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Renewable Energy 57 (2013) 43e50

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Renewable Energy

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Energy saving mechanism analysis of the absorptionecompressionhybrid refrigeration cycle

Xuelin Meng a,b, Danxing Zheng a,*, Jianzhao Wang a, Xinru Li a

aCollege of Chemical Engineering, Beijing University of Chemical Technology, Beijing 100029, ChinabGuangxi Vocational and Technical Institute of Industry, Nanning 530001, China

a r t i c l e i n f o

Article history:Received 8 April 2012Accepted 10 January 2013Available online 16 February 2013

Keywords:Energy saving mechanismThermodynamic analysisHybrid refrigeration cycleR134a-DMF

* Corresponding author. Tel./fax: þ86 10 6441 6406E-mail address: dxzh@mail.buct.edu.cn (D. Zheng)

0960-1481/$ e see front matter � 2013 Elsevier Ltd.http://dx.doi.org/10.1016/j.renene.2013.01.008

a b s t r a c t

Focusing on the effective use of low-grade solar heat as heat source to provide refrigeration for resi-dential and commercial space cooling, an absorption-compression hybrid refrigeration cycle has beenstudied on the basis of available data of working pair 1,1,1,2-tetrafluoroethane (R134a) and dime-thylformamide (DMF). In order to investigate their performance, the energy saving mechanism of thehybrid cycle was analyzed, by means of thermodynamic diagrams of log peT, log peh and Tes. Theresults show that the hybrid refrigeration cycle has a relatively high thermodynamic perfectibility andcan use low-grade heat to replace parts of mechanical work for obtaining lower temperature refriger-ation effect owing to its energy complement and cascade refrigerating configuration between the in-ternal sub-cycles. Moreover, on the basis of two new criteria, the heat powered coefficient ofperformance and the electricity saving rate, the competition behavior between the sub-cycles of thehybrid cycle, i.e. the trade-off effects of compressor pressure on the low-grade heat utilization perfor-mance were also investigated. It was found that the sub-cycles compete in their contribution to thehybrid refrigeration system and the cycle preferences depend on the dominance which one achieves. Inother words, there is an optimum compressor outlet pressure region under specified working conditions,where the hybrid refrigeration cycle has the maximum heat powered coefficient of performance andelectricity saving rate.

� 2013 Elsevier Ltd. All rights reserved.

1. Introduction

Energy plays an important role in our fast moving life andeconomic development. In recent years, with the sharp increase inthe energy cost and the high energy consumption exhausted byrefrigeration, the energy saving refrigeration system has receivedconsiderable attention [1]. Distributed energy systems generallyrefers to the distributed combined cooling, heat, and power system,which provide multiple terminal energies at the same time in orderto bring about cascaded utilization of energy. And it has beenalready considered as an energy saving technology and a potentialsolution to rationalizing energy supply [2,3].

The absorptionecompression refrigeration cycle consists of anabsorption refrigeration cycle and a compressor. It is driven by low-grade heat and mechanical work (electricity), which is also calledthe hybrid refrigeration cycle, and can obtain lower temperaturecooling effect compared with the absorption refrigeration cycle.

..

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Therefore, the hybrid refrigeration cycle can be used as a new formto provide refrigeration for residential and commercial spacecooling, especially to be used for the hybrid solar energy andelectricity powered refrigeration system. For example, it can beconsidered as one of the distributed energy utilization technology,and the hybrid refrigeration cycle has been paid more and moreattention.

Boer et al. [4] have analyzed the hybrid refrigeration cycle andproposed that there are three kinds of effects to involve a com-pressor in an absorption refrigeration cycle. The first is to increasethe concentration difference of solutions during the generatingprocess and decrease the solution circulation ratio, and thenimprove the coefficient of performance (COP). The second is tocause the absorption process that can be executed at higher tem-perature, so the cooling water with higher temperature can beused. And the third is to decrease the generation temperature.According to the feature of the generation temperature to bedecreased, the triple-effect H2OeLiBr hybrid refrigeration cyclewas studied and it was found that remarkably decreased the gen-erating temperature and the request for corrosion protection ofthe high temperature generator [5]. Moreover, Rameshkumar and

Fig. 1. The schematic diagram of the hybrid refrigeration cycle.

