Role of receiver on the performance of a transcritical CO2 based air-conditioning unit with single-stage and two-stage expansion

Full Terms & Conditions of access and use can be found at https://www.tandfonline.com/action/journalInformation?journalCode=uhvc21 Science and Technology for the Built Environment ISSN: (Print) (Online) Journal homepage: https://www.tandfonline.com/loi/uhvc21 Role of receiver on the performance of a transcritical CO2 based air-conditioning unit with single-stage and two-stage expansion Mihir Mouchum Hazarika, Maddali Ramgopal, Souvik Bhattacharyya & Rodrigo Llopis Doménech To cite this article: Mihir Mouchum Hazarika, Maddali Ramgopal, Souvik Bhattacharyya & Rodrigo Llopis Doménech (2021): Role of receiver on the performance of a transcritical CO2 based airconditioning unit with single-stage and two-stage expansion, Science and Technology for the Built Environment, DOI: 10.1080/23744731.2021.1902189 To link to this article: https://doi.org/10.1080/23744731.2021.1902189 © 2021 The Author(s). Published with license by Taylor & Francis Group, LLC Published online: 22 Mar 2021. Submit your article to this journal Article views: 68 View related articles View Crossmark data Role of receiver on the performance of a transcritical CO 2 based air-conditioning unit with single-stage and two-stage expansion MIHIR MOUCHUM HAZARIKA 1 , MADDALI RAMGOPAL 2 , SOUVIK BHATTACHARYYA 3 , and RODRIGO LLOPIS DOMÉNECH 4 1 Department of Energy and Process Engineering, Norwegian University of Science and Technology, Trondheim, Norway 2 Department of Mechanical Engineering, IIT Kharagpur, Kharagpur, India 3 Department of Mechanical Engineering, BITS Pilani, Pilani, India 4 Department of Mechanical Engineering and Construction, Jaume I University, Castell o, Spain Theoretical and experimental studies are carried out on a transcritical CO 2 based air-conditioning unit. Two configurations–a system with single-stage expansion (SSE) and a system with two-stage expansion (TSE) have been considered. Detailed numerical models have been developed for both configurations. Results obtained from numerical simulations are compared with experimental results. The performance of the system with single-stage expansion is compared with two-stage expansion under different conditions of refrigerant charge, receiver volume, and ambient temperature. Both numerical and experimental results show that the SSE configuration is highly sensitive to refrigerant charge as compared to the TSE configuration. The TSE configuration offers better system controllability provided the receiver is sized properly. Besides, it is found that for the same performance, the required charge can be reduced by about 15% for the TSE configuration as compared to the SSE configuration. Introduction With the rising concerns about global warming, the use of harmful synthetic refrigerants with high global warming potential (GWP) is no longer encouraged. Environment-friendly, natural refrigerants are seen as permanent replacements for the high GWP, synthetic refrigerants. Hydrocarbons have been widely used in recent years for small capacity refrigeration and air-conditioning applications. However, hydrocarbons are flammable. Though ammonia is an excellent refrigerant and is widely used, it is toxic and suffers from material compatibility issues. Carbon dioxide (CO 2 ), being non-flammable as well as nontoxic, is considered as a promising natural refrigerant due to its excellent thermophysical properties and material compatibility. However, CO 2 has a comparatively low critical temperature (31 C). Lorentzen and Pettersen (1993)have shown the possibility of using CO 2 in a transcritical refrigeration cycle by developing a system for air-conditioning application, thereby overcoming the operational difficulties associated with the low critical temperature of CO 2 . The performance of a single-stage transcritical CO 2 based system drops at high ambient temperature. To enhance the performance of transcritical CO 2 based system, extensive studies have been carried out in recent years. Different modifications have been proposed for the CO 2 system. The use of an ejector as an expansion work recovering device, use of parallel compression, and use of mechanical subcooling are considered as effective solutions to enhance the performance of the transcritical CO 2 system. Groll and Kim (2007) carried out an extensive review to analyze the potential benefits of using different modifications to enhance the performance of a transcritical CO 2 system. These modifications are the use of