- Email: [email protected]
Experimental investigation on minichannel parallel flow condenser performance with R22, R410A and R407C
Accepted Manuscript Title: Experimental investigation on minichannel parallel flow condenser performance with r22, R410A and R407C Author: Zhaogang Qi PII: DOI: Reference:
S0140-7007(16)30232-8 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.07.021 JIJR 3394
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
2-3-2016 21-5-2016 25-7-2016
Please cite this article as: Zhaogang Qi, Experimental investigation on minichannel parallel flow condenser performance with r22, R410A and R407C, International Journal of Refrigeration (2016), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.07.021. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Experimental Investigation on Minichannel Parallel Flow Condenser Performance with R22, R410A and R407C Zhaogang Qi* Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, No.800, Dongchuan Road, Shanghai 200240, China. *Corresponding author. Email: [email protected] , [email protected] (Z.G. Qi). Tel: +86-2134206259, Fax: +86-21-34206814.
Highlights
The performance of minichannel parallel flow (MCPF) condenser was evaluated.
The performance of MCPF condenser with R22, R410A and R407C were compared.
R410A condenser showed the best performance in heat rejection and pressure drop.
Abstract: A performance evaluation of minichannel parallel flow (MCPF) condenser in residential/commercial refrigeration system has been carried out in calorimeter room with wind tunnel in this paper. The heat rejection and pressure drop characteristics for heat exchangers were compared using R22, R410A and R407C as working fluids. The experimental results showed that heat rejection of MCPF condenser with R410A was higher than that of R22 and R407C by 15.6~26.3% and 12.3~22.7% under full and partial load conditions, respectively. The refrigerant side pressure drop trend of R410A in MCPF condenser was smaller than that of R22 and R407C under the same mass flow rate. Keywords: Minichannel, condenser, refrigerant, R410A, R407C
1. Introduction Refrigerant alternatives and phase-out process are always affected economically and politically. After the F-gas Regulation was approved in 2006 (European Commission,
Page 1 of 17
2006), the European Union (EU) has reviewed the Regulation in 2014 and decided to limit almost all the important F-gas in a relative low level from 2015 (European Commission, 2014). Mota-Babiloni et al. (2015) summarized the contents of EU's Regulation on refrigerants and analyzed the different refrigerant mixtures by experiments and theoretical calculations. They pointed out that currently there was no excellent mixture solution for high GWP refrigerant alternatives and R410A could be still used in commercial chillers (GWPR410A=2088<2050) (European Commission, 2014; Forster et al., 2007). R410A and R407C are the common alternatives for R22 whose efficiency were effected by subcooling and superheating temperatures (Sencan et al., 2006; Yang and Wu, 2013). For the fundamental condensation heat transfer coefficient and pressure drop, R410A showed a similar condensation heat transfer coefficient while R407C was 11-15% lower than that of R22 in a 9.52mm horizontal plain tube (Jung et al., 2004). The experimental results in 1.088mm and 1.289mm (inner diameter) round micro-tubes showed that condensation heat transfer coefficient of R410A was larger than that of R22 and R407C meanwhile pressure drop was smaller (Zhang et al., 2012). The authors concluded that R410A would be better than R407C as a substitute for R22. The overall heat transfer coefficient of R407C in a shell-and-tube condenser was decreased up to 70% comparing that of R22, especially at low mass flow rate (Gabrielii and Vamling, 1997). In the system level, Lee et al. (2002) studied round tube and fin condenser for R22 and R407C experimentally and numerically. The condenser performance of R22 was better than that of R407C in Z-type path configuration, but almost the same in U-type path configuration. R410A showed the advantages on system performance and Life Cycle
2 Page 2 of 17
Climate Performance (LCCP) characteristics in a medium-temperature air conditioning (AC) application (Spatz and Motta, 2004). The similar system performance was achieved in a “drop-in” test of R407C replacing R22 in air source heat pump under frosting and defrosting conditions (Liu et al., 2008). Chen (2008) experimentally studied four splittype residential ACs with R22 and R410A. It was concluded that the system coefficient of performance (COP) of R410A was about 20% higher than that of R22 system and meanwhile the system efficiency was not sensitive to refrigerant side pressure drop because of high operation pressure. Barve and Cremaschi (2012) carried out a “drop-in” experiments of R32 and HFO1234yf in a 5-ton R410A heat pump system which was commercially available. It was found that R32 had a comparable system performance both in heating and cooling modes as those of R410A but this feature would be deteriorated under the extremely high ambient temperature. R1234yf system COP was similar with R410A but cooling capacity was rather lower although the expansion valve has been optimized. Before an excellent solution coming to the market, more enhanced technologies should be developed for R410A and R407C applications to improve system COP and to reduce green-house gas emission. The experimental data showed that there was a significant reduction in refrigerant mass inside tube but refrigerant side pressure drop would be increased obviously when the round tube was flattened (Wilson et al., 2003). Compact heat exchanger with less internal volume could be a good way to reduce the total refrigerant charge amount (Wilson et al., 2003; Qi et al., 2010). Kim and Bullard (2003) experimentally studied microchannel condensers for a residential window AC system and compared with four conventional finned round-tube heat exchangers. The results revealed that heat transfer
