56 ASHRAE Journal ashrae.org June 2012 T his is the last of a series of articles discussing how to optimize the design and control of chilled water plants. The series summarizes ASHRAE’s Self Directed Learning (SDL) course called Fundamentals of Design and Control of Central Chilled Water Plantsand the research that was performed to support its development. 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 engineer- ing time compared to standard practice but at the same time result in signicantly reduced plant l ife-cycle costs. A procedu re was de veloped to provide ne ar-optimum pl an tde si gnfor m ostchil l- e r pl an ts including th e following ste ps: 1. Se l e ct chil led water distribu tion system. 2. Select chilled wa ter tem pera tures, fl ow rate, a nd prim ary pipe sizes. 3. Sel ect cond en se r water distribution system. 4. Select condense r water tem pera- tures, fl ow rate, an d prim ary pipe sizes. 5. Select cool ing tower type, spe ed control option, eff i cie ncy , app roach temperature, and make cool ing towe r selection. 6. Se le ct chil lers. 7.Fi na lizep ipi ngs yste m de si gn,ca lcu- late pum p hea d, and se lect pumps. 8. Deve lop a nd optim ize control se - quences. Each of the seste ps i s di scuss ed in thi s series of fi ve articles. This article dis- cusses step 8. Typical Chiller Plant Figure 1is a typical primary-only va riab le fl o w chille d wa te r plan t. The plant ha s two of ea ch m ajor compo- ne nt (chi llers, towe rs, conde nse r wate r pum ps, and chilled wa ter pum ps) each sized for 50% of the load. This plant de - sign is very comm on and was use d as the basis of the simulations and optimi- zation for this series of articles and the SDL course up on wh ich it is ba se d. Note tha t the condenser water (CW) pum ps in Figur e 1do not ha ve va ri ab le spe ed drives (V SDs). Seq ue nce s for variable spee d CW pum ps are al so ad- dressed in this article but, as discussed in Part 2 1 of this se ri es and in m ore de - tai l below, VSDs on cond e nse r wate r About the Author Steven T. Taylor, P .E., is a principal at Taylor Engineering in Alameda, Calif. By Steven T. Taylor, P .E., Fellow ASHRAE Optimizing Design & Control Of Chilled Water PlantsPart 5: Optimized Control Sequences This article was published in ASHRAE Journal, June 2012. Copyright 2012 ASHRAE. 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.
5 6 A S H R A E J o u r n a l a s h r a e . o r g J u n e 2 0 1 2
This is the last of a series of articles discussing how to optimize the
design and control of chilled water plants. The series summarizes
ASHRAE’s 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. 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 engineer-
ing time compared to standard practice but at the same time result in
signi cantly reduced plant life-cycle costs.
A procedure was developed to providenear-optimum plant design for most chill-er plants including the following steps:
1. Select chil led water distributionsystem.
2. Select chilled water temperatures,ow rate, and primary pipe sizes.
3. Select condenser water distribution
system.
4. Select condenser water tempera-tures, ow rate, and primary pipe sizes.
5. Select cooling tower type, speedcontrol option, eff iciency, approachtemperature, and make cooling towerselection.
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 thisseries of ve articles. This article dis-cusses step 8.
Typical Chiller PlantFigure 1 is a typical primary-only
variable ow chilled water plant. Theplant has two of each major compo-nent (chillers, towers, condenser waterpumps, and chilled water pumps) eachsized for 50% of the load. This plant de-sign is very common and was used asthe basis of the simulations and optimi-
zation for this series of articles and theSDL course upon which it is based.Note that the condenser water (CW)
pumps in Figure 1 do not have variablespeed drives (VSDs). Sequences forvariable speed CW pumps are also ad-dressed in this article but, as discussedin Part 2 1 of this series and in more de-tail below, VSDs on condenser water
About the AuthorSteven T. Taylor, P.E., is a principal at TaylorEngineering in Alameda, Calif.
By Steven T. Taylor, P.E., Fellow ASHRAE
Optimizing Design & ContrOf Chilled Water Plants Part 5: Optimized Control Sequences
This article was published in ASHRAE Journal, June 2012. Copyright 2012 ASHRAE. Posted at www.ashrae.org. This article may not becopied and/or distributed electronically or in paper form without permission of ASHRAE. For more information about ASHRAE Journal, visitwww.ashrae.org.
