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Energetic and exergetic investigation of a novel refrigeration system utilizing ejector integrated subcooling using different refrigerants
Energy 168 (2019) 712e727
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Energetic and exergetic investigation of a novel refrigeration system utilizing ejector integrated subcooling using different refrigerants Tuncay Yilmaz*, Mehmet Tahir Erdinç Department of Mechanical Engineering, Osmaniye Korkut Ata University, 80000 Osmaniye, Turkey
a r t i c l e i n f o
a b s t r a c t
Article history: Received 29 May 2018 Received in revised form 19 November 2018 Accepted 20 November 2018 Available online 24 November 2018
Increase in coefficient of performance of refrigeration systems is eminently important for various applications such as household refrigerator, tumble dryer, and more in practical applications. A new ejector subcooling system is described and investigated numerically without using an extra pump or compressor. After condenser, some part of the liquid (by-pass refrigerant) is expanded so that its temperature could be lower than the intended subcooling temperature. By using this by-pass refrigerant, the main flow of the refrigerant is subcooled. After evaporation, the by-pass refrigerant is expanded using an ejector. Seven different refrigerants are considered as working fluids (namely; R32, R1234yf, R290, R134a, R717, R600a and R245fa). R1234yf, R290, R600a and R717 are considered as low global warming potential refrigerants in European Union regulation. For each refrigerant, the values of coefficient of performance, relative increment in coefficient of performance and exergy efficiencies are calculated and demonstrated graphically for different condenser temperatures between 30 C=70 C and evaporator temperatures between 20 C=10 C. Results show that the best performance is obtained for R1234yf with an increment of about 20% in coefficient of performance and 18% in exergy efficiency. These efficiencies can be higher or lower approximately by 3% using ejector with higher or lower component efficiencies. © 2018 Elsevier Ltd. All rights reserved.
Keywords: Refrigeration Ejector Subcooling Energetic Exergetic
1. Introduction Refrigeration is perhaps the most important application for reducing the environmental impact of global energy use [1]. Therefore, European Union (EU) has issued regulations on the energy efficiency levels [2]. The importance of energy efficiency of household refrigerator is explained by Kosalay et al. [3]. They reported that these household refrigerators electricity bill constitutes over 30% of total domestic electricity bill in some countries. One of the most important method to increase the efficiency of the refrigeration systems is subcooling. Qureshi and Zubair [4] broadly described different possibilities for subcooling, such as ambient subcooling, subcooling with liquid-suction heat exchanger, subcooling with external heat exchanger and dedicated-integrated mechanical subcooling. As mentioned above, different methods are applicable for subcooling. In the first one, the liquid refrigerant can be cooled down after the condenser with the cold refrigerant which leaves the
* Corresponding author. E-mail address: [email protected] (T. Yilmaz). https://doi.org/10.1016/j.energy.2018.11.081 0360-5442/© 2018 Elsevier Ltd. All rights reserved.
evaporator. However, this method has the disadvantage of increasing the refrigerant temperature at the inlet of the compressor which causes some decrement in mass flow rate flowing through the compressor and hence decreases the capacity of refrigeration. Subcooling with external heat exchanger needs heat exchanger utilizing cooling water. However, this again needs cooling tower or similar devices with a circulating water pump which increase significantly the investment and maintenance costs. Another method is taking advantage of using a dedicated and integrated mechanical subcooling system in the cycle. A second refrigeration cycle is adopted for cooling the liquid after condensing process. However, this additional refrigeration cycle has the main drawback of increasing the investment cost of the system. Zubair [5] found that energy savings in between 20% and 40% can be obtained using mechanical subcooling. Khan and Zubair [6] found that the optimum subcooling temperature should be halfway between the condensation and evaporation temperatures. Dai et al. [7] studied mechanical subcooling refrigeration system using zeotropic mixtures. They found that optimum coefficient of performance (COP) is achieved at the optimum discharge pressure and the optimum subcooling degree. Roy and Mandal [8] investigated
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Nomenclature COP1 COP2 COP21 e ed h _ M _ M P Q_
q s T u w _ W x
coefficient of performance of the conventional system [] coefficient of performance of the ejector subcooling system [] coefficient of performance ratio [] specific exergy [kJ kg1] specific exergy destruction [kJ kg1] enthalpy [kJ kg1] mass flow rate [kg s1] mass flow rate ratio [] pressure [Pa] heat flow [kW] specific heat [kJ kg1] entropy [kJ kg1K1] temperature [ C ] velocity [ m s1] specific work [kJ kg1] power [kW] quality []
Greek Letters h efficiency [] 41 exergy efficiency of conventional system [] 42 exergy efficiency of ejector subcooling system [] 421 exergy efficiency ratio []
c comp1 comp2 cond1 cond2 d e1 e2 ev1 ev21 ev22 f he in m n out sc 0
713
compressor conventional system compressor ejector system compressor conventional system condenser ejector system condenser diffuser, destroyed, destruction conventional system evaporator ejector system evaporator conventional system expansion valve ejector system expansion valve 1 ejector system expansion valve 2 flashing point subcooling heat exchanger inlet mixing region nozzle outlet subcooling reference
Abbreviations COP coefficient of performance EU European Union ev expansion valve GWP global warming potential he subcooling heat exchanger HFC hydrofluorocarbons ODP ozone depleting potential
Subscripts BP by-pass
COP and exergetic efficiency of mechanical subcooling system and found increase of 16% in COP and decrease of 8.8% in exergy destruction rate at 10 C evaporator temperature for R134a. Yu et al. [9] described a refrigeration system with mechanical subcooling which uses an auxiliary liquidegas ejector and a liquid pump to enhance subcooling for the refrigerant leaving the condenser. This study proposes that ejector integrated system has advantages over other subcooling methods, therefore ejector technology has also been reviewed. There are many studies for ejector technology in literature. Because refrigeration systems which use solar heat at low temperatures are very interesting and have been investigated for many times. On this matter, extensive information was supplied by Pridasawas [10] and Kim and Infante Ferreira [11]. Abdullateef et al. [12] reported on the special topics of solar driven ejector refrigeration systems. New developments in ejector refrigeration systems (ERS) were depicted and explained by Chen et al. [13]. They explained especially geometric optimization of ejectors. Progress on mathematical modeling of ejectors were described and analyzed by He et al. [14]. They stated that finite difference method is recognized as the most universal technique for numerical solutions. Some researchers have focused on reduction of throttling losses in expansion valve by using ejector as an expansion valve in order to reduce the compressor work and increase COP of the system. Ersoy and Sag [15] investigated the utilization of ejector as expansion valve using R134a and found a COP enhancement in the range of 6e14%. Lin et al. [16,17] studied experimentally and numerically optimum geometric parameters of ejector. Sarkar [18]
analyzed and optimized four novel layouts of ejector driven multicompression three-evaporator refrigeration system using R32 as refrigerant. Ünal and Yılmaz [19] and Ünal et al. [20] studied two phase ejector and two evaporators using R134a as refrigerant in a bus refrigeration system. They found 15% increase in COP using the two-phase ejector in bus air-conditioning system compared with the conventional system. Besides this, reducing total weight of the bus, fuel saving can be achieved. Bai et al. [21] studied dualevaporator CO2 transcritical refrigeration cycle with two-stage ejector. They reported higher COP and exergy efficiency. In order to see the effect of the dimensions on the ejector system performance, numerical studies have been performed [22e28]. Comprehensive review on ejector refrigeration is given by Besagni et al. [29] and Chunnanond and Aphornratana [30]. Besagni et al. [31] carried out research work on working fluids for heat driven ejector refrigeration and found that each refrigerant has its own range of operating conditions. Gil and Kasperski [32] studied hydrofluoroolefins as working fluids in ejector refrigeration. They stated that the most benefit comes from the use of R-1234yf and R1234ze(E). The influence of generator temperature, working fluid and ejector component efficiency are presented by Besagni et al. [33]. They showed that ejector component efficiencies have great influence on COP. As stated above, the increase of efficiency and capacity of the conventional system is not very effective using subcooling of the liquid refrigerant at the exit of the condenser with the cold vapor from the exit of the evaporator. The subcooling with external heat exchanger needs in addition to the subcooling heat exchanger cooling tower and circulation pump. Dedicated mechanical
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subcooling needs besides subcooling heat exchanger, condenser and an other compressor. Similarly, for integrated mechanical subcooling, subcooling heat exchanger, receiver and an other compressor are necessary. Therefore, subcooling with external heat exchanger and dedicated and integrated mechanical subcooling systems increase significantly the capital and maintenance cost of the refrigeration system. In this work, a new ejector subcooling system [34] is described and numerically investigated without an extra pump or compressor. Besides subcooling heat exchanger, only an ejector and a valve are needed. Therefore, the most advantages of this new refrigeration system are its simplicity, its low capital and maintenance cost. 2. System description Thermodynamic analysis of the ejector subcooling refrigeration system is carried out by considering the following assumptions [19]:
Fig. 2. P-h diagram of conventional refrigeration system (system 1).
a) b) c) d)
Pressure losses of the system are neglected. Throttling process in expansion valve is isenthalpic. Isentropic efficiency of the compressor is prescribed. Isentropic efficiencies of the nozzle and the diffusor are prescribed. e) Efficiency of mixing section of the ejector is prescribed. f) The process in the mixing section takes place at constant pressure and constant cross-sectional area.
A conventional refrigeration system and its P-h diagram are presented in Figs. 1 and 2, respectively. For the sake of simplicity, compressor inlet is assumed as saturated vapor and condenser outlet is assumed as saturated liquid. The novel ejector subcooling refrigeration system [34] is shown schematically in Fig. 3 and its P-h diagram is presented in Fig. 4. After condensation process, some part (by-pass refrigerant) of the liquid refrigerant mass flow rate is expanded until it decreases to temperature T8, which is just below the subcooling temperature T5. The remaining liquid refrigerant is subcooled from temperature T4 till temperature T5 utilizing the enthalpy of evaporation of the refrigerant from point 4e to point 8. Afterwards, refrigerant pressure P7 is increased to the compressor inlet pressure P1 using the
Fig. 3. New subcooling refrigeration system (system 2).
ejector. T-Q_ diagram for the heat exchanger is presented in Fig. 5. Q_ is heat transferred in subcooling heat exchanger. he
3. Thermodynamic calculations 3.1. Conventional refrigeration cycle
Fig. 1. Conventional refrigeration system (system 1).
