Performance evaluation of a modified refrigeration cycle with parallel compression for refrigerator-freezer applications

Energy 188 (2019) 116093

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Performance evaluation of a modified refrigeration cycle with parallel compression for refrigerator-freezer applications Zhongcheng Fang, Chaochao Fan, Gang Yan, Jianlin Yu* Department of Refrigeration & Cryogenic Engineering, School of Energy and Power Engineering, Xi'an Jiaotong University, Xi'an, 710049, China

a r t i c l e i n f o

a b s t r a c t

Article history: Received 20 April 2019 Received in revised form 16 July 2019 Accepted 7 September 2019 Available online xxx

This paper proposes a modified refrigeration cycle (MRC) with zeotropic mixture R290/R600a for refrigerator-freezer applications. In the MRC, the use of two parallel compressors enables the two evaporators to be operated at two proper evaporation temperatures, which reduces the irreversibility of evaporator heat transfer processes. Noted that a subcooler is installed between the two loops for improving the system performance. Moreover, the MRC applies a phase separator to ensure more lowboiling point refrigerant to go to the freezer evaporator for increasing the freezer evaporation pressure. Energy and exergy performances of MRC are investigated based on a thermodynamic model. The simulation results indicate that compared with the conventional refrigeration cycle (CRC), the COP and exergy efficiency of MRC obtain the same improvement of 30.4% at a fixed condensation temperature of 35
C. In addition, from the viewpoint of exergy analysis, the high priority in need of improvement for the MRC is ordered as the compressors, condensers, expansion valves, two evaporators, subcooler and internal heat exchanger. Generally, applying the MRC could be an effective and practical method to enhance refrigerator-freezer energy efficiency. © 2019 Elsevier Ltd. All rights reserved.

Keywords: Refrigerator-freezer Parallel compression Zeotropic mixture Exergy analysis

1. Introduction In recent years, due to the improved living standards, refrigerator-freezers are widely used in households, which consume much electrical energy. According to the statistics, refrigerator-freezers consume about 6% of the electrical energy generation around the world [1]. Many approaches have been proposed to enhance the energy efficiency of refrigerator-freezers such as selecting alternative refrigerants, optimizing the structures of devices and applying novel refrigeration cycles, etc. [2,3]. In recent years, refrigerant substitution has been a research hotspot, especially, many literatures focused on applying zeotropic mixtures for further improving the performance of refrigerator, heat pump and air conditioning systems due to their temperature glide and composition separation features [4e6]. In detail, for refrigeratorfreezer applications, hydrocarbon zeotropic mixtures are the leading choices amongst the suitable working fluids owing to their environmentally friendly features [7e9]. Actually, among the hydrocarbon zeotropic mixtures, R290/R600a becomes a great choice

* Corresponding author. E-mail address: [email protected] (J. Yu). https://doi.org/10.1016/j.energy.2019.116093 0360-5442/© 2019 Elsevier Ltd. All rights reserved.

for refrigerator-freezer applications [10,11]. Rasti et al. [12] conducted an experimental study on substituting R134a with R290/ R600a in a refrigerator and then obtained about 5.3% lower energy consumption compared with the R134a system. He et al. [13] have applied the mixture of R290/R600a in a large capacity chest freezer and found that the electricity consumption can be reduced by 27.5% in comparison with the R134a system. Once the zeotropic mixture is applied, its composition separation feature could be used for further enhancing the system overall performance, for example, Yan et al. [14] have proposed a R290/R600a refrigeration cycle with separation condensation, which installs a phase separator in the middle of the condenser to enable more R290-enriched refrigerant to go to the freezer evaporator for increasing the evaporation pressure. Furthermore, Chen et al. [15] added an ejector on the basis of this above cycle and obtained 14% higher COP compared with the conventional refrigeration cycle. Generally, using zeotropic mixture and separation condensation technique is one of the promising ways to enhance the system performance of a refrigerator-freezer. Traditionally, the single stage vapor compression refrigeration cycle is usually used in refrigerator-freezers. As known, the most common refrigerator-freezer has two temperature compartments. From a thermodynamic viewpoint, the application of two proper evaporation temperatures in the refrigerator-freezer systems can

2

Z. Fang et al. / Energy 188 (2019) 116093

Nomenclature COP e E h m_ p q s t T w x y z MRC CRC IHX

coefficient of performance () specific exergy (kJ$kg1 ) exergy (W) specific enthalpy (kJ$kg1 ) refrigerant mass flow rate (kg,s1 ) refrigerant pressure (kPa) cooling capacity (W) specific entropy (kJ $kg1 $K1 ) refrigerant temperature (
C) temperature (K) compression work (W) refrigerant quality () payback period of investment (year) refrigerant mass fraction of R290 () modified refrigeration cycle conventional refrigeration cycle internal heat exchanger

