Experimental Study on the Pressure and the Power Input of a CO2 Air Conditioning System

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Experimental Study on the Pressure and the Power Input of a CO2 Air Conditioning System

Phuduc Nguyen1,a, Kimhang Vo1,b, and Thanhtrung Dang*1,c  

 

1 – Department of Thermal Engineering, HCMC University of Technology and Education, Vietnam

a – phuducspkttphcm@gmail.com

b – hangvk@cntp.edu.vn

c – trungdangt@hcmute.edu.vn

DOI 10.2412/ provided by Seo4U.link

 

Keywords: air conditioning, CO2 refrigerant, pressure, power input, evaporator, heat transfer.

 

ABSTRACT. This paper presented investigations on the pressure and the power input of a CO2 air conditioning system. In this study, the cycle changes from the vapour region to the two phase region for the conventional compressor and CO2 compressor, respectively. The suction and discharge pressures increase when the ambient temperature increases. When the ambient temperature increases from 31.8 °C to 40.5 °C, the compressor current increases from 2 A to 2.3 A. The suction pressure and the compressor current decrease as reducing the balance pressure of the system. Comparison between the gas cooler with heat transfer area of 6 m2 and the gas cooler with heat transfer area of 3 m2, the gas cooler with high heat transfer area has increased the cooling capacity of the system.

 

Introduction. The environmentally friendly refrigerants and high effectiveness heat exchangers are interesting topics for scientists. For natural refrigerants, the carbon dioxide (R744-CO2) is considered as a good candidate in order to replace HCFC (Hydrochlofluorocarbon) or HFC (Hydrofluorocarbon). Besides, the compact heat exchangers with high heat transfer efficiency would be used widely in the future. Regarding to CO2 and compact heat exchangers, Baheta el at. [1] simulated performance of transcritical carbon dioxide refrigeration cycle by using Excel program. In this study, the highest Coefficient of Performance (COP) was 3.24 at 10MPa gas cooler pressure. The results indicated that COP increases as rising the evaporative temperature. Using numerical simulation, Cheng and Thome [2] studied on cooling of microprocessors using flow boiling of CO2 in a micro-evaporator. Based on the analysis and comparison, CO2 appeared to be a promising coolant for microprocessors at low operating temperatures but also presented a great technological challenge like other new cooling technologies. However, the investigations in [1, 2] did not experimentally perform. An overview of the flow boiling heat transfer characteristics and the special thermo-physical properties of CO2 in a horizontal tube was investigated by Zhao and Bansal [3]. Due to the large surface tension, the boiling heat transfer coefficient of CO2 was found to be much lower at low temperatures but it increased with vapor quality (until dry out). However, this study was only reviewed for horizontal tube. With technical advantages [4], the carbon dioxide could be highlighted its high heat transfer coefficients in the supercritical region and its high pressure levels combined with low specific volumes. A comprehensive review of flow boiling heat transfer and two-phase flow of CO2  covers both macro-channel tests and micro-channel investigations was presented Thome and Ribatski [5]. The results showed that CO2 gives higher heat transfer coefficients than those of conventional refrigerants. Cheng et al. [6, 7] updated flow pattern map for CO2 evaporation inside tubes. The updated map was applicable for a wider range of conditions: tube diameters from 0.6 to 10 mm, mass velocities from 50 to 1500 kg/m2s, heat fluxes from 1.8 to 46 kW/m2, and saturation temperatures from – 28 to +25 oC. The new CO2 two-phase flow pressure drop model predicted the CO2 pressure drop was better than the former methods.

