugs, oil and oil filters, air cleaners, belts) that are not required for other chiller types. Manufacturers of these chillers can provide estimates of the frequency and cost of this work.

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Procurement Step #7: Calculate Life-Cycle Costs The life-cycle cost of a chiller plant is the present value of the total cost of owning and operating the plant over a specified period of time. A detailed description of the parameters used in calculating life-cycle cost is beyond the scope of this SDL. For more details, refer to the 2015 ASHRAE Handbook— HVAC Applications, Chapter 37 and other engineering manuals. Below is a brief summary of the relevant variables and formulas for calculating life-cycle costs. These formulas should be entered into a spreadsheet so that the sensitivity of various assumptions can be evaluated. (See the Case Study section for an example that uses these equations.) The life-cycle cost can be calculated using the following equation: N

LCC = FC +

UC j + MC j

 --------------------------1 + d  j

(6-1)

j=1

where LCC = FC

=

UCj

=

MCj d

= =

N

=

present value of the owning and operating costs of the chiller plant first costs of the plant. In this case, this is the cost of the chillers as proposed by the vendor adjusted for associated installation factors plant utility costs for year j. If there is more than one utility type (e.g., electricity, gas, water), this component is duplicated for each relative maintenance costs for year j discount rate, also called the cost of capital or the minimum rate of return. This rate is used to discount future cash flows, converting them to present costs. The higher the rate, the less future energy savings will help offset the first-cost penalty of a more expensive chiller plant. The discount rate for a typical business owner might be 8% to 15%, reflecting the cost of borrowing money plus a few percentage points to reflect the investment risk. A more progressive owner or government entity may be less conservative and use rates near 5% number of years of analysis, or life cycle. While called the life cycle, N is seldom equal to the actual number of years the chiller plant will operate. Most studies only look at the first 10 to 15 years because there are so many uncertainties in the utility costs and overall energy usage the further into the future one looks. Also, most businesses will expect a return on investment in well less than 10 years


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The life-cycle cost of Equation 6-1 does not include tax impacts, depreciation, investment tax credits, financing costs, or salvage value. Please refer to more detailed texts on life-cycle costing if these factors are considered significant enough to take into account. If energy and maintenance costs are constant, Equation 6-1 can be simplified to LCC = FC + PWFe · UC + PWFm · MC

(6-2)

where  1 + e'  N – 1 PWF e = -----------------------------e'  1 + e'  N d – e e' = ---------------1 + e  1 + m'  N – 1 PWF m = -------------------------------m'  1 + m'  N d – m m' = -----------------1 + m UC MC e

= = =

utility costs at current rates maintenance costs at current prices escalation (inflation) rate for electricity (or whatever fuel type is being used) above current rates. Estimating escalation rates is very difficult in this transitional market between regulated and unregulated utilities. See the discussion under the Utility Rates

section above m = maintenance cost escalation (inflation) rate. This can typically be assumed to be equal to the consumer price index (CPI), although, like energy escalation rate, this is difficult to predict with any certainty. A 1% escalation is usually reasonable PWF = present worth factor

Procurement Step #8: Final Chiller Selection Evaluating chiller options using the procedures recommended in this SDL requires making many assumptions and simplifications. To pick the “best” chiller plant option, it is important to test the sensitivity of various assumptions. For instance:

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If there is uncertainty about the loads the plant will need to handle, develop energy costs under various load profiles. For example, a plant may be expected to ultimately handle a large data center load, but it is possible, even likely, that the actual load will be smaller. Be sure that the plant can efficiently operate at low loads and that the plant that proved to be optimum under high-load conditions is also near optimum under low-load conditions.

•

Utility rates may change dramatically when utility markets are fully deregulated. Experiment with rates that include very high on-peak energy charges and demand charges. For mixed-fuel plants, experiment with escalating rates for one utility and deescalating rates for the other.

•

Life-cycle cost assumptions, such as discount rate and years of analysis, affect how much future energy savings are weighted. Experiment with various assumptions to see how they affect the results.

More than likely, after calculating results over a range of assumptions, the life-cycle costs of several chiller options will be close to the optimum option (the one with the lowest life-cycle cost). It is also possible that the optimum choice will vary depending on the assumptions made. In this case, the final selection must also consider “soft” factors (those to which a dollar amount cannot be easily attached) to break the tie. The final selection should be made by the entire design and ownership team, not just by the engineer. This will ensure that all soft issues have been considered and everyone had a fair chance to express any vendor preferences. These soft factors include the following: •

Reliability and reputation of the manufacturer and local representation: Most facility owners and operators will have a favorite chiller vendor based on past experience. This can be a very important tie breaker when making the final selection. If the owner’s favorite vendor offers an option that has close to the optimum life-cycle cost, it may be a politically sensible idea to choose that option.

•

Refrigerant type: Even if a range of refrigerant options is allowed in the bid specifications, the owner or engineer may have a preference for a given refrigerant type based on the refrigerant’s impact on the ozone layer and global warming, or impending production phase-out dates.

•

Redundancy: Typically, an analysis of chiller options assumes that the chiller plant operates normally. But what happens if a chiller fails? Different chiller options may accommodate outages better than others. Options offering the most chillers will generally offer the least exposure to chiller failure. However, other more complex failure considerations should be considered. For example, if a plant with unequally sized chillers loses the large CHW pump, will the flow ranges allow the large chiller to stay on-line using the small CHW pump?


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Procurement Step #9: Chiller Specification Once the final chiller selection is made, it must be made clear to contractors bidding the project at the 100% construction documents phase that the chiller selection is final. The following are guidelines to achieve this. •

•

Final equipment schedules and plans must, of course, reference the selected chillers. Add “NO EQUAL” to the remarks column to make it clear this is the final selection. Strip the chiller specification of the normal Part 2 Materials Section and most of Part 1 General; retain Part 3 Installation. In Part 1, state that the chiller must be the one scheduled and include chiller prices. For example: a. Provide the scheduled chillers, including the scheduled options for the following prices: 1. CH-1: $101,406 2. CH-2 and CH-3: $189,450 each 3. Price includes freight but does not include taxes b. Bids for other chillers will not be accepted.

Because this procedure may be unusual to some contractors, contractor bids should be reviewed to ensure compliance.

Case Study This case study illustrates the recommended chiller procurement approach discussed in this chapter. The case study building is 15 stories high, enclosing 540,000 ft2 in San Francisco, CA. The building primarily contains offices, with some assembly and retail space on the ground floor, a 5000 ft2 data center, and a large cafeteria. The data center and various small server rooms operate continuously and place roughly a 50 ton base load on the plant. Total plant load was calculated to be 1100 tons. Figure 2-2 in Chapter 2 shows the load profile for the building generated by a DOE-2.1 simulation. After the primary design temperatures and flow rates were determined (see Chapter 5), the performance specification and bid forms were developed (see the sample spreadsheet bid form [Chiller Bid Form] included with this SDL, available at ashrae.org/CHWSDL). Four chiller vendors were invited to bid. The data from the bid forms were used to generate regression coefficients of each chiller in each bid option. These coefficients were then input into the DOE-2.1 model of the building. Energy savings were calculated using various control scenarios such as condenser water set point reset and staging points. The lowest energy costs for each option were selected and life-cycle costs were calculated as shown in Table 6-2. Note that first-cost adjustments were made for additional piping and pumping system costs caused by the use of three-pass evaporators in some cases and multiple chillers in others. Additional exhaust fan capacity was added for open-drive machines due to the higher heat load they generate in the chiller room. (In more severe climates, this high heat load may require that mechanical cooling be added to the room.)


Description

1

400 ton, 0.50 kW/ton; 700 ton, 0.55 kW/ton

268,235

6000

0

274,235 6 712,293

9

6,371,092 142,016 10

2

400 ton with VFD, 0.50 kW/ton; 700 ton, 0.55 kW/ton

301,235

6000

0

307,235 9 694,427

2

6,251,168 22,092

3

365 ton, 0.56 kW/ton; 735 ton, 0.50 kW/ton

199,980

3000

0

202,980 1 724,350

12

6,403,038 173,962 12

4


240,130

0

0

240,130 3 702,168

5

6,250,322 21,246

3

5

365 ton with VFD, 0.56 kW/ton; 735 ton with VFD, 0.50 kW/ton

289,190

0

0

289,190 8 694,854

4

6,236,778

2

6

200 ton, 0.50 kW/ton; 900 ton dual 0.54 kW/ton

212,031

0

0

212,031 2 712,100

8

6,307,235 78,159

7

550 ton dual, 0.56 kW/ton; 550 ton dual, 0.56 kW/ton

256,485

0

0

256,485 5 714,269

11

6,370,255 141,179 9

8

400 ton dual, 0.53 kW/ton; 700 ton dual 0.53 kW/ton

254,533

0

0

254,533 4 711,137

7

6,341,495 112,419 8

9

200 ton, 0.53 kW/ton; 350 ton dual 0.57 kW/ton; 550 ton dual 0.59 kW/ton

252,485

25,000

0

277,485 7 712,547

10

6,376,516 147,440 11

10


307,106

6000

3000

316,106 11 702,969

6

6,333,154 104,078 7

11


305,151

6000

3000

314,151 10 691,038

1

6,229,076

365 ton with VFD, 0.52 kW/ton; 12 366 ton, 0.51 kW/ton 366 ton, 0.51 kW/ton

338,320

33,000

3000

374,320 12 694,806

3

6,321,497 92,421

Total Cost, $

Option

LCC Premium vs. Base, $ LCC Rank

Life-Cycle Cost Summary

Cost with Contractor Markup, $ Other First Cost Add/ Deduct, $ Added Exhaust Fan, $

Table 6-2

1st Cost Rank Total Building Energy Costs, $ Energy Cost Rank Life-Cycle Cost, $

200

7702

0

The total building energy cost column was taken from the DOE-2 model. Life-cycle costs were calculated based on a 15-year life, 8% discount rate, and 0% energy escalation rate. Option 11 has the lowest life-cycle cost, but Option 5 is very close and even Options 2 and 4 have life-cycle costs close enough to that of Option 11 that they need to be considered. Ultimately, Option 5 was selected for the project based on the following soft considerations:

4

5

1

6


•

•

•

•

201

This option uses R-134a, whereas Option 11 uses R-123. R-123 was considered acceptable, but R-134a was preferred by the owner because it has zero ODP and was not scheduled to be phased of production at the time of design for this project. Both chillers in Option 5 have VFDs, whereas only the small chiller in Option 11 has a VFD. Having two VFD chillers improves redundancy because either chiller should be able to handle the low nighttime data center loads. Both chillers in Option 5 have design CHW and condenser water flow rates and minimum flow rates that overlap so that the small CHW pump is large enough to keep the big chiller on-line. That is not the case with Option 11. In case of any pump failure, the large chiller can remain on-line with Option 5, whereas if one of the large pumps fails, only the small chiller can remain on-line with Option 11. Option 5 uses hermetic motors whereas Option 11 uses an open drive. Both have advantages and disadvantages, but the owner’s engineer preferred hermetic motors based on past seal problems experienced with open-drive machines. The hermetic motors also made the chiller room have a net neutral load so no conditioning was required. Option 5 is less expensive by about $15,000, making it easier to reach firstcost budgets.

After the bid, chiller vendors were given a modified version of Table 6-2 that included only the columns for first-cost rank, energy cost rank, life-cycle cost savings versus base, and life-cycle cost rank. This allowed the vendors to see how their proposals compared to their competitors without seeing actual pricing.

Simplified Procurement Procedure The selection procedure described in this chapter takes considerable time and cost on the part of both the design engineer and chiller vendors. For many projects, the schedule may not allow time for this rigorous analysis or the cost may not be in the design budget. On some projects, the design team may not have the expertise to perform the required energy modeling and life-cycle cost calculations. In these cases, instead of falling back on the traditional, almost arbitrary, selection process, we suggest using the following simplified procedure: 1. Once loads are confirmed (typically at the end of the design development phase), complete the simplified chiller bid form (available at ashrae.org/ CHWSDL) included with this SDL with project design data. In most cases, the designer would predetermine the size of each chiller, but that is not mandatory. For instance, pricing for a two-chiller plant could be solicited for both equally sized chillers and for 1/3- and 2/3-sized chillers.

202

Chapter 6 Chiller Procurement 2. Send the simplified bid form to vendors for pricing and performance data. 3. Select chillers based on first costs and a subjective judgment of energy costs based on AHRI rated data. 4. Hard-spec the chiller per Procurement Step #9: Chiller Specification above.

Example An office building chiller plant is composed of two 300 ton chillers serving a primary variable-flow distribution system. The simplified chiller bid form was completed by four chiller vendors with a total of seven options. Results are summarized in Table 6-3. From these options, by inspection based on first cost, energy efficiency, historical maintenance, acoustics, and experience with similar chillers on other projects, Chillers 2, 3, and 6 were deemed the best three options. Each has strengths and weaknesses: •

•

•

Option 2 is the most expensive and loudest but also has the highest published efficiency and, because it has a screw compressor, can have lower maintenance costs. Option 3 is the least expensive despite having magnetic bearings. Magnetic bearings reduce noise and maintenance costs and provide good low-load performance and stability. This chiller also needs no head pressure control because it can operate with negative lift. While its published efficiency is not as good as the other two options, it has the lowest minimum CHW flow rate (35%), which will result in lower CHW pump energy. Option 6 has good efficiency but has higher pressure drop (higher pump energy) and is physically much larger than the other two options. It also uses R-123, a refrigerant scheduled to be phased out of production during the chiller’s service life.

Ultimately, the design team chose Option 3, in part because it had lower first costs. This example demonstrates the following: •

•

This simplified approach is obviously not as comprehensive as the recommended procurement procedure, but it should be more effective than arbitrary chiller selection. Just by inspection, for example, it may be clear that the superefficient chiller the design team was considering (e.g., Option 7) is simply too expensive and not likely to be cost-effective. It also allows the design team to be able to design around a specific chiller. For instance, premium-efficiency chillers are often physically larger; if the premium-efficiency option is selected, the mechanical room can be sized accordingly. Conversely, if a smaller chiller is selected, the mechanical room can be reduced in size.

Fundamentals of Design and Control of Central Chilled-Water Plants I-P Table 6-3 Option

1

Price for one chiller (including freight)

203

Simplified Chiller Bid Summary 2

$118,750 $116,850

3

4

5

6

7

$99,595

$81,222

$108,850

$105,270

$154,760

Compressor type

Tri-rotor screw

Tri-rotor screw

Magnetic bearing centrifugal

Twin screw

Refrigerant type

R-134a

R-134a

R-134a

R-134a

R-123

R-123

R-123

Operating weight, lb

16,554

15,988

11,275

13,386

16,708

16,394

19,831

Refrigerant weight, lb

950

840

558

574

500

500

750

Delivery lead time, weeks

9

9

10

9

7

7

11

VFD, Y/N

Y

Y

Y

Y

Y

Y

Y

% of design capacity below which chiller cycles

10%

10%

10%

10%

10%

10%

6%

Minimum lift (condenser return temperature [CWRT] – chilled-water supply temperature [CHWST]) at minimum load, °F

12.6

12.6

–6.0

15.0

13.0

13.0

13.0

CHW pressure drop, ft

8.5

22.5

14.4

19.7

4.97

16.48

7.72

Minimum CHW flow rate, %

43%

29%

35%

53%

46%

31%

36%

Condenser water (CW) pressure drop, ft

7.6

6.8

11.8

9.38

16.91

16.91

17.03

Sound power dB(A)

82

82

74

88

77

77

70

NPLV

0.331

0.335

0.362

0.415

0.358

0.348

0.341

kW/ton at 25% load

0.346

0.347

0.364

0.375

0.373

0.356

0.342

Design kW/ton

0.537

0.542

0.525

0.609

0.552

0.554

0.487

•

Centrifugal Centrifugal

Ceramic bearing centrifugal

Even a subjective review like this can help make a better decision than the conventional arbitrary process. For instance, in this example, the three finalists were very different chillers (one screw, one magnetic bearing centrifugal, and one R-123 centrifugal). Using a conventional arbitrary selection, one type of chiller would have been specified, typically eliminating the other options.

204


•

•

Prior to bid, the design team thought magnetic bearing chillers would be prohibitively expensive. That was not the case. Again, using an arbitrary selection process, the chiller that was ultimately selected might not have even been considered. The bid process is fairly painless for the vendors because all of the data in the bid form are available from their standard selection software reports. Efficiency ratings are based on AHRI Standard 550/590 and not adjusted for zero tolerance, which greatly simplifies the completion of the form. There is little value to getting zero tolerance efficiency data because energy costs are based on judgment, not rigorous analysis. The lack of objective energy and life-cycle cost calculations can make the final selection more difficult, as shown in the example where it was not easy to make a selection among the three finalists. The more comprehensive procurement procedure is therefore the better option when possible within project time and budget constraints.

References AHRI. 2015. AHRI Standard 550/590-2015, Performance rating of water chilling packages using the vapor compression cycle. Arlington, VA: AirConditioning, Heating, and Refrigeration Institute. 2015. ASHRAE. Chapter 37, ASHRAE Handbook—HVAC applications. Atlanta: ASHRAE. ASHRAE. 2016. ANSI/ASHRAE/IES Standard 90.1-2016, Energy standard for buildings except low-rise residential buildings. Atlanta: ASHRAE.


205


The benefits of a performance-based chiller bid approach include which of the following: a. It can minimize life-cycle costs for the owner. b. It solidifies selections in the design phase, allowing the designer to accurately lay out the plant and avoid design rework by the contractor during the construction phase. c. The chiller selection process becomes based on objective criteria rather than subjective criteria that contractors may use when including chillers in their project bids. d. All of the above.

6-2

Which of the following parameters is not appropriate to include in a performancebased bid approach: a. The anticipated cooling load profile for the plant. b. Total required plant capacity. c. Chiller efficiency metrics. d. Sound power limits.

6-3

The following three selections are provided by a chiller manufacturer.

Chiller 1

Chiller 2

Chiller 3

200 ton centrifugal Two-pass evaporator Open-drive compressor CHW pressure drop: 5 ft Full-load kW/ton: 0.5 Native BACNet Variable speed

200 ton screw Two-pass evaporator Hermetic compressor CHW pressure drop: 10 ft Full-load kW/ton: 0.6 Native BACNet Variable speed

200 ton centrifugal Three-pass evaporator Hermetic compressor CHW pressure drop: 15 ft Full-load kW/ton: 0.55 BACNet gateway required Variable speed

Which option is likely to have the highest first cost, not including the cost of the chiller itself? a. Chiller 1 b. Chiller 2 c. Chiller 3 d. Indeterminate based on the provided data 6-4

When estimating annual plant utility costs with simulation modeling a. Using a flat effective annual rate that accounts for both electricity and demand is usually sufficiently accurate. b. Annual escalation of utility costs should be ignored because of the unknown future variation of utility prices.

206

Fundamentals of Design and Control of Central Chilled-Water Plants I-P c. Using real rate schedules that account for both electricity and demand independently should be used. d. Annual escalation of utility costs should be estimated based on forecasts provided by DOE/EIA. 6-5

Using zero tolerance chiller performance data in bid forms a. Greatly increases the time needed to complete the forms by vendors. b. Better ensures the best chillers are selected, particularly for plants expected to operate many hours at low load. c. Is not recommended when using the simplified procurement procedure. d. All of the above.

Controls


Introduction This chapter begins with a discussion of some general factors that should be considered when designing and installing a CHW plant BAS and concludes with suggested control sequences. It is assumed that the BAS is a DDC system. The two terms (BAS and DDC system) are used interchangeably in this chapter. To maintain stable and effective plant operation, designers must ensure that the control systems are of high quality and are sufficiently accurate and reliable, balanced with cost considerations. Also, the CHW plant’s control system must have sufficient programming and data manipulation and trending capabilities. Most importantly, the control system must be programmed with control sequences that optimize, or nearly optimize, energy performance while providing reliable operation. The following topics related to controls and instrumentation in CHW plants are covered: • • • • • •

Types of sensors available for energy monitoring and control Styles of and selection criteria for control valves Controller requirements and interfacing issues Performance monitoring Types and configuration of local instrumentation Control sequences for CHW plants

Industrial Versus Commercial Controls Many older CHW plants were designed using industrial control systems, including programmable logic controllers (PLCs) and industrial sensors and instrumentation. That was justified by the lack of robust and reliable commercial DDC systems in the 1980s and early 1990s. But almost all modern commercial DDC systems have the capability to optimally control and monitor CHW plants of virtually any size and complexity. Commercial systems are

208

Chapter 7 Controls usually easier to program and maintain (they are simpler and more operators are familiar with them) and always cost less to install, typically half the cost or less. But industrial controls do have one key advantage: the ability to integrate redundant controllers to allow seamless operation should a controller fail. It is possible to design commercial control systems with redundant controls, but it is complex and difficult to implement, in part because it is not common practice. The complexity of programming multiple controllers to provide backup operation in case of controller failure can ironically lead to unreliable performance. Typically with commercial control systems, controller failure is addressed solely by manual intervention—a building operator must be made aware of the failure (in itself a problem if alarms and alarm broadcasting are also provided by the failed controller) and have time to manually control pumps, chillers, valves, etc., before the system loses control of the cooling load. Manual intervention is a reasonable approach for noncritical plants such as those serving offices and other commercial occupancies. However, it may not be reasonable for plants serving critical data centers where loads are so high that failure of the chiller plant for even minutes may cause loss of temperature control and computer equipment failure. Industrial control systems may be appropriate for plants serving these and other very critical applications. Commercial controls are generally preferred otherwise.

Choosing Control and Monitor Points If the control and monitor points are not carefully chosen, the system may suffer from either under- or overinstrumentation. An underinstrumented system may be difficult to control optimally, whereas an overinstrumented system may be confusing to the operations staff, expensive, and difficult to maintain. The selection of control and monitor points should be based on a careful analysis of the chiller plant’s control and operating requirements. Each point must meet at least one of the following criteria: 1. Essential: The point is necessary for effective control of the chiller plant as required by the sequence of operations (SOO) established for the plant. 2. Desirable: The point provides information such as status and alarms that can be used by plant operators to ensure that the plant is operating properly or to indicate a problem has occurred or may soon occur. 3. Optional: The point may be used to gather accounting or administrative information such as energy use, efficiency, or runtime, if such monitoring is desired by the plant operator or owner. Because each plant has individual operating requirements for control logic, staffing, accounting, and operations and maintenance, there is no single standard for determining exactly what instrumentation is required for any plant. The decisions must be made by the designer with owner input and concurrence.


