Optimization of Water-cooled Chiller System With Load-based Speed Control

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Applied Energy 85 (2008) 931–950

www.elsevier.com/locate/apenergy

APPLIED

ENERGY

Optimization of water-cooled chiller systemwith load-based speed control

F.W. Yu *, K.T. Chan

Department of Building Services Engineering, The Hong Kong Polytechnic University, Hung Hom, Hong Kong, China

Received 19 December 2007; received in revised form 15 February 2008; accepted 18 February 2008Available online 2 April 2008

Abstract

This study investigates the energy performance of chiller and cooling tower systems integrated with variable condenserwater flow and optimal speed control for tower fans and condenser water pumps. Thermodynamic-behaviour chiller andcooling tower models were developed to assess how different control methods of cooling towers and condenser waterpumps influence the trade-off between the chiller power, pump power, fan power and water consumption under variousoperating conditions. Load-based speed control is introduced for the tower fans and condenser water pumps to achieveoptimum system performance. With regard to an example chiller system serving an office building, the optimal controlcoupled with variable condenser water flow could reduce the annual system electricity use by 5.3% and operating costby 4.9% relative to the equivalent system using constant speed fans and pumps with a fixed set point for cooling watertemperature control.� 2008 Elsevier Ltd. All rights reserved.

Keywords: Water-cooled chiller; Coefficient of performance; Electricity consumption; Operating costs

1. Introduction

Chiller and cooling tower systems are widely used to provide cooling energy for commercial and industrialfacilities but with considerable electricity and water consumption. There are many studies on how the controlof chillers and cooling towers can be optimized to enhance their performance [1–17], but most of them focuseither on electricity savings or on water savings without considering the likely trade-off between the two sav-ings. Hartman [18] launched the equal marginal performance principle (EMPP) to assist in optimizing theenergy performance of chiller systems. The prerequisites for implementing EMPP are that each of the systemcomponents should be driven by a variable speed drive (VSD) and that the mathematical relationship betweenthe system cooling energy output and the power demand of each component has to be identified. The EMPPmay be an extravagant application for most systems where VSDs are applied solely to secondary-loop pumpsand cooling tower fans.

0306-2619/$ - see front matter � 2008 Elsevier Ltd. All rights reserved.

doi:10.1016/j.apenergy.2008.02.008

* Corresponding author. Tel.: +852 27664560; fax: +852 27657198.E-mail address: [email protected] (F.W. Yu).

mailto:[email protected]

Nomenclature

Aimp impeller outlet area (m2)AU overall heat transfer coefficient (kW �C�1)BCL building cooling load (kW)COP coefficient of performance of chillerCpa specific heat capacity of air (assumed to be 1.02 kJ kg�1 �C�1)Cpaf fictitious specific heat capacity of saturation air (kJ kg�1 �C�1)Cpw specific heat capacity of water (assumed to be 4.19 kJ kg�1 �C�1)Cprg specific heat capacity of vapour refrigerant at evaporator (kJ kg�1 �C�1)Cprl specific heat capacity of liquid refrigerant at condenser (kJ kg�1 �C�1)E power input (kW)Echr rated power input to chiller compressor (kW)Ectr rated power consumption of the cooling tower fan (kW)hawb specific enthalpy of saturation air (kJ kg�1)hawbl specific enthalpy of saturation air leaving the cooling tower (kJ kg�1)hadbe specific enthalpy of air entering the cooling tower (kJ kg�1)hadbl specific enthalpy of air leaving the cooling tower (kJ kg�1)hi specific enthalpy of refrigerant at state point i (kJ kg�1)LMTD log mean temperature difference (�C)ma mass flow rate of air entering the cooling tower (kg s�1)mar nominal mass flow rate of air entering the cooling tower (kg s�1)mcdw mass flow rate of water entering the condenser or cooling tower (kg s�1)mctwl mass flow rate of water leaving the cooling tower (kg s�1)mr refrigerant mass flow rate (kg s�1)mw mass flow rate of chilled water (kg s�1)ni index of reversible polytropic expansionNtu number of transfer units of the cooling towerP saturated refrigerant pressure of the refrigeration circuit, absolute kPaPi refrigerant pressure at state point i, absolute kPaPLR chiller part load ratio given by Qcl/Qcr

Qcd heat rejection (kW)Qcl cooling capacity (kW)Qcr nominal cooling capacity (kW)Qm cooling capacity for motor’s heat dissipation (kW)qrf refrigeration effect (kJ kg�1)Sfan speed of cooling tower fan (rpm)Spump speed of condenser water pump (rpm)T temperature of saturated refrigerant within the refrigeration circuit (�C)Tadb dry bulb outdoor temperature (�C)Tadbe dry bulb temperature of air entering the cooling tower (�C)Tadbl dry bulb temperature of air leaving the cooling tower (�C)Tawb wet-bulb outdoor temperature (�C)Tcdwe condenser water entering temperature (equivalent of Tctwl) (�C)Tcdwl condenser water leaving temperature (equivalent of Tctwe) (�C)Tctwe cooling water entering temperature (equivalent of Tcdwl) (�C)Tctwl cooling water leaving temperature (equivalent of Tcdwe) (�C)Tchwr temperature of return chilled water (�C)Tchws temperature of supply chilled water (�C)Tevsh degree of superheat (�C)Tref reference temperature of water (�C)

932 F.W. Yu, K.T. Chan / Applied Energy 85 (2008) 931–950

Uimp tip speed of impeller (m s�1)Vr volumetric flow rate of refrigerant (m3 s�1)vi specific volume of refrigerant at state point i (m3 kg�1)Win Mechanical work input to compressor (kJ kg�1)gpol efficiency of polytropic compressiongm combined motor and transmission efficiency of compressorb angle of impeller (rad)qa air density (assumed to be 1.2 kg m�3)p system pressure ratiopimp impeller pressure ratiod throttling rateea airside heat transfer effectiveness of the cooling towerxadbe moisture content of air entering the cooling tower (kg kg�1 dry air)xadbl moisture content of air leaving the cooling tower (kg kg�1 dry air)xawbe moisture content of saturation air entering the cooling tower (kg kg�1 dry air)xawbl moisture content of saturation air leaving the cooling tower (kg kg�1 dry air)

Subscripts

cc compressorcd condensercdwp condenser water pumpch chillerct cooling towerev evaporatorimp impellermax maximumop optimumr rated condition

F.W. Yu, K.T. Chan / Applied Energy 85 (2008) 931–950 933

Variable flow of chilled water and condenser water is increasingly used to reduce pumping energy in chilledwater systems. According to Bahnfleth and Peyer [19], the successful application of chilled water systems withvariable primary flow depends on how the flow and chiller capacity can be adjusted to match changing loadconditions. This application is specific rather than generic, and the airside cooling coils are required to be fur-nished with two-way control valves in order to allow the flow of chilled water to drop under the reduced loadconditions. For a constant air volume (CAV) system where the latent load removal capacity is a concern, thepotential of reducing chilled water flow rate under part load conditions is rather limited and so is the pumpenergy savings from this application.

