Khaled YousefMechanical Engineering Department,

Menoufia University,

Menoufia 32511, Egypt

e-mail: khyousef@msu.edu

Abebayehu AssefaMechanical Engineering Department,

Addis Ababa University,

Addis Ababa 31490 AA, Ethiopia

e-mail: abebayehu_assefa@yahoo.com

Ahmed HegazyMechanical Engineering Department,

Menoufia University,

Menoufia 32511, Egypt

e-mail: Ahegazy7@yahoo.com

Abraham Engeda1

Mechanical Engineering Department,

Michigan State University,

East Lansing, MI 48824-1226

e-mail: Engeda@egr.msu.edu

Comparative Study of UsingR-410A, R-407C, R-22, andR-134a as Cooling Mediumin the Condenser of a SteamPower PlantA step-by-step technique has been implemented in the analytical study of heat transferand pressure gradient characteristics of refrigerants R-410A, R-407C, R-22, and R-134aused as cooling media in the condenser of a steam power plant. Refrigerants are opti-mized to replace water/air as coolant in the condenser of a steam power plant. Refriger-ants have much lower temperatures and much higher heat transfer rates than water orair. The thermal resistances that affect heat transfer characteristics and surface con-denser performance are included. The effect of inlet refrigerant temperature and massflow rate are reported for the four refrigerants. Calculations are performed at two inletrefrigerant temperatures �21 �C and �30 �C and mass flow rate ranging from 92.905 to132.905 kg/s. The results revealed that the overall heat transfer coefficient, heat transferrate, and condensation rate increased with refrigerant mass flow rate, with higher valuesat lower inlet refrigerant temperatures. For a given refrigerant mass flow rate and inlettemperature, the analytical study indicated that R-410A has higher values of overall heattransfer coefficient, heat transfer rate and condensation rate than R-407C, R-22, andR-314a, respectively. Moreover, it is found that R-410A, at �30 �C and 132.905 kg/s, issuperior in condensing all steam entering the condenser than the other refrigerants; thiscorresponds to higher exergy efficiency. The condenser pressure was observed to beslightly higher for R-410A than the other refrigerants. [DOI: 10.1115/1.4028266]

Keywords: R-410A, R-407C, R-22, R-134a, thermal resistance, heat transfer coefficient,condenser performance, SPPC

Introduction

Steam power plant condenser (SPPC) is a shell and tube heatexchanger and is considered as one of the most important devicesin the power plant. In the SPPC, vapor is converted to liquid andlatent heat of condensation is transferred to coolant. An indirectcontact condenser is a common type of condenser where the vapordoes not mix with the cooling fluid; the shell and tube type is thecommon one and the most widely used. Increasing the heat trans-fer rate in the SPPC increases the thermal efficiency of the powerplant and also reduces the condenser size.

Several researchers have tried to optimize the design and/oroperation of SPPC. Haseli et al. [1] presented a study to evaluatethe optimum cooling water temperature during condensation ofsaturated water vapor within a shell and tube condenser. A theo-retical study for simulation of a thermal power plant at various cir-culation water flow rates was developed by Anozie and Odejobi[2]. The study was conducted to determine the optimum con-denser cooling water flow rate to ensure maximum efficiency,heat transfer area requirements, fuel saving and minimum operat-ing cost of the plant. Zhang, and Bokil [3] presented a quasi three-dimensional two phase numerical model to predict the fluid flowand heat transfer features in water cooled shell and tube con-denser. The condensation heat transfer rate for ammonia–watermixtures in a horizontal shell and tube condenser is determined by

Philpott and Deans [4]. An ideal Rankine cycle with reheat wasdescribed to study the effect of varying the condenser pressure onenergy and exergy of the power plant [5]. This study concludedthat the condenser pressure is an important parameter that affectsthe output power, power potential, thermal, and exergy efficiencyof the cycle. Moreover, the minimum allowable condenser pres-sure that produces maximum efficiency and power output wasderived. A numerical analysis using a porous medium model waspresented to develop a new condenser tube arrangement [6]. Thistube arrangement imparts better condenser performance and highheat transfer coefficient than the Heat Exchange Institute (HEI)Standards [7]. This new tube arrangement can also be used withhigher power units. A design module for estimating the condenserperformance by determining the condenser pressure at variousloads, cooling water temperature and tube cleanliness is describedby Pearce [8]. This work is related to work done by Davidson andRowe [9] and Hu and Zhang [10] who carried out an assessmenton the performance of different inundation correlations and dis-cussed optimum circulation water flow rate that is essentially im-portant to ensure maximum efficiency and minimum operatingcost of the power plant. They proposed a new inundation correla-tion to improve the accuracy in predicting the heat transferbetween steam and cooling water in tube-and-shell condensers.References [11–13] discussed the performance of thermal powerplants in different countries. They concluded that energy lossesmainly occurred in a condenser to the environment and in theboiler due to high exergy destruction. A similar study which wasspecific to conditions in California indicated that the inlet coolingcondition has an important effect on the condenser and the wholepower plant efficiency [14].