209

The types and selection of sensors and control devices are discussed first. This is followed by an example of essential, desirable, and optional points.

Sensors Sensors are used in CHW plants for a variety of measurements, including: •

Temperature

•

Humidity

•

Liquid flow

•

Pressure

•

Electric current

•

Electric power

•

Gas flow

The sections that follow discuss the end-to-end accuracy requirements and the types of sensors available for each type of measurement, their measuring characteristics, and special factors to consider. Installation and calibration information is also provided.

End-to-End Accuracy Both accuracy (the ability to measure an actual value) and resolution (the ability to sense changes in the value) need to be considered together in selecting sensors. If accuracy is considered without looking at resolution, or vice versa, the measurement objective may not be satisfied. The inherent accuracy of the sensing device is the primary factor in determining the overall accuracy of the measurement. Other factors are •

the means of transferring the signal from the measurement device to the control system and

•

the signal span and number of bits used in the control system’s analog-todigital (A/D) converter.

Whenever possible, it is useful to specify end-to-end measurement property requirements. Typical and suggested end-to-end accuracies for common chiller plant control points are shown in Table 7-1. These accuracies are readily achievable with common commercial DDC systems and are usually all that is required for acceptable plant performance. Increased accuracy may be achieved but usually at added cost. Additional discussion on accuracy requirements is included in specific sensor sections.

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Chapter 7 Controls Table 7-1

Typical End-Use Accuracy

Measured Variable

Reported Accuracy

Outdoor air dry-bulb temperature

±1°F

CHW and condenser water temperatures at central plant mains

±0.2°F

CHW and condenser water temperatures

±0.5ºF

elsewhere Water T (supply to return)

±0.15ºF

Relative humidity

±5% rh

Water flow

±1% of full scale

Water pressure

±2% of full scale

Electrical power

1% of reading, 3 kHz response for VFD-driven equipment

Temperature Sensors Types of Sensors—Temperature Sensors The most common temperature sensing devices used in hydronic applications are • • •

Thermistors Resistance temperature detectors (RTDs) Integrated circuit (solid-state) temperature sensors

These all use materials whose resistance or impedance changes with temperature. Thermistor signals are nonlinear (Figure 7-1) and thus must be converted to a temperature reading from the resistance reading. Modern DDC systems include the conversion of standard thermistors either in firmware or software so that thermistors may be directly connected to them without a transmitter. The connection is then inexpensive and accurate because the preset scaling ranges on these systems can automatically adjust for the nonlinear signal. Thermistors generally cannot be calibrated on the device itself, although some DDC manufacturers allow a slope/intercept calibration within the point configuration software. RTDs, on the other hand, have linear resistances to temperature signals that do not require any conversion, but many RTDs have very low electrical resistance (e.g., 100 to 1000 ohms) such that the voltage drop in the wiring from the sensor to the DDC input can skew the signal. For these RTDs, a transmitter is required. The output of the transmitter is a linear current signal (e.g., 4–20 mA) that is independent of wiring length. The transmitters must be calibrated and limit the range of the output to a fixed temperature range.


Figure 7-1

211

Typical thermistor resistance versus temperature curve.

Integrated circuit temperature sensors are not commonly used in HVAC applications except with Btu meters. They are not very accurate (~±1.0°F) but they are extremely repeatable and linear. They also do not require calibration. Hence they are excellent for differential temperature measurement once the two sensors are matched and calibrated at the factory.

Issues and Recommendations—Temperature Sensors Thermistors and RTDs each have advantages and disadvantages, but when properly applied either may be used to measure temperature. The following issues may impact the choice of temperature sensor.

Cost RTDs generally are more expensive than thermistors, particularly those for which transmitters are required. Some high-resistance thin-film RTDs can have lower costs because transmitters are not required, but these types of RTDs are also much less accurate.

Accuracy Depending on the material (platinum is most common) and construction, RTDs can be purchased with a wide range of accuracies from very accurate Class A RTDs (as low as ±0.02°F) to modestly accurate thin film RTDs (typically ±1°F) over typical chiller plant temperature ranges. Matched-pair (sensor/transmitter) sensors are available to improve overall accuracy; they use the tight tolerance of Class A RTDs and a NIST traceable factory calibration procedure. Thermistors typically are offered with at least two options: standard and extra-precision. Standard thermistors have an accuracy of ±0.4°F over standard temperature ranges while extra-precision thermistors have an accuracy of about ±0.2°F.

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Chapter 7 Controls

Long-Term Stability RTDs are more stable in the long term than thermistors. However, the external transmitter circuits on most RTDs can cause readings to drift, so routine calibration is required. In the past, thermistors were known for drifting over short time periods, but modern thermistors are very stable. Most manufacturers guarantee drift of no more than 0.05°F over a five- or ten-year span. Because of their lower first costs, similar or higher accuracy, and reduced need for routine calibration, thermistors are usually preferred over RTDs for most temperature sensors in CHW plants. The exception may be those used for supply and return temperatures that are also used for plant load calculations. To achieve the ±0.15°F accuracy suggested in Table 7-1 for T measurement, each sensor must have an error of no more than ±0.1°F. This requires a highprecision, matched-pair RTD transmitter. Alternatively, a Btu meter can be used in lieu of these sensors and a standalone flowmeter; Btu meters are discussed in a later section.

Sensor Installation—Temperature Sensors Indirect well temperature sensors are recommended for measuring water temperature, so that the sensor may be removed from the well for calibration or testing. The wells should be brass or stainless steel. Recommended guidelines for installing well temperature sensors are as follows: • • •

• •

Place the temperature sensor in a well that penetrates the pipe by the lesser of half the pipe diameter or 4 in. Install the sensor in the well with a thermal-conducting grease or mastic. Use a closed-cell insulation patch that is integrated into the pipe insulation system to isolate the top of the well from ambient conditions but allows easy access to the sensor. Locate wells far enough downstream from regions of thermal stratification or mixing so that the fluid’s temperature is uniform at the well. For field calibration and testing, a test plug should be installed adjacent to the sensor. This will allow the sensor to be tested without removal using a handheld sensor connected to the test plug.

Calibration—Temperature Sensors There are two approaches to calibration; each has merit depending on how critical the measurement applications are.

Factory Calibrated For noncritical measurement applications, it is usually satisfactory to rely on the manufacturer’s temperature tolerance limit for the sensor supplied. Factory-established temperature tolerance limits generally range about ±0.5°F for RTD and thermistor sensors. Some manufacturers use the term


213

pre-calibrated to refer to the tolerance limit. However, this pre-calibration does not account for all the factors involved in the end-to-end measurement characteristics of the sensor (see End-to-End Accuracy above).

Field Calibrated It is recommended for more critical measurement applications (such as supply and return water temperatures used for plant load calculations) that each temperature sensor be field calibrated as part of the commissioning and start-up process (see Chapter 8 for information on commissioning). Because of the limited rangeability of most chiller plant temperature-sensing requirements, a single-point calibration is often adequate. An end-to-end field calibration of all critical devices can enhance the overall accuracy of the system. However, it is imperative that the temperature-sensing device used for calibration be at least twice as accurate as the respective field device. For example, if the temperature sensor is ±0.4°F accurate, test equipment must be ±0.2°F accurate over the same range. For highly precise sensors such as those used for CHW supply/return temperatures, a dry-well bath is typically required because handheld sensors with ±0.05°F accuracy are rare.

Liquid Flow Sensors Types of Sensors—Liquid Flow Sensors Fluid flow measurement can be a difficult and costly instrumentation item for a chiller plant. However, it can also be a critical measurement depending on control sequences and performance monitoring goals. Table 7-2 compares various types of common liquid flow sensing devices. Table 7-2

Comparison of Liquid Flow Sensing Devices

Flowmeter Type

Configuration

Range of Flow (Turndown Ratio)

Relative First Cost

Orifice

In-line

Low (5:1)

Low

±5%

Medium

Insertion turbine

Insertion

Medium (30:1)

Medium

±2%

Medium

Vortex shedding

Insertion

Medium (30:1)

High

±2%

Medium/low

Single-point magnetic

Insertion

High (50:1)

High

± 1%

Low

Transit time ultrasonic

External

High (100:1)

High

±0.5%

Low

±0.5%

Lowest

Full-bore magnetic

In-line

High up to Highest (1000:1) 12 in. pipe; Very high > 12 in.

Measurement Maintenance Accuracy Costs

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Chapter 7 Controls

Issues and Recommendations—Liquid Flow Sensors Generally, the most important flow measuring characteristics are range and accuracy. The design engineer must fully understand the expected error at various flow levels to be certain the device chosen will meet all requirements. For CHW and condenser water applications, full-bore magnetic flowmeters are often the best choice. Where costs are prohibitive on large pipes, single-point insertion magnetic flowmeters can be used. Full-bore magnetic meters are the most accurate, require the least maintenance, and are the least susceptible to errors from variations in flow profile (they require the least amount of space upstream and downstream of the meter) and the presence of particulates or air in the pipe. They also have no moving parts in the water stream that can get fouled, a critical feature with respect to open condenser water flow measurement. Ultrasonic meters share many of the positive qualities of magnetic flowmeters, but they can provide erroneous readings if air or other particles pass through the meter, such as on open-circuit condenser water systems. They are also more prone to installation errors; accuracy significantly depends on alignment of sensors and thickness of piping. However, they can be a good alternative to single-point magnetic flowmeters in large CHW piping where full-bore magnetic flowmeters are cost prohibitive. Turbine flowmeters are often used to reduce costs where budgets demand, but they should not be used on open condenser systems where fouling can render the sensor ineffective in a very short period of time.

Installation—Liquid Flow Sensors The manufacturer’s requirements for placement and installation must be carefully observed to ensure an accurate flow measurement. Because flow measurements often require specific upstream and downstream lengths of undisturbed straight piping, it may be necessary to pay special attention to the piping layout during plant design and construction. Full-bore magnetic flowmeters are the most forgiving in this regard; they will read flow accurately unless turbulence is so bad that flow reversal occurs within the bore. On the other hand, single-point magnetic and turbine insertion flowmeters can require as much as 15 to 30 pipe diameters upstream of the meter, which is difficult to achieve in most plants. Where flowmeters are intended to measure flow that may be bi-directional (e.g., installations in a common leg or for a TES tank), sensors must be capable of reading bi-directionally as well. All of the sensors in Table 7-2 are either inherently bi-directional or available in a bi-directional option, except for vortex shedding meters.

Calibration—Liquid Flow Sensors Calibrating flowmeters in the field can be very difficult and, more often than not, is impractical or impossible. Calibration of any device requires field measurement of the measured variable with a device that is substantially


215

more accurate than the device being calibrated. With flow measurements, it is often hard to find an acceptable location for the flowmeter, one with adequate length of straight piping both up- and downstream of the device. It is therefore very seldom possible to locate an additional, temporary flow sensor in the system in order to calibrate the installed sensor. Strap-on ultrasonic meters are often considered, but they can be difficult to properly use and, at least with current technology, are not usually more accurate than the installed flowmeter and therefore not appropriate for use in calibration. Use of pump curves, DP across known devices (such as chillers), or CBVs at pumps or coils also are not accurate enough for calibration. However, all or some of these can be used as a “reality check” to verify that the flowmeter is reading in the proper range, and often that is all that can be done in the field. Therefore it is important that flowmeters be accurately factory calibrated and preferably NIST traceable.

Btu Meters Types of Sensors—Btu Meters Measuring CHW load (in Btu/h, generically abbreviated Btu) is often required for chiller staging, chiller plant performance monitoring, or submetering CHW usage, such as at each building of a college campus. The load calculation can be performed by the BAS from flow and temperature sensors, or it can be done by a device called a Btu meter. The Btu meter generally is configured to send calculated Btu data, optionally along with individual temperature and flow measurement data, to the BAS or other data collection system for monitoring. It may also have a display for manual reading of internally stored energy usage data. Btu meters are generally designed to work with any flowmeter and temperature sensors, but more commonly the Btu meter manufacturer provides the meter, flowmeter, and temperature sensors as a package as shown in Figure 7-2. With this style, which is used for larger piping, the flowmeter and temperature sensors are all field mounted; with some smaller Btu meters, the flowmeter and one temperature sensor are built into the main meter housing. The temperature sensors are provided with the Btu meter so that they can be factory matched and calibrated for improved accuracy. The flowmeter can be any type depending on the desired accuracy (see the previous Liquid Flow Sensor section). The output of the Btu meter can be a pulse or analog output connected to a BAS or other data collection system. Modern Btu meters also include the ability to directly connect to common control networks such as BACnet/MSTP, BACnet/IP, Modbus/EIA485, LonWorks, and various proprietary networks.

Issues and Recommendations—Btu Meters The main advantage of the Btu meter is that temperature sensors are factory matched to minimize temperature difference calculation error. However, they generally cost more than using individual sensors connected to the BAS. Nevertheless, Btu meters are recommended due to their improved temperature

216

Chapter 7 Controls

Figure 7-2

Typical thermal energy meter.

measurement accuracy and stability and ease of data collection, particularly if energy is being metered for revenue purposes (e.g., allocating costs of CHW or hot-water usage per building).

Installation and Calibration—Btu Meters See the preceeding Temperature Sensors and Liquid Flow Sensors sections above for installation and calibration information.

Humidity Sensors Types of Sensors—Humidity Sensors The most common humidity sensors are either capacitance or resistive type measuring relative humidity. Common nominal accuracies are ±1%, 3%, or 5%


217

including hysteresis, linearity, and repeatability. These are manufacturer-listed accuracies. Actual accuracy will vary depending on the quality of the sensor and how well and how frequently the sensor has been calibrated. Temperature sensors tend to be very stable and remain accurate for many years (Edwards 1983; Lawton and Patterson 2002). Humidity sensors, on the other hand, are notorious for being difficult to maintain in calibration. A test of commercial humidity sensors (NBCIP 2004, 2005) showed that few of the sensors met manufacturers’ claimed accuracy levels out of the box and were even worse in real applications. Figure 7-3 and Figure 7-4 show the results of the NBCIP one year in situ tests of two brands of humidity sensors among the six brands tested. There were two sensors tested for each brand, represented by the dark and light gray dots. Figure 7-3 shows sensor data from the best performing manufacturer in the study, although even these top quality sensors did not meet the manufacturer’s claim of ±3% accuracy. Figure 7-4 shows sensor data from the worst performing manufacturer; both sensors generated almost random humidity readings.

Issues and Recommendations—Humidity Sensors In chiller plants, humidity sensing is generally not required unless the system has a WSE (see sequences below in the section Water-Side Economizer [WSE] Control), in which case outdoor wet-bulb temperature is needed to enable the economizer. Wet-bulb temperature is not typically measured directly; rather, relative humidity and temperature are measured and wet-bulb temperature is calculated from these readings in firmware or software. Relative humidity sensors are subject to error due to sensing

Figure 7-3

Iowa Energy Center NBCIP study—best humidity sensor.

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Figure 7-4

Iowa Energy Center NBCIP study—one of the worst humidity sensors.

technology, hysteresis, and drift. They are much more likely to drift out of calibration than temperature sensors and are often out of calibration when first installed. The humidity sensor should be specified for ±3% rh or lower. Because only one humidity sensor is needed for CHW plants, and because its accuracy will have a significant impact on WSE performance, a very high-quality sensor from a high-quality manufacturer should be used. Humidity sensors offered by major control system manufacturers, generally devices manufactured by original equipment manufacturer (OEM) mass production sensor companies, often are not of sufficient high quality despite their specifications; sensors from companies that specialize in humidity measurements are recommended instead. The NBCIP studies (2004, 2005) in the References should be consulted.

Installation—Humidity Sensors Outdoor air temperature and humidity must be measured in a location well protected from direct sunlight and also from exhaust outlets from building fans and exhaust from cooling tower fans, all of which can skew the readings. Where no such location can be found (as is typical), a fan-aspirated assembly can be used and located on the roof far from exhaust plumes. The fan assembly obviates the need to avoid direct solar exposure.


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Calibration—Humidity Sensors The outdoor air humidity sensor must be field tested initially (see Chapter 8 for information on commissioning) and at intervals recommended by the manufacturer (but no less than semi-annually) using a portable humidity-sensing device. Because humidity sensors are so prone to drift, it is recommended that DDC systems with Internet connectivity test sensor calibration by connecting to the closest National Oceanic and Atmospheric Administration (NOAA) website and comparing their dew-point temperature reading to that calculated at the site using outdoor air temperature and relative humidity. The dew point (absolute humidity) should be similar between the building and NOAA sites if they are not too far apart geographically; if they are not, an alarm can be generated alerting building engineers to possible sensor calibration problems.

Pressure Sensors Types of Sensors—Pressure Sensors Pressure is always measured as a DP, either the difference between the pressures of two fluids or the difference in pressure between a fluid and a reference pressure. When the reference pressure is atmospheric pressure, the sensor is referred to as a gage pressure sensor. The most common means of sensing pressure for fluid conditions are fastresponse capacitance type. Standard commercial-grade sensors offer excellent accuracy, usually 1% or less of the specified pressure range.

Issues and Recommendations—Pressure Sensors The following are common pressure sensor applications in CHW plants: •

Control of variable-speed pumps: The sensors should be located at the extreme ends of the system as discussed below. Pump speed is modulated to maintain DP at set point.

•

Control of flow: DP across a device of fixed geometry can be correlated to flow using known flow versus pressure drop curves. As discussed in Chapter 4, this is an alternative way to control minimum flow through chillers, rather than using a flowmeter. However, a flowmeter, which can also more accurately be used for plant load calculation, is a more common and usually preferred option.

•

Monitoring of system pressure: A gage pressure sensor can be installed at the CHW system expansion tank connection to sense a drop in pressure that can occur when there is a leak, generating an alarm accordingly. (Note that makeup water connections to CHW systems should be shut off after air is removed from the system to avoid exacerbating the damage from a leak.)

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Chapter 7 Controls Pressure need not be measured extremely accurately in these applications, so standard commercial sensors may be used without paying a premium for premium accuracy. Sensors must be purchased for the pressure range required by the application. As a general rule, the pressure range should be about two times the expected set point.

Installation—Pressure Sensors The sensor should be mounted in a location where it is accessible for maintenance and that is not subject to physical harm, such as from vibration or water damage. Often this means that the sensor is located in a temperature control panel and is piped to the remote piping taps rather than located remotely where it may not be readily accessible. DP transmitters are often installed in systems with pressures much higher than the DP being monitored. During installation, start-up, or shut-down, the pressure differential may exceed the transmitter DP rating, resulting in severe damage to the transmitter. A three-valve bypass assembly (Figure 7-5) should be used to minimize this possibility. The valve located in parallel with the DP transmitter is left opened until the sensor is ready for use, ensuring the DP across it is minimal. Test ports should be provided to allow a handheld gage to be used for field calibration. A five-valve assembly is similar but also includes two bleed valves for air removal; these additional ports can be used in lieu of test ports for field calibration.

Calibration—Pressure Sensors Generally, factory calibration of pressure sensors is adequate for most pressure measurement applications. It is recommended that each pressure-sensing device be field tested during start-up and balancing to confirm its accuracy at both zero pressure and at least one typical pressure condition.

Figure 7-5

Three-valve DP transmitter assembly with test ports.


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Flow Switches and Indicators Types of Sensors—Flow Switches and Indicators Historically, the most common flow indicating switch was a paddle switch. The switch contact is closed when fluid flow deflects the paddle. Another common sensor is the DP switch, which is activated when DP, such as across the evaporator, condenser, or pump, exceeds a set point. A relatively new flow indicator uses thermal or calorimetric principles. The sensor tip is heated to a few degrees above the liquid temperature. As the liquid flows across the tip, it is cooled by the liquid proportional to liquid velocity. The indicated liquid velocity is then compared to the set point programmed into the device and the contact is closed when the set point is exceeded.

Issues and Recommendations—Flow Switches and Indicators Pump status in the past was commonly determined using DP switches mounted across the pump. But the switch would provide a false flow indication if a valve in the system shut off flow—high DP would still exist across the switch. They are also expensive to install. Less expensive and more reliable status indicators for pumps are current switches for fixed-speed pumps and VFD status for variable-speed pumps, discussed in the section Issues and Recommendations—Electric Current Sensors. Chiller manufacturers often require flow indicators across evaporators and, to a lesser extent, condensers. Evaporator flow switches used to be essential because a sudden loss in flow could cause significant damage, such as freezing of the evaporator if the chiller was not shut off immediately upon loss of flow. Some manufacturers no longer require evaporator flow switches, relying instead on robust chiller controls to sense flow loss by the sudden drop in supply water temperature. Even more manufacturers have eliminated the requirement for condenser water flow indication, instead relying on high head pressure alarms to disable the chiller upon loss of flow. The one disadvantage of this approach is that loss-of-flow indication alarms are typically automatic reset alarms should flow resume; by contrast, high head pressure alarms sometimes require manual reset. This can often be avoided by the BAS simply not enabling chillers when flow indication at pumps or isolation valve positions or end switches indicate flow is not available. Flow indication switches should be avoided unless required by the chiller manufacturer because they can be a source of false trips, particularly for paddletype switches, which are sensitive to physical damage, dirt, and corrosion. DP switches are more reliable, but they can give false negative flow indication with variable-flow systems, particularly on the condenser water side on systems that vary flow for head pressure control (discussed in more detail below in the section Control Schematics). Calorimetric flow switches have no moving parts, are resistant to fouling, and can be set to very low-flow set points and are thus recommended, particularly on the condenser side.

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Chapter 7 Controls Flow indicators are not required by some manufacturers, are factory installed by others, and are required to be field installed by a few. The latter can be a hidden cost to the controls and piping contractors if the specified chiller had factory-installed flow indicators or none at all. To level the playing field in chiller specifications, flow indicators should be required to be factory installed or the chiller manufacturer must carry the cost of their installation.

Installation—Flow Switches and Indicators Flow switches, depending on the type, should be located well away from elbows or other sources of turbulence that may cause them to flutter or be damaged.

Calibration—Flow Switches and Indicators Flow indicators do not need to be calibrated per se, but they do have a set point that must be adjusted. On variable-flow systems, the set point must be adjusted so that flow is indicated even at the minimum flow rate.