Conventional refrigeration cycle is given by points 7-21-3-4-4a7 in Fig. 2. In this cycle the evaporator and the condenser temperatures are Te and Tc. T4a ¼ T7 ¼ Te, T3 ¼ T4 ¼ Tc are prescribed and therefore condenser and evaporator pressures (Pc and Pe) that are assumed constant are given as Pc ¼ P21¼P21s ¼ P2¼P2s ¼ P3¼P4 and Pe ¼ P6¼P4a ¼ P9¼P9s ¼ P10¼P7. In the conventional cycle,
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_ is work supplied where Q_ e is heat supplied to the evaporator and W to the compressor. For system 1 one can write:
COP1 ¼
qe1 w1
(7)
because refrigerant mass flow rates in evaporator and condenser are equal to each other. Specific heat qe1 and specific work w1 can be determined from the enthalpies as follows:
qe1 ¼ h7 h4a
(8)
w1 ¼ h21 h7
(9)
where h21 is calculated from the definition of isentropic compressor efficiency: Fig. 4. New subcooling refrigeration system (system 2) in P-h diagram.
h21 ¼ h7 þ
h21s h7
hc
(10)
Using eqs.(7)-(10), COP of the conventional cycle COP1 (system 1) is determined as follows:
COP1 ¼
h7 h4a h21 h7
(11)
3.2. Ejector integrated subcooling system Flashing occurs by the expansion from point 5 to point 6 as shown in Fig. 4. Flashing onset temperature is T5f. One can write approximately the equation below for T5f
T5f ¼ T5
(12)
because liquid refrigerant specific heat can be assumed independent of temperature. Temperature T4e must be below temperature T5, and it can be calculated as follows: Fig. 5. T-Q_ diagram for subcooling heat exchanger (he).
T8 ¼ T4e ¼ T5 DT58
(13)
DT58 is chosen between 2 5 C in heat exchangers. In this work D
superheating and subcooling are assumed to be zero. We can find enthalpies at points 7, 4 and 4a assuming saturated vapor at point 7, saturated liquid at point 4 and isenthalpic expansion from point 4 to point 4a:
T58 ¼ 3 C is assumed. Pressure P8 and temperature T8 for the subcooling are determined using the below expressions because all refrigerants are pure fluids:
h7 ¼ f ðP7 ; x7 ¼ 1Þ
(1)
T8 ¼ T4e
(14)
h4 ¼ f ðP4 ; x4 ¼ 0Þ
(2)
P8 ¼ P4e
(15)
h4a ¼ h4
(3)
Assuming saturated vapor at point 8, P8 is calculated by: Point 21s is compressor exit undergoing isentropic compression. Therefore:
P8 ¼ f ðT8 ; x8 ¼ 1Þ
(16)
The subcooling temperature is given by;
s21s ¼ s7
(4)
T5 ¼ T4 DTsc $ðT4 T7 Þ
h21s ¼ f ðP21s ; s21s Þ
(5)
where DTsc is dimensionless temperature which is defined as follows:
The COP of the refrigeration cycles is defined as follows:
Q_ COP ¼ e _ W
DTsc ¼ (6)
T5 T4 ðTc Te Þ
(17)
(18)
DTsc is a dimensionless parameter which changes between 0 and 1. DTsc is used as a parameter in the calculations. Because the
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temperature after subcooling T5 is known and the pressure at point 5 is equal to condenser pressure P4, one can write:
h5 ¼ fðT5 ; P4 Þ
(19)
The expansion process from point 5 to 6 is isenthalpic, therefore:
h6 ¼ h5
(20)
At the exit of the evaporator, refrigerant vapor is assumed as saturated. Then, enthalpy of the point 7 can be determined as:
h7 ¼ fðT7 ; x7 ¼ 1Þ
(21)
The vapor at the exit of the heat exchanger (point 8) is assumed to be saturated. Therefore, the following equation for the enthalpy at point 8 can be written:
h8 ¼ fðT8 ; x8 ¼ 1Þ
(23)
8
_ and M _ 5 are mass flow rate at ejector inlet at point 8 and where M 8 _ ¼ ∞ means no primary mass flow rate at point 5, respectively. M
_ the heat balance at the subsubcooling. For the calculation of M cooling heat exchanger (he) is used:
_ $ðh h Þ _ $ðh h Þ ¼ M M 5 5 8 8 4e 4
0:5 hm u10 ¼ u9 $ _ 1þM
(29)
u10 is the velocity at point 10. The velocities u at the points 7, 8 and evaporator exit are neglected. Besides these, the following equations are valid:
s9s ¼ s8
(30)
s8 ¼ fðP8 ; x8 ¼ 1Þ
(31)
h9s ¼ f ðP ¼ P9s ; s ¼ s9s Þ
(32)
P9s ¼ P7
(33)
Neglecting velocity u1, the energy equation between the points 10 and 1 can be written as follows:
. h1 ¼ h10 þ u210 2
(34)
Besides this, from the definition of the isentropic efficiency of the diffuser hd, the enthalpy at point 1 can be determined:
h1 ¼ h10 þ
_ is determined from the eqs. (23) and (24): M
ðh1s h10 Þ
hd
(35)
For the entropies s10 and s1s the following relations are valid:
(25)
Compressor inlet point 1 is determined using ejector (Fig. 6) equations which are written according to Ünal and Yılmaz [19]. From the definition of nozzle isentropic efficiency hn, the equation below can be written for the enthalpy at point 9:
h9 ¼ h8 hn $ðh8 h9s Þ
(28)
From the definition of mixing efficiency hm, the following equation is obtained:
(24)
_ ¼ h8 h4e M h4 h5
!
(22)
Mass flow rate ratio is defined as;
_ _ ¼ M5 M _ M
2 u2 _ $h7 ¼ 1 þ M _ $ h þ u10 h9 þ 9 þ M 10 2 2
(26)
s10 ¼ f ðh10 ; P7 Þ
(36)
s1s ¼ s10
(37)
The enthalpy at point 1s, temperature and entropy at point 1 and entropy at point 2s are determined from the relations below:
h1s ¼ f ðs1s ; P1 Þ
(38)
Neglecting the velocity at point 8, the energy equation between the points 8 and 9 yields:
T1 ¼ fðh1 ; P1 Þ
(39)
u9 ¼ ð2$ðh8 h9 ÞÞ0:5
s1 ¼ fðh1 ; P1 Þ
(40)
s2s ¼ s1
(41)
(27)
u9 is the velocity at point 9. Energy equation between the nozzle exit and the mixing region is given below:
The enthalpy at points 2 is calculated from the isentropic
Fig. 6. Detailed presentation of ejector.
T. Yilmaz, M.T. Erdinç / Energy 168 (2019) 712e727
calculate exergies at all points and exergy destructions at all components.
efficiency definition of the compressor:
h2 ¼ h1 þ
h2s h1
hc
(42)
with
h2s ¼ fðs2s ; P3 Þ
(43)
The isentropic efficiency of the compressor hc is assumed to be equal for both compressor of system 1 and system 2. The coefficient of performance COP2 for the new ejector subcooling system (system 2) can then be determined using eq. (6) as:
COP2 ¼
Q_ e2 _ W
717
(44)
3.3.1. Determination of exergy efficiency for conventional system (system 1) Specific exergies at the points 7, 21, 4 and 4a in Figs. 1 and 2 can be written according to eq. (49) as follows:
e7 ¼ h7 h0 T0 $ðs7 s0 Þ
(51)
e21 ¼ h21 h0 T0 $ðs21 s0 Þ
(52)
e4 ¼ h4 h0 T0 $ðs4 s0 Þ
(53)
e4a ¼ h4a h0 T0 $ðs4a s0 Þ
(54)
2
where
Q_ e2 ¼ M_ 5 $ðh7 h6 Þ
(45)
_ ¼ M_ 5 þ M_ $ðh h Þ W 2 8 2 1
(46)
ed; comp1 ¼ e7 e21 þ wcomp1
From eqs. (23) and (44)e(46) it follows:
COP2 ¼
_ $ðh7 h Þ M 6 _ $ðh h Þ 1þM 2 1
(47)
Ratio of COP2 and COP1 is defined as:
COP21
COP2 ¼ COP1
(48)
COP21 is a measure of the increase of the analyzed system (system 2) compared with the conventional system (system 1). 3.3. Exergy analysis Exergy is defined as a maximum amount of work which can be provided by a stream of matter, heat or work as it comes to equilibrium with the reference environment. In order to find out process inefficiencies and magnitudes of the exergy destruction, exergy analysis should be carried out [35]. One can write specific exergy flow as follows:
ei ¼ hi h0 T0 $ðsi s0 Þ
Specific exergy destructions in system 1 for the compressor ed;comp1 , the condenser ed;cond1, the expansion valve ed;ev1 and the evaporator ed;e1 are calculated according to the eq. (50) using the equations below, respectively:
(49)
Here, subscript “i” denotes the points in Figs. 1e4.T0 is the reference temperature and h0 and s0 are enthalpy and entropy at reference temperature T0 and reference pressure P0, respectively. Reference temperature and reference pressure are assumed as T0 ¼ 25 C and P0 ¼ 101:325 kPa. Exergy balance is then written as:
X X T T ed ¼ ein eout þ q 1 0 q 1 0 T T in out X X þ win wout
ed; cond1 ¼ e21 e4 qcond1 1
(55) T0
Tcond 1
(56)
ed; ev1 ¼ e4 e4a
(57)
T ed; e1 ¼ e4a e7 þ qe1 1 0 Te1
(58)
Here the work supplied to the compressor wcomp1 ; from the condenser transferred heat qcond1 and to the evaporator transferred heat qe1 are determined as follows:
wcomp1 ¼ h21 h7
(59)
qcond1 ¼ h21 h4
(60)
qe1 ¼ h7 h4a
(61)
According to the explanation above, one can write:
Tcond1 ¼ Tc 10 C
(62)
Te1 ¼ Te þ 10 C
(63)
Total exergy destruction in system 1 ed; 1 is then determined using the equation below:
ed; 1 ¼ ed; comp1 þ ed; cond1 þ ed; ev1 þ ed; e1
(64)
Exergy efficiency of a system is defined as:
(50) In the above equation, ed is denoted for destroyed exergy. Moreover, q is the specific heat flow transferred through the boundaries at temperature T. This equation is applied to all component in refrigeration cycles. The temperature T is assumed to be 10 C greater than the evaporation temperature in the evaporator and 10 C less than the condensation temperature in the condenser for convenient and economical heat transfer in evaporator and condenser [36]. Using the above equations, one can
4¼1
Exergy destroyed Exergy expended
(65)
Exergy efficiency in conventional system (system 1) is then yielded:
41 ¼ 1
ed;1 wcomp1
(66)
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3.3.2. Determination of exergy efficiency for new ejector subcooling system (system 2) Specific exergies at points 1, 2, 5, 6, 4e and 8 in Figs. 3 and 4 can be written according to the eq. (49) as follows:
e1 ¼ h1 h0 T0 $ðs1 s0 Þ
(67)
e2 ¼ h2 h0 T0 $ðs2 s0 Þ
(68)
e5 ¼ h5 h0 T0 $ðs5 s0 Þ
(69)
e8 ¼ h8 h0 T0 $ðs8 s0 Þ
(72)
Specific exergy destructions at the compressor ed;comp2, the condenser ed;cond2, the expansion valve 21 ed;ev21, the subcooling heat exchanger ed;he2 ; the expansion valve 22 ed;ev22 , the evaporator ed;e2 and the ejector ed;ej2 are calculated according to the eq. (50) as follows in system 2:
ed; comp2 ¼ e1 e2 þ wcomp2
e6 ¼ h6 h0 T0 $ðs6 s0 Þ
(70)
e4e ¼ h4e h0 T0 $ðs4e s0 Þ
(71)
Table 1 Flow chart for numerical calculations.
