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Performance optimization of a Lorenz–Meutzner cycle charged with hydrocarbon mixtures for a domestic refrigerator-freezer
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Performance optimization of a LorenzeMeutzner cycle charged with hydrocarbon mixtures for a domestic refrigerator-freezer Won Jae Yoon a, Kookjeong Seo b, Hyun Joon Chung a, Eun-Ji Lee a, Yongchan Kim a,* a b
Department of Mechanical Engineering, Korea University, Anam-Dong, Sungbuk-Ku, Seoul 136-713, Republic of Korea Digital Appliance Division, Samsung Electronics Inc. 416, Meatan-3 Dong, Yeongtong-Ku, Suwon 443-742, Republic of Korea
article info
abstract
Article history:
A LorenzeMeutzner cycle (called the “LM cycle”) for a domestic refrigerator-freezer (RF) has
Received 20 August 2011
energy saving potential because of lower entropy generation in the fresh food compart-
Received in revised form
ment (R)-evaporator and lower compression ratio due to higher mean evaporating
23 September 2011
temperature, compared to a conventional cycle using pure refrigerant. In this study,
Accepted 28 September 2011
a thermodynamic analysis for the optimum compositions of hydrocarbon (HC) mixtures
Available online 6 October 2011
and cycle specifications was performed. In addition, the effects of the refrigerant charge, capillary tube, compressor capacity, and mixture composition on the performance of the
Keywords:
LM cycle using R-290/R-600 were investigated experimentally. Based on the experimental
Hydrocarbon
data, the energy consumption of the optimized LM cycle using R-290/R-600 (40:60%) was
Non-azeotropic mixture
11.2% lower than that of a bypass two-circuit cycle using R-600a in the same RF platform. ª 2011 Elsevier Ltd and IIR. All rights reserved.
Performance Optimization Household refrigerator
Re´frige´rateur-conge´lateur domestique a` cycle LorenzMeutzner, charge´ avec un me´lange d’hydrocarbures : optimisation de la performance Mots cle´s : hydrocarbure ; me´lange non azeotropique ; performance ; optimisation ; re´frige´rateur domestique
1.
Introduction
Various types of refrigeration cycles have been investigated to improve the energy efficiency of domestic refrigeratorfreezers (RFs) because RFs can account for up to 11% of
household energy costs (Bansal et al., 2010). One of the promising cycles is the LorenzeMeutzner cycle (LM cycle) (Lorenz and Meutzner, 1975). The LM cycle uses a zeotropic mixture as a working fluid, which has an appropriate gliding temperature difference (GTD) during evaporation and
* Corresponding author. Tel.: þ82 2 3290 3366; fax: þ82 2 921 5439. E-mail address: [email protected] (Y. Kim). 0140-7007/$ e see front matter ª 2011 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2011.09.014
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Nomenclature A C COP Cp F GTD h HC HTHX LM LTHX _ m P PD Q_ R RF SBS
2
heat transfer area (m ) clearance volume ratio coefficient of performance specific heat at constant pressure (kJ kg1 K1) freezer compartment gliding temperature difference ( C) enthalpy (kJ kg1) hydrocarbon high temperature sub-cooler LorenzeMeutzner low temperature sub-cooler mass flow rate (kg s1) pressure (kPa) displacement rate (m3 s1) heat transfer rate (W) fresh food compartment refrigerator-freezer side-by-side
condensation processes. The LM cycle has two evaporating temperatures for the freezer compartment (F)-evaporator and the fresh food compartment (R)-evaporator, respectively, due to the GTD of zeotropic mixture, even though the F-and R-evaporators are connected serially. Therefore, the average temperature difference between the refrigerant and air in the R-evaporator becomes smaller during the evaporating process, resulting in lower entropy generation than that of conventional refrigerators using pure refrigerants (Simmons et al., 1996). Moreover, the GTD helps to increase the mean refrigerant temperature in the evaporators, resulting in the decrease in the compression ratio. In order to improve the energy efficiency of the RFs, it is required to adopt the LM cycle with optimized system configuration and operating conditions. Compressor
Condenser
SLHX T U v VC _ W
suction line heat exchanger temperature ( C) overall heat transfer coefficient (W m2 K1) specific volume (m3 kg1) volumetric capacity (kJ m3) power input (W)
Subscripts air air side comp compressor cond condenser side evap evaporator side in inlet LMTD log mean temperature difference out outlet ref refrigerant side sat, liquid saturated liquid state sat, vapor saturated vapor state suc suction
In the LM cycle, only the refrigerant temperature at the evaporator inlet has to be lower than the leaving air temperature, so long as the refrigerant temperature at the evaporator outlet is lower than the entering air temperature due to the GTD during the evaporation process. Therefore, it is essential to apply counter flow heat exchangers for the LM cycle (Didion and Bivens, 1990; Stoecker and Walukas, 1981). In addition, the F-evaporator should be located upstream of the R-evaporator so that it can use the lower evaporating temperature region. The performance of the LM cycle can be improved by using sub-coolers. As shown in Fig. 1, two sub-coolers are used in the conventional LM cycle. Both sub-coolers exchange heat between the condenser outlet and evaporator sections, decreasing the two-phase quality of the refrigerant entering the F-evaporator. As shown in Fig. 2, lower quality of the Compressor Condenser
3-way valve Suction pipe F-capillary tube
Suction pipe R-capillary tube
High temperature sub-cooler (HTHX) Capillary Low temperature tube sub-cooler (LTHX)
F-direction
R-direction
F-evaporator
a
R-evaporator
Bypass two-circuit cycle
37
F-evaporator
R-evaporator
b LM cycle
Fig. 1 e Schematic diagrams of (a) the bypass two-circuit cycle and (b) the LM cycle.
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Fig. 2 e Sub-cooling effect of the LM cycle.
refrigerant at a constant refrigerant temperature at the F-evaporator inlet can lead to the increase in the evaporating pressure. Consequently, the compression ratio of the LM cycle can be reduced (Jung and Radermacher, 1991), resulting in higher energy efficiency. In addition, a low temperature subcooler (LTHX) increases the temperature of the refrigerant entering the R-evaporator, which reduces the amount of entropy generated due to a better match of the GTD with the air temperature profile (Rose et al., 1992). A high temperature sub-cooler (HTHX), which is similar to a conventional suction line heat exchanger (SLHX), prevents the flow of the liquid refrigerant into the compressor. The energy saving of the LM cycle was evaluated by several researchers. Lorenz and Meutzner (1975) experimentally observed 20% energy savings compared to R-12 system when an R-22/R-11 mixture was used. Rose et al. (1992) reported 9% energy savings with a mixture of R-22/R-141b compared to an R-12 refrigeration cycle. Based on simulation results, Jung and Radermacher (1991) reported that COP increased significantly up to 15e18% with mixtures of R-22/R-123 and R-32/R-141b. However, most of the existing studies on the LM cycle were performed for halocarbon refrigerant mixtures, which have
ozone depletion potential (ODP) and/or high global warming potential (GWP). Some countries have already taken steps to curtail or prohibit the use of refrigerants with high GWPs, which is also being regulated by the Kyoto Protocol (Wang et al., 2010). Therefore, some natural refrigerants such as hydrocarbons (HCs) are receiving special attention these days. Many home appliance companies, in particular, apply R-600a for their household RFs. Simmons et al. (1996) demonstrated that the compartment temperature can be controlled independently in the LM cycle using HC mixtures with an additional refrigerant loop. However, there are hardly any studies on the performance optimization of the LM cycle using HC mixtures in the literature. This study aims to conduct a thermodynamic analysis for the optimum selection of HC mixtures and to optimize the performance of the LM cycle using R-290/R-600. In this study, the energy savings of the LM cycle using R-290/R-600, R-290/ R-600a, and R-600a/R-600 were evaluated numerically according to the mixture composition. The volumetric capacity and the linearity characteristics of the saturation temperature profile of these mixtures were also studied to improve cycle performance. In addition, the performance of the LM cycle using R-290/R-600 was measured by varying the refrigerant charge, capillary tube length, compressor capacity, and mixture composition.
2.
Experimental setup and test procedure
2.1.
