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This is the second of a series of articles discussing how to optimize the design and control of chilled water plants. The series will summarize ASHRAEs Self Directed Learning (SDL) course called Fundamentals of

Design and Control of Central Chilled Water Plants and the research

that was performed to support its development. See sidebar, Page 36

for a summary of the topics to be discussed. The articles, and the SDL

course upon which it is based, are intended to provide techniques for

plant design and control that require little or no added engineering

time compared to standard practice but at the same time result in sig-

nificantly reduced plant life-cycle costs.

A procedure was developed to provide near-optimum plant design for most chill-er plants including the following steps:

1. Select chilled water distribution system.

2. Select chilled water temperatures, flow rate, and primary pipe sizes.

3. Select condenser water distribution system.

4. Select condenser water tempera-tures, flow rate, and primary pipe sizes.

5. Select cooling tower type, speed con-trol option, efficiency, approach tempera-ture, and make cooling tower selection.

6. Select chillers.7. Finalize piping system design, calcu-

late pump head, and select pumps. 8. Develop and optimize control se-

quences.Each of these steps is discussed in this

series of five articles. This article dis-cusses Step 3: designing the condenser water distribution system. Steps 2 and 4 will be discussed in the next article.

Three common piping arrangements for condenser water pumps are:

Option A: Dedicate a pump for each condenser (Figure 1a);

Option B: Provide a common header at the pump discharge and two-way au-tomatic isolation valves for each con-denser (Figure 1b); and

Option C: Provide a common head-er with normally closed (NC) manual isolation valves in the header between pumps (Figure 1c).

The advantages of dedicated pumps for each condenser (Option A) include:

About the AuthorSteven T. Taylor, P.E., is a principal at Taylor Engineering in Alameda, Calif.

By Steven T. Taylor, P.E., Fellow ASHRAE

Optimizing Design & ControlOf Chilled Water Plants Part 2: Condenser Water System Design

This article was published in ASHRAE Journal, September 2011. Copyright 2011 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. Posted at www.ashrae.org. This article may not be copied and/or distributed electronically or in paper form without permission of ASHRAE. For more information about ASHRAE Journal, visit www.ashrae.org.

Sep t ember 2011 ASHRAE Jou rna l 27

1. The pump can be custom-selected for the condenser it serves. Pump selection can then account for variations in condenser pressure drop and flow rates when chillers are not identical. This can reduce pump energy compared to Option B where the head of each pump must be the same and sized for the condenser with the highest pressure drop; balance valves at the other condensers must be throttled to generate this same pressure drop.

2. Controls are a bit simpler because the pump can be con-trolled using the contact provided with the chiller controller. This ensures that the pump starts and stops when the chiller wants it to. With Option B, the control of the isolation valves and pumps is by the direct digital control (DDC) system and must be coordinated with the needs of the chiller controller to avoid nuisance trips. For instance, the pumps generally must run for several minutes after the command for the chiller to stop so that the chiller can pump down the refrigerant.

3. Pump failures do not cause multiple chiller trips. With dedicated pumps, if a pump fails, only the chiller it serves will see a flow disruption and trip. With Option B, all operating chillers will see a flow reduction when a pump fails, possibly causing more than one chiller to trip due to low flow or high refrigerant head. However if there is a lag or standby pump with Option B that can be started quickly, trips can usually be avoided because it takes some time for refrigerant head to rise.

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

1. Redundancy is improved. With Option A, if a pump fails and a chiller other than the one it serves also fails (albeit a rare event), then two chillers will be inoperative. With Option B, any pump can serve any chiller and under many conditions one pump can provide enough flow for two chillers to operate near full capacity.

2. Including a standby pump is much simpler. Adding a standby pump to Option A is cumbersome and expensive because it requires extensive piping and manual or automatic isolation valves. If standby pumps are desired, Option B is the best option.

3. Isolation valves can double as head pressure control valves. See discussion on head pressure control later. For Option A, head pressure control would require the addition of variable speed drives on condenser water pumps or tower bypass valves.

4. It is easier to integrate a water-side economizer. See discussion on waterside economizers below. Since waterside economizers are only operational in cold weather when loads are generally low, the condenser water side can use one (or more) of the condenser water pumps serving chillers rather than providing a dedicated pump. This reduces first costs.

Headered pumps with manual isolation valves (Option C) can have the advantages of Option A (although it works best with identical chillers) and it overcomes the redundancy dis-advantage of Option A but accommodating a pump failure requires manual manipulation of valves vs. the automatic response in Option B. Including a standby pump is possible with Option C but it only works (depending on which pump fails) with the header isolation valves open and chillers must be staged by manually opening and closing their isolation valves.