ed; cond2 ¼ e2 e4 qcond2 1
_ ðe e Þ ed; ev21 ¼ M 4e BP 4
(73) T0 Tcond2
(74)
(75)
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_ ðe e Þ þ M _ ðe e Þ ed; he2 ¼ 1 M 5 5 4 BP BP 4e
ed; ev22 ¼ 1
ed;
e2
ed; ej2
_ M BP
719
eq. (65) is then yielded:
(76) 42 ¼ 1
ðe5 e6 Þ
(77)
T0 _ ðe 1 ¼ 1M e Þ þ q 7 6 e2 BP Te2
(78)
_ $e þ 1 M _ e7 e ¼M 1 BP 8 BP
(79)
In the above equations, the following expressions are valid:
wcomp2 ¼ h2 h1
(80)
qcond2 ¼ h2 h4
(81)
qe2 ¼ h7 h6
(82)
_ _ ¼ M8 M BP _ M
(83)
ed;2 wcomp2
(89)
Exergy efficiency ratio is defined as follows:
421 ¼
42 41
(90)
421 shows the exergy efficiency increase of the analyzed system 2 compared to the conventional system 1. In this work specific exergies and specific exergy destructions _ are presented. These exergies are valid for total mass flow rate M 1
flowing through points 1e4. 4. Numerical results
1
_ is ratio of subcooling mass flow rate at point 8 to the total M BP mass flow rate at the compressor inlet at point 1. Similar to eqs. (54) and (55) one can write:
Tcond2 ¼ Tc 10 C
(84)
Te2 ¼ Te þ 10 C
(85)
For the isentropic efficiency of the compressor most commonly used values can be taken from literature as hc ¼ 0:8 [37]. Ejector nozzle, ejector mixing, and ejector diffuser efficiencies can be assumed as hn ¼ 0:9; hm ¼ 0:80 and hd ¼ 0:9, respectively [25,38]. Reference temperature and reference pressure are chosen as T0 ¼ 25 C and P0 ¼ 101:325 kPa, respectively. In order to evaluate thermodynamic properties and solve the equations EES [39] is used. Simple flowchart for numerical calculation is presented in Table 1. It is important to take different fluid properties into consideration for the working fluid selection. Besagni et al. [31] and Chen et al. [40] gave important information for screening of working fluids. R32, R245fa and R134a have higher GWP (Global Warming
_ is refrigerant mass flow rate between points 4e8 and M _ is the M 8 1
_ 5 can then be total refrigerant mass flow rate between points 1e4.M calculated as:
_ M _ _5¼M M 1 8
(86)
One can write the equation below using eqs. (23) and (86):
_ ¼ M BP
1
(87)
_ 1þM
Total exergy destruction in system 2 can then be determined using the following equation:
ed; 2 ¼ þed; comp2 þ ed; cond2 þ ed; ev21 þ ed; he2 þ ed; ev22 þ ed; e2 þ ed; ej2 (88) Exergy efficiency in ejector system (system 2) according to the
Fig. 7. Saturation dome of the analyzed refrigerants.
Table 2 Properties of different refrigerants [29,41]. Refrigerant Chemical name
Chemical Formula
Molecular weight
Normal boiling point [ C]
Latent heat of vaporization [kJ/kg] 20 C=10 C
Critical pressure Critical [kPa] temperature [ C]
Safety class
GWP wet/ dry
R32 R1234yf
CH2F2 CF3CF¼CH2
52 114
52 29.4
298.9/344 156.6/175.2
5782 3382.2
78.4 94.70
A2 A2L
675 4
wet wet
C3H8 CF3CH2F NH3 C4H10 CF3CH2CHF2
44.096 102.03 17.03 58.122 134.0
42.11 26.074 33.327 11.75 15.14
359.9/400.8 190.7/212.9 1226/1329 345.1/373.4 199/215.1
4251.2 4059.3 11333.0 3629.0 3651
96.67 101.1 132.25 134.67 154.01
A3 A1 B2 A3 A1
3 1430 0 3 1030
wet wet wet dry dry
R290 R134a R717 R600a R245fa
Difluoromethane 2,3,3,3-tetrafluoro-1propene Propane Tetrafluoroethane Ammonia Isobutane 1,1,1,3,3Pentafluoropropane
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Fig. 8. Effect of DTsc and ejector efficiencies on COP21 for different refrigerants (Tc ¼ 50 C , Te ¼ 20 C, DT58 ¼ 3 C) (HV, MV and LV values are given in Table 3).
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Potential), but they are non-flammable. R717 is added as a natural refrigerant with good physical properties, even if it is toxic. Hydrofluorocarbons (HFC) R290, R134a and R600a are investigated in both works [31,40]. R245fa is included in the work of Besagni et al. [31]. In this study, different refrigerants are used as working fluid which are given in Table 2. In this table; chemical name, chemical formula, molecular weight, normal boiling point, latent heat of evaporation at 20 C=10 C, critical pressure, critical temperature, safety class, GWP and wet/dry conditions are demonstrated [29,41]. All these refrigerants have zero ODP (Ozone Depleting Potential). On the other hand, natural refrigerant R717, hydrocarbons R600a and R290 have very low GWP. Synthetic fluid R1234yf has also very low GWP. The HFC R32, R245fa and R134a are also included to make comparisons with the low GWP refrigerants even if the usage of these refrigerants was limited by EU Regulation 517/2014 [42]. Fig. 11. Variation of COP21 with Tc for different refrigerants ðDTsc ¼ 0:5; Te ¼ 20 C; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
Table 3 High (HV), medium (MV) and low values (LV) of ejector efficiencies.
HV MV LV
hn
hm
hd
0.95 0.90 0.85
0.85 0.80 0.75
0.95 0.90 0.85
Fig. 12. Variation of COP21 with DTsc for different refrigerants (Tc ¼ 50 C, Te ¼ 20 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
Fig. 9. Variation of COP21 with DTsc for R1234yf (Tc ¼ 50 C; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9).
Fig. 10. Variation of COP21 with DT58 for R1234yf ðDTsc ¼ 0:5; Tc ¼ 50 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ.
Fig. 13. Variation of COP21 with Te for different refrigerants ðTc ¼ 50 C; DTsc ¼ 0:5; D T58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
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In Fig. 7 the different saturation dome of the analyzed refrigerants are shown in P-h diagram. This diagram is important, because the slopes of the vapor saturation lines are important for the values of COP and possibility of increase of COP. The slope of the vapor saturation lines increase beginning with R600a and ending with R32. Effect of ejector efficiencies on COP21 is shown in Fig. 8 for all analyzed refrigerants. As parameters, low values (LV), medium values (MV) and high values (HV) of ejector efficiencies are utilized. The values are given in Table 3. COP21 is 1 at DTsc ¼ 0, because there is no subcooling for this case. Near DTsc ¼ 1 we would again get COP21 ¼ 1, because the subcooling and evaporating temperatures would be equal to each other. If a limiting case temperature difference in heat exchanger DT58 ¼ 0 C is assumed, then COP21 ¼ 1 would be arrived at exactly DTsc ¼ 1. Because of these limiting values of COP21 for DTsc ¼ 0 and DTsc ¼ 1, COP21 must have an optimum value for a certain value of DTsc . Highest value of COP21 is achieved for approximately DTsc ¼ 0:5. One can see from Fig. 8 that the influence of ejector efficiency is not too high for all analyzed refrigerants. In Fig. 9, COP21 is demonstrated as a function of DTsc for different values of Te and at DT58 ¼ 3 C and Tc ¼ 50 C . It is seen
Fig. 14. Variation of COP1 with Te for different refrigerants (Tc ¼ 50 CÞ:
Fig. 15. Variation of COP2 with Te for different refrigerants ðTc ¼ 50 C; DT sc ¼ 0:5; D T58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
that, COP21 has a certain optimum value of DTsc which is in the vicinity of 0.5. COP21 increases with the decrease in evaporator temperature Te, because the enthalpy ratio of the h4-h5 to h7-h4a would be increased with the decrease of evaporating temperature Te. In Fig. 10, variation of COP21 with DT58 is shown for Tc ¼ 50 C and different values of Te : It can be seen that, COP21 decreases with the increase in DT58 , as expected. But the influence of DT58 on COP21 is not very important. Therefore, in our calculations DT58 ¼ 3 C is assumed. Results for the variation of COP21 with Tc for different refrigerants are illustrated in Fig. 11 for DTsc ¼ 0:5 and Te ¼ 20 C.COP21 increases with Tc because h7-h4 decreases with the increase of the condenser temperature Tc at a given evaporator temperature Te. In Fig. 12, COP21 is illustrated as a function of DTsc for different refrigerants, for Te ¼ 20 C; Tc ¼ 50 C and DT58 ¼ 3 C. It is seen that highest COP21 is obtained using R1234yf, because R1234yf has lower vapor saturation slope and lower latent heat of evaporation. In Fig. 13, variation of COP21 with Te is shown for DTsc ¼ 0:5; Tc ¼ 50 C and different refrigerants.COP21 decreases with the increase in evaporator temperature Te, because h7-h4 increases with
Fig. 16. Variation of 41 with Tc (Te ¼ 20 C, DT sc ¼ 0:5Þ:
Fig. 17. Variation of 42 with Tc (Te ¼ 20 C, DT sc ¼ 0:5; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
T. Yilmaz, M.T. Erdinç / Energy 168 (2019) 712e727
Fig. 18. Presentation of conventional cycle P-h diagram for different refrigerants (Tc ¼ 50 C and Te ¼ 20 C).