1Corresponding author.Contributed by the Industrial and Cogeneration Committee of ASME for

publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscriptreceived July 12, 2014; final manuscript received July 15, 2014; published onlineSeptember 10, 2014. Editor: David Wisler.

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Recently, the growing demand for water, for both domestic andindustrial use, has brought an increased attention in use of (dry)air cooled and hybrid (dry/wet) SPPC. There are few importantcurrent research and development works to improve heatexchanger geometries for finned tube bundles in dry-cooled con-densers [15,16], enhancement of dry-cooled condenser perform-ance with the use of limited water [17,18], the use of evaporativecondenser [19], the use of hybrid cooled condenser, where theheat of the wet condenser is dissipated into a cooling storage con-tainer [20]. The concept of ammonia dry cooling system isreported by Allemann et al. [21]. The system is an indirect type ofcondenser in which the circulating water loop is replaced by aphase-changing ammonia loop, where ammonia is evaporated inthe tubes of the steam condenser and condensed in an air-cooledcondenser. It is noted that most of the power generated duringpeak-load periods is used for driving the refrigerant compressor.This resulted in a great inefficiency and the demanded load wouldnot be covered during this period. For overcoming this problem,Hegazy [22] proposed the use of cooling thermal storage systemwith the combined steam plant and refrigerating cycle describedin Ref. [20]. This modification allows utilizing the discountedrates for energy during off peak loads to produce cooling that isstored in the cooling storage container. Over the period of thepeak-loads, the refrigeration machine is stopped and the coolingrefrigerant dissipates the heat absorbed from the steam condenserinto the cooling storage system. This allows the whole power gen-erated by the steam power plant to be exported to the grid.

The above described background leads to evaluating the con-cept of dry cooling by a refrigeration system in a SPPC; a studynecessary for testing the performance of the cooling method inconnection with the steam power plant performance and ambientconditions. Naturally, refrigerants have much lower temperaturesand much higher heat transfer rates than water or air. Further,refrigerants are optimized to replace water/air as a coolant forSPPC. The phasing-out of traditional chlorofluorocarbon (CFC)and hydrochlorofluorocarbon (HCFC) refrigerants results in thedevelopment and commercial application of hydrofluorocarbon(HFC) refrigerants. The CFC refrigerants contain chlorine thatcan combine with ozone in the stratosphere, thus depleting theozone layer. This has resulted in an intensified research forreplacement of HCFC-22 in recent years. Although HFCs havezero ozone depletion potential (ODP), they suffer from highglobal warming potential (GWP), and hence they are not particu-larly attractive from the environmental view point. New alterna-tive refrigerants should not only have low ODP but should alsohave low GWP, be safe, be reliable, be less flammable and be eco-nomical to be used in the existing facilities [23,24].

HFC-410A which is a mixture of HFC-32 and HFC-125 (50%and 50% by weight) has been considered as one of the primaryreplacements of R-22 in the air-conditioning system applicationsand in new equipment specifically designed for high operatingpressure because of its favorable thermodynamic properties ofheat transfer and environmental friendliness. HFC-134a replacedCFC-12 in domestic refrigeration and mobile air conditioning sys-tems, whereas HFC-407C became an alternative for HCFC-22 inchillers and heat pumps. HFC-407C, a nonazeotropic ternary mix-ture of HFC-32/HFC-125/HFC-134a (23/2025/52 wt.%), is theequivalent pressure replacement which can be used not only innew equipment but also in exiting equipment. R-410A has higherpressure and temperature than R-22 in both suction and dischargelines of the refrigeration system. Further, for the same refrigera-tion capacity, R-410A can only use 2/3 of the number of circuitsthat are required by R-22. Indeed R-410A requires small diametertubing, high heat transfer coefficient and lower refrigerant sideheat transfer exchange surface area. All these lead to compactheat exchanger size and less material requirement [23–25].

