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Performance improvements evaluation of an automobile air conditioning system using CO2-propane mixture as a refrigerant
Accepted Manuscript
Performance improvements evaluation of an automobile air conditioning system using CO2 -propane mixture as a refrigerant Binbin YU , Dandong WANG , Cichong LIU , Fuzheng JIANG , Junye Shi , Jiangping CHEN PII: DOI: Reference:
S0140-7007(18)30001-X 10.1016/j.ijrefrig.2017.12.016 JIJR 3858
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
24 October 2017 3 December 2017 30 December 2017
Please cite this article as: Binbin YU , Dandong WANG , Cichong LIU , Fuzheng JIANG , Junye Shi , Jiangping CHEN , Performance improvements evaluation of an automobile air conditioning system using CO2 -propane mixture as a refrigerant, International Journal of Refrigeration (2018), doi: 10.1016/j.ijrefrig.2017.12.016
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ACCEPTED MANUSCRIPT Highlights Performance of pure CO2 AMAC system is enhanced using CO2-propane mixture as a refrigerant. Effects of operating parameters on system performance have been investigated. Additional performance evaluation has been experimentally carried out with constant cooling capacity. A new optimum high pressure control algorithm has been proposed within a deviation of 5%.
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Performance improvements evaluation of an automobile air conditioning system using CO2-propane mixture as a refrigerant
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Binbin YUa, Dandong WANGa, Cichong LIUa, Fuzheng JIANGa, Junye Shia,b, Jiangping CHENa,b* a. Institute of Refrigeration and Cryogenics, Shanghai Jiaotong University, Shanghai, China b. Shanghai High Efficiency Cooling System Research Center, Shanghai, China *corresponding author: [email protected]. +(86)21 34206775
Abstract
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The main purpose of this work is to enhance the energy efficiency of CO2 automobile air conditioning system. Theoretical analysis demonstrated that the mixture of CO2 and propane can improve its performance, thus, experiments have been carried out to see effects of various CO2-propane mass fractions of 100/0, 90/10, 80/20, 70/30, 60/40, 50/50 on the system performance at different ambient temperatures and gas cooler frontal air velocities. Experimental results show similar trends with those from the theoretical results. It has been shown that under the same compressor speed, system COP reaches highest at 60% of CO2 mass fraction, which is 29.4% higher than pure CO2 system and even achieves equal level of the R134a system, the optimum pressure and discharge temperature are reduced up to a maximum of 40% and 47℃ during the research range. Furthermore, comparison was carried out under the same cooling capacity by adjusting compressor speed for different mass fraction of CO2, results demonstrate that the use of CO2-propane mixtures yields a maximum COP rise of 22%even when cooling capacity is kept constant. A new optimum high pressure control algorithm for the transcritical CO2-propane mixture cycle has been developed based on the experimental data within a deviation of 5%.
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Keywords: Zeotropic mixtures; CO2; propane; Automobile; Air conditioning; Performance
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Accumulator Coefficient of performance Specific heat (J kg-1 K-1) Critical pressure Critical temperature Critical density Electrical expansion valve Electric vehicle Gas cooler Enthalpy (kJ kg-1) Specific irreversibility (kJ kg-1) Internal combustion engine vehicle Internal heat exchanger Lower explosion limit Mass (g) Molar mass (g/mol) Automobile air conditioning National Institute of Standards and Technology Pressure (MPa) Specific cooling capacity (kJ kg-1) Cooling capacity (kW) Revolutions per minute Entropy (kJ kg-1 K-1) Temperature (℃) Specific work (kJ kg-1) Work（kW） Mass fraction of Upper explosion limit Volume flow rate (m3/h) Velocity（m/s） Molar volume of gas (L/mol)
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Nomenclature Accu COP Cp CP CT CD EXV EV G/C h i ICEV IHX LEL m M AMAC NIST P q Q RPM s T w W x UEL u v Vm Greek letters
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Effectiveness Pressure ration Volume ration of CO2 and propane
Subscripts a con dis env evp
Air Condenser Discharge Environment Evaporator
ACCEPTED MANUSCRIPT Experiment Gas cooler Inlet Mixture Outlet Optimum Pure Predict Refrigerant Saturation Suction Saturation vapor
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exp gc i mix o opt pur pre r Sa suc Sv
1 Introduction
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After the approval of Kigali amendment in the 28th Conference of the Parties to the Montreal Protocol, a historic limitation of greenhouse gas hydrofluorocarbons (HFCs) was reached (montreal-protocol, 2016). The widely used HFC refrigerant R134a in automobile air conditioning system is not questioned to be limited and banned gradually, interests for CO2 as an AMAC refrigerant have been promoted a step forward. The incomparable heating performance of CO2 heat pump has been experimentally demonstrated (Wang et al, 2017). The European automotive manufacture has announced to offer CO2 air conditioning system in production (Daimler, 2017). The performance of different components (Ayad et al., 2012; Baek et al., 2013; Pettersen et al., 1998) and AMAC systems (Kim et al., 2009a; Liu et al., 2005; Martin et al., 2005; Yang et al., 1998) demonstrate that the COP of a CO2 air conditioning system is lower in the 10% usage conditions at high ambient temperature (above 30℃). Brown et al., 2002 evaluated the performance merits of CO2 and R134a AMAC systems. The results from semi-theoretical models show that the COP of CO2 was lower by 21% at 32.2℃ and by 34% at 48.9℃. The COP disparity was even greater at high speeds and ambient temperatures. For the transcritical CO2 systems, many researchers have devoted to improving their efficiency through various methods. Aprea and Maiorino (2008), Torrella et al. (2011) and Sanchez et al. (2014) demonstrated that the use of internal heat exchanger (IHX) can improve the COP up to a maximum of 10%. Kawamoto et al. (2017), He et al. (2014) and Lee et al. (2014) investigated the ejectors to improve COP by reducing expansion loss. The use of expanders (Li et al., 2004; Yang et al., 2007) to reduce irreversibilities during the expansion process also offered COP increments. Llopis et al. (2015) applied a dedicated mechanical subcooling system, theoretical results showed 9.5%, 13.5% and 13.1% enhancements of COP at evaporation levels of 5 °C, -5 °C and -30 °C, respectively. Inagaki et al. (1997) found that the COP of a CO2 air conditioning system for moderate ambient temperature was improved by 20% using a two-stage split cycle, while 5% for higher ambient temperature conditions. Another improvement is to install a gas liquid separator downstream of the expansion device, bypass the vapor around the evaporator, feeding saturated liquid into the evaporator to eliminate the problems of two-phase distribution (Kim et al., 2004). Any new technology must demonstrate that it is workable in specific application areas, as for AMAC system, in the long run, however, the technologies above may be unrealistic in performance, size, weight, economic, etc. Another solution to improve the performance of CO2 systems is via using zeotropic mixture as a refrigerant such as CO2-propane. Propane is the refrigerant having a higher refrigerating effect and a much lower vapor
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density than CO2 (Kim and Kim, 2005). Thus, one can easily expect that adding propane to CO2 improves system efficiency. The earliest application analysis of zeotropic mixtures was in the autocascade system, Kim et al. (2002) investigated the performance of an autocascade refrigeration system using R744/134a and R744/R290 by experiment and simulation, it is found that the cycle has a merit of low operating pressure and small amount of charge. Kim et al. (2005) evaluated experimentally the performance potentials of CO2-propane mixtures in a vapor compression system using water as secondary heat transfer fluid. Results showed that the discharge pressures of CO2-propane mixtures of 85/15, 75/25 and 60/40 by charged mass percentage are 11.8%, 22.7% and 37.9% less than that of CO2, respectively. The COPs of CO2-propane mixtures of 85/15, 75/25 and60/40 are enhanced by 8.0%, 12.8% and 12.5%, respectively. CO2-propane mixtures of 85/15, 75/25 and 60/40 have lower cooling capacities in comparison with CO2, that is, 93.0%, 85.6% and 68.2% of that of CO2, respectively. In this study, safety issues and theoretical analysis of CO2-propane mixtures are evaluated, the possibilities of performance improvements of various CO2-propane mass fractions of 100/0, 90/10, 80/20, 70/30, 60/40 and 50/50 have been carried out experimentally to see their effects on the AMAC system performance varying with different ambient temperatures and gas cooler frontal air velocities. Furthermore, additional comparison was carried out under the same cooling capacity with pure CO2 system by adjusting compressor speed for different mass fraction of CO2. Based on the experimental data, a new optimum high pressure control algorithm for the transcritical CO2-propane mixture cycle has been developed within a deviation of 5%.
2 Operation of the CO2-propane mixture refrigeration cycle
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2.1 Security The addition of an inert gas suppresses the flammability of combustible gases, dilutes the concentration of combustible gas and narrows the range of explosive limits. Literature gives the formula to calculate the explosion limit (Niu, 2006). Lower explosion limit：
LEL 2.21995 2.23676 1.15607 2 0.96581 3 0.30005 4 0.03886 5 0.00181 6 (1)
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（ 0 8.21 ） Upper explosion limit：
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UEL 8.99719 6.01554 0.49461 2 0.01209 3 （ 0 8.21 ）
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𝐿𝐸𝐿(𝑈𝐸𝐿)𝑔/𝑚3
1000 × 𝐿𝐸𝐿(𝑈𝐸𝐿) = ×𝑀 𝑉𝑚
(2) (3)
Fig. 1 shows the explosion limit of CO2-propane mixtures with respect to CO2 mass fraction. The explosion limit of pure R290 is 2.20%/9.50%. When the mass fraction of R744 are 30% and 50% respectively, the explosion limit are 3.32%/12.03% and 4.86%/12.8%, respectively; According to ASHRAE34-2013 on the safety of refrigerant classification, when the mass fraction of the CO2 is above 60%, the lower explosion limit of mixed refrigerant is greater than 100 g/m3, the safety level rises from A3 to A2, when the mass fraction of the CO2 is equal to 89%, the upper explosion limit is equal to the lower explosion limit, when the mass fraction of the CO2 is greater than 89%, the corresponding security level rises from A2 to Al level, the mixture is already non-combustible gas.
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Fig. 1 Explosion limit of CO2-propane mixtures with respect to CO2 mass fraction.
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2.2. Temperature–composition relations for CO2-propane mixtures at a fixed pressure Fig. 2 is a typical temperature–composition diagram for CO2-propane mixtures at a fixed pressure of 4MPa with two lines and three regions, the dew point of the mixed refrigerant with CO2 mass fraction varying from 0 to 1 constitutes the dew line, and the bubble point constitutes the bubble line, this two lines divide the diagram into three regions: subcooled liquid region, superheated vapor region and two-phase region. As shown in Fig.2, the dew point temperature of the mixed refrigerant with a certain CO2 mass fraction is higher than the bubble point temperature, just because of this, the evaporation or condensing process in a zeotropic refrigerant system is no longer isothermal, but a process with temperature glide. When the CO2-propane mixture reaches the gas-liquid equilibrium, the gas phase consists of a large part of CO2 because of its volatility and low boiling temperature, while most of propane will stay in the liquid phase.