Ducoulombier et al. [8] studied carbon dioxide two-phase flow pressure drops in a single horizontal stainlesssteel micro-tube having the inner diameter of 0.529 mm. The apparent viscosity of the two-phase mixture was larger than the liquid viscosity at low vapor qualities, namely at the lowest temperatures. Boiling heat transfer of carbon dioxide inside a small-sized microfin tube was investigated by Dang et al. [9]. The experimental results indicated that heat flux has a significant effect on the heat transfer coefficient and the coefficient does not always increase with mass flux. In addition, the experimental results also shown that using microfin tubes may considerably increase the overall heat transfer performance. Design optimisation of CO2 gas cooler/condenser in a refrigeration system was done Ge et al. [10]. In this study, the design optimisation of the heat exchanger dealt with different structure designs, controls and system integration at different operating conditions in order to significantly enhance the performance in a CO2 refrigeration system. As a result, the effect of heat exchanger sizes on system performance can be enhanced with fan speed controls. Numerical analysis use the finite volume method on a microchannel evaporator for CO2 air-conditioning systems was fulfilled by Yun et al. [11]. Dang et al. [12-14] investigated the heat transfer and pressure drop phenomena of the microchannel and minichannel heat exchangers, both numerically and experimentally. At the same average velocity of water in the channels used in this study, the effectiveness obtained from the microchannel heat exchanger was 1.2 to 1.53 times of that obtained from the minichannel heat exchanger. Moreover, influences of gravity to heat transfer and pressure drop behaviors of the microchannel heat exchanger were presented by variation of the physical inclinations of the microchannel heat exchanger system used for experiments. However, in [12-14], the pure water was the working fluid; they did not study CO2 in these studies. Effect of inlet configuration on the refrigerant distribution in a parallel flow minichannel heat exchanger was studied by Kim et al. [15]; however, this study dealt with the refrigerant R134a, not CO2, as the working fluid.

From literature reviews above, the pressures such as suction pressure or discharge pressure, the power input and the ambient temperature did not indicate clearly. So, it is important to experiment on the pressure and the power input of CO2 air conditioning system for enhancing the performance. In this study, the cycle will be discussed with transcritical mode. With this system, an aluminium minichannel evaporator and two gas coolers were used to get thermodynamic parameters of the air conditioning cycle. In addition, the cycle using conventional compressor will be compared the cycle using CO2 compressor.

Methodology. Experimental setup.

The experimental test loop for CO2 air conditioning system is shown in Fig.1. This cycle has four main components: a CO2 compressor, a gas cooler, a throttle valve, and an evaporator. In this study, a conventional compressor and a CO2 compressor were used. The compressors belong to the reciprocating types. Aluminium minichannels were used to manufacture for the evaporator, as shown in Fig. 2. The desgin cooling capacity for this minichannel evaporator is 2700 W. For the cooler, the copper tubes with the diameter of 6.4 mm were used in the study. The gas cooler and evaporator were tested with the hydraulic testing method. The former and the latter did not tear or deform at the pressure of 150 bars and 90 bar, respectively. A photo for this test loop is shown in Fig. 3. Accuracies and ranges of testing apparatus are listed in Table 1 and equipments used for the experiments are listed as follows:

– Thermocouples, T-types

– Thermostat, EW – 181 H, made by Ewelly

– Infrared thermometer, AT 430L2, made by APECH

– Infrared thermometer, Raynger@ST, made by Raytek

– Thermal camera, Fluke Ti9, made by Fluke, USA

– Pressure gauge, made by Pro – Instrument

– Anemometer, AVM-03, made by Prova

– Clamp meter, Kyoritsu 2017, made by Kyoritsu.

 

Fig. 1. Schematic of the test loop for CO2 air conditioning system.

Fig. 2. Dimensions of the minichannel evaporator.

 

Table 1. Accuracies and ranges of testing apparatuses.

Testing apparatus Accuracy Range
Thermocouples ± 0.1 °C 0 ~100 °C
Thermal camera 2% -20~250°C
Infrared thermometer ± 1 °C of reading – 32 ~ 400 °C
Pressure gauge ± 1 FS 0~100 kgf/cm2
Clamp meter ± 1.5 % rdg 0 ~ 200 A
Anemometer ± 3 % 0 ~ 45 m/s
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

Fig. 3. A photo of the CO2 air conditioning system.

 

Governing equations. To analyze the thermodynamic properties of the CO2 air conditioning system, the governing equations were given below:

The heat transfer rate for gas cooler was calculated as  

(1)

The power input was determined using  

(2)

The isenthalpic process was presented by  

(3)

The heat transfer rate for evaporation was calculated as  

(4)

Finally the COP of the cycle was quantified by

 

(ignoring the heat transfer rate of the subcooling process by the throtle valve installed near the indoor unit)

where is the mass flow rate of carbon dioxide.

 

(5)

 

Results and discussion.

