Department of Automotive Engineering, Kocaeli University, Kocaeli, 41380, Turkey

Transcription

1 Effect of Ambient Temperature on the Performance Characteristics of Automotive Air Conditioning System Using R1234yf and R134a: Energy and Exergy-based approaches Mukhamad Suhermanto 1, a) Murat Hoşöz 2, b) M. Celil Aral 1, a) 1 Graduate School of Natural and Applied Sciences, Kocaeli University, Kocaeli, 4138, Turkey 2 Department of Automotive Engineering, Kocaeli University, Kocaeli, 4138, Turkey a) Corresponding author: suhermanto.mukhamad@gmail.com b) mhosoz@kocaeli.edu.tr Abstract. This paper presents the comparative experimental results of an automotive air conditioning (AAC) system charged with refrigerants R1234yf and R134a which were analyzed to find the effect of increasing ambient temperature using energy and exergy-based approaches. The urgencies for such investigation are due to the demand of finding refrigerant both satisfy the performance and also compatible with Kyoto Protocol and EU f-gas regulation. For this aim, a bench-top experimental AAC system was constructed and tested at four different compressor speeds, namely 1, 15, 2, and 25 rpm. For each compressor speed, the temperatures of the air streams at the inlets of the inside duct was 3 C and varies the outside duct, i.e. Tcond,ai, of 3, 35, and 4 C. It was revealed that the AAC system with R1234yf yield comparative performances shown by the higher mass flow rate, comparably less cooling capacity and COP than that of with R134a. The mass flow rates and power absorbed in the compressor get higher as T cond,ai increase. The biggest contributor to the total exergy destruction of the system is always the compressor and the heat exchangers, i.e. evaporator and condenser, respectively. Thus, in order to improve the system performance, those components should be enhanced. 1. INTRODUCTION Chlorofluorocarbons (CFCs) were banned for causing potentially most significant environmental problems as both tremendous contributor toward the destruction of the stratospheric ozone layer and greenhouse effect [1] after being prevalently used in the refrigeration and air conditioning systems, including automotive air conditioners for some decades. The prohibition was under Montreal Protocol in 1987 which enforcing that CFCs to be entirely phased out in the developed countries and developing countries starting from 1996 and 21, respectively [2]. The automotive industry was subsequently introduced to HFC134a (R134a), a family of Hydro fluorocarbons (HFCs), as an alternative whose ozone depletion potential (ODP) is zero [2], to fill the gap during the process of phasing out CFCs. However, it was later found that though HFCs might be harmless to ozone, it is still environmentally destructive for contributing global warming. Consequently, Kyoto Protocol 1997 determined to control not only high ODP gases but also gases with high global warming potential (GWP). R134a was one of gases listed as one of the controlled greenhouse gases for possessing zero ODP, it has a GWP-per-1-year of 143 [3]. Moreover, under