First costs are usually lowest with Option A if the chiller and pump pairs are close-coupled and the manual isolation valves between the two are eliminated (each chiller-pump pair is iso-lated for service as a pair). Option C is usually less expensive than Option B, but Option B is usually the best choice where head pressure control and standby pumps are required.

Refrigerant Head Pressure ControlAll chillers will require a minimum refrigerant head (lift)

between the evaporator and condenser. This can be quite high

Figure 1: Condenser water pump piping options. Option A (left): Dedicated pumps. Option B (center): Headered pumps with con-denser auto-isolation valves. Option C (right): Headered pumps with manual isolation valves.

Cooling Tower No. 1

Cooling Tower No. 2

Cooling Tower No. 3

Chiller No. 1

Chiller No. 2

Chiller No. 3

CHW Pump No. 1

CHW Pump No. 2

CHW Pump No. 3

Cooling Tower No. 1

Cooling Tower No. 2

Cooling Tower No. 3

Chiller No. 1

Chiller No. 2

Chiller No. 3

CHW Pump No. 1

CHW Pump No. 2

CHW Pump No. 3

Optional Standby Pump

Cooling Tower No. 1

Cooling Tower No. 2

Cooling Tower No. 3

Chiller No. 1

Chiller No. 2

Chiller No. 3

CHW Pump No. 1

CHW Pump No. 2

CHW Pump No. 3

N.C.

N.C.

A B C

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for most screw chillers and some hermetic centrifugal chill-ers, and very low for magnetic bearing chillers, which have no oil return considerations. There are two common reasons why low refrigerant head pressure can occur:

At start-up when water temperature in the cooling tower basins is cold. Some chillers can operate for a short period of time with low start-up head while others will trip on low head pressure safeties almost immediately. To determine if head pressure control is required, for cold starts, consult with the chiller manufacturer.

When integrated waterside economizers are used (dis-cussed later). Head pressure control is almost always manda-tory since cooling tower water temperatures are deliberately kept very cold for long periods.

Options to avoid low head pressure problem include: Tower three-way bypass valves. The bypass water is di-

verted around the tower fill into the cooling tower sump or into the suction piping, thus avoiding natural cooling that oc-curs across the tower fill even when tower fans are off. Piping the bypass to the suction line also avoids the mass of water in the basin for an even faster warm-up, but the design can be problematic: unless the bypass line is balanced to create a pressure drop equal to the height of the cooling tower, air will be drawn into the system backwards from the spray nozzles since piping above the basin will fall below atmospheric pres-sure. For staged or variable condenser water flow systems, the bypass must be balanced at the lowest expected flow rate. This creates a high pressure drop and reduced flow if more pumps operate, but reduced flow is acceptable when the intent of the bypass is to raise head pressure. The bypass valve is controlled by supply water temperature typically with a low limit setpoint well below the normal setpoint used to control tower fan on/off and speed. Tower bypass is most commonly used where towers must operate in very cold weather to avoid freezing in the fill. The following two options are less expensive and, therefore, preferred in other applications.

For systems with dedicated condenser water pumps (Op-tion A or C, Figure 1), variable speed drives on the pumps can be used to reduce water flow to the chiller. Head pressure can be maintained even with very cold supply water as long as the flow rate can be reduced so that the condenser refrigerant pressure can be high enough (head pressure depends on the

condenser water temperature leaving the chiller, not entering the chiller). Pump speed can be controlled by the temperature leaving the condenser at a setpoint that corresponds to mini-mum condenser pressure, or (preferably) by a signal from the chiller controller indicating head pressure needs; most chiller controllers have an analog output dedicated for this purpose.

For systems with headered pumps (Option B, Figure 1), the isolation valves can double as head pressure control valves by converting them from two-position to modulating. Valve position is typically controlled by the chiller controller head pressure con-trol analog output, either directly or through the DDC system. This signal will close the valve when the chiller shuts off.

The second two options mentioned previously reduce flow through the condenser. Many engineers are concerned that low condenser water flow will contribute to fouling of the con-denser tubes, but there is little definitive evidence to support the concept that high velocity keeps tubes clean; strainers and sidestream filters that prevent particles from entering the con-denser in the first place are preferred. But even if this is an is-sue, for most head pressure control applications there are few hours at reduced flowonly during cold startsso the impact on tube fouling should not be significant. Low flow through the cooling tower may also be an issue (see discussion later) but, again, it should not be given the short duration.