Test apparatus
A household side-by-side (SBS) RF with an internal volume of 0.74 m3 (740 l) was used in the experiments. The SBS RF originally adopts a bypass two-circuit cycle. Table 1 lists the specifications of the original bypass two-circuit cycle. After testing the original bypass two-circuit cycle, the RF was modified to operate under the LM cycle and then its performance was also measured. As shown in Fig. 1(a), in the bypass two-circuit cycle, a 3-way valve to control the refrigerant flow path and two capillary tubes are used for each operation
Table 1 e Specifications of the original bypass two-circuit cycle. Component
Parameter
Specification
Compressor
Type Motor Displacement volume Cooling capacity
R-evaporator
Type Heat transfer area Type Heat transfer area Type Heat transfer area Capillary tube Refrigerant SLHX length Control
Reciprocating, hermetically sealed Inverter-driven BLDC (3600e1600 RPM) 15 cc 171.5 W at AHAM conditions, 1800 RPM 329.1 W at AHAM conditions, 3600 RPM Plate fin-tube, aluminum, 1 column, 11 rows 0.359 m2 Plate fin-tube, aluminum, 2 columns, 15 rows 1.416 m2 Spiral fin-tube, steal, tube length 16.8 m 2.270 m2 Inner diameter 0.85 mm, length 3300 mm (F/R) R-600a, charge amount 96 g 2000 mm (soldered with F/R-capillary tube) Programmable micro-computer
F-evaporator Condenser Cycle
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mode. On the other hand, as shown in Fig. 1(b), in the LM cycle, only one refrigerant flow path and one capillary tube are used without the control valve in the refrigerant path. For the bypass two-circuit cycle, in the R/F simultaneous mode, the refrigerant flows along a serial flow path through the Rcapillary tube and R- and F-evaporators. When the system operates in the freezer-only mode (F-only mode), the refrigerant passes through the F-capillary tube and then enters the F-evaporator directly. For the LM cycle, in the R/F simultaneous mode, the refrigerant flows the same way as it does in the bypass two-circuit cycle. In the F-only mode, R-fan is turned off to reduce the cooling effects in the R-evaporator but the same flow path as used in the R/F simultaneous mode is used. The control logic and suction line assembly of the RF were changed when the bypass two-circuit cycle was modified into the LM cycle. The suction line assembly was installed outside of the RF cabinet with urethane panels so that the bypass twocircuit cycle could be modified easily into the LM cycle. As shown in Fig. 3, the suction line assembly of the bypass twocircuit cycle contained a suction pipe, F- and R-capillary tubes, and connecting component between the F- and R-evaporators, whereas that of the LM cycle included two sub-coolers (HTHX, LTHX). In the bypass two-circuit cycle, F- and R-capillary tubes were soldered to the suction pipe. On the other hand, the capillary tube of the LM cycle was located in the freezer compartment, so additional energy could be saved in the LM cycle (Rose et al., 1992). The HTHX and LTHX were a tube-intube type, which maximized the heat exchange between the warm and cold refrigerants (Simmons et al., 1996). In these subcoolers, the high-pressure refrigerant at the condenser outlet flows through an inside tube and the low-pressure refrigerant flows through an annulus between the tubes in the opposite direction. The lengths of the LTHX and HTHX were 1180 mm and 1400 mm, respectively, while the length of soldered capillary tube of the baseline system was 2000 mm. The condenser and R-evaporator in the original bypass two-circuit cycle used a counter flow type heat exchanger. Therefore, tube arrangement and internal volume were not changed in the LM cycle. However, the heat transfer area of
39
the R-evaporator in the LM cycle had to be enlarged because the mean evaporating temperature of the R-evaporator increased by the GTD of zeotropic mixture and the heat exchange in the LTHX. In this study, the required heat transfer area in the LM cycle, which was estimated by a thermodynamic analysis, was achieved by decreasing the fin pitches of the R-evaporator but not changing the tube arrangement. However, since the F-evaporator in the original bypass twocircuit cycle did not use a counter flow type heat exchanger, the F-evaporator was modified to have spirally-coiled circular fins. Spirally-coiled circular fin and tube heat exchanger easily allow a counter flow arrangement because, in a multi-column heat exchanger, each tube has individual fins. Meanwhile, the RF was controlled by a programmable micro-computer.
2.2.
Test procedure
The performance of the LM cycle was measured according to the refrigerant charge and the capillary tube length at an ambient temperature of 25 C and relative humidity of 70% (ISO 15502, 2005, N-class condition). The refrigerant charge was changed until the minimum energy consumption was observed in each capillary tube length. In addition, the capillary tube length was increased until the final cycle matching was achieved. The ambient condition was controlled by an environmental chamber. During the tests, the time-averaged air temperatures of the F- and R-compartments were maintained approximately at 18 C and 3 C, respectively. The internal air temperature of the F/R cabinet was determined as the average value of those measured by three temperature sensors located deep at the center of the compartment, equally spaced in the vertical direction. The cabinet air temperature was controlled by cyclic operation. In this study, the cut-in (set-on) and cut-out (set-off) temperatures were fixed at a constant value. For all tests, the compressor speed of the RF was maintained at 1800 RPM. This condition was appropriate for the energy saving mode operation of the RF. The performance of the RF was measured under the cyclic steady condition, in which the difference between the time-
Fig. 3 e Suction line assembly for (a) the bypass two-circuit cycle and (b) the LM cycle.
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averaged cabinet air temperatures of adjacent cycles was less than 0.2 C. All internal air temperatures of the cabinet were monitored and recorded every 30 s. The energy consumption of the RF, which indicates the total energy consumed by all components including the compressor and fan motors, was measured by a power meter with an uncertainty of 0.2% and integrated for four cycles. T-type copper-constantan thermocouples were used for temperature measurement. All thermocouples were calibrated to an accuracy of 0.2 C in a constant temperature bath.
3.
Thermodynamic cycle analysis
A thermodynamic analysis was conducted to find the optimum composition of the HC mixtures and the upper limit performance of the LM cycle. Furthermore, the energy saved by the GTD and the sub-cooling effects was also evaluated. In this study, binary mixtures of R-290/R-600, R-290/R-600a, and R-600a/R-600 were considered as the working fluid. The following assumptions were applied to the thermodynamic analysis: 1) pressure drops in the evaporator and condenser are neglected, 2) the ratio between the refrigerating effects of the F- and R- evaporators is 60:40, 3) the condenser outlet is in the saturated liquid state, 4) the R-evaporator outlet is in the saturated vapor state, 5) the UA of the F-evaporator is constant by practical considerations of space and cost (Smith et al., 1990), and 6) the UA of the condenser is proportional to the volumetric capacity of the refrigeration system with the mixtures. The condenser can be enlarged easily because it can be installed partially in the cabinet wall and operated under the no-frosting condition. The UA values of the F-evaporator and condenser were determined based on the baseline system of R-600a. The energy balance without latent cooling for a heat exchanger is given by Eq. (1). _ ref Dhref ¼ m _ air CP;air Tair;in Tair;out Q_ ¼ UADTLMTD ¼ m
DT1 DT2 lnðDT1 =DT2 Þ
(3)
DT1 ¼ Tair;in Tref;out ; DT2 ¼ Tair;out Tref;in
(4)
In Eq. (2), C is the clearance volume ratio and PD is the displacement rate of the compressor. ΔTLMTD of the F-evaporator was determined by the heat transfer rate because the UA of the F-evaporator was regarded as constant by practical considerations. The COP and volumetric capacity of the LM cycle are given by Eqs. (5) and (6), respectively. COP ¼ Q_ F þ Q_ R _ comp W VC ¼
(5)
hsat;vapor hsat;liquid vsuc
(6)
In the volumetric capacity calculation, the mean evaporating temperature and suction temperature were assumed to be 25 C and 25 C, respectively. Table 2 shows the input conditions for the thermodynamic analysis. The size of the LTHX was assumed to be 20% of the total evaporator UA. Jung and Radermacher (1991) reported that the maximum size of the LTHX was limited to roughly 20% of the total evaporator UA. Fig. 4 shows the flow chart of the simulation. The condensing and evaporating pressures were determined by the energy balance in the F-evaporator and condenser given in Eq. (1). After the condensing and evaporating pressures were determined, the required heat transfer area of the R-evaporator was calculated for each case.
4.
Results and discussion
4.1. Optimization of mixture composition using simulation results
(1)
_ ref and DTLMTD are calculated by Eqs. (2) and (3), In Eq. (1), m respectively. 3 1n P cond 5 PD ¼ 41 þ C C Pevap vsuc 2
_ ref m
DTLMTD ¼
(2)
The COP and volumetric capacity of the LM cycle using R-290/ R-600, R-290/R-600a, and R-600a/R-600 were analyzed to determine the optimum mixture composition. Fig. 5 shows the variations of the energy savings for the LM cycle with the HC mixtures according to the mixture composition. The energy saving potential was estimated by a percent decrease in the energy consumption of the LM cycle as compared with
Table 2 e Input conditions for cycle simulation. Parameter Inlet air temperature of the R-evaporator Inlet air temperature of the F-evaporator Air flow rate of the condenser Air flow rate of the F-evaporator Air flow rate of the R-evaporator UA of the LTHX Suction temperature Isentropic efficiency Clearance volume ratio
Specified value
Technical ground for the value
3 C 18 C 1.5 m3 min1 0.7 m3 min1 0.6 m3 min1 20% of total evaporator 25 C 70% 0.02
Internal cabinet air temperature Internal cabinet air temperature Measured data for the baseline system Measured data for the baseline system Measured data for the baseline system Reference data (Jung and Radermacher, 1991) Anti-sweat condition for suction Estimated value Estimated value
41
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Fig. 4 e Flow chart of the LM cycle simulation.
that of the bypass two-circuit cycle using R-600a (the baseline system). The R-290/R-600 showed the highest energy saving of 15.3% at the R-290 mass fraction of 0.4 mainly because it had the most suitable GTD among the mixtures. As mentioned earlier, the energy saving of the LM cycle increases with a decrease in the compression ratio due to higher mean evaporating temperature. Therefore, the efficiency of the LM cycle would be most improved when the refrigerant temperature profile closely matches the air temperature profile in the F-evaporator. In the lower GTD case, in which the GTD in the F-evaporator is smaller than the variation of air temperature through the F-evaporator, the increase in the mean evaporating temperature would be very small. However, an
excessive GTD can lead to the reduction of the cooling capacity at constant UA. The variation of air temperature through the F-evaporator was about 8 C at the air flow rate and cooling capacity of the baseline system, while the GTDs in the F-evaporator of R-290/R-600, R-290/R-600a, and R-600a/R600 were 8.6, 4.6, and 0.6 C, respectively, at the optimum composition of each mixture. The GTDs of R-290/R-600a and R-600a/R-600 were obviously relatively lower for domestic RFs. Therefore, R-290/R-600 was selected as the optimum binary mixture for the LM cycle. Fig. 6 represents the variation of volumetric capacity in the LM cycle according to the mixture composition when the evaporating temperature and suction temperature are 25 C and 25 C, respectively. The displacement rate (PD) of the compressor needs to be adjusted for each mixture composition so that the LM cycle can maintain a constant refrigeration capacity (Stoecker and Walukas, 1981). An excessive displacement volume of the compressor may increase cycling losses because the compressor would then have to turn on and off more frequently. In addition, it may lead to higher mean temperature differences between the refrigerant and air in the heat exchangers. The volumetric capacity of the LM cycle using R-290/R-600 with an R-290 mass fraction of 0.2 was similar to that using R-600a. When the baseline compressor for R-600a is directly used in the LM cycle, R-290/R-600 with the R-290 mass fraction of 0.2 would be the most recommendable mixture composition for the LM cycle. However, as mentioned in Fig. 5, the maximum energy saving for R-290/ R-600 was observed at the R-290 mass fraction of 0.4. The energy saving for R-290/R-600 at the R-290 mass fraction of 0.2 was estimated at 14.1%, which was 1.2% lower than the maximum value at the R-290 mass fraction of 0.4. Therefore, R-290/R-600 with the R-290 mass fraction of 0.4 is recommended to maximize the energy saving of the LM cycle. In this case, a down-sized compressor from the baseline system has to be adopted to output the same capacity as the baseline system. Fig. 7 represents the linearity characteristics in saturation temperature profile of R-290/R-600 at the mean evaporating temperature of 25 C. The normalized temperature is defined
1200 -3
15
Volumetric capacity kJ m
Energy saving potential (%)
20
10 5 0 -5 -10 0.0
R-290/R-600 R-290/R-600a R-600a/R-600 0.2
0.4
0.6
0.8
Mass fraction of lower NBP refrigerant Fig. 5 e Variation of the energy saving potential with mixture composition.
