Experimental investigation of a pilot compression chiller with alternatives refrigerants using compact heat exchanger

Energy Conversion and Management 66 (2013) 277–284

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Energy Conversion and Management

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Experimental investigation of a pilot compression chiller with alternativesrefrigerants using compact heat exchanger

M.A.M. Hassan ⇑Dept. of Mech. Power Eng., Faculty of Eng. (Mattaria), Masaken El-Helmia P.O., Cairo 11718, Egypt

a r t i c l e i n f o

Article history:Received 16 May 2011Received in revised form 20 September 2012Accepted 20 September 2012Available online 7 December 2012

Keywords:PerformanceWater cooled chillersR407CR410A

0196-8904/$ - see front matter � 2012 Elsevier Ltd. Ahttp://dx.doi.org/10.1016/j.enconman.2012.09.022

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a b s t r a c t

This paper reports the results of an experimental investigation on a pilot compression chiller workingwith R407C and R410A as R22 alternatives using compact heat exchanger. Experiments are conductedand the influences of the cooling water mass flow rate (200–2500 kg/h), cooling water inlet temperature(20–45 �C) and chilled water mass flow rate (200–1500 kg/h) on performance characteristics of chillersare evaluated for R22, R407C and R410A.

Increasing the cooling water mass flow rate or decreasing its inlet temperature causes the operatingpressures and electric input power to decrease while the cooling capacity and coefficient of performance(COP) increase. The pressure ratio is reduced while actual loads and COP are increased with the chilledwater mass flow rate. The effect of cooling water inlet temperature on the system performance is moresignificant than the effects of cooling and chilled water mass flow rates. Comparison between R22, R407Cand R410A under identical operating conditions has been done. An enhancement is obtained when usingcompact heat exchanger than shell and tube heat exchanger, this enhancement is about 8.5% in the coef-ficient of performance.

� 2012 Elsevier Ltd. All rights reserved.

1. Introduction

Chilled water systems are commonly used for comfort cooling,plastic processing, pulp and paper processing, the milk and foodindustries, etc. Among cooling systems, vapor compression chillersare the most widely used to produce chilled water with tempera-ture ranges from 6 to 18 �C [1–4] for such applications. These com-pression chillers can be categorized according to cooling mediuminto air-cooled chillers and water-cooled chillers. Air-cooled chill-ers utilize ambient air to condense refrigerant and use chilledwater to produce their cooling capacities. Water-cooled chillershave a flow of cooling water (which leaves a cooling tower ataround 26–32 �C [5] passing through their condensers and a flowof chilled water passing through their evaporators to deliver the re-quired cooling capacities. Scroll, screw, reciprocating and centrifu-gal compression water-cooled chillers are currently available in themarket. The present work focuses on water-cooled reciprocatingchillers that represent an overwhelming majority of chiller instal-lations and cooling energy consumption in the world.

HCFC-22, introduced about 60 years ago, is the world’s mostwidely used refrigerant [6]. It serves in both residential and com-mercial applications. It is particular combination of efficiency,capacity and pressure has made it a popular choice for equipment

ll rights reserved.

designers. Recently, extensive use of HCFC-22 has made it possibleto reduce the use of CFC refrigerants, because its Ozone DepletionPotential (ODP) is as much as 95% lower than CFCs. Nevertheless, itdoes have some ODP, so international law set forth in the MontrealProtocol and its Copenhagen and Vienna amendments have putHCFC-22 on a phase out schedule [7]. In developed countries, pro-duction of HCFC-22 will end no later than the year 2030. In inter-vening years, production is reduced in a series of specified steps.Detailed phase-out schedules vary from country to country. Obvi-ously, new substitutes should possess what HCFC-22 lacks from anenvironmental standpoint, while retaining its good thermody-namic characteristics. Environmentally, the ideal candidate musthave an ODP of zero and a low Global Warming Potential (GWP).Thermodynamic similarity is important when retrofitting existingchillers; otherwise, capacity and efficiency may be lost. In newchillers designs, thermodynamic similarity is less of an issue be-cause the equipment can be specifically designed for the replace-ment refrigerant’s particular characteristics. Currently, there aretwo leading replacement candidates: HFC-407C, and HFC-410A.