Greek symbols efficiency difference compression ratio f exergy destruction ratio 4 cooling capacity ratio

h D g

reduce heat transfer irreversibility of the evaporators due to the small heat transfer temperature difference and further improve the system energy efficiency. In addition, as the household refrigerators have become larger in terms of capacity and energy efficiency regulations have been strengthened, technologies of applying two compressors to the conventional refrigerator-freezer have been developed [16]. Yoon et al. [17] further carried out optimizations of dual loop cycles using R-600a and hydrocarbon (HC) mixtures, and confirmed that the optimized dual-loop cycles can significantly decrease the refrigerator/freezer energy consumptions due to using an additional compressor. Tang et al. [18] have conducted an experimental study on the performance of dual loop cycles with a subcooler installed between the two loops, and then indicated that using the subcooler can decrease energy consumption. Besides, applying parallel compression cycle with two compressors is an effective way to advance the improved energy efficiency in refrigerator-freezers. Yang et al. [19] conducted a comparison study on four kinds of two compressor cycles for domestic refrigeratorfreezer applications. Actually, many researches on parallel compression cycle for CO2 refrigeration application have been conducted [20,21]. From these above researches, it can be said that parallel compression cycle shows better performance than one compressor cycle [22]. As introduced above, in previous literatures [15e19], it is indicated that the separation condensation and parallel compression techniques are the promising technologies for enhancing the performance of a refrigerator-freezer. However, it is found that the research on combining these two techniques for refrigeratorfreezer applications is still lacking. Therefore, the objective of this paper is to propose a modified refrigeration cycle (MRC) with R290/ R600a by integrating of separation condensation and parallel compression techniques. Noted that a subcooler is installed between the two loops of MRC for further enhancing the system performance. Energy and exergy analysis methods are applied to evaluate the thermodynamic performance of MRC. Furthermore,

Superscripts T thermal M mechanical Subscripts 1e13 a1-a8 c D er ef r f F is j k L P comp cond evap OHX EV mix sepa

refrigerant state point air state point condensation destruction refrigerator evaporator freezer evaporator refrigerator compartment freezer compartment fuel isentropic state point component loss production compressor condenser evaporator other heat exchangers (subcooler and IHX) expansion valve mixing process separator

from the viewpoint of exergy analysis, the priority order that in need of improvement for MRC components is illustrated. Detailedly, the system critical parameters such as cooling capacity ratio, refrigerant mass fraction of R290 and condensation temperature are investigated, which can give a significant guidance for designing a MRC system. Additionally, the performance characteristics of MRC are compared with those of the conventional refrigeration cycle (CRC). 2. Cycle description The CRC system usually consists of a compressor, a condenser, an expansion valve, a subcooler, and two evaporators in series (one for freezer compartment and the other for refrigerator compartment). This system operates the two evaporators in series at a same evaporation pressure. Consequently, there is a large temperature difference between the refrigerator compartment and refrigerator evaporator, which creates great heat transfer irreversibility of refrigerator evaporator. Hence, the efficiency of this CRC system would be lower due to the use of an identical evaporation pressure for the two evaporators. However, this major disadvantage of the CRC could be overcome by a modified refrigeration cycle (MRC) proposed here. In contrary to CRC, the MRC applies two independent compressors to enable the two evaporators to be operated at different evaporation temperatures properly matched with the refrigerator compartment and freezer compartment requirements, respectively. Besides, in the MRC, zeotropic mixture R290/R600a is applied, which shows temperature glide and composition separation characteristics. Based on this point, a phase separator is set between the two condensers, which can ensure more low-boiling point refrigerant to go to the freezer evaporator for increasing the freezer evaporation pressure. Furthermore, in MRC, a subcooler is installed between the two loops, which is beneficial to the system performance improvement. Generally, the MRC may give a significant enhancement of system overall efficiency. The schematic

Z. Fang et al. / Energy 188 (2019) 116093

diagram and corresponding pressure-enthalpy diagram are shown in Fig. 1 (a) and Fig. 1 (b), respectively. As illustrated in Fig. 1, the MRC is operated in the following manner. The compressed vapor mixture R290/R600a is partially condensed in the condenser I and then split into two streams with different compositions by the separator (process 1-2-2v,2l). On one hand, the R290-enriched vapor refrigerant is completely condensed and deeply subcooled by passing through the condenser II, the subcooler and IHX, respectively (process 2v-3-4-5). Subsequently, this subcooled refrigerant is expanded by the EV II and then boiled in the freezer evaporator to absorb heat from freezer compartment (process 5-6-7), and this evaporated saturated vapor refrigerant would be superheated and compressed by passing through the IHX and compressor II, respectively (process 7-8-9). On the other hand, the R600a-enriched liquid refrigerant leaving the separator is expanded by the EV I and then partially evaporated by absorbing the heat from the refrigerator compartment (process 2l-10e11). This partially evaporated refrigerant is completely evaporated and superheated by passing through the subcooler for subcooling the refrigerant leaving condenser II (process 11e12). Then this superheated refrigerant goes to the compressor I (process 12e13). Finally, the compressed refrigerants at point 9 and point 13 are mixed to point 1 (process 9,13-1). In this way, the operating process of MRC is completed.