Table 1Specified simulation parameters of major equipments.

Item Value

Minimum temperature difference in SHX, �C 5Superheated degree of the refrigerant vapor in RHX exit, �C 5Compressor isentropic efficiency 0.72Pump efficiency 0.5Mass fraction of the R134a in the vapor phase of the rectifier top 0.999

X. Meng et al. / Renewable Energy 57 (2013) 43e5044

Udayakumar [6,7] investigated the effects of compression ratio onthe hybrid refrigeration cycle performance and found an optimalvalue of the concentration difference and the compression ratio.Yari et al. [8] studied and compared the GAX and GAX hybrid ab-sorption refrigeration cycles from the viewpoint of both first andsecond law of thermodynamics, and an increase of COP and exer-getic efficiencies in the GAX hybrid absorption refrigeration cyclewas found. The performance of the NH3eLiNO3 hybrid refrigerationcycle also was studied. The result showed that the hybrid refrig-eration cycle had 10% improvement compared with the vaporcompression refrigeration cycle [9,10]. Especially, basing onUA � DTlm models for separate regions of plate-type heat ex-changers, Ventas et al. [11] studied a single-effect hybrid refriger-ation cycle and pointed out that the hybrid refrigeration cycleallows for working at lower driving temperatures than that of thesingle-effect cycle and with low electricity consumption. Satapathyet al. [12e14] investigated the hybrid refrigeration cycle forsimultaneous cooling and heating through theory and experiment,and pointed out that the hybrid refrigeration cycle yields muchbetter overall performance than that of the traditional absorptioncycle.

However, the studies of the hybrid refrigeration cycle just aimedat the case calculations and cycle experiments and falls short ofenergy saving mechanism study for the hybrid refrigeration cycle,though revealing the internal configuration of the cycle. Such aslike as thework proposed by Tang et al. [15], through the simulationfor a hybrid refrigeration cycle, he indicated that it can increases thecycle’s COP. Any way, Zheng et al. [16,17] have analyzed the exergycoupling of GAX cycle and exergy coupling of a novel absorptionpower/cooling combined cycle by thermodynamic diagrams, butthere have been few reports on the hybrid refrigeration cycle.

In addition, these studies mostly focused on using NH3eH2O asworking pair. The demerits of working pair NH3eH2O limit itsapplication. Because, the operating pressure is relatively high whenNH3eH2O is adopted, and NH3 has slight toxicity as well as moredriving heat has to be consumed for separating working pair. R134ais an ozone-safe refrigerant and is the leading replacement in theapplication of refrigeration systems [18]. The absorbent DMF isa widely and commonly used organic solvent. It has a very lowpartial pressure in solutionwith halogenated hydrocarbons [19]. SoR134a-DMF is a potential working pair for the hybrid refrigerationcycle. Some researchers [20] have studied R134a-DMF solution ina single-stage solar-powered absorption refrigeration cycle, but theheat source temperature (below 90 �C) is too low to drive thesingle-effect absorption refrigeration cycle for making the refrig-eration below 0 �C.

This paper aims at revealing the energy saving mechanism ofthe hybrid refrigeration cycle driven by low-grade solar heat andmechanical work. The hybrid refrigeration cycle with R134a andDMF on the reported vaporeliquid equilibrium data is simulated.On the results of that, a thermodynamic analysis for the hybridrefrigeration cycle is performed.

2. Description of the hybrid refrigeration cycle

The hybrid refrigeration cycle powered by mechanical work andlow-grade solar heat with a flat solar collector is schematicallyillustrated in Fig. 1, which contains evaporator (E), compressor(COMP), absorber (A), generator (G), rectifier (R), condenser (C),solution heat exchanger (SHX), refrigeration heat exchanger (RHX),flat solar collector (SC), refrigerant throttle valve (VR), solutionthrottle valve (VS), triple valve (TV), pump (P) and the pipes toconnecting them.