The plant in Figure 1, serving a typical ofce building, wasmodeled with all permutations of the following designvariables:
Weather Oakland, Calif., Albuquerque, N.M., Chicago, Atlanta, Miami, Las VegasCHWST Reset by valve position from 42°F to 57°FChillers
• Two styles (two stage and open-drive)• Efciency at 0.35, 0.5, and 0.65 kW/ton at AHRI
conditionsTowers
• Approach: 3°F, 6°F, 9°F, and 12°F• Tower Range: 9°F, 12°F, and 15°F• Efciency: 50, 70, and 90 gpm/hp
Condenser water pumps With and without VSDs
The control equation coefcients were determined fromeach run, then these coefcients were themselves re-gressed against various design parameters and weatherindicators. The results are shown below. The developmentof these regressions is ongoing to include more weathersites and chiller variations.
1. Condenser water temperature control. Control CWreturn temperature to the setpoint determined from Equa-tion 1a:
These control sequences strictly apply to primary-onlyplants with centrifugal chillers serving air handlers withoutdoor air economizers in a typical ofce building. It is notknown how well they apply to other applications.
Modeling the Plant
APPROACH Design tower leaving water temperature mi-
nus WB, °F CHWFR Chilled water ow ratio, actual ow dividedby total plant design ow
CHWST Chilled water supply temperature (leavingevaporator temperature), °F
CWFR Condenser water ow ratio, actual ow di-vided by total plant design ow
CWRT Condenser water return temperature (leavingcondenser water temperature), °F
CDD65 Cooling degree-days base 65°F DP Differential pressure, feet H 2O KW/TON Chiller efciency at AHRI conditions, kW/ton D T Temperature difference, °F GPM/HP Tower efciency per ASHRAE Standard 90.1 IPLV Integrated part load value per AHRI 550/590,
kW/ton NPLV Non-standard part load value per AHRI
550/590, kW/ton RANGE Design tower entering minus leaving water
temperature, °FPLR Plant part load ratio, current load divided by
pumps are usually not life-cycle cost effective for plants serv-ing ofce building type loads.
Also note in Figure 1 that the cooling towers do not includeany isolation valves to shut off ow to allow one tower to op-erate alone. As discussed in Part 4 2 of this series, towers gen-
erally can be selected with nozzles and dams that allow halfow from one CW pump while still providing full coverage ofll and it is always most efcient to run as many tower cellsas possible. So whether one or two CW pumps are operat-ing, both tower cells are enabled and fans are controlled to thesame speed.
Determining Optimal Control SequencesChilled water plants have many characteristics that make
each plant unique so that generalized sequences of control thatmaximize plant efciency are not readily determined. Equip-ment and system variables that affect performance include:
Chillers. Each chiller has unique characteristics that affectfull-load and part-load efciency such as compressor design,evaporator and condenser heat transfer characteristics, un-loading devices (such as variable speed drives, slide valves,and inlet guide vanes), oil management systems, and internalcontrol logic.
Cooling towers. Tower efciency (gpm/hp) varies signi-cantly by almost an order of magnitude between a compactcentrifugal fan tower to an oversized propeller fan tower. Tow-ers can also be selected for a wide range of approach tempera-tures.
Chilled and condenser water pumps. Pumps and pipingsystems can be selected for a broad range of Δ T s and may ormay not include variable speed drives. Pump efciency alsovaries by pump type and size and pump head varies signi-cantly depending on physical arrangement and pipe sizingstandards.
Chilled water distribution systems. Distribution systemarrangements, such as primary-secondary vs. primary-onlyvariable ow, signicantly affect plant control logic.
Weather. Changes in outdoor air conditions affect loads andthe ability of cooling towers to reject energy.
Load prole. The size and consistency of loads will affectoptimum sequences. For instance, control sequences that areoptimum for an ofce building served by air-handling systems
with airside economizers may not be optimum for a data cen-ter served by systems without economizers.With so many variables, no single control sequence will
maximize the plant efciency of all plants in all climates forall building types.