1.0
1000
R-290/R-600 R-290/R-600a R-600a/R-600
800
600
Baseline: R-600a
400
200 0.0
0.2
0.4
0.6
0.8
1.0
Mass fraction of lower NBP refrigerant Fig. 6 e Variation of the volumetric capacity with mixture composition.
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20
1.0 Energy saving potential (%)
Normalized temperature T-Tsat.liquid /GTD
42
0.8 0.6 Mass fraction of R-290 o 0.1 GTD=7.6 C o 0.2 GTD=11.9 C o 0.3 GTD=14.3 C o 0.4 GTD=15.4 C
0.4 0.2 0.0
0.0
0.2
0.4
0.6
0.8
10 5 0 -5 -10 0.0
1.0
Case 1 Case 2 Case 3 Case 4
15
0.2
0.4
Vapor quality
as (TTsat, liquid)/GTD. The linearity of the refrigerant temperature with respect to vapor quality increased with the increase in the R-290 mass fraction. Higher linearity can lead to larger energy saving than that of the baseline system because it impedes refrigerant temperature rise and helps to remove the pinch point in the F-evaporator. As the R-290 mass fraction increased from 0.2 to 0.4, the mean evaporating temperature of the LM cycle increased by 1.6 C, which increased the energy saving. However, as the R-290 mass fraction increased beyond 0.4, the excessive volumetric capacity of R-290/R-600 increased ΔTLMTD in the F-evaporator, which reduced the energy saving. The GTD of R-290/R-600 showed a maximum value at the R-290 mass fraction of 0.5. Cycle simulations for four cases were carried out to investigate the energy saving of the LM cycle using R-290/ R-600 according to individual effect of the GTD and subcooling. Case 1 is the baseline for illustrating each effect. The mean evaporating temperature was fixed at 25 C, and the sub-cooling effect by the LTHX was not included. In Case 2, the GTD effect for higher mean evaporating temperature was considered. In Cases 3 and 4, both the GTD and the sub-cooling effect by the LTHX were taken into account. However, for Cases 3 and 4, the size of the LTHX was assumed to be 10% and 20% of the total evaporator UA, respectively. Fig. 8 shows the variation of the energy saving with mixture composition for each case. In Case 1, the energy saving was only 2.5% at the R-290 mass fraction of 0.2. In Case 2, an additional energy saving of 5.7% can be obtained by including the GTD effect with an increase in the mean evaporating temperature of 2.6 C. The energy savings for Cases 3 and 4 increased by 3.5% and 3.7% over Cases 2 and 3, respectively, due to the subcooling effect of the LTHX. In Cases 2e4, the optimum mass fraction of R-290 was 0.4 with respect to the energy saving. The mean evaporating temperatures for Cases 3 and 4 increased by 0.7 C in each case due to the sub-cooling effect of the LTHX. As mentioned earlier, the heat transfer area of the R-evaporator in the LM cycle should be increased because of the increase in the mean evaporating temperature of the
0.8
1.0
Fig. 8 e Variation of the energy saving potential with mixture composition for each case.
R-evaporator by the GTD and heat exchange in the LTHX. Fig. 9 represents the required heat transfer area of the R-evaporator in the LM cycle using R-290/R-600, compared with that of the baseline system. The results were obtained at a constant compressor capacity. As the LTHX size was increased, larger heat transfer area was required due to the higher evaporating temperature in the R-evaporator. In Case 4, the required heat transfer area of the R-evaporator at the R-290 mass fraction of 0.4 was estimated to be about 3 times that of the baseline system. However, when the compressor is down-sized to use the R-290 mass faction of 0.4, the heat transfer area needed to be only 2.2 times that of the baseline system. The volumetric capacity of R-290/R-600 at the R-290 mass faction of 0.4 was approximately 38% larger than that of R-600a, so the compressor size and required heat transfer area decreased for the same refrigeration capacity. Fig. 10 shows the R-evaporator for each cycle applied in the present experiments. The volumes occupied by each R-evaporator were very similar because a compact R-evaporator with smaller fin pitches was
5 Required area / Baseline area
Fig. 7 e Linearity characteristics of saturation temperature profiles of R-290/R-600.
0.6
Mass fraction of R-290
4
Case 2 Case 3 Case 4
3
2
1
0 0.0
0.2
0.4
0.6
0.8
1.0
Mass fraction of R-290 Fig. 9 e Ratio of the required heat transfer area of Revaporator in the LM cycle to that in the baseline system.
43
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Fig. 10 e R-evaporator for (a) the bypass two-circuit cycle and (b) the LM cycle.
4.2.
System optimization using experimental results
The performance of the LM cycle using R-290/R-600 was measured by varying the refrigerant charge amount, capillary tube length, mixture composition, and compressor capacity. Both a baseline reciprocating compressor and a down-sized reciprocating compressor were adopted for the LM cycle using R-290/R-600. The baseline compressor originally designed for R-600a had a displacement volume of 15 cc. The down-sized compressor for R-290/R-600 had a displacement volume of 11 cc, which was properly designed to have the same capacity at the R-290 mass faction of 0.4 as that of the baseline system. Both compressors had almost the same efficiency at 1800 RPM under the test conditions used in this study. Fig. 11 represents the effects of the refrigerant charge, R-290 mass fraction, and capillary tube length on the energy consumption of the LM cycle with the baseline compressor. The experiments were performed for the R-290 mass fractions of 0.2 and 0.3 because the volumetric capacity of the LM cycle for the R-290 mass fraction of 0.1 was considerably lower than the designed value. Optimum refrigerant charge in each capillary tube length was determined by the minimum energy consumption for each case. For the R-290 mass fraction of 0.2, the optimum refrigerant charge increased from 125 to 160 g as the capillary tube length increased from 13 to 17 m. This trend was very similar to that for the R-290 mass fraction of 0.3, but the range of the optimum refrigerant charge for the R-290 mass faction of 0.3 was lower than that for the R-290 mass fraction of 0.2 due to higher volumetric capacity of the LM cycle. As the capillary tube length increased at a fixed refrigerant charge, both the refrigerant flow rate through the LM
cycle and the evaporating temperature decreased, so the cooling capacity in the R-evaporator reduced. As the heat transfer rate in the F-evaporator increased with the decrease in the evaporating temperature, the cooling capacity of the R-evaporator decreased due to the limited remaining capacity in the R-evaporator during the R/F simultaneous operation. The decrease of the cooling capacity in the R-evaporator due to the longer capillary tube has to be compensated by the increase in the refrigerant charge. In addition, the minimum energy consumption for the R-290 mass fraction of 0.2 was 1.6% lower than that for the R-290 mass fraction of 0.3 because of excessive volumetric capacity for the baseline compressor in the R-290 mass fraction of 0.3. Therefore, the lowest energy consumption of the LM cycle with the baseline compressor was observed in the R-290 mass faction of 0.2 at the capillary tube length of 15 m and refrigerant charge of 140 g. As shown in Fig. 11, the optimum capillary tube length of 15 m in the LM cycle with the baseline compressor was significantly longer than the optimum value of 3.3 m in the bypass two-circuit cycle (Table 1) because of the larger sub-
34 Energy consumption (kWh/month)
applied for the LM cycle. A forced convection defrosting can be a solution for the frost accumulation in the compact R-evaporator. The forced convection defrosting was performed by using internal air circulation through the R-evaporator when the refrigerant did not pass through the R-evaporator. In addition, a counter flow and spirally-coiled circular fin type F-evaporator for the LM cycle was designed to have almost the same heat transfer area as that of the baseline system.
R-290 mass fraction 0.2 0.3
33
Capillary tube length 32
13 m
15 m
17 m
19 m
17 m
31
15 m 30 110
120
130
140
150
160
170
Optimum refrigerant charge (g) Fig. 11 e Variation of the energy consumption with the refrigerant charge, R-290 mass fraction, and capillary tube length for the baseline compressor.