To drive the hybrid refrigeration cycle, the low-grade solar heatis derived from the flat solar collector, and mechanical work

(electricity) is provided to the compressor and solution pumps,while heat rejection occurs on the absorber and condenser. Thenthe building users exhibit a freezing benefit from the evaporator.

The following description provides more details. The wet vaporfrom the evaporator gets into the compressor after superheated inthe RHX. The compressed vapor is absorbed in the absorber by poorsolution coming from the generator. The rich solution exiting fromthe absorber is pumped into the generator through the SHX. Toemit refrigerant vapor, the rich solution is heated by low-grade heatin the generator, and is then throttled into the absorber. Therefrigerant vapor is rectified in the rectifier and then condensed inthe condenser. After supercooled in the RHX, the condensedrefrigerant liquid is throttled into the evaporator by the VR. Theliquid refrigerant evaporates in the evaporator to make refriger-ation effect occur, and then the cycle closes. The flat solar collectoris used to provide the low-grade heat (below 90 �C) for thegenerator.

However, the above description is for cycle operation in thehybrid mode (the state points are 1, 2, 3, 4, 5, 6, 7, 8, 9, 10, 11, 12 and13). By shutting off the valve before the compressor, opening theby-pass, and connecting the evaporator and absorber directly, thecycle is turned into the absorptionmode (the state points are 1a, 2a,3a, 4a, 5a, 6a, 7a, 8a, 9a, 10a, 11a and 12a). Otherwise, the cycle is inthe compression mode (the state points are 8b, 9b,10b,11b,12b, 14)while allowing the refrigerant vapor from the compressor outlet toenter into the condenser directly, thus cutting off connections withthe generator and absorber.

3. Simulation of the hybrid refrigeration cycle

The major equipment specifications used in the simulation arelisted in Table 1. Other assumptions are described as follows:

(a) The cycle runs in steady-state;(b) Pressure that drops within the cycle, e.g. each pipe line, can be

neglected except passing through the throttle valves;(c) The heat loss is ignored;(d) The outlet temperatures of absorber and condenser are kept

the same;

X. Meng et al. / Renewable Energy 57 (2013) 43e50 45

(e) The solution that outlets of generator and absorber lays insaturated states, and etc.

In this study, the PengeRobinson (PR) equation of state isselected to calculate thermophysical properties of the R134a-DMFsolution. The interaction parameter of the PR equation of stateis �0.041274, regressed by a set of experimental vaporeliquidequilibrium data of R134a-DMF solution reported by Zehiouaet al. [21], and the relative deviations on pressure (MRDP ¼ (100/N)

Pj( pexp � pcal)/pexpj) is 1.81%.The mass balance, energy balance and vaporeliquid equilibrium

for each unit of the hybrid refrigeration cycle, i.e. generator,absorber and compressor, etc, are related as follows:

Xj

ms ¼ 0 (1)

Xj

ðmhÞs þXk

ðQ þWÞu ¼ 0 (2)

f�p; T ; x

�¼ 0 (3)

For example, the mass balance and energy balance for theabsorber can be described as

m6 þm13 ¼ m1 (4)

m6h6 þm13h13 ¼ m1h1 þ QA (5)

For the R134a-DMF binary system, Eq. (3) can be re-written as

yibfvi

�T ;p; y

�¼ xibfl

i

�T ; p; x

�ði ¼ 1 for R134a; 2 for DMFÞ

(6)

In this work, bfi has been predicted by the PR equation of state.Therefore, each unit of the hybrid refrigeration cycle can bedescribed.