HFC-407C is a high-pressure refrigerant, similar to HCFC-22with ODP of zero. As such, it is a long-term environmental solutionto ozone depletion and is not scheduled for production phase-out[8]. HFC-407C is a blend refrigerant with a low toxicity level. Inselecting a refrigerant to replace HCFC-22, HFC-407C was foundto be an ideal alternative for air-cooled DX chillers. Its operatingcharacteristics are so similar to HCFC-22 that it has been used as

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Nomenclature

C specific heat (kJ/kg K)COP coefficient of performance (�)M mass flow rate (kg/s)P pressure (bar)

P electric compression power (kW)PR pressure ratio (�)Q heat transfer rate (kW)T temperature (�C)

278 M.A.M. Hassan / Energy Conversion and Management 66 (2013) 277–284

a drop-in alternative in machines originally designed for HCFC-22.This allows the continued use of the proven design and compo-nents of air-cooled chillers [9–11]. Additionally, a unique featureof HFC-407C – the property known as ‘‘temperature glide’’ – pre-sents intriguing implications for DX chillers design. This tempera-ture glide phenomenon, when matched with DX cooler designmodifications, can improve energy efficiency. HFC-407C is a blendof three refrigerants: HFC-32, HFC-125 and HFC-134a. This compo-sition exhibits the characteristics of a zeotropic blend, meaningthat the resulting mixture does not act as a single substance. At agiven pressure, it evaporates over a range of temperatures, ratherthan at a single temperature. HFC-407C has a gliding temperaturedifference (GTD) of approximately 6 K [12], which can be leveragedto give opportunities for greater efficiency.

HFC-410A is a high-pressure refrigerant, having approximately60% higher pressure than HCFC-22. As a result, it can provide signif-icant capacity gain to a compressor designed to handle the pressure[13]. HFC-410A is a leading candidate for unitary residential andcommercial equipment. HFC-410A is a blend refrigerant consistingof HFC-125 and HFC-32 (50/50%). It has a low temperature glide of0.5 �C [14], small enough to have little or no effect on heat exchangerperformance. HFC-410A has an ODP of zero. As such, it is a long-termenvironmental solution to ozone depletion and is not scheduled forproduction phase-out. However, due to its high pressure, the exist-ing systems need to be redesigned, manufactured and certified forthe new pressure level. This in turn poses enormous challenges interms of engineering and manufacturing costs while try to stay com-petitive in a price driven market. Therefore, HFC-410A cannot beconsidered as a drop-in alternative for the existing R22 systems.Table 1 gives generally property data for R22, R407C and R410A.

The literature survey revealed that massive work on simulationor modeling of water-cooled chillers is available in the literaturewhile there is a lack of experimental data on water-cooled chillersworking with R22 alternatives such as R407C and R410A [15–17]using compact heat exchanger. Thus, the present work is carriedout with the aim of evaluating the performance characteristics ofa pilot vapor compression chiller (4 kW cooling capacity) workingwith R22, R407C and R410A under identical external operating con-ditions using compact heat exchanger. In order to achieve this aim, atest facility of a water cooled chiller is constructed and equippedwith the necessary instrumentation. The effects of the cooling watermass flow rate and its inlet temperature and chilled water mass flow

Table 1Property data for the considered refrigerants [18,19].

Property R410A R407C R22

NBP at 1 atm (�C) �51.58 �43.56 �40.81Critical temperature (�C) 72.13 86.74 96.15Critical pressure (kPa) 4926.1 4619.1 4990vcr (m3/kg) 0.0021 0.0019 0.0019qsl at 25 �C (kg/m3) 1062.4 1134 1191.8qsv at 25 �C (kg/m3) 65.9 41.9 44.23cp at 30 �C (kJ/kg K) 0. 832 0. 830 0.66Latent heat at NBP (kJ/kg) 234.3 188.25 233.95ODP (R11 = 1) 0 0 0.05GWP (R11 = 1) 0.825 0.4 0.425

rate on the performance characteristics of the water-cooled chillersworking with R22 alternatives are experimentally evaluated.

2. Description of test-rig

A schematic diagram of the test rig is shown in Fig. 1. It is com-prised of three distinct paths for the refrigerant, chilled water andcooling water. Each path has its own instrumentation and safetyand operational controls.