The heat transfer and flow characteristics of refrigerants havebeen studied by a number of researchers, and some examples ofthe most productive studies are described as follows. Lee et al.[23,26] presented an experimental study of heat transfer and

pressure gradient of hydrocarbon refrigerants R-290, R-600a, R-1270, and HCFC refrigerant R-22 during evaporating and conden-sation processes inside horizontal double pipe heat exchangers.Both the local evaporating and condensing heat transfer coeffi-cients of hydrocarbon refrigerants were higher than R-22. Further,increasing the mass flux increased the average heat transfer coeffi-cient. Higher values of pressure drop were also observed forhydrocarbon refrigerants than R-22. The heat transfer characteris-tics of R-410A, R-407C, and R-22 in a 6.4 mm inner diametertube was discussed by Ebisu and Torikoshi [27] while the pressuregradient for the same refrigerants is described in small diametertubes [28]. R-410A showed approximately 20% and 50% higherboiling heat transfer coefficients than R-22 and R-407C, respec-tively. R-410A achieved lower pressure gradient followed byR-407C and R-22, respectively. Christoffersen et al. [29] pre-sented a study for heat transfer coefficient and pressure drop dataof R-134a, R-22, and R-32/125 in smooth and microfin tubes. Forsmooth tubes, R-32/125 achieved higher heat transfer coefficientand lower pressure drop compared to R-22; R-134a, however,showed higher pressure drop and similar performance in heattransfer compared to R-22. Microfin tubes have higher heat trans-fer rates over the smooth ones. Experimental studies for heattransfer performance of R-22 and R-410A in smooth tubes, micro-fin tubes and horizontal tubes have been described Refs. [30–32].The study indicated that R-410 has higher heat transfer coefficientand lower pressure drop than R-22. Microfin tubes have higherheat transfer coefficient compared to smooth tubes but theenhancement in heat transfer performance was more effectivewith R-22 than R-410A, for microfin tubes.

In a series of experimental studies by Cavallini et al. [33–36], astudy of condensation heat transfer for R-134a, R-22, R-410A,and R-125 was carried out [33]. The studies revealed that R-410Ahas higher heat transfer coefficient followed by R-125, R-22, andR-134a, respectively. Heat transfer coefficients and pressure dropswere measured during condensation of R-134a, R-125, R-236ea,R-32, R-22, and R-410A inside a smooth tube [34]. It is foundthat heat transfer coefficient increases with increasing mass veloc-ity and vapor quality. R-236ea, R-134a, and R-22 gave better heattransfer performance than R-410A, which resulted in better per-formance than its components R-125 and R-32. Another compari-son is discussed between using R-134a and R-407C in microfintubes [35]. Microfin tubes gave better heat transfer coefficientover the smooth tubes at the same operating conditions. R-407Chad lower heat transfer coefficient than R-134a inside a microfintube. For multiport minichannel tube, the same authors presentedan experimental study for the pressure drop characteristics of R-134a, R-236ea, and R-410A [36]. The pressure drop of R-410Awas significantly lower than the other refrigerants under the sameconditions, while the pressure drop of R-236ea was the highestover the three refrigerants.

R-410A is near azeotropic refrigerant mixture, that means whenliquid and vapor are at equilibrium conditions, the composition ofvapor is very slightly different from the liquid composition. Thereare experimental studies that discussed heat transfer characteris-tics and pressure drop of R-410A in different kinds of tube surfa-ces (smooth/microfin, microchannel, multiport minichannel, flatand horizontal tube) [37–41]. The major findings of these studiesare heat transfer coefficient for R410A increased with mass fluxand with heat flux but it has higher values at lower saturation tem-perature. Pressure drop is directly proportional to the mass fluxand inversely to tube diameter. Also, it is found that small diame-ter tubes have higher heat transfer coefficient than tubes of biggerdiameter at the same conditions due to higher turbulence. More-over, microfin tubes, microchannel, and flat tube with higher flat-ness have higher heat transfer coefficient and pressure drop overthe smooth ones of the same hydraulic diameter. Experimentalinvestigations were carried out on heat transfer coefficient andpressure drop for R-410A [42] and for R-134a, R-410A, and R-236fa [43] inside a brazed plate heat exchanger. The results indi-cated that heat transfer coefficients have high sensitivity to heat

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flux and outlet conditions and weak sensitivity to saturation tem-perature. The frictional pressure drop also showed high sensitivityto refrigerant mass flux and weak sensitivity for both to saturationtemperature and outlet conditions. R-410A shows heat transfercoefficients of 40–50% higher than R-134a and 50–60% higherthan R-236fa and frictional pressure drops of 40–50% lower thanR-134a and 50–60% lower than R-236fa.