Gordon et al. [20] highlighted that the condenser water flow rate could be a control variable in improvingthe energy performance of chiller systems. They established an analytic semi-empirical chiller model to studyvariations of chiller COP (coefficient of performance) at different condenser water flow rates. No analysis wasmade on the system level involving the interaction between the compressor power, pump power and tower fanpower. The model serves well for fault detection and diagnosis purposes, but is incapable of accounting thecontrol variables of cooling towers. No control regime was generalized on how the condenser water flowshould be varied in response to various chiller load and wet-bulb conditions in order to achieve optimal energyperformance of the system.

Applying VSD to cooling tower fans can reduce their cycling frequency and allow better heat rejection con-trol for any given chiller load while maintaining the cooling water temperature set point. This application con-stitutes a requirement for all tower fans with rated power exceeding 3.7 kW in ASHRAE Standard 90.1-2004 –energy standard for buildings except low-rise residential buildings [21]. Yet the true benefits of using variablespeed tower fans hinge on how to control their speed to achieve system optimization. For conventional controlof cooling water leaving temperature, a fixed and high set point of 3–4 �C above the designed wet-bulb


temperature is often used to limit tower fan power, and hence the compressor needs to work at high condens-ing temperatures with a low chiller COP even when the chiller load or wet-bulb temperature drops. Some engi-neers tend to reduce the set point at the lowest possible level of around 15 �C for all operating conditions toincrease the chiller COP by minimizing the compressor power with more fan power. A fixed approach is analternative method in which the set point is adjusted by a fixed value of 3–4 �C above the changing wet-bulbtemperature. This method assumes that the chiller load would change linearly with the wet-bulb temperatureand that the tower fans can operate at lower speeds with power savings at the reduced wet-bulb temperatures.None of these temperature controls is a proven technique to achieve minimum energy use of chillers and cool-ing towers.

Computer simulation is an expeditious means of analysing chiller system performance. Lots of chiller andcooling tower models have been developed in various forms. Pure empirical models generally fail to performoptimization studies as most control variables are absent. Transient models, on the other hand, are toocomplicated for general engineers to master easily and robust numerical methods are required to solve thedifferential equations. Moreover, they are cumbersome for carrying out hour-by-hour energy analysis of mul-tiple-chiller systems frequently encountered. A general search for and comprehensive surveys [22,23] on manyexisting chiller and cooling tower models indicated that very few of them consider the interaction betweenchillers and cooling towers. Furthermore, none of individual models is capable of assessing power relation-ships of chillers, condenser water pumps and cooling towers together with water consumption in the heatrejection process, with regard to various control methods of cooling towers and condenser water pumps. Hyd-eman et al. [24] developed a modified DOE-2 regression-based chiller model for the variable condenser waterflow application. One major task along with the model development was to ascertain which modellingapproach would provide the most accurate simulation results with regard to the validation tests usingmanufacturer-supplied data and field-monitored data. Compared with other three public domain models –DOE-2 model [25], Gordon-Ng model [20] and ASHRAE Primary toolkit model [26], the modified modelprovided more accurate prediction over the chiller power, especially under low load conditions with variablecondenser water flow or variable speed drives. The improvement in the accuracy of the modified model is dueto the inclusion of the condenser water temperature along with the chiller part load ratio in the EIRFPLRfunction which expresses chiller efficiency under part load conditions. No control algorithm is included inthe model to govern the cooling tower characteristics in terms of changes in the cooling water temperatureunder various airflow and water flow conditions for any given heat rejection.

Lu et al. [27] presented a model-based optimization strategy for the condenser water loop of a chiller sys-tem. The strategy involved using a genetic algorithm to minimize the total power of the chillers, condenserwater pumps and cooling tower fans. Regarding the physical constraints of the cooling towers, the interactionbetween the varying air flow rate and water flow rate for a given heat rejection rate was considered. The influ-ence of condenser water entering temperature on both the chiller power consumption and cooling tower per-formance was considered. Yet they did not explain what set point should be used for the condenser waterentering temperature to control the fan speed for optimizing the system. Benton et al. [28] developed a regres-sion model to represent the improved cooling tower simulation algorithm (CTSA). No details were givenabout the mathematical expressions of the algorithm but the testing and comparison results were presentedregarding the accuracy of the model against five CTSAs with the use of manufacturers’ performance data.The algorithm considered the cooling tower approach (cooling water leaving temperature subtracted fromthe wet-bulb temperature) as a dependent variable which is determined by the independent variables of thewet-bulb temperature, range (temperature difference of cooling water), condenser water flow and fan power.It remains uncertain how the algorithm can be used to evaluate an optimal set point for cooling water tem-perature as the cooling temperature leaving temperature is an implicit variable in the evaluation of coolingtower characteristics.

Graves [29] reformulated the Gordon-Ng model in order to study more specifically the optimization of thewhole system including two chiller-pump pairs and one cooling tower satisfying the nominal heat rejectioncapacity of the two chillers. To properly describe the chiller-tower interaction, a condenser load and condenserwater leaving temperature was involved in the mathematical expression of the compressor power. The chillermodel was coupled with a NTU-effectiveness model for evaluating the cooling tower performance. The mod-ified model was capable of analysing how the system COP varied with the changing condenser water flow. Yet


the tower model discounted the water loss due to evaporation, and a single NTU value was assumed to rep-resent the tower performance at different water and air flow rates ranging from 50% to 100% of their nominallevels. Two correlations were identified to facilitate near-optimal system operation: one is the linear relation-ship between the cooling water set point and the wet-bulb temperature (an analogy to the fixed approachmethod); another is the linear relationship between the tower fan speed and pump speed.