two-stage compression, use of work recovering expander, use of vortex tube as an expansion device, use of ejector as an expansion device, and use of thermoelectric subcooling device to enhance the performance of the CO 2 system. They observed that the two-stage cycle with intercooling gave the highest energy efficiency. Lawrence and Elbel (2015) carried out studies to compare the performance of a CO 2 ejector with an R134a ejector. They observed that Received August 16, 2020; accepted March 4, 2021 Mihir Mouchum Hazarika, PhD, is a Postdoctoral Fellow. Maddali Ramgopal, PhD,isaProfessor.Souvik Bhattacharyya, PhD, is a Senior Professor and Vice-Chancellor. Rodrigo Llopis Dom enech, PhD,isaProfessor. Corresponding author e-mail: [email protected] This is an Open Access article distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/by/4.0/), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited. Science and Technology for the Built Environment, (2021) 0,1–28 Copyright #2021 The Author(s). Published with license by Taylor & Francis Group, LLC ISSN: print / online DOI: 10.1080/23744731.2021.1902189 the CO 2 ejector performed better as compared to the R134a ejector. Later, Elbel and Lawrence (2016) presented an extensive review on the use of an ejector as an expansion work recovering device in the transcritical CO 2 system. They reported that improvement in COP is in the range of 10 to 30% for transcritical CO 2 systems with ejectors. In a recent study, Zhu and Elbel (2020) introduced a new method in which a vortex was generated in the motive flow of CO 2 ejector to regulate the high-side pressure. They reported that the use of this vortex control method in the CO 2 ejector improved the system COP and capacity by 8.1% and 11.0%, respectively under off-design conditions. Cao, Ye, and Wang (2020) carried out an experimental study to analyze the effects of using an internal heat exchanger in a transcritical CO 2 heat pump. Karampour and Sawalha (2018) showed the potential benefits of using parallel compression and mechanical subcooling in a transcritical CO 2 booster system. Bush et al. (2017) conducted experimental studies to investigate the performance of a transcritical CO 2 booster system with mechanical subcooling. They reported that the use of mechanical subcooling significantly improved the performance and reduced the amount of flash gas in the flash tank. Catal an-Gil et al. (2019) carried out studies to compare the energy improvements gained by using integrated and dedicated mechanical subcooling in a transcritical CO 2 booster system. In a recent experimental study, Nebot-Andr es et al. (2020) estimated the optimum operating conditions of a transcritical CO 2 system with integrated mechanical subcooling. All these studies presented here showed that CO 2 systems could provide long term energy-efficient solutions, provided the issues related to performance are suitably addressed. Unlike conventional refrigeration cycles, the transcritical refrigeration cycle needs a suitable control scheme to optimize the high side pressure owing to the peculiar shape of the isotherm in the supercritical region. Several control schemes have been proposed to optimize the high side pressure for a CO 2 based system. Few studies reported methods to optimize the high side pressure by regulating the speed of the gas cooler fan (Baek et al. 2013). For such a system, the high side pressure can also be optimized by using an electronic expansion valve (Cho, Ryu, and Kim 2007; Hou et al. 2014). Although, single expansion valve can effectively maintain the optimum high side pressure of a transcritical CO 2 based system; it is not possible to maintain the desired refrigerant flowrate at the evaporator simultaneously to maintain the degree of superheat. Casson et al. (2003) suggested a novel method of using two expansion valves with a receiver in between for simultaneous control of gas cooler pressure as well as the degree of superheat. This particular control strategy has been found more reliable in optimizing the performance of CO 2 based systems (Boccardi et al. 2013; Cabello et al. 2008; Llopis et al. 2016). Most of the commercial CO 2 systems designed for supermarkets are now equipped with this control strategy using two-stage expansion (Gullo, Hafner, and Banasiak 2018). However, the detailed theoretical and experimental studies on systems using two EEVs and their performance comparison with single EEV are scarce in the literature. A detailed literature survey reveals that even though several mathematical models have been proposed to