There are a number of papers 3,4 on techniques to optimizecontrol sequences for chilled water plants. Almost all requiresome level of computer modeling of the system and systemcomponents, and the associated amount of engineering timethat most plant designers do not have. In writing this series ofarticles and the SDL upon which it is based, signicant model-ing was performed in an effort to determine generalized con-
trol sequences that account for most of the variation in plant
design parameters summarized above. The technique used todetermine optimized performance is described in a June 2007ASHRAE Journal article. 4
In brief, the technique involves developing calibrated simu-lation models of the plant and plant equipment that are run
against an annual hourly chilled water load prole with coin-cident weather data while parametrically modeling virtuallyall of the potential modes of operation at each hour. The oper-ating mode requiring the least amount of energy for each houris determined. The minimum hourly energy use summed forthe year is called the theoretical optimum plant performance(TOPP). Since all modes of operation were simulated, theplant performance cannot be better than the TOPP within theaccuracy of the component models.
TOPP modeling was performed for the chilled water plantshown in Figure 1 for a wide range of plant design options fortower range, approach, and efciency; different chiller typesand chiller efciency; and varying climates (see “Modelingthe Plant,” Page 57). The operating modes (e.g., number ofchillers, condenser water ow and pump speed, tower fanspeed and related condenser water temperatures) that result inthe TOPP for each plant design scenario were studied to seehow they relate to independent variables such as plant loadand weather (e.g., wet-bulb temperature) to nd trends thatcan be used to control the plant in real applications throughthe direct digital control (DDC) system.
Ideally, equipment should be controlled as simply as pos-sible; complex sequences are less likely to be sustained sinceoperators are more likely to disable them at the rst sign ofperceived improper operation. The remainder of this articlediscusses the TOPP modeling and the generalized sequencesthat were developed from the analysis for the chilled waterplant shown in Figure 1 serving a typical ofce building.
CHW Pump ControlChilled water pump speed is typically controlled to main-
tain supply-to-return differential pressure (DP) at setpoint.Standard 90.1 5 requires that the DP sensor(s) be located at themost remote coil(s). This is because the lower the DP setpoint,the lower the pump energy, as shown in Figure 2 .* If the DPsetpoint is reset by valve position, as discussed further below,pump energy can be close to the ideal curve in Figure 2 for
“DP setpoint =0.”Figure 3 shows the optimum number of CHW pumps as
a function of CHW ow ratio and as a function of pumpspeed for the chilled water plant shown in Figure 1 basedon TOPP modeling. Unlike cooling towers, the optimumsequence is not to run as many pumps as possible. This isbecause the pumps all pump through the same circuit (otherthan the pipes into and out of the each pump between head-ers) so there are not “cube-law” energy benets for eachpump individually.
* The curves in this gure assume pressure drop varies with ow to the 1.8 power since ow is typically in the transitional region between turbulent and laminar ow.They do not account for the impact of opening and c losing control valves, which change system geometry and hence the system ow characteristics. The curvesdo include reductions to the efciency of motors and VSDs at low load.
Figure 2: Variable speed performance at varying DP setpoint.
gpm (%)
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P u m p
k W ( %
)
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DP Setpoint = Design Head
DP Setpoint = Head × 0.75
DP Setpoint = Head ÷ 2
DP Setpoint = Head ÷ 3DP Setpoint = 0 (Reset)
Figure 3 clearly indicates that staging pumps off of owprovides better optimization than staging off of pump speed.
As suggested by Figure 3 , CHW pumps should be stagedas a function of CHW ow ratio (CHWFR =actual ow di-vided by total plant design ow) at a staging point of 47%,i.e., one pump should operate when the CHWFR is below47% and two pumps should operate when CHWFR is above47%, with a time delay to prevent short cycling. The 47%optimum staging point assumes DP setpoint is reset by valveposition; it will be somewhat higher at higher DP setpoints.
For very large pumps (>~100 hp [75 kW]), it may be worththe effort to determine the actual pump operating point (owvs. head) and optimize staging based on pump efciency de-termined by ow and pressure drop readings mapped to pumpcurves duplicated mathematically in the DDC system. 6 Thiscan allow pumps to operate closer to their design efciencyas the system operating curve varies from the ideal paraboliccurve due to modulating valves and minimum differential
pressure setpoint. But the small potential energy savings is notworth the effort for most chilled water plants.