Usually, the variables of the process feed streams are specifiedand information flows parallel to thematerial flows. In other words,the calculations proceed from unit to unit, beginning with units forwhich all of the feed streams have been specified. On the basis ofsolving the equations of mass balance, energy balance and vaporeliquid equilibrium, the determinations for the temperatures, flowrates, and compositions, as well as the enthalpies, of each streamhave been performed using some suitable approximate method,e.g. the Edmister method [22].

Table 2The value of COPCC of the R134a compression refrigeration cycle (at TE ¼ �10 �C).

Item TC (�C)

30 35 40 45

COPCC 3.87 3.34 2.88 2.52

4. Analysis and performance results of the hybridrefrigeration cycle

4.1. Evaluation criteria

For the hybrid refrigeration cycle, researchers have used differ-ent evaluation criteria. Fukuta et al. [23] had proposed an evaluationcriterion, dividing refrigeration capacity by mechanical work con-sumption; however, the point is that the utilization of low-gradeheat was omitted unfortunately. Our previous work [24] had pro-posed a new criterion for evaluating the low-grade heat utilizationperformance in the hybrid refrigeration cycle. It is called the heatpowered coefficient of performance (u), and is defined as following:

u ¼ ðQE �W$COPCCÞ=QG (7)

where W is the mechanical work consumption (involving com-pressor work and pump work), and COPCC is the coefficient ofperformance of a compression refrigeration cycle at the sametemperatures of evaporation and condensation as those of thehybrid refrigeration cycle. The values of COPCC in different con-densation temperatures are listed in Table 2. Explicitly, u, whichdescribes the benefit for replacing mechanical work with low-grade heat, is the ratio of pure heat powered refrigeration capac-ity (cutting off that mechanical work powered) to the generatorheat consumption.

To analyse and assess the solar-powered advantages of thehybrid refrigeration cycle, the electricity saving rate (h) is definedby Eq. (8)

h ¼ ðzCC � zHCÞ=zCC ¼ 1� zHC=zCC (8)

where zCC and zHC express the mechanical work consumptions perunit refrigeration capacity for the compression refrigeration cycleand the hybrid refrigeration cycle at the same temperature ofevaporation and condensation, respectively.

4.2. The heat utilization performance of the hybrid refrigerationcycle

Corresponding to the cycle scheme of Fig. 1, as simulated results,the stream parameters of the hybrid refrigeration cycle (HC) arelisted in Table 3, under the conditions of the temperatures ofevaporation, condensation, absorption and generation are �10 �C,35 �C, 35 �C and 90 �C, respectively, and the compressor outletpressure (pout) is 350 kPa. For comparison, the streamparameters ofthe absorption refrigeration cycle (AC) and the compressionrefrigeration cycle (CC) at the same condition are also simulatedand shown in Table 3, while the evaporation temperature (TE) of theabsorption refrigeration cycle is 5 �C.

Table 4 shows the overall performance of the cycles HC and CC atthe same temperature of evaporation and condensation. It can beseen that the mechanical work consumption and the mechanicalwork consumptions per unit refrigeration capacity of the HC are0.258 kW and 0.141, respectively, while that of the CC are 0.544 kWand 0.298, respectively. The difference is that the additional2.999 kW low-grade heats are consumed by the HC. And the u andh for HC are 0.322 and 0.527, respectively. It means that the HC cansave 52.7% of electricity comparing with the CC.

5. Thermodynamic diagram illustrations

5.1. Refrigeration characteristic analysis by the log peT diagram

Corresponding to the cycle scheme and the stream numbers inFig. 1. Fig. 2 illustrates the changes of the temperature, pressure andconcentration of each stream of the HC. It is plotted from the data ofTable 3. The auxiliary line aeb illustrates the pure R134a line. Thedistance between state points and line aeb indicates the phasestate and concentration. If a state point is to the left of the line aeb,e.g. state point 9, it signifies that it is in an overcooled state. If it is tothe right of the line aeb, it indicates the temperature, pressure andconcentration of these points.

Fig. 2. Cycle comparison and analysis by log peT diagram.