3. Mathematical model and simulation procedure 3.1. Assumptions To evaluate the thermodynamic performance of MRC in detail, a mathematical model based on mass and energy conservation is proposed. This theoretical research method has been used extensively for simulation studies of various refrigeration cycles in many literatures, thus, the presented simulation results could be used to indicate the performance characteristics of the cycle [19,20,23]. Some assumptions are made as follows for simplifying the simulation process [23].

3

(1) The compression process is irreversible and the corresponding isentropic efficiency related to the pressure ratios are taken into account [24]; (2) The expansion processes in expansion valves are isenthalpic; (3) Refrigerant pressure drop and heat losses in the cycle are neglected; (4) The refrigerant at refrigerator evaporator outlet is at twophase state; (5) The environmental reference state is at T0 ¼ 25+ C and p0 ¼ 101:325 kPa for exergy analysis; (6) The refrigerant mass flow rate at condenser inlet (point 1) is set as 1:5 g$s1 .

3.2. System energetic model Based on these above assumptions, the following equations for main components can be obtained. The refrigerant mass flow rates at the separator outlets is expressed as:

m_ 2v ¼ m_ 1 x2

(1)

m_ 2l ¼ m_ 1 ð1  x2 Þ

(2)

where m_ 2v and m_ 2l are the refrigerant mass flow rates at separator vapor outlet and liquid outlet, respectively; x2 is refrigerant quality at separator inlet. The cooling capacities and cooling capacity ratio are given as:

qer ¼ m_ 2l ðh11  h10 Þ

(3)

qef ¼ m_ 2v ðh7  h6 Þ

(4)

where qer and qef are cooling capacities of refrigerator evaporator and freezer evaporator, respectively; h10 , h11 , h6 and h7 are the specific enthalpies of the refrigerants at the inlet and outlet of the two evaporators, respectively. The cooling capacity ratio 4 is defined as the ratio between the

(a )

(b ) Fig. 1. (a) Schematic diagram of MRC; (b) Cycle pressure-enthalpy diagram of MRC.

4

Z. Fang et al. / Energy 188 (2019) 116093

cooling capacities of refrigerator evaporator and freezer evaporator, which can be expressed as follows. Noted that the variation of cooling capacity ratio is dependent on the heat loads of the two compartments, which are affected by the compartment structures, the types and quantities of storage goods as well as the environment conditions.

.

 i h ej ¼ ej T þ ej M ¼ hj  hpj :T0  T0 sj  spj :T0 p¼constant |fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl} ej T

 i h þ hpj :T0  h0  T0 spj :T0  s0 T0 ¼constant |fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl ffl} ej M

4 ¼ qer qef

(5)

(15)

The compression works of the two compressors can be calculated by

Actually, the thermal exergy ej T and mechanical exergy ej M are the obtainable parts due to the refrigerant temperature and pressure differences, respectively. Based the above definition, the exergy fuel EF:k and exergy production EP:k of each component can be calculated by the equations listed in Table 1 [27e29]. The exergy destruction ED:k of each component can be derived as:

w1 ¼ m_ 2l ðh13  h12 Þ ¼ m_ 2l ðh13:is  h12 Þ=his1

(6)

w2 ¼ m_ 2v ðh9  h8 Þ ¼ m_ 2v ðh9:is  h8 Þ=his2

(7)

w ¼ w1 þ w2

(8)

where h12 and h8 are the specific enthalpies at the inlets of the two compressors, respectively; h13:is and h9:is are refrigerant specific enthalpies at the two compressor outlets under isentropic compression process, respectively; his1 and his2 are the isentropic efficiencies of compressor I and compressor II, respectively. The compressor isentropic efficiencies are given as [24]:

p his1 ¼ 0:874  0:0135 13 p12

his2 ¼ 0:874  0:0135

p9 p8

(9)

(10)

ED:k ¼ EF:k  EP:k

(16)

Specially, the exergy loss EL of the system is the exergy transferred to the environment, which cannot be utilized further in the system, i.e., EL ¼ EP:cond . The exergy balance for the overall system can be expressed as [28]:

EF:comp ¼ EP:evap:r þ EP:evap:f þ ED:tot þ EL

(17)

To comparing the irreversibility during energy conversion process of each component, the exergy destruction ratio of each component is defined as:

fk ¼

ED:k EF:comp

(18)

where p12 , p8 , p13 and p9 are refrigerant pressures at the inlet and outlet of the two compressors, respectively. The COP is defined as the ratio between the total cooling capacity and compression work, which can be derived as follow.