It is worth promoting variable speed control as a standard environmental-friendly feature for condenserwater pumps and cooling tower fans in water-cooled chiller systems and strengthening simulation techniquesfor optimizing such application. The aim of this paper is to assess the energy savings and cost benefits ofapplying optimal control to the cooling tower fans and condenser water pumps of typical water-cooled chillersystems. This paper first presents how a thermodynamic chiller model interacts with a cooling tower model todetermine power relationships between the compressor, condenser water pump and tower fan under variousoperating conditions. The tower model, coupled with real control of cooling water temperature, is capable ofsimulating fan power and water loss by evaporation based on any given heat rejection and ambient condition.An assessment will be made on the extent of the increase in the system COP resulting from the optimum speedcontrol of the tower fans and condenser water pumps. Discussion will be given on how the fan and pumpspeed can be controlled directly based on the chiller load in order to achieve optimum system performance.Following that, an example chiller system serving a local office building will be considered to characterizeand compare annual electricity and water cost savings when applying various tower and pump control meth-ods. The significance of this study rests on providing more quantitative analysis to promote water-cooled chil-ler systems with optimal operating schemes in order to boost their environmental performance in terms ofannual electricity and water consumption, and, at the same time, to reduce their operating costs.

2. Description of the chiller and cooling tower models

The chiller and cooling tower models were developed using TRNSYS 15 [30]. Fig. 1 gives the vapour com-pression cycle of the water-cooled centrifugal chiller studied. The saturated vapour refrigerant leaves the evap-orator at point 1. From point 1 to point 10 superheating occurs when the refrigerant absorbs heat from thecompressor motor. Point 2 indicates a pressure drop in the refrigerant entering the compressor when someof the refrigerant is throttled through the partially opened inlet guide vanes at part load. After the polytropiccompression process the refrigerant at point 3 is desuperheated in the condenser to a saturated condition atpoint 30. From point 30 the refrigerant is further condensed to the saturated liquid state at point 4. The satu-rated liquid refrigerant expands isentropically to point 5 before entering the evaporator. No pressure loss inthe refrigerant pipelines was assumed, considering that changes in the condensing pressure and evaporatingpressure due to the loss do not cause an apparent variation in the chiller COP. Given this, P1, P 10 and P5

are equal to the evaporating pressure (Pev), and P3, P 30 and P4 are equal to the condensing pressure (Pcd).

Pcd

Pev

Pres

sure

(kP

a)

Refrigerant specific enthalpy, h (kJ/kg)

1

2

Polytropic compression

1’

33’4

5

Throttling of refrigerant at part load

Fig. 1. Vapour compression cycle of the chiller.


Each operating condition comprises seven inputs: dry bulb outdoor temperature (Tadb), wet-bulb outdoortemperature (Tawb), cooling capacity (Qcl), chilled water flow rate (mw), the temperature of supply chilledwater (Tchws), condenser water flow rate (mcdw) and the temperature of cooling water leaving the cooling tower(Tctwl). The cooling capacity (Qcl), Tadb and the coincident Tawb would come from the hourly load profile ofchillers operating for a given set of building cooling load and weather conditions. The chilled water flow rate(mw) was fixed at its designed level according to the interlocking operation of one chiller with one chilled waterpump at constant speed. The temperature of supply chilled water (Tchws) was set to be 7 �C in all operatingconditions. The reset of Tchws for chiller performance improvements is beyond this study, considering thatnot all airside systems and chilled water distribution systems are compatible with this control.

Unlike many past studies considering a fixed mcdw for all operating conditions, this analysis attempts toascertain how mcdw should be varied to minimize the sum of the compressor power, pump power and towerfan power. There is no generic method to regulate mcdw under various operating conditions for optimum sys-tem performance. Based on the velocity limits of condenser water passing through the condenser, the accept-able range of varying mcdw would be 50–100% of the nominal flow rate. One trial in this study is to vary mcdw

linearly with the chiller load in order to keep the temperature change of the condenser water roughly constant.This could enhance the stability of the cooling water temperature. The implementation of this control calls forthe interlocking operation of each chiller with a variable speed condenser pump and a controller which reg-ulates the pump speed as a function of the chiller load.

The outputs or operating variables of the chiller were determined by the following sets of algebraic equa-tions through an iterative procedure.

2.1. Evaporator model

The cooling capacity (Qcl) of an evaporator is expressed by Eqs. (1)–(4):

Qcl ¼ PLRQcr ð1ÞQcl ¼ mwCpwðT chwr � T chwsÞ ð2ÞQcl ¼ mrqrf ð3ÞQcl ¼ AU evLMTDev ð4Þ

where

qrf ¼ h1 � h5 ð5Þ

AU ev ¼1

c1m�0:8w þ c2Q�0:745

cl þ c3

ð6Þ

LMTDev ¼T chwr � T chws

lnððT chwr � T evÞ=ðT chws � T evÞÞð7Þ

The method of log mean temperature difference (LMTD) was used to model the heat transfer performance ofthe shell-and-tube flooded type evaporator. The overall heat transfer coefficient of the evaporator (AUev) isdescribed by a mechanistic relation in Eq. (6) [31], where c1, c2 and c3 are parameters to be evaluated basedon the performance data of the chiller at full load and part load conditions.

2.2. Compressor model

The chiller contained one refrigeration circuit and one hermetic centrifugal compressor within which inletguide vanes were used to modulate the cooling output at between 25% and 100% of the rated full load capacitywhile controlling the temperature of supply chilled water at its set point. The volumetric flow rate of the refrig-erant at the impeller outlet (Vr) can be evaluated from Eq. (8) [26,32]:

V r ¼Aimp

U imp

tanðbÞ nini� 1

P 2v2 pðni�1Þ=ni � 1� �

� U 2imp

� �ð8Þ


The angle (b), outlet area (Aimp) and tip speed (Uimp) of the impeller are constants with regard to the constantspeed compressor and they were determined by using the parameter identification program given by Bour-douxhe et al. [26]. The system pressure ratio p given by Eq. (9) accounts for the throttling rate d which is equalto 1 at full load. Under part load conditions, d is less than 1, indicating an increase in the system pressure ratio,

p ¼ P cd

dP ev

ð9Þ

It is difficult to find d at different part load ratios because the refrigerant state at the impeller inlet is not iden-tified. To reduce the computation effort to determine d based on fundamental equations, it is possible to cor-relate mr with its maximum mr,max at full load condition by a ratio c, as shown in Eq. (10). The ratio c wasidentified to be a function of chiller part load ratio (PLR), as shown in Eq. (13), where the constants a1 toa3 were determined based on the modelled results of mr under various operating conditions,

mr ¼ cmr;max ð10Þ

mr;max ¼V r

v2

p1=niimp ð11Þ

pimp ¼ 1þ ni� 1

2ni1

P 2v2

U 2imp �

V 2r

A2imp sin2ðbÞ

!" #ni=ðni�1Þ

ð12Þ

c ¼ a1PLR2 þ a2PLRþ a3 ð13Þ

Regarding the computation of the actual compressor power (Ecc) given by Eq. (14), the polytropic compres-sion efficiency (gpol) can be expressed by an empirical polynomial of chiller part load ratio and was determinedby correlating the modelled results of the mechanical work input to the compressor (Win) with the perfor-mance data of Ecc under part load conditions. The combined motor and transmission efficiency (gm) was takento be 0.8 for the constant speed compressor. The specific enthalpy of superheated refrigerant at the compressordischarge (h3) is given by Eq. (17),