3 Page 3 of 17
rates per unit core volume of microchannel condenser were higher those of the conventional ones by 14-331%. Park and Hrnjak (2008) experimentally studied the performances of two identical R410A residential AC systems, one with microchannel condenser and the other with round tube and fin condenser. The test results showed that COP and cooling capacity of microchannel condenser system were higher than that of round tube system while the two condenser had the same package volumes under ARI A, B and C conditions. On the other hand, the system refrigerant charge with microchannel condenser was smaller than that with round tube condenser by 9.2%. Brignoli et al. (2013) experimentally studied an air-water heat pump with micro-channel heat exchanger comparing with round tube heat exchanger, but the results showed that the system performance was penalized by the refrigerant mal-distribution inside microchannel heat exchanger. The mal-distribution of minichannel tube heat exchanger is a real obstacle for the future application in heat pump system (Qi, 2014) although this problem has attracted the academic attentions and focus in recent years (Zou et al., 2014; Li and Hrnjak, 2015a; Li and Hrnjak, 2015b). After a comprehensive literature review, Kandlikar (2007) summarized that design, performance evaluation and optimization of evaporators and condensers for specific products were strongly required for residential and commercial AC units. The present study compared the performance of minichannel parallel flow (MCPF) condenser used in residential/commercial air conditioning system under cooling mode using R22, R410A and R407C as working fluids. The heat exchanger was tested in a calorimeter room with wind tunnel. The condenser heat rejection and pressure drop
4 Page 4 of 17
characteristics under full and partial load conditions with variables refrigerants would be analyzed.
2. MCPF Condenser and Experimental Facility The MCPF condenser with corrugated louver fins in this paper was designed for residential/commercial AC unit. The condenser size was a little big. The prototype was subdivided into three slabs and assembled in a mobile AC system manufacturer as show in Fig. 1. The detailed condenser geometries were listed in Table 1. The slab was consisted of 31 extruded minichannel tubes and 32 corrugated louver fins. Each slab had 3 inlet/outlet pipes for uniform refrigerant distribution. Because of large air volume flow rate and heat rejection, a calorimeter test chamber for bus air conditioning system was used as shown in Fig. 2. The constant temperature and humidity environment was supplied by the additional heater/chiller and humidifier units. The extended duct with the tested condenser and fan was connected with the original wind tunnel as shown in Fig. 3. The air side inlet and outlet temperatures through the tested condenser were measured by sampling unit and thermocouple matrix, respectively. The air side pressure drop was measured by a differential pressure transducer and the air volume flow rate was calculated according to the nozzle pressure difference. The refrigerant side temperature and pressure were maintained by adjusting compressor speed and expansion valve and were measured by thermocouples and pressure transducers. The refrigerant mass flow rate was recorded by a mass flow meter. The uncertainty of heat transfer rate was ±3% based on Moffat’s method (Moffat, 1988). The heat balance of the tested condenser during the tests was shown in Fig. 4. The accuracy of air side temperature and pressure drop were ±0.15℃ and ±2.0Pa, respectively. 5 Page 5 of 17
The accuracy of refrigerant side temperature and pressure were ±0.15℃ and ±10.0kPa, respectively. The accuracy of refrigerant side mass flow rate was ±1.2kg hr-1. All data would be recorded in a computer through the data acquisition system. Experiments were performed under three different system load conditions: full load, 75% full load and 50% full load. The refrigerants were R22, R410A and R407C. The condensation temperature were maintained at about 52.0℃, 48.0℃ and 45.0℃ for full load, 75% full load and 50% full load, respectively (During the tests, the condensation temperature is not controlled directly. The refrigerant pressure at condenser inlet is controlled according to REFPROP V9.1 (Lemmon et al., 2013). So the condensation temperature is changed a little because of the moderate controlling precision and pressure drop in condenser). The average superheating and subcooling at condenser inlet and outlet were 21.0℃ and 5.0℃, respectively. The compressor and evaporator condition would be adjusted for the above refrigerant side conditions and system balance. The inlet air dry and wet-bulb temperatures were maintained at 35°C and 26°C, respectively. The air side volume flow rates were 9500, 10500 and 11500 m3 hr-1.