The system exergy efficiency is defined as the ratio between exergy production of the two evaporators and the total compression work, expressed as:

.  w COP ¼ qer þ qef

hex ¼

(11)

Neglecting the heat losses in the subcooler and IHX, the energy balances of the two heat exchangers are given as:

m_ 2v ðh3  h4 Þ ¼ m_ 2l ðh12  h11 Þ

(12)

h4  h5 ¼ h8  h7

(13)

3.3. System exergetic model Exergy is the maximum available work that could be obtained from an energy conversion system interacting with the environment. The exergy-based analysis method is always used to indicate the location of irreversibility generation and evaluate the thermodynamic imperfection quantitatively [25]. Therefore, the exergetic model is established to investigate the energy conversion behaviors of MRC. In a refrigeration system, the exergy of the refrigerant at a state point can be expressed as follows [26].

   Ej ¼ m_ j ej ¼ m_ j hj  h0  T0 sj  s0

(14)

where the ej can be split into thermal exergy ej T and mechanical exergy ej M, expressed as:

EP:evap:r þ EP:evap:r EF:comp

(19)

Using these above equations, the energy and exergy performance investigations of MRC can be carried out. The calculation program is written in Fortran Language. Zeotropic mixture R290/ R600a is used in this system and the corresponding refrigerant thermodynamic properties are calculated by using NIST database and subroutines [30]. The calculation procedure for the solution of MRC based on the above modeling equations is shown in Fig. 2. In this calculation procedure, dichotomy method is used to adjusted the refrigerant qualities and temperatures, and this calculation method has a fast convergence speed. Moreover, it should be noted that the refrigerant mass fractions of R290 at separator outlets are dependent on the refrigerant quality and pressure at separator inlet. In this study, the refrigerant quality at separator inlet is adjusted according to the cooling capacity ratio variations. Thus, from the viewpoint of system control aspect, it can be said that the refrigerant mass fractions of R290 at separator outlets are determined by the cooling capacity ratio variations. Based on practical application conditions, the temperatures in refrigerator and freezer chambers are always set as invariable. Therefore, the system efficiency is mainly dependent on the condensation temperature (temperature at condenser II outlet t3 ). In this paper, the condensation temperature ranging from 25
C to 55
C is selected as the input condition, which almost covers the whole application conditions of a refrigerator-freezer. The cooling capacity ratio between refrigerator evaporator and freezer evaporator is also selected as a critical input parameter, which is used to

Z. Fang et al. / Energy 188 (2019) 116093

5

Table 1 Definition of fuel and production exergy of the MRC cycle components. Component Comp Cond Evap.r Evap.f OHX

EV Mix Sepa

EF:k EF:comp ¼ w EF:cond ¼ m_ 1 ðe1  e2 Þ þ m_ 2v ðe2v  e3 Þ EF:evap:r ¼ m_ 10 ðe10  e11 Þ EF:evap:f ¼ m_ 6 ðe6  e7 Þ EF:subc ¼ MAX½jm_ 3 ðe3  e4 Þj; jm_ 11 ðe11  e12 Þj EF:IHX ¼ MAX½jm_ 4 ðe4  e5 Þj; jm_ 7 ðe7  e8 Þj EF:OHX ¼ EF:subc þ EF:IHX EF:EV ¼ m_ 10 ðe2l  e10 M Þ þ m_ 6 ðe5  e6 M Þ EF:mix ¼ m_ 13 e13 þ m_ 9 e9 EF:sepa ¼ m_ 2 e2

EP:k EP:comp ¼ m_ 12 ðe13  e12 Þ þ m_ 8 ðe9  e8 Þ EP:cond ¼ m_ a1 ðea2  ea1 Þ þ m_ a3 ðea4  ea3 Þ EP:evap:r ¼ m_ a5 ðea6  ea5 Þ EP:evap:f ¼ m_ a7 ðea8  ea7 Þ EP:subc ¼ MIN½jm_ 3 ðe3  e4 Þj; jm_ 11 ðe11  e12 Þj EP:IHX ¼ MIN½jm_ 4 ðe4  e5 Þj; jm_ 7 ðe7  e8 Þj EP:OHX ¼ EP:subc þ EP:IHX EP:EV ¼ m_ 10 e10 T þ m_ 6 e6 T EP:mix ¼ m_ 1 e1 EP:sepa ¼ m_ 2l e2l þ m_ 2v e2v