Electric Current Sensors Types of Sensors—Electric Current Sensors The universal means of sensing AC current is the current transformer (CT). As current passes through this device, a small voltage is generated proportional to the current that is being measured. There are both digital and analog versions of electric current sensors. The digital sensor (usually called a current switch) provides a binary signal (contact closure) as long as the current is above a fixed or adjustable set point. The analog sensor (usually called a current sensor or current transducer) provides an analog signal (usually 0–5 Vdc or 4–20 mA) that can be scaled to read the current draw. Current sensing alone is not very accurate for power measurement because it does not include the impact of simultaneous voltage and phase offsets. For power measurements, a true power meter should be used as discussed in the Electric Power Meters section.

Issues and Recommendations—Electric Current Sensors Adjustable set point current switches are very commonly used for on/off status of equipment. They are almost always better status indicators than DP sensors and flow switches for several reasons: •

Current switches are less expensive due to substantially lower costs for installation and wiring. DP and flow switches require installation into the piping and are usually a long way from the DDC panel. Current switches are more easily installed, particularly those with split core CTs, and are mounted in the starter panel, which is usually close to the DDC panel and also must be wired to anyway for on/off control of the motor starter. (For this reason, current switches are available packaged into the same enclosure with a pilot relay for on/off control of the starter.)


• •

•

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A DP switch for a pump may indicate pressure when in fact there is no flow. This can happen when a valve is closed but the pump is still on. The current switch set point can be set to a current higher than the power draw when the pump is deadheaded (no flow). DP switches cannot be used for cooling tower fan status because they do not generate enough DP to make a DP switch work reliably. The current switch set point can be set to a current higher than the motor’s no-load current so that it still can be a reliable indicator of a pump coupling or cooling tower fan belt failure. Current switches are solid-state devices with no moving parts or diaphragms, and are therefore much more reliable than DP switches or flow switches. They also require no maintenance and last longer.

Because of these advantages, current switches are recommended for constantspeed motor status. For variable-speed motors, current switches are usually not the best choice. For some applications such as cooling tower fans, current switches can indicate false failures due to very low current draw at low speed. However, a less expensive and more reliable status indicator for variable-speed motors is simply to use the status point that comes standard with the VFD. Like a current switch, it can be programmed to indicate failure at no-load power (broken coupling or belt) or deadhead power (closed valve).

Installation—Electric Current Sensors A current switch used for status of constant-speed motors should be located in the motor starter, mounted so that it is accessible and does not block access to other devices. In three-phase applications, a CT is mounted only on one leg of the power. The motor starter provides single-phase protection and will automatically shut down the motor if one phase is lost.

Calibration—Electric Current Sensors Factory calibration of current switches is adequate. However, as noted above, current switch set points must be field adjusted to indicate failure at noload and deadhead currents.

Electric Power Meters Types of Sensors—Electric Power Meters Two major types of meters are often used for power monitoring. The kW demand sensor provides an analog output (usually 0-5 Vdc or 4-20 ma) that indicates the instantaneous rate of electricity use. The kWh consumption sensor provides a pulse signal that indicates the number of kilowatt hours of electricity that have flowed since the last pulse. Both types of sensors have the

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Chapter 7 Controls same components. For three-phase applications, these sensors include CTs and voltage measurement taps for each leg. Modern power meters often have an option to transmit monitored information over a network such as BACnet/IP. Networked meters provide both kW and kWh information and commonly also provide power factor, kVAR, VA, and harmonic information. These data may be very useful in many facilities, especially if power quality is a concern.

Issues and Recommendations—Electric Power Meters Most kW and kWh meters provide better than 2% accuracy, which is suitable for verifying the plant’s energy use. However, in applications that involve monitoring the power of motors that are operated by VFDs or other wave distorting equipment, accuracy may be reduced unless the power sensor provides true root mean square (RMS) power sensing. Although there is no industry standard for true RMS sensing, there is agreement in the community that a minimum sampling or response rate of 3 kHz is required to get accurate measurement of nonlinear loads like VFDs. In most CHW plants where power monitoring is for performance monitoring (as opposed to monitoring for revenue purposes), there is no need for fieldinstalled power meters because almost every motor has a VFD, and VFDs include power monitoring inherently. The BAS simply needs to connect to the VFD (or chiller controller) network interface or hardwire to the power output point. For fixed-speed condenser water pumps, power meters are usually not needed because power draw is generally fairly constant, so power can be measured once with a handheld device during commissioning and calculated based on this measurement and pump status. While energy monitoring using VFD onboard meters and motor status for fixed-speed pumps is generally adequate for performance monitoring, it is not accurate enough for revenue metering. VFD power meters have on the order of ±3% to 5% accuracy at higher currents and many are less accurate at low currents. They also measure power output to the motor, so they do not include the inefficiency of the VFD; VFD efficiency is very high at full load but falls off at low loads (see Chapter 3). If very accurate power monitoring is required, a true RMS revenue-grade meter of the equipment or the power service to the whole plant is required.

Installation—Electric Power Meters Power meters are generally located in the motor starter panels or electrical distribution panels. Care must also be taken to ensure that sufficient ventilation is provided so that the manufacturer’s temperature limits for the equipment are not exceeded.


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Calibration—Electric Power Meters Factory calibration of electric kW and kWh sensors is nearly always adequate. However, reality checks should always be done during commissioning to verify that power meter scaling factors have been correctly configured.

Gas Flow Sensors Types of Sensors—Gas Flow Sensors There are several types of gas flow sensors that can monitor the natural gasflow of absorption or gas engine-driven chillers. The diaphragm gas meter is most widely used (this is the type that most utilities install as site meters). Other gas flow sensors are rotary and turbine meters; these are generally used when the maximum gas flow requirements exceed the capacity of a diaphragm meter. Table 7-3 compares various options for gas flow-sensing devices.

Issues and Recommendations—Gas Flow Sensors Volumetric (diaphragm and rotary) meters are generally recommended for most chiller plant applications. An important factor in choosing a gas flow meter is range of the flows that the meter must measure. Another consideration is available gas pressure; some diaphragm meters have high-pressure drops, so the available gas pressure must be high enough to accommodate the meter and still provide enough pressure for the chiller. The metering application should always be discussed with the utility supplying the gas. It is sometimes necessary to monitor the pressure in order to improve the measurement accuracy. Temperature compensation options should always be selected. Where the gas is serving only the chiller plant, the best option is to use the utility’s gas meter with an auxiliary transmitter to provide data (usually a pulse signal) to the BAS. This is usually the most economical and most accurate choice. The auxiliary transmitter is often an option, so coordination with the utility is recommended. Table 7-3

Comparison of Gas Flow-Sensing Devices

Type of Gas Meter

Range of Flow, Standard ft3/h

First Cost

Rangeability (Turndown Ratio)

Diaphragm

Up to 5000

Low

100:1

Rotary

100 to 50,000

High

40:1

Turbine

1500 to 200,000

Medium

15:1

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Installation—Gas Flow Sensors Because diaphragm and rotary gas flowmeters measure volume, not velocity, their placement is far less critical than for turbine meters. For turbine meters, the manufacturer’s requirements for placement and installation must be carefully observed.

Calibration—Gas Flow Sensors Typically diaphragm and rotary gas meters are fully factory tested and calibrated and require no further calibration.

Control Valves Valve Types Ball Valves Modern ball valves designed for control applications are inexpensive, effective, and reliable in smaller chiller plant piping. Ball valves are now available in up to 6 in. pipe sizes. Ball valves are well suited for isolation valves because they can be ported for full pipe size (i.e., the opening in the ball valve is the same as the inside diameter of the pipe, reducing pressure drop). Ball valves are also well suited for modulating control because they act with an equal percentage characteristic (see the Valve Selection Criteria section that follows) when fully ported or in special porting configurations. They usually have lower first costs than the globe valves that have been traditionally used for modulating duty. However, ball valves must be specifically designed for control applications; standard ball valve designs are not adequate for the continuous movement required for modulating control duty and usually suffer seal failures in a short period of time.

Butterfly Valves Butterfly valves are the most popular large-diameter control valves in chiller plants. Like ball valves, butterfly valves make excellent isolation valves because they offer nearly full pipe bore when open and thus have low pressure drop. Butterfly valves also have valve characteristics similar to equal percentage valves when used in modulating valve applications. However, they have very low pressure drop and thus have low valve authority, usually making them inappropriate for use in two-way modulating duty at cooling coils.

Globe Valves Globe valves have lost almost all their market share to ball valves for small modulating duty control valves, but for many years they were still the most common control valves for large (>3 in.) modulating control valves. They have relatively high-pressure drops and provide good authority for improved con-


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trollability. They are thus the valve of choice (along with ball valves for smaller coils) in two-way valve, variable-flow systems. However, now that ball valves are available up to 6 in. in size, globe valves will further fall from favor. Where both ball and globe valves are available, ball valves are preferred because they are less expensive and more reliable and have higher shutoff pressures while still providing equal percentage characteristics.

Pressure-Independent Valves Generally, pressure-independent valves are not required in typical chiller plant configurations but may be of use in solving special problems or system features. See Chapter 4 for a discussion of the use of pressure-independent valves.

Valve Selection Criteria Valve Sizing and Flow Coefficient Valve sizing in two-position (on/off) applications is straightforward: the valve is simply the same size as the piping it is installed in. However, valve sizing in modulating applications is more difficult and a fairly controversial subject. The valve size is based on its full open pressure drop, which in turn determines the valve’s authority and the ability of the control system to function as desired and expected. It is probably intuitively clear that an oversized valve will not be able to control flow well. As an extreme example, imagine trying to pour a single glass of water using a giant sluice gate at the Hoover Dam. However, undersizing a valve increases the system pressure drop, which leads to higher pump costs and higher energy costs. These two considerations must be balanced when making valve selections. The size of a valve is determined by its pressure drop when it is at full open. The question then is: what pressure drop should be used? Unfortunately, there is no right answer to this question and there are differing opinions and rules of thumb expressed by controls experts and manufacturers (discussion of which is beyond the scope of this SDL). While there is disagreement about the exact value of the desired pressure drop among these authorities, there is general agreement that the control valve pressure drop, whatever it is, must be a substantial fraction of the overall system pressure drop in order for stable control to be possible. With the advent of more sophisticated control algorithms such as proportionalintegral-differential (PID) and fuzzy logic, some designers have questioned the need for high valve pressure drops. However, while a well-tuned controller can certainly compensate for some valve oversizing, there is clearly a point where no control algorithm will help. For instance, getting a single glass of water out of a sluice gate will be impossible no matter how clever the control algorithm may be. Oversizing will also result in the valve operating near close-off most of the time. This can increase noise from flow turbulence and may accelerate wear on the valve seats. Therefore, relaxing old rules of thumb on valve selection is not recommended.

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Chapter 7 Controls One old rule of thumb that has been used successfully in valve sizing at coils is to select valve full open pressure drop between a minimum of half the coil pressure drop and a maximum of 5 psi. Two-way valves should be selected to result in a pressure drop near the maximum while three-way valves should be selected for pressure drops near the minimum. See also ASHRAE Handbook—HVAC Systems and Equipment (2016b), Chapter 47, Valves; Fundamentals of HVAC Control Systems; and valve manufacturers’ selection guides for valve sizing guidance. Once the pressure drop is determined, the valve can be selected using a rating called the valve flow coefficient, Cv. The valve flow coefficient is defined as the number of gallons per minute of fluid that will flow through the valve at a pressure drop of one psi with the valve in its wide-open position, expressed mathematically as s C v = Q ------P where Q =

(7-1)

flow rate in gpm

s

=

specific gravity of the fluid (the ratio of the density of fluid to that of pure water at 60°F)

P

=

pressure drop in psi

Specific gravity for water below about 200°F is nearly equal to 1.0, so this variable need not be considered for most HVAC applications other than those using brines and other freeze-protection solutions. Valve coefficients, which are a function primarily of valve size but also of the design of the valve body and plug, can be found in manufacturers’ catalogs.

Valve Characteristics For a more complete discussion of valve characteristics, refer to the ASHRAE Handbook—HVAC Systems and Equipment (2016b), Chapter 47, Valves; and Fundamentals of HVAC Control Systems. As a summary of the typical recommendations, see the following: Select equal percentage characteristics for • •

Any two-way valve Hot-water three-way valves Select linear characteristics for

•

CHW and condenser water three-way valves

Not all manufacturers offer choices of valve characteristics, so it may not be possible to always follow these recommendations.


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Valve Shut Off Modern valve seat materials provide zero leakage for many valves and pressure conditions. However, substantial variances in close-off rating exist between valves. For both ball and butterfly valves, the fluid pressure does not affect the closing force required, but fluid pressure is a factor with globe valves. With globe valves, it is important to carefully consider the valve operating conditions to ensure that the valve has adequate close-off capability. Many valves will have two close-off ratings, one for two-position duty and another for modulating duty that is sometimes called the dynamic close-off rating. The dynamic rating, which is always lower than the twoposition rating, is the maximum DP allowed for smooth modulation of the valve, particularly near shut off. Above this DP, the design turndown ratio will not be achieved. This is the rating that should be used when selecting a valve for modulating applications. In two-way valve systems, a common practice is to require that valves be capable of modulating and/or shutting off against the pump shutoff head plus a safety factor (typically 25% to 50%). This is conservative for systems with variable-speed driven pumps but still advisable because the pumps may be operated at fixed speed in case of VFD failure.

Valve Actuators Control valves and actuators for chiller plants should be purchased as a single unit that is designed to meet the specified requirements. Variations in breakaway torque as well as closing force variations due to fluid pressure may affect the size of the actuator required. Purchasing the valve and actuator as an assembly ensures that the performance of the assembly will meet specifications. Knowing the valve position of modulating valves is required if advanced reset logic is desired for CHW supply temperature set point and DP set point. Analog actuators, those controlled by analog outputs from the DDC system using 0–10 Vdc or 4–20 mA signals, are generally preferred because valve position is equal to the signal sent to the valve, other than the time delay for the motor as it adjusts the position to match the signal. Floating point actuators, which are controlled by two digital outputs (one to open the valve and one to close it), are less expensive, but actual valve position is not known unless feedback from the actuator is available and wired to a DDC analog input. With this feedback signal added, the cost is generally the same as for analog actuators. Hence, it is recommended that modulating valves have analog actuators. The actual position of two-position actuators is also desired to ensure failsafe operation. These actuators can be specified to include end switches that indicate full-open or full-closed position wired to DDC system digital input points.

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Controllers Minimum DDC Controller Requirements Programmability Designers should ensure that any DDC system that is to be used for chiller plant control have a powerful and flexible programming language. This language should permit a nearly unrestrained capacity to incorporate logic, mathematics, timing, and other functions. This is important not only for the initial development of the plant’s operations but also for future improvements to the plant’s performance.

Variables Chiller plants must be able to operate automatically under various operating conditions, including those caused by equipment failure and manual operator override. It is essential that the DDC system have the capacity for a large number of variables (sometimes called pseudo, virtual, or software points) so that features such as lead/lag sequences, automatic failure remedies, and operator disabling of equipment for maintenance can be implemented simply and effectively.

Flexible Input/Output (I/O) Point Capacity The DDC system must interface to I/O devices that use industry-standard interfaces. Also, each chiller plant may have a unique mix of digital and analog inputs and digital and analog outputs. Therefore, DDC controllers that have universal I/O points (those that can be configured as analog or digital) and have the capability of expanding point count with added I/O modules may provide lower cost and greater flexibility for future changes than those with dedicated I/O hardware.

Analog-to-Digital (A/D) and Digital-to-Analog (D/A) Resolution The A/D conversion must provide adequate resolution to read all analog inputs accurately to the number of significant digits desired. A 12-bit A/D resolution for analog inputs (4096 segments for the device span) is recommended for analog inputs. Similarly, D/A resolution must be high enough on analog outputs to provide good turndown control of valves and other devices; 8-bit or 10-bit D/A resolution for analog outputs is usually acceptable.

Automatic Network Each DDC controller used in a chiller plant must have the capacity to automatically and seamlessly share all point and variable information with other controllers in the plant. It must also be able to preserve the required analog


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value precision for those points whose value must be transmitted across the network. Also, the network characteristics must require that • •

the maximum change of value (COV) (where employed) for all points be the same as the specified precision or accuracy of those points, and the maximum scan time (where employed) be less than 30 seconds (this means that all controllers will use point data that is within at least 30 seconds of the current value).

Trend Logs For commissioning and long-term performance monitoring, controllers must have the capability to record historic data on the status of control points, to analyze this data, and produce trend logs that show the behavior of the control point relative to other variables. Plant operators should have the capability to identify the control points for which trend logs are generated, to set the time interval for taking data, and otherwise configure trend logs. It is very desirable to be able to trend all hardware points and key software points, such as set points and loop output points, in the system at short intervals (e.g., 1 minute) during commissioning and at longer intervals (e.g., 5 minutes) for long-term performance monitoring and diagnostics. The network must be designed to be able to pass this data robustly from controllers back up to the control system server where historical trend data may be stored for later analysis. Data should be stored in a format, such as an SQL database, that can allow for post-analysis by external performance analysis software. Proprietary data storage formats should be avoided because they can limit how the trend data may be used. Trending at short intervals can result in many megabytes of data, so a very large disk capacity at the control system server or operator workstation is recommended.

Network Interfaces Network Connections to Equipment Instrumentation Some equipment, such as chillers and VFDs, have built-in controls. These controls should include an interface, such as a BACnet gateway, that can be connected to the DDC system to allow it to access the built-in control and monitoring points and avoid the cost of installing redundant control points as part of the BAS. For instance, chillers have controls that include CHW and condenser water temperature sensors. If data from these sensors is accessible to the BAS system, the installation of additional sensors can be avoided. In general it is good practice to limit these network connections to monitoring and not control of critical components, as network connections often drop out or experience momentary loss of connection. This is particularly important for plants that serve mission-critical loads like manufacturing, hospitals, and data centers. Also, transferring control loop outputs (such as pump speed) across networks can result in unstable operation because of slow or inconsis-

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Chapter 7 Controls tent network speed. In general, controlled points (e.g., DP), controlled devices (e.g., VFD speed), and the associated PID loop should be connected to/reside in the same controller. Specifications should be specific as to which points should be hardwired and which can be transferred via the network. With all network connections it is important to make sure to coordinate the work of the equipment manufacturer and the BAS contractor in the specifications. Critical details include the network communication language (e.g., BACnet), physical link (e.g., RS 485), and points to be mapped from the network device to the BAS. Be aware that many gateways allow only a limited number of the available points to be mapped to the external BAS.

Chiller Interface The start/stop point should be hardwired for all plants. For critical plants such as those serving data centers, status and alarm points should also be hardwired. Status can often be deduced from other variables, but that can lead to time delays in starting backup equipment, which can result in critical loss of plant capacity or supply temperature control. Even in systems where CHW supply temperature set point is actively reset, it is usually not a critical point, so network speed and reliability concerns do not mandate that the point be hardwired. The exception is chiller plants serving a single coil where supply air temperature is controlled by CHW temperature directly. The following are the minimum chiller monitoring points that should be accessible from the chiller controller to the BAS via the network: • • • • • • • • • • • • • • • • • • •

Supply (leaving) CHW temperature Return CHW temperature Supply (entering) condenser water temperature Leaving condenser water temperature Evaporator refrigerant pressure Evaporator refrigerant temperature Condenser refrigerant pressure Condenser refrigerant temperature Compressor discharge refrigerant temperature Oil temperature Oil pressure Chiller electrical demand (power) Chiller electrical current CHW flow status Condenser water flow status Chiller operating status Chiller alarm status Inlet vane (centrifugal) or slide valve (screw) position Compressor speed (if variable speed)


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The following are the minimum chiller control points that should be controllable from the BAS to the chiller controller via the network (if not hardwired): • • •

Chiller start/stop Chiller demand limit set point CHW supply temperature set point

VFD Interface The following points should be hardwired for all plants: • •

Start/stop Speed

As indicated in the section Network Connections to Equipment Instrumentation, time delays in transmitting the speed signals over the network can cause control loops to be unstable or at least difficult to tune. Hence, speed should always be a hardwired point. For critical plants, such as those serving data centers, VFD status should also be hardwired to avoid time delays in starting backup equipment. The following are the minimum VFD monitoring points that should be accessible directly from each VFD to the BAS via the network: • • • •

Status Fault or alarm status Actual speed Power

Performance Monitoring Integrating Chiller-Plant Efficiency Monitoring with Control The Benefit of Performance Monitoring In many climates, chiller plants are responsible for a major portion of a facility’s energy use. Performance monitoring can help identify energy efficiency opportunities. Many chiller plants are not fully automated, and nearly all plants require ongoing maintenance to achieve top operating efficiencies. Integrating chiller plant monitoring with the control system helps the plant operating staff determine the most efficient equipment configuration and settings for various load conditions. It also helps the staff schedule maintenance activities at proper intervals so that maintenance is frequent enough to ensure the highest levels of efficiency but not so frequent that it incurs unnecessary expense.

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Monitoring Considerations It does not have to be expensive to integrate energy efficiency monitoring with chiller plant control. Most DDC systems capable of operating chiller plants effectively are well suited to provide monitoring capabilities. Because chiller plant efficiency is calculated by comparing the CHW energy output to the energy (electric, gas, or other) required to produce the chilled water, efficiency monitoring requires only the following three items: •

CHW output: Most plants will have the instrumentation needed for measuring plant capacity (CHW flow and supply and return temperatures) without additional cost because these sensors are needed for normal control.

•

Energy input: Again, most plants will already have the necessary power metering because power meters are inherent in VFDs and constant-speed condenser water pump energy can be estimated from on/off status, as discussed under the Electric Power Meters section. Fixed-speed chillers generally do not include power meters, so one must be added for each chiller or to the plant as a whole in this case.

•

DDC math and trend capabilities: The DDC system must have math functions so that the instrumentation readings can be easily scaled, converted, calculated, displayed, and stored in trend logs for future reference. Again, this is inherent in almost all DDC systems at no added cost.

In most cases, the only cost premium for performance monitoring is the programming required to calculate performance metrics (such as kW/ton) over various time intervals and to configure graphical displays and trends.