simulate the performance of CO 2 based transcritical systems (Lin et al. 2013), the effect of refrigerant charge on system performance is not considered by many. The impact of the refrigerant charge on system performance is very important, especially in mobile air conditioning applications where the refrigerant leakage rates are typically very high. Refrigerant leakages are also relatively high in small split air conditioning systems. Since depletion of refrigerant is expected to affect the performance of critically charged systems such as the ones mentioned, it is important to carry out studies to evaluate the same. Cho et al. (2005) carried out an experimental study to compare the performance of a transcritical CO 2 system with R22, R410A, and R407C systems. They reported that the CO 2 system showed the highest sensitivity in performance to change in refrigerant charge. Moreover, the performance was found to be deteriorating significantly at undercharged conditions. Hazarika, Ramgopal, and Bhattacharyya (2018) developed a numerical model for a transcritical CO 2 based air-conditioning system and carried out experiments to validate the model. The numerical results showed that the COP of the system drops marginally for ±18% change from the optimum refrigerant charge. However, any change in charge beyond this range affects the performance significantly. Wang et al. (2019) performed an experimental study to optimize the refrigerant charge for a transcritical CO 2 system for heating application. They also reported the effect of system transient on the migration of refrigerant charge. These studies show the importance of refrigerant charge for operating the system at optimum conditions. It is seen that very limited studies are available in the literature on the performance of single-stage expansion vis-a-vis two-stage expansion in transcritical CO 2 systems. The objective of the present study is to investigate the performance of a transcritical CO 2 based system. The intended applications of the system studied are in the field of small capacity (around 3.5 to 7 kW) systems for residential and mobile air conditioning. Two configurations are selected for the proposed CO 2 system: one configuration employs a single expansion valve while the other employs double expansion valves. The primary goal of this study is to compare the performance of the proposed configurations. For the configuration with two-stage expansion, a receiver must be located between the two expansion valves. The impact of the size of the receiver on system performance is also analyzed in the present study. To fulfill these objectives, experimental tests and numerical simulations are carried out for both configurations. The influences of refrigerant charge, receiver volume, and ambient temperature on system performance are presented and analyzed in the results. Control strategies with expansion device for transcritical CO 2 cycle One of the commonly used techniques to control the highside pressure for transcritical CO 2 cycle is to use an 2Science and Technology for the Built Environment electronic expansion valve driven by a feedback controller. The feedback controller adjusts the cross-sectional area of the expansion valve opening to maintain the desired highside pressure. However, with this control strategy, it is not possible simultaneously to feed the evaporator with the desired refrigerant flowrate and control the heat rejection pressure. For such configuration, an accumulator as a suction line heat exchanger should be located before the compressor and this accumulator could be flooded partially with liquid refrigerant. Lorentzen and Pettersen (1993) suggested that the accumulator helps in supplying the necessary refrigerant flowrate to the evaporator under varying operating conditions. Traditionally, a single expansion valve (mechanical or electronic) is used in a refrigeration system to maintain the desired degree of superheat at the evaporator exit. When this approach is adopted for the transcritical CO 2 cycle, to maintain the desired high-side pressure, a given system has to be charged with an optimum amount of refrigerant for a specific operating condition (Aprea, Greco, and Maiorino 2015; Cho et al. 2005; Hazarika, Ramgopal, and Bhattacharyya 2018). However, when the operating conditions change due to change in heat source or sink temperature and/or refrigeration load, the resulting high-side pressure may deviate from the desired optimum high-side pressure, resulting in suboptimal performance. Hence in a transcritical cycle, to control the