44
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Table 3 e Performances of the LM cycle and the bypass two-circuit cycle with the baseline compressor for R-600a. Parameter
Bypass two-circuit cycle
R/F-operation
F-operation
R/F-operation
F-operation
16.6% 77.7 W 46.2 C 28.6 C 24.0 C 1.6 C 382.6 kPa 75.4 kPa
39.8% 68.2 W 46.4 C 28.0 C 31.3 C 13.0 C 412.4 kPa 56.1 kPa
23.5% 74.4 W 46.3 C 30.8 C 19.1 C 19.3 C 413.7 kPa 71.9 kPa
43.1% 65.3 W 46.0 C 30.2 C 27.7 C 27.9 C 406.8 kPa 50.0 kPa
30.9 kWh/month
Fig. 13 shows the effects of the refrigerant charge, R-290 mass fraction, and capillary tube length on the energy consumption of the LM cycle with the down-sized compressor. As shown in Figs. 11 and 13, the refrigerant charge and capillary tube length had similar effects on the energy consumption for both compressors. For all R-290 mass
10 5
F-evaporator
R-evaporator
-5
Temperature
Connecting component
Air
o
C
0
-10 -15 Air
-20 -25
Refrigerant
Refrigerant -30 -35
R-evap. in R-evap. out F-evap. in F-evap. out Location of the cycle path
a
Bypass two-circuit cycle
10 5
LTHX
F-evaporator
R-evaporator
C
0 Air
-5
o
cooling of the LM cycle. Due to the heat transfer in the two sub-coolers of the LM cycle, the sub-cooling at the capillary tube inlet was large, as much as 40 C, while that of the conventional cycle was about 0 C. Fatouh (2007) reported that larger sub-cooling in the LM cycle can lead to a serious increase in the capillary tube length. In addition, the optimum refrigerant charge in the LM cycle was 46% higher than that in the bypass two-circuit cycle (Table 1) because of the condenser liquid lines in the two sub-coolers. Table 3 summarizes the performance of the LM cycle and the bypass two-circuit cycle for the same baseline compressor at the optimized refrigerant charge and capillary tube length. The energy consumption of the optimized LM cycle with the baseline compressor was 10.1% lower than that of the bypass two-circuit cycle, so it was 4% lower than the energy saving predicted by the simulation. This difference in the energy saving may be attributed to the increase in cycling losses of the LM cycle due to larger refrigerant charge and more frequent on/off cycling operations of the compressor. The increased refrigerant charge can cause an increase in the migration loss at compressor shut-off and redistribution loss at compressor start-up (Bjo¨rk and Palm, 2006). The compressor was turned on and off more frequently in the LM cycle due to the lower operating ratio resulting from larger cooling capacity in the LM cycle, as shown in Table 3. Fig. 12 represents the temperature distributions of the refrigerant and air through F- and R-evaporators for the bypass two-circuit cycle and the optimized LM cycle with the baseline compressor. All temperatures in the operating period are time-averaged values. In the LM cycle, the temperature profile of the refrigerant through the F-evaporator was almost parallel with the air temperature profile, so a higher energy saving was expected. The mean evaporating temperature in the LM cycle was approximately 3 C higher than that in the bypass two-circuit cycle, and the temperature of the air leaving the F-evaporator in the LM cycle was 1.9 C lower than that in the bypass two-circuit cycle, resulting in approximately 24% larger cooling capacity of the F-evaporator in the LM cycle. Based on the simulation results, the cooling capacity and compressor power input of the LM cycle with the baseline compressor were 23% and 8% higher, respectively, than those of the bypass two-circuit cycle. In addition, the ΔTLMTD in the R-evaporator of the LM cycle was significantly less than that of the bypass two-circuit cycle, allowing lower entropy generation.
34.4 kWh/month
Temperature
Operating ratio Average power input Discharge temperature Condenser outlet temperature Evaporator inlet temperature Evaporator outlet temperature Condensing pressure Evaporating pressure Energy consumption
LM cycle
-10 -15 -20
Air Refrigerant
-25 -30 -35
Refrigerant F-evap. in
F-evap. out R-evap. in R-evap. out Location of the cycle path
b
LM cycle
Fig. 12 e Temperature distributions of refrigerant and air through F- and R-evaporators for (a) the bypass two-circuit cycle and (b) the LM cycle.
45
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32
-12 o
Capillary tube length 20 m
31
22 m 24 m
18 m 18 m
22 m
20 m 22 m
20 m 30 110
120
130
Predicted data F-evap. inlet F-evap. outlet R-evap. inlet
C
R-290 mass fraction 0.35 0.40 0.45
Evaporating temperature
Energy consumpation (kWh/month)
33
-16
-20
-24
-28 Measured data
140
150
160
-32 0.0
170
Optimum refrigerant charge (g) Fig. 13 e Variation of the energy consumption with the refrigerant charge, R-290 mass fraction, and capillary tube length for the down-sized compressor.
factions, the optimum refrigerant charge increased with the increase in the capillary tube length. In addition, the range of the optimum refrigerant charge decreased with the increase in the R-290 mass fraction because of the increase in the volumetric capacity. The lowest energy consumption of the LM cycle with the down-sized compressor was observed for the R-290 mass faction of 0.4 at the capillary tube length of 20 m and refrigerant charge of 135 g. The best performance of the LM cycle for the R-290 mass fraction of 0.4 was due to proper design of the compressor capacity at the given mixture composition. The energy consumption of the optimized LM cycle with the down-sized compressor was 1.1% lower than that with the baseline compressor because of better linearity characteristics and larger GTD of R-290/R-600 for the R-290 mass fraction of 0.4. These results are consistent with the predictions from the simulations. The optimum capillary tube length (20 m) for the down-sized compressor was longer than that (15 m) for the baseline compressor. Generally, the pressure and sub-cooling at the capillary inlet are dominant factors for determining the mass flow rate at a given capillary tube geometry. Based on the results of Fatouh (2007), the optimum capillary tube length for the HC mixtures is directly proportional to the pressure at the capillary inlet. As the R-290 mass fraction increased from 0.2 to 0.4, the pressure at the capillary inlet was increased by approximately 30%, resulting in the increase in the capillary tube length. However, the variation of the sub-cooling with the increase in the R-290 mass fraction of 0.2e0.4 was relatively small. Fig. 14 compares the measured evaporating temperatures with simulation predictions. As the R-290 mass fraction increased from 0.2 to 0.4, the evaporating temperatures at each position averagely increased, as predicted in the simulation. The increase in the evaporating temperature allows higher energy saving of the LM cycle. The temperature difference of the refrigerant through the LTHX increased from 1.9 to 3.7 C with the increase in the R-290 mass fraction from 0.2 to 0.4, while that through the F-evaporator slightly decreased by 0.4 C due to the linearity characteristics. As shown in Fig. 7, the
0.2
0.4
0.6
0.8
1.0
Mass fraction of R-290 Fig. 14 e Comparison of the measured evaporating temperatures with simulation predictions.
evaporating temperature of R-290/R-600 for the R-290 mass fraction of 0.2 increased more rapidly than that for the R-290 mass fraction of 0.4 in the lower quality region. This rapid increase led to a relatively large evaporating temperature increase in the F-evaporator, which operates in the lower quality region, and small increase in the LTHX. The measured temperature differences through the LTHX were similar to the simulation results obtained for the optimized LTHX capacity.
5.
Conclusions
The performance of the LM cycle was measured and analyzed by varying the refrigerant charge, capillary tube length, compressor capacity, and mixture composition, and then compared with that of the bypass two-circuit cycle. From the simulation, the energy consumption of the LM cycle using R290/R-600 (40:60%) was 15.3% lower than that of the bypass two-circuit cycle using R-600a. However, for the optimum R290 mass faction of 0.4, the baseline compressor had to be down-sized by approximately 38% to allow proper refrigeration capacity in the LM cycle. In this case, the heat transfer area of the R-evaporator should be increased to be 120% greater than that of the bypass two-circuit cycle using R-600a. Based on the experimental results, the optimum refrigerant charge for the LM cycle using R-290/R-600 increased with the increase in the capillary tube length because of the decrease in the refrigerant flow rate and evaporating temperature. The optimum refrigerant charge and capillary tube length of the LM cycle using R-290/R-600 were significantly larger than those of the bypass two-circuit cycle using R-600a due to the condenser liquid lines in the two sub-coolers and larger subcooling, respectively. The energy consumption of the LM cycle using R-290/R-600 (40:60%) with the down-sized compressor was reduced to be 11.2% lower than that of the bypass two-circuit cycle using R-600a with the baseline compressor. In addition, the energy consumption of the LM cycle using R-290/R-600 (20:80%) with the baseline compressor was reduced to be 10.1% lower than that of the bypass twocircuit cycle using R-600a.
46
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Acknowledgments This work was supported by the Korea Institute of Energy Technology Evaluation and Planning (KETEP) grant funded by the Korea government Ministry of Knowledge Economy (No. 20103010080011) and Korea University.
references
Bansal, P., Fothergill, D., Fernandes, R., 2010. Thermal analysis of the defrost cycle in a domestic freezer. Int. J. Refrig. 33, 589e599. Bjo¨rk, E., Palm, B., 2006. Refrigerant mass charge distribution in a domestic refrigerator. Part _: transient conditions. Appl. Therm. Eng. 26, 829e837. Didion, D.A., Bivens, D.B., 1990. Role of refrigerant mixtures as alternatives to CFCs. Int. J. Refrig. 13, 163e175. Fatouh, M., 2007. Theoretical investigation of adiabatic capillary tubes working with propane/n-butane/iso-butane blends. Energy Convers. Manage. 48, 1338e1348.