Data Displays For the performance data to be useful to plant operators, the data must be collected over various time intervals so the current and prior performance can be compared. For instance, a graphical display might show monthly and year-to-date data for plant output, energy, and average/minimum/maximum efficiency (kW/ton) compared to the same data for prior years. Then, by inspection, the operator can see trends, such as plant efficiency degrading over time. For systems with interfaces to the chiller controllers, this performance graphic should also include trends for condenser and evaporator refrigerant-towater approach temperature differences, which are an indicator of tube fouling. Alarms should be generated when approach temperatures stray a few degrees from design values.


235

Control Schematics Control Schematic for a Typical Plant Control diagrams and schematics are represented in various ways with various levels of detail. Their primary purposes are to do the following: • •

Indicate the required control system sensors and their locations in the system Show the system schematically as a whole so that the relationship of the various components and equipment is clear

The latter goal is best achieved by limiting clutter in the schematic by showing only control devices; other system components that are unrelated to control, such as isolation valves, check valves, makeup water connections, expansion tanks, air separators, and pressure relief valves, should be shown on separate piping diagrams and details. Instrumentation such as thermometers, pressure gages, non-control water meters (such as at tower water makeup connections), test plugs, etc., should also be shown on piping diagrams (not control diagrams) because they are not dynamic control devices and they are also not provided by the BAS contractor. Figure 7-6 is a typical schematic of a primary-only CHW plant serving an office building. Some of the key elements of the design are as follows: •

•

•

•

Condenser water pumps are constant speed. As discussed in more detail below in the Condenser Water Pumps section, use of VFDs on condenser water pumps appears to be only marginally cost-effective if pump speed is optimally controlled. Non-optimal control logic can easily increase plant energy usage. So, VFDs on condenser water pumps are recommended only where sequences will be optimized through rigorous analysis or via realtime optimization add-on modules. Because CW pumps are constant speed, kilowatt meters are shown on them for overall plant performance monitoring. Meters are not required on other equipment because they have VFDs, which inherently include power meters. Pump power could also be estimated based on pump status and a one-time power measurement (performed after plant commissioning) when each pump is on alone and when both are on together. The VFDs to pumps and tower fans have hardwired start/stop and speed points with a network connection to the VFD interface for all other points. As noted previously, if the plant were serving critical systems such as data center air-handling systems, status points would also have been hardwired to allow faster and more reliable response to failed equipment. Towers do not include any isolation valves to shut off flow to allow one tower to operate alone. As noted in Chapter 2, towers can generally be selected with nozzles and dams to allow half flow while still providing full coverage of fill, and it is always most efficient to run as many cells as possible.

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Chapter 7 Controls

Figure 7-6 Typical CHW plant control schematic.


237

•

Makeup water to the condenser water system is shown to be controlled by the BAS using a level sensor mounted on a standpipe connected to the tower equalizer. This allows the tower level to include high- and low-level alarms and other benefits. See discussion in Chapter 3.

•

The condenser water isolation valves at each chiller are wired to the head pressure control output of the chiller controller, an output that is standard on most chiller controllers. (The term head pressure, a historical term for condenser pressure, is used here because it is common. In actuality, minimum lift, the difference between condenser and evaporator pressure, is what is being controlled by the chiller controller.) This allows the valve to do double duty: it can shut off flow when the chiller is off and also modulate flow to maintain head pressure when the system is first started and tower basin water may be cold. Head pressure control is usually required on screw chillers unless they have some separate means for oil control. Most centrifugal chillers can operate at low head long enough for the tower water to warm up so head pressure control is not needed unless the tower basin contains a very large volume of water or the chiller must operate in cold weather. Note that controlling head pressure in this manner obviates the need for a tower bypass control valve except in locations where the chiller is required to operate in freezing weather. Because valve position is not DDC controlled, a valve position feedback signal is wired to the DDC system so that position is known. Another option is to wire the chiller head pressure demand signal as an analog input to the DDC system, then control the condenser water valve as an analog output from the DDC system. This allows the DDC system to filter the logic of the chiller controller to ensure it is performing the desired function (including shutting off flow when the chiller is off). A similar approach can be used for chillers that do not have direct head pressure control outputs. Either actual refrigerant lift (difference between condenser and evaporator refrigerant pressure) can be passed to the DDC system from the chiller controller via the network interface or a lift indicator, such as the difference between leaving condenser water temperature and leaving CHW temperature, can be used as the controlled variable.

•

Flow switches are shown on the CHW side of the system but not on the condenser water side. Flow switches on the condenser water side are not needed and can cause false trips when head pressure control is active, as explained in Chapter 3.

•

Chillers are controlled by only a single hardwired point for on/off. All other points are mapped through the chiller controller network interface. That includes CHW and condenser water temperatures; separate fieldmounted temperature sensors hardwired to the DDC system are not required.

•

CHW isolation valves are modulating rather than two position. This ensures the chillers can be sequenced smoothly without rapid flow changes to prevent chiller trips when staging from one chiller to two. See the section

238

Chapter 7 Controls Control Sequences for further explanation. As chiller controllers become more robust, modulating valves may no longer be needed (two-position valves with electric actuators are already fairly slow moving), but the cost increase to use modulating rather than two-position valves is relatively small and can be justified by the flexibility they modulating valves add to control the speed at which valves open and close. •

CHW flow and supply and return water temperatures are shown to be part of a Btu meter package to improve the accuracy of plant load calculation. The Btu meter is shown connected to the BAS via a network connection through which flow, supply and return temperatures, and calculated Btu/h data are available. The flowmeter is also shown hardwired to a BAS analog input. This is required because the flowmeter is also being used by the BAS for minimum chiller flow control and controlled variables should be connected to the same controller as the controlled device (the minimum flow bypass valve) to avoid instability caused by slow and inconsistent network speeds. Minimum chiller flow could also be controlled by measuring DP across each chiller and correlating DP to flow, but a flowmeter is more accurate. The flowmeter also can be used directly for CHW pump staging and (with Btu calculation) for chiller staging and cooling tower set point reset. Therefore, a flowmeter is highly recommended and assumed to be installed in the control sequences discussed in the next section.

•

The minimum flow bypass valve used to ensure minimum chiller flow is maintained is located so that the flowmeter measures the flow through the chillers so it can be used for valve control. A separate CHW return temperature sensor is added upstream of the valve so that the temperature of the water from the coils can be monitored.

•

A DP sensor is shown remote from the plant. As noted in Chapter 3, the further out into the system the sensor is located, the lower the DP set point can be, which results in the lowest pump energy. In many plants where chilled water is piped to several different branches, multiple DP sensors are required, one for each branch. Each could have its own set point as determined by the balancing contractor to ensure adequate pressure is available downstream in each branch. Separate control loops would be executed for each sensor and the largest loop output would be used to control pump speed. These sensors should all be hardwired back to the plant controller rather than passed through the network to avoid control instability caused by slow and inconsistent network speeds.

•

Both CW supply and return temperature sensors are shown hardwired to the DDC system. Both temperatures are available individually for each chiller through the network interface. Whichever sensor is used for control (we recommend the return temperature in sequences below in the section Tower Fan Speed Control) must be hardwired to the DDC system because controlled points and controlled devices (tower fan speed in this case) should not rely on the network for control, as explained above in the sec-


•

•

239

tion Network Connections to Equipment Instrumentation. The other sensor is optional and could be eliminated because it is for monitoring only and the temperatures are available through the chiller interface. No CW flowmeter is shown. As discussed below in the section VariableSpeed Condenser Water Pumps, CW flow is desirable as a controlled variable for optimum control when VFDs are used on condenser water pumps. CW flow is not needed for constant-speed CW pump control. However, a side benefit of having a CW flowmeter is that it can also be used as a realtime instrumentation check: condenser heat rejection can be compared to the sum of compressor power (adjusted for motor losses if open drive) plus evaporator load with any discrepancy indicating that a sensor is faulty or out of calibration. Where CW flowmeters are used, full-bore magnetic or ultrasonic types are recommended to avoid measurement errors and added maintenance due to fouling and corrosion. The plant is instrumented sufficiently for total plant efficiency measurement, typically characterized as kW/ton . The thermal energy meter provides accurate plant load data. All variable-speed chillers (assumed in this example, Figure 7-6) and many constant-speed chillers include power meters. Pump and tower fan VFDs also include power meters. Chiller and VFD power data can be transmitted through the network interface. (Note that power measured by VFDs is output power, not input power, so VFD inefficiency is not included. However, VFD efficiency is very high except at low speeds where power is also very low, so the error should be small; see Chapter 3.) If condenser water pumps are constant speed, either power meters can be installed (as shown in this example) or pump power can be calculated based on pump status as described above in the section Electric Power Meters.

Control Sequences Determining Optimal Control Sequences CHW plants have many characteristics that make each plant unique so that generalized sequences of control that maximize plant efficiency are not readily determined. Equipment and system variables (see Chapter 5) that affect performance include the following: •

•

Chillers: Each chiller has unique characteristics that affect full-load and part-load efficiency, such as compressor design, evaporator and condenser heat transfer characteristics, unloading devices (such as VFDs, slide valves, and inlet guide vanes), and internal control logic. Cooling towers: Tower efficiency (gpm/hp) varies significantly by almost an order of magnitude between a compact centrifugal fan tower to an oversized propeller fan tower. Towers can also be selected for a wide range of approach temperatures.

240

Chapter 7 Controls •

CHW and condenser water pumps: Pumps and piping systems can be selected for a broad range of Ts and may or may not include VFDs. Pump efficiency also varies by pump type and size, and pump head varies significantly depending on physical arrangement and pipe sizing standards.

•

CHW distribution systems: Distribution system arrangements, such as primary/secondary versus primary-only variable flow, significantly affect plant control logic.

•

Weather: Changes in outdoor air conditions affect loads and the ability of cooling towers to reject energy.

•

Load profile: The size and consistency of loads affect optimum sequences. For instance, control sequences that are optimum for an office building served by air-handling systems with air-side economizers may not be optimum for a data center served by systems without economizers.

It is clear that no single control sequence will maximize the plant efficiency of all plants in all climates. There are a number of articles (Hartman 2005; Hydeman and Zhou 2007) on techniques to optimize control sequences for CHW plants. Almost all require some level of computer modeling of the system and system components and an associated amount of engineering time that most plant designers do not have. In developing this SDL, significant modeling was performed in an effort to determine generalized control sequences that accounted for the variation in plant design parameters summarized above. The technique used to determine optimized performance is described in a June 2007 ASHRAE Journal Article (Hydeman and Zhou 2007). In brief, the technique involves developing a calibrated simulation model of the plant and plant equipment that is run against an annual hourly CHW load profile with coincident weather data while parametrically modeling all of the potential modes of operation at each hour using multiple nested iterative loops. See Figure 7-7. The operating mode requiring the least amount of energy for each is determined. The minimum hourly energy usage is summed for the year—this is the TOPP. Because all modes of operation were simulated, the plant performance cannot be better than the TOPP within the accuracy of the component models. The operating modes (e.g., number of chillers, condenser water flow and pump speed, tower fan speed and related condenser water temperatures) that result in the TOPP are then studied using scatter plots, frequency charts, etc., to see how they relate to independent variables such as plant load, operating temperatures, and wet-bulb temperature in order to find trends that can be used to develop simple sequences to control the plant in real applications through the DDC system. Ideally, equipment should be controlled as simply as possible—complex sequences are less likely to be sustained because operators are more likely to disable them at the first sign of perceived improper operation.


Figure 7-7 Theoretical optimum plant performance (TOPP) model flow chart.

241

242

Chapter 7 Controls Equipment models include the following: •

Chillers ° Regression-based electric chiller model used in EnergyPlus (Hydeman et al. 2002) ° Multipoint calibration using zero tolerance manufacturer’s data (see Chapter 6 of this course book) • Towers ° DOE-2.2 model ° Calibrated using manufacturer’s data for real tower selections • Pumps ° Multiple piping sections assuming near turbulent flow (P = C · gpm1.8) ° Pump efficiency based on regression models of manufacturer’s data for real pump selections • VFD and motor efficiency ° Part-load curves from manufacturer’s data (see Chapter 3 of this course book) The sequences described in the sections below were developed from the TOPP modeling for the all-variable-speed CHW plant shown in Figure 7-6 serving a typical office building for a wide range of plant design options for tower range, approach, and efficiency; different chiller types and chiller efficiencies; and varying climates (see Appendix A for details). Also included are control sequences for constant-speed chillers and pumps developed from past experience.

Cooling Towers Tower Staging As noted in Chapter 4, cooling towers are most efficient when the most cells are operated within the flow limits of the towers. Plants with two, and sometimes three, tower cells can be designed so that all cells are active under all load conditions by selecting tower distribution pans and nozzles for the flow of only one condenser water pump. Where cells must be staged using isolation valves (discussed in Chapter 4), the maximum number of cells possible should be enabled while still maintaining minimum flow through each active cell.

Tower Fan Speed Control A common approach to controlling cooling towers is to reset condenser water supply temperature based on outdoor air wet-bulb temperature. However, our simulations seldom indicated a good fit; as shown in Figure 7-8, the correlation of optimum condenser water supply temperature versus wet-bulb temperature was fairly good in Miami but not in Oakland or most other climates.


Figure 7-8

243

TOPP optimum condenser water supply temperature versus wet-bulb temperature.


For plants serving typical office buildings,1 good correlations were found in all TOPP simulations between plant part-load ratio (PLR, actual plant load divided by total plant design capacity) and the difference between the CWRT (leaving the condenser) and the CHW supply temperature (CHWST, leaving the evaporator). Examples are shown in Figure 7-9. The CWRT – CHWST difference is a direct indicator of the refrigerant lift (the condenser and evaporator leaving water temperatures are determined by the condenser and evaporator temperatures), which drives chiller efficiency. The data in Figure 7-9 can be fit to a straight line: LIFT = CWRT – CHWST = A · PLR + B

(7-2a)

where A and B are coefficients that vary with climate and plant design (see Appendix A). LIFT calculated from Equation 2a must be no larger than the maximum lift at design conditions (design CWRT minus design CHWST from equipment schedules) and must be no less than the minimum lift required at minimum load (obtained from the chiller manufacturer). The latter varies among chiller types and manufacturers. For frictionless (magnetic) bearing chillers, the minimum lift is typically less than 5°F because these chillers have no oil, eliminating one of the primary reasons for maintaining minimum lift. Some manufacturers of these chillers even claim to have 0°F minimum lift requirement. The minimum lift at minimum load for standard bearing variable-speed centrifugal chillers (e.g., for the chiller modeled in Figure 7-9) is typically around 9°F to 20°F. 1. For plants with more consistent loads that do not vary with weather, such as those serving data centers and those located in consistently humid climates, such as Miami, correlation of load with CWRT/CHWST temperature difference is poor. For these plants, optimum CWST versus wet-bulb temperature was found to have better correlation. However, for office buildings in general, the correlations in Figure 7-9 were more consistent.

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Chapter 7 Controls

Figure 7-9

TOPP (CWRT – CHWST) versus plant load ratio.


Minimum lift is typically highest for screw chillers at 25°F to 30°F, but it may be lower for variable-speed screw chillers, which generally have separate oil separators and pumps to avoid having to rely on refrigerant lift for oil return. The lower this minimum is, the lower the annual chiller plant energy will be, particularly in mild climates. Equation 2a can be solved for the optimum CWRT set point given the current CHWST: CWRT = CHWST + A · PLR + B

(7-2b)

Cooling tower fans are then modulated to maintain condenser water return temperature at this set point. Controlling tower fan speed based on return temperature rather than supply temperature is nonconventional, but it makes sense because it is the temperature leaving the condenser that determines chiller lift, not the entering (supply) water temperature. Chiller efficiency is not sensitive to entering chilled- or condenser water temperature. Controlling fans off of return temperature rather than supply temperature is even more critical with variable-flow condenser water pumps because CWRT rises as CW flow falls. See the Variable-Speed CW Pump section.


245

Controlling tower fans off of CWRT can be unstable if the towers and chillers are greatly separated, such as towers located on the roof and chillers located in the basement. This is due to time delays caused by the relatively slow-moving water mass in the piping and is exacerbated by control network delays if the sensors are not all wired to the same controller. Where towers and chillers are separated, cascading control loops are recommended: the tower fans would be controlled by the water temperature leaving the tower (CWST) using the controller near the towers, but the set point for this loop would be reset by a second control loop maintaining the water temperature leaving the chillers at the set point from Equation 2a using the controller located near the chillers. As with all cascading loops, the loop resetting the CWST set point must be slower than the loop controlling tower fan speed for stable operation.

Chilled-Water Pumps Primary CHW Pumps on Primary/Secondary Systems Constant-speed primary pumps on primary/secondary systems generally are staged with chillers: if a chiller is started, then a pump is started; if a chiller is stopped, a pump is stopped. Where primary pumps have VFDs, they must be controlled to maintain primary flow at least equal to the secondary flow (to avoid the death spiral discussed in Chapter 4) and above the chiller minimum flow rate. One of two sequences is commonly used: •

•

If the primary and secondary circuits both have flowmeters, primary pump speed can be controlled to maintain primary flow ~10% larger than secondary flow but no lower than the sum of the minimum flow rates of all active chillers. This logic requires reliable flowmeters, such as full-bore magnetic flowmeters discussed above in the section Liquid Flow Sensors. If there are no flowmeters, primary pump speed can be controlled to maintain secondary CHW supply temperature equal to the primary CHW supply temperature down to a minimum speed that provides minimum chiller flow as determined during the test and balance phase. This is best done using trim and respond logic (Taylor 2015): primary pump speed is steadily dropped until secondary CHW temperature rises above primary CHW temperature, which generates a “request” for higher speed. Examples of trim and respond logic are provided in the example sequences in Appendix B.

Secondary CHW Pumps on Primary/Secondary Systems The speed of secondary pumps in primary/secondary systems is typically controlled to maintain DP measured far out in the system at set point. The DP sensor should be as far out in the system as possible, as discussed in Chapter 3.

246

Chapter 7 Controls The design DP set point for variable-speed pumps should be determined by the system balancer using one of the following procedures depending on whether or not the system is balanced (see Balancing Variable-Flow Systems in Chapter 4): •

•

For systems that are fully balanced, the set point is simply the DP reading when all coils are balanced and operating at design flow rates. The most remote coil should have its balance valve fully open if the system is properly balanced. For systems that are not balanced, the set point is determined by first closing all control valves except those downstream of the DP sensor, then, manually decreasing pump speed until flow through one of these downstream coils just reaches its design rate. The DP reading at this condition becomes the set point.

The DP set point can be reset to minimize pump energy usage under lowload conditions. This is done by monitoring valve position and resetting the set point as required to maintain the most open valve near full wide open. See the discussion in the Chilled-Water Temperature and Differential Pressure (DP) Set Point Reset section. For large buildings or campuses, DP sensors may be located far from the plant. Pump control in this case can be unstable due to time delays between a change in pump speed and the resulting DP change, and greatly exacerbated by control network delays if the DP sensors are not wired to the controller controlling pump speed. Where DP remote sensors are required, cascading control loops are recommended: the pumps would be controlled by a DP sensor located at the plant tied to the controller controlling pump speed, but the set point for this loop would be reset by the output of a second control loop maintaining the remote DP sensor at its set point with the loop residing in the controller to which the DP sensor is connected. This logic allows the use of multiple DP sensors; the highest output of each remote loop would be used to control the pump. As with all cascading loops, the loops resetting the set point must be slower than the loop controlling pump speed for stable operation.

Pump Staging Figure 7-10 shows the optimum number of CHW pumps as a function of CHW flow ratio (CHWFR, actual flow divided by design flow) and as a function of pump speed for a two chiller constant primary/variable secondary system with two secondary pumps plant based on TOPP modeling. Unlike cooling towers, the optimum sequence is not to run as many pumps as possible. This is because the pumps all pump through the same circuit (other than the pipes into and out of each pump between headers), so there are not cube law energy benefits for each pump individually. Because of the minimum DP being maintained at coils (which causes the system curve to bend off of the ideal curve at low flow, reducing pump efficiency) and because motor efficiency falls rapidly at low loads, running excess pumps will increase energy use. The optimum


247

(a)

(b)

Figure 7-10

(a) TOPP CHW pump staging versus (b) CHWFR and pump speed.


staging in Figure 7-10 is about 47% flow as opposed to 50% flow because of the piping dedicated to each pump as noted above; the pressure drop in these piping sections is minimized when more pumps are operating. As suggested by Figure 7-10, the optimum number of CHW pumps should be staged as a function of CHW flow, not CHW speed as is common practice. Specifically, stage up from one pump to two pumps above 47% CHWFR; stage down to one pump below 47% CHWFR, where CHWFR = actual flow divided by design flow. Provide a time delay (e.g., 10 minutes) between each stage to prevent short cycling. The above logic applies to a two-pump plant but it can be extended to an Npump plant: • • • •

Stage up from one pump to two pumps at (1/N – 3%) CHWFR and vice versa Stage up from two pumps to three pumps at 2 · (1/N – 3%) CHWFR and vice versa Stage up from three pumps to four pumps at 3 · (1/N – 3%) CHWFR and vice versa Etc.

All operating pumps run at the same speed. Time delays must be provided between stages. (The time delay could also be provided by creating a deadband between staging up and staging down set points, but using actual timers is more direct and can be more energy efficient.)

248

Chapter 7 Controls Using flow to control pump staging, rather than speed, is also beneficial for plants with three or more pumps because using pump speed alone can result in running too few pumps, which in turn can result in pumps running off the ends of their curves and cavitating. For very large pumps (≳ 100 hp), it may be worth the effort to determine the actual pump operating point (flow versus head) and optimize staging based on pump efficiency determined by flow and pressure drop readings mapped to pump curves duplicated mathematically in the DDC system (Rishel 2001). This can allow pumps to operate closer to their design efficiency as the system operating curve varies from the ideal parabolic curve due to modulating valves and minimum DP set point. However, the small potential energy savings is not worth the effort for most CHW plants.

Primary-Only CHW Pumps Staging primary pumps in a primary-only variable-flow system is identical to staging secondary pumps as described above in the Pump Staging section. The pumps must respond to the flow and pressure requirements of the system, not to the load. For headered variable-speed pumps (see Figure 4-17b), it is not necessary to start a pump when a chiller starts. For instance, two pumps may be able to meet the flow requirements for three chillers over a wide flow range. Conversely, if there is significant T degradation, three pumps could operate to meet flow requirements while running only two chillers to meet the load. This is one of the advantages of this design, and it is therefore recommended for primary-only variable-flow systems over dedicated pumps piped directly to each chiller (see Figure 4-17a).