3. Results and Discussion During the present study, the condensation temperature was mainly controlled indirectly and kept at the same level for 3 refrigerants. Fig. 5 showed the heat rejection of MCPF condenser with variable refrigerants under 100% full load conditions. It was indicated that condenser with R410A showed the highest heat rejection under all three air volume flow rates. The heat rejection was a little constant with air volume increasing under this condition since air side thermal resistance reduction was very small although
6 Page 6 of 17
the air face velocity was increased from 2.56m s-1 (9500m3 hr-1) to 3.10m s-1(11500m3 hr1
). The average heat rejection of R410A condenser was larger than that of R22 and
R407C condenser by 21.9% and 12.3%, respectively. Heat rejection of R407C condenser was larger than that of R22 by 8.6%. As shown in Fig. 5(b), all the refrigerant side pressure drops were increased with refrigerant mass flow rate. As the saturated properties of R22, R410A and R407C at 55℃ from REFPROP V9.1 (Lemmon et al., 2013), R410A has the best thermodynamic and transport properties among the three refrigerants. It was implied that the mass flow rate of R410A condenser was the largest with the same heat rejection because of the smallest latent heat among three refrigerants. From the trend development of pressure drop with refrigerant mass flow rate, it seemed the refrigerant side pressure drop of R410A was smaller than that of R22 condenser under the same mass flow rate. Fig. 6 showed the condenser performance under 75% full load conditions. The heat rejection characteristics was the same with that of full load conditions. R410A condenser had the largest heat rejection and it was higher than that of R22 and R407C condenser by 26.3% and 22.7%, respectively. And heat rejection difference of R407C and R22 became close and it was larger than that of R22 by 2.9%. The refrigerant mass flow rates and pressure drop trend were the same with that under 100% full load conditions. As shown in Fig. 7, the average heat rejection of R410A still was larger than that of R22 and R407C by 15.6% and 20.5% under 50% full load conditions, respectively. But under this condition, heat rejection of R407C was smaller than that of R22 condenser by 4.1%. Here one of the possible explanations is the temperature glide of R407C negatively influence heat transfer performance more under 50% full load conditions. As refrigerant
7 Page 7 of 17
side pressure drop shown in Fig. 7(b), it should be pointed out that the refrigerant mass flow rate is about 0.09 kg/s under 50% full load conditions for three refrigerants and the measured refrigerant side pressure drop is less than 10 kPa which is beyond the accuracy of pressure transducer in the present study. At these conditions, the uncertainty would be a problem. As Zhang et al. (2012) work, R410A showed the better heat transfer and pressure drop characteristics and thermodynamic and transport properties than R22 and R407C. Through the above comparisons, R410A MCPF condenser showed the better performance than that in R22 and R407C. It was revealed that the advantages of minichannel tube are not only in single tube but also in a real condenser application. It was concluded that MCPF condenser could be applied in residential/commercial AC units to improve system efficiency for the next refrigerant alternatives (Kim and Bullard, 2003; Park and Hrnjak, 2008). Fig. 8 showed all the air side pressure drops during three conditions. It was implied that the pressure drop of air side was only influenced by air volume flow rate which was a quadratic relationship with air face velocity as the solid curve in Fig. 8 and heat transfer performance has little influences on air side pressure drop.
4. Conclusions A performance evaluation of minichannel parallel flow (MCPF) condenser in residential/ commercial air conditioning system was performed using R22, R410A and R407C as working fluids. The tests were carried out in a calorimeter room under full load, 75% and 50% full load conditions. The experimental results showed that R410A condenser had a higher heat rejection amount than R22 and R407C condenser by 8 Page 8 of 17
15.6~26.3% and 12.3~22.7%, respectively. Meanwhile R407C condenser had a better heat transfer performance than R22 condenser under full and 75% full load conditions but smaller under 50% full load conditions. From the results, the condenser heat rejection was kept at constant level when air face velocity was increased from 2.56m s-1 to 3.10m s-1 because air side thermal resistance changes were very small. Although latent heat of R410 was smaller and refrigerant mass flow rate was larger, refrigerant pressure drop development trend in R410A MCPF condenser was smaller at the same mass flow rate because of a better transport properties. It was concluded that MCPF condenser could be applied in residential/ commercial AC units to improve system efficiency when R22 was replaced by R410A and R407C.
Acknowledgement The author is grateful for the financial supports from the National Natural Science Foundation of China (Grant No. 51406113).
