COMPACT HEAT EXCHANGERS FOR MOBILE CO 2 SYSTEMS

Transcription

1 COMPACT HEAT EXCHANGERS FOR MOBILE CO 2 SYSTEMS A. Hafner SINTEF Energy Research Refrigeration and Air Conditioning Trondheim (Norway) ABSTRACT The natural refrigerant carbon dioxide (CO 2 ) with all its advantages offers new possibilities of efficient heating and cooling at different climates. In reversible air conditioning systems the capacity and efficiency in cooling mode is seen to be most important. The efficiency of the reversible interior heat exchanger depends on the design of the heat exchanger. Efficiency reduction is among other factors caused by refrigerant- side pressure drop, heat conduction, refrigerant- and air maldistribution. Uniform air temperatures at the outlet of the heat exchanger are an important aspect regarding comfort and control of the mobile HVAC system. A prototype CO 2 system with a separator, refrigerant pump and a single pass heat exchanger was tested under varies conditions. The tests were performed at low compressor revolution speed i.e. idle conditions. Compared with a baseline heat exchanger of equal size, the use of an external operated pump to circulate the refrigerant through the heat exchanger, increased the cooling capacity at equal pressure levels by up to 14 %. At equal cooling capacities the COP increased by up to 23 %. Airside temperature distribution was more uniform with this way of operating the system. 1 INTRODUCTION Several authors for example Hrnjak et al 22, Nekså et al. 21 describe CO 2 systems where flash gas was bypassed the evaporator and thereafter reunified with the leaving gas of the evaporator. With such a system concept only liquid refrigerant enters the evaporator which has an positive h, Wm-2K G = 19 G = 28 G = 38 G = vapour fraktion Figure 1.1 Influence of the mass flux on the heat transfer coefficient, from Pettersen. 22. effect on distribution of the refrigerant into the microchannels. Studies performed at SINTEF Energy Research (Figure 1.1) show a dramatic reduction on heat transfer coefficients inside microchannel tubes at higher vapour fraction and mass flux. On the other hand the heat transfer coefficients at lower vapour fraction had been in the range of 12. W/m 2 K nearly independent of the mass flux. This was the basis for this study of an evaporator inside an entire CO 2 system, where an external pump was used to circulate the refrigerant through the evaporator. In this case the refrigerant side heat transfer should not drop and the capacity of the heat exchanger should increase.

2 2 MEASUREMENT EQUIPMENT The test facility consists of two climate chambers with wind tunnels containing the prototype heat exchangers, and a compressor drive arrangement. (Hafner, 2). The refrigerant circuiting around the interior heat exchanger was modified and a gear pump was used to circulate the liquid CO 2 from a separator. The baseline evaporator was a prototype CO 2 microchannel heat exchanger with 2 rows and 7 passes. This was a unit built from extruded microchannel tubes of 41mm width (Pettersen et al., 1998). To get a single pass heat exchanger with reduced pressure drop in mind, the connectors of a similar heat exchanger (2 rows 6 passes, 1% reduced front surface) was connected in a way to get a equally bottom fed 2row-1pass heat exchanger. Exterior HX High side pressure Suction line HX control valve Separator Compressor A made-to order sight-glass was used as the liquid separator / lowpressure receiver in the CO 2 circuit. The internal volume was 1 dm 3. The separator/receiver, the gear pump and the mass flow meter were located inside the climate chamber. The revolution speed of the gear pump was controlled with a 12V DC power supply. The maximum refrigerant mass flow rate of the pump at 5 bar evaporation pressure was around 3.2 kg/min. In front of the separator a video camera was installed to observe the liquid level inside. The compressor for the test program was a 2.7 cm 3 reciprocating prototype machine Evaporator The prototype CO 2 microchannel parallel-flow heat exchanger served as the exterior heat exchanger. Gear pump Mass flow meter m Figure 2.1 Flow circuit of the CO 2 prototype system The suction line heat exchanger used was an extruded aluminium profile made by Hydro Aluminium. The total length of the heat exchanger was 1.5 m and it was bent into a U-shape, so the inlet and outlet were side by side. A manual metering valve was used as the expansion device during the test series. A grid of 2 thermocouples was installed 1 cm behind the interior heat exchanger to measure temperature distribution on the air side.