Condenser Water Pumps Constant-Speed CW Pumps Constant-speed condenser water pumps generally are staged with chillers: if a chiller is started, then a pump is started; if a chiller is stopped, a pump is stopped. In larger plants with three or more pumps, efficiency can be improved by operating fewer CW pumps than chillers (a quasi-variable-flow system), but actual sequences must be developed from energy models to avoid increasing plant energy. See the Variable-Speed CW Pumps section that follows.

Variable-Speed CW Pumps Optimum control of variable-speed CW pumps is challenging because flow reduction reduces pump and tower fan energy but simultaneously causes an increase in chiller energy due to higher lift. Because of these offsetting factors, energy savings even with TOPP optimization are small, as shown in Figure 7-11. (The x-axis of this figure is made up of codes for various simulations with different climates, tower efficiencies, and chiller types.) The savings are even smaller when real sequences are used based on the TOPP models because we could not find any strong correlation between the TOPP optimized flow rate and any independent variables from which a good real sequence


Figure 7-11

249

Energy use of plants with constant speed versus variable speed.

could be developed. Optimum condenser water pump speed and flow were plotted against various parameters such as PLR, wet-bulb temperature, chiller percent power, and lift with no consistent relationships. The best correlation was flow versus PLR. Figure 7-12 shows optimized CW flow versus plant load, both expressed as a percentage of design flow and load. (Pump flow was varied in 10% increments so the data points are not continuous.) However, the correlations were seldom strong (R2 was typically less than 0.85 and some were as low as 0.5). The correlations were significantly weaker for pump speed than for flow, so a condenser water flowmeter should be added if one is not already part of the design. The data in Figure 7-12 can be fit to a straight line: CWFR = C · PLR + D

(7-3a)

where CWFR is condenser water flow ratio (percentage of design flow), and C an D are coefficients that vary with climate and plant design (see Appendix A). Pump flow set point, condenser water flow set point (CWFSP), can then be calculated from the design condenser water flow rate, CWFd: CWFSP = CWFd · CWFR

(7-3b)

This set point must be bounded by the minimum required CW flow rate obtained from the chiller manufacturer. The minimum flow from most manufacturers correlates to the onset of laminar flow and will be on the order of 40% to 70% of design flow depending on the design T, number of tubes, number of passes, and tube design (e.g., smooth versus enhanced). Higher rates are reputed to discourage fouling of condenser tubes, but, to our knowledge, no studies have been done to support that notion (Li and Webb 2001). Pump speed is then modulated to maintain measured flow at this set point. Optimum staging for variable-speed CW pumps was found to correlate very well to CW flow with 60% of the total design flow as the staging point (i.e., one pump should operate when the CWFR is below 60% and two pumps should operate when CWFR is above 60%, with a time delay to prevent short cycling).

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Chapter 7 Controls

Figure 7-12

TOPP optimum percent CW loop flow versus percent plant load.


When C and D coefficients determined for specific plants were fed back into the energy model, actual performance ranged from 101% to 110% of the TOPP. This performance gets worse when C and D are determined from the regression equations based on plant design (see Appendix A) rather than from actual plant performance modeling (e.g., Figure 7-12). Figure 7-13 shows life-cycle costs (based on LCCA parameters described in Chapter 5—essentially a 14-year simple payback period) for an Oakland, CA, office building with both constant-speed and variable-speed CW pumps with pump speed both optimally controlled and controlled using Equation 3b with coefficients C and D determined from curve-fits in Figure 7-12. The VFDs are barely cost-effective even with perfect controls and even assuming life-cycle cost parameters that equate to a 14-year simple payback period. In other words, even with perfect controls, the VFDs will barely pay for themselves in their typical 15-year service life. And with the reduced performance of real control sequences, CW pump VFDs are not cost-effective, as shown on the right side of Figure 7-13. Cost-effectiveness was even worse in more humid climates like Miami and a bit better in dry


Figure 7-13

251

Life-cycle costs for an Oakland office building with constant-speed (CS) and variable-speed (VS) CW pumps.

climates like Albuquerque. For plants seeing higher loads over more widely varying weather conditions, such as 24/7 data centers, variable-speed condenser water pumps can be cost-effective but only if the controls are optimized using plant simulation. Of even greater concern, our studies found that variable-speed CW pumps can increase the energy usage of the plant if not optimally controlled. For example, Figure 7-14 shows energy usage for a plant serving an office building in Denver, CO, and Figure 7-15 shows the same plant and building in Miami, FL, each using three control strategies: •

•

•

TOPP: This is the theoretical optimum plant performance of the plant with variable-speed condenser water pumps determined using the technique described in the section Determining Optimal Control Sequences). This is the theoretical best performance possible. Standard (STD): This is the performance of the plant with constant-speed condenser water pumps and cooling tower fans controlled to reset supply water temperature per AHRI Standard 550/590 condenser water relief curves (2015). This is most indicative of conventional practice. Oakland, CA (OAK): This is the performance of the plant with variablespeed pumps controlled using control sequences that were optimized for the same plant located in Oakland, CA (see Appendix A), instead of Denver, CO.

The figures show that energy usage is highest using control sequences that provided near-ideal performance for the same plant in another climate zone, significantly higher energy usage than the plant without VFDs on condenser

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Chapter 7 Controls

Figure 7-14

Denver CHW plant energy usage using three control strategies.


Figure 7-15

Miami CHW plant energy usage using three control strategies.


253

water pumps. This demonstrates how sensitive plant performance is to the details of the control logic. Based on this analysis, VFDs on CW pumps should not be used for most commercial applications for the following reasons: • • • • • •

Energy use may be increased if not optimally controlled. VFDs are unlikely to be cost-effective even if optimally controlled. A condenser water flowmeter is required for optimized control. Higher lift can cause chillers to operate in surge (see in the section VariableSpeed Chiller Staging for more details). Low flow rates may cause fouling of condenser tubes due to low tube velocities. Low flow rates may cause scaling in towers if rates are below minimum flow or tower isolation valves plus low-flow nozzles must be provided.

For plants that operate 24/7, VFDs may be cost-effective but should only be considered if the control logic is optimized using TOPP type simulations.

Chilled-Water Temperature and Differential Pressure (DP) Set Point Reset Chillers are more efficient at higher CHWST because this reduces lift. Resetting the CHW temperature set point upward when loads are low is always an effective energy-saving strategy for constant-flow systems. Reset may or may not be an energy-saving strategy in variable-flow systems depending on the plant design. High CHW temperature reduces coil performance, so coils in two-way valve systems will demand more chilled water for the same load, degrading T and increasing flow and pump energy requirements. Whether the net energy savings (chiller energy decrease less pump energy increase) is positive and sufficient to offset the cost of implementing the reset strategy depends on chiller performance characteristics and the nature of coil loads. This is discussed further below. CHWST set point reset strategies include the following: • •

•

Resetting inversely proportional to outdoor air temperature Resetting from return water temperature, either indirectly by maintaining a constant return water temperature or resetting the set point proportional to return water temperature Resetting from CHW valve position. See a detailed discussion on how this is implemented below

The last strategy (reset of valve demand) was once impractical with pneumatic controls and distributed controls in large campus buildings. However, it is readily done in systems with DDC for all control valves. It is by far the most reliable and efficient strategy in that it ensures that no coil is starved. The other strategies are indirect and cannot assure all coils will be satisfied unless they

254

Chapter 7 Controls are very conservative (i.e., will yield little actual reset). Using valve position also ensures that humidity control will be maintained. Contrary to conventional wisdom, the impact of reset on the dehumidification capability of air handlers is quite small and should not be a concern. Space humidity is a function of the supply air humidity ratio, which in turn is a function of the coil leaving drybulb temperature set point. Regardless of CHWST, the air leaving a wet cooling coil is nearly saturated; lowering CHWST only slightly reduces the supply air humidity ratio. So as long as the supply air temperature can be maintained at the desired set point, as can be indicated by valve position, resetting CHWST will not impact space humidity. Valve position can also be used to reset the DP set point used to control pump speed. In fact this is required by ANSI/ASHRAE/IES Standard 90.1 (2016a). (ANSI/ASHRAE/IES Standard 90.1-2016 allows either CHW temperature set point reset, DP set point reset, or both.) The logic is similar to CHWST set point reset: the DP set point is reset upwards until the valve controlling the coil that requires the highest DP is wide open. So, we have a dilemma: valve position can be used to reset either CHWST set point or DP set point but not both independently—it is not possible to know if the valve is starved from lack of pressure or from lack of cold enough water. We must decide which of the two set points to favor. While resetting CHWST set point upward reduces chiller energy use, it increases pump energy use in variable-flow systems. Higher CHW temperature causes coils to require more chilled water for the same load, degrading CHW T and increasing flow and pump energy requirements. Degrading T can also affect optimum chiller staging; however, this is not generally an issue in primary-only plants with variable-speed chillers (see Chapter 4). Furthermore, our simulations have shown that the positive impact of resetting CHW temperature on chiller efficiency is much greater than the negative impact on pump energy even when distribution losses are high for plants that have variablespeed chillers. Figure 7-16 shows a DOE2.2 simulation of a primary-only plant with variable-speed chillers and CHW pumps with high pump head (150 ft) using three reset strategies based on valve position: reset of CHW temperature alone, reset of DP set point alone, and a combination of the two that first resets CHW temperature then resets DP set point. The simulations were done in several climate zones (Houston and Oakland results are shown in the figure) and in all cases, resetting CHW temperature was a more efficient strategy than resetting DP set point. Sequencing the two (resetting CHW temperature first then DP set point) was the best approach, although only slightly better than resetting CHW temperature alone. Figure 7-17 shows how this sequenced reset strategy can be implemented. The x-axis is a software point called CHW Plant Reset that varies from 0% to 100% using trim and respond logic. The coil valve controllers generate “requests” for colder CHW temperature or higher pump pressure when the valve is full open. When valves are generating requests, CHW Plant Reset increases; when they are not, CHW Plant Reset steadily decreases. When CHW Plant Reset is 100%, the CHWST set point is at Tmin and the DP set


Figure 7-16

255

Plant energy with CHWST set point reset, CW DP set point reset, and a combination of the two.


Figure 7-17

CHWST set point reset sequenced with CW DP set point reset off of CHW valves.


256

Chapter 7 Controls

Figure 7-18

CW DP set point reset sequenced with CHWST set point reset off of CHW valves.

point is at DPmax. Tmin is typically the design CHW temperature for plants with variable-speed chillers but should be 1°F to 2°F lower for constant-speed plants; this allows the operating chillers to fully max out their capacity before staging on another chiller (see the Constant-Speed Chiller Staging section). DPmax is the design DP set point determined as described above for variablespeed CHW pumps. As the load backs off, the trim and respond logic reduces the CHW Plant Reset point. As it does, CHW temperature is increased, first up to a maximum Tmax (equal to the lowest air-handler supply air temperature set point less 2°F), then DP set point is reduced down to a minimum value DPmin (such as 5 psi). In practice, this logic seldom results in much reset of the DP set point—the CHWST reset is aggressive enough to starve the coils first—so it is important to locate the DP sensor(s) at the most remote coil(s) so that DPmax can be as low as possible to minimize pump energy. The opposite was found to be true for plants with constant-speed chillers. Their efficiency benefits less from the reduced lift, so the increase in CHW pump energy more than offsets the chiller savings. For these plants, the reset logic from valve position is the same but the DP set point is reset preferentially instead of CHWST. This is shown in Figure 7-18.

Chiller Staging Constant-Speed Chiller Staging For a plant composed of single-speed chillers, the most efficient logic is to operate no more chillers than required to meet the load. Chiller efficiency actually improves as load and lift fall (see Figure 7-19), so it may seem to make


Figure 7-19

257

Typical chiller part-load performance with and without VFDs (includes condenser water relief, as defined in AHRI Standard 550/590).

sense to run, say, two chillers at 40% load rather than one at 80% load, but this chiller savings is offset by the added energy of an additional CW pump (and an additional primary CHW pump on primary/secondary systems). Logic for staging chillers on is straightforward: a new chiller is started when the operating chillers are no longer able to meet the load, as indicated by plant leaving water temperature rising above set point by 1°F or 2°F. For plants that have the ability to reset CHW temperature, it is important that the CHW temperature of operating chillers be reset 2°F or so below design CHW temperature to ensure chillers are fully loaded before starting the next chiller. This reduces the efficiency of the operating chiller, but, for most singlespeed chiller plants, the total plant energy use will be less than if another chiller were started. See Chilled-Water Temperature Reset logic above in the section Chilled-Water Temperature and Differential Pressure (DP) Set Point Reset section. Logic for staging chillers off is trickier. First, it must be conservative because staging a chiller off prematurely will cause it to stage back on too quickly, causing excessive wear on the chiller and starters. The logic must determine that the load can be comfortably handled with one less chiller. For primary-only variable-flow plants (which should have a flowmeter for minimum flow control and pump staging), load can be determined from a Btu meter or calculated from flow and supply/return water temperatures. For constant-volume systems, flow is typically presumed from design data or mea-

258

Chapter 7 Controls sured once during balancing and assumed constant thereafter, thereby making a flowmeter unnecessary. For primary/secondary systems, stagedown logic cannot just be based on load; it must also ensure that primary flow always exceeds secondary flow to avoid the death spiral described in Chapter 4. This is best accomplished by installing a flowmeter in the secondary loop. A chiller can be staged off only if the load determined from this meter is below the capacity of the remaining operating chillers and the secondary flow rate will be lower than the primary flow rate (determined from a flowmeter or as described above for constant-flow systems) with the remaining operating primary CHW pumps.

Variable-Speed Chiller Staging Figure 7-19 shows the performance of fixed-speed versus variable-speed chillers with so-called condenser water relief—condenser water temperatures fall with the load per AHRI Standard 550/590. Note that without the reduction in lift provided by the reduced condenser water temperatures, the part-load efficiency of both constant-speed and variable-speed chillers gets worse at part load. The efficiency of fixed-speed chillers with condenser water relief improves as the load falls from peak, but, for most chillers, efficiency will start to rise above design efficiency at about 40% to 50% load. However, variablespeed chillers do not suffer from this degradation in efficiency until the load is very low, about 20% of full load with condenser relief. Because of the operation of ancillary equipment, such as condenser and primary CHW pumps, the overall plant efficiency will start to degrade at an even higher part-load point, as shown in Figure 7-21, but still well below 50%. This figure is for a plant with fixed-speed condenser water and primary CHW pumps. For systems with variable-speed primary-only pumps, the staging point is even lower, as in Figure 7-20. Figure 7-21 also corroborates the conventional wisdom that efficiency when staging fixed-speed chillers is maximized when operating chillers are maxed out before starting the next stage. Figure 7-20 shows the optimum number of chillers that should be run plotted against plant load for variable-speed centrifugal chillers in a plant with two equally sized chillers and variable primary distribution. The graph shows that it is often optimum to operate two chillers as low as 25% of overall plant load. This result may seem somewhat counterintuitive—conventional wisdom is to run as few chillers as possible. That is true for fixed-speed chillers but not for variable-speed chillers, which are more efficient at low loads when condenser water temperatures are low. Figure 7-20 shows that staging chillers based on load alone does not optimize performance because there is a fairly wide range where either one or two chillers should be operated. There is also another problem with staging based on load alone: it can cause the chillers to operate in surge. This can be seen in Figure 7-22, which schematically shows centrifugal chiller load versus lift, defined as the difference between condenser and evaporator refrigerant temperatures. If two chillers are operated when the lift is high (upper horizontal


Figure 7-20

TOPP variable-speed chiller staging versus plant load ratio (Albuquerque).


Figure 7-21

Two-chiller plant part-load performance with and without VFDs.

259

260

Chapter 7 Controls

Figure 7-22

Possible surge problem staging by load only.


line), the chillers will operate in the surge region. To avoid surge, the chiller controllers will speed up the compressors and throttle inlet guide vanes to control capacity. This reduces chiller efficiency so that it would then be more efficient to operate one chiller rather than two. But if the lift is low (lower horizontal line in Figure 7-22), the chillers would not be in surge, so operating two chillers would be more efficient than operating one. Therefore, in addition to load, chiller staging must take chiller lift into account. (This consideration applies only to centrifugal chillers; surge does not occur with positive displacement chillers, such as those with screw compressors.) Figure 7-23 shows the optimum number of operating chillers (light gray dots indicate one chiller while dark grey dots indicate two chillers) as determined by TOPP simulations. For all plant design options and for all climate zones, good correlations were found for the optimum staging point described by a straight line: SPLR = E · (CWRT – CHWST) + F

(7-4)

where staging part-load ratio (SPLR) is the PLR staging and E and F are coefficients that vary with climate and plant design as shown in Appendix A. If the actual measured PLR is less than the SPLR, one chiller should operate; if the PLR is larger than the SPLR then two chillers should operate, with a time delay to prevent short cycling.


Figure 7-23

Optimum staging versus (CWRT– CHWST) and plant PLR.


261

262

Chapter 7 Controls Primary-only variable-flow plants also require coordinated staging sequences and minimum flow control to avoid chiller trips. These sequences and why they are needed are discussed in more detail in Chapters 4 and 5.

Water-Side Economizer (WSE) Control Recommended control sequences for integrated WSEs are as follows: •

•

Reset CHW supply temperature set point based on valve demand, as described in the Chilled-Water Temperature and Differential Pressure (DP) Set Point Reset section. Enable the economizer if the CHW return temperature is greater than the predicted HX leaving water temperature (PHXLWT) plus 2°F. The 2°F differential is needed to avoid expending a lot of cooling tower fan energy for only minimal economizer load reduction. PHXLWT is estimated using the equation below: PHXLWT = T WB + P A HX + P A CT P A HX = D A HX  PLR HX P A CT = m   DT WB – T WB  + D A CT where TWB

=

current wet-bulb temperature

PAHX

=

predicted heat exchanger approach

PACT

=

predicted cooling tower approach

DAHX

=

design heat exchanger approach

PLRHX =

•

predicted heat exchanger part-load ratio (current chilledwater flow rate divided by design heat exchanger chilledwater flow rate)

DTWB

= design wet-bulb temperature

DACT

= design cooling tower approach

m

= slope developed from the manufacturer’s cooling tower selection program or empirically after the plant is operational (typical values are 0.2 to 0.5 for near-constant load applications like data centers; for office type applications, m is typically in the range of −0.2 to 0)

Disable the WSE if it is not reducing the CHW return temperature by at least 1°F.

Fundamentals of Design and Control of Central Chilled-Water Plants I-P • •

• •

•

263

Disable chillers when HX leaving water temperature (HXLWT) is at or below the desired CHW supply temperature set point. Enable chillers when CHW supply temperature is greater than the desired set point. Note that multiple chillers may need to be enabled if the current CHW flow is well above the design flow of a single chiller. Run as many tower cells as tower minimum flow limits allow. Control condenser water flow to roughly match the current CHW flow but reduce flow (within tower minimum flow constraints) as needed to maintain a minimum 5°F range. The lower flow and higher range improves tower efficiency and reduces pump power. Flow can be controlled by staging pumps, modulating speed on variable-speed pumps, and/or modulating isolation valves on the HX. Flow rate can be measured directly with a flowmeter (fullbore magnetic or ultrasonic types are recommended to prevent fouling) or deduced from HX pressure drop. Tower speed control logic varies based on WSE and chiller status: °

° °

When WSE is disabled: control fan speed to maintain normal condenser water temperatures, which should be reset from load or wet-bulb temperature as described above. When WSE and chillers are enabled: run tower fans at 100% speed. When chillers are disabled: control speed to maintain HXLWT at desired CHW supply temperature set point.

The only complex sequence above is predicting when the economizer should be enabled. Fortunately, if the prediction calculation is off and the economizer is enabled prematurely, it will shortly be disabled and the plant will see no disruptions in CHW flow or supply temperature. This contrasts with nonintegrated economizers where switching from economizers to chillers can be disruptive and guessing wrong about economizer performance can result in chiller short cycling and temporary loss of CHW supply temperature control.

Real-Time Optimization The TOPP modeling technique used to develop the sequences recommended above in this chapter has several disadvantages: • •

• •

It requires significant modeling effort, which not only increases engineering time but requires expertise not all design engineers have. Sequences are only valid for the range of conditions in the model of the plant load profile and concurrent weather conditions. The sequences may not be optimum or even stable for unexpected operating conditions. It relies on equipment models being accurate and cannot account for any degradation in equipment performance over time. The simplified correlations and curve fits are not always optimal and, in fact, can be so poor in the case of variable-speed condenser water pump control that VFDs are not recommended for these pumps.

264

Chapter 7 Controls An alternative to TOPP modeling is to perform real-time TOPP modeling in the DDC system or on a computer tied to the DDC system network. For every time step (e.g., 5 minutes or a bit longer depending on transitions such as equipment start/stop made in the previous time step) based on the current weather, load, and other operating conditions, the program would iterate through all possible operating set points, staging, etc. to determine the optimum. These would be implemented in real time and the process would be repeated for the next time step. Data on equipment performance could also be collected in real time to allow the program to automatically adjust equipment models to maintain their accuracy over time (e.g., to account for reduced chiller performance as tubes foul). This type of control system would truly provide the most optimum control possible provided equipment models are accurate and maintain their accuracy over time. There is currently a very limited number of companies providing real-time, model-based optimization, but it is expected to be standard practice one day, given its advantages.

Appendix A—TOPP Model Coefficients The plant in Figure 7-6 (with the optional addition of VFDs on condenser water pumps) serving a typical office building was modeled with all permutations of the following design variables: • • •

Weather: Oakland, CA; Albuquerque, NM; Chicago, IL; Atlanta, GA; Miami, FL; Las Vegas, NV CHWST: reset by valve position from 42°F to 57°F Chillers: ° °

•

Towers: ° ° °

•

Two styles (two stage and open drive) Efficiency at 0.35, 0.5, and 0.65 kW/ton at AHRI conditions Approach: 3°F, 6°F, 9°F, and 12°F Tower range: 9°F, 12°F, and 15°F Efficiency: 50, 70, and 90 gpm/hp

Condenser water pumps: with and without VFDs

The control equation coefficients were determined from each run, then these coefficients were themselves regressed against various design parameters and weather indicators. The results are shown in the subsections that follow. The development of these regressions is ongoing to include more weather sites and chiller variations. Because of the limited range of variables, the coefficients calculated using the equations that follow can be invalid if any of the variables are out of range from those used in the regressions, so these equations must be used with care. In each case, we provide simplified coefficients that we have found to be stable in most applications, although of course they are not optimized.