International standard, ISO 15502, 2005. Household Refrigerating Appliance-characteristics and Test Methods. Jung, D.S., Radermacher, R., 1991. Performance simulation of a two-evaporator refrigerator-freezer charged with pure and mixed refrigerants. Int. J. Refrig. 14, 254e263. Lorenz, A., Meutzner, K., 1975. On Application of Non-azeotropic Two Component Refrigerants in Domestic Refrigerators and Home Freezers. XIV Int. Congress Refrigeration, Moscow. IIR, Paris. Rose, R.J., Jung, D., Radermacher, R., 1992. Testing of domestic two-evaporator refrigerators with zeotropic refrigerant mixtures. ASHRAE Trans. 98 (2), 216e226. Simmons, K.E., Haider, I., Radermacher, R., 1996. Independent compartment temperature control of Lorenz-Meutzner and modified Lorenz-Meutzner cycle refrigerators. ASHRAE Trans. 102 (1), 1085e1092. Smith, M.K., Heun, M.C., Crawford, R.R., Newell, T.A., 1990. Thermodynamic performance limit considerations for dualevaporator, non-azeotropic refrigerant mixture-based domestic refrigerator-freezer systems. Int. J. Refrig. 13, 237e242. Stoecker, W.F., Walukas, D.J., 1981. Conserving energy in domestic refrigerators through the use of refrigerant mixtures. Int. J. Refrig. 4, 201e208. Wang, K., Eisele, M., Hwang, Y., Radermacher, R., 2010. Review of secondary loop refrigerant systems. Int. J. Refrig. 33, 212e234.
Available online at www.sciencedirect.com
w w w . i i fi i r . o r g
journal homepage: www.elsevier.com/locate/ijrefrig
Performance optimization of a LorenzeMeutzner cycle charged with hydrocarbon mixtures for a domestic refrigerator-freezer Won Jae Yoon a, Kookjeong Seo b, Hyun Joon Chung a, Eun-Ji Lee a, Yongchan Kim a,* a b
Department of Mechanical Engineering, Korea University, Anam-Dong, Sungbuk-Ku, Seoul 136-713, Republic of Korea Digital Appliance Division, Samsung Electronics Inc. 416, Meatan-3 Dong, Yeongtong-Ku, Suwon 443-742, Republic of Korea
article info
abstract
Article history:
A LorenzeMeutzner cycle (called the “LM cycle”) for a domestic refrigerator-freezer (RF) has
Received 20 August 2011
energy saving potential because of lower entropy generation in the fresh food compart-
Received in revised form
ment (R)-evaporator and lower compression ratio due to higher mean evaporating
23 September 2011
temperature, compared to a conventional cycle using pure refrigerant. In this study,
Accepted 28 September 2011
a thermodynamic analysis for the optimum compositions of hydrocarbon (HC) mixtures
Available online 6 October 2011
and cycle specifications was performed. In addition, the effects of the refrigerant charge, capillary tube, compressor capacity, and mixture composition on the performance of the
Keywords:
LM cycle using R-290/R-600 were investigated experimentally. Based on the experimental
Hydrocarbon
data, the energy consumption of the optimized LM cycle using R-290/R-600 (40:60%) was
Non-azeotropic mixture
11.2% lower than that of a bypass two-circuit cycle using R-600a in the same RF platform. ª 2011 Elsevier Ltd and IIR. All rights reserved.
Performance Optimization Household refrigerator
Re´frige´rateur-conge´lateur domestique a` cycle LorenzMeutzner, charge´ avec un me´lange d’hydrocarbures : optimisation de la performance Mots cle´s : hydrocarbure ; me´lange non azeotropique ; performance ; optimisation ; re´frige´rateur domestique
1.
Introduction
Various types of refrigeration cycles have been investigated to improve the energy efficiency of domestic refrigeratorfreezers (RFs) because RFs can account for up to 11% of
household energy costs (Bansal et al., 2010). One of the promising cycles is the LorenzeMeutzner cycle (LM cycle) (Lorenz and Meutzner, 1975). The LM cycle uses a zeotropic mixture as a working fluid, which has an appropriate gliding temperature difference (GTD) during evaporation and
* Corresponding author. Tel.: þ82 2 3290 3366; fax: þ82 2 921 5439. E-mail address: [email protected] (Y. Kim). 0140-7007/$ e see front matter ª 2011 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2011.09.014
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Nomenclature A C COP Cp F GTD h HC HTHX LM LTHX _ m P PD Q_ R RF SBS
2
heat transfer area (m ) clearance volume ratio coefficient of performance specific heat at constant pressure (kJ kg1 K1) freezer compartment gliding temperature difference ( C) enthalpy (kJ kg1) hydrocarbon high temperature sub-cooler LorenzeMeutzner low temperature sub-cooler mass flow rate (kg s1) pressure (kPa) displacement rate (m3 s1) heat transfer rate (W) fresh food compartment refrigerator-freezer side-by-side
condensation processes. The LM cycle has two evaporating temperatures for the freezer compartment (F)-evaporator and the fresh food compartment (R)-evaporator, respectively, due to the GTD of zeotropic mixture, even though the F-and R-evaporators are connected serially. Therefore, the average temperature difference between the refrigerant and air in the R-evaporator becomes smaller during the evaporating process, resulting in lower entropy generation than that of conventional refrigerators using pure refrigerants (Simmons et al., 1996). Moreover, the GTD helps to increase the mean refrigerant temperature in the evaporators, resulting in the decrease in the compression ratio. In order to improve the energy efficiency of the RFs, it is required to adopt the LM cycle with optimized system configuration and operating conditions. Compressor
Condenser
SLHX T U v VC _ W
suction line heat exchanger temperature ( C) overall heat transfer coefficient (W m2 K1) specific volume (m3 kg1) volumetric capacity (kJ m3) power input (W)
Subscripts air air side comp compressor cond condenser side evap evaporator side in inlet LMTD log mean temperature difference out outlet ref refrigerant side sat, liquid saturated liquid state sat, vapor saturated vapor state suc suction
In the LM cycle, only the refrigerant temperature at the evaporator inlet has to be lower than the leaving air temperature, so long as the refrigerant temperature at the evaporator outlet is lower than the entering air temperature due to the GTD during the evaporation process. Therefore, it is essential to apply counter flow heat exchangers for the LM cycle (Didion and Bivens, 1990; Stoecker and Walukas, 1981). In addition, the F-evaporator should be located upstream of the R-evaporator so that it can use the lower evaporating temperature region. The performance of the LM cycle can be improved by using sub-coolers. As shown in Fig. 1, two sub-coolers are used in the conventional LM cycle. Both sub-coolers exchange heat between the condenser outlet and evaporator sections, decreasing the two-phase quality of the refrigerant entering the F-evaporator. As shown in Fig. 2, lower quality of the Compressor Condenser
3-way valve Suction pipe F-capillary tube
Suction pipe R-capillary tube
High temperature sub-cooler (HTHX) Capillary Low temperature tube sub-cooler (LTHX)
F-direction
R-direction
F-evaporator
a
R-evaporator
Bypass two-circuit cycle
37
F-evaporator
R-evaporator
b LM cycle
Fig. 1 e Schematic diagrams of (a) the bypass two-circuit cycle and (b) the LM cycle.
38
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Fig. 2 e Sub-cooling effect of the LM cycle.
refrigerant at a constant refrigerant temperature at the F-evaporator inlet can lead to the increase in the evaporating pressure. Consequently, the compression ratio of the LM cycle can be reduced (Jung and Radermacher, 1991), resulting in higher energy efficiency. In addition, a low temperature subcooler (LTHX) increases the temperature of the refrigerant entering the R-evaporator, which reduces the amount of entropy generated due to a better match of the GTD with the air temperature profile (Rose et al., 1992). A high temperature sub-cooler (HTHX), which is similar to a conventional suction line heat exchanger (SLHX), prevents the flow of the liquid refrigerant into the compressor. The energy saving of the LM cycle was evaluated by several researchers. Lorenz and Meutzner (1975) experimentally observed 20% energy savings compared to R-12 system when an R-22/R-11 mixture was used. Rose et al. (1992) reported 9% energy savings with a mixture of R-22/R-141b compared to an R-12 refrigeration cycle. Based on simulation results, Jung and Radermacher (1991) reported that COP increased significantly up to 15e18% with mixtures of R-22/R-123 and R-32/R-141b. However, most of the existing studies on the LM cycle were performed for halocarbon refrigerant mixtures, which have
ozone depletion potential (ODP) and/or high global warming potential (GWP). Some countries have already taken steps to curtail or prohibit the use of refrigerants with high GWPs, which is also being regulated by the Kyoto Protocol (Wang et al., 2010). Therefore, some natural refrigerants such as hydrocarbons (HCs) are receiving special attention these days. Many home appliance companies, in particular, apply R-600a for their household RFs. Simmons et al. (1996) demonstrated that the compartment temperature can be controlled independently in the LM cycle using HC mixtures with an additional refrigerant loop. However, there are hardly any studies on the performance optimization of the LM cycle using HC mixtures in the literature. This study aims to conduct a thermodynamic analysis for the optimum selection of HC mixtures and to optimize the performance of the LM cycle using R-290/R-600. In this study, the energy savings of the LM cycle using R-290/R-600, R-290/ R-600a, and R-600a/R-600 were evaluated numerically according to the mixture composition. The volumetric capacity and the linearity characteristics of the saturation temperature profile of these mixtures were also studied to improve cycle performance. In addition, the performance of the LM cycle using R-290/R-600 was measured by varying the refrigerant charge, capillary tube length, compressor capacity, and mixture composition.
2.
Experimental setup and test procedure
2.1.