Chilled Water Temperature and DP Setpoint ResetChillers are more efcient at higher leaving water temper-
atures so, in general, optimum efciency is achieved whenthe chilled water supply temperature (CHWST) setpoint isas high as possible. (The impact of CHWST on CHW pumpenergy is discussed below.) Where all zones are controlledby the DDC system, the best reset strategy is based on valve
6 0 A S H R A E J o u r n a l a s h r a e . o r g J u n e 2 0 1 2
Figure 4: Plant energy with CHWST Setpoint Reset, CW DPSetpoint Reset and a combination of the two.
4,000
3,500
3,000
2,500
2,000
1,500
1,000
500
0
W h / f t 2 · y r
CHWST Reset,Fixed DP
Fixed CHWST,DP Reset
Reset CHWSTThen DP
4,000
3,500
3,0002,500
2,000
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1,000
500
0
W h / f t 2 · y r
CHWST Reset,Fixed DP
Fixed CHWST,DP Reset
Reset CHWSTThen DP
Houston
Oakland
Fans ChillerTowersPumps
Figure 3: TOPP CHW pump staging vs. CHW ow ratio and pump speed.
Percent of Design Flow
Percent of Pump Speed
2
1
0 O p
t i m u m
N u m
b e r o
f P u m p s
O p t i m u m
N u m
b e r o f
P u m p s
0 10 20 30 40 50 60 70 80 90 100
2
1
00 10 20 30 40 50 60 70 80 90
position where the CHWST setpoint isreset upwards until the valve control-ling the coil that requires the coldestwater is wide open. This strategy en-sures that no coil is starved; all are able
to maintain their desired supply air orspace temperature setpoints. † Valve position can also be used to
reset the DP setpoint used to controlpump speed. In fact, this is required byStandard 90.1. The logic is similar toCHWST setpoint reset: the DP setpointis reset upwards until the valve control-ling the coil that requires the highest DPis wide open.
So we have a dilemma: Valve positioncan be used to reset either CHWST set-point or DP setpoint, but not both inde-pendently; it is not possible to know ifthe valve is starved for lack of pressureor from lack of cold enough water. Wemust decide which of the two setpointsto favor.
While resetting CHWST setpoint upward reduces chillerenergy use, it will increase pump energy use in variableow systems. Higher chilled water temperature will causecoils to require more chilled water for the same load, de-grading CHW Δ T and increasing ow and pump energy re-quirements. Degrading Δ T can also affect optimum chillerstaging; however, this is not generally an issue in primary-only plants with variable speed chillers. 7 Furthermore, oursimulations have shown that the positive impact of reset-ting chilled water temperature on chiller efciency is muchgreater than the negative impact on pump energy even whendistribution losses are high.
Figure 4 shows a DOE2.2 simulation of a primary-onlyplant with variable speed chillers and CHW pumps withhigh pump head (150 ft [450 kPa]) using three reset strate-gies based on valve position: reset of chilled water tempera-ture alone; reset of differential pressure setpoint alone; anda combination of the two that rst resets chilled water tem-perature then resets DP setpoint. The simulations were done
in several climate zones (Houston and Oakland results areshown in the gure) and in all cases, resetting chilled watertemperature was a more efcient strategy than resetting DPsetpoint. Sequencing the two (resetting chilled water temper-ature rst then DP setpoint) was the best approach, althoughonly slightly better than resetting chilled water temperaturealone.
† Contrary to conventional wisdom, the impact of reset on the dehumidicationcapability of air handlers is quite small and should not be a concern. Space hu-midity is a function of the supply air humidity ratio, which in turn, is a function ofthe coil leaving dry-bulb temperature setpoint. Regardless of CHWST, the air leav-ing a wet cooling coil is nearly saturated; lowering CHWST only slightly reducessupply air humidity ratio. As long as the supply air temperature can be main-tained at the desired setpoint, resetting CHWST will not impact space humidity.
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Figure 5 shows how this sequenced reset strategy can beimplemented. The x-axis is a software point called “CHWPlant Reset” that varies from 0% to 100% using “trim andrespond” logic. 8 The coil valve controllers generate “requests”for colder chilled water temperature or higher pump pressure
when the valve is full open. When valves are generating “re-quests,” CHW Plant Reset increases; when they are not, CHWPlant Reset steadily decreases.