However, before these refrigerants can be accepted in SPPCcooling in the industry, fundamental heat and mass transfer andpressure gradient characteristics of these refrigerants need to beinvestigated for the optimal design of the heat exchangers.Recently, the authors had performed an extensive study on theheat transfer and pressure gradient characteristics of using wateror R-134a in SPPC cooling. R-134a achieved good performancecompared to water. Although the above studies, dealing with ei-ther the fundamental aspects of refrigerant heat transfer or heattransfer data can be implemented for optimizing SPPC cooling,the information on using refrigerants in SPPC is practically non-existent. The current study, therefore, attempts to fill in this gapby presenting the study of heat transfer characteristics and pres-sure gradient using four refrigerants, namely R-134a, R-22, R-407C, and R-410A in the SPPC cooling. The prediction of heattransfer characteristics in condensers is, therefore, a prerequisiteto increase the power plant cycle performance through the correctsizing of each power plant component. The effects of the refriger-ant mass flow rate and inlet temperature on heat transfer rate,overall heat transfer coefficient, condensation rate, exergy effi-ciency, and condenser pressure are presented.

Analytical Model Analysis

In this paper, four refrigerants, namely R-134a, R-22, R-407C,and R-410A, are investigated to evaluate the heat transfer charac-teristics, condensation rate and pressure gradient characteristics inthe SPPC cooling. In the process, steam is condensed as its latentheat of condensation is removed while the cooling fluid thatabsorbs this heat gets heated. Heat absorbed by the cooling fluid isdissipated in the evaporator of a refrigeration cycle. This cycle isnot included in the analysis given in this paper.

In actual cycles, condensate is collected at the bottom of thecondenser and pumped again into the cycle. From the review forSPPC cooling, there are six main thermal resistances that affectthe heat transfer rate and surface condenser performance[3,7,8,12,13,44–47]. These are the inner and outer tube film ther-mal resistance, tube wall thermal resistance, noncondensablegases resistance, and fouling resistances on the inside and outsideof the tubes. In this study, the condenser tubes are assumed cleanand the fouling resistances on the inside and outside of the tubesare neglected [8]. Also, it is assumed that during condensation ofthe steam in the steam condenser both water and steam are in ther-modynamic equilibrium and there is no freezing of water. Theoverall thermal resistances are the sum of the above four thermalresistances which are discussed as follows.

First. When the refrigerant flows inside the tube, it creates athin boundary layer on the tube wall; this layer will build a ther-mal resistance on the direction of the heat flow. There are manyempirical correlations in the literature that estimate the convectiveheat transfer coefficient. One of these correlations reported inRefs. [8] and [10] is described as follows:

Rf�in ¼1

hc�in¼ 0:036

kcl

din

Re0:8cl Pr

1=3cl

L

din

� ��0:054" #�1

(1)

where

Recl ¼qcldinv

lcl

; Prcl ¼lclCpcl

kcl

; v ¼ _mcl

qclNpd2in=4

(2)

This equation is valid only for 10 < L=din < 400 and all proper-ties are evaluated at Tb given by Tb ¼ ðTcl�inlet þ Tcl�outletÞ=2.

Second. When the hot steam comes in contact with the coldexterior surface of the condenser tubes, the temperature of thesteam drops to the wall temperature. Because of the phase change,a thin film of water (condensate) is formed on the outer surface ofthe tubes. This film creates thermal resistance for heat transferfrom the steam to the coolant. Nusselt approximated the thermalresistance for heat transfer for a single horizontal smooth tube forlaminar flow from the correlation [10]

Nucs ¼hc�outdout

kcs

¼ 0:728dout

kcs

qcs qcs � qsð Þgh�fgk3cs

doutlcs Tf � Twð Þ

" #�0:25

(3)

The modified latent heat of condensation h�fg accounts for the cool-ing of the liquid below the saturation temperature [13].

h�fg ¼ hfg þ 0:68Cpcs Tsat � Twð Þ (4)

A simple correlation, using the row number of the tube as themain parameter, is presented to account for the inundation effectsat row i [9,10].