Figure.2 Temperature–composition diagram of CO2-propane mixtures at a fixed pressure
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2.3 Variation of critical properties and saturation vapor pressure Fig. 3 shows that with the decrease of mass fraction of CO2, the critical pressure and critical density decrease to a large extent, when the critical temperature increases gradually. Fig. 4 gives the CO2-propane mixtures bubble point pressure and pure refrigerants saturation vapor pressure varying with temperature. As it can be seen, when the mass fraction of CO2 decreases, the saturation pressure shows a downward trend, thus using CO2-propane mixtures can contribute a lot to reduce system pressure.
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Figure.3 Critical properties of CO2-propane mixtures varying with CO2 mass fraction
Figure.4 Saturation vapor pressure of CO2-propane mixtures varying with temperature
2.4 Theoretical analysis To evaluate the possibilities of improving the performance of AMAC system using CO2-propane mixture as a refrigerant, we consider a simple theoretical analysis using the thermodynamic properties from NIST Refprop 9 database (Lemmon et al., 2010). Fig. 5 shows the pressure-enthalpy (a) and temperature-entropy (b) diagram of CO2-propane mixture refrigeration cycle varying with mass fraction of CO2, it is obvious that the system
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gradually moves from the subcritical cycle to the transcritical cycle with increasing of CO2 mass fraction in the CO2-propane mixture. The following assumptions are made for the analysis of CO2-propane mixture system: 1) All components are assumed to be a steady-state and steady-flow process; 2) All pressure drops and heat losses to the environment are neglected; 3) The throttling processes are isenthalpic; 4) The evaporation temperature and a constant compressor suction superheat of 5℃ are fixed; 5) The gas cooler outlet temperature for transcritical cycle or the condensing temperature for subcritical cycle are obtained by an approach temperature of 5℃ to the environment as Equation (4); 6) The compression process is irreversible and the isentropic efficiency related to its pressure ratio is taken into account, Equation (5) of compressor isentropic efficiency was used (Liao et al., 2000).
Tgc,o / con Tenv 5℃
is 1.003 0.121
(4) (5)
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For the transcritical cycle, all the calculations are carried out for the optimum high pressure through an iteration process. The specific cooling capacity is obtained with Equation (6) considering the equal enthalpy difference between the high temperature side and low side of IHX. The specific compressor work is calculated by Equation (7), where the hdis, s represents discharge isentropic enthalpy, and the COP is obtained by Equation (8).
hdis,s hsuc q w
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q hsuc hgc,o
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(7)
(8)
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（a）
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（b） Fig. 5 Pressure-enthalpy (a) and Temperature-entropy (b) diagram of CO2-propane mixture refrigeration cycle. The theoretical results of CO2-propane mixture refrigeration cycle are shown in Fig. 6 with mass fraction of CO2 varying from 1 to 0.5. It is demonstrated that the use of CO2-propane can continuously improve the COP of pure CO2 system with reduction of the CO2 mass fraction, when the CO2 mass fraction is 0.5, both discharge pressure and temperature drops significantly to a maximum value of 48% and 20%, respectively. Additionally, CO2-propane mixtures have lower cooling capacity compared with pure CO2. The theoretical results obtained are undoubtedly of great value for the experiments in following sections.
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Fig. 6 Theoretical results with respect to mass fraction of CO2.
3 Experiments
3.1. Experimental setup description The Schematic diagram of the experimental CO2-propane mixture AC system and test rig for vehicle is shown in Fig. 7. The AC system consists of two parallel flow type micro-channel heat exchangers involving an
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evaporator and a gas cooler, a plate IHX, a rotary type compressor and an electrical expansion valve. All components were successfully designed to bear the extremely high pressure due to the high operating pressure of CO 2 system, we designed a plate IHX which can bear the extreme burst pressure of 100MPa in our test and both microchannel heat
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exchangers can withstand a pressure of 30Mpa with our new structure and design. The specifications of the
CO2-propane mixture AC components are listed in Table 1. The compressor, with a displacement volume of 6 cm3 rev-1, is a fixed displacement and variable speed type driven by a variable frequency inverter which can be adjusted
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from 30hz to 100hz. The maximum allowable compressor discharge temperature and pressure are 120℃ and 12MPa, respectively. The gas cooler and evaporator in this system are both microchannel heat exchagers made of aluminum
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with louvered fin, their core size are 540mm, 325mm, 12.5mm and 275mm, 235mm, 36mmin width, length and depth, respectively. The internal heat exchanger is a stainless steel plate heat exchanger with a core size of
111W*310L*28.2D (mm). An electric expansion valve that can be controlled from 0-500 steps with step motor
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controller was installed as expansion device, with which the system mass flow rate and high pressure can be precisely regulated. Additionally, an accumulator with a volume of 600ml was installed after evaporator to serve as a system refrigerant charge buffer. Fig. 7 also shows the layout of enthalpy potential method for the system performance test. The calorimeter facility
consists of two chambers with a wind tunnel in each side. The left one installed with a gas cooler was used to simulate the environmental conditions inside the cabin and the other installed with an evaporator was used to modulate the environmental conditions outside the vehicle. Both chambers were equipped with a blower, a cooling system, a PTC heating unit and humidity controller to adjust the air flow rate, ambient temperatures and relative humidity according to the requested operating conditions. Measurement devices showed in Fig. 7 are instrumented to measure the pressure and temperature of refrigerant and air. Refrigerant side temperatures are measured by platinum resistance temperature sensors with an uncertainty of ±0.2°C. The environmental dry bulb and wet bulb temperature can be controlled within ±0.2°C. Refrigerant side absolute pressure and pressure drop of air loops are measured by piezoresistive type transducer
ACCEPTED MANUSCRIPT with an uncertainty of ± 0.5%. The system mass flow rate is accurately measured by a Coriolis type flow meter with an accuracy of ±0.5%. The electric compressor power input is measured by digital power meter with an accuracy of ±0.5%including the inverter loss. Table 3 presents the accuracy of the experimental variables measured by various
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Fig. 7 Schematic diagram of the experimental CO2-propane mixture AMAC system and test rig. Table 1. The specifications of the CO2-propane mixture AMAC components Specifications
Compressor
DC-driven, rolling rotor type, variable speed (1800-6000rpm), Displacement: 6 cc/rev Microchannel, Parallel flow type, 4-pass, 2-slab, Core size: 275W*235L*36D(mm) Microchannel, Parallel flow type, 3-pass, 1-slab, Core size: 540W*325L*12.5D(mm) Plate type heat exchanger, Stainless steel, Size: 111W*310L*28.2D(mm) EXV, copper, orifice diameter:1.4mm, driven by step motor 600ml
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Gas cooler
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Evaporator
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Components
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IHX Expansion device Accumulator
3.2. Test method and Test conditions To evaluate the performance characteristics of CO2-propane mixture AC system, several CO2 mass fractions in the mixtures have been test at various operating conditions showed in Table 2. The evaluated test methods and conditions are: Proper charge amount: the system charge amount was firstly adjusted to reach the maximum COP and compressor suction superheat was ensured above 5℃ to avoid liquid shock, and then all experiments were carried out at this charge amount.
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Outdoor air velocity: the system has been tested at four different gas cooler frontal air velocities: 1.5, 2, 3 and 4.5m/s. These velocities are supplied by the blower in the outdoor chamber to simulate different driving speed, 1.5 and 2m/s for idling condition, 3 and 4.5m/s for normal driving condition. Ambient temperature: the system has been tested at four different outdoor temperatures: 27, 35, 42, 45℃, when the indoor temperature keeps at 27℃. An extreme high temperature condition of both 40℃ for indoor and outdoor has also been considered. EXV opening adjustment: all test conditions have been performed within a wide range of gas cooler pressure to reach the optimum COP point by manually adjusting the EXV opening. Compressor speed: for condition 1-3 in Table 2, the compressor speed has been kept at 4800rpm, which can supply sufficient cooling capacity for pure CO2 system. However, theoretical results show that the cooling capacity of CO2-propane AC system decreases with reduction of CO2 mass fraction, thus, the condition 4 in Table 2 has been designed to evaluate the CO2-propane AC system performance at the same cooling capacity of pure CO2 system.
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Table 2. Experimental conditions Ta,e,in(℃)*
1
27 100%,90%,80%, 70%,60%,50%
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ua,e,in（m3.h-1） Ta,gc,in（℃）
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27
va,gc,in（m/s） RPM
420
35
1.5,2,3,4.5
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420
27,35,42,45
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4800
420
40
2,4.5
4800
420
27,35,42,45
2
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*：the relative humidity was controlled at 50%. “/” represents that the compressor speed is adjusted to keep the cooling capacity constant with the pure CO2 system. The cooling capacity Q of the experiments is calculated by Equation (9), where the Qa representing air side heat
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transfer is obtained by the heat balanced method using Equation (10), the Qr representing refrigerant side heat transfer is obtained by the enthalpy difference method using Equation (11), then the overall COP is determined by Equation (12). Modifications of COP due to the use of CO2-propane mixtures are evaluated as a percentage variation expressed by
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Equation (13). The relative uncertainty of the cooling capacity and COP of the system are calculated as5.5% and 6.3%, respectively, by the sequential perturbation method (Moffat, 1988).