The refrigerant path consists of a vapor compression refrigera-tion system with a single cylinder reciprocating hermetic compres-sor; compact plate heat exchangers are used as an evaporator andcondenser which offer a unique mounting system and the highestlevel of thermal efficiency and durability in a compact. The uniqueintegral mounting tabs and connection locations allow for easyinstallation. The corrugated plate design provides very high heattransfer coefficients, resulting in a more compact design. The unit’sstainless steel plates are vacuum brazed together to form a dura-ble, integral piece that can withstand high pressure and tempera-ture. A thermostatic expansion valve (TEV) with externalpressure equalizer, and different accessories such as a receiver, asight glass and a filter-drier. R22, R407C and R410A are used asan internal working fluid while water is used as an external heattransfer fluid at the evaporator and condenser. In order to mini-mize the heat exchange with the surroundings, the refrigerant cir-cuit is thermally insulated with polyurethane foam and fiberglass.

The chilled water path includes a chilled water stainless steeltank (0.5 m3), a chilled water pump, the compact heat exchangerevaporator and the control valves. The cooling water path containsa cooling water stainless steel tank (0.5 m3), a cooling water pump,the compact heat exchanger condenser and the control valves. Afeed water system is used to guarantee constant water levels inboth water stainless steel tanks. Drainpipes are provided to helpthe draining and cleaning of the system. In order to reduce the heatexchange between the environment and the chilled water path orthe cooling water path, both paths are thermally insulated.

The necessary instrumentation system is arranged to measurevarious operating parameters such as refrigerant and water temper-atures, refrigerant pressures, water flow rates and electric power.Copper�Constantan (type-T) thermocouples are installed at differ-ent locations in the refrigerant and water paths for measurementof temperatures which are monitored by a digital thermometercapable of reading 0.1 �C. Digital pressure gauges (bourdon tubetype) are connected to the refrigerant path to indicate the pressureat the inlet and outlet of both the evaporator and the condenser. Theflow rate of water is manually measured using a digital balancecapable of reading 0.001 g and a stopwatch capable of reading0.01 s. Electric input power supplied to the compressor is measuredusing a digital watt meter capable of reading 0.1 W. All of the instru-mentation devices are calibrated before their use.

2.1. Experimental procedure and data reduction

After leakage tests were successfully completed, the refrigerantcircuit has been evacuated using a double stage vacuum pump.

p5, T5

Chilled water

Electric heater

Drain

V12

Make-upwater

tank

V7

Chilled water pump

T10

T9V9

V8

V10

Evaporator

T6

V11 V1

Cooling coil

Cooling water tank

Condenser

Receiver

V5 Filter drier

F1

V4

Sight glass

Thermal expansion

valve

V6

Compressor

p4,T4

V3

T3

Cooling water pump

V16

Drain

V13T7

V14

T8

p2, T2

p1, T1

V2 V15

Condensing unit

Chilled water line

Measured ParametersT TemperatureP Pressure

Make-upwater

Cooling water line

Refrigerant line

Fig. 1. Schematic diagram of the test rig.

Table 3Ranges of operating parameters.

Primary working fluids R22, R407C and R410A

Cooling water mass flow rate 200–2500 kg/hCooling water inlet temperature 20–45 �CChilled water mass flow rate 200–1500 kg/h

M.A.M. Hassan / Energy Conversion and Management 66 (2013) 277–284 279

This vacuum was held for one day and after that the refrigerant cir-cuit was charged by the correct amount of refrigerant, which wasdetermined using both the input current and sight glass. Theexperimental procedure was then conducted as follows: desiredchilled and cooling water temperatures were reached within±0.1 �C, water pumps were switched on to pump the water intothe evaporator and the condenser, the refrigerant compressorwas switched on and refrigerant superheating degrees at evapora-tor outlet was set to 5 ± 0.5 �C using a thermal expansion valve(TEV). The system was allowed to run until the steady state ofthe operating conditions was reached, then data were recordedfour times during one hour and the average value was used inthe data reduction. Experiments were started with R22 followedby R407C and R410A under controlled external operating condi-tions. Validation of the test facility is done by comparing R22 datawith available data in literature. Two compressors were used in theexperimental runs; one was used for R22 and R407C, the specifica-tions of the two compressors are shown in Table 2. No modificationwas made to the test facility before the experiments for R407C andR22 except lubricant change in the case of R407C. The other com-pressor was used for R410A. Table 3 lists the ranges of operatingparameters.

The actual cooling capacity (Qe) is calculated based on the mea-sured chilled water mass flow rate and its temperature differencethrough the evaporator, i.e.