Ecc ¼W in

gpolgm

ð14Þ

where

W in ¼ mrP 2v2

nini� 1

pni�1

ni � 1� �

ð15Þ

gpol ¼ a4PLR2 þ a5PLRþ a6 ð16Þ

h3 ¼ h2 þEcc

mr

ð17Þ

2.3. Expansion valve model

An orifice plate with a fixed opening was used to throttle proper level of the refrigerant in the evaporator atvarious load conditions. Assuming that the motor’s heat dissipation (Qm) accounts for 5% of Ecc as given byEq. (19), the degree of superheat (Tevsh) can then be calculated by using Eq. (20),

Qm ¼ mrðh10 � h1Þ ð18ÞQm ¼ 0:05Ecc ð19Þh10 ¼ h1 þ CprgT evsh ð20Þ


2.4. Condenser model

Heat rejection (Qcd) involves the energy and mass balance in the condenser and is given by Eqs. (21)–(26).The overall heat transfer coefficient (AUcd) was predicted by Eq. (25) [31], where the constant parameters c4–c6

were evaluated using the performance data of the chiller at full load and part load conditions. The condenserwater entering temperature (Tcdwe) is the temperature of cooling water leaving the cooling tower (Tctwl) ofwhich the set point is used to control the cooling tower fan while satisfying Qcd,

Qcd ¼ Qcl þ Ecc ð21ÞQcd ¼ mrðh3 � h4Þ ð22ÞQcd ¼ mcdwCpwðT cdwl � T cdweÞ ð23ÞQcd ¼ AU cdLMTDcd ð24Þ

where

AU cd ¼1

c4m�0:8cdw þ c5Q1=3

cd þ c6

ð25Þ

LMTDcd ¼T cdwl � T cdwe

lnððT cd � T cdweÞ=ðT cd � T cdwlÞÞð26Þ

2.5. General control of cooling tower fans by use of cooling water leaving temperature

A robust cooling tower model with actual control of cooling water leaving temperature is developed toillustrate how all operating variables in the heat rejection process can be computed. Modelling equations froman effectiveness model given by Braun [33] was adopted but the problem-solving approach was redefined tomatch the real fan control logic. The model is capable of evaluating the states of air passing through a coolingtower, and thus water loss due to evaporation can be estimated. As Eq. (27) illustrates, the heat rejectioncapacity of a cooling tower correlates with the airside heat transfer effectiveness (ea), the air mass flow rate(ma) and the hypothetical difference between the enthalpy of saturation air leaving the tower and the enthalpyof air entering the tower. ea of a cross-flow cooling tower is given by Eq. (28), assuming that the Lewis numberis one. The variable Cpaf is the fictitious specific heat which is defined as the enthalpy difference over the tem-perature difference of the leaving and entering conditions of the saturation air. Regardless of the entering orleaving conditions, the saturation enthalpy (hawb) can be determined from its wet-bulb temperature (Tawb) byEq. (31), where the constants k0 to k3 are 9.3625 kJ kg�1, 1.7861 kJ kg�1 �C�1, 0.01135 kJ kg�1 �C�2 and0.00098855 kJ kg�1 �C�3, respectively.

The number of transfer units (Ntu) was used to determine the cooling tower performance, which is given byEq. (32) with the characteristic constants c and n. It varied with the air mass flow rate (ma) for a constant orvariable mass flow rate of water leaving from the condenser (mcdw).

Qcd ¼ eamaðhawbl � hadbeÞ ð27Þ

ea ¼1

m�ð1� expð�m�ð1� expð�NtuÞÞÞÞ ð28Þ

where

m� ¼ maCpaf

mcdwCpw

ð29Þ

Cpaf ¼hawbe � hawbl

T awbe � T awbl

ð30Þ

hawb ¼ k0 þ k1T awb þ k2T 2awb þ k3T 3

awb ð31ÞNtu ¼ cðmcdw=maÞ1þn ð32Þ


The states (enthalpy hadbl and moisture content xadbl) of air leaving the cooling tower were determined by Eqs.(33) and (34). Given xadbl, the mass flow rate of water leaving the cooling tower (mctwl) can be evaluated usingthe mass balance in Eq. (35), where the water loss rate due to evaporation is given by (mcdw � mctwl). The cool-ing water leaving temperature (Tctwl) is related to the cooling water entering temperature (Tctwe) by Eq. (36).Tref is the water reference temperature and 0 �C was used,

hadbl ¼ hadbe þ eaðhawbl � hadbeÞ ð33Þxadbl ¼ xawbl þ ðxadbe � xawblÞ expð�NtuÞ ð34Þmctwl ¼ mcdw � maðxadbl � xadbeÞ ð35Þ

T ctwl ¼mcdwCpwðT ctwe � T refÞ � Qcd

mctwlCpw

þ T ref ð36Þ

According to the conventional approach to evaluating cooling tower performance, the heat rejection airflowrate (ma) is one of the inputs required to calculate the temperatures of cooling water, states of ambient air, etc.This is in contrary to the real situation where ma is a variable depending on how the tower fan is controlled bya set point of cooling water leaving temperature (Tctwl) or condenser water entering temperature (Tctwe).Where a constant speed fan is used, it will be cycled on and off to maintain Tctwl within the tolerable rangeof the set point. For a variable speed tower fan, the air mass flow rate varies by modulating the fan speedto satisfy the required Tctwl. Based on the aforementioned equations, an iteration procedure shown inFig. 2 is required to solve ma based on a given Tctwl.