Table 2 The operation parameters in the simulation. Parameter

Value

Refrigerator evaporation temperature (t11 ) Freezer evaporation temperature (t7 ) Air inlet temperature of refrigerator evaporator (ta5 ) Air outlet temperature of refrigerator evaporator (ta6 ) Air inlet temperature of freezer evaporator (ta7 ) Air outlet temperature of freezer evaporator (ta8 ) Compressor I suction temperature (t12 )

5
C 28
C 5
C 1
C 23
C 28
C 10
C

Table 3 The performance comparison between MRC and CRC. Cycle

COP

hex

per ðkPaÞ

pef ðkPaÞ

g

MRC

2.53

37.3%

215

105

CRC Improvement

1.94 30.4%

28.6% 30.4%

92 123

92 13

4.5 (CP I) 9.3 (CP II) 9.8 d

operated at different evaporation pressures, which can decrease the irreversibility during heat transfer process of refrigerator evaporator and reduce the total compression work. As shown in Table 3, the refrigerator evaporation pressure of MRC is 123 kPa higher than that of CRC. Moreover, owing to the application of separation condensation technique, the freezer evaporation pressure of MRC is higher than that of CRC, which can reduce the compression ratio of compressor II in MRC. As mentioned above, the MRC can give a significant performance enhancement for refrigerator-freezer applications. Actually, the given conditions such as condensation temperature, refrigerant mass fraction of R290 and cooling capacity ratio can significantly affect the system performance, which are investigated detailedly as follows.

Fig. 2. The calculation procedure for the solution of MRC.

represent different cases of the heat load in two evaporators. Noted that the heat transfer temperature difference at the subcooler cold end is set as 5 K. Additionally, the other operation parameters in the simulation are listed in Table 2. 4. Results and discussion The performance simulation results of both cycles are listed in Table 3, which are calculated under the basic conditions where z ¼ 0:6, tc ¼ 35+ C, ter ¼  5+ C, tef ¼  28+ C, 4 ¼ 0:8. Table 4 displays the state of each point in MRC under the above conditions. It is indicated that MRC yields about 30.4% higher COP and exergy efficiency over CRC. The MRC allows the two evaporators to be

4.1. The effect of condensation temperature on system performance The effect of tc on COP and w variations of both cycles is shown in Fig. 3. It is observed that as tc increases, due to the increasing compression ratios, the COPs of both cycles decrease whereas the compression works show increasing trends, and the COP of MRC is always superior to that of CRC. Compared with CRC, the corresponding COP improvement of MRC varies from 26% to 35% in the tc range of 25e55
C. It is seen that as tc rises, the corresponding COP improvement of MRC over CRC is increased. The main reason is that when tc rises, the heat transfer capacity of subcooler is increased since the temperature at point 11 is set as 5
C, in another word, the subcooling degree at point 4 increases. Hence, the refrigerant quality at point 2 should be reduced to match the increasing subcooling degree at point 4, which leads to the increasing refrigerant mass flow rate of compressor I. In this way, due to the lower

6

Z. Fang et al. / Energy 188 (2019) 116093

Table 4 The state of each point in MRC system. point

tð+ CÞ

pðkPaÞ

hðkJ$kg1 Þ

sðkJ$kg1 $K1 Þ

x

zR290

1 2 2l 2v 3 4 5 6 7 8 9 10 11 12 13

77.8 41.0 41.0 41.0 35.0 0.0 14.9 34.9 28.0 5.0 88.9 9.6 5.0 10.0 69.6

976 976 976 976 976 976 976 105 105 105 976 215 215 215 976

689.8 436.1 303.4 618.0 288.9 200.2 165.1 165.1 535.2 570.3 716.9 303.4 519.7 584.4 669.9

2.68 1.89 1.85 1.94 1.41 1.10 0.97 0.98 2.52 2.65 2.76 1.51 2.32 2.56 2.61

superheated 0.422 0 1 0 subcooled subcooled 0.117 1 superheated superheated 0.346 0.890 superheated superheated

0.55 0.55 0.48 0.65 0.65 0.65 0.65 0.65 0.65 0.65 0.65 0.48 0.48 0.48 0.48

2.6

270

2.5

240 210

2.0

1.0

ter z 25

30

5 , tef 0.55, 35

25

40

45

50

55

230

2.3

ter

2.2

tc

35 z

25

220

0.55

180 170

1.9

120

1.8

tc ( ) Fig. 3. The COP and compression work variations of both cycles with condensation temperature.

5 , tef

2.0

150

0.8

240

2.1

180 1.5

MRC CRC

250

w

2.4

w(W)

COP

2.5

300

COP

w(W)

3.0

2.7

160 0.2

0.4

0.6

0.8

1.0

1.2

1.4

Fig. 4. The COP and compression work variations of both cycles with the cooling capacity ratio.

compression ratio and specific work of compressor I, the compression work increasing rate of MRC is reduced, which can be observed in Fig. 3. Therefore, the corresponding COP improvement of MRC over CRC shows an increasing trend when increasing tc . As mentioned above, the MRC outperforms the CRC and it appears to be very promising for refrigerator-freezer applications.