Table 3Stream parameters of the cycles HC, AC and CC.

Cycle Stream T (K) p (kPa) Mass fractionof R134a

mi

(kg h�1)hi(kJ kg�1)

si(kJ kg�1 K�1)

HC 1 308.15 350 0.490 208.326 �6038.8 �4.5832 308.55 885 0.490 208.326 �6037.7 �4.5813 348.15 885 0.490 208.326 �5965.6 �4.3634 363.15 885 0.370 168.722 �5251.1 �4.6645 313.55 885 0.370 168.722 �5340.0 �4.9276 313.85 350 0.370 168.722 �5340.0 �4.9257 317.25 885 0.999 39.604 �8771.7 �2.4168 308.15 885 0.999 39.604 �8952.8 �3.0039 293.95 885 0.999 39.604 �8973.8 �3.073

10 263.15 200 0.999 39.604 �8973.8 �3.06211 265.45 200 0.999 39.604 �8807.8 �2.43112 288.45 200 0.999 39.604 �8786.7 �2.35513 311.65 350 0.999 39.604 �8768.9 �2.339

AC 1a 308.15 350 0.490 208.326 �6038.8 �4.5832a 308.55 885 0.490 208.326 �6037.7 �4.5813a 348.15 885 0.490 208.326 �5965.6 �4.3634a 363.15 885 0.370 168.722 �5251.1 �4.6645a 313.55 885 0.370 168.722 �5340.0 �4.9276a 313.85 350 0.370 168.722 �5340.0 �4.9257a 317.25 885 0.999 39.604 �8771.7 �2.4168a 308.15 885 0.999 39.604 �8952.8 �3.0039a 296.25 885 0.999 39.604 �8970.5 �3.061

10a 278.35 350 0.999 39.604 �8970.5 �3.05711a 280.75 350 0.999 39.604 �8797.7 �2.43612a 298.95 350 0.999 39.604 �8780.0 �2.375

CC 8b 308.15 885 1.000 44.500 �8956.3 �3.0019b 306.55 885 1.000 44.500 �8958.7 �3.009

10b 263.15 200 1.000 44.500 �8958.7 �2.99011b 263.15 200 1.000 44.500 �8811.0 �2.42812b 266.15 200 1.000 44.500 �8808.6 �2.41914 328.55 885 1.000 44.500 �8764.6 �2.381

X. Meng et al. / Renewable Energy 57 (2013) 43e5046

In order to analyse conveniently, the hybrid cycle configurationcan be considered as it consists of an absorption sub-cycle anda compression sub-cycle. In Fig. 2, the state change1 / 2 / 3 / 4 / 5 / 6 / (1) denotes the solution sub-cycle ofthe HC, i.e. the absorption sub-cycle, and the state change(3) / 7 / 8 / 9 / 10 / 11 / 12 / 13 / (1) denotes therefrigerant sub-cycle, i.e. the compression sub-cycle, where thestates 3 and 1 comes from and returns to the solution sub-cycle. Thecompressor boosts the pressure only from evaporation pressure200 kPa (point 12) to mid-pressure 350 kPa (point 13). However,the further boosting pressure up to 885 kPa (point 7) is completedby the absorption cycle; this process is usually called “thermalcompression” [25]. If the AC can be driven by the low-grade solarheat, e.g. 90 �C only, its refrigerant sub-cycle has state change(3a) / 7a / 8a / 9a / 10a / 11a / 12a / (1a). The corre-sponding temperature of point 10a is 5 �C, i.e. the refrigerationtemperature of the cycle. Obviously, it is higher than that of the HC.In other words, a better cooling benefit can be obtained using 90 �Clow-grade solar heat in the HC. Usually, to achieve this target, themore mechanical work must be consumed by the CC.

Table 4Overall performance of the cycles HC and CC.