FI ¼ i3=4 � i� 1ð Þ34 (5)

The vapor shear correlation or correction term are proposed as fol-lows [48]:

Cvg ¼ 1þ 0:0095 Re11:8=

ffiffiffiffiffiffiffiNucs

p

mix (6)

The correlation for the condensate thermal resistance that com-bines both the inundation and vapor shear effects is

Rf�out ¼kcs

dout

FI Nucsð1þ 0:0095 Re11:8=

ffiffiffiffiffiffiffiNucs

p

mix � ��1

(7)

Remix ¼qmixdoutv1

lmix

(8)

The free stream velocity v1 depends on the condenser tube row[13] as given below:

at first row v1;1 ¼_ms þ _ma

qsAvd

; Avd ¼ wtsL (9)

at the other rows v1;i>1 ¼ð _ms þ _maÞ �

Xn

i¼1

_mcs;i

qmixAmv

;

Amv ¼ LðNTR� 1Þ Pt �pd2

out

2ffiffiffiffiffiffiffi3Pt

p� �

(10)

Third. Existence of the noncondensable gases in the powerplant condenser increases the condenser pressure and thus lowersthe power output. Moreover, noncondensable gases cover theouter surface of the condenser tubes which decrease the heattransfer from hot steam to the cooling refrigerant. There are twosources of noncondensable gases. The first source is air from theatmosphere that leaks into the vacuum condenser. The othersource of noncondensable gases is the decomposition of water anddissolved gases in the feed water [5,49]. An empirical expressionfor the thermal resistance due to the presence of noncondensablegases is given by Refs. [3,10,48],

Ra ¼ a Res

D

dout

1

ðTsat � TfÞ1=3

pc

pc � ps

� �b

p1=3c qs

L

Tsat

� �2=3" #�1

(11)

where a and b are constants that depend on the Res value

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a ¼ 0:52 and b ¼ 0:7 when Res � 350

a ¼ 0:82 and b ¼ 0:6 when Res > 350

D is the diffusion coefficient which is calculated from [13],

D ¼ 0:926

pc

¼ T2:5

T þ 245

� �and T ¼ Tsat þ Tf

2(12)

pc is the condenser pressure and equals pc ¼ ps þ pa, where pa isassumed constant. The calculation of pc depends on the number oftube rows as follows [13,50]:

For the first row; pcði¼1Þ ¼ ps�inlet 1þ 0:622eð Þ; e ¼ _ma= _ms

(13)

For the other rows;

pcðiþ1Þ ¼ pcðiÞ 0:2C2r

€m2v

qs

� �� €m2

v

qs

� �i

� €m2v

qs

� �iþ1

� �� �(14)

where Cr ¼ Avd=Amv and €mv ¼ qsv1;i.The vacuum pressure inside the condenser can be calculated as

follows [13,50]:

pvac ¼patm � pc

patm

� �� 750:061 (15)

In this study, the noncondensable gases fraction is assumed to beconstant throughout the condenser shell side [10].

Finally. The wall thermal resistance depends on the thermalconductivity of the tube material. From the heat transfer textbooks[51], Rw can be calculated as follows:

Rw ¼dout

2km

lndout

din

� �(16)

From the above discussed thermal resistances, the total condenserthermal resistance, overall heat transfer coefficient and heat trans-fer rate are calculated as follows [3,51]:

Rtot ¼ Rf�in

dout

din

þ Rf�out þ Ra þ Rw (17)

U ¼ 1

Rtot

(18)

_Q ¼ UAoutLMTD (19)

where LMTD ¼ ðTcl�outlet � Tcl�inletÞ=ln ðTsat � Tcl�inletÞ=ðTsatð�Tcl�outletÞÞ and Aout ¼ pdoutLNTR.

To assess the condenser performance, destruction energy rate,Blasius friction factor and exergy efficiency have to be calculatedfrom the following equations [7,41]:

_Einlet ¼ _mclecl�inlet þ _ms�inletes�inlet (20)

and

_Eoutlet ¼ _mclecl�outlet þ _mcs�outletecs�outelt þ _ms�outletes�outlet

where e ¼ h� ho � Toðs� soÞ (21)

_Ed ¼ _Einlet � _Eoutlet (22)

Considering no exergy loss across the condenser, the exergy effi-ciency is defined as [45]

gex ¼ 1�_Ed

_Einlet

(23)

Condenser Configuration

The condenser configuration model which is used in this studyis adapted from Hu and Zhang study [10] which was in turnadapted from a previous experimental study made by Al-Saneaet al. [47] and Bush et al. [52]. In existing configuration, the re-frigerant flows inside the tubes of the condenser while the steamflows on the shell side from up to down. The condensate is col-lected at the bottom of the condenser. The geometric and operat-ing parameters of the condenser model are shown in Table 1.