2 the European Union s f-gas regulation, Europe forbids the selling of vehicle equipped by air conditioning system containing refrigerant with a GWP higher than 15 by January 217 [4]. Thus, the search of R134a replacements is, therefore, indisputable. Among the prospective substitutes to R134a are R444A (AC5), CO 2, R152a, and R1234yf. The latter, which is also recognized as Hydrofluoroolefin (HFO) 1234yf, has been regarded as the best potential subtitutes for AAC use due to its low liquid viscosity and small number of GWP [5]. In addition, Society of Automotive Engineering Cooperative Research Project (SAE-CRP), known as CRP1234, conducted an extensive period of collaborative research in an effort to identify a suitable alternative refrigerants unanimously concluded it as acceptable alternative for automotive applications [6]. Since R1234yf shows a good drop-in performance, it has been approved to replace R134a in AAC systems. Hosoz and Karabektas [9] found in their theoretical study that R1234yf has % lower COP than that of R134a. They also reported that the compressor discharge temperature of R1234yf is % lower than that of R134a. Mota-Babiloni et al. [8] determined that the coefficient of performance (COP) and cooling capacity of R1234yf compared to R134a was up to 3-11% and 9% lower, respectively. On the other hand, Lee and Jung [3] obtained 2.7% lower COP and 4.% lower cooling capacity. Consequently, they suggested that the R1234yf system must employ a larger stroke volume compressor to circulate higher refrigerant mass flow rate in the cycle. Mathur conducted an experimental study in a bench-top AAC system, by maintaining the cabin interior condition constant at 2 C and 5% of relative humidity and varying the ambient temperature from 25 to 45 C with the interval of 5 C [9]. The energy-based method revealed that the system with R1234yf yielded 9 21% lower operating pressure and 5 9 o C lower compressor discharge temperature than that of the baseline. Exergy analysis is also required for investigating some important aspects in resource utilization, since energetic performance evaluation alone is considered to be less illustrative [1]. Presently, exergy analysis of AAC systems with R1234yf can only be found in a limited number of literature. Ansari et al. [11] and Yataganbaba et al. [12] performed theoretical exergy analysis of a simple vapor compression refrigeration system with single and dual evaporators, respectively, using R1234yf and R1234ze in replacement to R134a. They determined that R1234yf and R1234ze could be good replacements of R134a. This paper enriches the energy and exergy studies literature about the comparative results of R1234yf and R134a within a certain range of operating conditions in a fully monitored bench-top AAC system. Using experimental data obtained from the tests, the pressure ratio, compressor discharge temperature, refrigerant mass flow rate, cooling capacity, compressor power, COP, pressure ratio across the compressor, have been evaluated for both R1234yf and R134. Moreover, exergy-based approach was applied and the contribution percentage of each component to the total exergy destruction of the AAC system are presented for different ambient temperature. The results are presented in comparative graphics and discussed in detail. 2. EXPERIMENTAL SETUP AND TEST PROCEDURE The schematic illustration and the components specifications of the experimental AAC system is indicated by Fig. 1 and Table 1, respectively. It consists of a five-cylinder wobble-plate compressor, a plate-and-fin laminated type indoor unit serving as evaporator, a thermostatic expansion valve (TXV), a receiver, a filter/drier and a parallelflow micro-channel type outdoor unit as condenser. The compressor was driven at the required speed by a 4 kw three phase electric motor via a frequency inverter. All piping of the refrigeration circuit was made of copper tubing with elastomeric insulator. In order to provide adequate air streams, a centrifugal fan was attached to the inlet of the evaporator air duct providing an air flow rate of.11 m 3 s 1. On the other hand, an axial fan was attached to the inlet of the condenser air duct providing an air flow rate of.18 m 3 s 1. Furthermore, to test the system at required air inlet temperatures, both air ducts were equipped with electric heaters embedded upstream of the evaporator and condenser. All measurement instruments utilized in the bench-top AAC system are shown in Fig. 1, while their characteristics are described in Table 2. A potentiometer connected to the control circuit of the frequency inverter was used for adjusting the compressor speed. A Coriolis flow meter installed downstream of the condenser was activated for measuring the refrigerant mass flow rate. The magnitude of refrigerant temperatures were taken by type-k thermocouples at the inlets and outlets of each essential component. Furthermore, the air streams were measured for both dry and wet bulb temperatures at the T and T w points, respectively, as shown in Fig. 1. All temperature and pressure data were gained by a data acquisition system having 16 bit- 2 KHz of frequency with 56 input channels of thermocouple module and 8 channels of transducer interface module. Prior to the thermodynamic analysis, the thermodynamic properties of the refrigerants were obtained from REFPROP.