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Q (Qa Qr ) / 2
Qa maC p (Ta,o Ta,i )
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COPr （COPmix COPpur）/ COPpur
(9) (10) (11) (12) (13)
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Table 3. Accuracy of the experimental parameters with various sensor devices Device
Range
Accuracy
Temperature Pressure Mass flow rate Compressor work Cooling capacity COP
RTD-type temperature sensors Pressure transducers
-50~150℃ 0~10MPa/0~20MPa 0~600kg.h 0~6kw / /
±0.2℃ ±0.5% ±0.15% of reading ±0.5% of reading 5.5% 6.3%
Coriolis type flow meter
Digital power meter / /
4 Experimental results and discussion
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Variable
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4.1 System performance at various CO2 mass fractions 4.1.1 The influence of G/C frontal air velocity on system performance Figs. 8-12 show the experimental results for condition 1 in Table 2. It can be observed in Fig. 8 that at 35/27 condition, the COP enhancement reaches highest when the mass fraction of CO2 is 60% for all G/C frontal air velocities. When the G/C frontal air velocity is 4.5m/s, the exact COPs of CO2 mass fraction of 100%, 90%, 80%, 70%, 60%, 50% are1.81, 1.90, 2.07, 2.17, 2.35 and 2.32, the corresponding percent increments are 4.8%, 14.5%, 20% and 29.4%, 28.3%, relatively. The COP trend can be explained by the decreasing of specific compressor work and specific capacity when the mass fraction of CO2 reduces, and the compressor work decreases more quickly, so the system COPs increase gradually. The COP enhancement by using CO2-propane mixture will contribute a lot to improve fuel economy for ICEVs and extend driving range for EVs. However, the biggest difference between the theoretical and experimental results for COPs is that the COP of CO2 mass fraction of 50% is a little lower than that of 60%, this trend may result from the low compressor efficiency and the poor matching of temperature profiles between CO2-propane mixture of 50/50 and air, and for this study, the detailed reason will be given in subsection of 4.1.3. Fig. 9 and Fig. 10 show impacts of G/C frontal air velocity on the optimum high pressure and discharge temperature of CO2-propane mixture system varying with mass fraction of CO2. The optimum pressures are reduced by 40% from 10.4MPa to 6.3MPa and discharge temperature by 47℃ from 114℃ to 67℃ when the mass fraction of CO2 is reduced to 50%, on the average for different air velocities. For the maximum COP point, the optimum pressure is 7MPa and discharge temperature is 75℃. The reduction of operating pressure and discharge temperature will effectively release the CO2 compressor load such as high pressure leakage and thermal deformation.
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Fig. 8 Impacts of G/C frontal air velocity on the COP of CO2-propane mixture system varying with mass fraction of CO2.
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Fig. 9 Impacts of G/C frontal air velocity on the optimum high pressure of CO2-propane mixture system varying with mass fraction of CO2.
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Fig. 10 Impacts of G/C frontal air velocity on the discharge temperature of CO2-propane mixture system varying with mass fraction of CO2.
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Fig. 11 shows that the cooling capacity decreases with reduction of CO2 mass fraction, and the slope of cooling capacity increases with reducing CO2 mass fraction, results show that the CO2 mass fraction of 70% is a significant point, for if the CO2 mass fraction is higher than this point, the cooling capacity declines slightly only within a 5.7% level on the average of different air velocities, but when the CO2 mass fraction is smaller than 70%, the cooling capacity declines appreciably by an average of 38%. Thus, the balance between cooling capacity and COP at different air velocities should be weighed deliberately. Fig. 12 shows other significant parameters, besides COP, cooling capacity, Popt, Tdis, which are described before, as the mass fractions of CO2 decrease, the compressor work reduce almost linearly to the reduction of CO2 mass fraction, refrigerant mass flow rates are reduced by 30% on the average as the suction density and volume efficiency reduce simultaneously, and the gas cooler refrigerant outlet temperatures are almost kept constant.
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Fig. 11 Impacts of G/C frontal air velocity on the cooling capacity of CO2-propane mixture system varying with mass fraction of CO2.
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(b) Fig. 12 Various parameters of CO2-propane mixture system varying with mass fraction of CO2 (Ta,ev,i=27℃, Ta,gc,i=35℃, va,gc,i=4.5m/s).
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4.1.2 The influence of ambient temperatures on system performance Figs. 13 and Figs. 14 show the experimental results for condition 2 in Table 2. It is clear that the COP and cooling capacity decrease with the increasing outdoor temperature. Figs. 13 shows that at the outdoor temperature of 27℃, COPs of CO2-propane mixtures of 90%, 80%, 70%, 60%, 50% are 2.05, 2.27, 2.49, 2.64 and 2.66, the corresponding percent increments are 0.5%, 11.3%, 22.1%, 29.4%, 30.4%, respectively, but for outdoor temperature of 35℃, 42℃ and 45℃, the COPs of CO2-propane mixtures reach to a maximum value at the 60% of CO2 mass fraction, which are higher than pure CO2 by 26.66%, 24.87% and 19.2%, respectively. The impacts of outdoor temperature on the cooling capacity varying with mass fraction of CO2 is similar to that of gas cooler frontal air velocity illustrated in subsection 2.1.1. As shown in Fig. 14, the cooling capacity decreases with reduction of CO2 mass fraction, when the CO2 mass fraction is higher than 70%, the cooling capacity declines slightly only within a 4.8% level on the average of different outdoor temperatures, but when the CO 2 mass fraction is smaller than 70%, the cooling capacity declines appreciably by 41% on the average. The balance between cooling capacity and COP at different outdoor temperature should be weighed deliberately.
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Fig. 13 Impacts of outdoor temperature on the COP of CO2-propane mixture system varying with mass fraction of CO2 (va,gc,i=2m/s).
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Fig. 14 Impacts of outdoor temperature on the cooling capacity of CO2-propane mixture system varying with mass fraction of CO2 (va,gc,i=2m/s). At the condition of both 40℃ for outdoor and indoor temperature, it can be observed in Fig. 15 that COPs of CO2-propane mixtures increase continuously with the reduction of CO2 mass fraction, though the curves slope decrease gradually, the COP of 50% CO2 mass fraction is enhanced by 45.2% and 31.7% compared with pure CO2 at the gas cooler frontal air velocity of 2m/s and 4.5m/s, respectively. As the CO2 mass fraction varied from 100% to 50%, the cooling capacity is decreased by 19.8% and 39.7% at the gas cooler frontal air velocity of 2m/s and 4.5m/s, respectively.
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Fig. 25 Impacts of G/C frontal air velocity on the CO2-propane mixture system performance varying with mass fraction of CO2 at 40/40 condition.
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4.1.3 Analysis of COP loss at 50% of CO2 mass fraction in some conditions From the theoretical results of Fig. 6, we know that the use of CO2-propane mixtures can continuously improve the COP of pure CO2 system with reduction of CO2 mass fraction, but an interesting observation can be made in the experiments mentioned above is that the COP enhancement reaches highest when the mass fraction of CO2 is 60% except for the conditions of 27/27 and 40/40 for outdoor/ indoor temperature in Table 2. As shown in Fig. 16, the loss of compressor efficiency and different decline trends may account for this, when the mass fraction of CO2 reduces from 100% to 50%, the total compressor efficiency calculated by Equation(14) are decreased by 49.0%, 30.5% and 26.8% for 35/27, 40/40 and 27/27 condition, respectively. Therefore, due to the two contrary factors of higher propane refrigerating effect and compressor efficiency loss, the COP enhancements exist a maximum value for most conditions.
(14)
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Fig. 3 Impacts of ambient conditions on the total compressor efficiency varying with mass fraction of CO2.
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After expansion of CO2-propane mixture in the electric expansion valve, the working fluid turns into two-phase flow and enters the evaporator, at the evaporator inlet, the refrigerant qualities are between 0.3~0.4 during our research range, due to the zeotropic effect, there exists a temperature glide resulting from the temperature difference between the bubble point and dew point at the same CO2 mass fraction, and there are more CO2 in the gas phase because of its volatility and low boiling temperature than propane, this situation is contrary in the liquid phase. Kim illustrated the temperature glide effect of CO2-propane mixtures in his research, when the temperature glide occurring in the evaporator of CO2-propane mixture with reduction of CO2 fraction exceeds the temperature difference of secondary fluid, the mean temperature difference in the evaporator of CO2-propane mixture increases. This results in an increased temperature difference in the evaporator and reduced system efficiency. As a result, when zeotropic mixtures are used in applications with the same temperature gradients of two heat transfer fluids in heat exchangers, the benefits derived from temperature gliding effect can contribute to an improvement of the system capacity and efficiency (Kim et al ., 2008). For AMAC system using CO2-propane mixture as a refrigerant, since the heat exchanger area can not be changed, the entropy generation in the evaporator can be reduced if the temperature glides of the CO2-propane mixture and air are well matched, as shown by Equation (15), Equation (16-18) illustrated that the requirement of perfect glide matching can be achieved by adjusting the air flow rate through HVAC blower.
Qa Qr
mr c pr dtr ma c padta ma mr c pr / c pa
(15) (16) (17) (18)
4.2 Performance comparison at constant cooling capacity CO2-propane mixtures have lower cooling capacities in comparison with pure CO2 at constant compressor
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speed, thus, additional performance comparison at constant cooling capacity is essential to meet the comfort requirement of cabin. Fig.17 shows the impacts of mass fraction of CO2 on the COP at constant cooling capacity with pure CO2 system varying with outdoor temperatures, the cooling capacities of pure CO2 system are 4.2kw, 3.8kw, 3.4kw and 3.2kw for the outdoor temperature of 27℃, 35℃, 42℃ and 45℃, respectively. It is clear that COP improvements all reach highest at the CO2 mass fraction of 70% for different outdoor temperatures, that is, 22.1%, 18.6%, 19.6% and 18.4% of pure CO2 system, the corresponding exact COPs are 2.49, 1.92, 1.58 and 1.44, respectively. For CO2 mass fraction of 60%, its COP reduced by 16% on the average compared with that of constant compressor speed, but is still higher than pure CO2 by 8%. Therefore, if taking security, cooling capacity and COP into consideration, the best choice for mass fraction of CO2-propane mixtures is 70/30.
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Fig. 17 Impacts of mass fraction of CO2 on the COP at constant cooling capacity with pure CO2system(Ta,ev,i=27℃, va,gc,i=2m/s).
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4.3 Correlation for predicting optimum high pressure for the transcritical CO2-propane mixture cycle 4.3.1 Verification of existing pure CO2 system correlations University of Illinois at Urbana-Champaign (UIUC) did experimental investigation and model analysis of the control and operating parameters of a transcritical CO2 mobile A/C system under idle and medium driving operating conditions. The results demonstrated that the COP of the transcritical CO2 system is a strong function of the gas cooling pressure. Furthermore, the strong relation between the optimum pressure and the gas cooler outlet temperature was revealed, which suggests a simple controller of only one parameter for the transcritical CO2cycle (Park YC et al ., 1999). So far there is no optimum high pressure correlation for CO2-propane mixtures, the available correlations from open literature are all developed for pure CO2 as listed in Table 4, the errors of these correlations for predicting the optimum high pressure of transcritical CO2-propane mixtures cycle are verified in Table 4 by experimental data.
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Table 4 Summary of existing optimum high pressure correlations with only one parameter Literature
Correlation
Deviation(%)
Popt 1.938Tgc,r ,o 9.872
Yang et al.
-6~15
Popt 2.6Tgc,r ,o 7.54
Kauf
-26~-12
Popt 2.56Tgc,r ,o 4.2
Yahia et al.