1.8 m2l

1.6

m2v

350

qer qef

300 250

1.4

200

4.2. The effect of cooling capacity ratio on system performance The cooling capacity ratio is another critical parameter affecting the system performance of MRC. Fig. 4 displays the COP and w variations of both cycles with 4 under the above conditions. As shown in Fig. 4, the COP of MRC rises and the compression work w decreases with an increasing 4. The main reason is that the evaporation pressure of refrigerator evaporator is higher than that of freezer evaporator in MRC, which means that the compression ratio of compressor I is lower than that of compressor II, thus, the specific compression work of compressor I is much lower. Actually, when 4 increases, the refrigerant mass flow rate of compressor I increases whereas that of compressor II is in the opposite trend, which can be observed in Fig. 5. Therefore, the total compression work of MRC would decrease with an increasing 4, which results in the increasing COP. Fig. 5 also displays the cooling capacity variations for the two evaporators with 4. It is seen that the cooling capacity is in direct proportion to the corresponding refrigerant mass flow

m(g/s)

1.2

150

1.0

100

0.8

50

0.6

ter

0.4

tc

0.2

0.4

5 , tef 35 z

0.6

0

25

-50

0.55

0.8

qe(W)

MRC CRC

330

w

COP

COP

3.5

1.0

1.2

1.4

-100

Fig. 5. The refrigerant mass flow rate and cooling capacity variations for the two evaporators with the cooling capacity ratio.

rate, which can be illustrated by Eqs. (3)e(5). Generally, increasing the cooling capacity ratio is beneficial to the performance improvement of MRC. However, for the CRC system, due to the

Z. Fang et al. / Energy 188 (2019) 116093

single evaporation pressure, the COP and w are invariable when 4 changes. In comparison to CRC, the COP of MRC is improvement by 17%e35% in the 4 range of 0.2e1.4.

7

8

MRC CRC

7 4.3. The effect of R290 mass fraction on system performance

MRC CRC

2.7

COP

2.6 2.5 ter

2.1

tc

2.0

5 , tef 35 

25 0.8

1.9 0.2

0.3

0.4

0.5

z

0.6

0.7

Fig. 6. The COP variations of both cycles with R290 mass fraction.

tef

0.8

t (K)

6

Zeotropic mixture R290/R600a is used and the R290 mass fraction z significant affects the system performance of both cycles. Fig. 6 displays the COP variations of both cycles with z under the above conditions. Noted that when comparing the performance between the two cycles, zeotropic mixture R290/R600a is also used in the CRC, and the refrigerant mass fractions for both cycles are set at the same value. It is observed that the COPs of both cycles first decrease and then increase when z rises. Actually, for both cycles, as R290 mass fraction z increases, the condensation and evaporation pressures would increase since R290 is the low boiling-point refrigerant in this mixture. It is known that increasing evaporation pressure is beneficial to the system performance improvement, whereas the condensation pressure increment would lead to negative effect. As shown in Fig. 6, when z is less than 0.45, the effect of condensation pressure increment on COP is dominant, however, once z is larger than 0.45, the evaporation pressure increment have a much higher effect on COP. Therefore, the COPs of both cycles first decrease and then increase. As z ranges from 0.2 to 0.8, the MRC yields 28%e35% higher COP over CRC. Additionally, using the zeotropic mixture can ensure a better matching characteristic between the temperature variations of refrigerant and air for the evaporator due to its temperature glide feature. Hence, the evaporator temperature glide is a critical parameter for evaluating the system performance. Fig. 7 shows the evaporator temperature glide variations with z for both cycles. It is indicated that the evaporator temperature glides of both cycles first increase and then decrease, which is contrary to the COP variation trend. As shown in Fig. 7, the temperature glides of the two evaporators in MRC are much higher than those of CRC under the basic conditions, especially of freezer evaporator. There are two main reasons for this phenomenon. On one hand, the MRC contains two independent evaporators, and each evaporator contains an entire phase change process, however, the evaporation process in CRC are divided into two parts for refrigerator and freezer evaporators. Hence, the temperature glides of the two evaporators in MRC is higher than those of CRC. On the other hand, the MRC applies a

2.8

ter

5 4 3 2

ter

1

tc

0.2

0.3

0.4

5 , tef 35 

0.5

z

0.6

25 0.8

0.7

0.8

Fig. 7. The effect of R290 mass fraction on the evaporator temperature glides for both cycles.

subcooler to increase the refrigerant subcooling degree at EV II inlet (point 5), which results in a higher temperature glide for the freezer evaporator in MRC. Consequently, compared with CRC, the MRC can ensure a much better matching characteristic between temperature variations of the two fluid in evaporator. According to these above investigations, as z increases, the evaporator temperature glide variation trend are contrary to that of COP, which means that z should be properly selected according to the trade-off between COP and evaporator temperature glides. In this study, the basic condition where z equals 0.55 is selected.