The CO2 air conditioning system was tested more times in order to collect the stable data for transcritical mode. The conventional compressor was made by TECUMSEH, with the power input of 200 W. For this compressor, relationships between the balance pressure of the system (when the system has not worked yet) versus the suction pressure and the discharge pressure are shown in Fig. 4 under the ambient temperature of 31ºC ¸ 33ºC. It is observed that when the balance pressure increases, the difference pressure between the suction pressure and the discharge pressure is not more distinguishable. At the balance pressure of 40 bar, there is a little bit of difference, it is due to the experimental error. At the low balance pressure, the discharge pressure is near with the balance pressure; however, the discharge pressure is higher than the balance pressure at the high balance pressure. The experimental results show that the current increases from 1.8 A to 3.5 A as increasing the balance pressure from 30 bar to 60 bar. However, the conventional compressor was out of work when the current is over 3.7 A.

Fig. 4. Balance pressure vs. suction pressure and discharge pressure.

 

A comparison on the thermodynamic cycle between the conventional and CO2 compressors in the system is shown in Table 2 and Fig. 5. The Table 2 shows the thermodynamic parameters of the cycle for three cases: case 1 with conventional compressor, case 2 and case 3 with CO2 compressor. The CO2 compressor was made by SANDEN, with the power input of 450 W. For case 2, the throttle valve installed near the outdoor unit; however, the throttle valve installed near the indoor unit for case 3. The data in Table 2 were drawn on the p-h diagram of CO2, using the EES (Engineering Equation Solver) software.

 

Table 2. Thermodynamic parameters of the CO2 cycle with three cases.

Case p1

(bar)

t1

(°C)

p2

(bar)

t2

(°C)

p3

(bar)

t3

(°C)

p4

(bar)

t4

(°C)

w1-2

(kJ/kg)

q4-1

(kJ/kg)

COP
1 20 20.2 45 83.2 45 31.5 20 0 43.5 19.6 0.45
2 34 20.8 86 105 85 40.2 35 -3.1 45.2 89.5 1.99
3 45 9.2 77 55 77 29.4 47 13.5 20.6 150.2 7.28

 

The Fig. 5 shows a change of the cycle for the conventional compressor to CO2 compressor for compressing high pressure. In case 1, the cycle from the conventional compressor belongs the vapour region which has low COP. In case 2 and 3, the evaporation processes belong to the two phase region. In case 3, when the throttle valve is near the indoor unit, the gas cooler outlet from the point 3’ (32.2 °C) moves to the point 3 (29.4 °C). The results show that the case 3 is the highest COP and the lowest power input. The outlet pressure of the evaporator is lower than the inlet pressure of the evaporator. It is due to the pressure drop of the evaporator and the suction force of the compressor. At the evaporator pressure of 47 bar and the cooler pressure of 77 bar, the power input is 20.6 kJ/kg and the COP is 7.28. It is noted that the COP of the cycle is determined for the refrigerant side.

 

Fig. 5. Experimental comparison between the conventional and CO2 compressors.

 

A relationship between the ambient temperature and pressures is shown in Fig. 6. The results show that the suction and discharge pressures increase when the ambient temperature increases. When the ambient temperature increases from 31.8 °C to 40.5 °C, the suction pressure increases from 42 bar to 48 bar and the discharge pressure increases from 80 bar to 84 bar. The increasing pressures affect to increase the compressor current from 2 A to 2.3 A, as shown in Fig. 7.  The results in Figs. 6 and 7 obtained at the balance pressure of 54 bar. Figure 8 shows the ambient temperature versus pressure and current at the balance pressure of 50 bar. With the results in Figs. 6-8, the SANDEN CO2 compressor has the power input of 450W and the gas cooler has the heat transfer area of 3 m2. From Figs. 6-8, it is observed that the suction pressure and the compressor current decrease as reducing the balance pressure of the system. At the same the ambient temperature of 37.5 °C, the cycle with the 54 bar balance pressure has the suction pressure of 47 bar and the current of 2.2 A; while the cycle with the 50 bar balance pressure has the suction pressure of 42 bar and the current of 1.9 A.

 

Fig. 6. Ambient temperature versus pressure at the balance pressure of 54 bar.

 

Fig. 7. Ambient temperature versus compressor current at the balance pressure of 54 bar.