The following results are obtained using engineering equationsolver (EES) program (version 6.804) for R-410A, R-407C, R-22,and R-134a at two inlet refrigerant temperatures: �21 and �30 �Cand mass flow rate range from 92.905 to 132.905 kg/s. The high-est inlet refrigerant flow rate was set to 132.905 kg/s as this is therequired refrigerant mass flow rate to condense all steam for thebest refrigerant “R-410A” at inlet temperature of �30 �C.

Solution Procedure

In this study, step by step technique is used to evaluate heattransfer and pressure gradient characteristics. The calculations arestarted from the condenser top, where the steam enters, and con-tinued to its bottom. The solution procedures implemented for cal-culating thermal properties and condenser parameters aresummarized as follows:

(1) Geometric and operating parameters are specified for boththe vapor and the refrigerant in the code.

(2) An iterative approach, as described in Refs. [46] and [47],is implemented to calculate the wall and film temperaturesat each row using EES procedures and functions.

(3) Thermodynamic properties for all refrigerants and steamare obtained using EES functions at each row as a functionof the local properties.

(4) Once the solution has converged for this row, all thermalresistances, total resistance, overall heat transfer coeffi-cient, heat transfer rate, condensate mass flow rate, con-denser and vacuum pressures are calculated for each row.

(5) The exit condition from a tube row is assumed to be theentering condition for the next row and so on until the exitfrom the last tube row.

(6) The properties are updated and steps 2 to 5 are repeated forthe next vertical rows.

(7) Based on the thermodynamic properties at the first and finalrows of the condenser, for both the cooling refrigerant andthe vapor, the exergy efficiencies are computed.

Table 1 Geometrical and operational parameters of thecondenser model

Geometric and operational parameters Steam/refrigerants

Condenser length (L) m 1.219Condenser depth or width of tube sheet (wts) m 1.02Condenser height, m 0.78Tube outer diameter (dout) mm 25.4Tube inner diameter (din) mm 22.9Tube pitch (pt) mm 34.9Tube material (titanium G-2)thermal conductivity, W/m K

17

Number of rows (N) 18NTR 22Inlet refrigerant temperature (Tcl�inlet)

�C �21/�30

Inlet refrigerant pressure (pcl) bar 4Inlet refrigerant flow rate ( _mref ) kg/s 92.905�132.905Inlet steam pressure (ps�inlet) bar 0.27670Inlet steam flow rate ( _ms) kg/s 2.032Inlet air flow rate ( _ma) kg/s 2.48� 10�4

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Comparative Results and Discussion

In this paper, a comparative study of using R-410A, R-407C,R-22, and R-134a as cooling fluids in SPPC is performed. Themain parameters that are used to investigate the condenser per-formance are calculated using the above mentioned equations foreach refrigerant using EES software and reported in Tables 2 and3. Also, this technique is used to simulate the same experimentaloperating and geometrical conditions and tube material as in Refs.[47] and [52] for water as a coolant. The validation gives a reason-able agreement where the predicted accumulated condensate was1.925 kg/s but the experimental value was 2.021 kg/s.

Figures 1(a) and 1(b) and 2(a) and 2(b) describe comparativeheat transfer rate and overall heat transfer coefficient across thecondenser rows for two different inlet refrigerant temperaturesand mass flow rates, respectively. As can be seen, heat transfercharacteristics decrease continuously across the condenser rowsirrespective of the refrigerant type. Decreasing heat transfer rateand overall heat transfer coefficient across the condenser rowsfrom top to bottom are observed which are results of the inunda-tion phenomena and the condensate blanket created on condensertubes. As expected, the heat transfer rate and overall heat transfercoefficient increase with the refrigerant mass flow rate; they are

strongly dependent on mass flow rate. The heat transfer rate hashigher values at lower inlet refrigerant temperature while theoverall heat transfer coefficient has lower values. This is attributedto the fact that all the refrigerants have higher heat capacity atinlet refrigerant temperature of �21 �C than at �30 �C. Adecrease in other thermal properties results in higher thermal re-sistance at �30 �C and hence lower overall heat transfer coeffi-cient. However, the increase in the logarithmic mean temperaturedifference is higher than the decrease in the overall heat transfercoefficient which reveals higher heat transfer rate at �30 �C thanat �21 �C. For a given refrigerant mass flow rate and inlet temper-ature, it can be noted that R-410A shows higher heat transfer char-acteristics followed by R-407C, R-22, and R-134a, respectively.This higher heat transfer performance is attributed to the superiorthermophysical properties of R-410A over the other refrigerants.For the same conditions, R-410A has higher thermal conductivityand heat capacity but lower viscosity than R-407C, R-22, and R-134a, respectively.