-18~5
Qi, He, Wang and Meng
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Popt 2.68Tc 6.797
Chen and Gu
-20~2
Popt 132.2 8.4Tgc,r ,o 0.3Tgc2 ,r ,o 27.7 104 Tgc3 ,r ,o
-16~5
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4.3.2 Development of a new optimum high pressure control algorithm for transcritical CO2-propane mixture cycle By verification of the available correlations, the conclusion can be drawn that none of these existing correlations can predict the experimental data of CO2-propane mixtures satisfactorily as they were obtained from pure CO2 system. Therefore, it is necessary to develop a new optimum pressure control algorithm for transcritical CO2-propane mixture system. In the present study, a new correlation was proposed based on the existing model by considering strong effect of gas cooler outlet temperature and effect of CO2-propane mixtures, as shown in Equation (19), where the CO2 mass fraction is limited to ensure transcritical cycle. Table 5 lists all the 92 validation data of the new correlation, and Fig. 18 presents the comparisons of the experimental data with the predicted values of the new correlation and literature. It can be seen that the new correlation predictions agree with all of the experimental data within a deviation of ±5%.
Popt （MPa） 0.16629Tgc,out （℃） 3.20186 xCO2 （ 0.7 xCO2 1 ）
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Table 5 validation of the new developed optimum high pressure control algorithm
x
Tgc,out
Pexp,opt
Ppre,opt
error
45.8
10.83
10.81
0.20%
0.8
45.2
9.90
10.08
-1.85%
1
44.5
10.62
10.61
0.11%
0.8
43.1
9.38
9.72
-3.47%
1
42.3
10.30
10.24
0.56%
0.8
40.7
9.14
9.32
-1.97%
1
40.8
10.00
9.99
0.03%
0.8
38.5
8.90
8.97
-0.83%
1
45.3
10.80
10.73
0.61%
0.8
45.1
9.95
10.06
-1.10%
1
44.2
10.62
10.55
0.66%
0.8
42.8
9.43
9.68
-2.57%
1
42.0
10.30
10.19
1.07%
0.8
40.6
9.12
9.31
-2.07%
1
40.5
10.00
9.94
0.60%
0.8
38.3
8.92
8.93
-0.12%
1
38.2
9.48
9.55
-0.76%
0.8
35.8
8.5
8.52
-0.52%
1
42.1
10.29
10.20
0.90%
0.8
37.2
8.93
8.75
2.09%
CO2
1
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CO2-propane
CO2
Tgc,out
Pexp,opt Ppre,opt
error
10.62
10.61
0.11%
0.8
43.1
9.4
9.72
-3.47%
1
48.8
11.27
11.31
-0.41%
0.8
46.8
9.98
10.34
-3.52%
1
50.3
11.42
11.57
-1.28%
0.8
50.1
10.5
10.89
-3.38%
1
53.0
11.75
12.02
-2.24%
0.8
52.1
10.8
11.22
-4.23%
1
37.9
9.48
9.50
-0.26%
0.8
35.6
8.53
8.48
0.57%
1
42.0
10.29
10.19
1.01%
0.8
37.5
8.92
8.80
1.39%
1
44.1
10.62
10.54
0.82%
0.8
43.3
9.46
9.76
-3.09%
1
48.2
11.27
11.22
0.43%
0.8
47
10.52
10.38
1.38%
1
50.6
11.42
11.62
-1.71%
0.8
49.8
10.52
10.84
-2.98%
1
52.8
11.75
11.98
-1.93%
0.8
51.8
10.83
11.18
-3.09%
1
52.5
12.08
11.93
1.27%
0.8
49.1
11.3
10.72
4.91%
1
45.8
10.91
10.82
0.84%
0.8
44.1
10.1
9.89
1.84%
1
52.2
12.08
11.88
1.70%
0.8
48.9
11.23
10.69
5.02%
1
45.8
10.91
10.82
0.84%
0.8
44.0
10.13
9.88
2.55%
0.9
45
10.70
10.36
3.23%
0.7
45.1
9.2
9.75
-5.74%
0.9
44.1
10.44
10.22
2.16%
0.7
41.8
8.9
9.20
-2.99%
0.9
41.6
9.96
9.81
1.59%
0.7
39.9
8.6
8.87
-2.54%
0.9
39.8
9.59
9.50
0.94%
0.7
37.8
8.4
8.53
-1.08%
0.9
44.8
10.70
10.33
3.57%
0.7
45
9.26
9.72
-4.78%
0.9
44.3
10.44
10.25
1.85%
0.7
41.6
8.95
9.16
-2.28%
0.9
41.8
9.96
9.83
1.31%
0.7
40.1
8.62
8.91
-3.25%
0.9
39.6
9.59
9.47
1.32%
0.7
37.6
8.43
8.49
-0.75%
0.9
41.9
9.92
9.84
0.77%
0.7
41.8
8.9
9.20
-2.99%
0.9
44.1
10.44
10.22
2.16%
0.7
47.6
9.8
10.16
-3.05%
0.9
48.7
11.08
10.98
0.92%
0.7
50.9
10.2
10.71
-4.30%
0.9
53.5
11.72
11.78
-0.47%
0.7
41.6
8.93
9.16
-2.50%
0.9
37.0
0.9
41.7
0.9
43.8
9.03
2.86%
0.7
47.5
9.82
10.14
-3.16%
9.92
9.82
1.03%
0.7
50.3
10.23
10.61
-3.54%
10.44
10.17
2.68%
0.7
48.1
9.8
10.24
-4.20%
48.3
11.08
10.91
1.51%
0.7
43.7
9.6
9.50
1.21%
52.1
11.51
11.55
-0.32%
0.7
48.3
9.82
10.27
-4.41%
0.9 0.9
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9.29
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0.9
53.4
11.72
11.76
-0.34%
0.7
43.2
9.63
9.43
2.17%
0.9
52.6
11.68
11.62
0.50%
0.7
47.6
9.9
10.16
-3.02%
0.9
45.2
10.87
10.40
4.58%
0.7
43.2
9.6
9.43
2.17%
0.9
52.4
11.68
11.60
0.75%
0.7
48.0
9.80
10.22
-4.14%
0.9
45.0
10.87
10.36
4.90%
0.7
43.5
9.65
9.47
1.85%
12
+5
%
11 -5 %
10
Yang Kauf Yahia Chen Qi Present
9 8 8
9
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Predicted optimum high pressure/MPa
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Experimental optimum high pressure/MPa
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Fig. 18 Comparison of predicted optimum high pressure with experimental data
5 Conclusion
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In present study, the thermodynamic simulation of CO2-propane binary mixture refrigeration cycle has been performed firstly. Next, based on the simulation results, the performance characteristics of an AMAC system using CO2-propane mixture as a refrigerant have been experimentally evaluated, experimental conditions include CO2-propane mass fractions of 100/0, 90/10, 80/20, 70/30, 60/40, a wide range of outdoor temperatures, from 27℃ to 45℃, and gas cooler frontal air velocity, from 1.5m/s to 4.5m/s. The work accomplished will be of great value for promotion of natural refrigerants to protect our environment. The main conclusions obtained are as follow: The AMAC system benefits greatly from using CO2-propane mixture as a refrigerant, under the condition of same compressor speed with pure CO2 system, COPs of CO2-propane mixtures reach highest at 60% of CO2 mass fraction, which is 29.4% higher than pure CO2 system and even has achieved equal level of the R134a prototype system, this will contribute a lot to improve fuel economy for ICEVs and extend driving range for EVs. The optimum pressure and discharge temperature are reduced up to a maximum of 40% and 47℃ during the research range, and this will effectively release the CO2 compressor load such as high pressure leakage and thermal deformation. Cooling capacity decreases with reduction of CO2 mass fraction, and the slope of cooling capacity increases with reducing CO2 mass fraction, results show that the CO2 mass fraction of 70% is a significant point, the cooling capacity declines slightly only within a 5.7% level if the CO2 mass fraction is higher than this point. Reducing CO2 mass fraction results in the decrease of CO2 compressor efficiency and temperature glide in the CO2 evaporator, and just because of this, the COP of 50% CO2 mass fraction is a little lower than that of 60% CO2 mass fraction, which is the biggest difference compared with thermodynamic analysis. In our research, for AMAC system this problem can be solved to some extent by adjusting evaporator air flow rate to match the temperature glide. COPs of CO2-propane mixtures are still higher than pure CO2 system even when the cooling capacities are kept constant, the CO2-propane mass fraction of 70/30 can improve the COP up to a maximum of 22%. Taking consideration of security, cooling capacity and COP, the best choice for mass fraction of CO2-propane mixtures is 70/30.
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6 Acknowledgements
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This project was supported by National Nature Science Foundation of China (NO.51776119).