3 FIGURE 1. Schematic diagram of the experimental automotive air conditioning system TABLE 1. Specifications of the components of the automotive air conditioning system Component Specification Component Specification Compressor Type:fixed-capacity wobble-plate Stroke volume:138 cm 3 No. of cylinders: 5 Max. Speed: 6 rpm TXVs Outdoor unit Electric heater: outdoor unit Type: parallel-flow micro-channel coil Capacity: 6.44 kw Dimensions: (mm 3 ) Hydraulic diameter: 1.2 mm No. of Channels: 31 Heat transfer area: 4.24 m 2 Power range: 6 kw Indoor unit Electric heater: indoor unit TABLE 2. Characteristics of the instrumentation Type: Internally equilized with bulb Capacity: 5.27 kw Type: laminated coil Capacity: 4.39 kw Dimensions: (mm 3 ) Hydraulic diameter: 9.71 mm No. of Channels: 8 Heat transfer area: 3.6 m 2 Power range: 2 kw Measured Variable Instrument Range Uncertainty Temperature Type-K thermocouple -5 5 C ±.3 absolute Pressure Pressure transmitter Bourdon gauge 25 bar 1 1, 3 bar ±.2 % full scale.1,.5 bar Air flow speed Anemometer.1 15 m s -1 ± 3. % full scale Refrigerant mass flow rate Coriolis flow meter 35 kg h -1 ±.1 % full scale Compressor speed Photoelectric tachometer 1 1 rpm ± 2 % full scale A 2.2 kg of R134a was first loaded into the AAC system, and the first set of tests were performed. Then, the R134a was recovered and 2. kg of R1234yf was charged into the system for the second set of tests. Tests for both refrigerants were performed at four different compressor speeds, namely 1, 15, 2, and 25 rpm with a tolerance of ±1 rpm. Each compressor speed test was performed for one and three air stream temperatures entering

4 the evaporator and condenser, respectively. The temperatures at the air inlet of evaporator (T evap,ai) were maintained at 3 C, while the condenser air inlet temperatures (T cond,ai) were adjusted 3 C, 35 C and 4 C, which are called Set1, Set2 and Set3, respectively. The analysis were done for steady state condition in which the data were taken after the tenth minute of total 15 minutes of each test. The steady-state operation was assumed occurring by the time the refrigerant temperature deviations remained at the compressor discharge were within.5 C for 3 minutes [13]. In addition, an interlude of 1 minutes was given to the AAC system prior to the following test to confirm the similarity of starting conditions existed. 3. THERMODYNAMIC ANALYSIS In this study, the cooling capacity and power absorbed in the compressor were evaluated using the principle of first law of thermodynamics (FLT) applied to control volume system as defined in Hosoz and Karabektas [7] using the common formula as shown in the Eq. (1). 2 V Q W cv cv m h g( z) 2 (1) Assuming that the system operates in steady state and ignoring the potential and kinetic energies change, the first law of thermodynamics can be implemented to the AAC system components individually to evaluate the their performances. Then, the COP of the AAC system is defined as the magnitude of cooling capacity per unit compressor power. The magnitude of the thermodynamic inefficiency in each component of the AAC system can be evaluated through an exergy analysis. To this end, the equation of exergy rate balance for control volumes can be utilized as following [1]. The specific flow exergy can be determined from T 1 Q j W cv m in in m out out Ex (2) d Tj (3) ( h h ) T ( s s) Assuming that the compressor operates adiabatically, the exergy destruction rate in this component can be obtained from E x compin compout W, comp d, comp mr, The exergy destruction rate in the condenser can be calculated from (4) E x m ) m ( ) (5) d, cond r ( cond, in cond, out a, cond E F The specific flow exergies of air in Eq. (4) at the spots E and F can be determined from [14] ( c a p, a c p, v ) T ( T / T ) -1- ln ( T / T ) (1 678)( R T ( ) ln ( ) /( ) RaT ln ( / ) It the TXV is assumed adiabatic, its exergy destruction rate can be obtained from a ln ( p / p ) E x d m ) (7), TXV r ( TXV, in TXV, out Using Eq. (3) and considering that enthalpy remains constant in the TXV, Eq. (7) yields E x m T s s ) (8) d, TXV r ( TXV, in TXV, out Assuming that the refrigerant and air streams in evaporator exchange heat only between each other, the exergy destruction rate in this component can be evaluated from (6)

