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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9/14/2017 10:43:25 AM

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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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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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.

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

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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.

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

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

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

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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-

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

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.

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

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.

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

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

Instructions Read the material in Chapter 3. 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 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.

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

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

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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.

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

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

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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.

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

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.

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

•

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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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Chapter 3 Chilled-Water Plant Equipment

Figure 3-8

Single-effect absorption.

Figure 3-9

Double-effect absorption.

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29

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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Chapter 3 Chilled-Water Plant Equipment

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

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

Figure 3-11

31

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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Chapter 3 Chilled-Water Plant Equipment

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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Chapter 3 Chilled-Water Plant Equipment

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

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•

•

•

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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.

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•

•

•

•

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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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•

•

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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.

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

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.

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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 12 in.

Measurement Maintenance Accuracy Costs

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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

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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

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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%

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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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• •

•

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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.

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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.

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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

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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-

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•

•

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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.

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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.

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

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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.

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Figure 7-8

243

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

Source: Taylor 2012.

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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Figure 7-9

TOPP (CWRT – CHWST) versus plant load ratio.

Source: Taylor 2012.

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.

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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

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(a)

(b)

Figure 7-10

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

Source: Taylor 2012.

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.)

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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

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

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).

250

Chapter 7 Controls

Figure 7-12

TOPP optimum percent CW loop flow versus percent plant load.

Source: Taylor 2012.

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

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

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

252

Chapter 7 Controls

Figure 7-14

Denver CHW plant energy usage using three control strategies.

Source: Taylor 2011.

Figure 7-15

Miami CHW plant energy usage using three control strategies.

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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

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

Figure 7-16

255

Plant energy with CHWST set point reset, CW DP set point reset, and a combination of the two.

Source: Taylor 2012.

Figure 7-17

CHWST set point reset sequenced with CW DP set point reset off of CHW valves.

Source: Taylor 2012.

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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

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

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

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

Figure 7-20

TOPP variable-speed chiller staging versus plant load ratio (Albuquerque).

Source: Taylor 2012.

Figure 7-21

Two-chiller plant part-load performance with and without VFDs.

259

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Chapter 7 Controls

Figure 7-22

Possible surge problem staging by load only.

Source: Taylor 2012.

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.

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

Figure 7-23

Optimum staging versus (CWRT– CHWST) and plant PLR.

Source: Taylor 2012.

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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.

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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.

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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.

Source: Taylor 2012.

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.

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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

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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.

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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