Item HC CC

Heat input into generator, kW 2.999 e

Mechanical work input into compressor, kW 0.196 0.544Mechanical work input into pump, kW 0.062 e

Cooling energy output from evaporator, kW 1.827 1.826Mechanical work consumptions per unit

refrigeration capacity, z0.141 0.298

Heat powered coefficient of performance, u 0.322 e

Electricity saving rate, h 0.527 e

5.2. Energy saving mechanism analysis by log peh diagram and Tes diagram

For the cycle comparison and analysis, Figs. 3 and 4, the dia-grams of log peh and Tes, are plotted by the listed data in Table 3. InFig. 3, there are three sets of phase equilibrium auxiliary lineswhich the mass fractions are 0.999, 0.490 and 0.370 from left toright, respectively. In each set of phase equilibrium auxiliary line,the left side is the bubble-point line, while the right side is the dew-point line. Similar illustration is given in Fig. 4, i.e. the change orderof the mass fraction is from right to left, however the locations ofbubble line and dew line do not change. For comparison andanalysis to the HC, the illustration of the CC with same operatingcondition as the HC is added in Figs. 3 and 4, i.e. the state change8b / 9b / 10b / 11b / 12b / 14 / (8b).

In principle, the solution sub-cycle of the AC boosts the pressureof refrigerant using the low-grade heat, i.e. “thermal compression”of refrigerant. Therefore, it can be considered that there is no me-chanical work consumption except for a small amount of solutionpump work consumption in the AC. It makes cooling by low-gradeheat only. The extension of this understanding, in Fig. 3, the solu-tion sub-cycle of the HC is also performing the “thermal com-pression” with help of the input and output of the low-grade heat.Therefore, the differences of the mechanical work consumptionbetween the cycles HC and CC come into being. It can be seen that,

Fig. 3. Cycle comparison and analysis by log peh diagram.

Fig. 4. Cycle comparison and analysis by Tes diagram.

X. Meng et al. / Renewable Energy 57 (2013) 43e50 47

in Fig. 3, the length that the process line 12 / 13 projects onto theh-axis denotes the mechanical work consumption by the com-pressor of the HC. Similarly, that of the process line 12b / 14 de-notes the mechanical work consumption of the CC. It can be seenthat the result is consistent with the data listed in Table 4.

As shown as Fig. 4, the solution sub-cycle of the HC lies at theupper left in Tes diagram owing to it works between the drivingheat source temperature and ambient temperature; while therefrigerant sub-cycle lies at the bottom right, because it worksbetween the ambient temperature and refrigeration temperature.

Fig. 5. Effect of pout on u for the hybrid refrigeration cycle at TE ¼ �10

The relative position of two sub-cycles in Tes diagram describes thesituation of energy complement between the solution sub-cyclepowered by low-grade heat and refrigerant sub-cycle powered bymechanical work in the HC [16]. If only using low-grade solar heat,e.g. 90 �C, to power an absorption refrigeration cycle, the higherrefrigeration temperature, e.g. 5 �C, can be obtained, but the lowerrefrigeration temperature, e.g. �10 �C, cannot be obtained. How-ever, in the HC, due to the presence of energy complement betweenthe sub-cycles, the lower refrigeration temperature, e.g. �10 �C,effect can be obtained.

It can be inferred that the solution sub-cycle is responsible formaking the refrigeration temperature from ambient temperaturedown to 5 �C, while the compression sub-cycle is responsible formaking the refrigeration temperature down to lower tempera-ture �10 �C. On the second law of thermodynamics of view, it isa cascade refrigerating configuration [26]. Though the refrigeratingability of the solution sub-cycle is weak than that of the com-pression sub-cycle, however, through the energy complement be-tween them, not only the effect of lower refrigeration temperatureis achieved, but also the benefit of less energy consumption isobtained. On the view of the mechanism of the energy complementand cascade refrigerating configuration for the HC, it can be un-derstood why the HC exhibits relatively higher performance, i.e.higher values of u and h.