Q e ¼ mchcchðTch;i � Tch;oÞ ð1Þ

The condenser heat load (Qc) is estimated from the measuredcooling water mass flow rate and temperature rise through it asfollows:

Q c ¼ mcwccwðTcw;o � Tcw;iÞ ð2Þ

where m is the water mass flow rate, c is the water specific heat andT is the water temperature at the respective state points. Subscripts‘‘ch’’ and ‘‘cw’’ stand for chilled water and cooling water. The actual

Table 2Compressors specifications.

Compressor R407C and R22 R410

Nominal (hp) 2 2Cooling capacity (kW) 4.5 5.1COP 2.9 2.7Displacement (m3/h) 5.3 3.9

coefficient of performance of the compression chiller (COP) wascomputed by the following equation:

COP ¼ Q e

Pð3Þ

where P is the measured electric input power supplied to the com-pressor. The uncertainty in various measured quantities has beenanalyzed using the procedure proposed by Moffat [20] as follow:

If a certain quantity (M) is a function of variables (x,y,z . . . t),

M ¼ f ððx; y; z . . . tÞ ð4Þ

And if there are errors Dx, Dy, Dz . . . Dt, in determining the valuesof the corresponding quantities x,y,z . . . t, then the value (M) com-puted from the inexact values of the arguments will be obtainedwith an error.

ðM þ DMÞ ¼ f ½ðxþ DxÞ; ðyþ DyÞ; ðzþ DzÞ . . . ðt þ DtÞ� ð5Þ

Then for sufficiently small absolute of the quantities (Dx, Dy,Dz . . . Dt) the increment can be approximately replaced by the to-tal differential:

DM ¼ @M@x

Dxþ @M@y

Dyþ @M@z

Dzþ � � � þ @M@t

Dt ð6Þ

The maximum absolute error of quantity (x) is the ratio of theminimum reading of the device used in the measurements andthe minimum reading of the measuring value plus the accuracyof the device denoted by (k) and is denoted by dx, i.e.,

dx ¼ jDxjjxj þ j ð7Þ

The uncertainty in the experimental data of cooling capacity,condenser heat load, compressor power and coefficient of perfor-mance are estimated to be about 1.5%, 2.4%, 1.8% and 3.4%,respectively.

6

8

10

Pres

sure

rat

io (

-)

6

9

12

15

ter

outl

et te

mpe

ratu

re (

°C)

T chw,i =11°C

T cw,i =27°C

R410AR407CR22

PR


3. Results and discussion

Experimental results of the compression chiller working withR22, R407C and R410Q are obtained under a wide range of identi-cal external operating parameters. The influences of cooling watermass flow rate and its inlet temperature and chilled water massflow rate on performance characteristics of a small compressionchiller, using R22, R407C or R410A as a working fluid, are discussedin the next sections.

0 500 1000 1500 2000 2500 3000

Cooling water mass flow rate (kg/hr)

2

4

0

3

Chi

lled

waT chw,o

Fig. 3. Variation of performance characteristics with cooling water mass flow rate.

0 500 1000 1500 2000 2500 3000


1

2

3

4

5

6

7

8

Act

ual h

eat l

oads

(kW

) T chw,i =11°C

T cw,i =27°C

R410A

R407CR22

Qc

Qe


1

2

3

4

Ele

ctri

c po

wer

(kW

)

2

4

6

CO

P

Tchw,i =11°C

Tcw,i =27°C

R410AR407C

R22

Pel

COP

3.1. Effect of cooling water mass flow rate

Fig. 2 shows the measured operating pressures of R22, R407Cand R410A versus the cooling water mass flow rate. As the coolingwater mass flow rate increases, the operating pressures of theinvestigated refrigerants decrease. Clearly, condensing pressure ismore sensitive than evaporator pressure to the variation in thecooling water mass flow rate. This is shown by the higher slopeof the condensing pressure line than that of the evaporatingpressure.

Fig. 3 presents the pressure ratio and measured chilled wateroutlet temperature against the cooling water mass flow rate. Sincethe slope of the condensing pressure line is higher than that of theevaporating pressure as shown in Fig. 2, the pressure ratio (pc/pe) isreduced when the cooling water mass flow rate rises. It is seen thatmeasured chilled water outlet temperature decreases as the cool-ing water mass flow rate increases. This can be attributed to thereduction in evaporator pressure with the cooling water mass flowrate as shown in Fig. 2.