723
724
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Fig. 19. Presentation of ejector cycle P-h diagram for different refrigerants (Tc ¼ 50 C, DT sc ¼ 0:5, Te ¼ 20 C; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9).
T. Yilmaz, M.T. Erdinç / Energy 168 (2019) 712e727
the increase of the evaporator temperature Te at a given condenser temperature Tc. In Figs. 14 and 15, COP1 and COP2 values are demonstrated as a function of evaporator temperature for different refrigerants at Tc ¼ 50 C . As expected, COP1 and COP2 values increase with the increase of the evaporator temperature Te. It is seen that COP2 values of new refrigeration cycle are always higher than the conventional cycle values (COP1) which shows the benefit of the new refrigeration system. The highest COP1 is obtained for R717, because it has high value of latent heat of evaporation. The highest COP2 values are obtained for R245fa. This is because R245fa has, similar to R1234yf, lower slope of vapor saturation line and lower value of latent heat of evaporation. In Figs. 16 and 17, exergy efficiencies 41 and 42 are demonstrated as a function of condenser temperature for different refrigerants. It is seen that exergy efficiencies of new refrigeration cycle 42 are higher than those for the conventional cycle 41 which shows again the effectiveness of the proposed new refrigeration system. The highest exergy efficiency 41 values are obtained with R717 and the highest 42 values are obtained for R245fa. These results are similar to those values obtained for COP1 and COP2. Exergy efficiencies have an optimum point with respect to condenser temperature Tc.
725
In Figs. 18 and 19, P-h diagrams of conventional system and ejector system are demonstrated for different refrigerants and Tc ¼ 50 C, DT sc ¼ 0:5 and Te ¼ 20 C. One can see from Fig. 19 that the pressure at points 1 is much higher than the pressure at point 7, which is one of the causes for increase of COP2 and 42 . In Fig. 20, the exergy efficiency ratio 421 is demonstrated as a function of DT sc for different refrigerants at Tc ¼ 50 C and Te ¼ 20 C. It is seen that the maximum value is around DT sc ¼ 0:5 similar to Fig. 9 for COP 21. In Fig. 21, the variation of exergy efficiency ratio 421 with Te is presented for different refrigerants at Tc ¼ 50 C and DT sc ¼ 0:5. Like COP 21 , 421 is highest for R1234yf and lowest for R717. 421 increases with decrease of the evaporator temperature, because the same amount of subcooling is more effective at lower evaporator temperatures. Figs. 22 and 23 show exergy destruction of components for conventional and ejector cycle. In conventional system exergy destruction is highest in expansion valve and lowest in evaporator and condenser. In ejector cycle, the exergy destruction in the compressor is higher than the sum of exergy destructions in both expansion valves 1 and 2.
Fig. 20. Variation of 421 with DT sc for different refrigerants (Tc ¼ 50 C, Te ¼ 20 C; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
Fig. 22. Variation of exergy destructions ed of conventional system's different components with Tc (Te ¼ 20 CÞ.
Fig. 21. Variation of 421 with Te (Tc ¼ 50 C, DT sc ¼ 0:5; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ.
Fig. 23. Variation of exergy destructions ed of new ejector system's different components with Tc (Te ¼ 20 C, DTsc ¼ 0:5; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
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5. Conclusıon In this work, a new ejector subcooling refrigeration system without an extra pump or compressor has been modeled and calculated numerically for condenser temperatures between 30 C= 70 C and evaporator temperatures between 20 C=10 C. Seven different refrigerants (namely; R32, R1234yf, R290, R134a, R717, R600a and R245fa) are considered as working fluids. According to the obtained results, the following conclusions can be drawn: Highest COP1 value is obtained with R717 because of high latent heat of evaporation. Highest COP2 value is obtained with R245fa, because of low latent heat of evaporation and lower slope of vapor saturation line. Highest 41 value is obtained with R717 and highest 42 value is obtained with R245fa, similar to COP1 and COP2 values. As condenser temperature increases at constant evaporator temperature, the enthalpy difference h7-h4 decreases. Therefore, the effect of subcooling becomes very important. Because of this, COP21 increases with the increase in condenser temperature. Enthalpy difference h7-h4 decreases with the decrease of the evaporator temperature Te at constant condenser temperature Tc. Because of this, COP21 increases with the decrease in evaporator temperature. The highest value of COP21 ð20%Þ and 421 ð18%Þ are obtained for the refrigerant R1234yf, because it has the lowest latent heat of evaporation and lower slope of vapor saturation line. The lowest value of COP21 ð5%Þ and 421 ð5%Þare obtained for the refrigerant R717, because it has the highest latent heat of evaporation and higher slope of vapor saturation line. Exergy efficiencies 41 and 42 have a maximum value for a certain condenser temperature. COP21 and 421 approach to 1 for DTsc /0 and for DTsc /1. There exists an optimum point of COP21 and 421 around DTsc ¼ 0:5 for all refrigerants. Influence of DT58 on COP21 is not very important. The novel refrigeration system has shown that a considerable increase in coefficient of performance (COP) and exergy efficiency (4) can be achieved with the addition of a valve and ejector which do not significantly increase the investment cost of the refrigeration system. For the validation of the presented results, future experimental study is important. Acknowledgment The authors are very grateful to Prof. Dr. Alper YILMAZ for his valuable comments, which have been utilized to improve the quality of this paper. References [1] Hong WJ, Alhussan K, Zhang H, Garris Jr CA. A novel thermally driven rotorvane/pressure-exchange ejector refrigeration system with environmental benefits and energy efficiency. Energy 2004;29:2331e45. [2] Commission Delegated Regulation (EU) No:1060/2010 of 28 September 2010 Supplementation Directive 201030/EU of the European Parliament and of the Council with regard to energy labeling of household refrigerating appliances. [3] Kosalay I, Ilk HG. Experimental studies and test results on the energy efficiency of household refrigerating appliances. J Sci Ind Res 2017;76:575e80. [4] Qureshi BA, Zubair SM. Mechanical sub-cooling vapor compression systems: current status and future directions. Int. J. Refrigeration 2013;36:2097e110. [5] Zubair SM. Improvement of refrigeration/air-conditioning performance with mechanical sub-cooling. Energy 1990;15:427e33.