5 E x m ) m ( ) (9) d, evap r ( evap, in evap, out a, evap B C Finally, the total exergy destruction rate in the AAC system can be determined by calculating all of the exergy destruction rates in the compressor, condenser, TXV, and evaporator. 4. RESULTS AND DISCUSSIONS According to the results, the comparison of limit criteria of usability of AAC using R1234yf and R134a is illustrated by Fig.2 as a function of compressor speed. The criteria includes pressure ratio across the compressor and temperature at the compressor discharge. The pressure ratio has affirmative affects the discharge temperature. Moreover, the higher the discharge temperature, the greater the risk of impairing lubrication, which eventually affecting directly to the mechanical parts of the compressor [15]. Both the compressor ratio and discharge temperature get higher with the increasing T cond,ai. The compressor ratio and discharge temperature of AAC system with R1234yf generally lower than that of R134a except the compessor ratio of the system with T cond,ai =4 o C. It takes place due to the thermophysical properties of R1234yf compared to R134a which is higher for certain temperature as stated by Koban [16]. The compressor ratio of AAC system with R1234yf are %,.1 4.1% lower for Set1, Set2, and.3 5.9% higher for Set3 than that of with R134a. The temperature at compressor discharge of AAC system with R1234yf are o C, o C, and o C for Set1, Set2, and Set3 lower than that of with R134a, respectively. FIGURE 2. Comparison of limit criteria of usability of AAC using R1234yf and R134a: (a) pressure ratio across the compressor, (b) temperature at the compressor discharge Fig. 3 and 4 indicate the energetic performance comparison of AAC system using R1234yf and R134a as a function of compressor speed. Mass flow rate of AAC system with R1234yf, as shown in Fig. 3 (a), are %, %, and % higher than that of with R134a for Set1, Set2, and Set3, respectively. Through Fig. 3 (a), it can be understood that m r gets higher as T cond,ai and compressor speed increase. Higher vapor density of R1234yf provides the reason of its greater mass flow rate compared with R134a. The thermostatic expansion valve of the system, which is originally designed and adjusted for R134a, also has a possibility of contributing the higher mass flow rate of R1234yf. The greater m r of AAC system using R1234yf in some extent is able to compensate its lower enthalpy difference in actual cycle, as depicted in the Fig. A1 of appendix. Thus, it can minimize the performance gap between the AAC system using R1234yf and R134a. The cooling capacity of AAC system with R1234yf, as shown in Fig. 3 (b), are %, %, and % lower than that of with R134a for Set1, Set2, and Set3, respectively. It generally gets lower with increasing T cond,ai and usually gets higher with increasing compressor speed. Fig. 4 (a) and (b) shows the variation of the power absorbed in the compressor (W comp) and the coefficient of performance (COP), respectively. W comp rises as the compressor speed and T cond,ai increase. Recall that the COP is

6 the ratio of cooling capacity to W comp, the COP decreases gradually due to the cooling capacities of both refrigerants increase more sluggish than the compressor powers do. Besides, the COP of both refrigerants gets lower as T cond,ai increases because the trends of cooling capacity which decreases along with the increasing T cond,ai while the power absorbed in the compressor gets higher in accordance with increasing T cond,ai. For Set1, Set2, and Set3, the W comp of AAC system with R1234yf are %, %, and % lower than that of with R134a and the COP of AAC system with R1234yf are ,5 %, %, and % lower than that of with R134a, respectively. FIGURE 3. Comparison of energetic performance I of AAC using R1234yf and R134a: (a) mass flow rate and (b) cooling capacity FIGURE 4. Comparison of energetic performance II of AAC using R1234yf and R134a: (a) power absorbed in the compressor and (b) COP Fig. 5 (a) and (b) depict the comparison of total exergy destruction rate ( E xd ) of AAC system using R1234yf and R134a and contribution percentages of components as a function of compressor speed with T cond,ai =3 o C (Set1) and T cond,ai =4 o C (Set2), respectively. The total E xd of the system using R1234yf generally lower than that of with R134a. Furthermore, it can be observed that the contribution percentage of the condenser and compressor to the total of E x for both refrigerants increase in accordance with the rising T cond,ai. This occurrence is due to the increasing d