3 3 EXPERIMENTS AND RESULTS Results from earlier measurements of the 2row-7pass heat exchanger were used as baseline for the comparison. The baseline CO 2 system included a conventional low-pressure receiver. In the same CO 2 system the 2row-7pass heat exchanger was replaced by a 2row-1pass heat exchanger (identical tubes, fins and headers), a liquid separator and an external controlled refrigerant pump (Fig.2.1) were added to perform the test program. The performance of the 2row-1pass was studied at varying refrigerant mass flow rates and a test with natural refrigerant circulation was performed. Most of the tests were performed at an air inlet temperature to the interior heat exchanger of 41 C at a relative humidity of %. The air mass flow rate through the interior heat exchanger was kept to 8.1 kg/min (42m 3 /h). During most of the tests the high side pressure was controlled to be at 1 bar, except the test where the optimum high side pressure was investigated. Under these COP-tests, the airside velocity of the exterior heat exchanger was 2.4 m/s at a temperature of 39 C. The baseline test data where taken with a high performance gascooler, therefore the inlet vapour fraction to the evaporator was x=.2, see Figure 3.1. Equal airside temperature reduction at: equal air mass flow rates, equal air face velocities equal evaporation pressure were applied in the direct comparisons between the heat exchangers. 12 baseline 1 pass kg/min Pressure [bar] Evaporator inlet outlet Pressure [bar] Evaporator inlet outlet Specific enthalpy [kj/kg] Specific enthalpy [kj/kg] Figure 3.1 Process values of the systems in a p, h-diagram. Equal air side capacity at 8.1 kg/min air mass flow rate. Figure 3.2 Process values of the pump circulating system in a p, h-diagram. Equal air side capacity at 8.1 kg/min air mass flow rate and varying refrigerant mass flow rates. A refrigerant liquid overfeeding factor through the evaporator of 2 was reached with the gear pump at maximum revolution speed; see the leftmost columns of Figure 3.3. The refrigerant mass flow rate through the compressor was in the same range as for the baseline test at 1.55 kg/min during the tests at equal air face velocity and equal air mass flow rate through the evaporator, while it was 15% higher for the test at equal evaporation pressure.

4 The refrigerant side pressure drop was 3.6 times lower for the tests with pump circulation even with twice the refrigerant mass flow rate through the evaporator (3.1 kg/min) ref. mass flow rate, compressor pressure drop ref. mass flow rate, evap. vapour fraction, evap. outlet Refrigerant mass flow rate [kg/min] vapour fraction separator inlet: Pressure drop [bar] / Vapour fraction [-]. 7 pass baseline eq_air side velocity eq._air mass flow rate equal_p 3.2 kg/min (ref) 2.6 kg/min (ref) 2 kg/min (ref) natural ref. circ. = 8.1 kg/min; t air, interior = 41 C (% rel. hum.) Figure 3.3 Refrigerant mass flow rate (through evaporator dark filled bar +amount through compressor lighter filled bar), refrigerant side pressure, vapour fraction at inlet to separator and outlet from the evaporator for the different test series.. Figure 3.4 shows the effect of liquid overfeeding on the evaporation pressure. For the same capacity it raised from 5 bar for the baseline system to 54 bar for the tests at equal air face velocity and air mass flow rate. When keeping the evaporation pressure at the same level the cooling capacity increased by 14% at an equal amount of air through the evaporators. In addition, the air inlet temperature to the passenger compartment could be further reduced from 25.2 C to 22.5 C. The COP of the system including pump power for this case was 6% higher as can be seen in Fig. 3.5 (fist group of columns on the leftmost side). For tests at equal air face velocity the COP was 19% higher while it increased with 23% at equal air mass flow rates. The amount of condensed water was reduced by 3-4 g/min for the tests at equal air face velocity and equal air mass flow rate. It increased a few grams/minute for the test at equal evaporation pressure. Unfortunately, the humidifier of the test rig was by that time not able to keep the inlet relative humidity at %, only 38% was reached.