high-side pressure and degree of superheat simultaneously for maximum COP, the use of two-stage expansion with a receiver in between is suggested by Casson et al. (2003). It is shown that the first expansion valve plays the role of a differential expansion valve that maintains the receiver pressure as well as the differential pressure drop. The receiver is always maintained at saturation pressure such that saturated liquid exists at the outlet of the differential valve. The second expansion valve maintains the degree of superheat by feeding the evaporator with the required refrigerant flowrate. Thus it is possible to maintain the high-side pressure as well as the degree of superheat simultaneously. During operation, the receiver is kept partially flooded with liquid refrigerant. Casson et al. (2003) suggested that, with this control strategy, it is possible to achieve optimum performance for the transcritical cycle under varying operating conditions. However, it is essential to predict the size of the receiver as well as the amount of refrigerant charge appropriately to maintain sufficient liquid in the receiver under varying operating conditions. These issues are addressed in the present work. Experimental test facility An experimental test facility is built to study the performance of the system with single and two-stage expansion. The test-rig comprises of fin-and-tube type evaporator and gas cooler, electronic expansion valves, a receiver, an accumulator, and a semi-hermetic compressor. A photograph and schematic drawing of the test-rig are shown in Figure 1. Tests are conducted to investigate the sensitivity of the performance of the system to refrigerant charge with two-stage and single-stage expansion. Then by comparing the results, it is observed that the numerical results agree reasonably well with the experimental results. The maximum uncertainty in cooling capacity and gas cooler heat rejection rates are estimated to be 8.7% and 9.6%, respectively. The detailed description of the test facility is presented in earlier communication by the authors (Hazarika, Ramgopal, and Bhattacharyya 2018). Simulation model A mathematical model is developed to simulate the performance of the experimental test-rig shown in Figure 1. The steady-state model presented in this manuscript has similarities with the model presented in the earlier communication (Hazarika, Ramgopal, and Bhattacharyya 2018). This model is developed by integrating the models for individual components. The individual component models are adopted from the earlier manuscript (Hazarika, Ramgopal, and Bhattacharyya 2018). In the earlier case, the steady-state model is developed considering a single expansion valve that is used to maintain the desired degree of superheat. In the present study, the steady-state model for the CO 2 system is upgraded considering double expansion valvesthe first one maintains the desired high-side pressure while the second one maintains the desired degree of superheat. Finally, numerical simulations are carried out for both configurations: single-stage expansion system and two-stage expansion system. In case of the configuration with two-stage expansion, the high-pressure fluid that leaves the gas cooler is expanded in the first-stage (EEV1). After the first stage of expansion, the fluid enters the receiver. The fluid exiting the receiver is then expanded in the second-stage (EEV2). On the other hand, in the case of the configuration with a single-stage of expansion, the first stage of expansion is bypassed as shown in Figure 1b. Compressor The compressor is modeled assuming an irreversible, but adiabatic compression process. Empirical correlations for isentropic and volumetric efficiencies of the compressor suggested by Wang et al. (2013) are used to predict the power input to the compressor and mass flow rate of the refrigerant. For the sake of simplification, the refrigerant present in the compressor is neglected in the present model. However, the amount of refrigerant present in the compressor can be estimated by considering the single-phase vapor present in the shell side of the compressor and the liquid refrigerant dissolved in the lubricating oil (He et al. 2020). It is observed that this amount is less than 2% of the total refrigerant charge. This is due to the presence of two pressure vessels, the receiver and accumulator, which contain a major portion of the refrigerant charge. Hence, this amount Volume 0, Number 0, Month 2021 3 Fig. 1. (a) Photograph and (b) schematic of the CO 2 based air conditioning unit. 4Science