Test apparatus
A household side-by-side (SBS) RF with an internal volume of 0.74 m3 (740 l) was used in the experiments. The SBS RF originally adopts a bypass two-circuit cycle. Table 1 lists the specifications of the original bypass two-circuit cycle. After testing the original bypass two-circuit cycle, the RF was modified to operate under the LM cycle and then its performance was also measured. As shown in Fig. 1(a), in the bypass two-circuit cycle, a 3-way valve to control the refrigerant flow path and two capillary tubes are used for each operation
Table 1 e Specifications of the original bypass two-circuit cycle. Component
Parameter
Specification
Compressor
Type Motor Displacement volume Cooling capacity
R-evaporator
Type Heat transfer area Type Heat transfer area Type Heat transfer area Capillary tube Refrigerant SLHX length Control
Reciprocating, hermetically sealed Inverter-driven BLDC (3600e1600 RPM) 15 cc 171.5 W at AHAM conditions, 1800 RPM 329.1 W at AHAM conditions, 3600 RPM Plate fin-tube, aluminum, 1 column, 11 rows 0.359 m2 Plate fin-tube, aluminum, 2 columns, 15 rows 1.416 m2 Spiral fin-tube, steal, tube length 16.8 m 2.270 m2 Inner diameter 0.85 mm, length 3300 mm (F/R) R-600a, charge amount 96 g 2000 mm (soldered with F/R-capillary tube) Programmable micro-computer
F-evaporator Condenser Cycle
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mode. On the other hand, as shown in Fig. 1(b), in the LM cycle, only one refrigerant flow path and one capillary tube are used without the control valve in the refrigerant path. For the bypass two-circuit cycle, in the R/F simultaneous mode, the refrigerant flows along a serial flow path through the Rcapillary tube and R- and F-evaporators. When the system operates in the freezer-only mode (F-only mode), the refrigerant passes through the F-capillary tube and then enters the F-evaporator directly. For the LM cycle, in the R/F simultaneous mode, the refrigerant flows the same way as it does in the bypass two-circuit cycle. In the F-only mode, R-fan is turned off to reduce the cooling effects in the R-evaporator but the same flow path as used in the R/F simultaneous mode is used. The control logic and suction line assembly of the RF were changed when the bypass two-circuit cycle was modified into the LM cycle. The suction line assembly was installed outside of the RF cabinet with urethane panels so that the bypass twocircuit cycle could be modified easily into the LM cycle. As shown in Fig. 3, the suction line assembly of the bypass twocircuit cycle contained a suction pipe, F- and R-capillary tubes, and connecting component between the F- and R-evaporators, whereas that of the LM cycle included two sub-coolers (HTHX, LTHX). In the bypass two-circuit cycle, F- and R-capillary tubes were soldered to the suction pipe. On the other hand, the capillary tube of the LM cycle was located in the freezer compartment, so additional energy could be saved in the LM cycle (Rose et al., 1992). The HTHX and LTHX were a tube-intube type, which maximized the heat exchange between the warm and cold refrigerants (Simmons et al., 1996). In these subcoolers, the high-pressure refrigerant at the condenser outlet flows through an inside tube and the low-pressure refrigerant flows through an annulus between the tubes in the opposite direction. The lengths of the LTHX and HTHX were 1180 mm and 1400 mm, respectively, while the length of soldered capillary tube of the baseline system was 2000 mm. The condenser and R-evaporator in the original bypass two-circuit cycle used a counter flow type heat exchanger. Therefore, tube arrangement and internal volume were not changed in the LM cycle. However, the heat transfer area of
39
the R-evaporator in the LM cycle had to be enlarged because the mean evaporating temperature of the R-evaporator increased by the GTD of zeotropic mixture and the heat exchange in the LTHX. In this study, the required heat transfer area in the LM cycle, which was estimated by a thermodynamic analysis, was achieved by decreasing the fin pitches of the R-evaporator but not changing the tube arrangement. However, since the F-evaporator in the original bypass twocircuit cycle did not use a counter flow type heat exchanger, the F-evaporator was modified to have spirally-coiled circular fins. Spirally-coiled circular fin and tube heat exchanger easily allow a counter flow arrangement because, in a multi-column heat exchanger, each tube has individual fins. Meanwhile, the RF was controlled by a programmable micro-computer.
2.2.
Test procedure
The performance of the LM cycle was measured according to the refrigerant charge and the capillary tube length at an ambient temperature of 25 C and relative humidity of 70% (ISO 15502, 2005, N-class condition). The refrigerant charge was changed until the minimum energy consumption was observed in each capillary tube length. In addition, the capillary tube length was increased until the final cycle matching was achieved. The ambient condition was controlled by an environmental chamber. During the tests, the time-averaged air temperatures of the F- and R-compartments were maintained approximately at 18 C and 3 C, respectively. The internal air temperature of the F/R cabinet was determined as the average value of those measured by three temperature sensors located deep at the center of the compartment, equally spaced in the vertical direction. The cabinet air temperature was controlled by cyclic operation. In this study, the cut-in (set-on) and cut-out (set-off) temperatures were fixed at a constant value. For all tests, the compressor speed of the RF was maintained at 1800 RPM. This condition was appropriate for the energy saving mode operation of the RF. The performance of the RF was measured under the cyclic steady condition, in which the difference between the time-
Fig. 3 e Suction line assembly for (a) the bypass two-circuit cycle and (b) the LM cycle.
40
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 5 ( 2 0 1 2 ) 3 6 e4 6
averaged cabinet air temperatures of adjacent cycles was less than 0.2 C. All internal air temperatures of the cabinet were monitored and recorded every 30 s. The energy consumption of the RF, which indicates the total energy consumed by all components including the compressor and fan motors, was measured by a power meter with an uncertainty of 0.2% and integrated for four cycles. T-type copper-constantan thermocouples were used for temperature measurement. All thermocouples were calibrated to an accuracy of 0.2 C in a constant temperature bath.
3.
Thermodynamic cycle analysis
A thermodynamic analysis was conducted to find the optimum composition of the HC mixtures and the upper limit performance of the LM cycle. Furthermore, the energy saved by the GTD and the sub-cooling effects was also evaluated. In this study, binary mixtures of R-290/R-600, R-290/R-600a, and R-600a/R-600 were considered as the working fluid. The following assumptions were applied to the thermodynamic analysis: 1) pressure drops in the evaporator and condenser are neglected, 2) the ratio between the refrigerating effects of the F- and R- evaporators is 60:40, 3) the condenser outlet is in the saturated liquid state, 4) the R-evaporator outlet is in the saturated vapor state, 5) the UA of the F-evaporator is constant by practical considerations of space and cost (Smith et al., 1990), and 6) the UA of the condenser is proportional to the volumetric capacity of the refrigeration system with the mixtures. The condenser can be enlarged easily because it can be installed partially in the cabinet wall and operated under the no-frosting condition. The UA values of the F-evaporator and condenser were determined based on the baseline system of R-600a. The energy balance without latent cooling for a heat exchanger is given by Eq. (1). _ ref Dhref ¼ m _ air CP;air Tair;in Tair;out Q_ ¼ UADTLMTD ¼ m
DT1 DT2 lnðDT1 =DT2 Þ
(3)
DT1 ¼ Tair;in Tref;out ; DT2 ¼ Tair;out Tref;in
(4)
In Eq. (2), C is the clearance volume ratio and PD is the displacement rate of the compressor. ΔTLMTD of the F-evaporator was determined by the heat transfer rate because the UA of the F-evaporator was regarded as constant by practical considerations. The COP and volumetric capacity of the LM cycle are given by Eqs. (5) and (6), respectively. COP ¼ Q_ F þ Q_ R _ comp W VC ¼
(5)
hsat;vapor hsat;liquid vsuc
(6)
In the volumetric capacity calculation, the mean evaporating temperature and suction temperature were assumed to be 25 C and 25 C, respectively. Table 2 shows the input conditions for the thermodynamic analysis. The size of the LTHX was assumed to be 20% of the total evaporator UA. Jung and Radermacher (1991) reported that the maximum size of the LTHX was limited to roughly 20% of the total evaporator UA. Fig. 4 shows the flow chart of the simulation. The condensing and evaporating pressures were determined by the energy balance in the F-evaporator and condenser given in Eq. (1). After the condensing and evaporating pressures were determined, the required heat transfer area of the R-evaporator was calculated for each case.
4.
Results and discussion
4.1. Optimization of mixture composition using simulation results
(1)
_ ref and DTLMTD are calculated by Eqs. (2) and (3), In Eq. (1), m respectively. 3 1n P cond 5 PD ¼ 41 þ C C Pevap vsuc 2
_ ref m
DTLMTD ¼
(2)
The COP and volumetric capacity of the LM cycle using R-290/ R-600, R-290/R-600a, and R-600a/R-600 were analyzed to determine the optimum mixture composition. Fig. 5 shows the variations of the energy savings for the LM cycle with the HC mixtures according to the mixture composition. The energy saving potential was estimated by a percent decrease in the energy consumption of the LM cycle as compared with
Table 2 e Input conditions for cycle simulation. Parameter Inlet air temperature of the R-evaporator Inlet air temperature of the F-evaporator Air flow rate of the condenser Air flow rate of the F-evaporator Air flow rate of the R-evaporator UA of the LTHX Suction temperature Isentropic efficiency Clearance volume ratio
Specified value
Technical ground for the value
3 C 18 C 1.5 m3 min1 0.7 m3 min1 0.6 m3 min1 20% of total evaporator 25 C 70% 0.02
Internal cabinet air temperature Internal cabinet air temperature Measured data for the baseline system Measured data for the baseline system Measured data for the baseline system Reference data (Jung and Radermacher, 1991) Anti-sweat condition for suction Estimated value Estimated value
41
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 5 ( 2 0 1 2 ) 3 6 e4 6
Fig. 4 e Flow chart of the LM cycle simulation.