When CHW Plant Reset is 100%, the CHWST setpoint is atT min (the design chilled water temperature) and the DP setpointis at DP max (the design DP setpoint). As the load backs off, thetrim and respond logic reduces the CHW Plant Reset point. Asit does, chilled water temperature is increased rst up to a maxi-mum T max (equal to the lowest air handler supply air tempera-ture setpoint less 2°F [1°C]), then DP setpoint is reduced downto a minimum value DP min (such as 3 psi [21 kPa]).
In practice, this logic seldom results in much reset of the DPsetpoint—the CHWST reset is aggressive enough to starve thecoils rst—so it is important to locate the DP sensor(s) at themost remote coil(s) so that DP max can be as low as possible tominimize pump energy ( Figure 2 ).
Tower Fan Speed ControlA common approach to controlling cooling towers is to
reset condenser water supply temperature based on outdoorair wet-bulb temperature. But our simulations seldom indi-cated a good t; as shown in Figure 6 , the correlation wasfairly good in Miami but not in Oakland and most otherclimates.
For plants serving typical ofce buildings, ‡ good correla-tions were found in all TOPP simulations between plant partload ratio (PLR, actual plant load divided by total plant design
Figure 5: CHWST Setpoint and CW DP Setpoint Reset se -quenced off of CHW valves.
DP max
DPSetpoint
DP min0 50% 100%
DPSetpoint
CHWSTSetpoint
CHWSTSetpoint
T max
CHW Plant Reset
T min
‡ For plants with more consistent loads that do not vary with weather, such asthose serving data centers and those located in consistently humid climatessuch as Miami, correlation of load with CWRT/CHWST temperature differenceis poor. For these plants, optimum CWST vs. wet-bulb temperature was foundto have better correlation. But for ofce buildings in general, the correlations inFigure 7 were more consistent.
Figure 6: TOPP optimum condenser water supply temperature vs. wet-bulb temperature.
90858075
706560555045403530
C W S T
OAWB
30 35 40 45 50 55 60 65 70 75 80 85 90
90858075
706560555045403530
C W S T
OAWB
30 35 40 45 50 55 60 65 70 75 80 85 90
Miami Oakland
y = 0.9133x + 12.997R2 = 0.927
y = 0.1223x + 58.823R2 = 0.0119
capacity) and the difference between the condenser tempera-ture return temperature (CWRT, leaving the condenser) andthe CHWST. Examples are shown in Figure 7. 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 efciency.
A and B are coefcients that vary with climate and plant design(see “Modeling the Plant,” Page 57). Equation 1 can be solvedfor the optimum CWRT setpoint given the current CHWST:
CWRT =CHWST +A ×PLR +B (1a)
This setpoint must be bounded by the minimum CWRT-CHWST difference at low load prescribed by the chillermanufacturer. This minimum (9°F [5°C] for the chiller inFigure 7 ) is a function primarily of the chiller’s oil manage-ment design and can range from only a few degrees for oil-free chillers (e.g., those with magnetic or ceramic bearings)to as high as about 20°F [11°C]. The lower this minimum is,the lower annual chiller plant energy will be, particularly inmild climates.
So near-optimum tower performance can be achieved bycontrolling tower fan speed based on condenser water return temperature to the setpoint determined from Equation 1a.Controlling tower fan speed based on return temperature rath-er than supply temperature is non-conventional but it makessense because it is the temperature leaving the condenser thatdetermines chiller lift, not the entering (supply) water tem-perature. Chiller efciency is not sensitive to entering chilledor condenser water temperature.
Condenser Water Pump ControlNo good correlations were found for control of VSDs on
condenser water pumps. Optimum condenser water pumpspeed and ow were plotted against various parameters suchas PLR, wet-bulb temperature, chiller percent power, and lift
with no consistent relationships. The best correlation wasow vs. PLR as shown in Figure 8, but the correlations wereseldom strong (R 2 typically less than 0.85 and some as lowas 0.5). The correlations were signicantly weaker for pumpspeed than for ow so a condenser water ow meter should beadded if one is not already part of the design.
The curve t can be expressed as follows
CWFR =C ×PLR +D (2)
where CWFR is the ratio of desired CW ow setpoint to thedesign CW ow. The CW ow setpoint is then calculated as:
CWFSP =CWFR ×CWDF (2a)
where CWDF is the design CW ow rate for the plant(both pumps). This setpoint must be bounded by the mini-mum required CW ow rate obtained from the chiller manu-facturer. The minimum ow from most manufacturers cor-
6 8 A S H R A E J o u r n a l a s h r a e . o r g J u n e 2 0 1 2
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–30
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Figure 8: TOPP CW ow vs. plant load ratio.