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References Aprea, C., Greco, A., Maiorino, A., 2015. The application of a desiccant wheel to increase the energetic performances of a transcritical cycle. Energy Convers. Manag 89, 222–230. Aprea, C., Maiorino, A., 2008. An experimental evaluation of the transcritical CO2 refrigerator performances using an internal heat exchanger. Int. J. Refrigeration 31, 1006–1011. Dandong Wang, Binbin Yu, Jichao Hu, Liang Chen, Junye Shi, Jiangping Chen, 2017. Heating performance characteristics of CO2 heat pump system for electrical vehicle in a cold climate. Int. J. Refrigeration. http://dx.doi.org/doi:10.1016/j.ijrefrig.2017.09.009. Designation and Safety Classification of Refrigerants. ANSI/ASHRAE Standard 34–2016. Friedrich Kauf, 1999. Determination of the optimum high pressure for transcritical CO2-refrigeration cycles. Int. J. Therm. Sci. 38, 325–330. He, Y., Deng, J., Zhang, Z., 2014. Thermodynamic study on a new transcritical CO2 ejector expansion refrigeration system with two-stage evaporation and vapor feedback. HVAC&R Res. 20, 655–664. Lee, J.S., Kim, M.S., Kim, M.S., 2014. Studies on the performance of a CO2 air conditioning system using an ejector as an expansion device. Int. J. Refrigeration 38, 140–152. https//www.daimler.com/sustainability/product/further-environmental-technologies/co2-air-conditioning-system.ht ml (Daimler official website). http://conf.montreal-protocol.org/meeting/mop/mop-28/final-report/English/MOP28-Dec1.docx(Ozone Secretariat Conference Portal） Inagaki M, Sasaya H, Ozakli Y, 1997. Pointing to the future: two stage CO2 compression. Heat transfer issues in natural refrigerants, International Institute of Refrigeration, 131–140. J. Sarkar, S. Bhattacharyya, M.R. Gopal, 2004. Optimization of a transcritical CO2 heat pump cycle for simultaneous cooling and heating applications, Int. J. Refrigeration 27, 830–838. Ju Hyok Kim, Jin Min Cho, Min Soo Kim, Cooling performance of several CO2-propane mixtures and glide matching with secondary heat transfer fluid, Int. J. Refrig. 31 (2008) 800– 806. Kim, J.H., Kim, M.S., 2005. Vapor–liquid equilibria for the carbon dioxide propane system over a temperature range from 253.15 to 323.15 K. Fluid Phase Equilib. 238 (1), 13–19. Kawamoto, Y., Ogata, G., and Shan, Z., 2017. Ejector Energy-Saving Technology for Mobile Air Conditioning Systems, SAE Int.J. Passeng. Cars - Mech. Syst. 10(1), doi:10.4271/2017-01-0120. Lemmon, E.W., Huber, M.L., McLinden, M.O., 2010. REFPROP, NIST Standard Reference Database 23, v.9. National Institute of Standards, Gaithersburg, MD, U.S.A. Llopis, R., Cabello, R., Sánchez, D., Torrella, E., 2015. Energy improvements of CO2 transcritical refrigeration cycles using dedicated mechanical subcooling. Int. J. Refrigeration 55, 129–141. Li, M., Ma, Y., Guan, H., Li, L., 2004. Development and experimental study of CO2 expander in CO2 supercritical refrigeration cycles. Int. J. Green Energy 1, 89-99. Man-Hoe Kim, Jostein Pettersen, Clark W. Bullard, 2004. Fundamental process and system design issues in CO2 vapor compression systems, Progress in Energy and Combustion Science 30, 119–174. Niu Baolian, 2006. Theory and experiment of natural refrigerant mixture used in low temperature circuit of cascade refrigeration. Tianjin university. Park YC., Yin JM., Bullard CW., Hrnjak PS., 1999. Experimental and model analysis of control and operating parameters of transcritical CO2 mobile A/C system. ImechE, 163–170. P.-C. Qi, Y.-L. He, X.-L. Wang, X.-Z. Meng, 2013. Experimental investigation of the optimal heat rejection pressure for a transcritical CO2 heat pump water heater, Appl. Therm. Eng. 56, 120–125. R.J. Moffat, 1988. Describing the uncertainties in experimental results, Exp. Therm.Fluid. Sci. 1, 3–17.
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Performance improvements evaluation of an automobile air conditioning system using CO2 -propane mixture as a refrigerant Binbin YU , Dandong WANG , Cichong LIU , Fuzheng JIANG , Junye Shi , Jiangping CHEN PII: DOI: Reference:
S0140-7007(18)30001-X 10.1016/j.ijrefrig.2017.12.016 JIJR 3858
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
24 October 2017 3 December 2017 30 December 2017
Please cite this article as: Binbin YU , Dandong WANG , Cichong LIU , Fuzheng JIANG , Junye Shi , Jiangping CHEN , Performance improvements evaluation of an automobile air conditioning system using CO2 -propane mixture as a refrigerant, International Journal of Refrigeration (2018), doi: 10.1016/j.ijrefrig.2017.12.016
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ACCEPTED MANUSCRIPT Highlights Performance of pure CO2 AMAC system is enhanced using CO2-propane mixture as a refrigerant. Effects of operating parameters on system performance have been investigated. Additional performance evaluation has been experimentally carried out with constant cooling capacity. A new optimum high pressure control algorithm has been proposed within a deviation of 5%.
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Performance improvements evaluation of an automobile air conditioning system using CO2-propane mixture as a refrigerant
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Binbin YUa, Dandong WANGa, Cichong LIUa, Fuzheng JIANGa, Junye Shia,b, Jiangping CHENa,b* a. Institute of Refrigeration and Cryogenics, Shanghai Jiaotong University, Shanghai, China b. Shanghai High Efficiency Cooling System Research Center, Shanghai, China *corresponding author: [email protected]. +(86)21 34206775
Abstract
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The main purpose of this work is to enhance the energy efficiency of CO2 automobile air conditioning system. Theoretical analysis demonstrated that the mixture of CO2 and propane can improve its performance, thus, experiments have been carried out to see effects of various CO2-propane mass fractions of 100/0, 90/10, 80/20, 70/30, 60/40, 50/50 on the system performance at different ambient temperatures and gas cooler frontal air velocities. Experimental results show similar trends with those from the theoretical results. It has been shown that under the same compressor speed, system COP reaches highest at 60% of CO2 mass fraction, which is 29.4% higher than pure CO2 system and even achieves equal level of the R134a system, the optimum pressure and discharge temperature are reduced up to a maximum of 40% and 47℃ during the research range. Furthermore, comparison was carried out under the same cooling capacity by adjusting compressor speed for different mass fraction of CO2, results demonstrate that the use of CO2-propane mixtures yields a maximum COP rise of 22%even when cooling capacity is kept constant. A new optimum high pressure control algorithm for the transcritical CO2-propane mixture cycle has been developed based on the experimental data within a deviation of 5%.
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Keywords: Zeotropic mixtures; CO2; propane; Automobile; Air conditioning; Performance
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Accumulator Coefficient of performance Specific heat (J kg-1 K-1) Critical pressure Critical temperature Critical density Electrical expansion valve Electric vehicle Gas cooler Enthalpy (kJ kg-1) Specific irreversibility (kJ kg-1) Internal combustion engine vehicle Internal heat exchanger Lower explosion limit Mass (g) Molar mass (g/mol) Automobile air conditioning National Institute of Standards and Technology Pressure (MPa) Specific cooling capacity (kJ kg-1) Cooling capacity (kW) Revolutions per minute Entropy (kJ kg-1 K-1) Temperature (℃) Specific work (kJ kg-1) Work（kW） Mass fraction of Upper explosion limit Volume flow rate (m3/h) Velocity（m/s） Molar volume of gas (L/mol)
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Nomenclature Accu COP Cp CP CT CD EXV EV G/C h i ICEV IHX LEL m M AMAC NIST P q Q RPM s T w W x UEL u v Vm Greek letters
AC
Effectiveness Pressure ration Volume ration of CO2 and propane
Subscripts a con dis env evp
Air Condenser Discharge Environment Evaporator
ACCEPTED MANUSCRIPT Experiment Gas cooler Inlet Mixture Outlet Optimum Pure Predict Refrigerant Saturation Suction Saturation vapor
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exp gc i mix o opt pur pre r Sa suc Sv
1 Introduction
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After the approval of Kigali amendment in the 28th Conference of the Parties to the Montreal Protocol, a historic limitation of greenhouse gas hydrofluorocarbons (HFCs) was reached (montreal-protocol, 2016). The widely used HFC refrigerant R134a in automobile air conditioning system is not questioned to be limited and banned gradually, interests for CO2 as an AMAC refrigerant have been promoted a step forward. The incomparable heating performance of CO2 heat pump has been experimentally demonstrated (Wang et al, 2017). The European automotive manufacture has announced to offer CO2 air conditioning system in production (Daimler, 2017). The performance of different components (Ayad et al., 2012; Baek et al., 2013; Pettersen et al., 1998) and AMAC systems (Kim et al., 2009a; Liu et al., 2005; Martin et al., 2005; Yang et al., 1998) demonstrate that the COP of a CO2 air conditioning system is lower in the 10% usage conditions at high ambient temperature (above 30℃). Brown et al., 2002 evaluated the performance merits of CO2 and R134a AMAC systems. The results from semi-theoretical models show that the COP of CO2 was lower by 21% at 32.2℃ and by 34% at 48.9℃. The COP disparity was even greater at high speeds and ambient temperatures. For the transcritical CO2 systems, many researchers have devoted to improving their efficiency through various methods. Aprea and Maiorino (2008), Torrella et al. (2011) and Sanchez et al. (2014) demonstrated that the use of internal heat exchanger (IHX) can improve the COP up to a maximum of 10%. Kawamoto et al. (2017), He et al. (2014) and Lee et al. (2014) investigated the ejectors to improve COP by reducing expansion loss. The use of expanders (Li et al., 2004; Yang et al., 2007) to reduce irreversibilities during the expansion process also offered COP increments. Llopis et al. (2015) applied a dedicated mechanical subcooling system, theoretical results showed 9.5%, 13.5% and 13.1% enhancements of COP at evaporation levels of 5 °C, -5 °C and -30 °C, respectively. Inagaki et al. (1997) found that the COP of a CO2 air conditioning system for moderate ambient temperature was improved by 20% using a two-stage split cycle, while 5% for higher ambient temperature conditions. Another improvement is to install a gas liquid separator downstream of the expansion device, bypass the vapor around the evaporator, feeding saturated liquid into the evaporator to eliminate the problems of two-phase distribution (Kim et al., 2004). Any new technology must demonstrate that it is workable in specific application areas, as for AMAC system, in the long run, however, the technologies above may be unrealistic in performance, size, weight, economic, etc. Another solution to improve the performance of CO2 systems is via using zeotropic mixture as a refrigerant such as CO2-propane. Propane is the refrigerant having a higher refrigerating effect and a much lower vapor
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density than CO2 (Kim and Kim, 2005). Thus, one can easily expect that adding propane to CO2 improves system efficiency. The earliest application analysis of zeotropic mixtures was in the autocascade system, Kim et al. (2002) investigated the performance of an autocascade refrigeration system using R744/134a and R744/R290 by experiment and simulation, it is found that the cycle has a merit of low operating pressure and small amount of charge. Kim et al. (2005) evaluated experimentally the performance potentials of CO2-propane mixtures in a vapor compression system using water as secondary heat transfer fluid. Results showed that the discharge pressures of CO2-propane mixtures of 85/15, 75/25 and 60/40 by charged mass percentage are 11.8%, 22.7% and 37.9% less than that of CO2, respectively. The COPs of CO2-propane mixtures of 85/15, 75/25 and60/40 are enhanced by 8.0%, 12.8% and 12.5%, respectively. CO2-propane mixtures of 85/15, 75/25 and 60/40 have lower cooling capacities in comparison with CO2, that is, 93.0%, 85.6% and 68.2% of that of CO2, respectively. In this study, safety issues and theoretical analysis of CO2-propane mixtures are evaluated, the possibilities of performance improvements of various CO2-propane mass fractions of 100/0, 90/10, 80/20, 70/30, 60/40 and 50/50 have been carried out experimentally to see their effects on the AMAC system performance varying with different ambient temperatures and gas cooler frontal air velocities. Furthermore, additional comparison was carried out under the same cooling capacity with pure CO2 system by adjusting compressor speed for different mass fraction of CO2. Based on the experimental data, a new optimum high pressure control algorithm for the transcritical CO2-propane mixture cycle has been developed within a deviation of 5%.