For a given set of inputs (mcdw, Tctwl and Tawb) and Qcd computed by the condenser model, the range andapproach of the cooling tower can be evaluated. Tctwl has to be set at above Tawb to achieve a positiveapproach (Tctwl � Tawb) in order to calculate a logical heat rejection airflow rate (ma). The initial trial forthe hypothetical wet-bulb temperature of air leaving the cooling tower (Tawbl) is taken to be the cooling waterentering temperature (Tctwe) subtracted by 0.01 �C. The fictitious specific heat Cpaf can then be determined,given the enthalpies and temperatures of the saturation air calculated at the leaving and entering conditions.The LMTD method is applied to describe the effective heat transfer coefficient (AUct) of the cooling tower, andma is then computed by using the given relationship between Ntu and AUct. A new Tawbl can be calculatedbased on the ma computed and checked with its previous value. If the difference between the values of Tawbl

in successive calculations is within ±0.005 �C, then the convergent criterion of the solutions can be met and allother variables can be determined explicitly. Otherwise Tawbl needs to be deducted by 0.01 �C in each iterationto repeat the calculation steps until the convergent criterion is satisfied. In cases where the computed ma isgreater than its nominal level mar, mar should be used to determine the smallest approach and all other vari-ables for the given set of inputs.

The water evaporation rate of a cooling tower working under various operating conditions was computedaccording to the mass flow rate of air entering the cooling tower and the difference between the leaving andentering moisture contents of that air, i.e. ma (xadbl � xadbe). The power of a constant speed fan (Ect) is therated power Ectr multiplied by the factor (ma/mar), considering that the average cycling period of the fan dur-ing an hour is directly proportional to the required ma and assuming that the fan power increases as a linearfunction of the tower airflow. If the fan operates continuously at its full flow (mar) throughout an hour, thefactor will be equal to one and its rated power will be taken up over that hour. While a variable speed fanis used, the fan law is applied and the factor becomes (ma/mar)

3 based on the fan law with the cube relationshipbetween the power and flow rate. In this analysis, the variable speed drive was assumed to account for 3% ofthe power of the fan operating at any speed and the minimum flow rate delivered could be down to 10% of therated flow rate.

2.6. Evaluation of operating variables

The chiller and cooling tower models were used to evaluate all the operating variables based on any giveninput data and constant parameters. The flow charts in Fig. 3 show the relationships between the componentmodels and how the operating variables were evaluated. The evaporator model started to compute PLR,Tchwr, AUev and LMTDev based on the inputs: cooling capacity (Qcl), chilled water flow rate (mw) and the

Fig. 2. Procedure for evaluating the required air mass flow rate under the actual control of cooling tower fans.


temperature of supply chilled water (Tchws). It then calculated the evaporating temperature (Tev) and pressure(Pev), and refrigerant properties at the evaporator discharge (h1, v1). Given that the condensing temperature(Tcd) linked the compressor and condenser models, the operating variables of these models had to be com-puted altogether at a specific accuracy through an iterative procedure. The iterative procedure started withan initial condensing temperature (Tcdo) of 45 �C in calculating the variables of the compressor model (thecase: ITERo = 0). A degree of superheat (Tevsho) of, say, 3 �C was initially assumed in order to computethe compressor power (Ecc). Iterations were carried out on each Tevsh calculated by the expansion valve modeluntil Tevsh converged to within a specified tolerance of ±0.005 �C.

In the condenser model, the inputs consisted of cooling water leaving temperature (Tctwl), compressorpower (Ecc), cooling capacity (Qcl) and the refrigerant mass flow rate (mr) satisfying the requirement of the

Fig. 3. Procedure for determining the operating variables of the chiller model.


evaporator and compressor models. Based on the heat rejection (Qcd) computed by the condenser model, oper-ating variables related to the cooling tower were then determined by solving the modelling equations and usingthe control algorithm of operating the tower fan. If the difference between the condensing temperature (Tcd)evaluated from the condenser model and its previous value calculated from the compressor model was within±0.005 �C, all operating variables would be calculated logically with the required accuracy; otherwise the nextvalue of Tcd would substitute for its previous one to perform the next iteration until the accuracy was met.

The COP of a chiller is defined as the cooling capacity (Qcl) divided by the compressor power (Ecc). In thecalculation of a system COP, the power input means the total power of the compressor (Ecc), condenser waterpump (Ecdwp) and cooling tower fan (Ect). When the condenser water pump operates at constant speed, Ecdwp

is fixed at its rated value Ecdwp,r for the chiller operating. When a VSD is applied to the pump, Ecdwp is Ecdwp,r

multiplied by a factor (mcdw/mcdwr)3 based on the cube relationship between the power and flow rate. mcdw


varies linearly with the chiller part load ratio (PLR), from 0.5 mcdwr at a PLR of 0.5 to mcdwr at a PLR of 1.The variable speed drive was assumed to take up 3% of Ecdwp at any speed.

3. Results and discussion

3.1. Improved performance from optimal tower fan control with variable condenser water flow

Table 1 summarizes the details of the chiller system studied. The total capacity of the system was designedto satisfy the peak cooling load of an office building (see Section 3.3 for details). The chiller model was cal-ibrated for the part load performance of the chiller shown in Fig. 4. Each performance curve represents onlythe variation of the chiller COP at different part load ratios for a given condenser water entering temperature(Tcdwe), irrespective of how Tcdwe or the cooling water leaving temperature (Tctwl) is actually controlled by thecooling tower at any given wet-bulb temperature. The parameters of the modelled chiller given in Table 1 wereidentified from the model validation exercise and the characteristic parameters (c and n) used for the coolingtower are representative of typical cross-flow cooling towers.

The operating schemes shown in Table 2 were analysed with respect to different controls of the condenserwater pump and cooling tower. Schemes 1–4 represent a constant speed configuration of the cooling tower fanand condenser water pump. Under this configuration, the tower fan is cycled on and off to deliver the heatrejection airflow required to meet a given Tctwl for any heat rejection. The condenser water pump is stagedcontinuously to provide the operating chiller with the rated condenser water flow rate for all loading condi-tions. For schemes 5–8, variable speed control is applied to the fan and pump. Under this control, the fan

Table 1Details of the chiller system

Total plant cooling capacity (kW) (four identical sets of chillers, pumps and cooling towers) 6400For each chiller

Refrigerant type R134aNominal cooling capacity (kW) 1600Nominal compressor power (kW) 280.7COP at full load 5.7Design chilled water supply/return temperature (�C) 7/12.5Design chilled water flow rate (l/s) 72Design condenser water entering/leaving temp. (�C) 33/38Design condenser water flow rate (l/s) 87Chiller model parameters

Evaporator (c1, c2, c3) (units omitted) 0.1172, 0.0593, 0.0001Compressor (Aimp, Uimp, b) 0.0059 m2, 200 m/s,