 

With the results in Fig. 9, the SANDEN CO2 compressor has the power input of 450W and the gas cooler has the heat transfer area of 6 m2. When the suction pressure increases from 46 bar to 50 bar, the current decreases from 2.25 A to 2.07 A. The results are the same with the results obtained from the cycle using the gas cooler 3 m2. However, the outlet temperature of the gas cooler 6 m2 is lower 2 °C than that obtained from the gas cooler 3 m2. The gas cooler with high heat transfer area has increased the cooling capacity of the system. From the results shown in Figs. 4-9, they are emphasized that these results contribute to the significant data availability as well as increase efficiency for studies on CO2 air conditioning systems.

 

Fig. 8. Ambient temperature versus pressure and current at the balance pressure of 50 bar.

Fig. 9. Suction pressure versus discharge pressure and compressor current.

 

Summary. Experimental studies on the pressure and the power input of a CO2 air conditioning system were done. The study indicated a change of the cycle for the conventional compressor to CO2 compressor for compressing high pressure: from the vapour region to the two phase region. In this study, at the evaporator pressure of 47 bar and the cooler pressure of 77 bar, the power input is 20.6 kJ/kg and the COP is 7.28 for the refrigerant side.

The suction and discharge pressures increase when the ambient temperature increases. When the ambient temperature increases from 31.8 °C to 40.5 °C, the suction pressure increases from 42 bar to 48 bar and the discharge pressure increases from 80 bar to 84 bar, leading to the compressor current increases from 2 A to 2.3 A.

The suction pressure and the compressor current decrease as reducing the balance pressure of the system. The cycle with the 54 bar balance pressure has the suction pressure of 47 bar and the current of 2.2 A. While the cycle with the 50 bar balance pressure has the suction pressure of 42 bar and the current of 1.9 A, at the same the ambient temperature of 37.5 °C.

Comparison between the cooler with heat transfer area of 6 m2 and the cooler with heat transfer area of 3 m2, the compressor currents are the same; however, the outlet temperature of the gas cooler 6 m2 is lower 2 °C than that obtained from the gas cooler 3 m2. So, the gas cooler with high heat transfer area has increased the cooling capacity of the system.

In summary, the results of pressure and power input provide the significant data for studies on air conditioning systems using the CO2 refrigerant.

Acknowledgment.

The supports of this work by the projects (No. T2017-37 TĐ sponsored by the specific research fields at HCMUTE and No.B2015.22.01 sponsored by Vietnam Ministry of Education and Training) are deeply appreciated.

Nomenclature.

w : power input

q  : heat transfer rate

m: mass flow rate

h   : enthalpy

p   : pressure

t   : temperature

Subscripts.

c      : carbon dioxide

1      : exit of evaporator

2      : exit of compressor

3’      : exit of gas cooler

3     : exit of subcooler

4      : exit of throttle valve

COP: Coefficient of Performance.

References

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[2] Cheng, L. and Thome, J.R., Cooling of microprocessors using flow boiling of CO2 in a micro-evaporator: Preliminary analysis and performance comparison, Applied Thermal Engineering, Vol. 29, 2009, pp. 2426 – 2432

[3] Zhao, X. and Bansal, P.K., Flow boiling heat transfer characteristics of CO2 at low temperatures, International Journal of Refrigeration, Vol. 30, 2007, pp. 937 – 945

[4] Manhoe Kim, Jostein Pettersen, Clark W. Bullard, Fundamental process and system design issues in CO2 vapor compression systems. Progress in Energy and Combustion Science, Vol. 30, 2004, pp. 119 – 174

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[12] Dang, T.T. and Teng, J.T., Comparison on the heat transfer and pressure drop of the microchannel and minichannel heat exchangers, Heat and Mass Transfer, Vol. 47, 2011, pp. 1311-1322

[13] Dang, T.T. and Teng, J.T., The effects of configurations on the performance of microchannel counter-flow heat exchangers – An experimental study, Applied Thermal Engineering, Vol. 31, 2011, pp. 3946-3955

[14] Dang, T.T., et al., A study on the simulation and experiment of a microchannel counter-flow heat exchanger, Applied Thermal Engineering, Vol. 30, 2010, pp. 2163-2172

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