that of the bypass two-circuit cycle using R-600a (the baseline system). The R-290/R-600 showed the highest energy saving of 15.3% at the R-290 mass fraction of 0.4 mainly because it had the most suitable GTD among the mixtures. As mentioned earlier, the energy saving of the LM cycle increases with a decrease in the compression ratio due to higher mean evaporating temperature. Therefore, the efficiency of the LM cycle would be most improved when the refrigerant temperature profile closely matches the air temperature profile in the F-evaporator. In the lower GTD case, in which the GTD in the F-evaporator is smaller than the variation of air temperature through the F-evaporator, the increase in the mean evaporating temperature would be very small. However, an
excessive GTD can lead to the reduction of the cooling capacity at constant UA. The variation of air temperature through the F-evaporator was about 8 C at the air flow rate and cooling capacity of the baseline system, while the GTDs in the F-evaporator of R-290/R-600, R-290/R-600a, and R-600a/R600 were 8.6, 4.6, and 0.6 C, respectively, at the optimum composition of each mixture. The GTDs of R-290/R-600a and R-600a/R-600 were obviously relatively lower for domestic RFs. Therefore, R-290/R-600 was selected as the optimum binary mixture for the LM cycle. Fig. 6 represents the variation of volumetric capacity in the LM cycle according to the mixture composition when the evaporating temperature and suction temperature are 25 C and 25 C, respectively. The displacement rate (PD) of the compressor needs to be adjusted for each mixture composition so that the LM cycle can maintain a constant refrigeration capacity (Stoecker and Walukas, 1981). An excessive displacement volume of the compressor may increase cycling losses because the compressor would then have to turn on and off more frequently. In addition, it may lead to higher mean temperature differences between the refrigerant and air in the heat exchangers. The volumetric capacity of the LM cycle using R-290/R-600 with an R-290 mass fraction of 0.2 was similar to that using R-600a. When the baseline compressor for R-600a is directly used in the LM cycle, R-290/R-600 with the R-290 mass fraction of 0.2 would be the most recommendable mixture composition for the LM cycle. However, as mentioned in Fig. 5, the maximum energy saving for R-290/ R-600 was observed at the R-290 mass fraction of 0.4. The energy saving for R-290/R-600 at the R-290 mass fraction of 0.2 was estimated at 14.1%, which was 1.2% lower than the maximum value at the R-290 mass fraction of 0.4. Therefore, R-290/R-600 with the R-290 mass fraction of 0.4 is recommended to maximize the energy saving of the LM cycle. In this case, a down-sized compressor from the baseline system has to be adopted to output the same capacity as the baseline system. Fig. 7 represents the linearity characteristics in saturation temperature profile of R-290/R-600 at the mean evaporating temperature of 25 C. The normalized temperature is defined
1200 -3
15
Volumetric capacity kJ m
Energy saving potential (%)
20
10 5 0 -5 -10 0.0
R-290/R-600 R-290/R-600a R-600a/R-600 0.2
0.4
0.6
0.8
Mass fraction of lower NBP refrigerant Fig. 5 e Variation of the energy saving potential with mixture composition.
1.0
1000
R-290/R-600 R-290/R-600a R-600a/R-600
800
600
Baseline: R-600a
400
200 0.0
0.2
0.4
0.6
0.8
1.0
Mass fraction of lower NBP refrigerant Fig. 6 e Variation of the volumetric capacity with mixture composition.
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 5 ( 2 0 1 2 ) 3 6 e4 6
20
1.0 Energy saving potential (%)
Normalized temperature T-Tsat.liquid /GTD
42
0.8 0.6 Mass fraction of R-290 o 0.1 GTD=7.6 C o 0.2 GTD=11.9 C o 0.3 GTD=14.3 C o 0.4 GTD=15.4 C
0.4 0.2 0.0
0.0
0.2
0.4
0.6
0.8
10 5 0 -5 -10 0.0
1.0
Case 1 Case 2 Case 3 Case 4
15
0.2
0.4
Vapor quality
as (TTsat, liquid)/GTD. The linearity of the refrigerant temperature with respect to vapor quality increased with the increase in the R-290 mass fraction. Higher linearity can lead to larger energy saving than that of the baseline system because it impedes refrigerant temperature rise and helps to remove the pinch point in the F-evaporator. As the R-290 mass fraction increased from 0.2 to 0.4, the mean evaporating temperature of the LM cycle increased by 1.6 C, which increased the energy saving. However, as the R-290 mass fraction increased beyond 0.4, the excessive volumetric capacity of R-290/R-600 increased ΔTLMTD in the F-evaporator, which reduced the energy saving. The GTD of R-290/R-600 showed a maximum value at the R-290 mass fraction of 0.5. Cycle simulations for four cases were carried out to investigate the energy saving of the LM cycle using R-290/ R-600 according to individual effect of the GTD and subcooling. Case 1 is the baseline for illustrating each effect. The mean evaporating temperature was fixed at 25 C, and the sub-cooling effect by the LTHX was not included. In Case 2, the GTD effect for higher mean evaporating temperature was considered. In Cases 3 and 4, both the GTD and the sub-cooling effect by the LTHX were taken into account. However, for Cases 3 and 4, the size of the LTHX was assumed to be 10% and 20% of the total evaporator UA, respectively. Fig. 8 shows the variation of the energy saving with mixture composition for each case. In Case 1, the energy saving was only 2.5% at the R-290 mass fraction of 0.2. In Case 2, an additional energy saving of 5.7% can be obtained by including the GTD effect with an increase in the mean evaporating temperature of 2.6 C. The energy savings for Cases 3 and 4 increased by 3.5% and 3.7% over Cases 2 and 3, respectively, due to the subcooling effect of the LTHX. In Cases 2e4, the optimum mass fraction of R-290 was 0.4 with respect to the energy saving. The mean evaporating temperatures for Cases 3 and 4 increased by 0.7 C in each case due to the sub-cooling effect of the LTHX. As mentioned earlier, the heat transfer area of the R-evaporator in the LM cycle should be increased because of the increase in the mean evaporating temperature of the
0.8
1.0
Fig. 8 e Variation of the energy saving potential with mixture composition for each case.
R-evaporator by the GTD and heat exchange in the LTHX. Fig. 9 represents the required heat transfer area of the R-evaporator in the LM cycle using R-290/R-600, compared with that of the baseline system. The results were obtained at a constant compressor capacity. As the LTHX size was increased, larger heat transfer area was required due to the higher evaporating temperature in the R-evaporator. In Case 4, the required heat transfer area of the R-evaporator at the R-290 mass fraction of 0.4 was estimated to be about 3 times that of the baseline system. However, when the compressor is down-sized to use the R-290 mass faction of 0.4, the heat transfer area needed to be only 2.2 times that of the baseline system. The volumetric capacity of R-290/R-600 at the R-290 mass faction of 0.4 was approximately 38% larger than that of R-600a, so the compressor size and required heat transfer area decreased for the same refrigeration capacity. Fig. 10 shows the R-evaporator for each cycle applied in the present experiments. The volumes occupied by each R-evaporator were very similar because a compact R-evaporator with smaller fin pitches was
5 Required area / Baseline area
Fig. 7 e Linearity characteristics of saturation temperature profiles of R-290/R-600.
0.6
Mass fraction of R-290
4
Case 2 Case 3 Case 4
3
2
1
0 0.0
0.2
0.4
0.6
0.8
1.0
Mass fraction of R-290 Fig. 9 e Ratio of the required heat transfer area of Revaporator in the LM cycle to that in the baseline system.
43
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 5 ( 2 0 1 2 ) 3 6 e4 6
Fig. 10 e R-evaporator for (a) the bypass two-circuit cycle and (b) the LM cycle.
4.2.
System optimization using experimental results
The performance of the LM cycle using R-290/R-600 was measured by varying the refrigerant charge amount, capillary tube length, mixture composition, and compressor capacity. Both a baseline reciprocating compressor and a down-sized reciprocating compressor were adopted for the LM cycle using R-290/R-600. The baseline compressor originally designed for R-600a had a displacement volume of 15 cc. The down-sized compressor for R-290/R-600 had a displacement volume of 11 cc, which was properly designed to have the same capacity at the R-290 mass faction of 0.4 as that of the baseline system. Both compressors had almost the same efficiency at 1800 RPM under the test conditions used in this study. Fig. 11 represents the effects of the refrigerant charge, R-290 mass fraction, and capillary tube length on the energy consumption of the LM cycle with the baseline compressor. The experiments were performed for the R-290 mass fractions of 0.2 and 0.3 because the volumetric capacity of the LM cycle for the R-290 mass fraction of 0.1 was considerably lower than the designed value. Optimum refrigerant charge in each capillary tube length was determined by the minimum energy consumption for each case. For the R-290 mass fraction of 0.2, the optimum refrigerant charge increased from 125 to 160 g as the capillary tube length increased from 13 to 17 m. This trend was very similar to that for the R-290 mass fraction of 0.3, but the range of the optimum refrigerant charge for the R-290 mass faction of 0.3 was lower than that for the R-290 mass fraction of 0.2 due to higher volumetric capacity of the LM cycle. As the capillary tube length increased at a fixed refrigerant charge, both the refrigerant flow rate through the LM
cycle and the evaporating temperature decreased, so the cooling capacity in the R-evaporator reduced. As the heat transfer rate in the F-evaporator increased with the decrease in the evaporating temperature, the cooling capacity of the R-evaporator decreased due to the limited remaining capacity in the R-evaporator during the R/F simultaneous operation. The decrease of the cooling capacity in the R-evaporator due to the longer capillary tube has to be compensated by the increase in the refrigerant charge. In addition, the minimum energy consumption for the R-290 mass fraction of 0.2 was 1.6% lower than that for the R-290 mass fraction of 0.3 because of excessive volumetric capacity for the baseline compressor in the R-290 mass fraction of 0.3. Therefore, the lowest energy consumption of the LM cycle with the baseline compressor was observed in the R-290 mass faction of 0.2 at the capillary tube length of 15 m and refrigerant charge of 140 g. As shown in Fig. 11, the optimum capillary tube length of 15 m in the LM cycle with the baseline compressor was significantly longer than the optimum value of 3.3 m in the bypass two-circuit cycle (Table 1) because of the larger sub-
34 Energy consumption (kWh/month)
applied for the LM cycle. A forced convection defrosting can be a solution for the frost accumulation in the compact R-evaporator. The forced convection defrosting was performed by using internal air circulation through the R-evaporator when the refrigerant did not pass through the R-evaporator. In addition, a counter flow and spirally-coiled circular fin type F-evaporator for the LM cycle was designed to have almost the same heat transfer area as that of the baseline system.