2.269 radCompressor (a1, a2, a3 for c) 0.0704, 0.931, 0.00006Compressor (a4, a5, a6 for gpol) �0.8131, 1.537, 0.0057Condenser (c4, c5, c6) (units omitted) �0.0005, 0.3443, 0.0002

For each cooling tower

Type Cross-flowHeat rejection capacity (kW) 2004Design entering/leaving temperature (�C) 34.4/29.4Water flow rate (l/s) 87Motor Staged or variable speedAir volume flow rate (m3/s) 63Air mass flow rate (kg/s) 75.6Fan motor power (kW) 22.8Drift loss (% of nominal flow) 0.2Design water evaporation rate (% of nominal flow) 0.53Design wet-bulb outdoor temperature (�C) 28Tower model parameters

c, n 2.3, �0.7Rated power of each chilled water pump (kW) 47.0Rated power of each condenser water pump (kW) 21.6

3

4

5

6

7

8

0 0.2 0.4 0.6 0.8 1Chiller part load ratio

Chi

ller

CO

P

T cdwe (oC)

3330282624222018

Fig. 4. Part load performance of the water-cooled centrifugal chiller.

Table 2Operating schemes of different controls of condenser water pump and cooling tower

Scheme Notation Description

1 (baseline) Tctwl29.4CS Tctwl set at 29.4 �C with constant speed configurations2 Tctwl18.3CS Tctwl set at 18.3 �C with constant speed configurations3 App4CS Approach (Tctwl–Tawb) fixed at 4 �C with constant speed configurations4 Tctwl,opCS Optimum Tctwl for maximum system COP with constant speed configurations5 Tctwl29.4VS Tctwl set at 29.4 �C with variable speed configurations6 Tctwl18.3VS Tctwl set at 18.3 �C with variable speed configurations7 App4VS Approach (Tctwl–Tawb) fixed at 4 �C with variable speed configurations8 Tctwl,opVS Optimum Tctwl for maximum system COP with variable speed configurations


speed is modulated in response to a given Tctwl to continuously adjust the airflow for any heat rejection withfan power savings. The variable speed pump provides varying condenser water flow which changes in directproportion to the chiller load in order to maintain a constant temperature difference across the condenser.

The set points of 29.4 �C and 18.3 �C used for Tctwl in the schemes 1, 2, 5 and 6 are based on the upper andlower limits of part load rating conditions given in ARI standard 550/590 [34]. A fixed approach of 4 �C wasconsidered in the schemes 3 and 7, considering that it helps offer fan power savings when the chiller load dropsand that this control provides more energy savings compared with the fixed set point control [12]. For schemes4 and 8, the models contain an algorithm which searches for an optimal Tctwl to minimize the sum of compres-sor power, condenser water pump power and cooling tower fan power for any given operating condition.

Fig. 5 shows the part load performance of the system (one chiller dedicated with one condenser water pumpand one cooling tower) operating at different combinations of part load ratios and wet-bulb temperatures in

2

3

4

5

6

0.2 0.4 0.6 0.8 1.0Chiller part load ratio

Syst

em C

OP

16202428

Wet bulb outdoor temperature (oC)

Fig. 5. System part load performance at different wet-bulb temperatures in scheme 1.


the scheme 1 (the baseline). There is no significant difference in the trend of system COPs at part load withdifferent wet-bulb temperatures when the set point of Tctwl is fixed at 29.4 �C for all operating conditions.For a given wet-bulb temperature, the system COP is maximized at a part load ratio of 0.8–0.9, instead ofat full load, and it drops considerably when the chiller load reduced from such part load ratios. This variationin the chiller COP is governed by compressor efficiency and the capacity control done by the inlet guide vanes.

Comparisons were made on how the system COP varied under various tower fan and pump controls. AsFig. 6 illustrates, lowering Tctwl to 18.3 �C could not bring an absolute increase in the system COP, though thecompressor power was lowest for most operating conditions among the other operating schemes. Indeed, thesystem COP could drop by up to 10% when the chiller load reduced from a part load ratio of 0.8. This isbecause the tower fans operated at full speed continuously and the increase of the fan power exceeded thereduction of the compressor power. Except the 18.3 �C Tctwl control, the system COP could be improvedby reducing the condenser water flow rate via modulating the pump speed in response to the chiller load. Thisimplies that keeping the temperature difference of the condenser water at the design level is a viable means tooptimize the trade-off between the compressor power and condenser water pump power.

The system COP could increase by various degrees for all operating conditions when using the 4 �Capproach or implementing the optimal Tctwl, regardless of whether the on/off or variable speed control wasused for the tower fans and condenser water pumps. When the wet-bulb temperature was below 28 �C, thepercentage increase in the system COP was more noticeable at higher chiller loads when the 4 �C approachcontrol was in place of the 29.4 �C Tctwl control. The system COP increased by 1.4–16.1% by using the opti-mum Tctwl and VSDs for both the fans and pumps. Such an increase tends to be significant at the minimum

-15

-10

-5

0

5

10


Perc

enta

ge c

hang

e in

sys

tem

CO

P (%

)

Tctwl18.3CS App4CSTctwl,opCS Tctwl29.4VS

Tctwl18.3VS App4VSTctwl,opVS

Operating schemes

-10

-5

0

5

10

15

20


Perc

enta

ge c

hang

e in

sys

tem

CO

P (%

)

Tctwl18.3CS App4CS

Tctwl,opCS Tctwl29.4VS

Tctwl18.3VS App4VS

Tctwl,opVS

Operating schemes

-5

0

5

10

15

20


Perc

enta

ge c

hang

e in

sys

tem

CO

P (%

)

Tctwl18.3CS App4CS


Tctwl18.3VS App4VS

Tctwl,opVS

Operating schemes

-15

-10

-5

0

5

10

0.2 0.4 0.6 0.8 1.0

Chiller part load ratio

Perc

enta

ge c

hang

e in

sys

tem

CO

P (%

)

Tctwl18.3CS App4CS


Tctwl18.3VS App4VS

Tctwl,opVS

Operating schemes

Fig. 6. Percentage change of system COP in schemes 2–8 in relation to the baseline.


and maximum loading conditions and be moderate at a part load ratio of around 0.5 with a wet-bulb temper-ature of below 28 �C.