[6] Khan J, Zubair SM. Design and rating of an integrated mechanical-subcooling vapor-compression refrigeration system. Energy Convers Manag 2000;41(11): 1201e22. [7] Dai B, Liu S, Li H, Sun Z, Song M, Yang Q. Energetic performance of transcritical CO2 refrigeration cycles with mechanical subcooling using zeotropic mixture as refrigerant. Energy 2018;150:205e21. [8] Roy R, Mandal BK. Thermodynamic analysis of a vapor compression refrigeration system integrated with a subcooler cycle. Int J Renew Energy Technol 2017;8:334e45. [9] Yu J, Ren Y, Chen H, Li Y. Applying mechanical subcooling to ejector refrigeration cycle for improving the coefficient of performance. Energy Convers Manag 2007;48:1193e9. [10] Pridasawas W. Solar-driven refrigeration systems with focus on the ejector cycle. PhD Thesis. Stockholm: Royal Institute of Technology; 2006. [11] Kim DS, Infante Ferreira CA. Solar refrigeration options-a state-of-the-art review. Int. J. Refrigeration 2008;31:3e15. [12] Abdullateef JM, Sopian K, Alghoul MA, Sulaiman MY. Review on solar driven ejector refrigeration technologies. Renew Sustain Energy Rev 2009;13: 1338e49. [13] Chen X, Omer S, Worall M, Riffat S. Recent developments in ejector refrigeration technologies. Renew Sustain Energy Rev 2013;19:629e51. [14] He S, Li Y, Wang RZ. Progress of mathematical modeling on ejectors. Renew Sustain Energy Rev 2009;13:1760e80. [15] Ersoy HK, Sag NB. Preliminary experimental results on the R134a refrigeration system using a two-phase ejector as an expander. Int. J. Refrigeration 2014;43:97e110. [16] Lin C, Li Y, Cai W, Yan J, Hu Y. Experimental investigation of the adjustable ejector in a multi-evaporator refrigeration system. Appl Therm Eng 2013;61: 2e10. [17] Lin C, Cai W, Li Y, Yan J, Hu Y. Numerical investigation of geometry parameters for pressure recovery of an adjustable ejector in multi-evaporator refrigeration system. Appl Therm Eng 2013;61:649e56. [18] Sarkar J. Performance analyses of novel two-phase ejector enhanced multievaporator refrigeration systems. Appl Therm Eng 2017;110:1635e42. [19] Ünal S¸, Yılmaz T. Thermodynamic analysis of the two-phase ejector airconditioning system for buses. Appl Therm Eng 2015;79:108e16. [20] Ünal S¸, Erdinç MT, Kutlu Ç. Optimal thermodynamic parameters of two-phase ejector refrigeration system for buses. Appl Therm Eng 2017;124:1354e67. [21] Bai T, Yan G, Yu J. Thermodynamics analysis of a modified dual-evaporator CO2 transcritical refrigeration cycle with two-stage ejector. Energy 2015;84: 325e35. [22] Yifei W, Hongxia Z, Cunquan Z, Lei W. Optimization analysis of structure parameters of steam ejector based on CFD and orthogonal test. Energy 2018;151:79e93. [23] Michal H, Jacek S, Armin H, Ziemowit O, Michal P, Andrzej JN, Krzysztof B. System model derivation of the CO2 two-phase ejector based on the CFDbased reduced-order model. Energy 2018;144:941e56. [24] Ünal S¸. Determination of the ejector dimensions of a bus air-conditioning system using analytical and numerical methods. Appl Therm Eng 2015;90: 110e9. [25] Besagni G, Mereu R, Chiesa P, Inzoli F. An integrated lumped parameter-CFD approach for off-design ejector performance evaluation. Energy Convers Manag 2015;105:697e715. [26] Besagni G, Mereu R, Inzoli F, Chiesa P. Application of an integrated lumped parameter-CFD approach to evaluate the ejector-driven anode recirculation in a PEM fuel cell system. Appl Therm Eng 2017;121:628e51. [27] Varga S, Armando CO, Bogdan D. Numerical assessment of steam ejector efficiencies using CFD. Int. J. Refrigeration 2009;32(6):1203e11. [28] Besagni G, Inzoli F. Computational fluid-dynamics modeling of supersonic ejectors: screening of turbulence modeling approaches. Appl Therm Eng 2017;117:122e44. [29] Besagni G, Mereu R, Inzoli F. Ejector refrigeration: a comprehensive review. Renew Sustain Energy Rev 2016;53:373e407. [30] Chunnanond K, Aphornratana S. Ejectors: applications in refrigeration technology. Renew Sustain Energy Rev 2004;8:129e55. [31] Besagni G, Mereu R, Di LG, Inzoli F. A study of working fluids for heat driven ejector refrigeration using lumped parameter models. Int. J. Refrigeration 2015;58:154e71. [32] Gil B, Kasperski J. Efficiency evaluation of the ejector cooling cycle using a new generation of HFO/HCFO refrigerant as a R134a replacement. Energies 2018;11:2136. [33] Besagni G, Croci L, Nesa R. A screening of Low-GWP refrigerant for ejector refrigeration. Chem Eng Trans 2018;70:1291e6. [34] Yılmaz T. Ejector enhanced subcooling vapor compression machine. Patent application. Turkish Patent Office; April 2018. [35] Dincer I, Rosen MA. Exergy, energy, environment and sustainable development. first ed. Oxford, U.K: Elsevier; 2007. [36] Lawrence N, Elbel S. Theoretical and practical comparison of two-phase ejector refrigeration cycles including first and second law analysis. Int. J. Refrigeration 2013;36:1220e33. 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Contents lists available at ScienceDirect
Energy journal homepage: www.elsevier.com/locate/energy
Energetic and exergetic investigation of a novel refrigeration system utilizing ejector integrated subcooling using different refrigerants Tuncay Yilmaz*, Mehmet Tahir Erdinç Department of Mechanical Engineering, Osmaniye Korkut Ata University, 80000 Osmaniye, Turkey
a r t i c l e i n f o
a b s t r a c t
Article history: Received 29 May 2018 Received in revised form 19 November 2018 Accepted 20 November 2018 Available online 24 November 2018
Increase in coefficient of performance of refrigeration systems is eminently important for various applications such as household refrigerator, tumble dryer, and more in practical applications. A new ejector subcooling system is described and investigated numerically without using an extra pump or compressor. After condenser, some part of the liquid (by-pass refrigerant) is expanded so that its temperature could be lower than the intended subcooling temperature. By using this by-pass refrigerant, the main flow of the refrigerant is subcooled. After evaporation, the by-pass refrigerant is expanded using an ejector. Seven different refrigerants are considered as working fluids (namely; R32, R1234yf, R290, R134a, R717, R600a and R245fa). R1234yf, R290, R600a and R717 are considered as low global warming potential refrigerants in European Union regulation. For each refrigerant, the values of coefficient of performance, relative increment in coefficient of performance and exergy efficiencies are calculated and demonstrated graphically for different condenser temperatures between 30 C=70 C and evaporator temperatures between 20 C=10 C. Results show that the best performance is obtained for R1234yf with an increment of about 20% in coefficient of performance and 18% in exergy efficiency. These efficiencies can be higher or lower approximately by 3% using ejector with higher or lower component efficiencies. © 2018 Elsevier Ltd. All rights reserved.
Keywords: Refrigeration Ejector Subcooling Energetic Exergetic
1. Introduction Refrigeration is perhaps the most important application for reducing the environmental impact of global energy use [1]. Therefore, European Union (EU) has issued regulations on the energy efficiency levels [2]. The importance of energy efficiency of household refrigerator is explained by Kosalay et al. [3]. They reported that these household refrigerators electricity bill constitutes over 30% of total domestic electricity bill in some countries. One of the most important method to increase the efficiency of the refrigeration systems is subcooling. Qureshi and Zubair [4] broadly described different possibilities for subcooling, such as ambient subcooling, subcooling with liquid-suction heat exchanger, subcooling with external heat exchanger and dedicated-integrated mechanical subcooling. As mentioned above, different methods are applicable for subcooling. In the first one, the liquid refrigerant can be cooled down after the condenser with the cold refrigerant which leaves the
* Corresponding author. E-mail address: [email protected] (T. Yilmaz). https://doi.org/10.1016/j.energy.2018.11.081 0360-5442/© 2018 Elsevier Ltd. All rights reserved.
evaporator. However, this method has the disadvantage of increasing the refrigerant temperature at the inlet of the compressor which causes some decrement in mass flow rate flowing through the compressor and hence decreases the capacity of refrigeration. Subcooling with external heat exchanger needs heat exchanger utilizing cooling water. However, this again needs cooling tower or similar devices with a circulating water pump which increase significantly the investment and maintenance costs. Another method is taking advantage of using a dedicated and integrated mechanical subcooling system in the cycle. A second refrigeration cycle is adopted for cooling the liquid after condensing process. However, this additional refrigeration cycle has the main drawback of increasing the investment cost of the system. Zubair [5] found that energy savings in between 20% and 40% can be obtained using mechanical subcooling. Khan and Zubair [6] found that the optimum subcooling temperature should be halfway between the condensation and evaporation temperatures. Dai et al. [7] studied mechanical subcooling refrigeration system using zeotropic mixtures. They found that optimum coefficient of performance (COP) is achieved at the optimum discharge pressure and the optimum subcooling degree. Roy and Mandal [8] investigated
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Nomenclature COP1 COP2 COP21 e ed h _ M _ M P Q_
q s T u w _ W x
coefficient of performance of the conventional system [] coefficient of performance of the ejector subcooling system [] coefficient of performance ratio [] specific exergy [kJ kg1] specific exergy destruction [kJ kg1] enthalpy [kJ kg1] mass flow rate [kg s1] mass flow rate ratio [] pressure [Pa] heat flow [kW] specific heat [kJ kg1] entropy [kJ kg1K1] temperature [ C ] velocity [ m s1] specific work [kJ kg1] power [kW] quality []
Greek Letters h efficiency [] 41 exergy efficiency of conventional system [] 42 exergy efficiency of ejector subcooling system [] 421 exergy efficiency ratio []
c comp1 comp2 cond1 cond2 d e1 e2 ev1 ev21 ev22 f he in m n out sc 0
713
compressor conventional system compressor ejector system compressor conventional system condenser ejector system condenser diffuser, destroyed, destruction conventional system evaporator ejector system evaporator conventional system expansion valve ejector system expansion valve 1 ejector system expansion valve 2 flashing point subcooling heat exchanger inlet mixing region nozzle outlet subcooling reference
Abbreviations COP coefficient of performance EU European Union ev expansion valve GWP global warming potential he subcooling heat exchanger HFC hydrofluorocarbons ODP ozone depleting potential
Subscripts BP by-pass
COP and exergetic efficiency of mechanical subcooling system and found increase of 16% in COP and decrease of 8.8% in exergy destruction rate at 10 C evaporator temperature for R134a. Yu et al. [9] described a refrigeration system with mechanical subcooling which uses an auxiliary liquidegas ejector and a liquid pump to enhance subcooling for the refrigerant leaving the condenser. This study proposes that ejector integrated system has advantages over other subcooling methods, therefore ejector technology has also been reviewed. There are many studies for ejector technology in literature. Because refrigeration systems which use solar heat at low temperatures are very interesting and have been investigated for many times. On this matter, extensive information was supplied by Pridasawas [10] and Kim and Infante Ferreira [11]. Abdullateef et al. [12] reported on the special topics of solar driven ejector refrigeration systems. New developments in ejector refrigeration systems (ERS) were depicted and explained by Chen et al. [13]. They explained especially geometric optimization of ejectors. Progress on mathematical modeling of ejectors were described and analyzed by He et al. [14]. They stated that finite difference method is recognized as the most universal technique for numerical solutions. Some researchers have focused on reduction of throttling losses in expansion valve by using ejector as an expansion valve in order to reduce the compressor work and increase COP of the system. Ersoy and Sag [15] investigated the utilization of ejector as expansion valve using R134a and found a COP enhancement in the range of 6e14%. Lin et al. [16,17] studied experimentally and numerically optimum geometric parameters of ejector. Sarkar [18]
analyzed and optimized four novel layouts of ejector driven multicompression three-evaporator refrigeration system using R32 as refrigerant. Ünal and Yılmaz [19] and Ünal et al. [20] studied two phase ejector and two evaporators using R134a as refrigerant in a bus refrigeration system. They found 15% increase in COP using the two-phase ejector in bus air-conditioning system compared with the conventional system. Besides this, reducing total weight of the bus, fuel saving can be achieved. Bai et al. [21] studied dualevaporator CO2 transcritical refrigeration cycle with two-stage ejector. They reported higher COP and exergy efficiency. In order to see the effect of the dimensions on the ejector system performance, numerical studies have been performed [22e28]. Comprehensive review on ejector refrigeration is given by Besagni et al. [29] and Chunnanond and Aphornratana [30]. Besagni et al. [31] carried out research work on working fluids for heat driven ejector refrigeration and found that each refrigerant has its own range of operating conditions. Gil and Kasperski [32] studied hydrofluoroolefins as working fluids in ejector refrigeration. They stated that the most benefit comes from the use of R-1234yf and R1234ze(E). The influence of generator temperature, working fluid and ejector component efficiency are presented by Besagni et al. [33]. They showed that ejector component efficiencies have great influence on COP. As stated above, the increase of efficiency and capacity of the conventional system is not very effective using subcooling of the liquid refrigerant at the exit of the condenser with the cold vapor from the exit of the evaporator. The subcooling with external heat exchanger needs in addition to the subcooling heat exchanger cooling tower and circulation pump. Dedicated mechanical
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subcooling needs besides subcooling heat exchanger, condenser and an other compressor. Similarly, for integrated mechanical subcooling, subcooling heat exchanger, receiver and an other compressor are necessary. Therefore, subcooling with external heat exchanger and dedicated and integrated mechanical subcooling systems increase significantly the capital and maintenance cost of the refrigeration system. In this work, a new ejector subcooling system [34] is described and numerically investigated without an extra pump or compressor. Besides subcooling heat exchanger, only an ejector and a valve are needed. Therefore, the most advantages of this new refrigeration system are its simplicity, its low capital and maintenance cost. 2. System description Thermodynamic analysis of the ejector subcooling refrigeration system is carried out by considering the following assumptions [19]:
Fig. 2. P-h diagram of conventional refrigeration system (system 1).