4.4. Exergetic analysis of MRC and CRC Fig. 8 displays the component exergy destruction comparison between the two cycles. It is observed that in comparison with CRC, the exergy destructions of the compressor, condenser and refrigerator evaporator within MRC are much lower, especially of refrigerator evaporator. As mentioned earlier, the MRC applies two independent compressors, and the compressor I is operated at low compression ratio conditions, thus, the isentropic efficiency of compressor I is higher, which can decrease the compressor exergy destruction of MRC. Besides, due to the application of parallel compression, the refrigerant superheat degree at condenser inlet decreases, which results in the lower heat transfer capacity of the condensers in MRC. Hence, the condenser exergy destruction of MRC is lower than that of CRC. Additionally, in MRC system, the heat transfer temperature difference of refrigerator evaporator is lower, which is beneficial for reducing the heat transfer irreversible loss. However, in CRC system, there is an identical evaporation pressure of the two evaporators, and this leads to the higher exergy destruction of refrigerator evaporator. In this way, the refrigerator evaporator exergy destruction of MRC is much lower than that of CRC. Furthermore, the MRC uses two-stage heat exchangers (subcooler and IHX) to reduce the average heat transfer temperature difference, thus, the ED:OHX of MRC is lower. At last, due to the higher refrigerator evaporation pressure of MRC, the irreversible loss during expansion process is decreased. As investigated above, compared with CRC, the exergy destruction of each component in MRC is lower, which means that the MRC can reduce the irreversible loss of each component and improve the system overall

8

Z. Fang et al. / Energy 188 (2019) 116093

MRC CRC

50

ter

ED (W)

40

tc

5 , tef 35 , z

25 0.55,

0.8

30

20

10

0

EL

ED.comp ED.cond ED.evap.r ED.evap. f ED.OHX ED. EV

Fig. 8. The component exergy destruction comparison between the two cycles.

performance. It is indicated that the exergy destruction reduction of compressor and refrigerator evaporator is the main reason for the higher COP of MRC. The component exergy destruction ratios of both cycles are listed in Table 5 under the basic conditions. It is observed that the refrigerator evaporator exergy destruction ratio of CRC is 12.6%, whereas that of MRC is only 4.9%, which means that applying MRC can significantly reduce the irreversible loss during heat transfer process within refrigerator evaporator. As shown in Table 5, the compressor contributes the largest exergy destruction ratio in both cycles due to the high irreversibility during compression process. For MRC, the second largest exergy destruction occurs in condenser with fComp ¼ 12:5%, followed by the expansion valves, which account for 9.2% of the system total compression work. Moreover, the exergy destruction ratios of the refrigerator and freezer evaporators are 4.9% and 4.4%, respectively. From the viewpoint of exergy analysis, the component with highest exergy destruction should be improved first. Hence, the high priority in need of improvement for the MRC is ordered as the compressors, condensers, expansion valves, two evaporators, subcooler and IHX. The effect of cooling capacity ratio 4 on component exergy destructions of MRC is shown in Fig. 9. It is found that as 4 increases, the ED:evap:r and ED:EV show increasing trends. The main reason is that when 4 rises, the heat transfer capacity and refrigerant mass flow rate of refrigerator evaporator increase, which results in the increasing ED:evap:r . Besides, MRC applies a subcooler and an IHX in series to increase the refrigerant subcooling degree at EV II inlet (point 5), which can reduce the irreversible loss during expansion

process of EV II. Therefore, though the expansion ratio of EV I is lower, the specific exergy destruction of EV I is higher than that of EV II, and the refrigerant mass flow rate of EV I rises whereas that of EV II shows the opposite trend when 4 rises, consequently, the ED:EV would increase. As shown in Fig. 9, the cooling capacity ratio has little effect on the condenser exergy destruction. It is also seen that the exergy destructions of compressor, freezer evaporator, subcooler and IHX decrease with an increasing 4. As mentioned earlier, the specific exergy destruction of compressor I is lower than that of compressor II due to its lower compression ratio. Therefore, increasing refrigerant mass flow rate of compressor I is beneficial to the total compression exergy destruction decrement. Furthermore, as 4 rises, the heat transfer capacities of subcooler, IHX and freezer evaporator are decreased, which results in the decreasing ED:evap:f and ED:OHX . Generally, increasing 4 can significantly reduce the total exergy destruction of MRC and improve the system overall performance. 4.5. Economic evaluation of the MRC The economic advantages of MRC are investigated on the basis of a selected refrigerator-freezer prototype with the model of BCD750 W. The specifications of this refrigerator-freezer are listed in Table 6. Actually, on the basis of CRC, the MRC extra applies a compressor, a subcooler, a phase separator and some additional tubes. In a refrigerator-freezer, the subcooler is made of capillary winding. Detailed cost information of CRC and MRC is displayed in