3.2. Load-based speed control for system optimization

The simulation results showed that for a chiller operating at above half load, it is possible to adjust thecondenser water flow rate linearly with the part load ratio (PLR) to optimize the trade-off between the com-pressor power and condenser water pump power while maintaining the temperature difference at the con-denser side at its design level. To achieve such flow modulation, the variable speed drive should regulatethe pump speed (Spump,op) based on Eq. (37), given that speed is direct proportional to flow rate accordingto the pump laws. Spump,full in Eq. (37) is the full speed of the pump,

Fig. 7.ratio.

Spump;op ¼PLRSpump;full if PLR > 0:5

0:5 Spump;full if PLR 6 0:5

�ð37Þ

To perform optimum temperature control for cooling tower fans, it is essential to have a dedicated controllerwhich calculates the optimum set point (Tctwl,op) based on signals of wet-bulb temperatures (Tawb) and thechiller load in terms of a part load ratio (PLR), and operates the fans at the right speed to meet that set point.Eq. (38) gives the basic mathematical expression of Tctwl,op, where b0 to b2 are coefficients to be determinedbased on the power relationships between the chillers, tower fans and condenser water pumps for a given sys-tem design. In this analysis, b0, b1 and b2 were identified to be 2.9178, 4.8636 and �2.095, respectively, accord-ing to the plot of (Tctwl,op–Tawb) against PLR given in Fig. 7. The data in the plot refer to Tctwl,op coincidedwith the maximum system COP at various combinations of PLR from 0.2 to 1 at 0.1 intervals and Tawb from16 to 28 �C at 4 �C intervals. The Tctwl,op predicted by using Eq. (38) deviates from its actual value by �0.50 to0.49 �C only, based on its possible change from 20.0 to 33.2 �C. This slight prediction error would bring nosignificant effect on attaining the maximum system COP,

T ctwl;op ¼ T awb þ b0 þ b1PLRþ b2PLR2 ð38Þ

The calculation of PLR involves measurements of three variables: the flow rate (mw), supply temperature(Tchws) and return temperature (Tchwr) of chilled water. Overall there are four measured variables affectingthe uncertainty of the calculated Tctwl,op, in addition to its prediction error. A high quality relative humiditysensor should be used to limit the uncertainty of the measured Tawb. Furthermore, the measurements shouldhave response time which is short enough to trace the individual changing patterns of PLR and Tawb.

For most existing cooling water temperature control, it would be difficult to incorporate the reset algorithminto the controllers. Alternatively, it is expected to have more simple and direct control for the tower fans in

y = -2.095x 2 + 4.8636x + 2.9178

R2 = 0.8497

2

3

4

5

6

7

8


(Tct

wl,o

p -

Taw

b) a

t max

imum

sys

tem

CO

P

Relationship between the optimum cooling water leaving temperature (Tctwl,op), wet-bulb temperature (Tawb) and chiller part load

Near-optimum % fan speed = 72.808 PLR + 17.762

R2 = 0.9215

0

20

40

60

80

100


% f

ull s

peed

of

fans

at m

axim

um s

yste

m C

OP

Fig. 8. Load-based speed control for cooling tower fans.


order to achieve maximum system COP for all operating conditions. As Fig. 8 illustrates, it is possible to cor-relate the fan speed (in % full speed) with the chiller part load ratio (PLR). The data set shown in Fig. 8 refersto the optimum fan speed at which the maximum system COP took place for a set of operating conditions interms of various combinations of PLRs from 0.2 to 1 at 0.1 intervals and wet-bulb temperatures from 16 to28 �C at 4 �C intervals. Based on the regression curve with the coefficient of determination (R2) of 0.9215, thenear-optimum fan speed (Sfan,op) can be predicted by Eq. (39), where Sfan,full is the full speed of the fan,

TableCompoptimu

Wet-b

16202428

Sfan;op ¼ ð0:7281PLRþ 0:1776ÞSfan;full ð39Þ

Table 3 gives a comparison between the percentage increase of the system COP from baseline under theload-based speed control and that under the optimum Tctwl control of the tower fans, with regard to variousoperating conditions. For any given operating condition, the results of the two controls are quite comparable,suggesting that it is viable to use the load-based speed control to achieve near-optimum operation of the sys-tem while eliminating sophisticated controllers and sensors for the cooling water temperature reset. When theload-based speed control is recognized for the tower fans and condenser water pumps, the optimal control ofthe whole system can be highly simplified. This is because the sequencing of chillers, pumps and tower fans andtheir individual speed controls can be based entirely on the chiller load conditions only. Such load-based sys-tem control serves to emphasize the relevance of chiller load measurement and verification to system optimi-zation. It is expected that such control is generic for all types of chiller systems with multiple chillerarrangements and with full or partial use of VSDs for the system components.

3.3. Potential benefits from the load-based speed control

The cooling load profile of a reference office building in Hong Kong was considered in order to assess thepotential operating cost savings when the load-based speed control was applied to the chiller plant designed

3arison between the percentage increase in system COP from baseline under the load-based fan speed control (SC) and that under them cooling water temperature control (TC)

ulb (�C) Chiller part load ratio

0.2 0.4 0.6 0.8 1

SC TC SC TC SC TC SC TC SC TC

10.21 10.30 5.27 5.28 5.44 5.52 9.61 9.71 16.01 16.088.19 8.21 3.79 3.83 3.11 3.29 5.60 5.86 9.51 9.736.89 7.12 2.99 3.06 1.54 1.82 2.46 2.87 4.03 4.43

10.52 11.00 6.48 6.57 3.52 3.95 2.31 2.93 1.26 1.81


with four identical sets of the chillers described in Table 1. Detailed features of the building are given in Ref.[35]. The building has 40 storeys and a total air-conditioned floor area of 42840 m2. Fig. 9 is a histogram show-ing the building cooling load profile. There are 2834 cooling hours in total which account for 90.5% of thetotal office hours (3131 h a year). The annual cooling energy for the building is 7,423,883 kW h based on localweather conditions of an example year for building energy analysis. To meet the changing building coolingloads, conventional chiller sequencing was considered, under which all the chillers are operating at the sameload, and no additional chillers start to operate until each of the running chillers is operating at its full load of1600 kW. Within the chiller system, a single-loop pumping system with a differential pressure by-pass pipe wasused to control the amount of chilled water flowing from the operating chillers to cooling coils of the airsideequipment. There are four constant speed pumps, each dedicated to one chiller to provide a constant chilledwater flow of 72 l/s. Each pump has a rated power of 47 kW.