265

These control sequences strictly apply to primary-only plants with centrifugal chillers serving air handlers with outdoor air economizers in a typical office building. It is not known how well they apply to other applications.

Condenser Water Temperature Control Control cooling tower fan speed to maintain CW return temperature to the set point (CWRTsp) determined from Equation A-1 LIFT = A · PLR + B

(A-1a)

CWRTsp = CHWST + LIFT

(A-1b)

Lift calculated from Equation A-1b should be no smaller than LIFTm and no larger than LIFTd. Regressed coefficients (use with care): A = –63 + 0.0053 · CDD65 – 0.0087 · WBDD55 + 1.67 · WB + 0.52 · APPROACH – 0.029 · gpm/hp B = 18 – 0.0033 · CDD65 + 0.0053 · WBDD55 – 0.26 · WB + 0.15 · APPROACH – 0.014*gpm/hp Simplified coefficients (recommended): A = (LIFTd – LIFTm)/0.9 B = LIFTd – A

Variable-Speed Condenser Water Pumps Control CW pump speed to maintain CW flow at the set point determined from Equation A-2: CWFR = C · PLR + D

(A-2a)

CWFSP = CWFd · CWFR

(A-2b)

C and D coefficients must be determined through modeling. VFDs on CW pumps are not recommended otherwise.

Chiller Staging Use one chiller when the PLR is less than the SPLR determined from Equation A-3; use two chillers otherwise: SPLR = E · (CWRT – CHWST) + F

(A-3)

266

Chapter 7 Controls Regressed coefficients (use with care): E = 0.057 – 0.000569 · WB – 0.0645 · IPLV – 0.000233 · APPROACH – 0.000402 · RANGE + 0.0399 · kW/ton F = –1.06 + 0.0145 · WB + 2.16 · IPLV + 0.0068 · APPROACH + 0.0117 · RANGE – 1.33 · kW/ton Simplified coefficients: E = 0.45/(LIFTd – LIFTm) F = E · (0.4 · LIFTd – 1.4 · LIFTm)

Variables APPROACH = design tower leaving water temperature minus WB, °F CHWFR = chilled-water flow ratio, actual flow divided by total plant design flow CHWST = current chilled-water supply temperature (leaving evaporator temperature), °F = design condenser water flow rate CWFd CWFR = condenser water flow ratio, actual flow divided by total plant design flow CWRT = current condenser water return temperature (leaving condenser water temperature), °F CDD65 = cooling degree-days base 65°F (see Table 7-A) DP = differential pressure, ft H2O kW/ton = chiller efficiency at AHRI conditions T = temperature difference, °F = tower efficiency per ANSI/ASHRAE/IES Standard 90.1 IPLV = integrated part-load value per AHRI 550/590, kW/ton LIFT = CWRT – CHWST LIFTd = lift at design conditions (from equipment schedule) = CWRTdesign – CHWSTdesign LIFTm = minimum lift at minimum load (from chiller manufacturer) NPLV = nonstandard part-load value per AHRI Standard 550/590, kW/ton RANGE = design tower entering minus leaving water temperature, °F PLR = plant part-load ratio, current load divided by total plant design capacity TOPP = theoretical optimum plant performance WB = design wet-bulb temperature, ASHRAE 1% design condition, °F WBDD55 = wet-bulb cooling degree-days base 55°F (see Table 7-A)

Fundamentals of Design and Control of Central Chilled-Water Plants I-P Table 7-A

267

ASHRAE Climate Zone Weather Data

Location

Zone

CDD65

WBDD55

Miami

1A

4147

5255

Houston

2A

2898

3842

Phoenix

2B

4290

1554

Atlanta

3A

1646

2100

Los Angeles

3B (LA)

564

1127

Las Vegas

3B (LV)

3308

683

San Francisco

3C

198

214

Baltimore

4A

1272

1608

Albuquerque

4B

1406

418

Seattle

4C

272

248

Chicago

5A

948

1094

Boulder

5B

922

225

Minneapolis

6A

803

964

Helena

6B

561

130

Duluth

7A

284

374

Fairbanks

8A

146

87

Example TOPP modeling was performed for a plant serving an Oakland, CA office building. The following slopes and intercepts were determined from curve-fits: A = 47, B =5.2 C = 1.3, D = 0.13 E = 0.009, F = 0.21 Figure 7-A shows the theoretical optimum performance for both variablespeed (VS) and constant speed (CS) CW pumps compared to our proposed real sequences using the coefficients listed above. Despite their simplicity, our sequences resulted in only about 1% higher energy use than the TOPP. VFDs on the CW pumps saved 3% versus constant-speed pumps, but this was not enough savings to make them cost-effective at a 14-year simple payback period for this plant. Also shown in the figure for comparison is plant performance using the AHRI Standard 550/590 condenser water relief curve to reset condenser water temperature (4% higher energy use than our sequences) and performance assuming CWST set point is fixed at the design temperature (16% higher than our sequences).

268

Chapter 7 Controls

Figure 7-A

TOPP versus real sequences for both constant-speed and variable-speed CW pumps.


Appendix B—Detailed Sequence of Operation (SOO) The following is a detailed sequence of operation for the plant in Figure 7-6.

Sequence of Operation (SOO) 1.01 SEQUENCES OF OPERATION A. General 1. The term proven (i.e., “proven on”/“proven off”) shall mean that the equipment’s DI status point matches the state set by the equipment’s DO command point. 2. The term PID loop or control loop is used generically for all control loops and shall not be interpreted as requiring proportional plus integral plus derivative gains on all loops. Unless specifically indicated otherwise, the following guidelines shall be followed: a. Use proportional only (P only) loops for limiting loops (such as zone CO2 limiting loops, etc.) to ensure there is no integral windup.


269

b. Do not use the derivative term on any loops unless field tuning is not possible without it. 3. Trim and respond set point reset logic a. Trim and respond set point reset logic and zone/system reset requests where referenced in sequences shall be implemented as described below. b. “Requests” are pressure, cooling, or heating set point reset requests generated by zones or air-handling systems. 1. For each zone or system, and for each set point reset request type listed for the zone/system, provide the following software points: a. Importance multiplier (default = 1). This point is used to scale the number of requests the zone/system is generating. A value of zero causes the zone/ system’s requests to be ignored. A value greater than zero can be used to effectively increase the number of requests from the zone/system based on the critical nature of the spaces served or to increase the requests beyond the number of ignored requests (defined below) in the trim and respond reset block. b. Request hours 1. This point accumulates the integral of requests (prior to adjustment of importance multiplier) to help identify zones/systems that are driving the reset logic. Every x minutes (adjustable, default 5 minutes), add x/60 times the current number of requests to this request-hours accumulator point. 2. The request-hours point is reset to zero upon a global command from the system/plant serving the zone/system—this global point simultaneously resets the request-hours point for all zones/ systems served by this system/plant. 3. Cumulative percent-request-hours is the zone request hours divided by the zone run hours (the hours in any mode other than unoccupied mode) since the last reset, expressed as a percentage. 4. A Level 4 alarm is generated if the zone importance multiplier is greater than zero, the zone percent-request-hours exceeds 70%, and the total number of zone run hours exceeds 40. 2. See zone and air-handling system control sequences for logic to generate requests.

270

Chapter 7 Controls 3. Multiply the number of requests determined from zone/system logic times the importance multiplier and send to the system/plant that serves the zone/system. See system/plant logic to see how requests are used in trim and respond logic. c. Variables. All variables below shall be adjustable from a reset graphic accessible from a hyperlink on the associated system/plant graphic. Initial values are defined in system/plant sequences below. Values for trim, respond, time step, etc. shall be tuned to provide stable control. Device = associated device (e.g., fan, pump) SP0

= initial set point

SPmin

= minimum set point

SPmax

= maximum set point

Td

= delay timer

T

= time step

I

= number of ignored requests

R

= number of requests from zones/systems

SPtrim

= trim amount

SPres

= respond amount

SPres-max = maximum response per time interval d. Trim and respond logic shall reset set point within the range SPmin to SPmax. When the associated device is off, the set point shall be SP0. The reset logic shall be active while the associated device is proven on, starting Td after initial device start command. When active, every time step T, trim the set point by SPtrim. If there are more than I requests, respond by changing the set point by SPres times (R – I), that is (the number of requests minus the number of ignored requests). But the net response shall be no more than SPresmax. The sign of SPtrim must be the opposite of SPres and SPres-max. For example, if SPtrim = –0.1, SPres = +0.15, SPres-max = +0.35, R = 3, I = 2, then each time step, the set point change = –0.1 + (3 – 2) · 0.15 = +0.05. If R = 10, then set point change = –0.1 + (10 –2) · 0.15 = 1.1 but limited to a maximum of 0.35. If R2, the set point change is –0.1. 4. Lead/lag and lead/standby alternation a. Even wear 1. Lead/lag. Unless otherwise noted, parallel staged devices (such as pumps, towers) shall be lead/lag alternated when more than one is off or more than one is on so that the device with the most operating hours is made the later stage device and the one with the least number of hours is made the earlier stage device. For example, assuming there are


271

three devices, if all three are off or all are on, the staging order will simply be based on run hours from lowest to highest. If two devices are on, the one with the most hours will be set to be Stage 2 while the other is set to Stage 1; this may be the reverse of the operating order when the devices were started. If two devices are off, the one with the most hours will be set to be Stage 3 while the other is set to Stage 2; this may be the reverse of the operating order when the devices were stopped. 2. Lead/standby. Unless otherwise noted, parallel devices (such as pumps and towers) that are 100% redundant shall be lead/standby alternated when more than one is off so that the device with the most operating hours is made the later stage device and the one with the least number of hours is made the earlier stage device. For example, assuming there are three devices, if all three are off, the staging order will be based on run hours from lowest to highest. If devices run continuously, lead/standby shall switch at an adjustable runtime; standby device shall first be started and proven on before former lead device is changed to standby and shutoff. b. Exceptions 1. Operators shall be able to manually fix staging order via software points on graphics overriding the even wear logic above but not overriding the failure or hand operation logic below. 2. Failure: If the lead device fails or has been manually switched off, the device shall be placed into high-level alarm (Level 2) and set to the last stage position in the lead/ lag order until alarm is reset by operator. Staging position of remaining devices shall follow the even wear logic. A failed device in alarm can only automatically move up in the staging order if another device fails. Note that a device in alarm will be commanded to run if the sequence calls for it to run. In this way the BAS will keep trying to run device(s) until it finds enough that will operate. Failure is determined by: 3. Variable-speed fans and pumps 1. VFD critical fault is ON or 2. Status point not matching its ON/OFF point for 3 seconds after a time delay of 15 seconds when device is commanded ON or 3. Supervised HOA at control panel in tion or

OFF

posi-

4. Loss of power (e.g., VFD DC bus voltage = zero)

272

Chapter 7 Controls b. Constant-speed fans and pumps 1. Status point not matching its ON/OFF point for 3 seconds after a time delay of 15 seconds when device is commanded ON or 2. Supervised HOA at control panel in OFF position c. Chillers 1. Chiller alarm contact or 2. Chiller is manually shut off as indicated by the status of the local/auto switch from chiller gateway or 3. Chiller status remains off 5 minutes after command to start. 4. Hand operation. If a device is on in hand (for example via an HOA switch or local control of VFD), the device shall be set to the lead device and a low-level alarm (Level 4) shall be generated. The device will remain as lead until the alarm is reset by the operator. Hand operation is determined by a. Variable-speed fans and pumps 1. Status point not matching its ON/OFF point for 15 seconds when device is commanded OFF or 2. VFD in local hand mode or 3. Supervised HOA at control panel in ON position. b. Constant-speed fans and pumps 1. Status point not matching its ON/OFF point for 15 seconds when device is commanded off or 2. Supervised HOA at control panel in ON position. c. Chillers 1. Chiller is manually turned on as indicated by the status of the local/auto switch from chiller gateway. B. Chiller plant 1. Chillers shall be lead/lag alternated per Paragraph 1.01A.4. If a chiller is in alarm, its CHW isolation shall be closed. 2. Chillers are staged in part based on load calculated from thermal energy meter except for 15 minutes after a stage-up or stage-down transition, freeze-calculated load at the value at the initiation of the transition. This allows steady state to be achieved and ensures a minimum ON and OFF time before changing stages.


273

3. Staging shall be as follows. Timers shall reset to zero after every stage change. Each stage shall have a minimum runtime of 15 minutes (including Stage 0). Plant PLR is calculated load divided by total chiller design load as scheduled on drawings. Lockout temperature (LOT) shall be 60°F (adjustable). The chiller plant shall include an enabling schedule that allows operators to lock out the plant during off hours (e.g., to allow off-hour operation of HVAC systems, except the chiller plant; the default schedule shall be 24/7 [adjustable]). The staging part-load ratio shall be calculated every 5 minutes as SPLR = E(TCWR – TCHWS) + F where E = xx and F = ?? (see Appendix A.)

Stage

Chillers On

Nominal Capacity

Stage Up to Next Stage if Either

Stage Down to Lower Stage if

0

—

Any chiller plant requests and OAT > LOT and schedule is active

2

Lead chiller

50%

for 15 minutes PLR is greater than SPLR

CHW plant reset = 100 for 15 minutes and PLR greater than 30%

No chiller plant requests for 5 minutes or OAT < (LOT –5°F) or schedule is inactive

3

Both chillers

100%

—

—

for 15 minutes PLR less than SPLR

0

All off

—

4. Whenever there is a stage-up command: a. Command operating chillers to reduce demand to 50% of their current load. Wait until actual demand <55% up to a maximum of 5 minutes before proceeding. b. Slowly change the minimum bypass controller set point to that appropriate for the stage as indicated in Paragraph 1.01B.8 below. (If the bypass valve suddenly opens then the chiller load will suddenly drop and the chiller(s) could trip. Also, the DP control loop will be unstable). After new set point is achieved wait 1 minute to allow loop to stabilize. c. Start the next CW pump and when the CW isolation/head pressure control valve feedback indicates the valve is more than 10% open, wait 30 seconds. d. Slowly open CHW isolation valve of the chiller that is to be started. The purpose of opening slowly is to prevent sudden disruption to flow through active chillers. Valve timing to be determined in the field as that required to prevent nuisance trips.

274

Chapter 7 Controls e. Start the next stage chiller after CHW isolation valve is full open. f. Release the demand limit. 5. Whenever there is a stage-down command: a. Shut-off last stage chiller. b. When the controller of the chiller being shut off indicates no request for chilled-water flow, slowly close the chiller’s CHW isolation valve to avoid sudden change in flow through other operating chiller. c. When the CW isolation/head pressure control valve feedback indicates the valve is fully closed, shut off last-stage condenser water pump. d. Change the minimum bypass controller set point to that appropriate for the stage as indicated in Paragraph B.8 of Appendix B. 6. Condenser water pumps: a. Condenser water pumps shall be lead/lag alternated per Paragraph B.4 of Appendix B. b. See Paragraph B.4 of Appendix B and Paragraph B5 of Appendix B for ON/OFF staging sequence. 7. Chilled-water pumps: a. Chilled-water pumps shall be lead/lag alternated per Paragraph B.4 of Appendix B. b. CHW pumps shall be staged as a function of CHW flow ratio (CHWFR = actual flow divided by total plant design flow). The lead pump shall run whenever the plant is enabled. When CHWFR is above 47% for 10 minutes, start the lag pump. When CHWFR is below 47% for 15 minutes, shut off the lag pump. c. When any pump is proven on, pump speed will be controlled by a PID loop maintaining the DP signal at a set point determined by the reset scheme described below. All pumps receive the same speed signal. Minimum speed set point in VFDs shall be 10%. 8. Bypass valve: When any CHW pump is proven ON, the bypass valve shall be enabled, and opened otherwise. Bypass valve shall be modulated to maintain minimum flow as measured by the flowmeter. Minimum flow rates are as follows (based on manufacturers’ minimum flow rates plus 10% to ensure control variations do not cause flow to go below actual minimum):

Chiller Stage

Minimum Flow

0

0

1

xx

2

xx


275

9. Chilled-water supply temperature set point and pump differential static pressure set point shall be reset based on the figure below and the value CHW plant reset determined as described below. DPmax shall be determined under Section 230593 Testing, Adjusting and Balancing. Tmin is the design chilled-water temperature as scheduled on Drawings.

Figure 7-B

CHW plant reset.

a. CHW Plant Reset Shall be reset using trim and respond logic (see Paragraph 1.01A.3) with the following parameters:

Variable

Value

Device

Any CHW pump

SP0

0%

SPmin

0%

SPmax

100%

Td

15 min

T

5 min

I

2

R

Cooling CHWST reset requests

SPtrim

–2%

SPres

+3%

SPres-max

+7%

276

Chapter 7 Controls b. CHW plant reset logic shall be disabled and value fixed at its last value for 15 minutes after the plant stages up or down. 10. Cooling tower a. Fans are controlled off of CW return temperature (leaving chiller) rather than supply. Tower fans are enabled when any CW pump is proven on and CWRT rises above set point by 1°F. If CWRT drops below set point and fans have been at minimum fan speed for 5 minutes, fans shall cycle off for at least 3 minutes and until CWRT rises above set point by 1°F. b. Condenser water return temperature set point shall be CWRTsp = CHWST + LIFTx LIFTx = A · PLR + B where PLR is the plant PLR (actual chiller load divided by total plant design capacity), A =?? and B =?? (See Appendix A), but in no case shall LIFTx be less than the minimum lift at low load from the manufacturer (12°F) nor more than the design lift (45°F). c. Fan speed shall be modulated to maintain CWRT at set point. All operating fans receive the same speed signal. 11. Tower makeup water a. Makeup water valve shall cycle based on tower water fill level sensor. The valve shall open when water level falls below the minimum fill level recommended by the tower manufacturer. It shall close when the water level goes above the maximum level recommended by the tower manufacturer. 12. Performance Monitoring a. Total plant power. Calculate total plant power as the sum of chiller power, pump power, and cooling tower fan power. For motors with VFDs, power shall be actual power as indicated by the VFD. For fixed-speed motors (e.g., CW pumps), power shall be assumed to be fixed at brake power (from equipment schedule) · 0.746/0.93 (approximate motor efficiency). b. Summary Data. For each chiller and total plant, statistics shall be retained and displayed on graphic for runtime, average actual efficiency (kW/ton), and average demand (tons) and load (ton-hours). Show on chiller plant graphic: instantaneous values, year-to-date totals/averages and previous year totals/averages. 13. Alarms a. Maintenance interval alarm when pump has operated for more than 1500 hours: Level 5. Reset interval counter when alarm is acknowledged.


277

b. Maintenance interval alarm when chiller has operated for more than 1000 hours: Level 5. Reset interval counter when alarm is acknowledged. c. Chiller alarm: Level 2. d. Tower level: 1. If tower water level sensor indicates low water level, generate a Level 2 alarm. 2. If tower water level sensor indicates high water level, generate a Level 3 alarm. e. High chiller leaving chilled-water temperature (more than 5°F above set point) for more than 15 minutes when chiller has been enabled for longer than 15 minutes: Level 3. f. Pump or tower fan alarm is indicated by the status input being different from the output command after a period of 15 seconds after a change in output status. 1. Commanded on, status off: Level 2 2. Commanded off, status on: Level 4 g. Head pressure CW valve is partly closed and tower fan is on for more than 10 minutes: Level 3. h. Excessive CW approach indicating water side fouling: If leaving condenser water temperature is more than 3°F below refrigerant condensing temperature for 15 minutes at least 15 minutes after chiller start. i. Excessive CHW approach indicating water side fouling: If leaving chilled-water temperature is more than 3°F above refrigerant evaporator temperature for 15 minutes at least 15 minutes after chiller start.

References AHRI. 2015. AHRI Standard 550/590, 2015 Standard for performance rating of water-chilling and heat pump water-heating packages using the vapor compression cycle. Arlington, VA: Air-Conditioning, Heating, and Refrigeration Institute. ASHRAE. 2011. Fundamentals of HVAC control systems. Atlanta: ASHRAE. ASHRAE. 2012. ASHRAE Guideline 22-2012, Instrumentation for monitoring central chilled-water plant efficiency. Atlanta: ASHRAE. ASHRAE. 2016a. ANSI/ASHRAE/IES Standard 90.1-2016, Energy standard for buildings except low-rise residential buildings. Atlanta: ASHRAE. ASHRAE. 2016b. Chapter 47, Valves, ASHRAE Handbook—HVAC Systems and Equipment. Atlanta: ASHRAE. Edwards, T.J. 1983. Observations on the stability of thermistors. Review of Scientific Instruments 54:613; doi:10.1063/1.1137423. Hartman, T. 2005. Designing efficient systems with the equal marginal performance principle. ASHRAE Journal 47(7):64–70.

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Chapter 7 Controls Hydeman, M., and G. Zhou. 2007. Optimizing chilled water plant control. ASHRAE Journal 48:45–54. Hydeman, M., N. Webb, S. Blanc, and P. Sreedharan. 2002. Development and testing of a reformulated regression-based electric chiller model. ASHRAE Transactions 108(1). Lawton, K.M., and S.R. Patterson. 2002. Long-term relative stability of thermistors. Precision Engineering 26(3):340–45. Li, W., and R. Webb. 2001. Fouling characteristics of internal helical-rib roughness tubes using low-velocity cooling tower water. International Journal of Heat and Mass Transfer (6). NBCIP. 2004. Product testing report: Duct-mounted relative humidity transmitters. Iowa Energy Center, April 2004, National Building Controls Information Program. http://www.iowaenergycenter.org/wp-content/uploads /2017/02/PTR_Humidity_Rev.pdf. NBCIP. 2005. Product testing report supplement: Duct-mounted relative humidity transmitters. Iowa Energy Center, July 2005, National Building Controls Information Program. http://www.iowaenergycenter.org/wp -content/uploads/2017/02/NBCIP_S.pdf. Taylor, S. 2011. Optimizing design & control of chilled water plants: Part 2: Condenser water system design. ASHRAE Journal 9:14–25. Taylor, S. 2012. Optimizing design & control of chilled water plants: Part 5: Optimized control sequences. ASHRAE Journal 6:56–74. Taylor, Steve. 2015. Setpoint reset using trim & respond logic. ASHRAE Journal 57(6). Rishel, J.B. 2001. Wire-to-water efficiency of pumping systems. ASHRAE Journal 43(4).