Variations of actual cooling capacity and actual condenser heatload with cooling water mass flow rate are shown in Fig. 4. It is evi-dent from the figure that both cooling capacity and condenser heatload increase with cooling water mass flow rate. For given chilledwater mass flow rate and its inlet temperature, as the coolingwater mass flow rate increases, the chilled water outlet tempera-ture decreases causing the chilled water temperature differenceto increase and thereby the evaporator cooling capacity increases.It is also observed that the cooling water temperature differencethrough the condenser reduces when the cooling water mass flowrate rises. Thus, the increase in the condenser heat load is mainlydue to an increase in the cooling water mass flow rate.

Measured electric power input to the compressor and the actualcoefficient of performance as a function of the cooling water flowrate are illustrated in Fig. 5. When the cooling water flow rate in-creases, the electric power input to the compressor decreaseswhereas the actual COP increases. A lower electric power inputto the compressor can be attributed to a lower pressure ratio at a

0 1000 2000 3000


0

10

20

30

40

Pres

sure

s (b

ar)

Tchw,i =11°C

Tcw,i =27°C

R410A

R407C

R22 pc

pe


0 500 1000 1500 2000 2500 3000


0 0


higher cooling water mass flow rate. The actual coefficient of per-formance sharply increased due to the combined effect of theincreasing evaporator cooling capacity and reducing compressorpower input as cooling water mass flow rate increases.

3.2. Effect of cooling water inlet

Dependency of performance characteristics of a small compres-sion chiller upon the cooling water inlet temperature for the con-sidered refrigerants is shown in Figs. 6–9. Measured operating

24 28 32 36 40 44Cooling water inlet temperature (0C)

0

10

20

30

40

Pres

sure

s (b

ar)

m chw =790kg/hr

m cw =662 kg/hr

R410AR407CR22 pc

pe

Fig. 6. Variation of performance characteristics with cooling water inlettemperature.

24 28 32 36 40 44

Cooling water inlet temperature (0C)

2

4

6

8

10

Pres

sure

rat

io (

-)

0

3

6

9

12

15

18

21

Chi

lled

wat

er o

utle

t tem

pera

ture

(°C

)

m chw =790kg/hr

m cw =662 kg/hr R410A

R407C

R22 PRTchw,o


24 28 32 36 40 44

Cooling water inlet temperature (OC)

0

2

4

6

8

Act

ual h

eat l

oads

(kW

)

mchw =790kg/hr

mcw =662 kg/hr

R410A

R407CR22

QcQe


24 28 32 36 40 44

Cooling water inlet temperature (OC)

0

1

2

3

4

Ele

ctri

c po

wer

(kW

)

1

2

3

4

5

CO

P

m chw =790kg/hr

m cw =662 kg/hr R410A

R407C

R22Pel

COP


0 400 800 1200 1600

Chilled water mass flow rate (kg/hr)

0

10

20

30

40

Pres

sure

s (b

ar)

Tchw,i =17.5 °C

Tcw,i =27.5 °C

R410A

R407C

R22

pcpe



pressures against the cooling water inlet temperature are pre-sented in Fig. 6. As expected, both the evaporator and condenserpressures increase with cooling water inlet temperature.

Variations of pressure ratio and measured chilled water outlettemperature with the cooling water inlet temperature are illus-trated in Fig. 7, which reveals that the pressure ratio increases withthe cooling water inlet temperature. This trend is mainly due to alarge increase in the condenser pressure compared to a small in-

crease in the evaporator pressure. It should be noted that the watertemperature differences through both the condenser and evapora-tor decrease as the cooling water inlet temperature increases. As aresult, for given water mass flow rates, a high cooling water inlettemperature yields low evaporator and condenser heat loads asshown in Fig. 8.

The effect of cooling water temperature on electric compressorpower and actual COP can be predicted from Fig. 9. A higher cool-ing water inlet temperature needs a higher compressor power andattains a lower actual COP. The large required compressor power ismainly because of the higher pressure ratio at a high cooling waterinlet temperature as shown in Fig. 8. The increase in compressorpower along with the reduction in evaporator cooling capacitywith the cooling water inlet temperature, leads to a decrease inthe actual COP when the cooling water inlet temperature increasesas shown in Fig. 9.