Minimum Flow RatesWhen water enters the cooling tower, it is distributed uni-

formly across the fill through spray nozzles via a piping head-er or gravity distribution basin. Each cell has a minimum flow rate to ensure that tower fill is fully wetted along the face of the air entering the fill. If there are dry spots along this face, air will bypass the wetted fill due to lower pressure drop and, more importantly, cause scale to build up at the boundary be-tween the wet and dry fill as water is evaporated and dissolved solids remain. So it is important to maintain minimum tower cell flow rates, particularly in areas with hard makeup water.

In plants with multiple cooling towers and chillers, it is desirable to operate one condenser water pump at low loads, which will reduce the flow rate through cooling towers. Op-tions for maintaining minimum flow rates (Figure 2) include:

Option A: Select tower weir dams and/or nozzles to allow one pump to serve all towers. For systems with two or three

Figure 2: Cooling tower cell isolation options. Option A (left): Weir dams and/or low flow nozzles. Option B (center): Auto-isolation valves on supply only. Option C (right): Auto-isolation valves on supply and suction.

Cooling Tower No. 1

Cooling Tower No. 2

Cooling Tower No. 3

Cooling Tower No. 1

Cooling Tower No. 2

Cooling Tower No. 3

Cooling Tower No. 1

Cooling Tower No. 2

Cooling Tower No. 3

A B C

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Sep t ember 2011 ASHRAE Jou rna l 29

tower cells, this can eliminate the need for isolation valves, which cost much more than the weir dams and nozzles. This option is also the most efficient; tower energy use is mini-mized by operating as many cells as possible, particularly when tower fans are controlled by variable speed drives. This is because fan speed is reduced (reducing fan power by almost the cube of the speed) and cooling is achieved through tower cells even when fans are off. With most man-ufacturers and tower types, nozzles and dams are available to reduce flow by 50%, and many can go down to 33% or even 25% depending on the selection and design flow rate. Because of low cost and high efficiency, this option should always be the first choice. When a plant has many tower cells and automatic isolation valves are unavoidable, the dams and nozzles should still be selected to allow as many cells to operate as possible.

Option B: Install automatic isolation valves on supply lines only. This option uses the equalizer to keep basin levels between overflow and fill lines and will require that equal-izers be oversized from that required by normal duty. For example, assume there are three tower cells, and only one is active; supply flow to the others is shut off. But water is drawn out of all three cell basins since the suction lines have no automatic isolation valves. The water level in the basin

of the cell that is supplied will rise while the other two ba-sin levels will fall. The difference in the two elevations must provide enough head for water to transfer from the supplied cell to the others through the equalizer. If the equalizer is undersized, water will overflow in the supplied cell, and the others will be drawn so low that makeup water valves open, wasting water and water treatment chemicals. There are only a few inches of elevation difference between the overflow and fill lines, so it is imperative that the equalizer be properly sized for this option to work. Another approach is to elimi-nate the basins at each tower and use a common sump, often located indoors in cold climates. This avoids the need for equalizer lines entirely but is much more expensive.

Option C: Install automatic isolation valves on both sup-ply and suction lines. This is usually the most expensive option since automatic valves are expensive relative to an incremental increase in equalizer size. This design also in-creases exposure to a valve failure; an oversized equalizer line has no failure modes. It also increases the risk of freez-ing (or increases the energy used by basin heaters) in the basins of inactive cells in systems that must operate in cold weather. But this is often the best option when there are many tower cells that are not located close together (long equalizer lines).

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stead, waterside economizers must use an integrated piping arrangement shown in Figure 4 for a primary-secondary system and Figure 5 for a primary-only system. Integrated systems, which cost only slightly more than non-integrated

Piping for Waterside EconomizersWaterside economizers are an alternative to airside econ-

omizers. Airside economizers are usually more energy ef-ficient, but they are not always practical and can be much more expensive. Applications where waterside economiz-ers are often preferred include floor-by-floor air handlers in a high-rise office building or computer room air handlers serving a large data center. A waterside economizer uses cold water generated at the cooling tower to produce chilled water without, or with reduced, mechanical refrigeration. This is accomplished by running the cooling towers to produce water temperatures typically 45F (7C) and less during periods of low ambient wet-bulb temperatures. The cold water is pumped through a high effectiveness water-to-water heat exchanger, usually a plate and frame type, to produce chilled water at temperatures of 50F (10C) or less. The heat exchanger protects the chilled water system from the corrosion, dirt and debris typical of open circuit condenser water.