References Barve, A., Cremaschi, L., 2012. Drop-in Performance of Low GWP Refrigerants in a Heat Pump System for Residential Applications. In: Proceedings of International Refrigeration and Air Conditioning Conference at Purdue, West Lafayette, IN, July 1619, 2012, Paper No. 2197. Brignoli, R., Cecchinato, L., Zilio, C., 2013. Experimental analysis of an air-water heat pump with micro-channel heat exchanger. Appl. Therm. Eng. 50, 1119-1130. Chen, W., 2008. A comparative study on the performance and environmental characteristics of R410A an R22 residential air conditioners. Appl. Therm. Eng. 28, 1-7. 9 Page 9 of 17
European Commission, 2006. Directive 2006/40/EC of The European Parliament and of the Council of 17 May 2006 relating to emissions from air conditioning systems in motor vehicles and amending Council Directive 70/156/EC. Official Journal of the European Union. European Commission, 2014. Regulation (EU) No 517/2014 of the European Parliament and the Council of 16 April 2014 on fluorinated greenhouse gases and repealing Regulation (EC) No 842/2006. Official Journal of the European Union. Forster, P., V. Ramaswamy, P. Artaxo, T. Berntsen, R. Betts, D.W. Fahey, J. Haywood, J. Lean, D.C. Lowe, G. Myhre, J. Nganga, R. Prinn, G. Raga, M. Schulz and R. Van Dorland, 2007: Changes in Atmospheric Constituents and in Radiative Forcing. In: Climate Change 2007: The Physical Science Basis. Contribution of Working Group I to the Fourth Assessment Report of the Intergovernmental Panel on Climate Change [Solomon, S., D. Qin, M. Manning, Z. Chen, M. Marquis, K.B. Averyt, M.Tignor and H.L. Miller (eds.)]. Cambridge University Press, Cambridge, United Kingdom and New York, NY, USA. Gabrielii, C., Vamling, L., 1997. Replacement of R22 in tube-and-shell condensers: experiments and simulations. Int. J. Refrig. 20, 165-178. Jung, D., Cho, Y., Park, K., 2004. Flow condensation heat transfer coefficients of R22, R134a, R407C, and R410A inside plain and microfin tubes. Int. J. Refrig. 27, 25-32. Kandlikar, S., 2007. A Roadmap for Implementing Minichannels in Refrigeration and Air-Conditioning Systems-Current Status and Future Directions. Heat Transfer Eng. 28, 973-985.
10 Page 10 of 17
Kim, M., Bullard, C., 2003. Performance evaluation of a window room air conditioner with microchannel condensers. J. Energy Res. Technol. 124, 47-55. Lee, J., Bae, S., Bang, K., Kim, M., 2002. Experimental and numerical research on condenser performance for R-22 and R-407C refrigerants. Int. J. Refrig. 25, 372-382. Lemmon, E., Huber, M., McLinden, M., 2013. Reference Fluid Thermodynamic and Transport Properties - REFPROP Ver. 9.1. 2013, National Institute of Standards and Technology, Boulder, CO, USA. Li H.Z., Hrnjak, P., 2015a. Quantification of liquid refrigerant distribution in parallel flow microchannel heat exchanger using infrared thermography. Appl. Therm. Eng. 78, 410-418. Li H.Z., Hrnjak, P., 2015b. An experimentally validated model for microchannel heat exchanger incorporating lubricant effect. Int. J. Refrig. 59, 259-268. Liu, Z., Li, X., Wang, H., Peng, W., 2008. Performance comparison of air source heat pump with R407C and R22 under frosting and defrosting. Energy Convers. Manage. 49, 32-239. Moffat, R., 1988. Describing the Uncertainties in Experimental Results. Exp. Therm. Fluid Sci. 1, 3-17. Mota-Babiloni, A., Navarro-Esbrí, J., Barragán-Cervera, Á., Molés, F., Peris B., 2015. Analysis based on EU Regulation No 517/2014 of new HFC/HFO mixtures as alternatives of high GWP refrigerants in refrigeration and HVAC systems. Int. J. Refrig. 52, 21-31.
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Park, C., Hrnjak P., 2008. Experimental and numerical study on microchannel and roundtube condensers in a R410A residential air-conditioning system. Int. J. Refrig. 31,822831. Qi, Z., 2014. Advances on air conditioning and heat pump system in electric vehicles – A review. Renew. Sustain. Energy Rev. 38, 754–764. Qi, Z., Zhao, Y., Chen, J., 2010. Performance enhancement study of mobile air conditioning system using microchannel heat exchangers. Int. J. Refrig. 33, 301-312. Sencan, A., Selbas, R., Kizilkan, O., Kalogirou, S., 2006. Thermodynamic analysis of subcooling and superheating effects of alternative refrigerants for vapor compression refrigeration cycles. Int. J. Energy Res. 30, 323-347. Spatz, M., Motta, S., 2004. An evaluation of options for replacing HCFC-22 in medium temperature refrigeration systems. Int. J. Refrig. 27, 475-483. Wilson M., Newell, T., Chato, J., Ferreira, C., 2003. Refrigerant charge, pressure drop, and condensation heat transfer in flattened tubes. Int. J. Refrig. 26, 442-451. Yang, Z., Wu, X., 2013. Retrofits and options for the alternatives to HCFC-22. Energy 59, 1-21. Zou, Y., Tuo, H.F., Hrnjak P., 2014. Modeling refrigerant maldistribution in microchannel heat exchangers with vertical headers based on experimentally developed distribution results. Appl. Therm. Eng. 64, 172-181. Zhang, H., Li, J., Liu, N., Wang, B., 2012. Experimental investigation of condensation heat transfer and pressure drop of R22, R410A and R407C in mini-tubes. Int. J. Heat Mass Transfer 55, 3522-3532.