Table 5 The component exergy destruction ratios of both cycles. Component Cycle

CRC MRC

Comp

Cond

Evap.r

Evap.f

OHX

EV

20.9% 19.1%

13.1% 12.5%

12.6% 4.9%

3.9% 4.4%

4.7% 3.0%

9.0% 9.2%

Z. Fang et al. / Energy 188 (2019) 116093

40

ED.comp ED.cond ED .OHX

35

ED (W)

30 25

9

ter tc

5 , tef 35 , z

ED . EV ED.evap.r ED.evap. f

25 0.55

20 15 10 5 0

0.2

0.4

0.6

0.8

1.0

1.2

1.4

Fig. 9. The component exergy destruction variations of MRC with the cooling capacity ratio.

Table 6 Specifications of the selected refrigerator-freezer prototype. Components (Model: BCD-750 W)

Specifications

Machine dimensions Refrigerator compartment Freezer compartment Compressor Capillary tube

1912  850  908 mm (Height  Depth  Width) Volume: 450 L (920  670  730 mm) Volume: 300 L (660  640  710 mm) Piston compressor (Displacement: 11.0 cm3 ) 1.8  0.7  3200 mm (out-tube diameter  in-tube diameter  length)

Table 7. In China, the average price of resident electricity is about 0.5 RMB=kW,h. The simulation results show that the MRC can reduce about 30.4% power consumption in comparison with CRC. According to Table 7, the payback period of the investment is calculated as follow [31].

y ¼ ð417:3  329:3Þ÷½ð1:06  0:739Þ  0:5 ¼ 549 day ¼ 1:5 year The payback period of the investment for MRC is 1.5 year, and the operating lifetime of a refrigerator-freezer is larger than 10 years. Thus, from the viewpoint of economic evaluation, the MRC is much suitable for refrigerator-freezer applications.

5. Conclusions A modified refrigeration cycle (MRC) with zeotropic mixture R290/R600a is proposed for refrigerator-freezer applications. In contrast to the conventional refrigeration cycle (CRC), the MRC applies two parallel compressors to enable the two evaporators to be operated at different evaporation temperatures properly matched with the requirements of refrigerator and freezer compartments, respectively. A subcooler is installed between the two loops for further improving the system performance. Moreover, a phase separator is set between the two condensers, which can ensure more low-boiling point refrigerant to go to the freezer evaporator for increasing the freezer evaporation pressure. On the basis of a thermodynamic model, the energy and exergy performance of MRC is investigated and then compared with the CRC

Table 7 The cost comparison between CRC and MRC. Cycles

Compressor (cost/model)

Phase separator

Capillary tube

Additional tube

Total cost

Power consumption (kW,h=24h)

CRC MRC

315 RMB (DONPER VFL110CY1) 150 RMB (DONPER VFL070CY1) Compressor I 230 RMB (DONPER VFA090CY1) Compressor II

___ 13.5 RMB

4.8 RMB (3.2 m) 8.3 RMB (5.5 m)

9.5 RMB (2.6 m) 15.5 RMB (4.2 m)

329.3 RMB 417.3 RMB

1.06 0.739

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Z. Fang et al. / Energy 188 (2019) 116093

under different given conditions. The primary findings can be summarized as follows: (1) The simulation results indicate that compared with CRC system, the MRC yields 26%e35% higher COP and exergy efficiency in the condensation temperature range of 25e55
C. Besides, the refrigerant temperature glides of the two evaporators in MRC are much higher than those of CRC, especially of the freezer evaporator. This means that in comparison with CRC, the MRC can ensure better matching characteristics between the temperature variations of refrigerant and air for the two evaporators. (2) Increasing condensation temperature or cooling capacity ratio 4 is beneficial to the performance improvement of MRC over CRC. Hence, the MRC becomes more suitable for the applications under high environment temperature or large refrigerator cooling capacity requirement conditions. (3) From the viewpoint of exergy analysis, the component with highest exergy destruction should be improved first. Hence, the high priority in need of improvement for the MRC is ordered as the compressors, condensers, expansion valves, two evaporators, subcooler and IHX. In general, this preliminary investigation indicates that the MRC could obtain a significant performance enhancement for refrigeratorfreezer applications, and the relevant experimental work will be conducted in the future.

[8] [9]

[10]

[11]

[12]

[13]

[14]

[15]

[16] [17]

[18] [19] [20]

Acknowledgements This study is financially supported by the National Natural Science Foundation of China (NSFC) under the grant No. 51576149. The authors would like to thank NSFC for the sponsorship.

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