a) b) c) d)
Pressure losses of the system are neglected. Throttling process in expansion valve is isenthalpic. Isentropic efficiency of the compressor is prescribed. Isentropic efficiencies of the nozzle and the diffusor are prescribed. e) Efficiency of mixing section of the ejector is prescribed. f) The process in the mixing section takes place at constant pressure and constant cross-sectional area.
A conventional refrigeration system and its P-h diagram are presented in Figs. 1 and 2, respectively. For the sake of simplicity, compressor inlet is assumed as saturated vapor and condenser outlet is assumed as saturated liquid. The novel ejector subcooling refrigeration system [34] is shown schematically in Fig. 3 and its P-h diagram is presented in Fig. 4. After condensation process, some part (by-pass refrigerant) of the liquid refrigerant mass flow rate is expanded until it decreases to temperature T8, which is just below the subcooling temperature T5. The remaining liquid refrigerant is subcooled from temperature T4 till temperature T5 utilizing the enthalpy of evaporation of the refrigerant from point 4e to point 8. Afterwards, refrigerant pressure P7 is increased to the compressor inlet pressure P1 using the
Fig. 3. New subcooling refrigeration system (system 2).
ejector. T-Q_ diagram for the heat exchanger is presented in Fig. 5. Q_ is heat transferred in subcooling heat exchanger. he
3. Thermodynamic calculations 3.1. Conventional refrigeration cycle
Fig. 1. Conventional refrigeration system (system 1).
Conventional refrigeration cycle is given by points 7-21-3-4-4a7 in Fig. 2. In this cycle the evaporator and the condenser temperatures are Te and Tc. T4a ¼ T7 ¼ Te, T3 ¼ T4 ¼ Tc are prescribed and therefore condenser and evaporator pressures (Pc and Pe) that are assumed constant are given as Pc ¼ P21¼P21s ¼ P2¼P2s ¼ P3¼P4 and Pe ¼ P6¼P4a ¼ P9¼P9s ¼ P10¼P7. In the conventional cycle,
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715
_ is work supplied where Q_ e is heat supplied to the evaporator and W to the compressor. For system 1 one can write:
COP1 ¼
qe1 w1
(7)
because refrigerant mass flow rates in evaporator and condenser are equal to each other. Specific heat qe1 and specific work w1 can be determined from the enthalpies as follows:
qe1 ¼ h7 h4a
(8)
w1 ¼ h21 h7
(9)
where h21 is calculated from the definition of isentropic compressor efficiency: Fig. 4. New subcooling refrigeration system (system 2) in P-h diagram.
h21 ¼ h7 þ
h21s h7
hc
(10)
Using eqs.(7)-(10), COP of the conventional cycle COP1 (system 1) is determined as follows:
COP1 ¼
h7 h4a h21 h7
(11)
3.2. Ejector integrated subcooling system Flashing occurs by the expansion from point 5 to point 6 as shown in Fig. 4. Flashing onset temperature is T5f. One can write approximately the equation below for T5f
T5f ¼ T5
(12)
because liquid refrigerant specific heat can be assumed independent of temperature. Temperature T4e must be below temperature T5, and it can be calculated as follows: Fig. 5. T-Q_ diagram for subcooling heat exchanger (he).
T8 ¼ T4e ¼ T5 DT58
(13)
DT58 is chosen between 2 5 C in heat exchangers. In this work D
superheating and subcooling are assumed to be zero. We can find enthalpies at points 7, 4 and 4a assuming saturated vapor at point 7, saturated liquid at point 4 and isenthalpic expansion from point 4 to point 4a:
T58 ¼ 3 C is assumed. Pressure P8 and temperature T8 for the subcooling are determined using the below expressions because all refrigerants are pure fluids:
h7 ¼ f ðP7 ; x7 ¼ 1Þ
(1)
T8 ¼ T4e
(14)
h4 ¼ f ðP4 ; x4 ¼ 0Þ
(2)
P8 ¼ P4e
(15)
h4a ¼ h4
(3)
Assuming saturated vapor at point 8, P8 is calculated by: Point 21s is compressor exit undergoing isentropic compression. Therefore:
P8 ¼ f ðT8 ; x8 ¼ 1Þ
(16)
The subcooling temperature is given by;
s21s ¼ s7
(4)
T5 ¼ T4 DTsc $ðT4 T7 Þ
h21s ¼ f ðP21s ; s21s Þ
(5)
where DTsc is dimensionless temperature which is defined as follows:
The COP of the refrigeration cycles is defined as follows:
Q_ COP ¼ e _ W
DTsc ¼ (6)
T5 T4 ðTc Te Þ
(17)
(18)
DTsc is a dimensionless parameter which changes between 0 and 1. DTsc is used as a parameter in the calculations. Because the
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temperature after subcooling T5 is known and the pressure at point 5 is equal to condenser pressure P4, one can write:
h5 ¼ fðT5 ; P4 Þ
(19)
The expansion process from point 5 to 6 is isenthalpic, therefore:
h6 ¼ h5
(20)
At the exit of the evaporator, refrigerant vapor is assumed as saturated. Then, enthalpy of the point 7 can be determined as:
h7 ¼ fðT7 ; x7 ¼ 1Þ
(21)
The vapor at the exit of the heat exchanger (point 8) is assumed to be saturated. Therefore, the following equation for the enthalpy at point 8 can be written:
h8 ¼ fðT8 ; x8 ¼ 1Þ
(23)
8
_ and M _ 5 are mass flow rate at ejector inlet at point 8 and where M 8 _ ¼ ∞ means no primary mass flow rate at point 5, respectively. M
_ the heat balance at the subsubcooling. For the calculation of M cooling heat exchanger (he) is used:
_ $ðh h Þ _ $ðh h Þ ¼ M M 5 5 8 8 4e 4
0:5 hm u10 ¼ u9 $ _ 1þM
(29)
u10 is the velocity at point 10. The velocities u at the points 7, 8 and evaporator exit are neglected. Besides these, the following equations are valid:
s9s ¼ s8
(30)
s8 ¼ fðP8 ; x8 ¼ 1Þ
(31)
h9s ¼ f ðP ¼ P9s ; s ¼ s9s Þ
(32)
P9s ¼ P7
(33)
Neglecting velocity u1, the energy equation between the points 10 and 1 can be written as follows:
. h1 ¼ h10 þ u210 2
(34)
Besides this, from the definition of the isentropic efficiency of the diffuser hd, the enthalpy at point 1 can be determined:
h1 ¼ h10 þ
_ is determined from the eqs. (23) and (24): M
ðh1s h10 Þ
hd
(35)
For the entropies s10 and s1s the following relations are valid:
(25)
Compressor inlet point 1 is determined using ejector (Fig. 6) equations which are written according to Ünal and Yılmaz [19]. From the definition of nozzle isentropic efficiency hn, the equation below can be written for the enthalpy at point 9:
h9 ¼ h8 hn $ðh8 h9s Þ
(28)
From the definition of mixing efficiency hm, the following equation is obtained:
(24)
_ ¼ h8 h4e M h4 h5
!
(22)
Mass flow rate ratio is defined as;
_ _ ¼ M5 M _ M
2 u2 _ $h7 ¼ 1 þ M _ $ h þ u10 h9 þ 9 þ M 10 2 2
(26)
s10 ¼ f ðh10 ; P7 Þ
(36)
s1s ¼ s10
(37)
The enthalpy at point 1s, temperature and entropy at point 1 and entropy at point 2s are determined from the relations below:
h1s ¼ f ðs1s ; P1 Þ
(38)
Neglecting the velocity at point 8, the energy equation between the points 8 and 9 yields:
T1 ¼ fðh1 ; P1 Þ
(39)
u9 ¼ ð2$ðh8 h9 ÞÞ0:5
s1 ¼ fðh1 ; P1 Þ
(40)
s2s ¼ s1
(41)
(27)
u9 is the velocity at point 9. Energy equation between the nozzle exit and the mixing region is given below:
The enthalpy at points 2 is calculated from the isentropic
Fig. 6. Detailed presentation of ejector.