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Which of the following statements regarding commercial versus industrial controls is not true: a. Commercial controls are not suitable for most CHW plant applications due to their inferior reliability. b. Industrial controls should be used for mission-critical applications requiring parallel operation of redundant controllers. c. Commercial control systems have lower first costs. d. Commercial control systems cost less to maintain.

7-2

You are in the process of specifying control sensors for a CHW plant used to serve both comfort cooling and mission-critical loads. The plant has variablespeed primary CHW pumps. Which of the follow options is most appropriate for monitoring CHW pump status? a. Hard wire status from the VFD. b. Read status over the network from the VFD. c. Install a current transducer on one phase on the line side of the VFD. d. Install a DP switch across the pump.

7-3

Which of the following statements regarding calibration of water temperature sensors is true? i. Field calibration is typically not necessary for noncritical sensors. ii. Field calibration reference temperature sensors should be at least twice as accurate as the field-installed sensors. iii. Dry well baths are always required to field calibrate the temperature sensors used for CHW plant load measurement.

a. b. c. d. 7-4

iv. RTDs are more stable than thermistors and therefore require less frequent repeat calibration. (i), (ii) (i), (ii), (iii), (iv) (i), (ii), (iv) (ii), (iv)

You are specifying two-way control valves for air-handler CHW coils served by a primary, variable secondary plant. The valve sizes vary between 1.5 and 4 in. What valve type(s) are most appropriate for this application? a. Globe valves for all sizes. b. Ball valves for 3 in. and less; globe valves for all greater than 3 in. c. Ball valves for all sizes. d. Ball valves for 3 in. and less; butterfly valves for all greater than 3 in.

280

Chapter 7 Controls 7-5

You are designing a variable-speed primary loop that controls pump speed based on DP measured by a DP sensor installed far out in the loop, remote from the mechanical room where the pumps are to be installed. A colleague recommends specifying that the DP sensor be wired to a controller located near the sensor, then passing the sensor reading to the pump controller over the network. This approach a. Can save on wiring costs relative to hardwiring the DP feedback to the pump controller. b. May result in control loop instability due to network latency or limited feedback polling frequency. c. Increases network traffic and may slow control system performance. d. Will lead to loss of pump feedback control if network communications between controllers are lost. e. All of the above.

7-6

Monitoring CHW plant efficiency using modern DDC control systems a. Typically requires installation of expensive true-RMS power meters. b. Typically requires installation of CHW loop instrumentation beyond that necessary for CHW plant control. c. Requires additional programming on the part of the DDC controls contractor. d. All of the above.

7-7

Plant optimization modeling indicates that the most efficient approach to control cooling tower fans serving a plant with typical office loads, short of using real-time optimization, is a. Controlling fans to maintain a condenser water supply temperature set point fixed to the minimum allowed by the chiller minimum lift requirement. b. Controlling fans to maintain a condenser water supply temperature set point fixed at the design temperature to minimize tower fan energy. c. Controlling fans to maintain a condenser water supply temperature reset as a function of ambient wet-bulb temperature. d. Controlling fans to maintain a condenser water return temperature set point reset based on plant load and CHW supply temperature.

7-8

Variable-speed control of condenser water pumps a. Can easily increase plant energy use if not controlled optimally. b. May be cost-effective in high load applications, such as data centers that require operation under a variety of load and ambient wet-bulb conditions, but modeling or real-time optimization is needed to determine the optimal sequence. c. Is highly climate dependent: the optimal strategy in one climate is unlikely to be optimal in another. d. All of the above.


281

The optimal approach to resetting CHW supply temperature set point and CHW DP set point for a plant with variable-speed chillers is: a. Dependent on climate. b. Dependent on plant load profile. c. Resetting CHW supply temperature set point first, then CHW DP set point. d. Resetting CHW DP set point first, then resetting CHW supply temperature set point.

Commissioning Instructions Read the material in Chapter 8. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the Skill Development Exercises without referring to the text. Review those sections of the chapter as needed to complete the exercises.

Introduction This chapter focuses on elements of the commissioning (Cx) process that are key to ensuring that chiller plants meet their design intent.

Commissioning Overview ASHRAE defines commissioning as “a systematic process of ensuring that systems are designed, installed, functionally tested, and capable of being operated and maintained to perform in conformity with the design intent” (ASHRAE 2007). Commissioning is intended to achieve the following objectives: • • • • •

Ensure that equipment and systems are properly installed and receive adequate operational checkout by the installing contractors. Verify and document proper operation and performance of equipment and systems. Ensure that the design intent for the project is met. Ensure that the project is thoroughly documented. Ensure that the facility operating staff is adequately trained.

Commissioning may be considered a quality control process that adds value by helping to meet the project’s design intent and achieve top performance of a CHW plant. This chapter is not intended to be a primer on the commissioning process and its costs and benefits. The cost-effectiveness of commissioning HVAC systems, in particular complex systems such as CHW plants, is well documented. Instead, this chapter focuses on key commissioning activities applied to CHW plants: • •

SOO review Point-to-point checkout

284

Chapter 8 Commissioning • •

Functional testing Trend review

Other commissioning activities, such as design and submittal reviews, operator training, and issues logging and resolution, are not addressed in detail. For further information on the commissioning process, refer to ASHRAE Guideline 1.1-2007, HVAC&R Technical Requirements for the Commissioning Process, and other ASHRAE materials on commissioning. This chapter also does not address who should serve as the commissioning authority (CxA), the person or group of people who coordinate and oversee Cx activities. There are debates about whether the engineer of record (EOR) (supported by project subcontractors) can successfully serve as the CxA or whether the CxA should be a third party, independent of the design and construction teams. The pluses and minuses of each approach are outside the scope of this SDL. The Cx activities addressed in this chapter apply well to either approach.

Commissioning Focus While following a comprehensive Cx process (as outlined in the many ASHRAE Cx documents and guidelines) will improve plant performance, it is important that the Cx process focus on the most likely source of plant underperformance: the control system. Occasionally, plant performance may be compromised by a chiller, pump, or tower that does operate as the manufacturer claims, but this is rare. Instead, when a plant has problems, the cause is almost always due to the controls, primarily due to improper control sequences and/or improper implementation of the sequences (BAS programming). The two most effective tools for identifying these control problems are as follows: •

•

Functional tests: Functional tests are intended primarily to ensure that the control sequences specified by the design engineer have been properly programmed into the BAS. The tests essentially mimic the control sequences by simulating an operating scenario and observing the BAS response. Trend reviews: Trend reviews consist of collecting data of actual system operation under normal conditions using the BAS trending capabilities then analyzing the data to see if the BAS is operating properly. In addition to complimenting functional testing to verify proper BAS programming, trend reviews can also identify flaws in the control sequences (e.g., when the system reacts inappropriately to an operating condition that was not anticipated in the control sequences).

Likewise, the two most effective tools for minimizing the potential for control system issues to arise during functional testing and trend review are as follows: •

SOO review: SOO review entails analyzing the specified sequence in detail to identify any logic flaws that may prevent the sequence from operating the plant without issue. Because the CxA will ultimately develop functional


•

285

tests to validate the sequence, this also provides an early opportunity to identify any unclear aspects of the written sequence that may ultimately impact the control system programmer’s ability to code to the design intent. Point-to-point checkout: As discussed in Chapter 7, the accuracy required of temperature sensors used for CHW plant control exceeds that required of most other commercial HVAC system components. As such, the point-topoint checkout process is particularly critical to ensure that (1) all points are mapped into the BAS properly and (2) all points are calibrated properly, with appropriate documentation, to ensure that proper plant control can be achieved.

Sequence of Operation (SOO) Review SOO review provides the CxA an opportunity to identify issues with plant control logic before it is implemented and tested in the field. The value of finding issues before they exist cannot be overstated: changing a few lines of text in the written SOO is much faster than (1) identifying an issue during functional testing or trend review, (2) getting sign off from the EOR on modifying the sequence, (3) having the controls contractor implement the sequence revision, (4) repeating functional tests for the revised portion of the sequence, (5) reviewing trends again, and (6) tracking the entire process through the issues log. If the EOR is acting as the CxA for the project, then a second set of eyes, either from within the EOR’s firm or a third party hired specifically for a peer review, should be engaged to complete a review of the SOOs. Key questions to keep in mind while reviewing the SOOs include: • • • • •

Does the plant controls system include all of the necessary hardware points to execute the sequence as specified? Is the logic too vaguely expressed to be executed as intended by a controls programmer? Are there aspects of the sequence that may yield control instability, equipment cycling, or damage? Can the logic be simplified in any way to achieve the expressed intent with less complication? Are there clear sections of the logic that will “work,” but unnecessarily waste energy while doing so? If so, how might they be amended?

The first question is typically easiest to answer and simply requires due diligence on the part of the CxA. For example, if the sequence for a variable primary/variable secondary plant calls for resetting the primary pump speed if the secondary flow rate exceeds the primary flow rate but does not include a secondary loop flowmeter, this is an easy red flag to the CxA. Is the design intent to include a secondary loop flowmeter, or should the SOO instead be modified to reset primary loop flow based on temperature sensors (e.g., the secondary and primary return temperature sensors)?

286

Chapter 8 Commissioning The second question is typically partially answered during an initial readthrough of the SOOs. Going through the thought process of how each section of the sequence should be functionally tested often further reveals aspects of the sequence that are unclear but upon initial inspection seemed fine. If there are aspects of the sequence that are unclear to the CxA, then the controls contractor tasked with implementing the sequence is likely to face the same problem. The third question requires more detailed inspection of the sequence and thought experiments to work through how the sequence will function in practice. For each control loop and each piece of staging logic called for in the sequence, the CxA should evaluate how it will work in the field. As an example, consider a sequence that calls for staging primary pumps based on flow rate. Does the sequence include minimum stage runtime delays to prevent short cycling, or does it include hysteresis on the stage-up/stage-down set points to achieve the same effect? If neither, then there is a problem. As another example, does the sequence stage chillers based on plant load but lack logic to reset the CWRT set point based on plant load as well? If so, the sequence may predispose the chillers to surge. The fourth question should be kept in mind at all times because the written sequence ultimately has be implemented in programming. The more complicated the programming, the more likely issues will reveal themselves during functional testing and trend review. The goal of answering this question is not to simplify the logic just for simplicity’s sake but to simplify it as much as possible while still achieving the design intent. The fifth question more commonly yields conflict between the CxA and the EOR than the others because it comes down to design preference, not operational integrity. However, if the sequence unnecessarily wastes energy by running equipment suboptimally, it is the CxA’s responsibility to bring these findings to the owner’s and EOR’s attention. Does the sequence call for staging towers even though minimum tower flow can be achieved with all towers running? If so, tower fan energy is wasted. Does the sequence call for always running as many primary CHW pumps as possible due to a misapplication of the affinity laws? If so, pump energy is wasted.

Point-to-Point Checkout The point-to-point checkout procedure for CHW plants is no different in implementation than the point-to-point checkout for any other commissioned HVAC equipment—it is just more critical that it is done correctly. To that end, the CxA should be thoroughly engaged in the process. Instead of just reviewing and cataloging point-to-point forms completed by the contractor, the CxA should be actively involved in crafting the forms to ensure that all necessary information is documented. At a minimum, point-to-point checkout forms should include the following:


•

•

•

•

287

Sensor calibration entries for all field-calibrated sensors. Recorded data should include reference instrumentation measured value, BAS connected sensor value, implemented calibration offset, and implemented calibration slope (as applicable). Binary input and output status verification. Regardless of whether an input device (e.g., an isolation valve end switch) is identified as normally open or normally closed in submittal data, the actual condition corresponding to a “hot” (energized) input should be verified on the point-to-point form. Similarly, for output points, the operational condition corresponding to a “hot” output should be recorded. Analog output configuration. Forms should include fields to record how controller outputs scale to controlled device outputs. This typically requires recording the operational state at the low and high ends of the control signal output range. Using VFD hardwired speed control as an example, a 10 V controller output may correspond to 6 Hz at the drive, while a 10 V output may correspond to 60 Hz at the drive. Alternatively, 0 V at the controller may correspond to 0 Hz at the drive. The data in this example scenario provide valuable indication to the future operator, and the CxA during testing and trend review, of whether minimum speed control is implemented at the drive or through the BAS. Analog input configuration. As with analog outputs, forms should include entries for how controller inputs from transducers scale to outputs from sensors. As an example, for a 0–20 psi DP sensor, the form should note the corresponding input range—likely 4 to 20 mA. For sensors that do not utilize transducers, such as thermocouples, the device type (e.g., Type K) should be noted for verification of correct output mapping by the BAS. Trend configuration. For each I/O point, forms should indicate whether a trend has been configured (Yes/No) and the trending interval or COV configuration as appropriate.

In addition to the above field-completed forms, the CxA should also prepare a list of required factory calibration certifications and NIST-traceable calibration certificates for all sensors called out as such in the controls specification. Point-to-point checkout should not be considered complete until the field forms are completed by the controls contractor and reviewed by the CxA and all required calibration certificates are in hand. The above data provide a configuration record for all hardwired components in the controls system. These data are useful to plant operators and control technicians tasked with modifying the plant controls in the future. Perhaps more importantly, requiring detailed point-to-point checkout documentation inherently requires the controls contractor to execute a detailed point-to-point checkout procedure. This in turn increases the probability that issues with faulty sensors, or improperly mapped inputs and outputs, are identified early in the Cx process before they become bigger issues during plant start-up, functional testing, and trend review.

288

Chapter 8 Commissioning

Functional Testing Role of Functional Testing Functional testing is intended to verify that the BAS has been properly programmed to execute the specified sequence of controls (Chapter 7). Tests should be performed after the plant has completed traditional start-up and testing, adjusting, and balancing (TAB) procedures and the BAS has been fully programmed. All plant control sequences and alarms should be tested. This is done by creating a test script that challenges the BAS with different operating scenarios and observing and recording the response. If the actual response matches the expected response, the test is successful. It is not at all uncommon for tests to be performed several times before all are completed successfully. Typically, Cx procedures require the BAS contractor to perform the tests on their own prior to CxA witness testing. The tests, or a subset of key tests, should also be conducted in the presence of plant operating engineers so that they can learn firsthand how the plant is expected to operate. Functional tests also serve a secondary role: they interpret the written control sequences. Because of the imprecise nature of written language, and perhaps the lack of detail common to many written sequences, the specifics of how the system is to respond to various operating scenarios is seldom precisely clear. The functional tests then can serve as a formal interpretation of the designer’s intent. It is with this added function in mind that thought should be given to how sequences will be tested during the initial SOO review, as noted above.

Creating Proper Testing Conditions Functionally testing chiller plants has an added complication that does not exist with most other mechanical systems: there may not be any actual load when tests are being performed. For example, servers have not been installed in a data center, the plant serves a speculative office building that does not yet have tenants, or testing is being conducted in the winter. Short of waiting until real loads exist, there are a few strategies that can be used, discussed in the following subsections.

Option 1: Simulated Loads Plant staging logic and set point reset logic are generally functions of load and flow, as discussed in Chapter 7. Design and variable-flow rates can generally be achieved without actual loads by opening coil control valves. (If there are no valves yet installed [e.g., if all coils are installed by future tenants], then the plant simply cannot be tested.) Varying load can then be simulated by varying flow by manually overriding pump speed and overriding supply and return temperatures. The chillers are locked out from actually starting at their control panels, so no real cooling is done, but the commands to start the chiller from the BAS can be observed and cor-


289

rect staging logic can be confirmed. Chiller status may also have to be overridden to ON when the chiller is commanded on to simulate proper operation. Chiller failures can be simulated by overriding status and/or leaving water temperatures. Pump start/stop and staging logic can also be readily tested in this way. The plant’s response to degrading T (Chapter 4) can also be simulated by overriding return CHW temperature to lower values. Cooling tower set point reset and low-load fan cycling can also be tested, but fan speed control to maintain set point cannot; that would have to wait for the trend review phase, discussed in a later section.

Option 2: Artificial Loads It may be possible to create loads on the CHW system using the heating system. Example strategies include the following: •

•

For a VAV system, cooling loads at AHU CHW coils can be created by resetting all zones to maximum airflow, resetting zonal discharge air temperature set points to maximum, locking out the outdoor air economizer, and setting the air-handler supply air temperature (SAT) set points low. Varying AHU SAT set points and manually modulating AHU fan speed can be used to achieve the desired load for testing. Loads on AHUs with preheat coils can be created by overriding the preheat coil valve to create a load on the CHW coil. Again, SAT set points and AHU fan speed can be adjusted to achieve the desired plant load.

Similar approaches for achieving simultaneous heating and cooling exist for most other system types as well, and are effective means of creating the load necessary for plant testing. In most cases, there is no need to physically interconnect the CHW and hot-water systems. Doing so can lead to complications due to varying system pressures or water treatment regimens, as well as possible equipment damage if entering water temperatures to chillers or boilers are too extreme.

Simulator Factory Witness Testing One effective means of increasing the probability that testing works the way it should in the field is to set up a plant simulator in the controls contractor’s facility. Simulator tests are run as part of a factory witness test procedure conducted before the sequence is ever implemented in the field. The simulator is constructed by loading the actual programming to be installed in the field on controllers in the laboratory and recreating the field network of controllers in the laboratory. Hardwired binary I/Os can be connected to relays such that command outputs and status feedback can be verified in the laboratory. Analog output commands can be verified in software as well, but understandably the 0–10 V and 4–20 mA outputs for all points cannot reasonably be wired in the laboratory. Analog inputs (temperatures, pressures, etc.) must be overridden to fixed values in software that are then adjusted throughout testing to represent desired field conditions.

290

Chapter 8 Commissioning The CxA is responsible for generating a factory witness test script that takes the simulated system through all operational scenarios the same way a field test script would. This process is necessarily complicated by the fact that the CxA must define the value of every hardwired analog input during each test step to ensure that the plant conditions being simulated represent the desired real-world conditions. The entire factory witness test process adds considerable overhead to Cx for both the CxA and the controls contractor that must be accounted for when budgeting. The CxA must complete another round of testing with a second script that usually takes longer to assemble than the actual field script, and the controls contractor must assemble a laboratory test bench and participate in additional testing. The benefit of all this additional work is that the control sequence is much more likely to work the first time in the field because the majority of critical flaws should be identified during factory testing. This has the dual benefits of minimizing the potential for realworld downtime resulting from flaws in the controls logic and reducing the amount of time spent testing and retesting conditions that may only present themselves occasionally in the field. Simulator factory witness testing is particularly invaluable for testing sequences that will be implemented in already live mission-critical facilities, such as data centers and laboratories, that cannot have any downtime. It generally does not make sense on less critical facilities such as commercial buildings.

Trend Review Trend review is an invaluable final step in the central plant Cx process because it • •

• •

validates operating conditions that were not observed during functional testing; validates logic that only can truly be observed when buildings are operating under real load conditions, such as optimum start and many set point reset algorithms; allows proper tuning and stability of control loops and reset blocks to be verified; and reveals control logic flaws due to conditions that were not anticipated when SOOs were created.

Trend Review Preparation Before a trend review can be properly executed, the groundwork for that review needs to be established. In particular, the CxA should provide the controls contractor with a list of desired trend review points during implementation of the sequence. This ensures that in addition to hardwired points, which are captured on the point-to-point forms discussed previously, software values of interest such as set points and trim and respond requests are preconfigured


291

for trending. (It is best if trending requirements are established in the bid specifications to avoid BAS contractor change orders; with some BASs, the time to set up trends can be significant.) If the EOR’s control specification does not clearly lay out a point-naming convention to be used by the control contractor, the CxA should work with the controls contractor to develop a point-naming standard to ease the identification of trend points when a trend database export is provided for trend review. In addition to helping with the near-term trend review, this step also helps make trend data more usable for engaged operators, contractors, and engineers that may complete future work in the building. An effective naming convention includes the following information: • • • •

Building name tag (if more than one apply) System type (e.g., CHW) Equipment or system component identifier (e.g., CH-1 or Primary loop) Property (e.g., CHWST)

The above point might be tagged as 1.CHW.CH1.CHWST, for example. With all points clearly named, and all points preconfigured upfront, the overhead involved with later adding missing points and deciphering and cleaning up poorly named trends can be avoided.

Trend Review Execution The first goal of the trend review process should be to assess whether the plant is achieving its basic intent of supplying chilled water at the temperature it is commanded to, when it should, and with limited fluctuation in control to set point. The second goal should be verifying the proper implementation of all control sequences in much the same way that was done during functional testing, albeit with an added focus on control loop stability and tuning. To ensure that no sequences go without review, an effective strategy is to step through the written sequence line by line and identify the type of graph (time series or scatter plot) and points that need to be included therein to validate each sequence. After the review is completed, the graphical results along with detailed explanations of identified issues and proposed resolutions should be provided to the owner in the form of a detailed report for review. The trend review graphs serve as the evidence that is then brought to bear when discussing proposed resolutions with the owner, EOR, and controls contracting team. The two most common types of graphs used to observe trends are timeseries plots and scatter plots. Time-series plots are often the easiest because almost all BASs have the capability built into their interface software. Several points can be plotted together, such as set point, controlled point, and loop output to the controlled device. Figure 8-1 is a time-series plot that shows pump speed as it modulates to control DP set point. There is the typical erratic control on start-up, but the loop eventually settles.

292

Chapter 8 Commissioning

Figure 8-1

Time-series plot.

Figure 8-2

Scatter plot.

The ability to create scatter plots is not always possible using BAS interface software. If not, the data need to be downloaded and plotted using a spreadsheet. Scatter plots are also useful in evaluating loop tuning by plotting the set point versus the controlled point. The desired result is a tight fit along a 45° line. Figure 8-2 is a scatter plot that shows a cooling tower fan system that is not well tuned.

References ASHRAE. 2007. Guideline 1.1-2007, HVAC&R technical requirements for the commissioning process. Atlanta: ASHRAE.