3.3. Effect of chilled water mass flow rate

The effect of the chilled water mass flow rate on the perfor-mance characteristics of a small water-cooled chiller working withR22, R407C or R410A is shown in Figs. 10–13. Variations of mea-sured evaporator and condenser pressures with the chilled waterflow rate are presented in Fig. 10. It can be seen that both evapora-tor and condenser pressures are increased with chilled water massflow rate. Clearly, the influence of chilled water mass flow rate onboth the evaporator and condenser pressures is minor.

0 400 800 1200 1600


2

4

6

8

10Pr

essu

re r

atio

(-)

0

3

6

9

12

15

18

21

24

27

Chi

lled

wat

er o

utle

t tem

pera

ture

(°C

)

Tchw,i =17.5 °C

Tcw,i =27.5 °C

R410A

R407C

R22

PR

Tchw,o


0 400 800 1200 1600Chilled water mass flow rate (kg/hr)

1

2

3

4

5

6

Act

ual h

eat l

oads

(kW

)

Tchw,i =17.5 °C

Tcw,i =27.5 °C

R410A

R407C

R22Qc

Qe


0 400 800 1200 1600


0

0.5

1

1.5

2

2.5

Ele

ctri

c po

wer

(kW

)

0

1

2

3

4

5

CO

P

Tchw,i =17.5 °C

Tcw,i =27.5 °C R410A

R407C

R22 Pel

COP



Fig. 11 illustrates pressure ratio and measured chilled wateroutlet temperature as a function of chilled water mass flow rate.Clearly, the pressure ratio decreases when the chilled water massflow rate increases. This behavior can be explained by the fact thathigher flow rates result in a higher change in evaporator pressurethan that for condenser pressure and thereby the pressure ratiois reduced. The measured chilled water outlet temperature slightlyincreases due to an increase in evaporation pressure with thechilled water mass flow rate.

Actual evaporator and condenser heat loads versus chilledwater mass flow rate are presented in Fig. 12. The increase in evap-orator cooling capacity is mainly due to an increase in convectiveheat transfer coefficient as chilled water mass flow rate increases,whereas the increase in temperature difference through the con-denser is the main reason behind the increase in the condenserheat load.

Measured electrical power input to the compressor and actualcoefficient of performance against chilled water mass flow rateare indicated in Fig. 13. Both electrical power input to the com-pressor and actual coefficient of performance slightly increasedwith chilled water mass flow rate. As mentioned earlier inFig. 10, when the chilled water mass flow rate increases, evapora-tor pressure rises, causing the vapor density at the compressor in-let to increase. This leads to an increase in refrigerant mass flowrate through the compressor. In contrast, a reduction in pressureratio yields a small specific compression work. It was observedthat the rate of increase in refrigerant mass flow rate is higherthan that of decrease in specific work, thereby electrical power in-put to the compressor increases. Since the rate of increase in ac-tual evaporator heat load is higher than that in compressorpower input, the actual coefficient of performance increases asshown in Fig. 13.

3.4. Comparison between R22, R407C and R410A

Comparison between R22, R407C and R410A is carried out overa wide range of cooling water mass flow rates (Figs. 2–5), coolingwater inlet temperatures (Figs. 6–9) and chilled water mass flowrates (Figs. 10–13).

3.4.1. Operating pressures and pressure ratioComparing the variation of condensation and evaporation pres-

sures with various external operating parameters revealed thatR410A yields the highest condensation pressure whereas R22 pro-duces the lowest condensation pressure as shown in Figs. 2, 6 and10. Also, R410A produces the highest evaporation pressure whilethe evaporation pressure for R22 and R407C is comparable.

R22, R407C and R410A yield average condensation pressures of13.4, 15.4 and 20.1 bar while their average evaporation pressuresare about 4.7, 5.4 and 6.7 bar, respectively, over the entire rangeof cooling water mass flow rate (Fig. 2). It can be noted that aver-age condensation pressure of 15.8, 18.2 and 22.7 along with meanevaporation pressure of 4.6, 5.3 and 6.7 are obtained for R22,R407C and R410A, respectively, when the cooling water inlet tem-perature is varied between 20 and 45 �C (Fig. 6). As chilled watermass flow rate increases from 200 to 1500 kg/h, average condensa-tion pressures of R407C and R410A are higher than that of R22 byabout 15% and 44%, respectively, while the average evaporationpressures of R407C and R410A are higher than that of R22 by about16.1% and 44.4%, respectively (Fig. 10).