6. Competition behavior between the sub-cycles

In this work, the competition behavior between the sub-cyclesof the HC, i.e. the trade-off effects of compressor pressure on the

�C. (a) TG ¼ 90 �C; (b) TG ¼ 80 �C; (c) TG ¼ 70 �C; (d) TG ¼ 60 �C.

Table 5The values of popt, umax and hmax of the HC (at TE ¼ �10 �C).

TG (�C) TC ¼ TA (�C) popt (kPa) umax hmax Electricityconsumptionsper unit refrigerationcapacity, (kW)

90 30 270 0.418 0.671 0.08735 350 0.322 0.527 0.13940 480 0.273 0.417 0.20245 580 0.225 0.351 0.257

80 30 320 0.385 0.574 0.11035 420 0.308 0.450 0.16540 540 0.248 0.359 0.22345 650 0.198 0.297 0.279

70 30 370 0.340 0.485 0.13335 480 0.266 0.373 0.18840 580 0.209 0.296 0.24445 740 0.161 0.233 0.305

60 30 430 0.284 0.389 0.15835 540 0.213 0.295 0.21140 710 0.166 0.218 0.27245 880 0.121 0.161 0.333

X. Meng et al. / Renewable Energy 57 (2013) 43e5048

low-grade heat utilization performance under different generationtemperature (TG ¼ 60e90 �C) and condensation temperature(TC ¼ 30 �Ce45 �C) were investigated.

Fig. 5 shows the effect of pout on u for the HC. It can be seen that,at the given TG, TC, and TE, with the increasing of pout, u increasesspeedily to a maximum value at first, and then decreases. Thatmeans that during the increasing of pout, an optimum value of thepout (popt) exists and the maximum value of u (umax) appears cor-respondingly. In addition, with the given TG and TE, umax decreaseswith the increasing of TC, while the pout corresponding to umaxincreases; with the given TC and TE, umax increases with theincreasing of TG, while the pout corresponding to umax decreases.

Fig. 6 shows the effect of pout on h for the HC. Comparing Figs. 5and 6, it can be seen that the change of h is the same as that of uwith the change of pout, and umax and hmax almost depend on thesame value of pout under specified TG, TC and TE. In other words,there is an optimum pout (popt) that the HC has the maximum u andh with specified TG, TC and TE. The values of popt, umax and hmax ofthe HC at different conditions can be seen in Table 5. It can beseen that, in conditions of TG ¼ 60 �Ce90 �C, TC ¼ 30 �Ce45 �C andTE¼�10 �C, the values of umax and hmax are 0.121e0.418 and 0.162e0.671, respectively. It means that the HC can save the electricityfrom 16.1% to 67.1% compared with the CC at this condition.

For the varying curves of u and h shown in Figs. 5 and 6, it can beexplained by the virtue of competition behavior of the HC. Asmentioned above, the HC can be considered as a coupling config-uration of an absorption sub-cycle and a compression sub-cycle.However, two sub-cycles compete in their contribution to theoverall performance of the system.

Fig. 6. Effect of pout on h for the hybrid refrigeration cycle at TE ¼ �10

Before pout reaching to popt, with the increasing of pout, the so-lution concentration difference increases gradually from infin-itesimal and the circulation ratio decreases sharply from infinitevalue. In this case, the cycle performance of the absorption sub-cycle raises sharply; however, the compressor work consumptionincreases steadily. Therefore, the absorption sub-cycle becomes

�C. (a) TG ¼ 90 �C; (b) TG ¼ 80 �C; (c) TG ¼ 70 �C; (d) TG ¼ 60 �C.

X. Meng et al. / Renewable Energy 57 (2013) 43e50 49

dominant during this process, and u of the HC increases with theincreasing of pout.

After pout reaching to popt, with the increasing of pout, though thesolution concentration difference increases steadily, but thedecreasing of circulation ratio becomes gradual. Also, the increas-ing of cycle performance of the absorption sub-cycle is gettinggradual. However, the compressor work consumption increasescontinually, and u describes the benefit for replacing mechanicalwork with low-grade heat. In this case, the compression sub-cyclecontrols dominance in the HC, and u of the HC decreases with theincreasing of pout.