293


Which aspect of a central plant is most prone to issues during start-up and commissioning? a. Cooling towers b. Chillers c. Control system hardware d. Control system configuration and sequence implementation

8-2

You are reviewing a preliminary control sequence for a plant with two equally sized variable-speed centrifugal chillers and two cooling towers with the clauses below. What types of issues do you envision with this sequence? Cooling Towers: Control cooling towers to maintain a fixed condenser water supply temperature equal to the design temperature of 85°F. Stage towers optimally to minimize energy use. Chillers: Stage the lag chiller ON when plant load exceeds 30% of design plant capacity. Stage the lag chiller OFF when the plant load drops below 30% of design plant capacity. a. The sequence predisposes the chillers to surge. b. The sequence predisposes the chillers to cycling. c. The sequence for cooling tower staging is too vague to be implemented in practice. d. All of the above.

8-3

In which of the following scenarios is factory witness testing with a simulator appropriate? i.

Installation of a new plant in a new ten-story commercial office building. ii. Updating the sequences of an existing plant serving a data center. iii. Retrofit of an existing plant serving a college campus. iv. Installation of a new plant in a new laboratory facility. a. (i), (ii), (iii), (iv) b. (ii), (iii), (iv) c. (ii), (iii) d. (ii), (iv) 8-4

Which of the following are key benefits of functional testing? a. Test scripts serve as a formal interpretation of the written SOOs. b. Test script development provides an opportunity to identify unclear sequences that are likely to impact the controls contractor as well. c. Functional testing verifies implementation of the sequence consistent with the design intent. d. Functional testing decreases the probability that issues with the plant occur following commissioning due to controls programming and point-to-point configuration issues. e. All of the above.

294


Which of the following plant control issues are more likely to be identified in trend review than in functional testing? a. Unstable control loops b. Improperly implemented staging logic c. Improperly configured set point resets d. Proper response to equipment failure

Skill Development Exercises To receive full continuing education credit, all questions must be answered and submitted at www.ashrae.org/sdlonline. Please log in using your student ID number and the SDL number. Your student ID number is composed of the last five digits of your Social Security Number or another unique five-digit number you create when first registering online. The SDL number for this course can be located near the top of the copyright page of this book.


Total number of questions: 6 2-1

2-2

2-3

2-4

2-5

Why is the shape of a CHW plant’s cooling load profile a critical factor in plant design? a. It dictates the conditions under which the plant must operate efficiently to minimize energy costs. b. It impacts the selection of chillers because the plant must be able to handle the full range of expected load conditions stably. c. It drives the peak capacity required of the plant. d. Both (a) and (b). e. All of the above. Which of the following are true regarding the impact of air-side economizing on the annual load profile of a plant serving an office building? i. It reduces the total annual kWh-hours served by the plant. ii. It shifts the most common load percentage to a lower value. iii. It reduces the peak load of the plant. iv. It reduces the plant’s run hours. a. (i), (ii), (iv) b. (i), (ii) c. (i), (iii), (iv) d. (i), (ii), (iii), (iv) ASHRAE Handbook—Fundamentals supports which of the following load calculations methodologies? a. RTS b. HBM c. Transfer function method d. Only (a) and (b) e. All of the above Oversizing CHW plants: a. Typically yields more efficient pumping in variable-speed applications due to lower friction losses. b. Usually leads to more efficient chiller operation. c. May cause controllability issues if chillers are not properly selected for stable low-load operation. d. Is problematic when the condenser and CHW pumps are variable speed. e. Both (a) and (c). You are replacing oversized chillers in an existing CHW plant with modern direct digital control (DDC) controls, trending capabilities, and recently calibrated instrumentation. Which of the following is the recommended approach for determining peak load to size the new chillers?

Chapter 2 Skill Development Exercises

Skill Development Exercises for Chapter 2



2-6

a. Develop a load model of the facility using a simulation tool and utilize the peak load estimated therefrom. b. Develop a load model of the facility using a simulation tool and calibrate the model on an annual basis using utility billing data, then assess peak load with the model. c. Utilize the DDC system’s primary CHW loop flowmeter and supplyand return-temperature sensors to trend load. Use a few months of trended load and local weather data from the summer and/or swing seasons to develop a load profile and predict peak load therefrom. d. Install temporary National Institute of Standards and Technology (NIST)-calibrated instrumentation, including an ultrasonic flowmeter and supply- and return-temperature sensors to trend load. Use the same approach as option (c) to predict peak load. True or False: In early design, developing a prototypical model of the proposed building is usually too cost prohibitive to assist in plant design development. a. True b. False



In the vapor compression refrigeration cycle, refrigerant flows from the compressor discharge, then a. through the condenser, evaporator, expansion valve, and back to the compressor. b. through the condenser, expansion valve, evaporator, and back to the compressor. c. through the expansion valve, condenser, evaporator, and back to the compressor. d. through the evaporator, expansion valve, condenser, and back to the compressor.

3-2

Which of the following are valid reasons for a building owner to prefer selecting a chiller utilizing R-134a instead of R-123 refrigerant?

a. b. c. d.

i. ODP ii. Purge unit requirements iii. GWP iv. Future refrigerant availability (i), (ii) (i), (ii), (iv) (i), (ii), (iii), (iv) (iii), (iv)

3-3

You are selecting a 300 ton water-cooled chiller. The available compressor types in this size range include which of the following? a. Scroll, screw, and centrifugal b. Scroll and screw c. Screw and centrifugal d. Centrifugal

3-4

Manufacturers typically limit evaporator flow rates on the low end to avoid _________ and on the high end to avoid _________. a. Surge; noise b. Deposit formation; erosion c. Laminar flow; erosion d. Deposit formation; noise

3-5

Benefits of gear-driven centrifugal chillers relative to direct-drive centrifugal chillers include: a. Smaller physical footprint b. The more common incorporation of multiple stages, and thus refrigerant economizers c. Greater flexibility in optimizing compressor efficiency




Fundamentals of Design and Control of Central Chilled-Water Plants I-P d. All of the above e. (b) and (c) f. (a) and (c) 3-6

You are designing a 300 ton primary-only CHW plant with a 20°F CHW -T. The pump type with the best combination of efficiency and cost for this application is most likely to be a. Close coupled, end suction b. Base mounted, double suction c. Base mounted, end suction d. Close coupled, in-line

3-7

A customer expresses concern that the condenser water pumps in a plant located in San Diego, CA make a gurgling noise. The pumps are installed 3 ft below the basin water level. The strainers are located downstream of the pumps and the towers have vortex shedders. What is the most likely cause of the issue? a. Cavitation b. Dissolved air coming out of solution c. Vortices forming at the tower suction d. None of the above

3-8

You are designing a 700 ton central plant for a 20-story high-rise building. The cooling towers are to be located in a well on the roof where the primary constraint is roof area. Acoustics are not a concern. The type of tower best suited to this application is a. Field erected, induced draft, counterflow b. Packaged, induced draft, cross flow c. Packaged, induced draft, counterflow d. Packaged, forced draft, counterflow

3-9

A cooling tower has been selected for a 15°F range with a 6°F approach when the ambient wet-bulb temperature is 72°F. At design flow rate and 15°F range, but with an ambient wetbulb temperature of 62°F, the tower is capable of providing a. A better approach b. A worse approach c. The same approach

3-10

Pump and tower VFD minimum speed set points of 6 Hz or less a. Result in significant energy savings relative to a minimum speed set point of 12 Hz. b. Are not feasible, because they cause motors to overheat. c. Improve controllability under low-load situations. d. None of the above.



Aggressive CHW temperature reset a. Causes issues with thermal comfort because higher supply water temperature yields significantly higher supply air moisture content for a given set of entering air conditions and thus higher space humidity. b. Typically increases pump energy usage significantly enough to offset the benefit of the decrease in chiller energy use. c. Has little to no impact on space humidity control. d. Both (a) and (b).

4-2

A plant consists of two identical fixed-speed centrifugal chillers, each with a dedicated constant-speed primary CHW pump. The chillers supply chilled water to one large built-up air handler that primarily serves daytime commercial office space loads and another large air handler that serves an auditorium space most frequently occupied in the evening. Both air handlers have threeway CHW control valves and are of approximately equal size. Which of the following are true? i. The design will require two chillers to operate when the auditorium air handler is at full load, even if the office air handler is off. ii. The plant will operate least efficiently in the rare instances that both the office air handler and auditorium air handler are at near design load. iii. The plant will operate most efficiently in the rare instances that both the office air handler and auditorium air handler are at near design load.

a. b. c. d. 4-3

iv. Headering the primary pumps would increase the controls complexity with no benefit in system redundancy. (i), (iii), (iv) (i), (iii) (ii), (iv) (i), (ii)

For three-way valve systems a. The flow rate through the branch serving the coil is constant, irrespective of valve position. b. The flow rate through the branch serving the coil peaks when the valve is fully open to the coil. c. The flow rate through the branch serving the coil peaks when the valve is 50% open. d. Balancing of the bypass leg is never necessary.



Fundamentals of Design and Control of Central Chilled-Water Plants I-P The back-loaded position of the common leg shown below a. Causes one chiller to be almost fully loaded with the remainder of the load handled by the other chiller. b. Causes unbalanced flow through the two chillers. c. Results in the same energy performance as a system with a common leg located in the normal position just upstream of the secondary pumps. d. Is a reasonable location for most plants if it is less expensive due to the physical layout of the plant.

4-5

Construction of a variable primary CHW plant is just about complete when it is discovered that the design includes only two-way valves and no means to maintain minimum flow. Which of the following last-minute design change options will resolve the problem at minimum cost? a. Install a CHW bypass locally at the CHW plant. Measure DP across the chillers to indirectly measure flow. b. Install a CHW bypass locally at the CHW plant. Install a flowmeter in the main return line at the plant to measure flow. c. Install a CHW bypass at the end of the line. Install a flowmeter in the main return line at the plant to measure flow. d. Install enough three-way valves at end-of-line coils to maintain minimum flow.

4-6

True or false? Both headered and dedicated pump per chiller configurations are equally appropriate for plants requiring a standby pump. a. True b. False


4-4

Fundamentals of Design and Control of Central Chilled-Water Plants I-P Which balancing method is most appropriate for all but very large distribution systems? a. Manual balancing using CBVs to measure and adjust flow. b. No balancing. c. AFLVs at all coils. d. Reverse-return piping.

4-8

True or False: Air-side economizing systems can contribute to low CHW T issues in systems with high T designs. a. True b. False

4-9

What factors constrain the number of cooling towers that can be operated with a given number of constant-speed condenser water pumps enabled? a. Minimum per-tower flow requirements. b. Maximum per-tower flow requirements. c. Neither (a) nor (b). d. Both (a) and (b).

4-10

Isolating cooling towers by means of isolation valves on the tower supply piping only a. Requires that the equalizer be oversized. b. Does not require the equalizer to be oversized. c. Requires that all towers be operated whenever the plant is enabled. The isolation valves are only installed to prevent tower overflow upon plant shutdown. d. Is not a viable design option.

4-11

True or false? Most modern centrifugal chillers can operate with an integrated WSE without any means of head pressure control. a. True b. False


4-7



You are responsible for developing the master plan for a new college campus that will include multiple buildings. Which of the following CHW distribution approaches will be the most energy efficient? a. Variable primary flow b. Primary, variable secondary c. Primary, variable secondary with tertiary bridge connected pumps at the buildings d. Primary, distributed secondary where the secondary pumps are sized for the pressure drop from and back to the primary loop

5-2

Your customer is building a sprawling six-story, 600,000 ft2 building with three wings, each of which is served by a large AHU. The building also has multiple computer rooms served by FCUs. All of the high density rooms are localized in the same area of the building. What is the likely to be the most energy-efficient CHW distribution design approach? a. Variable primary with a plant bypass leg and two-way valves at all AHU and FCU coils b. Constant primary, variable secondary with two-way valves at all AHU and FCU coils c. Variable primary with distributed variable secondary coil pumps for all AHUs and a separate variable secondary pump serving all FCUs with two-way valves d. Variable primary with distributed variable secondary pumps at all AHUs and all FCUs

5-3

When sizing piping to minimize life-cycle costs, which of the following are critical considerations? a. Expected system operating hours b. Variable speed versus constant-speed operation c. Utility rates d. All of the above

5-4

Which of the following are benefits resulting from selecting condenser water loop T for a larger range? i.Reduced chiller energy use due to lower chiller lift ii.Reduced pumping energy use iii.Increased cooling tower capacity for a given tower selection iv.Reduced cooling tower scaling potential a. (ii.) and (iii.) b. (i.), (ii.) and (iii.) c. (ii.), (iii.), and (iv.) d. (i.), (ii.), and (iv.)





Which of the following are true when selecting cooling tower fan controls? a. Constant-speed towers often prove life-cycle cost-effective because of their low upfront costs. b. VFD fan speed controls are prohibitively expensive relative to the small energy benefit gained relative to two-speed motor controls. c. To achieve the energy benefits provided by variable-speed tower control, it is only necessary to install a VFD on the lead cooling tower. d. VFDs on all tower cells generally result in lowest life-cycle costs.

5-6

You are retrofitting an existing central plant serving a large office building with a WSE in a climate where freeze conditions are not a concern. The plant currently maintains minimum chiller head pressure in cool dry weather using a cooling tower bypass. How does this arrangement need to be modified to accommodate the WSE? a. No changes are needed. b. A separate cooling tower without condenser water bypass needs to be added to accommodate the WSE HX. c. Modulating actuators should be added to the condenser isolation valves on each chiller and the cooling tower bypass should be permanently shut. d. The WSE HX should be piped in series with the cooling towers and the cooling tower bypass should be shut permanently.



The benefits of a performance-based chiller bid approach include which of the following: a. It can minimize life-cycle costs for the owner. b. It solidifies selections in the design phase, allowing the designer to accurately lay out the plant and avoid design rework by the contractor during the construction phase. c. The chiller selection process becomes based on objective criteria rather than subjective criteria that contractors may use when including chillers in their project bids. d. All of the above.

6-2

Which of the following parameters is not appropriate to include in a performancebased bid approach: a. The anticipated cooling load profile for the plant. b. Total required plant capacity. c. Chiller efficiency metrics. d. Sound power limits.

6-3

The following three selections are provided by a chiller manufacturer.

Chiller 1

Chiller 2

Chiller 3

200 ton centrifugal Two-pass evaporator Open-drive compressor CHW pressure drop: 5 ft Full-load kW/ton: 0.5 Native BACNet Variable speed

200 ton screw Two-pass evaporator Hermetic compressor CHW pressure drop: 10 ft Full-load kW/ton: 0.6 Native BACNet Variable speed

200 ton centrifugal Three-pass evaporator Hermetic compressor CHW pressure drop: 15 ft Full-load kW/ton: 0.55 BACNet gateway required Variable speed

Which option is likely to have the highest first cost, not including the cost of the chiller itself? a. Chiller 1 b. Chiller 2 c. Chiller 3 d. Indeterminate based on the provided data 6-4

When estimating annual plant utility costs with simulation modeling a. Using a flat effective annual rate that accounts for both electricity and demand is usually sufficiently accurate. b. Annual escalation of utility costs should be ignored because of the unknown future variation of utility prices.




Fundamentals of Design and Control of Central Chilled-Water Plants I-P c. Using real rate schedules that account for both electricity and demand independently should be used. d. Annual escalation of utility costs should be estimated based on forecasts provided by DOE/EIA. 6-5

Using zero tolerance chiller performance data in bid forms a. Greatly increases the time needed to complete the forms by vendors. b. Better ensures the best chillers are selected, particularly for plants expected to operate many hours at low load. c. Is not recommended when using the simplified procurement procedure. d. All of the above.



Which of the following statements regarding commercial versus industrial controls is not true: a. Commercial controls are not suitable for most CHW plant applications due to their inferior reliability. b. Industrial controls should be used for mission-critical applications requiring parallel operation of redundant controllers. c. Commercial control systems have lower first costs. d. Commercial control systems cost less to maintain.

7-2

You are in the process of specifying control sensors for a CHW plant used to serve both comfort cooling and mission critical loads. The plant has variablespeed primary CHW pumps. Which of the follow options is most appropriate for monitoring CHW pump status? a. Hard wire status from the VFD. b. Read status over the network from the VFD. c. Install a current transducer on one phase on the line side of the VFD. d. Install a DP switch across the pump.

7-3

Which of the following statements regarding calibration of water temperature sensors is true? i. Field calibration is typically not necessary for noncritical sensors. ii. Field calibration reference temperature sensors should be at least twice as accurate as the field-installed sensors. iii. Dry well baths are always required to field calibrate the temperature sensors used for CHW plant load measurement.

a. b. c. d.

iv. RTDs are more stable than thermistors and therefore require less frequent repeat calibration. (i), (ii) (i), (ii), (iii), (iv) (i), (ii), (iv) (ii), (iv)

7-4

You are specifying two-way control valves for air-handler CHW coils served by a primary, variable secondary plant. The valve sizes vary between 1.5 and 4 in. What valve type(s) are most appropriate for this application? a. Globe valves for all sizes. b. Ball valves for 3 in. and less; globe valves for all greater than 3 in. c. Ball valves for all sizes. d. Ball valves for 3 in. and less; butterfly valves for all greater than 3 in.

7-5

You are designing a variable-speed primary loop that controls pump speed based on DP measured by a DP sensor installed far out in the loop, remote




Fundamentals of Design and Control of Central Chilled-Water Plants I-P from the mechanical room where the pumps are to be installed. A colleague recommends specifying that the DP sensor be wired to a controller located near the sensor, then passing the sensor reading to the pump controller over the network. This approach a. Can save on wiring costs relative to hardwiring the DP feedback to the pump controller. b. May result in control loop instability due to network latency or limited feedback polling frequency. c. Increases network traffic and may slow control system performance. d. Will lead to loss of pump feedback control if network communications between controllers are lost. e. All of the above. 7-6

Monitoring CHW plant efficiency using modern DDC control systems a. Typically requires installation of expensive true-RMS power meters. b. Typically requires installation of CHW loop instrumentation beyond that necessary for CHW plant control. c. Requires additional programming on the part of the DDC controls contractor. d. All of the above.

7-7

Plant optimization modeling indicates that the most efficient approach to control cooling tower fans serving a plant with typical office loads, short of using real-time optimization, is a. Controlling fans to maintain a condenser water supply temperature set point fixed to the minimum allowed by the chiller minimum lift requirement. b. Controlling fans to maintain a condenser water supply temperature set point fixed at the design temperature to minimize tower fan energy. c. Controlling fans to maintain a condenser water supply temperature reset as a function of ambient wet-bulb temperature. d. Controlling fans to maintain a condenser water return temperature set point reset based on plant load and CHW supply temperature.

7-8

Variable-speed control of condenser water pumps a. Can easily increase plant energy use if not controlled optimally. b. May be cost-effective in high load applications, such as data centers that require operation under a variety of load and ambient wet-bulb conditions, but modeling or real-time optimization is needed to determine the optimal sequence. c. Is highly climate dependent: the optimal strategy in one climate is unlikely to be optimal in another. d. All of the above.



The optimal approach to resetting CHW supply temperature set point and CHW DP set point for a plant with variable-speed chillers is: a. Dependent on climate. b. Dependent on plant load profile. c. Resetting CHW supply temperature set point first, then CHW DP set point. d. Resetting CHW DP set point first, then resetting CHW supply temperature set point.



Which aspect of a central plant is most prone to issues during start-up and commissioning? a. Cooling towers b. Chillers c. Control system hardware d. Control system configuration and sequence implementation

8-2

You are reviewing a preliminary control sequence for a plant with two equally sized variable-speed centrifugal chillers and two cooling towers with the clauses below. What types of issues do you envision with this sequence? Cooling Towers: Control cooling towers to maintain a fixed condenser water supply temperature equal to the design temperature of 85°F. Stage towers optimally to minimize energy use. Chillers: Stage the lag chiller ON when plant load exceeds 30% of design plant capacity. Stage the lag chiller OFF when the plant load drops below 30% of design plant capacity. a. The sequence predisposes the chillers to surge. b. The sequence predisposes the chillers to cycling. c. The sequence for cooling tower staging is too vague to be implemented in practice. d. All of the above.

8-3

In which of the following scenarios is factory witness testing with a simulator appropriate? i. ii. iii. iv. a. b. c. d.

8-4

Installation of a new plant in a new ten-story commercial office building. Updating the sequences of an existing plant serving a data center. Retrofit of an existing plant serving a college campus. Installation of a new plant in a new laboratory facility.

(i), (ii), (iii), (iv) (ii), (iii), (iv) (ii), (iii) (ii), (iv)

Which of the following are key benefits of functional testing? a. Test scripts serve as a formal interpretation of the written SOOs. b. Test script development provides an opportunity to identify unclear sequences that are likely to impact the controls contractor as well. c. Functional testing verifies implementation of the sequence consistent with the design intent.




Fundamentals of Design and Control of Central Chilled-Water Plants I-P d. Functional testing decreases the probability that issues with the plant occur following commissioning due to controls programming and point-to-point configuration issues. e. All of the above. 8-5

Which of the following plant control issues are more likely to be identified in trend review than in functional testing? a. Unstable control loops b. Improperly implemented staging logic c. Improperly configured set point resets d. Proper response to equipment failure

ASHRAE LEARNING INSTITUTE Self-Directed Learning Course Evaluation Form Course Title: Fundamentals of Design and Control of Central Chilled-Water Plants (I-P) (2017) On a scale of 1 to 5, circle the number that corresponds to your feeling about the statements below. (1 = strongly agree, 5 = strongly disagree, 3 = undecided) Strongly Agree

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Flexible and Effective Continuing Education for HVAC&R Professionals

ASHRAE’s Fundamentals of Design and Control of Central Chilled-Water Plants, I-P Edition, self-directed learning course book focuses on optimizing the design and control of chilled-water plants to minimize lifecycle costs. This work is an invaluable tool for HVAC designers of various backgrounds and an introduction for those new to chilled-water plants. Plant operators, energy engineers, and control system designers will also find information on loads, equipment, distribution, chiller procurement, controls, and commissioning. Supplemental online files are included to promote understanding of real-world scenarios. Skill Development Exercises at the end of each chapter help readers assess their understanding of the material and apply what they learn to real-world situations. Answers to these exercises can be submitted online to earn PDH or LU credits.

1791 Tullie Circle Atlanta, GA 30329-2305 Telephone: 404/636-8400 Fax: 404/321-5478 E-mail: [email protected] www.ashrae.org/ali ISBN 978-1-939200-66-2 (paperback) ISBN 978-1-939200-67-9 (PDF)

9 781939 200662

Product Code: 98014

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