As illustrated in Figs. 3, 7 and 11, the influence of various exter-nal operating parameters on the pressure ratio of R22, R407C andR410A showed that the average pressure ratios of R410A andR407C are higher than that of R22 by about 6% and 4%, respectively,over the considered range of cooling water conditions (Figs. 3 and7). R22, R407C and R410A yield average pressure ratios of 3.09, 3.1and 3.26, respectively, as the chilled water mass flow rate increasesfrom 200 to 1500 kg/h (Fig. 11).

3.4.2. Chilled water outlet temperatureVariations of external operating parameters yield an average

chilled water outlet temperatures between 7.3 and 14.0 �C (Figs. 3,7 and 11), which is within the recommended range for differentapplications. The chilled water outlet temperatures of R407C andR410A are comparable. However, R22 achieves chilled water outlet


temperature lower than those of R407C and R410A by about 0.4–1.4 �C. R22, R407C and R410A achieve average chilled water outlettemperatures of 9.0, 8.3 and 9.3 �C, respectively, over the entirerange of cooling water mass flow rate (Fig. 3). Due to the changein cooling water inlet temperatures from 20 to 45 �C, R22 yieldsaverage chilled water outlet temperature of 7.3 �C while R407Cand R410A achieve comparable average chilled water outlet tem-perature of nearly 9.8 �C as shown in Fig. 7. Average chilled wateroutlet temperature of R22 is approximately 12.8 �C and that ofR407C or R410A is about 14.0 �C over the considered chilled watermass flow rate range (Fig. 11).

3.4.3. Heat loads and electric compressor powerComparison between the considered refrigerants from the view

point of actual heat loads confirmed that, in general, the condenserload for R22 is larger than those of R407C and R410A (Figs. 4, 8 and12) while the condenser heat loads for R22 and R410A are compa-rable. The average condenser loads for, R22, R407C and R410A are4.8, 4.6 and 4.6 kW respectively, over the considered range of cool-ing water mass flow rates (Fig. 4). Also, the average condenser loadof R22 is higher than that of R407C by about 5% while that ofR410A is approximately same as R407C when the cooling water in-let temperature is changed from 27 to 43 �C (Fig. 8). However, theaverage condenser loads of R22, R407C and R410A are 3.3, 3.2 and3.2 kW respectively; over the considered range of the chilled watermass flow rate (Fig. 12).

It is seen that the cooling capacities for R22 is higher than thatof R407C by about 8.8% while that of R401 is approximately 5%lower than of R22 over the considered range of cooling water massflow rates (Fig. 4). The average cooling capacities of R22 is higherthan that of R22 and R410A by about 5% as illustrated in Fig. 8when the cooling water inlet temperature is changed from 20 to45 �C. Also, the average cooling capacity of all the investigatedrefrigerants is comparable under the investigated range of chilledwater mass flow rate (Fig. 12).

Comparing the required electric power for refrigerants underinvestigation operating conditions revealed that, R22, R407C andR410A need approximately similar electric compressor power overthe entire range of the considered cooling water mass flow rate(Fig. 5). The electric input power of R22 is larger than that ofR407C and R410A over the whole range of cooling water inlet tem-peratures (Fig. 9). The electric input power of 410A is larger than

Table 4Numerical constants for COP, Eq. (8).

Refrigerant Constants

A0 A1 A2 A3 A4

R22 6.01 0.00051 0.00031 �0.1712 �0.0295R407C 5.32 0.00049 0.00019 �0.1501 �0.0601R410A 4.79 0.00061 0.00029 �0.1611 �0.0301

Table 5Comparison between compact heat exchanger and shell and tube using R22.

Cooling water mass flow rate (kg/h) Compact heat exchanger

COP Qev (kW)

205 2.79 3.32576 3.28 3.671098 3.34 3.671569.6 3.73 3.911890 3.98 4.092510 3.40 3.73

that of R407C and R22 by 4% when chilled water mass flow ratechanged from 200 to 1500 kg/h (Fig. 13).

3.4.4. Actual COPVariations of external operating parameters yield average coef-

ficient of performance (COP) between 3.4 and 1.6 (Figs. 5, 9 and13). Actual COP of R410A is lower than that of R22 by about 10%while R22 yields the highest actual COP and actual COP of R22 ishigher than of R407C by 6.25% over the considered range of coolingwater mass flow rates (Fig. 5). In general, average COP of 2.7, 2.6,and 2.57 is achieved for R22, R407C and R410A, respectively, whenthe cooling water inlet temperature is changed from 20 to 45 �C(Fig. 9). Actual COP of R410A is lower than that of R22 by about12.5% while that of R410A is approximately same as R407C overthe considered range of the chilled water mass flow rate (Fig. 13).