R-290 mass fraction 0.2 0.3
33
Capillary tube length 32
13 m
15 m
17 m
19 m
17 m
31
15 m 30 110
120
130
140
150
160
170
Optimum refrigerant charge (g) Fig. 11 e Variation of the energy consumption with the refrigerant charge, R-290 mass fraction, and capillary tube length for the baseline compressor.
44
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 5 ( 2 0 1 2 ) 3 6 e4 6
Table 3 e Performances of the LM cycle and the bypass two-circuit cycle with the baseline compressor for R-600a. Parameter
Bypass two-circuit cycle
R/F-operation
F-operation
R/F-operation
F-operation
16.6% 77.7 W 46.2 C 28.6 C 24.0 C 1.6 C 382.6 kPa 75.4 kPa
39.8% 68.2 W 46.4 C 28.0 C 31.3 C 13.0 C 412.4 kPa 56.1 kPa
23.5% 74.4 W 46.3 C 30.8 C 19.1 C 19.3 C 413.7 kPa 71.9 kPa
43.1% 65.3 W 46.0 C 30.2 C 27.7 C 27.9 C 406.8 kPa 50.0 kPa
30.9 kWh/month
Fig. 13 shows the effects of the refrigerant charge, R-290 mass fraction, and capillary tube length on the energy consumption of the LM cycle with the down-sized compressor. As shown in Figs. 11 and 13, the refrigerant charge and capillary tube length had similar effects on the energy consumption for both compressors. For all R-290 mass
10 5
F-evaporator
R-evaporator
-5
Temperature
Connecting component
Air
o
C
0
-10 -15 Air
-20 -25
Refrigerant
Refrigerant -30 -35
R-evap. in R-evap. out F-evap. in F-evap. out Location of the cycle path
a
Bypass two-circuit cycle
10 5
LTHX
F-evaporator
R-evaporator
C
0 Air
-5
o
cooling of the LM cycle. Due to the heat transfer in the two sub-coolers of the LM cycle, the sub-cooling at the capillary tube inlet was large, as much as 40 C, while that of the conventional cycle was about 0 C. Fatouh (2007) reported that larger sub-cooling in the LM cycle can lead to a serious increase in the capillary tube length. In addition, the optimum refrigerant charge in the LM cycle was 46% higher than that in the bypass two-circuit cycle (Table 1) because of the condenser liquid lines in the two sub-coolers. Table 3 summarizes the performance of the LM cycle and the bypass two-circuit cycle for the same baseline compressor at the optimized refrigerant charge and capillary tube length. The energy consumption of the optimized LM cycle with the baseline compressor was 10.1% lower than that of the bypass two-circuit cycle, so it was 4% lower than the energy saving predicted by the simulation. This difference in the energy saving may be attributed to the increase in cycling losses of the LM cycle due to larger refrigerant charge and more frequent on/off cycling operations of the compressor. The increased refrigerant charge can cause an increase in the migration loss at compressor shut-off and redistribution loss at compressor start-up (Bjo¨rk and Palm, 2006). The compressor was turned on and off more frequently in the LM cycle due to the lower operating ratio resulting from larger cooling capacity in the LM cycle, as shown in Table 3. Fig. 12 represents the temperature distributions of the refrigerant and air through F- and R-evaporators for the bypass two-circuit cycle and the optimized LM cycle with the baseline compressor. All temperatures in the operating period are time-averaged values. In the LM cycle, the temperature profile of the refrigerant through the F-evaporator was almost parallel with the air temperature profile, so a higher energy saving was expected. The mean evaporating temperature in the LM cycle was approximately 3 C higher than that in the bypass two-circuit cycle, and the temperature of the air leaving the F-evaporator in the LM cycle was 1.9 C lower than that in the bypass two-circuit cycle, resulting in approximately 24% larger cooling capacity of the F-evaporator in the LM cycle. Based on the simulation results, the cooling capacity and compressor power input of the LM cycle with the baseline compressor were 23% and 8% higher, respectively, than those of the bypass two-circuit cycle. In addition, the ΔTLMTD in the R-evaporator of the LM cycle was significantly less than that of the bypass two-circuit cycle, allowing lower entropy generation.
34.4 kWh/month
Temperature
Operating ratio Average power input Discharge temperature Condenser outlet temperature Evaporator inlet temperature Evaporator outlet temperature Condensing pressure Evaporating pressure Energy consumption
LM cycle
-10 -15 -20
Air Refrigerant
-25 -30 -35
Refrigerant F-evap. in
F-evap. out R-evap. in R-evap. out Location of the cycle path
b
LM cycle
Fig. 12 e Temperature distributions of refrigerant and air through F- and R-evaporators for (a) the bypass two-circuit cycle and (b) the LM cycle.
45
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 5 ( 2 0 1 2 ) 3 6 e4 6
32
-12 o
Capillary tube length 20 m
31
22 m 24 m
18 m 18 m
22 m
20 m 22 m
20 m 30 110
120
130
Predicted data F-evap. inlet F-evap. outlet R-evap. inlet
C
R-290 mass fraction 0.35 0.40 0.45
Evaporating temperature
Energy consumpation (kWh/month)
33
-16
-20
-24
-28 Measured data
140
150
160
-32 0.0
170
Optimum refrigerant charge (g) Fig. 13 e Variation of the energy consumption with the refrigerant charge, R-290 mass fraction, and capillary tube length for the down-sized compressor.
factions, the optimum refrigerant charge increased with the increase in the capillary tube length. In addition, the range of the optimum refrigerant charge decreased with the increase in the R-290 mass fraction because of the increase in the volumetric capacity. The lowest energy consumption of the LM cycle with the down-sized compressor was observed for the R-290 mass faction of 0.4 at the capillary tube length of 20 m and refrigerant charge of 135 g. The best performance of the LM cycle for the R-290 mass fraction of 0.4 was due to proper design of the compressor capacity at the given mixture composition. The energy consumption of the optimized LM cycle with the down-sized compressor was 1.1% lower than that with the baseline compressor because of better linearity characteristics and larger GTD of R-290/R-600 for the R-290 mass fraction of 0.4. These results are consistent with the predictions from the simulations. The optimum capillary tube length (20 m) for the down-sized compressor was longer than that (15 m) for the baseline compressor. Generally, the pressure and sub-cooling at the capillary inlet are dominant factors for determining the mass flow rate at a given capillary tube geometry. Based on the results of Fatouh (2007), the optimum capillary tube length for the HC mixtures is directly proportional to the pressure at the capillary inlet. As the R-290 mass fraction increased from 0.2 to 0.4, the pressure at the capillary inlet was increased by approximately 30%, resulting in the increase in the capillary tube length. However, the variation of the sub-cooling with the increase in the R-290 mass fraction of 0.2e0.4 was relatively small. Fig. 14 compares the measured evaporating temperatures with simulation predictions. As the R-290 mass fraction increased from 0.2 to 0.4, the evaporating temperatures at each position averagely increased, as predicted in the simulation. The increase in the evaporating temperature allows higher energy saving of the LM cycle. The temperature difference of the refrigerant through the LTHX increased from 1.9 to 3.7 C with the increase in the R-290 mass fraction from 0.2 to 0.4, while that through the F-evaporator slightly decreased by 0.4 C due to the linearity characteristics. As shown in Fig. 7, the
0.2
0.4
0.6
0.8
1.0
Mass fraction of R-290 Fig. 14 e Comparison of the measured evaporating temperatures with simulation predictions.
evaporating temperature of R-290/R-600 for the R-290 mass fraction of 0.2 increased more rapidly than that for the R-290 mass fraction of 0.4 in the lower quality region. This rapid increase led to a relatively large evaporating temperature increase in the F-evaporator, which operates in the lower quality region, and small increase in the LTHX. The measured temperature differences through the LTHX were similar to the simulation results obtained for the optimized LTHX capacity.
5.
Conclusions
The performance of the LM cycle was measured and analyzed by varying the refrigerant charge, capillary tube length, compressor capacity, and mixture composition, and then compared with that of the bypass two-circuit cycle. From the simulation, the energy consumption of the LM cycle using R290/R-600 (40:60%) was 15.3% lower than that of the bypass two-circuit cycle using R-600a. However, for the optimum R290 mass faction of 0.4, the baseline compressor had to be down-sized by approximately 38% to allow proper refrigeration capacity in the LM cycle. In this case, the heat transfer area of the R-evaporator should be increased to be 120% greater than that of the bypass two-circuit cycle using R-600a. Based on the experimental results, the optimum refrigerant charge for the LM cycle using R-290/R-600 increased with the increase in the capillary tube length because of the decrease in the refrigerant flow rate and evaporating temperature. The optimum refrigerant charge and capillary tube length of the LM cycle using R-290/R-600 were significantly larger than those of the bypass two-circuit cycle using R-600a due to the condenser liquid lines in the two sub-coolers and larger subcooling, respectively. The energy consumption of the LM cycle using R-290/R-600 (40:60%) with the down-sized compressor was reduced to be 11.2% lower than that of the bypass two-circuit cycle using R-600a with the baseline compressor. In addition, the energy consumption of the LM cycle using R-290/R-600 (20:80%) with the baseline compressor was reduced to be 10.1% lower than that of the bypass twocircuit cycle using R-600a.
46
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 5 ( 2 0 1 2 ) 3 6 e4 6
Acknowledgments This work was supported by the Korea Institute of Energy Technology Evaluation and Planning (KETEP) grant funded by the Korea government Ministry of Knowledge Economy (No. 20103010080011) and Korea University.
references
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