2 Operation of the CO2-propane mixture refrigeration cycle
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2.1 Security The addition of an inert gas suppresses the flammability of combustible gases, dilutes the concentration of combustible gas and narrows the range of explosive limits. Literature gives the formula to calculate the explosion limit (Niu, 2006). Lower explosion limit：
LEL 2.21995 2.23676 1.15607 2 0.96581 3 0.30005 4 0.03886 5 0.00181 6 (1)
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（ 0 8.21 ） Upper explosion limit：
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UEL 8.99719 6.01554 0.49461 2 0.01209 3 （ 0 8.21 ）
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𝐿𝐸𝐿(𝑈𝐸𝐿)𝑔/𝑚3
1000 × 𝐿𝐸𝐿(𝑈𝐸𝐿) = ×𝑀 𝑉𝑚
(2) (3)
Fig. 1 shows the explosion limit of CO2-propane mixtures with respect to CO2 mass fraction. The explosion limit of pure R290 is 2.20%/9.50%. When the mass fraction of R744 are 30% and 50% respectively, the explosion limit are 3.32%/12.03% and 4.86%/12.8%, respectively; According to ASHRAE34-2013 on the safety of refrigerant classification, when the mass fraction of the CO2 is above 60%, the lower explosion limit of mixed refrigerant is greater than 100 g/m3, the safety level rises from A3 to A2, when the mass fraction of the CO2 is equal to 89%, the upper explosion limit is equal to the lower explosion limit, when the mass fraction of the CO2 is greater than 89%, the corresponding security level rises from A2 to Al level, the mixture is already non-combustible gas.
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Fig. 1 Explosion limit of CO2-propane mixtures with respect to CO2 mass fraction.
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2.2. Temperature–composition relations for CO2-propane mixtures at a fixed pressure Fig. 2 is a typical temperature–composition diagram for CO2-propane mixtures at a fixed pressure of 4MPa with two lines and three regions, the dew point of the mixed refrigerant with CO2 mass fraction varying from 0 to 1 constitutes the dew line, and the bubble point constitutes the bubble line, this two lines divide the diagram into three regions: subcooled liquid region, superheated vapor region and two-phase region. As shown in Fig.2, the dew point temperature of the mixed refrigerant with a certain CO2 mass fraction is higher than the bubble point temperature, just because of this, the evaporation or condensing process in a zeotropic refrigerant system is no longer isothermal, but a process with temperature glide. When the CO2-propane mixture reaches the gas-liquid equilibrium, the gas phase consists of a large part of CO2 because of its volatility and low boiling temperature, while most of propane will stay in the liquid phase.
Figure.2 Temperature–composition diagram of CO2-propane mixtures at a fixed pressure
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2.3 Variation of critical properties and saturation vapor pressure Fig. 3 shows that with the decrease of mass fraction of CO2, the critical pressure and critical density decrease to a large extent, when the critical temperature increases gradually. Fig. 4 gives the CO2-propane mixtures bubble point pressure and pure refrigerants saturation vapor pressure varying with temperature. As it can be seen, when the mass fraction of CO2 decreases, the saturation pressure shows a downward trend, thus using CO2-propane mixtures can contribute a lot to reduce system pressure.
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Figure.3 Critical properties of CO2-propane mixtures varying with CO2 mass fraction
Figure.4 Saturation vapor pressure of CO2-propane mixtures varying with temperature
2.4 Theoretical analysis To evaluate the possibilities of improving the performance of AMAC system using CO2-propane mixture as a refrigerant, we consider a simple theoretical analysis using the thermodynamic properties from NIST Refprop 9 database (Lemmon et al., 2010). Fig. 5 shows the pressure-enthalpy (a) and temperature-entropy (b) diagram of CO2-propane mixture refrigeration cycle varying with mass fraction of CO2, it is obvious that the system
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gradually moves from the subcritical cycle to the transcritical cycle with increasing of CO2 mass fraction in the CO2-propane mixture. The following assumptions are made for the analysis of CO2-propane mixture system: 1) All components are assumed to be a steady-state and steady-flow process; 2) All pressure drops and heat losses to the environment are neglected; 3) The throttling processes are isenthalpic; 4) The evaporation temperature and a constant compressor suction superheat of 5℃ are fixed; 5) The gas cooler outlet temperature for transcritical cycle or the condensing temperature for subcritical cycle are obtained by an approach temperature of 5℃ to the environment as Equation (4); 6) The compression process is irreversible and the isentropic efficiency related to its pressure ratio is taken into account, Equation (5) of compressor isentropic efficiency was used (Liao et al., 2000).
Tgc,o / con Tenv 5℃
is 1.003 0.121
(4) (5)
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For the transcritical cycle, all the calculations are carried out for the optimum high pressure through an iteration process. The specific cooling capacity is obtained with Equation (6) considering the equal enthalpy difference between the high temperature side and low side of IHX. The specific compressor work is calculated by Equation (7), where the hdis, s represents discharge isentropic enthalpy, and the COP is obtained by Equation (8).
hdis,s hsuc q w
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q hsuc hgc,o
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(7)
(8)
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（a）
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（b） Fig. 5 Pressure-enthalpy (a) and Temperature-entropy (b) diagram of CO2-propane mixture refrigeration cycle. The theoretical results of CO2-propane mixture refrigeration cycle are shown in Fig. 6 with mass fraction of CO2 varying from 1 to 0.5. It is demonstrated that the use of CO2-propane can continuously improve the COP of pure CO2 system with reduction of the CO2 mass fraction, when the CO2 mass fraction is 0.5, both discharge pressure and temperature drops significantly to a maximum value of 48% and 20%, respectively. Additionally, CO2-propane mixtures have lower cooling capacity compared with pure CO2. The theoretical results obtained are undoubtedly of great value for the experiments in following sections.
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Fig. 6 Theoretical results with respect to mass fraction of CO2.
3 Experiments
3.1. Experimental setup description The Schematic diagram of the experimental CO2-propane mixture AC system and test rig for vehicle is shown in Fig. 7. The AC system consists of two parallel flow type micro-channel heat exchangers involving an
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evaporator and a gas cooler, a plate IHX, a rotary type compressor and an electrical expansion valve. All components were successfully designed to bear the extremely high pressure due to the high operating pressure of CO 2 system, we designed a plate IHX which can bear the extreme burst pressure of 100MPa in our test and both microchannel heat
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exchangers can withstand a pressure of 30Mpa with our new structure and design. The specifications of the
CO2-propane mixture AC components are listed in Table 1. The compressor, with a displacement volume of 6 cm3 rev-1, is a fixed displacement and variable speed type driven by a variable frequency inverter which can be adjusted
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from 30hz to 100hz. The maximum allowable compressor discharge temperature and pressure are 120℃ and 12MPa, respectively. The gas cooler and evaporator in this system are both microchannel heat exchagers made of aluminum
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with louvered fin, their core size are 540mm, 325mm, 12.5mm and 275mm, 235mm, 36mmin width, length and depth, respectively. The internal heat exchanger is a stainless steel plate heat exchanger with a core size of
111W*310L*28.2D (mm). An electric expansion valve that can be controlled from 0-500 steps with step motor
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controller was installed as expansion device, with which the system mass flow rate and high pressure can be precisely regulated. Additionally, an accumulator with a volume of 600ml was installed after evaporator to serve as a system refrigerant charge buffer. Fig. 7 also shows the layout of enthalpy potential method for the system performance test. The calorimeter facility
consists of two chambers with a wind tunnel in each side. The left one installed with a gas cooler was used to simulate the environmental conditions inside the cabin and the other installed with an evaporator was used to modulate the environmental conditions outside the vehicle. Both chambers were equipped with a blower, a cooling system, a PTC heating unit and humidity controller to adjust the air flow rate, ambient temperatures and relative humidity according to the requested operating conditions. Measurement devices showed in Fig. 7 are instrumented to measure the pressure and temperature of refrigerant and air. Refrigerant side temperatures are measured by platinum resistance temperature sensors with an uncertainty of ±0.2°C. The environmental dry bulb and wet bulb temperature can be controlled within ±0.2°C. Refrigerant side absolute pressure and pressure drop of air loops are measured by piezoresistive type transducer
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Fig. 7 Schematic diagram of the experimental CO2-propane mixture AMAC system and test rig. Table 1. The specifications of the CO2-propane mixture AMAC components Specifications
Compressor
DC-driven, rolling rotor type, variable speed (1800-6000rpm), Displacement: 6 cc/rev Microchannel, Parallel flow type, 4-pass, 2-slab, Core size: 275W*235L*36D(mm) Microchannel, Parallel flow type, 3-pass, 1-slab, Core size: 540W*325L*12.5D(mm) Plate type heat exchanger, Stainless steel, Size: 111W*310L*28.2D(mm) EXV, copper, orifice diameter:1.4mm, driven by step motor 600ml
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Gas cooler
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Evaporator
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Components
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IHX Expansion device Accumulator
3.2. Test method and Test conditions To evaluate the performance characteristics of CO2-propane mixture AC system, several CO2 mass fractions in the mixtures have been test at various operating conditions showed in Table 2. The evaluated test methods and conditions are: Proper charge amount: the system charge amount was firstly adjusted to reach the maximum COP and compressor suction superheat was ensured above 5℃ to avoid liquid shock, and then all experiments were carried out at this charge amount.
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Outdoor air velocity: the system has been tested at four different gas cooler frontal air velocities: 1.5, 2, 3 and 4.5m/s. These velocities are supplied by the blower in the outdoor chamber to simulate different driving speed, 1.5 and 2m/s for idling condition, 3 and 4.5m/s for normal driving condition. Ambient temperature: the system has been tested at four different outdoor temperatures: 27, 35, 42, 45℃, when the indoor temperature keeps at 27℃. An extreme high temperature condition of both 40℃ for indoor and outdoor has also been considered. EXV opening adjustment: all test conditions have been performed within a wide range of gas cooler pressure to reach the optimum COP point by manually adjusting the EXV opening. Compressor speed: for condition 1-3 in Table 2, the compressor speed has been kept at 4800rpm, which can supply sufficient cooling capacity for pure CO2 system. However, theoretical results show that the cooling capacity of CO2-propane AC system decreases with reduction of CO2 mass fraction, thus, the condition 4 in Table 2 has been designed to evaluate the CO2-propane AC system performance at the same cooling capacity of pure CO2 system.