Condensation rate is a very effective parameter in condenserperformance evaluation. As higher the condensation rate is thebetter condenser performance will be. R-410A provides highercondensation and total condensation rates than R-407C, R-22, andR-134a, respectively as depicted in Figs. 3 and 4. That is because

Table 2 Comparison of refrigerants parameters at _mref 5 100 kg/s and two different temperatures

R-134a R-22 R-407C R-410A

221 �C 230 �C 221 �C 230 �C 221 �C 230 �C �21 �C �30 �C

gex %ð Þ 74.44 75.07 82.281 82.96 83.99 84.27 87.4 87.5_Ed 105W� �

8.657 8.736 8.757 8.858 8.822 8.921 9.067 9.183f*10�3 5.712 5.9 5.316 5.458 5.344 5.504 5.081 5.228Total condensation rate (%) 56.939 60.24 62.75 67.323 65.0 69.24 77.165 82.78

Table 3 Comparison of refrigerants parameters at _mref 5 132.905 kg/s and two different temperatures

R-134a R-22 R-407C R-410A

221 �C 230 �C 221 �C 230 �C 221 �C 230 �C �21 �C �30 �C

gex %ð Þ 78.65 79.24 86.01 86.09 86.97 87.16 89.86 91.87_Ed 105W� �

8.907 8.987 9.01 9.15 9.114 9.255 9.363 9.533f*10�3 5.32 5.495 4.921 5.083 4.976 5.127 4.733 4.869Total condensation rate (%) 69.39 73.97 76.132 82.33 78.84 84.55 92.815 100

Fig. 1 Heat transfer rate across condenser rows

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condensation rate is directly proportional to the heat transfer rate.It is obvious from Fig. 4 and Table 3 that R-410A has the capabil-ity of condensing all the steam at temperature �30 �C and massflow rate of 132.905 kg/s. Consequently, R-410A proves to be thebest refrigerant that requires minimum mass flow rate to condenseall the steam entering the condenser. Decreasing the refrigerantmass flow rate decreases the condenser size which results inhigher condenser performance and cycle efficiency. The effect ofinundation phenomena appears on both graphs especially onFig. 3.

Figures 5 and 6 show the condenser and vacuum pressuresacross the condenser rows for all four refrigerants at two inletrefrigerants temperatures of �21 �C and �30 �C, respectively.The condenser pressure is the sum of partial pressure of steam andthe partial pressure of the noncondensable gas. Since the noncon-densable gas pressure is assumed to be constant, the trends of thecurves of the steam pressure and condenser pressure are similar. Itis observed that the condenser pressure decreases slightly fromcondenser inlet to its exit and vice versa for the vacuum pressure.This may be attributed to following possible explanation. Steamenters the condenser in a saturated or wet vapor state at saturatedpressure. Saturation temperature is a direct function of the steampressure while flow resistance is directly proportional to steam ve-locity. Steam pressure decreases from condenser inlet to its exitbecause of the flow resistance and results in the decrease of the

Fig. 2 Overall heat transfer coefficient across condenser rows

Fig. 3 Condensation rate across condenser rows Fig. 4 Total condensation rate across condenser rows

Fig. 5 Condenser pressure across condenser rows

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saturation temperature as well. A condenser is characterized bythe decreasing vapor/liquid ratio from the first row to the last rowdue to condensing of steam. R-410A has lower vapor/liquid ratiocompared to the other refrigerants as inferred from Fig. 4, for thesame flow area and inlet steam flow rate. As a consequence of theuncondensed steam that remains in each row, which is lower forR-410A than the other refrigerants, lower steam flow rate andpressure drop are observed for R-410A. Vacuum pressure isdirectly proportional to the condenser pressure as shown in Eq.(15). This relation can be used to calculate the vacuum pressuregenerated from rapid condensation of R-410A. As a result ofdecreasing of condenser pressure, as discussed above, the vacuumpressure increases across condenser rows with the same trend.Consequently, slightly higher condenser pressure and lower vac-uum pressure are produced for R-410A compared to R-407C, R-22, and R-134a, respectively.