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a. Single slab
b. Sketch of the tested condenser
Fig. 1 Slab and condenser assembly
Fig. 2 Schematic diagram of calorimeter test room
a. Condenser with fan
b. Connection with wind tunnel
Fig. 3 Modifications on wind tunnel 13 Page 13 of 17
Refrigerant Side Heat Rejection(kW)
45.0
40.0
R407C R410A R22
+5% -5%
35.0
30.0
25.0
20.0
15.0 15.0
20.0
25.0
30.0
35.0
40.0
45.0
Air Side Heat Rejection (kW)
Fig. 4 Energy balance between air and refrigerant side
35.0
Refri. Side Pressure Drop (kPa)
43.00
Heat Rejection (kW)
41.00 39.00
37.00 35.00
33.00 31.00
R407C
29.00 8500
9500
R410A
10500
R22
11500
Air Volume Flow Rate (m3 hr-1)
a. Heat rejection
12500
30.0
25.0
20.0
15.0
R407C
10.0 600.0
650.0
700.0
R410A
R22
750.0
800.0
Refri. Mass Flow Rate (kg hr-1)
b. Refrigerant side pressure drop
Fig. 5 MCPF condenser performance at full load conditions
14 Page 14 of 17
30.0
Refri. Side Pressure Drop (kPa)
Heat Rejection (kW)
33.00 31.00 29.00 27.00 25.00 23.00 21.00 8500
R407C
9500
R410A
10500
R22
11500
25.0
20.0 15.0 10.0 R407C
5.0 0.0 400.0
12500
Air Volume Flow Rate (m3 hr-1)
450.0
500.0
550.0
R410A
600.0
R22
650.0
700.0
Refri. Mass Flow Rate (kg hr-1)
a. Heat rejection
b. Refrigerant side pressure drop
22.00
16.0
21.00
14.0
Refri. Side Pressure Drop (kPa)
Heat Rejection (kW)
Fig. 6 MCPF condenser performance at 75% full load conditions
20.00 19.00
18.00 17.00
16.00
R407C
15.00 8500
9500
R410A
10500
R22
11500
Air Volume Flow Rate (m3 hr-1)
a. Heat rejection
12500
12.0 10.0 8.0 6.0 4.0 R407C
R410A
R22
2.0 0.0 300.0
320.0
340.0
360.0
380.0
400.0
Refri. Mass Flow Rate (kg hr-1)
b. Refrigerant side pressure drop
Fig. 7 MCPF condenser performance at 50% full load conditions
15 Page 15 of 17
Air Side Pressure Drop (Pa)
55.0
50.0
R407C R410A R22
45.0
40.0
35.0
30.0 8500
9500
10500
11500
12500
Air Volume Flow Rate (m3 hr-1)
Fig. 8 Air side pressure drop under different volume flow rates
Table 1. The detailed geometries of the tested condenser Parameters (Unit)
Size
Core width (mm)
1042
Core height (mm)
987
Core depth (mm)
26
Minichannel tube(mm×mm)
26.0×2.0 with 12 ports
Port size (mm×mm)
1.2×1.75
Fin width (mm)
26.0
Fin height (mm)
8.52
Fin pitch (mm)
1.4
Fin thickness (mm)
0.1
Louver length (mm)
7.1
Louver pitch (mm)
1.1
Louver angle (o)
27
16 Page 16 of 17
Passes Configuration
1st pass:62; 2nd pass:31
17 Page 17 of 17
S0140-7007(16)30232-8 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.07.021 JIJR 3394
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
2-3-2016 21-5-2016 25-7-2016
Please cite this article as: Zhaogang Qi, Experimental investigation on minichannel parallel flow condenser performance with r22, R410A and R407C, International Journal of Refrigeration (2016), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.07.021. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Experimental Investigation on Minichannel Parallel Flow Condenser Performance with R22, R410A and R407C Zhaogang Qi* Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, No.800, Dongchuan Road, Shanghai 200240, China. *Corresponding author. Email: [email protected] , [email protected] (Z.G. Qi). Tel: +86-2134206259, Fax: +86-21-34206814.