C W
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Percent Total Plant Design Capacity
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Miami Atlanta
Oakland Chicago
y = 0.8431x + 0.3286R2 = 0.8284
y = 0.9833x +0.2667R2 = 0.8356
y = 1.335x + 0.1784R2 = 0.8488
y = 1.0327x + 0.2503R2 = 0.8333
relates to the onset of laminar ow and will be on the orderof 40% to 70% of design ow depending on the number oftubes, number of passes, and tube design (e.g., smooth vs.enhanced). Higher rates are reputed to discourage foulingof condenser tubes but to the author’s knowledge, no studieshave been done to support that notion. 9 Once the ow rateis determined, CW pump speed is modulated to maintain theCW ow at setpoint.
When C and D coefcients determined for specic plantswere fed back into the energy model, actual performanceranged from 101% to 110% of the TOPP. With this less thanoptimum performance, VSDs were found to be marginallylife-cycle cost effective in dry climates (Albuquerque, N.M.)and not cost effective elsewhere. This performance getsworse when C and D are determined from the regressionequations based on plant design (see “Modeling the Plant”),rather than from actual plant performance modeling (e.g.,Figure 8 ). In some cases, particularly in humid climates, theCW pump control logic caused energy use to increase vs.constant speed CW pumps. Therefore, VSDs on CW pumps
are recommended only on plants in dry climates and only if
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Optimum staging for variable speed CW pumps was foundto correlate very well to CW ow with 60% of the total de-sign ow as the staging point, i.e., one pump should operatewhen the CWFR is below 60% and two pumps should operatewhen CWFR is above 60%, with a time delay to prevent short
cycling.Optimum staging for constant speed pumps was found tovary with both CWRT-CHWST difference and with PLR, butwith fairly weak correlations and relatively small energy im-pact regardless of logic. For simplicity, constant speed CWpumps should simply be staged with the chillers.
Chiller StagingFigure 9 (Page 68) shows the optimum number of chillers
that should be run plotted against plant load for variable speedcentrifugal chillers. The graph shows that it is often optimumto operate two chillers as low as 25% of overall plant load.
This result may seem somewhat counterintuitive; convention-al wisdom is to run as few chillers as possible. That is true forxed speed chillers, but not for variable speed chillers, whichare more efcient at low loads when condenser water tempera-tures are low.
Figure 9 shows that staging chillers based on load alonewill not optimize performance since there is a fairly widerange where either one or two chillers should be operated.
Figure 10: Possible surge problem staging by load only.
Two Chillers
Two Chillers
Surge Region
R e f r i g e r a n t L i f t
100%
90%One Chiller
One Chiller70%
60%
Load
80%
Speed
There is also another problem with staging based on loadalone: it can cause the chillers to operate in surge. This canbe seen in Figure 10, which schematically shows centrifu-gal chiller load vs. lift, dened as the difference betweencondenser and evaporator refrigerant temperature. If twochillers are operated when the refrigerant lift is high (redline), the chillers will operate in the surge region. To avoid
Figure 11: Optimum staging vs. (CWRT-CHWST) and plantpart load ratio.
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One Chiller Two Chil lers
surge, the chiller controllers will speed up the compres-sors and throttle inlet guide vanes to control capacity. Thisreduces chiller efciency so that it would then be more ef-cient to operate one chiller rather than two. But if the liftis low (green line in Figure 10 ), the chillers would not be
in surge so operating two chillers would be more efcientthan operating one. So in addition to load, chiller stagingmust take chiller lift into account. (This consideration ap-plies only to centrifugal chillers; surge does not occur withpositive displacement chillers such as those with screwcompressors.)
Figure 11 shows the optimum number of operating chill-ers (blue dots indicate one chiller while red dots indicate twochillers) for example TOPP simulations. For all plant designoptions and for all climate zones, good correlations werefound for the optimum staging point described by a straightline:
SPLR =E ×(CWRT – CHWST) +F (3)
where SPLR is the staging PLR and E and F are coefcientsthat vary with climate and plant design (see “Modeling thePlant”). If the actual measured PLR is less than SPLR, onechiller should operate; if the PLR is larger than SPLR thentwo chillers should operate, with a time delay to prevent shortcycling.