For detailed design guidance on sizing waterside economiz-er heat exchangers and flow rates, see Stein.1

Figure 3 shows a non-integrated waterside economizer where the economizer heat exchanger is piped in parallel with the chiller evaporators on the chilled water side. This design allows the economizer to operate only if the chill-ers are not operating and vice versa; they cannot operate together. This design was the most common when water-side economizers first became popular in the 80s, but it is not very efficient and is no longer allowed to be used by energy standards such as ASHRAE Standard 90.1.2 In-

Figure 3: Waterside economizer, non-integrated.

Cooling Tower No. 1

Cooling Tower No. 2

Chiller No. 1

Chiller No. 2

Plate and Frame Heat Exchanger

Figure 4: Waterside economizer, integrated, primary-secondary.

Figure 5: Waterside economizer, integrated, primary-only.

Cooling Tower No. 1

Cooling Tower No. 2

Chiller No. 1

Chiller No. 2

Plate and Frame Heat Exchanger

Cooling Tower No. 1

Cooling Tower No. 2

Chiller No. 1

Chiller No. 2

Plate and Frame Heat Exchanger

Either Pump Or Valve

(Not Both)

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32 A SHRA E Jou rna l S e p t e m b e r 2 0 1 1

systems, allow simultaneous operation of the chillers and the economizer because the heat exchanger is piped in se-ries with the chiller evaporators on the chilled water side. The economizer can provide some pre-cooling of the return chilled water temperature even if it cannot provide all of the cooling. This substantially extends the number of hours the economizer can be operational.

Figure 4 shows two options for how to provide flow through the heat exchanger. The least expensive option is to place a two-position valve in the chilled water return line. The valve closes when the economizer is enabled and is open otherwise. This option requires that secondary pumps have variable speed drives so that they can slow down when the heat exchanger is out of the circuit and vice versa. The secondary pumps generally do not need to be sized for the added head of the heat exchanger since the heat exchanger will be in the loop only when the economizer is active and cooling loads (and flows) are low. If secondary pumps are constant speed (rarely true in modern plants) or if the design flow rate through the heat exchanger is much lower than the expected chilled water flow during economizer operation, a sidestream pump should be used instead of the two-position valve. This sidestream pump is sized with enough head to

Figure 6: All-variable speed primary-only chilled water plant.

VSD

Cooling Tower No. 1

Chiller No. 1

Chiller No. 2

Cooling Tower No. 2

VSD

VSD

VSD

VSD

VSDVSD

VSD

draw water out of the secondary return, pump it through the heat exchanger then back to the return.

In both the integrated and non-integrated designs, the heat exchanger is generally not provided with its own con-denser water pumps. Since the load will be low when the weather is cold enough for the towers to deliver cold water, it should not be necessary to run all chillers, so one or more of the chiller condenser water pumps can serve the heat ex-changer. The heat exchanger should be selected so that its pressure drop is similar to the pressure drop across chiller condensers.

When using waterside economizers, refrigerant head pres-sure control is required because of the cold water coming off the cooling tower. See the earlier discussion regarding head pressure control options.

Variable Speed Condenser Water PumpsWith the ever-increasing drive to improve plant effi-

ciency, there is more interest in all-variable speed chilled water plants,3 which refers to plants with variable speed drives on all components, including condenser water pumps (Figure 6). It is common to find variable speed drives on cooling towers and chilled water pumps and, in fact, they are required with few exceptions by energy stan-dards such as Standard 90.1. Using variable speed drives on

34 A SHRA E Jou rna l ash rae .o rg S e p t e m b e r 2 0 1 1

chiller compressors is also more and more common as the cost premium vs. fixed speed continues to fall. But vari-able speed drives on condenser water pumps are relatively rare, and for good reason: it is not clear that they are cost effective and the required control logic is not self-apparent. For instance, as condenser water flow falls, both pump energy and cooling tower energy (for the same condenser water supply tem-perature) will fall, but chiller energy will rise as leaving condenser water temperature rises. The condenser leav-ing water temperature is indicative of chiller condensing temperature and, therefore, chiller efficiency; efficiency will vary little with changes to con-denser supply water temperature at the same leaving water temperature. With variable speed drives on the chiller compressor, the impact of condenser

Figure 7: Denver chilled water plant energy use using three control strategies.

1.4 Million

1.2 Million

1 Million

800,000

600,000

400,000

200,000

0

Ann

ual C

hill

ed W

ater

Pla

nt E

nerg

y U

se (

kWh

)

Chiller Tower CHWP CWP Plant Total

TOPPSTDOAK

temperature is even stronger and, in fact, these drives will save no energy at all if leaving condenser water tempera-tures are not driven down at low loads.