Highlights
The performance of minichannel parallel flow (MCPF) condenser was evaluated.
The performance of MCPF condenser with R22, R410A and R407C were compared.
R410A condenser showed the best performance in heat rejection and pressure drop.
Abstract: A performance evaluation of minichannel parallel flow (MCPF) condenser in residential/commercial refrigeration system has been carried out in calorimeter room with wind tunnel in this paper. The heat rejection and pressure drop characteristics for heat exchangers were compared using R22, R410A and R407C as working fluids. The experimental results showed that heat rejection of MCPF condenser with R410A was higher than that of R22 and R407C by 15.6~26.3% and 12.3~22.7% under full and partial load conditions, respectively. The refrigerant side pressure drop trend of R410A in MCPF condenser was smaller than that of R22 and R407C under the same mass flow rate. Keywords: Minichannel, condenser, refrigerant, R410A, R407C
1. Introduction Refrigerant alternatives and phase-out process are always affected economically and politically. After the F-gas Regulation was approved in 2006 (European Commission,
Page 1 of 17
2006), the European Union (EU) has reviewed the Regulation in 2014 and decided to limit almost all the important F-gas in a relative low level from 2015 (European Commission, 2014). Mota-Babiloni et al. (2015) summarized the contents of EU's Regulation on refrigerants and analyzed the different refrigerant mixtures by experiments and theoretical calculations. They pointed out that currently there was no excellent mixture solution for high GWP refrigerant alternatives and R410A could be still used in commercial chillers (GWPR410A=2088<2050) (European Commission, 2014; Forster et al., 2007). R410A and R407C are the common alternatives for R22 whose efficiency were effected by subcooling and superheating temperatures (Sencan et al., 2006; Yang and Wu, 2013). For the fundamental condensation heat transfer coefficient and pressure drop, R410A showed a similar condensation heat transfer coefficient while R407C was 11-15% lower than that of R22 in a 9.52mm horizontal plain tube (Jung et al., 2004). The experimental results in 1.088mm and 1.289mm (inner diameter) round micro-tubes showed that condensation heat transfer coefficient of R410A was larger than that of R22 and R407C meanwhile pressure drop was smaller (Zhang et al., 2012). The authors concluded that R410A would be better than R407C as a substitute for R22. The overall heat transfer coefficient of R407C in a shell-and-tube condenser was decreased up to 70% comparing that of R22, especially at low mass flow rate (Gabrielii and Vamling, 1997). In the system level, Lee et al. (2002) studied round tube and fin condenser for R22 and R407C experimentally and numerically. The condenser performance of R22 was better than that of R407C in Z-type path configuration, but almost the same in U-type path configuration. R410A showed the advantages on system performance and Life Cycle
2 Page 2 of 17
Climate Performance (LCCP) characteristics in a medium-temperature air conditioning (AC) application (Spatz and Motta, 2004). The similar system performance was achieved in a “drop-in” test of R407C replacing R22 in air source heat pump under frosting and defrosting conditions (Liu et al., 2008). Chen (2008) experimentally studied four splittype residential ACs with R22 and R410A. It was concluded that the system coefficient of performance (COP) of R410A was about 20% higher than that of R22 system and meanwhile the system efficiency was not sensitive to refrigerant side pressure drop because of high operation pressure. Barve and Cremaschi (2012) carried out a “drop-in” experiments of R32 and HFO1234yf in a 5-ton R410A heat pump system which was commercially available. It was found that R32 had a comparable system performance both in heating and cooling modes as those of R410A but this feature would be deteriorated under the extremely high ambient temperature. R1234yf system COP was similar with R410A but cooling capacity was rather lower although the expansion valve has been optimized. Before an excellent solution coming to the market, more enhanced technologies should be developed for R410A and R407C applications to improve system COP and to reduce green-house gas emission. The experimental data showed that there was a significant reduction in refrigerant mass inside tube but refrigerant side pressure drop would be increased obviously when the round tube was flattened (Wilson et al., 2003). Compact heat exchanger with less internal volume could be a good way to reduce the total refrigerant charge amount (Wilson et al., 2003; Qi et al., 2010). Kim and Bullard (2003) experimentally studied microchannel condensers for a residential window AC system and compared with four conventional finned round-tube heat exchangers. The results revealed that heat transfer
3 Page 3 of 17
rates per unit core volume of microchannel condenser were higher those of the conventional ones by 14-331%. Park and Hrnjak (2008) experimentally studied the performances of two identical R410A residential AC systems, one with microchannel condenser and the other with round tube and fin condenser. The test results showed that COP and cooling capacity of microchannel condenser system were higher than that of round tube system while the two condenser had the same package volumes under ARI A, B and C conditions. On the other hand, the system refrigerant charge with microchannel condenser was smaller than that with round tube condenser by 9.2%. Brignoli et al. (2013) experimentally studied an air-water heat pump with micro-channel heat exchanger comparing with round tube heat exchanger, but the results showed that the system performance was penalized by the refrigerant mal-distribution inside microchannel heat exchanger. The mal-distribution of minichannel tube heat exchanger is a real obstacle for the future application in heat pump system (Qi, 2014) although this problem has attracted the academic attentions and focus in recent years (Zou et al., 2014; Li and Hrnjak, 2015a; Li and Hrnjak, 2015b). After a comprehensive literature review, Kandlikar (2007) summarized that design, performance evaluation and optimization of evaporators and condensers for specific products were strongly required for residential and commercial AC units. The present study compared the performance of minichannel parallel flow (MCPF) condenser used in residential/commercial air conditioning system under cooling mode using R22, R410A and R407C as working fluids. The heat exchanger was tested in a calorimeter room with wind tunnel. The condenser heat rejection and pressure drop
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characteristics under full and partial load conditions with variables refrigerants would be analyzed.