5 6 evaporating pressure air temp. reduction 19 Evaporation pressure [bar] Air side temperature reduction [K] 7 pass baseline eq_air side velocity eq._air mass flow rate equal_p 3.2 kg/min (ref) 2.6 kg/min (ref) 2 kg/min (ref) natural ref. circ. = 8.1 kg/min; t air, interior = 41 C (% rel. hum.) 13 Figure 3.4 Evaporation pressure and air side temperature reduction for the different test series. ϕ air in : 38 %* 1% humidif ier capacity condensed water COP 4 Condensed water [g/min] System COP 7 pass baseline eq_air side velocity eq._air mass flow rate equal_p 3.2 kg/min (ref) 2.6 kg/min (ref) 2 kg/min (ref) natural ref. circ. = 8.1 kg/min; t air, interior = 41 C (% rel. hum.*) Figure 3.5 Amount of condensed water from the evaporator and system COP for the different test series. The performance of the 2row-1pass heat exchanger was studied at varying refrigerant mass flow rates. Figure 3.2 shows the process in a p, h-diagram for a test series with different refrigerant mass flow rates through the evaporator. The airside temperature reduction was 15.5 K for each test.

6 The refrigerant mass flow rate through the compressor was approximately 2.4 kg/min, as shown in Figure 3.3 (center group of columns). The exit vapour fraction increased when less refrigerant was pumped through the heat exchanger. The refrigerant side pressure drop increased by 38 % from.8 to.11 bar when increasing the mass flow rate from 2 to 3.2 kg/min. The evaporation pressure went up from 49 to 55.5 bar still reaching the equal air exit temperatures (Fig. 3.2 & Fig. 3.4). The amount of condensed water drained from the evaporator went down from 36 to 22 g/min when the refrigerant mass flow rate was increased. The system COP increased by 19% from the lowest to the highest refrigerant mass flow rate. A test with natural refrigerant circulation was performed with the 2row-1pass heat exchanger, the distance (altitude difference) between separator liquid level and evaporator inlet was about.7 m. It was not possible to reach equal capacity like obtained before. The evaporation pressure was limited to a minimum of 35 bar, to prevent the evaporator from freezing. However, the results are shown in the rightmost columns of Fig. 3.3, Fig. 3.4 & Fig The shape of the optimum COP curve versus high side pressure was investigated at constant airside capacity of 3.8 kw. Figure 3.6 shows that maximum COP under these conditions was obtained when the high side pressure reached 113 bar. The evaporation pressure was 53 bar for all the tests in this series. Cooling capacity Pow er cons. comp. COP Temp. ex. device m_ref, interior HX vapour fraction m_ref, comp P / P max.cop ; Q o / Q o,max COP ; COP / max COP = 8.1 kg/min v air, exterior = 2.4 m/s High side pressure [bar] Temp. throttling valve inlet [ C]. t air, interior = 41 C t air, exterior = 39 C Figure 3.6 Relative shaft power to the compressor, rel. COP and refrigerant temperature at expansion valve inlet versus the high side pressure at constant cooling capacity. Constant airside conditions on the gascooler side. Ref. mass flow rate [kg/min] High side pressure [bar] = 8.1 kg/min v air, exterior = 2.4 m/s Ref. vapour fraction sep. inlet [-]. t air, interior = 41 C t air, exterior = 39 C Figure 3.7 Refrigerant mass flow rates and refrigerant vapour fraction at separator inlet versus the high side pressure at constant cooling capacity. During these tests the air inlet temperature of the exterior heat exchanger was 39 C, and the air face velocity was 2.4 m/s. The reduced refrigerant mass flow rate through the compressor at higher discharge pressure was the reason for the reduction of the refrigerant inlet temperature to the expansion device, as shown in Fig. 3.6 & Fig. 3,7. The refrigerant mass flow rate through the interior heat exchanger was around 3 kg/min.