Clearly the optimum control logic will not be the same for all plants. For instance, a plant with very efficient (high gpm/horsepower) cooling towers will operate more effi-ciently by driving condenser water temperatures down fur-ther than a plant with inefficient towers. So what is the best control strategy? The answer is it depends. A few authors have proposed theoretical approaches to determining the optimum logic,4,5 but the techniques are either difficult and time consuming to implement or require proprietary con-trol logic.

As part of the development of the ASHRAE SDL refer-enced earlier, studies were conducted to develop generalized optimum control sequences for all-variable speed plants and to determine life-cycle costs of various design alternatives. Our studies led to two important conclusions about variable speed drives on condenser water pumps:

1. They are life-cycle cost effective if optimum control se-quences are used.

2. They can increase the energy use of the plant if not opti-mally controlled.

The second conclusion is disturbing, in particular, because we found that the difference was very subtle between the con-trol logic that minimized energy use and that which increased use above constant speed pumps. For example, Figure 7 shows energy use for a plant serving an office building in Denver, using three control strategies:

TOPP. This is the theoretical optimum plant performance of the plant with variable speed condenser water pumps deter-mined using the technique described in Reference 5. This is the theoretical best performance possible.

STD. This is the performance of the plant with constant speed condenser water pumps and cooling tower fans con-trolled to reset supply water temperature per ARI Standard 550/5906 condenser water relief curves. This is most indica-tive of conventional practice.

OAK. This is the performance of the plant with variable speed pumps controlled using control sequences that were found to provide near-TOPP level performance for the same plant located in Oakland, Calif., instead of Denver.

The figure shows that energy use is highest using control sequences that provided near-ideal performance for the same plant in another climate zone, significantly higher energy use than the plant without pump variable speed drives. This dem-onstrates how sensitive plant performance is to the details of the control logic. So, variable speed drives should only be used on condenser water pumps if the designer takes the time to en-sure that control sequences are near-optimum. These sequenc-es will be discussed in detail in Part 5 of this series of articles.

SummaryThis article is the second in a series of five that summarize

chilled water plant design techniques intended to help engi-neers optimize plant design and control with little or no added engineering effort. In this article, condenser water system pip-ing and distribution system options were discussed. In the next article pipe sizing and optimizing T will be addressed.

References1. Stein, J. 2009. Waterside Economizing in Data Centers:

Design and Control Considerations. ASHRAE Transactions 115(2):192 200.

2. ANSI/ASHRAE/IES Standard 90.1-2010, Energy Standard for Buildings Except Low-Rise Residential Buildings.

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36 A SHRA E Jou rna l S e p t e m b e r 2 0 1 1

This series of articles will summarize the upcoming Self Directed Learning (SDL) course called Fundamentals of Design and Control of Central Chilled Water Plants and the research that was performed to support its develop-ment. The series includes five segments. Part One: Chilled Water Distribution System Selection was published in July.

Pipe sizing and optimizing T. This article will discuss how to size piping using life-cycle costs then how to use pipe sizing to drive the selection of chilled water and con-denser water temperature differences (Ts).

Chillers and cooling tower selection. This article will address how to select chillers using performance bids and how to select cooling tower type, control devices, tower efficiency, and wet-bulb approach.

Central Chilled Water Plants Series Optimized control sequences. The series will conclude with a discussion of how to optimally control chilled water plants, focusing on all-variable speed plants.

The intent of the SDL (and these articles) is to provide simple yet accurate advice to help designers and oper-ators of chilled water plants to optimize life-cycle costs without having to perform rigorous and expensive life-cycle cost analyses for every plant. In preparing the SDL, a significant amount of simulation, cost estimating, and life-cycle cost analysis was performed on the most common water-cooled plant configurations to determine how best to design and control them. The result is a set of improved design parameters and techniques that will provide much higher performing chilled water plants than common rules-of-thumb and standard practice.

3. Hartman, T. 2001. All-Variable Speed Centrifugal Chiller Plants. ASHRAE Journal 43(9):43 51.

4. Hartman, T. 2005. Designing Efficient Systems with the Equal Marginal Performance Principle. ASHRAE Journal 47(7):64 70.

5. Hydeman, M., G. Zhou. 2007. Optimizing Chilled Water Plant Control. ASHRAE Journal 49(6):44 54.

6. ARI Standard 550/590-2003, Performance Rating of Water Chill-ing Packages Using the Vapor Compression Cycle.