2. MCPF Condenser and Experimental Facility The MCPF condenser with corrugated louver fins in this paper was designed for residential/commercial AC unit. The condenser size was a little big. The prototype was subdivided into three slabs and assembled in a mobile AC system manufacturer as show in Fig. 1. The detailed condenser geometries were listed in Table 1. The slab was consisted of 31 extruded minichannel tubes and 32 corrugated louver fins. Each slab had 3 inlet/outlet pipes for uniform refrigerant distribution. Because of large air volume flow rate and heat rejection, a calorimeter test chamber for bus air conditioning system was used as shown in Fig. 2. The constant temperature and humidity environment was supplied by the additional heater/chiller and humidifier units. The extended duct with the tested condenser and fan was connected with the original wind tunnel as shown in Fig. 3. The air side inlet and outlet temperatures through the tested condenser were measured by sampling unit and thermocouple matrix, respectively. The air side pressure drop was measured by a differential pressure transducer and the air volume flow rate was calculated according to the nozzle pressure difference. The refrigerant side temperature and pressure were maintained by adjusting compressor speed and expansion valve and were measured by thermocouples and pressure transducers. The refrigerant mass flow rate was recorded by a mass flow meter. The uncertainty of heat transfer rate was ±3% based on Moffat’s method (Moffat, 1988). The heat balance of the tested condenser during the tests was shown in Fig. 4. The accuracy of air side temperature and pressure drop were ±0.15℃ and ±2.0Pa, respectively. 5 Page 5 of 17
The accuracy of refrigerant side temperature and pressure were ±0.15℃ and ±10.0kPa, respectively. The accuracy of refrigerant side mass flow rate was ±1.2kg hr-1. All data would be recorded in a computer through the data acquisition system. Experiments were performed under three different system load conditions: full load, 75% full load and 50% full load. The refrigerants were R22, R410A and R407C. The condensation temperature were maintained at about 52.0℃, 48.0℃ and 45.0℃ for full load, 75% full load and 50% full load, respectively (During the tests, the condensation temperature is not controlled directly. The refrigerant pressure at condenser inlet is controlled according to REFPROP V9.1 (Lemmon et al., 2013). So the condensation temperature is changed a little because of the moderate controlling precision and pressure drop in condenser). The average superheating and subcooling at condenser inlet and outlet were 21.0℃ and 5.0℃, respectively. The compressor and evaporator condition would be adjusted for the above refrigerant side conditions and system balance. The inlet air dry and wet-bulb temperatures were maintained at 35°C and 26°C, respectively. The air side volume flow rates were 9500, 10500 and 11500 m3 hr-1.