and Technology for the Built Environment of charge present in the compressor is neglected in this study. Gas cooler The gas cooler is a fin-and-tube heat exchanger with spiral fins as shown in Figure 2. Air and refrigerant approach the gas cooler from the opposite direction as the counter-cross flow arrangement is chosen. This gas cooler is modeled using energy balance equations for the air and refrigerant side with suitable empirical correlations for heat transfer and pressure drop available in the literature. Considering the steep property variation on the CO 2 side, the entire gas cooler is discretized into finite elements (Yin, Bullard, and Hrnjak 2001) and the governing equations are applied across each element. An iterative procedure is adopted to find the outlet conditions of the gas cooler from known inlet conditions on air and refrigerant sides. The equations are solved numerically using an element size that leads to a grid-independent solution. For each element, heat transfer rate is expressed as: DQ¼eCmin Tref i,j,k ðÞ Tai,j,k ðÞ  ¼_ maCpaTa,ex i,j,k ðÞ Ta,in i,j,k ðÞ  ¼_ mref href iþ1, j,k ðÞ href i,j,k ðÞ  (1) for Cmax ¼CCO2;e¼1exp cCmax Cmin (2) where, c¼1exp UA Cmax  (3) for Cmin ¼CCO2;e¼Cmax Cmin 1exp cCmin Cmax  (4) where, c¼1exp UA Cmin  (5) 1 UA ¼1 aagoverall anode þ1 aref ai (6) The heat transfer coefficient of CO 2 during the heat exchange process in the gas cooler is estimated using the correlation proposed by Pitla, Groll, and Ramadhyani (2002): ai¼Nu di kref (7) Nu ¼Nuwall þNubulk 2  kwall kbulk (8) where “Nu wall ”and “Nu bulk ”are Nusselt numbers calculated at the wall and bulk temperatures. These two parameters “Nu wall ”and “Nu bulk ”are calculated using Gnielinski correlation within the range 2300 <Re <10 6 and 0.6 <Pr <10 5 : Nu ¼f=8 ðÞ Re 1000 ðÞ Pr 12:7ffiffiffiffiffiffiffiffi f=8 pPr2=31 ðÞ þ1:07 (9) For Re >10 6 , Petukhov–Popov–Kirilov correlation is used to calculate “Nu wall ”and “Nu bulk ”: Nu ¼f=8 ðÞ RePr 12:7ffiffiffiffiffiffiffiffi f=8 pPr2=31 ðÞ þ1:07 (10) where friction factor ‘f’is given by: f¼0:79 ln Re ðÞ 1:64 ðÞ 2(11) To estimate the heat transfer coefficient and pressure drop on air-side, the Colburn factor “j”and friction factor “f”are estimated. Pongsoi et al. (2013) suggested correlations to estimate these parameters for a fin-and-tube heat exchanger with spiral fins: j¼0:2150 Re0:4059 do(12) f¼0:4852 Re0:2156 dofp do  0:4771 (13) Evaporator Similar to the gas cooler, the evaporator is also modeled using a finite element approach. As the evaporator operates at subcritical pressure with a finite degree of superheat at Fig. 2. Schematic of (a) arrangement of tubes, and (b) single tube row of a spiral-fin-and-tube heat exchanger. Volume 0, Number 0, Month 2021 5 the exit, both single-phase and two-phase zones exist inside the evaporator tubes. Also depending upon the outer surface temperature, moisture from the air can condense (wet coil) or does not condense (dry coil). The heat and mass transfer phenomena occurring on the evaporator coil is modeled using Threlkeld (1970) method. Suitable equations are used to estimate the heat transfer coefficients and pressure drop in single as well as two-phase regions. Using the model, the latent and sensible heat transfer rates on the air side are estimated, which are required to calculate the dry-bulb temperature and moisture content of air at the evaporator exit. For each wet element, heat transfer rate is estimated from (Threlkeld 1970): DQ¼_ maha,ex ha,in ðÞ ¼_ mref href i,j,k ðÞ href iþ1, j,k ðÞ  ¼UA ðÞ wet Dh (14) where, Dh¼ ha,in hfictitiousjT¼Tref ði,j,kÞ  ha,exit hfictitiousjT¼Tref ðiþ1,j,kÞ  log ha,inhfictitiousjT¼Tref ði,j,kÞ ha,exithfictitiousjT¼Tref ðiþ1,j,kÞ (15) 1 UA ðÞ wet ¼bw goverall ao,wanode þbref aref ai (16) ao,w¼bwaa Cpa þyw kw (17) For each dry element, heat transfer rate is estimated from: DQ¼_ maha,ex ha,in ðÞ ¼_ mref href i,j,k ðÞ href iþ1, j,k ðÞ  ¼UA ðÞ dry DT (18) where, DT¼Ta,in Tref i,j,k ðÞ  Ta,exit Tref iþ1, j,k ðÞ  log Ta,inTref i,j,k ðÞ Ta,exitTref iþ1,j,k ðÞ hi(19) 1 UA ðÞ dry ¼1 goverall aaanode þ1 aref ai (20) To estimate the heat transfer coefficient and pressure drop during the phase change process, it is important to understand the two-phase flow characteristics through flow-pattern maps. Cheng, Ribatski, and Thome (2008) and Cheng, Ribatski, Moreno Quib en, et al. (2008) developed a flowpattern map of CO 2 to model correlations for frictional pressure drop and heat transfer coefficient during evaporation. These correlations are used in this study to estimate the heat transfer coefficient