In order to provide the actual COP as function of user parame-ters (variables can be controlled by the user such as cooling waterinlet temperature, cooling water mass flow rate and chilled watermass flow rate, etc.), the experimental data are correlated asfollows:

COPact ¼ A0 þ A1mcw þ A2mch þ A3Tchþ4Tcw ð8Þ

Table 4 gives the constants A0, A1, A2, A3 and A4 of the aboveequation for the investigated refrigerants. Validity range of coolingwater mass flow rate is 200–2500 kg/h, cooling water inlet temper-ature is 20–45 �C, chilled water mass flow rate is 200–1500 andchilled water inlet temperature is 10–18 �C. It can be noted thatthe predicted COP using Eq. (8) is about ±10% of the experimentalCOP.

It should be noted that when chilled water mass flow rate, cool-ing water mass flow rate and cooling water inlet temperature in-creased by 50% of their corresponding values, the actual COPvaried by 3.4%, 6% and 9%, respectively. This confirms that the ef-fect of cooling water inlet temperature is larger than those of cool-ing and chilled water mass flow rates on the chiller performance.

3.5. Effect of using compact heat exchanger

Compact heat exchangers are gaining increasing attention inthe industrial practice, and they are being considered as cost-effec-tive alternatives in applications. Many of the features of compactheat exchanger when dealing with different refrigerants have beennoted during this study, for example an enhancement in the coef-ficient of performance is obtained when using compact heat ex-changer, and this enhancement is about 8.5%. The comparisonhas been done by comparing the results of the existing test-rigwith the previous result in the same test-rig but using shell andtube heat exchanger for R22 as working fluid and is shown in Ta-ble 5. Also this table shows an increasing in the actual load ofthe compact heat exchangers compared to the shell and tube evap-orator. The enhancement of the coefficient of performance and theheat load is due to improvement of the heat transfer characteristicsduring evaporation and condensation of the refrigerant inside thecompact heat exchanger.

Shell and tube heat exchanger

Qc (kW) COP Qev (kW) Qc (kW)

4.51 2.57 3.03 4.194.78 3.02 3.35 4.454.77 3.07 3.35 4.434.95 3.44 3.57 4.615.11 3.67 3.73 4.754.82 3.139 3.40 4.49


4. Conclusions

In this paper, performance characteristics of a small water-cooled chiller working with R22, R407C or R410A as primary refrig-erant have been experimentally determined and compared over awide range of external operating conditions.

Based on the reported results, the following conclusions aredrawn:

– R410A and R22C produce the highest and the lowest condensa-tion pressures, respectively.

– R410A yields the highest evaporation pressure while the evap-oration pressures for R22 and R407C are comparable.

– R22 achieve the lowest pressure ratios and is similar to R407C.The pressure ratio of R410A is higher than that of R407C andR22 by 5.8%.

– Average chilled water outlet temperatures between 7.04 and15.3 �C are obtained over the considered range of the externaloperating parameters. However, the chilled water outlet tem-peratures of R407C and R410A are comparable.

– R22 yields higher condenser heat load compared to those of R22and R401, which have approximately the same condenser heatloads.

– Cooling capacity of R407C is similar to R410A and lower thanthat of R22 over the considered range of investigated operatingconditions.

– The compressor working with R410A and R22 consume nearlythe same electric power and lower than that of R22 over theentire range of the considered operating parameters.

– Variations of the actual coefficient of performance of the inves-tigated refrigerants are insignificant due to changes of coolingand chilled water mass flow rates. R22 yields the highest actualCOP over the whole range of cooling water inlet temperatures.

– An enhancement of the COP and heat exchanger load usingcompact heat exchangers than shell and tube heat exchangers.

In conclusion, R407C can replace R22 as a drop-in refrigerant inwater-cooled chillers with comparable performance characteristicsand R410A cannot replaced R22 as a drop in refrigerant but can beused as alternative as a long term solution while R407C can beused as a short term solution.

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Documents

Experimental investigation of a pilot compression chiller with alternatives refrigerants using compact heat exchanger