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Table 2. Experimental conditions Ta,e,in(℃)*
1
27 100%,90%,80%, 70%,60%,50%
2 3
27 40
4
ua,e,in（m3.h-1） Ta,gc,in（℃）
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27
va,gc,in（m/s） RPM
420
35
1.5,2,3,4.5
4800
420
27,35,42,45
2
4800
420
40
2,4.5
4800
420
27,35,42,45
2
/
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*：the relative humidity was controlled at 50%. “/” represents that the compressor speed is adjusted to keep the cooling capacity constant with the pure CO2 system. The cooling capacity Q of the experiments is calculated by Equation (9), where the Qa representing air side heat
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transfer is obtained by the heat balanced method using Equation (10), the Qr representing refrigerant side heat transfer is obtained by the enthalpy difference method using Equation (11), then the overall COP is determined by Equation (12). Modifications of COP due to the use of CO2-propane mixtures are evaluated as a percentage variation expressed by
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Equation (13). The relative uncertainty of the cooling capacity and COP of the system are calculated as5.5% and 6.3%, respectively, by the sequential perturbation method (Moffat, 1988).
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Q (Qa Qr ) / 2
Qa maC p (Ta,o Ta,i )
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Qr mr (hsuc hgc,o ) COP Q / W
COPr （COPmix COPpur）/ COPpur
(9) (10) (11) (12) (13)
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Table 3. Accuracy of the experimental parameters with various sensor devices Device
Range
Accuracy
Temperature Pressure Mass flow rate Compressor work Cooling capacity COP
RTD-type temperature sensors Pressure transducers
-50~150℃ 0~10MPa/0~20MPa 0~600kg.h 0~6kw / /
±0.2℃ ±0.5% ±0.15% of reading ±0.5% of reading 5.5% 6.3%
Coriolis type flow meter
Digital power meter / /
4 Experimental results and discussion
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4.1 System performance at various CO2 mass fractions 4.1.1 The influence of G/C frontal air velocity on system performance Figs. 8-12 show the experimental results for condition 1 in Table 2. It can be observed in Fig. 8 that at 35/27 condition, the COP enhancement reaches highest when the mass fraction of CO2 is 60% for all G/C frontal air velocities. When the G/C frontal air velocity is 4.5m/s, the exact COPs of CO2 mass fraction of 100%, 90%, 80%, 70%, 60%, 50% are1.81, 1.90, 2.07, 2.17, 2.35 and 2.32, the corresponding percent increments are 4.8%, 14.5%, 20% and 29.4%, 28.3%, relatively. The COP trend can be explained by the decreasing of specific compressor work and specific capacity when the mass fraction of CO2 reduces, and the compressor work decreases more quickly, so the system COPs increase gradually. The COP enhancement by using CO2-propane mixture will contribute a lot to improve fuel economy for ICEVs and extend driving range for EVs. However, the biggest difference between the theoretical and experimental results for COPs is that the COP of CO2 mass fraction of 50% is a little lower than that of 60%, this trend may result from the low compressor efficiency and the poor matching of temperature profiles between CO2-propane mixture of 50/50 and air, and for this study, the detailed reason will be given in subsection of 4.1.3. Fig. 9 and Fig. 10 show impacts of G/C frontal air velocity on the optimum high pressure and discharge temperature of CO2-propane mixture system varying with mass fraction of CO2. The optimum pressures are reduced by 40% from 10.4MPa to 6.3MPa and discharge temperature by 47℃ from 114℃ to 67℃ when the mass fraction of CO2 is reduced to 50%, on the average for different air velocities. For the maximum COP point, the optimum pressure is 7MPa and discharge temperature is 75℃. The reduction of operating pressure and discharge temperature will effectively release the CO2 compressor load such as high pressure leakage and thermal deformation.
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Fig. 8 Impacts of G/C frontal air velocity on the COP of CO2-propane mixture system varying with mass fraction of CO2.
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Fig. 9 Impacts of G/C frontal air velocity on the optimum high pressure of CO2-propane mixture system varying with mass fraction of CO2.
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Fig. 10 Impacts of G/C frontal air velocity on the discharge temperature of CO2-propane mixture system varying with mass fraction of CO2.
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Fig. 11 shows that the cooling capacity decreases with reduction of CO2 mass fraction, and the slope of cooling capacity increases with reducing CO2 mass fraction, results show that the CO2 mass fraction of 70% is a significant point, for if the CO2 mass fraction is higher than this point, the cooling capacity declines slightly only within a 5.7% level on the average of different air velocities, but when the CO2 mass fraction is smaller than 70%, the cooling capacity declines appreciably by an average of 38%. Thus, the balance between cooling capacity and COP at different air velocities should be weighed deliberately. Fig. 12 shows other significant parameters, besides COP, cooling capacity, Popt, Tdis, which are described before, as the mass fractions of CO2 decrease, the compressor work reduce almost linearly to the reduction of CO2 mass fraction, refrigerant mass flow rates are reduced by 30% on the average as the suction density and volume efficiency reduce simultaneously, and the gas cooler refrigerant outlet temperatures are almost kept constant.
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Fig. 11 Impacts of G/C frontal air velocity on the cooling capacity of CO2-propane mixture system varying with mass fraction of CO2.
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(b) Fig. 12 Various parameters of CO2-propane mixture system varying with mass fraction of CO2 (Ta,ev,i=27℃, Ta,gc,i=35℃, va,gc,i=4.5m/s).
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4.1.2 The influence of ambient temperatures on system performance Figs. 13 and Figs. 14 show the experimental results for condition 2 in Table 2. It is clear that the COP and cooling capacity decrease with the increasing outdoor temperature. Figs. 13 shows that at the outdoor temperature of 27℃, COPs of CO2-propane mixtures of 90%, 80%, 70%, 60%, 50% are 2.05, 2.27, 2.49, 2.64 and 2.66, the corresponding percent increments are 0.5%, 11.3%, 22.1%, 29.4%, 30.4%, respectively, but for outdoor temperature of 35℃, 42℃ and 45℃, the COPs of CO2-propane mixtures reach to a maximum value at the 60% of CO2 mass fraction, which are higher than pure CO2 by 26.66%, 24.87% and 19.2%, respectively. The impacts of outdoor temperature on the cooling capacity varying with mass fraction of CO2 is similar to that of gas cooler frontal air velocity illustrated in subsection 2.1.1. As shown in Fig. 14, the cooling capacity decreases with reduction of CO2 mass fraction, when the CO2 mass fraction is higher than 70%, the cooling capacity declines slightly only within a 4.8% level on the average of different outdoor temperatures, but when the CO 2 mass fraction is smaller than 70%, the cooling capacity declines appreciably by 41% on the average. The balance between cooling capacity and COP at different outdoor temperature should be weighed deliberately.
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Fig. 13 Impacts of outdoor temperature on the COP of CO2-propane mixture system varying with mass fraction of CO2 (va,gc,i=2m/s).
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Fig. 14 Impacts of outdoor temperature on the cooling capacity of CO2-propane mixture system varying with mass fraction of CO2 (va,gc,i=2m/s). At the condition of both 40℃ for outdoor and indoor temperature, it can be observed in Fig. 15 that COPs of CO2-propane mixtures increase continuously with the reduction of CO2 mass fraction, though the curves slope decrease gradually, the COP of 50% CO2 mass fraction is enhanced by 45.2% and 31.7% compared with pure CO2 at the gas cooler frontal air velocity of 2m/s and 4.5m/s, respectively. As the CO2 mass fraction varied from 100% to 50%, the cooling capacity is decreased by 19.8% and 39.7% at the gas cooler frontal air velocity of 2m/s and 4.5m/s, respectively.
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Fig. 25 Impacts of G/C frontal air velocity on the CO2-propane mixture system performance varying with mass fraction of CO2 at 40/40 condition.
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4.1.3 Analysis of COP loss at 50% of CO2 mass fraction in some conditions From the theoretical results of Fig. 6, we know that the use of CO2-propane mixtures can continuously improve the COP of pure CO2 system with reduction of CO2 mass fraction, but an interesting observation can be made in the experiments mentioned above is that the COP enhancement reaches highest when the mass fraction of CO2 is 60% except for the conditions of 27/27 and 40/40 for outdoor/ indoor temperature in Table 2. As shown in Fig. 16, the loss of compressor efficiency and different decline trends may account for this, when the mass fraction of CO2 reduces from 100% to 50%, the total compressor efficiency calculated by Equation(14) are decreased by 49.0%, 30.5% and 26.8% for 35/27, 40/40 and 27/27 condition, respectively. Therefore, due to the two contrary factors of higher propane refrigerating effect and compressor efficiency loss, the COP enhancements exist a maximum value for most conditions.
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Fig. 3 Impacts of ambient conditions on the total compressor efficiency varying with mass fraction of CO2.
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After expansion of CO2-propane mixture in the electric expansion valve, the working fluid turns into two-phase flow and enters the evaporator, at the evaporator inlet, the refrigerant qualities are between 0.3~0.4 during our research range, due to the zeotropic effect, there exists a temperature glide resulting from the temperature difference between the bubble point and dew point at the same CO2 mass fraction, and there are more CO2 in the gas phase because of its volatility and low boiling temperature than propane, this situation is contrary in the liquid phase. Kim illustrated the temperature glide effect of CO2-propane mixtures in his research, when the temperature glide occurring in the evaporator of CO2-propane mixture with reduction of CO2 fraction exceeds the temperature difference of secondary fluid, the mean temperature difference in the evaporator of CO2-propane mixture increases. This results in an increased temperature difference in the evaporator and reduced system efficiency. As a result, when zeotropic mixtures are used in applications with the same temperature gradients of two heat transfer fluids in heat exchangers, the benefits derived from temperature gliding effect can contribute to an improvement of the system capacity and efficiency (Kim et al ., 2008). For AMAC system using CO2-propane mixture as a refrigerant, since the heat exchanger area can not be changed, the entropy generation in the evaporator can be reduced if the temperature glides of the CO2-propane mixture and air are well matched, as shown by Equation (15), Equation (16-18) illustrated that the requirement of perfect glide matching can be achieved by adjusting the air flow rate through HVAC blower.
Qa Qr
mr c pr dtr ma c padta ma mr c pr / c pa
(15) (16) (17) (18)
4.2 Performance comparison at constant cooling capacity CO2-propane mixtures have lower cooling capacities in comparison with pure CO2 at constant compressor
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speed, thus, additional performance comparison at constant cooling capacity is essential to meet the comfort requirement of cabin. Fig.17 shows the impacts of mass fraction of CO2 on the COP at constant cooling capacity with pure CO2 system varying with outdoor temperatures, the cooling capacities of pure CO2 system are 4.2kw, 3.8kw, 3.4kw and 3.2kw for the outdoor temperature of 27℃, 35℃, 42℃ and 45℃, respectively. It is clear that COP improvements all reach highest at the CO2 mass fraction of 70% for different outdoor temperatures, that is, 22.1%, 18.6%, 19.6% and 18.4% of pure CO2 system, the corresponding exact COPs are 2.49, 1.92, 1.58 and 1.44, respectively. For CO2 mass fraction of 60%, its COP reduced by 16% on the average compared with that of constant compressor speed, but is still higher than pure CO2 by 8%. Therefore, if taking security, cooling capacity and COP into consideration, the best choice for mass fraction of CO2-propane mixtures is 70/30.