T. Yilmaz, M.T. Erdinç / Energy 168 (2019) 712e727
calculate exergies at all points and exergy destructions at all components.
efficiency definition of the compressor:
h2 ¼ h1 þ
h2s h1
hc
(42)
with
h2s ¼ fðs2s ; P3 Þ
(43)
The isentropic efficiency of the compressor hc is assumed to be equal for both compressor of system 1 and system 2. The coefficient of performance COP2 for the new ejector subcooling system (system 2) can then be determined using eq. (6) as:
COP2 ¼
Q_ e2 _ W
717
(44)
3.3.1. Determination of exergy efficiency for conventional system (system 1) Specific exergies at the points 7, 21, 4 and 4a in Figs. 1 and 2 can be written according to eq. (49) as follows:
e7 ¼ h7 h0 T0 $ðs7 s0 Þ
(51)
e21 ¼ h21 h0 T0 $ðs21 s0 Þ
(52)
e4 ¼ h4 h0 T0 $ðs4 s0 Þ
(53)
e4a ¼ h4a h0 T0 $ðs4a s0 Þ
(54)
2
where
Q_ e2 ¼ M_ 5 $ðh7 h6 Þ
(45)
_ ¼ M_ 5 þ M_ $ðh h Þ W 2 8 2 1
(46)
ed; comp1 ¼ e7 e21 þ wcomp1
From eqs. (23) and (44)e(46) it follows:
COP2 ¼
_ $ðh7 h Þ M 6 _ $ðh h Þ 1þM 2 1
(47)
Ratio of COP2 and COP1 is defined as:
COP21
COP2 ¼ COP1
(48)
COP21 is a measure of the increase of the analyzed system (system 2) compared with the conventional system (system 1). 3.3. Exergy analysis Exergy is defined as a maximum amount of work which can be provided by a stream of matter, heat or work as it comes to equilibrium with the reference environment. In order to find out process inefficiencies and magnitudes of the exergy destruction, exergy analysis should be carried out [35]. One can write specific exergy flow as follows:
ei ¼ hi h0 T0 $ðsi s0 Þ
Specific exergy destructions in system 1 for the compressor ed;comp1 , the condenser ed;cond1, the expansion valve ed;ev1 and the evaporator ed;e1 are calculated according to the eq. (50) using the equations below, respectively:
(49)
Here, subscript “i” denotes the points in Figs. 1e4.T0 is the reference temperature and h0 and s0 are enthalpy and entropy at reference temperature T0 and reference pressure P0, respectively. Reference temperature and reference pressure are assumed as T0 ¼ 25 C and P0 ¼ 101:325 kPa. Exergy balance is then written as:
X X T T ed ¼ ein eout þ q 1 0 q 1 0 T T in out X X þ win wout
ed; cond1 ¼ e21 e4 qcond1 1
(55) T0
Tcond 1
(56)
ed; ev1 ¼ e4 e4a
(57)
T ed; e1 ¼ e4a e7 þ qe1 1 0 Te1
(58)
Here the work supplied to the compressor wcomp1 ; from the condenser transferred heat qcond1 and to the evaporator transferred heat qe1 are determined as follows:
wcomp1 ¼ h21 h7
(59)
qcond1 ¼ h21 h4
(60)
qe1 ¼ h7 h4a
(61)
According to the explanation above, one can write:
Tcond1 ¼ Tc 10 C
(62)
Te1 ¼ Te þ 10 C
(63)
Total exergy destruction in system 1 ed; 1 is then determined using the equation below:
ed; 1 ¼ ed; comp1 þ ed; cond1 þ ed; ev1 þ ed; e1
(64)
Exergy efficiency of a system is defined as:
(50) In the above equation, ed is denoted for destroyed exergy. Moreover, q is the specific heat flow transferred through the boundaries at temperature T. This equation is applied to all component in refrigeration cycles. The temperature T is assumed to be 10 C greater than the evaporation temperature in the evaporator and 10 C less than the condensation temperature in the condenser for convenient and economical heat transfer in evaporator and condenser [36]. Using the above equations, one can
4¼1
Exergy destroyed Exergy expended
(65)
Exergy efficiency in conventional system (system 1) is then yielded:
41 ¼ 1
ed;1 wcomp1
(66)
718
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3.3.2. Determination of exergy efficiency for new ejector subcooling system (system 2) Specific exergies at points 1, 2, 5, 6, 4e and 8 in Figs. 3 and 4 can be written according to the eq. (49) as follows:
e1 ¼ h1 h0 T0 $ðs1 s0 Þ
(67)
e2 ¼ h2 h0 T0 $ðs2 s0 Þ
(68)
e5 ¼ h5 h0 T0 $ðs5 s0 Þ
(69)
e8 ¼ h8 h0 T0 $ðs8 s0 Þ
(72)
Specific exergy destructions at the compressor ed;comp2, the condenser ed;cond2, the expansion valve 21 ed;ev21, the subcooling heat exchanger ed;he2 ; the expansion valve 22 ed;ev22 , the evaporator ed;e2 and the ejector ed;ej2 are calculated according to the eq. (50) as follows in system 2:
ed; comp2 ¼ e1 e2 þ wcomp2
e6 ¼ h6 h0 T0 $ðs6 s0 Þ
(70)
e4e ¼ h4e h0 T0 $ðs4e s0 Þ
(71)
Table 1 Flow chart for numerical calculations.
ed; cond2 ¼ e2 e4 qcond2 1
_ ðe e Þ ed; ev21 ¼ M 4e BP 4
(73) T0 Tcond2
(74)
(75)
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_ ðe e Þ þ M _ ðe e Þ ed; he2 ¼ 1 M 5 5 4 BP BP 4e
ed; ev22 ¼ 1
ed;
e2
ed; ej2
_ M BP
719
eq. (65) is then yielded:
(76) 42 ¼ 1
ðe5 e6 Þ
(77)
T0 _ ðe 1 ¼ 1M e Þ þ q 7 6 e2 BP Te2
(78)
_ $e þ 1 M _ e7 e ¼M 1 BP 8 BP
(79)
In the above equations, the following expressions are valid:
wcomp2 ¼ h2 h1
(80)
qcond2 ¼ h2 h4
(81)
qe2 ¼ h7 h6
(82)
_ _ ¼ M8 M BP _ M
(83)
ed;2 wcomp2
(89)
Exergy efficiency ratio is defined as follows:
421 ¼
42 41
(90)
421 shows the exergy efficiency increase of the analyzed system 2 compared to the conventional system 1. In this work specific exergies and specific exergy destructions _ are presented. These exergies are valid for total mass flow rate M 1
flowing through points 1e4. 4. Numerical results
1
_ is ratio of subcooling mass flow rate at point 8 to the total M BP mass flow rate at the compressor inlet at point 1. Similar to eqs. (54) and (55) one can write:
Tcond2 ¼ Tc 10 C
(84)
Te2 ¼ Te þ 10 C
(85)
For the isentropic efficiency of the compressor most commonly used values can be taken from literature as hc ¼ 0:8 [37]. Ejector nozzle, ejector mixing, and ejector diffuser efficiencies can be assumed as hn ¼ 0:9; hm ¼ 0:80 and hd ¼ 0:9, respectively [25,38]. Reference temperature and reference pressure are chosen as T0 ¼ 25 C and P0 ¼ 101:325 kPa, respectively. In order to evaluate thermodynamic properties and solve the equations EES [39] is used. Simple flowchart for numerical calculation is presented in Table 1. It is important to take different fluid properties into consideration for the working fluid selection. Besagni et al. [31] and Chen et al. [40] gave important information for screening of working fluids. R32, R245fa and R134a have higher GWP (Global Warming
_ is refrigerant mass flow rate between points 4e8 and M _ is the M 8 1
_ 5 can then be total refrigerant mass flow rate between points 1e4.M calculated as:
_ M _ _5¼M M 1 8
(86)
One can write the equation below using eqs. (23) and (86):
_ ¼ M BP
1
(87)
_ 1þM
Total exergy destruction in system 2 can then be determined using the following equation:
ed; 2 ¼ þed; comp2 þ ed; cond2 þ ed; ev21 þ ed; he2 þ ed; ev22 þ ed; e2 þ ed; ej2 (88) Exergy efficiency in ejector system (system 2) according to the
Fig. 7. Saturation dome of the analyzed refrigerants.
Table 2 Properties of different refrigerants [29,41]. Refrigerant Chemical name
Chemical Formula
Molecular weight
Normal boiling point [ C]
Latent heat of vaporization [kJ/kg] 20 C=10 C
Critical pressure Critical [kPa] temperature [ C]
Safety class
GWP wet/ dry
R32 R1234yf
CH2F2 CF3CF¼CH2
52 114
52 29.4
298.9/344 156.6/175.2
5782 3382.2
78.4 94.70
A2 A2L
675 4
wet wet
C3H8 CF3CH2F NH3 C4H10 CF3CH2CHF2
44.096 102.03 17.03 58.122 134.0
42.11 26.074 33.327 11.75 15.14
359.9/400.8 190.7/212.9 1226/1329 345.1/373.4 199/215.1
4251.2 4059.3 11333.0 3629.0 3651
96.67 101.1 132.25 134.67 154.01
A3 A1 B2 A3 A1
3 1430 0 3 1030
wet wet wet dry dry
R290 R134a R717 R600a R245fa
Difluoromethane 2,3,3,3-tetrafluoro-1propene Propane Tetrafluoroethane Ammonia Isobutane 1,1,1,3,3Pentafluoropropane
720
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Fig. 8. Effect of DTsc and ejector efficiencies on COP21 for different refrigerants (Tc ¼ 50 C , Te ¼ 20 C, DT58 ¼ 3 C) (HV, MV and LV values are given in Table 3).
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721
Potential), but they are non-flammable. R717 is added as a natural refrigerant with good physical properties, even if it is toxic. Hydrofluorocarbons (HFC) R290, R134a and R600a are investigated in both works [31,40]. R245fa is included in the work of Besagni et al. [31]. In this study, different refrigerants are used as working fluid which are given in Table 2. In this table; chemical name, chemical formula, molecular weight, normal boiling point, latent heat of evaporation at 20 C=10 C, critical pressure, critical temperature, safety class, GWP and wet/dry conditions are demonstrated [29,41]. All these refrigerants have zero ODP (Ozone Depleting Potential). On the other hand, natural refrigerant R717, hydrocarbons R600a and R290 have very low GWP. Synthetic fluid R1234yf has also very low GWP. The HFC R32, R245fa and R134a are also included to make comparisons with the low GWP refrigerants even if the usage of these refrigerants was limited by EU Regulation 517/2014 [42]. Fig. 11. Variation of COP21 with Tc for different refrigerants ðDTsc ¼ 0:5; Te ¼ 20 C; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
Table 3 High (HV), medium (MV) and low values (LV) of ejector efficiencies.