7 temperature of air intake leads the higher mean temperature difference of air and run-through refrigerant, thus, enlarging the heat transfer and so does the. Despite of the insignificance, the contribution percentage of E x d E x d, TXV was seen to be greater with low ambient air. On the other hand, the biggest contribution percentage to the total the irreversibility of the AAC system using R1234yf. Ex d were always the compressor. Thus, this certain type component needs to be paid more attention to increase FIGURE 5. Comparison of exergetic performance of AAC system using R1234yf and R134a and contribution percentages of components: (a) the system with T cond,ai =3 o C (Set1) and (b) the system with T cond,ai =4 o C (Set2) 5. CONCLUSSIONS This paper deals with experimental evaluation and comparison of the steady-state performance parameters of an AAC system employing R1234yf and R134a for different ambient temperature. For this aim, it was tested at four different compressor speeds, namely 1, 15, 2, and 25 rpm, and for each compressor speed, the temperatures of the air streams at the inlets of the inside duct of 3 C and varies the outside duct, i.e. T cond,ai, of 3, 35, and 4 C. The safety of R1234yf for further design of vehicle use can be confirmed by lower the pressure ratio across the compressor and compressor discharge temperature of the AAC system. Both the compressor ratio and discharge temperature get higher with the increasing T cond,ai. The mass flow rate of R1234yf is always higher which is positive in terms of compensating the less enthalpy difference in the AAC cycle compared to that of with R134a. However, it is noticed that the TXV needs further investigations since it is designed for R134a. The mass flow rates and power absorbed in the compressor get higher as T cond,ai increase. Higher vapor density of R1234yf provides the reason of its

8 greater mass flow rate compared with R134a. Despite the lower cooling capacity and COP, the performance of AAC system with R1234yf can be seen as comparable to that of R134a. The exergy-based approach reveals that the biggest contributor to the total exergy destruction of the system is always the compressor and the heat exchangers, i.e. evaporator and condenser, respectively. Thus, in order to improve the system performance, those components should be enhanced. There are few other type of compressors that can be investigated in virtue of finding the best suitable replacement for R1234yf Moreover, the exergy destructions in the condenser and compressor get higher as ambient temperature rises due to the increasing temperature of air intake leads the higher mean temperature difference of air and run-through refrigerant, thus, E x d enlarging the heat transfer and so does the. In order to enhance the heat transfer in the heat exchangers, it can be done through enlarging the heat transfer area and/or air flow through the fins. REFERENCES 1. I. Dinçer, M. Kanoglu, Refrigerant System and Application (Wiley, Chennai, 211), pp United Nations Environmental Programme, Montreal Protocol on Substances that Deplete the Ozone Layer, Final Act, United Nations, New York (1987). 3. Y. Lee, D. Jung, Appl. Therm. Eng. 35, (212). 4. European Union Directive 26/4/EC of the European Parliament and of the Council of 17 May 26 relating to emissions from air-conditioning systems in motor vehicles and amending Council Directive 7/156/EC, Official Journal of the European Union (26). 5. A. G. Devecioglu, V. Oruc, Energy Procedia 75, (215). 6. Lewandowski, T., Additional Risk Assessment of Alternative Refrigerant R-1234yf, Gradient Corporation, July 24, 213 http: // 7. M. Hosoz, M. Karabektas, Comparative performance of an automotive air conditioning system using R1234yf and R134a 13th International Conference on Sustainable Energy Technologies, Paper ID: SET214-E482, Geneva, Switzerland (214). 8. A. Mota-Babiloni, J. Navarro-Esbri, A. Barragan, F. Moles, B. Peris, Appl. Therm. Eng. 71, (214). 9. G. Mathur, SAE Technical Paper, DOI:1.4271/ M.J. Moran, H.N. Shapiro, Fundamentals of Engineering Thermodynamics, 5th ed. (John Wiley and Sons, West Sussex, 26), pp N.A. Ansari, B. Yadav, J. Kumar, Int. J. Sci. Eng. 4, (213). 12. A. Yataganbaba, A. Kilicarslan, I. Kurtbas, 215. Int. J. Refrigeration 6, (215). 13. A. Alkan, M. Hosoz, Int. J. of Refrig. 33, (21). 14. O. Ozgener, A. Hepbasli, Energ Buildings, 39, (27). 15. H. Cho, H. Lee, C. Park, C. Appl. Therm. Eng., 61, (213). 16. M. Koban., SAE Technical Paper, DOI:1.4271/ APPENDIX FIGURE A1. Comparison of Actual Cycle of AAC System Using R1234yf And R134a for 1 rpm