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Fig. 17 Impacts of mass fraction of CO2 on the COP at constant cooling capacity with pure CO2system(Ta,ev,i=27℃, va,gc,i=2m/s).
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4.3 Correlation for predicting optimum high pressure for the transcritical CO2-propane mixture cycle 4.3.1 Verification of existing pure CO2 system correlations University of Illinois at Urbana-Champaign (UIUC) did experimental investigation and model analysis of the control and operating parameters of a transcritical CO2 mobile A/C system under idle and medium driving operating conditions. The results demonstrated that the COP of the transcritical CO2 system is a strong function of the gas cooling pressure. Furthermore, the strong relation between the optimum pressure and the gas cooler outlet temperature was revealed, which suggests a simple controller of only one parameter for the transcritical CO2cycle (Park YC et al ., 1999). So far there is no optimum high pressure correlation for CO2-propane mixtures, the available correlations from open literature are all developed for pure CO2 as listed in Table 4, the errors of these correlations for predicting the optimum high pressure of transcritical CO2-propane mixtures cycle are verified in Table 4 by experimental data.
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Table 4 Summary of existing optimum high pressure correlations with only one parameter Literature
Correlation
Deviation(%)
Popt 1.938Tgc,r ,o 9.872
Yang et al.
-6~15
Popt 2.6Tgc,r ,o 7.54
Kauf
-26~-12
Popt 2.56Tgc,r ,o 4.2
Yahia et al.
-18~5
Qi, He, Wang and Meng
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Chen and Gu
-20~2
Popt 132.2 8.4Tgc,r ,o 0.3Tgc2 ,r ,o 27.7 104 Tgc3 ,r ,o
-16~5
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4.3.2 Development of a new optimum high pressure control algorithm for transcritical CO2-propane mixture cycle By verification of the available correlations, the conclusion can be drawn that none of these existing correlations can predict the experimental data of CO2-propane mixtures satisfactorily as they were obtained from pure CO2 system. Therefore, it is necessary to develop a new optimum pressure control algorithm for transcritical CO2-propane mixture system. In the present study, a new correlation was proposed based on the existing model by considering strong effect of gas cooler outlet temperature and effect of CO2-propane mixtures, as shown in Equation (19), where the CO2 mass fraction is limited to ensure transcritical cycle. Table 5 lists all the 92 validation data of the new correlation, and Fig. 18 presents the comparisons of the experimental data with the predicted values of the new correlation and literature. It can be seen that the new correlation predictions agree with all of the experimental data within a deviation of ±5%.
Popt （MPa） 0.16629Tgc,out （℃） 3.20186 xCO2 （ 0.7 xCO2 1 ）
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(19)
Table 5 validation of the new developed optimum high pressure control algorithm
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Tgc,out
Pexp,opt
Ppre,opt
error
45.8
10.83
10.81
0.20%
0.8
45.2
9.90
10.08
-1.85%
1
44.5
10.62
10.61
0.11%
0.8
43.1
9.38
9.72
-3.47%
1
42.3
10.30
10.24
0.56%
0.8
40.7
9.14
9.32
-1.97%
1
40.8
10.00
9.99
0.03%
0.8
38.5
8.90
8.97
-0.83%
1
45.3
10.80
10.73
0.61%
0.8
45.1
9.95
10.06
-1.10%
1
44.2
10.62
10.55
0.66%
0.8
42.8
9.43
9.68
-2.57%
1
42.0
10.30
10.19
1.07%
0.8
40.6
9.12
9.31
-2.07%
1
40.5
10.00
9.94
0.60%
0.8
38.3
8.92
8.93
-0.12%
1
38.2
9.48
9.55
-0.76%
0.8
35.8
8.5
8.52
-0.52%
1
42.1
10.29
10.20
0.90%
0.8
37.2
8.93
8.75
2.09%
CO2
1
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CO2-propane
CO2
Tgc,out
Pexp,opt Ppre,opt
error
10.62
10.61
0.11%
0.8
43.1
9.4
9.72
-3.47%
1
48.8
11.27
11.31
-0.41%
0.8
46.8
9.98
10.34
-3.52%
1
50.3
11.42
11.57
-1.28%
0.8
50.1
10.5
10.89
-3.38%
1
53.0
11.75
12.02
-2.24%
0.8
52.1
10.8
11.22
-4.23%
1
37.9
9.48
9.50
-0.26%
0.8
35.6
8.53
8.48
0.57%
1
42.0
10.29
10.19
1.01%
0.8
37.5
8.92
8.80
1.39%
1
44.1
10.62
10.54
0.82%
0.8
43.3
9.46
9.76
-3.09%
1
48.2
11.27
11.22
0.43%
0.8
47
10.52
10.38
1.38%
1
50.6
11.42
11.62
-1.71%
0.8
49.8
10.52
10.84
-2.98%
1
52.8
11.75
11.98
-1.93%
0.8
51.8
10.83
11.18
-3.09%
1
52.5
12.08
11.93
1.27%
0.8
49.1
11.3
10.72
4.91%
1
45.8
10.91
10.82
0.84%
0.8
44.1
10.1
9.89
1.84%
1
52.2
12.08
11.88
1.70%
0.8
48.9
11.23
10.69
5.02%
1
45.8
10.91
10.82
0.84%
0.8
44.0
10.13
9.88
2.55%
0.9
45
10.70
10.36
3.23%
0.7
45.1
9.2
9.75
-5.74%
0.9
44.1
10.44
10.22
2.16%
0.7
41.8
8.9
9.20
-2.99%
0.9
41.6
9.96
9.81
1.59%
0.7
39.9
8.6
8.87
-2.54%
0.9
39.8
9.59
9.50
0.94%
0.7
37.8
8.4
8.53
-1.08%
0.9
44.8
10.70
10.33
3.57%
0.7
45
9.26
9.72
-4.78%
0.9
44.3
10.44
10.25
1.85%
0.7
41.6
8.95
9.16
-2.28%
0.9
41.8
9.96
9.83
1.31%
0.7
40.1
8.62
8.91
-3.25%
0.9
39.6
9.59
9.47
1.32%
0.7
37.6
8.43
8.49
-0.75%
0.9
41.9
9.92
9.84
0.77%
0.7
41.8
8.9
9.20
-2.99%
0.9
44.1
10.44
10.22
2.16%
0.7
47.6
9.8
10.16
-3.05%
0.9
48.7
11.08
10.98
0.92%
0.7
50.9
10.2
10.71
-4.30%
0.9
53.5
11.72
11.78
-0.47%
0.7
41.6
8.93
9.16
-2.50%
0.9
37.0
0.9
41.7
0.9
43.8
9.03
2.86%
0.7
47.5
9.82
10.14
-3.16%
9.92
9.82
1.03%
0.7
50.3
10.23
10.61
-3.54%
10.44
10.17
2.68%
0.7
48.1
9.8
10.24
-4.20%
48.3
11.08
10.91
1.51%
0.7
43.7
9.6
9.50
1.21%
52.1
11.51
11.55
-0.32%
0.7
48.3
9.82
10.27
-4.41%
0.9 0.9
CE
9.29
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0.9
53.4
11.72
11.76
-0.34%
0.7
43.2
9.63
9.43
2.17%
0.9
52.6
11.68
11.62
0.50%
0.7
47.6
9.9
10.16
-3.02%
0.9
45.2
10.87
10.40
4.58%
0.7
43.2
9.6
9.43
2.17%
0.9
52.4
11.68
11.60
0.75%
0.7
48.0
9.80
10.22
-4.14%
0.9
45.0
10.87
10.36
4.90%
0.7
43.5
9.65
9.47
1.85%
12
+5
%
11 -5 %
10
Yang Kauf Yahia Chen Qi Present
9 8 8
9
10
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Predicted optimum high pressure/MPa
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12
Experimental optimum high pressure/MPa
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Fig. 18 Comparison of predicted optimum high pressure with experimental data
5 Conclusion
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In present study, the thermodynamic simulation of CO2-propane binary mixture refrigeration cycle has been performed firstly. Next, based on the simulation results, the performance characteristics of an AMAC system using CO2-propane mixture as a refrigerant have been experimentally evaluated, experimental conditions include CO2-propane mass fractions of 100/0, 90/10, 80/20, 70/30, 60/40, a wide range of outdoor temperatures, from 27℃ to 45℃, and gas cooler frontal air velocity, from 1.5m/s to 4.5m/s. The work accomplished will be of great value for promotion of natural refrigerants to protect our environment. The main conclusions obtained are as follow: The AMAC system benefits greatly from using CO2-propane mixture as a refrigerant, under the condition of same compressor speed with pure CO2 system, COPs of CO2-propane mixtures reach highest at 60% of CO2 mass fraction, which is 29.4% higher than pure CO2 system and even has achieved equal level of the R134a prototype system, this will contribute a lot to improve fuel economy for ICEVs and extend driving range for EVs. The optimum pressure and discharge temperature are reduced up to a maximum of 40% and 47℃ during the research range, and this will effectively release the CO2 compressor load such as high pressure leakage and thermal deformation. Cooling capacity decreases with reduction of CO2 mass fraction, and the slope of cooling capacity increases with reducing CO2 mass fraction, results show that the CO2 mass fraction of 70% is a significant point, the cooling capacity declines slightly only within a 5.7% level if the CO2 mass fraction is higher than this point. Reducing CO2 mass fraction results in the decrease of CO2 compressor efficiency and temperature glide in the CO2 evaporator, and just because of this, the COP of 50% CO2 mass fraction is a little lower than that of 60% CO2 mass fraction, which is the biggest difference compared with thermodynamic analysis. In our research, for AMAC system this problem can be solved to some extent by adjusting evaporator air flow rate to match the temperature glide. COPs of CO2-propane mixtures are still higher than pure CO2 system even when the cooling capacities are kept constant, the CO2-propane mass fraction of 70/30 can improve the COP up to a maximum of 22%. Taking consideration of security, cooling capacity and COP, the best choice for mass fraction of CO2-propane mixtures is 70/30.
ACCEPTED MANUSCRIPT A new optimum high pressure control algorithm for the transcritical CO2-propane mixture cycle has been developed based on the experimental data, the new correlation agrees with all of the experimental data within a deviation of ±5%. The optimum pressure should be met fundamentally at any conditions to enhance the performance of CO2-propane mixture AMAC system if operating as a transcritical cycle.
6 Acknowledgements
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This project was supported by National Nature Science Foundation of China (NO.51776119).
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