Note that the number of operating chilled water pumps andthe number of operating chillers may not match. The pumpsmust respond to the ow and pressure requirements of the sys-tem, not to the load, and thus are staged independently fromchillers.
Primary-only variable ow plants like this also will require“soft staging” and minimum ow control. These sequencesand why they are needed are discussed in more detail in theSDL and in Reference 10.
Example The TOPP model results for an Oakland plant were plotted
per Figures 7, 8, and 11 and the following slopes and inter-cepts were determined from curve-ts:
A =47, B =5.2 C =1.3, D =0.13
E =0.009, F =0.21Figure 12 shows the theoretical optimum performance for
both variable speed (VS) constant speed (CS) CW pumpscompared to our proposed “real” sequences using the coef-cients listed above. Despite their simplicity, our sequencesresulted in only about 1% higher energy use than the TOPP.Variable speed drives on the CW pumps saved 3% vs. constantspeed pumps, but this was not enough savings to make themcost effective at a 15 scalar ratio (simple payback period) forthis plant. Also shown in the gure for comparison is plantperformance using the AHRI 550/590 condenser water reliefcurve to reset condenser water temperature (4% higher energy
use than our sequences) and performance assuming CWST
setpoint is xed at the design temperature (16% higher thanour sequences).
Summary This article is the last in a series of ve that summarize
chilled water plant design techniques intended to help engi-neers optimize plant design and control with little or no addedengineering effort. In this article, optimized control logic wasaddressed. The logic is very simple and easily programmedinto any DDC system controlling the plant. With these se-quences properly implemented, chiller plants can performwithin a few percent of their theoretical optimum.
References1. Taylor, S. 2011. “Optimizing design and control of chilled water plants
part 2: condenser water system design.” ASHRAE Journal 53(9):26 – 36.2. Taylor, S. 2012. “Optimizing design and control of chilled water
plants part 4: chiller and cooling tower selection.” ASHRAE Journal
54(3):60 – 70.3. Hartman, T. 2005. “Designing efcient systems with the equalmarginal performance principle.” ASHRAE Journal 47(7):64 – 70.
4. Hydeman, M., G. Zhou. 2007. “Optimizing chilled water plantcontrol.” ASHRAE Journal 49(6):45 – 54.
Figure 12: TOPP vs. real sequences for both constant speedand variable speed CW pumps.
7. Taylor, S. 2002. “Degrading chilled water plant Δ T : causes andmitigation.” ASHRAE Transactions 108(1):641 – 653.
8. Taylor, S. 2007. “Increasing efciency with VAV system static
pressure setpoint reset.” ASHRAE Journal 49(6): 24 – 32.9. Li, W., R. Webb. 2001. “Fouling characteristics of inter-
nal helical-rib roughness tubes using low-velocity coolingtower water.” Int ernational Journal of H eat and Mass Transfer 45(8):1685 – 1691.10. Taylor, S. 2002. “Primary-only vs. primary-secondary variable
ow systems.” ASHRAE Journal 44(2):25 – 29.
This series of articles summarizes the upcoming Self DirectedLearning (SDL) course Fundamentals of Design and Control of
Central Chilled Water Plants and the research that was performedto support its development. The series includes ve segments.
Part 1: “Chilled Water Distribution System Selection” (July 2011),Part 2: “Condenser Water System Design” (September 2011),Part 3: “Pipe Sizing and Optimizing DT” (December 2012), andPart 4: “Chiller & Cooling Tower Selection” (March 2012).
Optimized control sequences. The series concludedwith a discussion of how to optimally control chilled waterplants, focusing on all-variable speed plants.
Central Chilled Water Plants Series The intent of the SDL (and these articles) is to providesimple yet accurate advice to help designers and operatorsof chilled water plants to optimize life-cycle costs withouthaving to perform rigorous and expensive life-cycle costanalyses for every plant.
In preparing the SDL, a significant amount of simula-tion, cost estimating, and life-cycle cost analysis wasperformed on the most common water-cooled plant con-figurations to determine how best to design and controlthem. The result is a set of improved design parametersand techniques that will provide much higher perform-ing chilled water plants than common rules-of-thumb andstandard practice.