7 Air side temperature distribution was measured with a grid of 2 thermocouples right after the air outlet of the interior heat exchanger during all tests. Figure 3.8 shows the air temperatures after the baseline heat exchanger. The upper part on the Baseline 2row-7pass heat exchanger left hand side had the lowest temperatures while C top upper lower the bottom part to the left hand side had up to 1 K higher temperatures. It shows also the direction of refrigerant flow through the heat exchanger, entering the heat exchanger where 4 tubes are in parallel, changing the row on the right hand side and leaving the heat exchanger after passing through the last 8 parallel tubes. left air flow CO 2 2row-1pass heat exchanger C left air flow left air flow right bottom air inlet: 41 C, % rel.hum.; 8.1 kg/min Figure 3.8 Air side temperature distribution of the baseline heat exchanger top upper lower bottom right number of parallel number tubes of tubes CO 2 bottom feed, se also figure 2.1 Figure 3.9 Air side temperature distribution of the 2row-1pass heat exchanger, pump circulation 3.2 kg/min 2row-1pass heat exchanger C C CO 2 bottom feed, se also figure 2.1 right number of parallel of tubes tubes top upper lower bottom number number of parallel tubes of tubes Figure 3.1 Air side temperature distribution of the 2row-1pass heat exchanger, natural circulation 1.29 kg/min The 2row-1pass heat exchanger was bottom fed. Due to the original 2row-6pass design of the heat exchanger, the separation baffles were still inside the manifolds. Therefore the refrigerant lines were connected to each corner-connector of the heat exchanger (see also Figure 2.1). The temperature distribution on the airside was more vertical than with the baseline system. A hot spot was observed in most of the test in the upper part on the left hand side, as shown in Figure 3.9. During the natural circulation test most of the refrigerant passed the heat exchanger through the first tubes on each side. The air stream in the middle of the heat exchanger was not cooled at all on the upper part, as shown in Figure DISCUSSION The three different test conditions, equal air face velocity, air mass flow rate and equal evaporation pressure were chosen to establish a platform for the comparison of the two heat exchanger concepts. The power input to the gear pump was in the range of 3 W. Since the entire heat exchanger was connected as a single pass heat exchanger with 42 tubes in parallel (35 microchannes / tube) the mass flux was around 7 kg/m 2 s and the Reynolds Number at 7 at the entrance of the channels for the refrigerant mass flow rate of 3 kg/min. Nucleate boiling was the dominating heat transfer on the inside of the microchannel tubes. Due to the very low refrigerant side pressure drop the

8 temperature on the tube surface was quite equal. However, an area of warmer air leaving the heat exchanger at the upper part on the left hand side. One reason might be the still existing baffles inside the manifolds causing some restriction the refrigerant flow, on the other hand some microchannels might be blocked by brazing material from that time the heat exchanger was build. The performance reached at the natural circulation test showed that this kind of concept will be depended on a sufficient liquid column. In this case a height of m would have been sufficient. A special design of a tube and manifold combination is needed to ensure proper refrigerant distribution on the inlet plus a way to remove the vapour with a very low-pressure drop in mind. The airside temperature distribution was more uniform for the tests with high refrigerant overfeeding of the evaporator. Due to a better distribution of the refrigerant inside the heat exchanger the wall temperatures are more uniform and no superheated areas were present. 5 CONCLUSION A prototype CO 2 system with a separator, refrigerant pump and a single pass heat exchanger was tested under varied conditions. The tests were performed at low compressor revolution speed i.e. idle conditions. Compared with a baseline heat exchanger of equal size, the use of an external operated pump to circulate the refrigerant through the heat exchanger, increased the cooling capacity at equal pressure levels by up to 14 %. At equal cooling capacities the COP increased by up to 23 %, including pump power. Airside temperature distribution was more uniform when the refrigerant overfeeding rate was high. A new generation of heat exchanger, designed for this kind of operation is under construction. REFERENCES Pettersen, J., 22. Flow Vaporization of CO 2 in Microchannel Tubes, Dr. Techn. Thesis, Faculty of Engineering Science and Technology, Norwegian University of Science and Tech., February 22. Pettersen, J., Hafner, A. Skaugen, G. Rekstad, H., 1998, Development of compact heat exchangers for CO2 air-conditioning systems. International Journal of Refrigeration, Vol. 21, No. 3 pp , Hafner, A. Pettersen, J. Skaugen G. Nekså, P An Automotive HVAC System with CO 2 as the Refrigerant. IIR-Gustav Lorentzen Conference on Natural Working Fluids, Oslo, Norway (1998) June 2-5, Hrnjak, P. 22 Some issues in R744 Heat Exchangers. Proceedings of the VDA Alternative Refrigerant Winter Meeting 22, Saalfelden, Austria (22), January Neksaa, P. Pettersen, J. and Skaugen G. (21): Heat Transfer and Pressure Drop of Evaporating CO 2 in Microchannels, and System Design Implications of Critical Heat Flux Conditions,21 ASME Int. Mech. Eng. Congress, Nov , New York, US. Hafner, A. 2, Experimental Study on Heat Pump Operation of Prototype CO 2 mobile air conditioning system,. IIR-4 th Gustav Lorentzen Conference on Natural Working Fluids, Purdue, USA (2) July 25-28,