and pressure drop. The general expression to estimate the heat transfer coefficient in two-phase flow: ai¼atp ¼hdryagþ2phdry  awet 2p(21) where h dry represents the portion in the tube covered with vapor. For annular, intermittent and bubbly flow, the tube is completely covered with liquid layer and hence h dry ¼0. h dry changes from zero to its maximum value for stratifiedwavy flow and h dry ¼h strat for stratified flow. For the portion of the tube covered with vapor, heat transfer coefficient is estimated from: ag¼0:023Re0:8 gPr0:4 g kg di (22) For the portion of the tube covered with a liquid layer, the heat transfer coefficient is estimated from: awet ¼SaNB ðÞ 3þa3 CB hi 13 = (23a) where, aNB ¼131P0:0063 log P ðÞ 0:55M0:5q}0:58 (23b) aCB ¼0:0133Re0:69 dPr0:4 l kl dfilm (23c) S¼ 1, if x <xIA 11:14 di 0:00753  21d dIA  2:2,if x xIA 8 < : (23d) For mist flow, the heat transfer coefficient is estimated from: ai¼aM¼2108Re1:97 HPr1:06 gY1:83 kg di (24a) where, Y¼10:1ql qg 1  1x ðÞ  0:4(24b) In the dryout region, the heat transfer coefficient is estimated from: ai¼adryout ¼atp xdi ðÞ xxdi xde xdi atp xdi ðÞ aMxde ðÞ½ (25) atp xdi ðÞ is the heat transfer coefficient estimated at the dryout inception quality with Equation 21 and aMxde ðÞ is the heat transfer coefficient estimated at the dryout completion quality with Equation 24a. The total pressure drop of CO 2 during evaporation is estimated from: DPtotal ¼DPstatic þDPmomentum þDPfrictional (26) where, DPstatic ¼0 (27) DPmomentum ¼G21x ðÞ 2 ql1n ðÞ þx2 qgn "# out 1x ðÞ 2 ql1n ðÞ þx2 qgn "# in 8 < :9 = ; (28) To estimate the frictional pressure drop, different models are proposed for different flow regimes and are mentioned here. For annular flow: DPfrictional ¼DPA¼4fA dL di qgu2 g 2(29a) fA¼3:128Re0:454 gWe0:0308 l(29b) For slug and intermittent flow: 6Science and Technology for the Built Environment DPfrictional ¼DPslugþI¼DPLO 1n nIA  þDPA n nIA  (30a) where, DPLO ¼4fLO dL di G2 2ql (30b) fLO ¼0:079 Re0:25 LO (30c) For mist flow: DPfrictional ¼DPM¼4fM dL di G2 2qH (31a) fM¼91:2 Re0:832 M (31b) In the dryout region: DPfrictional ¼DPdryout ¼DPtp xdi ðÞ xxdi xde xdi DPtp xdi ðÞ DPMxde ðÞ  (32) where, DPtp xdi ðÞ and DPMxde ðÞ are frictional pressure drops estimated at the dryout inception quality and dryout completion quality, respectively. Void fraction is estimated from: n¼x qg"1þ0:12 1 þx ðÞðÞ x qg þ1x ðÞ ql ! þ 1:18 1 x ðÞ grq lqg ðÞ½ 1=4  Gq1=2 l#1(33) To evaluate the air-side heat transfer coefficient and pressure drop for the wet condition of the evaporator, Colburn factor “j”and friction factor “f”are estimated (Nuntaphan, Kiatsiriroat, and Wang 2005): j¼0:1970 Re0:1295 do ft fpft  0:1452 Pl Pt  1:1874 Pt do  0:8238 df do  0:0010 (34) f¼2:1768 Re0:2679 do ft fpft  0:2468 Pl Pt  1:8680 Pt do  0:3011 df do  0:4470 (35) Expansion valves Both the electronic expansion valves are driven by feedback controllers to maintain the desired parameters. The feedback controller drives a stepper motor to adjust the longitudinal movement of a needle to maintain the desired valve opening. The expansion valves are modeled assuming an isenthalpic expansion process and the refrigerant flow rate is estimated using the flow characteristic of the expansion valves. Calculation of total charge To estimate the total charge in the system, the charge distribution in all the components and tubes is estimated assuming it to be a pure refrigerant. The charge in the gas cooler is estimated from: Chargegc ¼Xqi,j,kacrossdLi,j,k(36) The charge in the evaporator is estimated from: Chargeev ¼Charge2/þCharge1/(37) Charge2/¼Xqgnþql1n ðÞ  i,j,kacrossdLi,j,k(38) Charge1/¼Xqi,j,kacrossdLi,j,k(39) The charge in the receiver is estimated from: Chargerec ¼qg Vrec Hrec,gþHrec,l  Hrec,g þql Vrec Hrec,gþHrec,l  Hrec,l; when partially filled with liquid (40) Chargerec ¼qlVrec;(41) when completely filled with liquid Chargerec ¼qgVrec;(42) when completely filled with vapor During simulations, constant superheat is maintained at the exit of the evaporator. Therefore, the charge in the accumulator is estimated from: Chargeacc ¼qaccVacc (43) The connecting tubes are considered as adiabatic to estimate the amount of charge present in the tubes. Solution procedure The simulation model for the entire system is developed by integrating the models for individual components on MATLAB (2008) platform. The refrigerant properties are obtained from REFPROP (2010) which is integrated with the MATLAB code. To estimate the refrigerant charge, the pressure and temperature are recorded initially when the system is in standstill (switch-off mode). As the internal volume is known, from the initial pressure and temperature, the refrigerant charge is estimated. Then numerical simulations are carried out to study the effect of charge. During simulations, the calculations are initialized based on the guessed values of pressures at the discharge and suction of the compressor for both configurations. These pressures are updated by an iterative procedure. The suction pressure is updated based on the requirement of the desired superheat for both Volume 0, Number 0, Month 2021 7 cases. For the configuration with single-stage expansion, the discharge pressure is updated to get the desired amount of charge in the system. On the other hand, for the configuration with two-stage expansion, the receiver is maintained at a pressure corresponding to the saturated liquid state after the first stage of expansion. Hence, the discharge pressure is updated based on the differential pressure drop across the first expansion valve. Results and discussion Effect of refrigerant charge Numerical simulations, as well as experimental tests, are performed to analyze the effect of refrigerant charge. In the experiments, the equalized pressure of the system when it is at a standstill condition (not in operation) is taken as an indication of the system charge. The standstill system pressure is recorded at identical surrounding temperatures (32 C). The test conditions are shown in Table 1. It may be mentioned here that in the experimental test-rig, modeled on standard laboratory test-rigs, the state of the air at the inlet of evaporator and gas cooler is identical as the same ambient air flows thorough both these components. The experimental results reported here are for monsoon conditions, hence the air at the inlet of the evaporator is close to saturation as shown in Table 1. Figure 3 shows the effect of refrigerant charge on gas cooler pressure and suction pressure. It is observed that for single-stage expansion, gas cooler pressure increases while suction pressure decreases with refrigerant charge. As the refrigerant charge increases, the charge accumulated in the gas cooler increases, and hence gas cooler pressure increases. With an increase in gas cooler pressure, there is an improvement in refrigerant quality at the inlet of the evaporator. This results in the requirement of a lesser refrigerant flow rate to achieve the desired superheat. Hence the flow area of the expansion valve opening decreases resulting in lesser refrigerant flow rate and lower suction pressure. On the other hand, with two-stage expansion, gas cooler pressure and suction pressure remain constant as long as the total charge is within a specific range. This is occurring because of the presence of the receiver which acts as a charge buffer. During operation, the receiver holds the charge in the form of liquid and gas. This condition must be fulfilled during operation to maintain the desired gas cooler pressure and suction pressure. There exists a range of charge over which this condition can be maintained to keep the receiver partially filled with liquid. Any variation in charge within this range will have a negligible effect on the performance of the system; as a result, the desired gas cooler pressure and the degree of superheat are effectively maintained by two EEVs. However, any change in charge beyond this range will affect the performance of the system, and no longer will the desired gas cooler pressure and the degree of superheat be maintained. For a charge higher than the upper charge limit, the liquid refrigerant completely occupies the receiver and its pressure starts increasing beyond the desired limit. While for a charge lesser than the lower charge limit, the liquid refrigerant completely disappears in the receiver and its pressure starts decreasing beyond the desired limit. With the deviation of receiver pressure from the desired Table 1. Ambient conditions, degree of superheat, amount of charge, and standstill system pressure for experimental tests. Standstill system pressure (bar) at 32 C Amount of charge (g) Degree of superheat (K) Expansion stages Air inlet temperature Gas cooler Evaporator DBT (C) DBT (C) WBT (C) 57.0 940 Double 28.61 29.74 28.88 60.2 1090 Double 28.73 29.93 28.77 61.4 1170 Double 28.65 29.74 28.84 57.0 940 12 Single 28.72 30.02 29.05 60.2 1090 Single 28.88 29.84 28.62 61.4 1170 Single 28.75 29.81 28.66 63.1 1250 Single 28.85 29.99 28.70 Fig. 3. Effect of refrigerant charge on gas cooler pressure and suction pressure. 8Science and Technology for the Built Environment