HV MV LV
hn
hm
hd
0.95 0.90 0.85
0.85 0.80 0.75
0.95 0.90 0.85
Fig. 12. Variation of COP21 with DTsc for different refrigerants (Tc ¼ 50 C, Te ¼ 20 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
Fig. 9. Variation of COP21 with DTsc for R1234yf (Tc ¼ 50 C; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9).
Fig. 10. Variation of COP21 with DT58 for R1234yf ðDTsc ¼ 0:5; Tc ¼ 50 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ.
Fig. 13. Variation of COP21 with Te for different refrigerants ðTc ¼ 50 C; DTsc ¼ 0:5; D T58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
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In Fig. 7 the different saturation dome of the analyzed refrigerants are shown in P-h diagram. This diagram is important, because the slopes of the vapor saturation lines are important for the values of COP and possibility of increase of COP. The slope of the vapor saturation lines increase beginning with R600a and ending with R32. Effect of ejector efficiencies on COP21 is shown in Fig. 8 for all analyzed refrigerants. As parameters, low values (LV), medium values (MV) and high values (HV) of ejector efficiencies are utilized. The values are given in Table 3. COP21 is 1 at DTsc ¼ 0, because there is no subcooling for this case. Near DTsc ¼ 1 we would again get COP21 ¼ 1, because the subcooling and evaporating temperatures would be equal to each other. If a limiting case temperature difference in heat exchanger DT58 ¼ 0 C is assumed, then COP21 ¼ 1 would be arrived at exactly DTsc ¼ 1. Because of these limiting values of COP21 for DTsc ¼ 0 and DTsc ¼ 1, COP21 must have an optimum value for a certain value of DTsc . Highest value of COP21 is achieved for approximately DTsc ¼ 0:5. One can see from Fig. 8 that the influence of ejector efficiency is not too high for all analyzed refrigerants. In Fig. 9, COP21 is demonstrated as a function of DTsc for different values of Te and at DT58 ¼ 3 C and Tc ¼ 50 C . It is seen
Fig. 14. Variation of COP1 with Te for different refrigerants (Tc ¼ 50 CÞ:
Fig. 15. Variation of COP2 with Te for different refrigerants ðTc ¼ 50 C; DT sc ¼ 0:5; D T58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
that, COP21 has a certain optimum value of DTsc which is in the vicinity of 0.5. COP21 increases with the decrease in evaporator temperature Te, because the enthalpy ratio of the h4-h5 to h7-h4a would be increased with the decrease of evaporating temperature Te. In Fig. 10, variation of COP21 with DT58 is shown for Tc ¼ 50 C and different values of Te : It can be seen that, COP21 decreases with the increase in DT58 , as expected. But the influence of DT58 on COP21 is not very important. Therefore, in our calculations DT58 ¼ 3 C is assumed. Results for the variation of COP21 with Tc for different refrigerants are illustrated in Fig. 11 for DTsc ¼ 0:5 and Te ¼ 20 C.COP21 increases with Tc because h7-h4 decreases with the increase of the condenser temperature Tc at a given evaporator temperature Te. In Fig. 12, COP21 is illustrated as a function of DTsc for different refrigerants, for Te ¼ 20 C; Tc ¼ 50 C and DT58 ¼ 3 C. It is seen that highest COP21 is obtained using R1234yf, because R1234yf has lower vapor saturation slope and lower latent heat of evaporation. In Fig. 13, variation of COP21 with Te is shown for DTsc ¼ 0:5; Tc ¼ 50 C and different refrigerants.COP21 decreases with the increase in evaporator temperature Te, because h7-h4 increases with
Fig. 16. Variation of 41 with Tc (Te ¼ 20 C, DT sc ¼ 0:5Þ:
Fig. 17. Variation of 42 with Tc (Te ¼ 20 C, DT sc ¼ 0:5; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
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Fig. 18. Presentation of conventional cycle P-h diagram for different refrigerants (Tc ¼ 50 C and Te ¼ 20 C).
723
724
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Fig. 19. Presentation of ejector cycle P-h diagram for different refrigerants (Tc ¼ 50 C, DT sc ¼ 0:5, Te ¼ 20 C; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9).
T. Yilmaz, M.T. Erdinç / Energy 168 (2019) 712e727
the increase of the evaporator temperature Te at a given condenser temperature Tc. In Figs. 14 and 15, COP1 and COP2 values are demonstrated as a function of evaporator temperature for different refrigerants at Tc ¼ 50 C . As expected, COP1 and COP2 values increase with the increase of the evaporator temperature Te. It is seen that COP2 values of new refrigeration cycle are always higher than the conventional cycle values (COP1) which shows the benefit of the new refrigeration system. The highest COP1 is obtained for R717, because it has high value of latent heat of evaporation. The highest COP2 values are obtained for R245fa. This is because R245fa has, similar to R1234yf, lower slope of vapor saturation line and lower value of latent heat of evaporation. In Figs. 16 and 17, exergy efficiencies 41 and 42 are demonstrated as a function of condenser temperature for different refrigerants. It is seen that exergy efficiencies of new refrigeration cycle 42 are higher than those for the conventional cycle 41 which shows again the effectiveness of the proposed new refrigeration system. The highest exergy efficiency 41 values are obtained with R717 and the highest 42 values are obtained for R245fa. These results are similar to those values obtained for COP1 and COP2. Exergy efficiencies have an optimum point with respect to condenser temperature Tc.
725
In Figs. 18 and 19, P-h diagrams of conventional system and ejector system are demonstrated for different refrigerants and Tc ¼ 50 C, DT sc ¼ 0:5 and Te ¼ 20 C. One can see from Fig. 19 that the pressure at points 1 is much higher than the pressure at point 7, which is one of the causes for increase of COP2 and 42 . In Fig. 20, the exergy efficiency ratio 421 is demonstrated as a function of DT sc for different refrigerants at Tc ¼ 50 C and Te ¼ 20 C. It is seen that the maximum value is around DT sc ¼ 0:5 similar to Fig. 9 for COP 21. In Fig. 21, the variation of exergy efficiency ratio 421 with Te is presented for different refrigerants at Tc ¼ 50 C and DT sc ¼ 0:5. Like COP 21 , 421 is highest for R1234yf and lowest for R717. 421 increases with decrease of the evaporator temperature, because the same amount of subcooling is more effective at lower evaporator temperatures. Figs. 22 and 23 show exergy destruction of components for conventional and ejector cycle. In conventional system exergy destruction is highest in expansion valve and lowest in evaporator and condenser. In ejector cycle, the exergy destruction in the compressor is higher than the sum of exergy destructions in both expansion valves 1 and 2.
Fig. 20. Variation of 421 with DT sc for different refrigerants (Tc ¼ 50 C, Te ¼ 20 C; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
Fig. 22. Variation of exergy destructions ed of conventional system's different components with Tc (Te ¼ 20 CÞ.
Fig. 21. Variation of 421 with Te (Tc ¼ 50 C, DT sc ¼ 0:5; DT58 ¼ 3 C; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ.
Fig. 23. Variation of exergy destructions ed of new ejector system's different components with Tc (Te ¼ 20 C, DTsc ¼ 0:5; hn ¼ 0:9; hm ¼ 0:8; hd ¼ 0:9Þ:
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5. Conclusıon In this work, a new ejector subcooling refrigeration system without an extra pump or compressor has been modeled and calculated numerically for condenser temperatures between 30 C= 70 C and evaporator temperatures between 20 C=10 C. Seven different refrigerants (namely; R32, R1234yf, R290, R134a, R717, R600a and R245fa) are considered as working fluids. According to the obtained results, the following conclusions can be drawn: Highest COP1 value is obtained with R717 because of high latent heat of evaporation. Highest COP2 value is obtained with R245fa, because of low latent heat of evaporation and lower slope of vapor saturation line. Highest 41 value is obtained with R717 and highest 42 value is obtained with R245fa, similar to COP1 and COP2 values. As condenser temperature increases at constant evaporator temperature, the enthalpy difference h7-h4 decreases. Therefore, the effect of subcooling becomes very important. Because of this, COP21 increases with the increase in condenser temperature. Enthalpy difference h7-h4 decreases with the decrease of the evaporator temperature Te at constant condenser temperature Tc. Because of this, COP21 increases with the decrease in evaporator temperature. The highest value of COP21 ð20%Þ and 421 ð18%Þ are obtained for the refrigerant R1234yf, because it has the lowest latent heat of evaporation and lower slope of vapor saturation line. The lowest value of COP21 ð5%Þ and 421 ð5%Þare obtained for the refrigerant R717, because it has the highest latent heat of evaporation and higher slope of vapor saturation line. Exergy efficiencies 41 and 42 have a maximum value for a certain condenser temperature. COP21 and 421 approach to 1 for DTsc /0 and for DTsc /1. There exists an optimum point of COP21 and 421 around DTsc ¼ 0:5 for all refrigerants. Influence of DT58 on COP21 is not very important. The novel refrigeration system has shown that a considerable increase in coefficient of performance (COP) and exergy efficiency (4) can be achieved with the addition of a valve and ejector which do not significantly increase the investment cost of the refrigeration system. For the validation of the presented results, future experimental study is important. Acknowledgment The authors are very grateful to Prof. Dr. Alper YILMAZ for his valuable comments, which have been utilized to improve the quality of this paper. References [1] Hong WJ, Alhussan K, Zhang H, Garris Jr CA. A novel thermally driven rotorvane/pressure-exchange ejector refrigeration system with environmental benefits and energy efficiency. Energy 2004;29:2331e45. [2] Commission Delegated Regulation (EU) No:1060/2010 of 28 September 2010 Supplementation Directive 201030/EU of the European Parliament and of the Council with regard to energy labeling of household refrigerating appliances. [3] Kosalay I, Ilk HG. Experimental studies and test results on the energy efficiency of household refrigerating appliances. J Sci Ind Res 2017;76:575e80. [4] Qureshi BA, Zubair SM. Mechanical sub-cooling vapor compression systems: current status and future directions. Int. J. Refrigeration 2013;36:2097e110. [5] Zubair SM. Improvement of refrigeration/air-conditioning performance with mechanical sub-cooling. Energy 1990;15:427e33.
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