Drawing on the chiller and cooling tower models, calculations were made on the annual electricity con-sumption of the chiller plant with the eight operating schemes listed in Table 2. The load-based speed control(L-VS) was considered as a substitute for the optimum Tctwl control in the scheme 8. The electricity consump-tion is normalized by the total air-conditioned floor area of the building in terms of kW h m�2. Table 3 showsthe results of the normalized electricity consumption of the system components along with the annual waterconsumption of the cooling towers. The water consumption is made up of three parts of water loss: evapora-tion, drift and bleed-off. The cooling tower model was used to estimate the loss rate by evaporation. The driftrate and bleed-off rate were assumed to be 0.2% and 0.6%, respectively, of the cooling water circulation rate,with regard to the use of traditional chlorination water treatment.

Table 4 shows that the annual electricity consumption of chilled water pumps is 6.6 kW h m�2 for all theoperating schemes because the same sequencing of the chillers was applied. The use of variable speed controlfor the condenser water pumps enables their annual electricity consumption to drop by 46.5%. A low set pointof the cooling water leaving temperature helps lower the annual chiller electricity consumption (as in schemes2 and 6), but with much higher fan electricity consumption. Using the load-based speed control for the towerfans and condenser water pumps, the trade-off between the chiller power, pump power and fan power wasoptimized, resulting in a drop of 5.3% in the annual plant electricity consumption. The reduced condenserwater flow along with the advanced control could bring about a 12.8% reduction or 63.8 l m�2 drop in theannual water consumption.

To investigate the cost effectiveness of the eight operating schemes for the chiller system, the annual oper-ating costs and life cycle costs were calculated and compared. Local tariff structures were considered to eval-uate the operating costs associated with the annual electricity and water consumption. Operating costs otherthan the electricity and water charges were assumed to be the same value for all the schemes and they were

0

100

200

300

400

500

400 800 1200 1600 2000 2400 2800 3200 3600 4000 4400 4800 5200 5600 6000 6400

Building cooling load (kW)

No.

of

oper

atin

g ho

urs

1 chiller operating 2 chillers operating 3 chillers operating 4 chillers operating

Fig. 9. Building cooling load profile.

Table 4Annual chiller plant electricity and water consumption per unit of air-conditioned floor area of the building

Scheme 1 (baseline) 2 3 4 5 6 7 8 (% saving from baseline)

Normalized annual electricity consumption (kW h m�2)

Chiller 30.6 29.0 30.1 29.9 31.4 29.8 31.0 30.8 (�0.7%)Chilled water pump 6.6 6.6 6.6 6.6 6.6 6.6 6.6 6.6 (0%)Condenser water pump 3.1 3.1 3.1 3.1 1.6 1.6 1.6 1.6 (46.5%)Cooling tower fan 1.8 3.1 2.0 2.1 0.7 3.0 0.7 0.7 (57.6%)Total 42.0 41.8 41.7 41.7 40.4 41.2 39.9 39.8 (5.3%)

Normalized annual water consumption (l m�2)

497.4 598.5 513.0 519.7 413.3 519.6 424.5 433.6 (12.8%)

Table 5Annual operating costs and life cycle costs per unit of air-conditioned floor area of the building for the chiller plant with eight operatingschemes

Scheme 1 (baseline) 2 3 4 5 6 7 8 (saving from baseline)

Normalized annual operating cost (HK$ m�2)

Electricity 51.7 51.3 51.4 51.3 49.9 50.9 49.6 49.5 (2.2)Water 2.6 3.1 2.7 2.7 2.1 2.6 2.2 2.2 (0.4)Total 54.3 54.4 54.0 54.0 52.1 53.5 51.8 51.7 (2.6)

Normalized life cycle operating cost (HK$ m�2)

Electricity 393.2 390.5 390.7 389.9 379.9 387.1 377.0 376.4 (16.8)Water 19.7 23.3 20.3 20.5 16.3 20.0 16.7 17.0 (2.7)Total 413.0 413.8 411.0 410.4 396.2 407.1 393.7 393.4 (19.5)


excluded in this cost analysis. Regarding the calculation of the life cycle operating costs, a life span of 15 yearswas assumed for the chiller plant and a discount rate of 10% was considered throughout the lifespan. Table 5gives the calculation results. The annual operating cost could drop by HK$2.6 to HK$51.7 m�2 by using theload-based speed control for the condenser water pumps and cooling tower fans instead of the conventionalon/off control with a fixed cooling water leaving temperature. The actual annual savings of HK$111,384 helpsdecide how much investment cost the load-based speed control is worth taking. If the building owner accepts amaximum payback of up to 2 years for a chiller plant improvement program, the ceiling on the investmentcost for the advanced control could be HK$222,768 which is expected to be sufficient for purchasing eightVSD controllers for the tower fans and condenser water pumps and the associated control system. Afterthe operating cost savings fully recoup the implementation cost, the building owner can enjoy profits broughtfrom the optimum control throughout the rest of the system’s functional life. It is envisaged that the imple-mentation of load-based speed control would be more economically attractive if the chiller plant needs tooperate at part load conditions for longer hours per year or if there is a subsidized scheme for purchasingenergy efficient equipment.

4. Conclusions

This paper presents the use of load-based speed control to enhance the energy performance of water-cooledchiller systems. Thermodynamic-behaviour chiller and cooling tower models have been developed to investi-gate how the energy and water uses vary for a chiller system operating under various controls of condenserwater pumps and cooling tower fans. The optimum operation of the system can be achieved simply anddirectly by the load-based speed control under which the speed of the tower fans and condenser water pumpsis regulated as a linear function of the chiller part load ratio. The superiority of such control rests on its coher-ence with typical sequencing of chillers based entirely on their load conditions and on eliminating the need ofhigh quality humidity sensors for the reset of cooling water temperature. The system COP under the optimalcontrol could increase by 1.4–16.1% relative to the equivalent system with fixed temperature and flow rate


controls for the cooling water leaving from cooling towers. A case study showed that the optimum controlcould bring about a drop of 5.3% in the operating cost of a chiller plant running for the cooling load profileof a local office building. It is envisaged that the implementation of the optimum control is economically via-ble, as the operating cost savings would fully recoup the investment cost associated with the control in twoyears or less. The findings of this research highlight the need to widen the use of variable speed drives withload-based speed control for water-cooled chiller systems serving air-conditioned buildings in order toenhance their sustainability.

Acknowledgement

This study was supported by a grant from the central research grant of The Hong Kong Polytechnic Uni-versity, Project A/C Code: G-U272.

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Optimization of Water-cooled Chiller System With Load-based Speed Control

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