The effect of refrigerant mass flow rate on the total condensa-tion rate is depicted for two different temperatures in Fig. 7. ForR-410A, increasing the mass flow rate increases the total conden-sation rate linearly until the steam is fully condensed at �30 �Cand 132.905 kg/s. As a result, the best refrigerant R-410A will beconsidered as the cooling refrigerant in the experimental setupwhich is under construction to simulate a SPPC. Figure 8 showsthe condenser pressure drop across the condenser inlet and exit asa function of refrigerant mass flow rate. It is obvious that the con-denser pressure drop decreases with the mass flow rate of the re-frigerant. The condenser pressure drop shows a linear dependenceon the refrigerant mass flow rate. At the same flow rate, all refrig-erants achieved lower condenser pressure drop at lower inlet re-frigerant temperature. This is because of the relation betweencondenser pressure and total condensation rate as discussed above.R-410A has lower pressure drop at �30 �C and 132.905 kg/swhich is related to its condenser pressure as shown in Fig. 5.

Tables 2 and 3 provide the quantitative assessment of using R-134a, R-22, R-407C, and R-410A as a cooling fluid in SPPC. It isshown that, R-410A has higher exergy efficiency, destructionenergy rate, and total condensation rate compared with R-407C,R-22, and R-134a, respectively, at all operating conditions. Inaddition, R-410A has lower friction coefficient which revealslower pressure drop in refrigerant side. The refrigerant side pres-sure drop is small due to lower flow velocity in this flow region.Exergy efficiency, destruction energy rate, and total condensationrate increase with refrigerant mass flow rate until all the steam arefully condensed at �30 �C and mref¼ 132.905 kg/s which corre-sponds to higher exergy efficiency for R-410A.

Conclusion

In the present study, an analytical comparison between usingfour different refrigerants, namely, R-134a, R-22, R-407C, and R-410A as a cooling fluid in a SPPC is presented. These refrigerantsproved to be alternative coolants in a SPPC to air/water. The heattransfer characteristics, condensation rates, and condenser pres-sures are analyzed with respect to the condenser rows, refrigerantinlet temperatures, and mass flow rates. The results indicated thatR-410A have better heat transfer characteristics and condensationrates followed by R-407C, R-22, and R-134a, reflecting better per-formance as a SPPC coolant. R-410A reveals higher exergy effi-ciency and destruction energy rate than the other refrigerants.Moreover, lower friction coefficient has been realized for R-410Awhich reflects lower pressure drop on the refrigerant side. Theseconcluded remarks shall be experimentally verified on the test rigthat is under construction to analyze the SPPC in which R-410A isused as a cooling medium.

Fig. 6 Vacuum pressure across condenser rows Fig. 7 Total condensation rate versus refrigerant mass flowrate

Fig. 8 Condenser pressure drop versus refrigerant mass flowrate

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Nomenclature

Amv ¼ mean void area in the vapor space (m2)Asur ¼ total heat transfer area at a given

control volume (m2)Avd ¼ cross-sectional area of the vapor duct m2

cp ¼ specific heat (J/k kg)cvg ¼ vapor shear correction term

d ¼ tube diameter (m)D ¼ diffusion coefficient (m2/s)e ¼ specific exergy flow (J/kg)_E ¼ exergy rate (W)

FI ¼ inundation correction factorg ¼ gravity (m/s2)h ¼ enthalpy (J/kg K)

hc ¼ convective heat transfer coefficient (W/m2 K)h*

fg ¼ modified latent heat of vaporization (J/kg K)k ¼ thermal conductivity (W/m K)L ¼ length of the condenser tube (m)

LMTD ¼ log mean temperature difference (�C)_m ¼ mass flow rate (kg/s)

€mv ¼ mass velocity (kg/m2 s)N ¼ number of tubes inside the condenser

Nu ¼ Nusselt numberNTR ¼ number of tubes in each row

p ¼ pressure (bar)pt ¼ tube pitch (m)Pr ¼ Prandtl numberQ ¼ heat transfer rate within the condenser (W)R ¼ thermal resistance (m2 K/W)

Re ¼ Reynolds numbers ¼ specific entropy (J/kg K)T ¼ temperature (K)U ¼ overall heat transfer coefficient (W/m2 K)v ¼ flow velocity (m/s)

wts ¼ width of the tube sheet (m)

Greek Symbols

e ¼ relative content of the gasesgex ¼ exergy efficiency

l ¼ dynamic viscosity (Pa s)q ¼ density (kg/m3)

Subscripts

a ¼ noncondensable gasesatm ¼ atmosphere

b ¼ bulkc ¼ condenser

cl ¼ coolantcs ¼ condensated ¼ destructionF ¼ filmin ¼ innerm ¼ metal

mix ¼ mixture (steam/air)o ¼ dead state conditions

out ¼ outers ¼ steam

sat ¼ saturatedtot ¼ total

vac ¼ vacuumw ¼ wall1¼ free stream

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