However, when pout is equal to popt, the contribution of thetwo sub-cycles is in balance, and u of the HC has the maximumvalue umax. The changing mechanism of h is the same as that ofu. Hence, u and h increase speedily to a maximum value at first,and then decrease, as shown in Figs. 5 and 6. In other words,the condition, pout ¼ popt, is the optimal operating condition,which exhibits optimal cycle performance and electricity savingbenefit.

7. Conclusions

In this paper, the absorptionecompression hybrid refrigerationcycle with R134a-DMF working pair has been studied, and someconclusions can be made as follows.

On the basis of two new criteria, the heat powered coefficient ofperformance (u) and the electricity saving rate (h), the low-gradeheat utilization performance in the hybrid refrigeration cycle isinvestigated.When the temperatures of evaporation, condensation,absorption and generation are �10 �C, 35 �C, 35 �C and 90 �C,respectively, and compressor outlet pressure is 350 kPa, the cor-responding u and h are 0.322 and 0.527, respectively. It means thatthe hybrid refrigeration cycle can save 52.7% of electricity com-paring with the compression refrigeration cycle.

The energy saving mechanism for the hybrid refrigeration cyclehas been analyzed by the thermodynamic diagrams of log peT,log peh and Tes. The results show that the low-grade heat can beused to replace parts of mechanical work for making lower tem-perature refrigeration effect in the hybrid refrigeration cycle, whereit cannot be achieved in the independent absorption refrigerationcycle. In fact, this is due to the energy complement and cascaderefrigeration configuration between the internal sub-cycles, i.e. itcan be considered as a coupling configuration of a compressionsub-cycle and an absorption sub-cycle.

The study result of the cycle coupling behavior between the sub-cycles indicates that the two sub-cycles compete in their con-tribution to the whole refrigeration system. There is an optimumcompressor outlet pressure region with specified working condi-tions, where the hybrid refrigeration cycle has themaximumvaluesof u and h, exhibiting the benefit of electricity saving. WhenTG ¼ 60 �Ce90 �C, TC ¼ 30 �Ce45 �C and TE ¼ �10 �C, the values ofumax and hmax are 0.121e0.418 and 0.162e0.671, respectively. Itmeans that the hybrid refrigeration cycle can save the electricityfrom 16.1% to 67.1% at this condition.

This work may prove that the hybrid refrigeration cycle hasresearch and development potential to use low-grade solar heat asheat source to supply refrigeration for residential and commercialspace in the future, e.g. some new technology in the like of dis-tributed energy system.

Acknowledgments

This work is supported by the National Natural Science Foun-dation of China (50890184, 51276010) and the National BasicResearch Program of China (2010CB227304).

Nomenclature

A absorberAC absorption refrigeration cycleC condenserCC compression refrigeration cycleCOP coefficient of performance [e]COMP compressorE evaporatorG generatorh specific enthalpy [kJ kg�1]HC hybrid refrigeration cyclem mass flow rate [kg h�1]p pressure [kPa]P pumpQ heat duty [kJ]R rectifierRHX refrigerant heat exchangers specific entropy [kJ kg�1 K�1]SC flat solar collectorSHX solution heat exchangerT Temperature [K] [�C]TV triple valveVR refrigerant throttle valveVS solution throttle valveW mechanical work [kW]x mole fraction for liquid phase [e]x one-dimensional array [e]y mole fraction for vapor phase [e]

Greek symbolsz mechanical work consumptions of unit refrigeration

capacity [e]bfi fugacity coefficient of species i [e]h electricity saving rate [e]u heat powered coefficient of performance [e]

SubscriptsA absorberC condenserCC compression refrigeration cycleE evaporatorG generatorHC hybrid refrigeration cyclei speciesj stream numberk unit-numberopt optimumout outlets streamu unit

Superscriptsl liquid phasev vapor phase

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