3. Results and Discussion During the present study, the condensation temperature was mainly controlled indirectly and kept at the same level for 3 refrigerants. Fig. 5 showed the heat rejection of MCPF condenser with variable refrigerants under 100% full load conditions. It was indicated that condenser with R410A showed the highest heat rejection under all three air volume flow rates. The heat rejection was a little constant with air volume increasing under this condition since air side thermal resistance reduction was very small although
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the air face velocity was increased from 2.56m s-1 (9500m3 hr-1) to 3.10m s-1(11500m3 hr1
). The average heat rejection of R410A condenser was larger than that of R22 and
R407C condenser by 21.9% and 12.3%, respectively. Heat rejection of R407C condenser was larger than that of R22 by 8.6%. As shown in Fig. 5(b), all the refrigerant side pressure drops were increased with refrigerant mass flow rate. As the saturated properties of R22, R410A and R407C at 55℃ from REFPROP V9.1 (Lemmon et al., 2013), R410A has the best thermodynamic and transport properties among the three refrigerants. It was implied that the mass flow rate of R410A condenser was the largest with the same heat rejection because of the smallest latent heat among three refrigerants. From the trend development of pressure drop with refrigerant mass flow rate, it seemed the refrigerant side pressure drop of R410A was smaller than that of R22 condenser under the same mass flow rate. Fig. 6 showed the condenser performance under 75% full load conditions. The heat rejection characteristics was the same with that of full load conditions. R410A condenser had the largest heat rejection and it was higher than that of R22 and R407C condenser by 26.3% and 22.7%, respectively. And heat rejection difference of R407C and R22 became close and it was larger than that of R22 by 2.9%. The refrigerant mass flow rates and pressure drop trend were the same with that under 100% full load conditions. As shown in Fig. 7, the average heat rejection of R410A still was larger than that of R22 and R407C by 15.6% and 20.5% under 50% full load conditions, respectively. But under this condition, heat rejection of R407C was smaller than that of R22 condenser by 4.1%. Here one of the possible explanations is the temperature glide of R407C negatively influence heat transfer performance more under 50% full load conditions. As refrigerant
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side pressure drop shown in Fig. 7(b), it should be pointed out that the refrigerant mass flow rate is about 0.09 kg/s under 50% full load conditions for three refrigerants and the measured refrigerant side pressure drop is less than 10 kPa which is beyond the accuracy of pressure transducer in the present study. At these conditions, the uncertainty would be a problem. As Zhang et al. (2012) work, R410A showed the better heat transfer and pressure drop characteristics and thermodynamic and transport properties than R22 and R407C. Through the above comparisons, R410A MCPF condenser showed the better performance than that in R22 and R407C. It was revealed that the advantages of minichannel tube are not only in single tube but also in a real condenser application. It was concluded that MCPF condenser could be applied in residential/commercial AC units to improve system efficiency for the next refrigerant alternatives (Kim and Bullard, 2003; Park and Hrnjak, 2008). Fig. 8 showed all the air side pressure drops during three conditions. It was implied that the pressure drop of air side was only influenced by air volume flow rate which was a quadratic relationship with air face velocity as the solid curve in Fig. 8 and heat transfer performance has little influences on air side pressure drop.
4. Conclusions A performance evaluation of minichannel parallel flow (MCPF) condenser in residential/ commercial air conditioning system was performed using R22, R410A and R407C as working fluids. The tests were carried out in a calorimeter room under full load, 75% and 50% full load conditions. The experimental results showed that R410A condenser had a higher heat rejection amount than R22 and R407C condenser by 8 Page 8 of 17
15.6~26.3% and 12.3~22.7%, respectively. Meanwhile R407C condenser had a better heat transfer performance than R22 condenser under full and 75% full load conditions but smaller under 50% full load conditions. From the results, the condenser heat rejection was kept at constant level when air face velocity was increased from 2.56m s-1 to 3.10m s-1 because air side thermal resistance changes were very small. Although latent heat of R410 was smaller and refrigerant mass flow rate was larger, refrigerant pressure drop development trend in R410A MCPF condenser was smaller at the same mass flow rate because of a better transport properties. It was concluded that MCPF condenser could be applied in residential/ commercial AC units to improve system efficiency when R22 was replaced by R410A and R407C.
Acknowledgement The author is grateful for the financial supports from the National Natural Science Foundation of China (Grant No. 51406113).
References Barve, A., Cremaschi, L., 2012. Drop-in Performance of Low GWP Refrigerants in a Heat Pump System for Residential Applications. In: Proceedings of International Refrigeration and Air Conditioning Conference at Purdue, West Lafayette, IN, July 1619, 2012, Paper No. 2197. Brignoli, R., Cecchinato, L., Zilio, C., 2013. Experimental analysis of an air-water heat pump with micro-channel heat exchanger. Appl. Therm. Eng. 50, 1119-1130. Chen, W., 2008. A comparative study on the performance and environmental characteristics of R410A an R22 residential air conditioners. Appl. Therm. Eng. 28, 1-7. 9 Page 9 of 17
European Commission, 2006. Directive 2006/40/EC of The European Parliament and of the Council of 17 May 2006 relating to emissions from air conditioning systems in motor vehicles and amending Council Directive 70/156/EC. Official Journal of the European Union. European Commission, 2014. Regulation (EU) No 517/2014 of the European Parliament and the Council of 16 April 2014 on fluorinated greenhouse gases and repealing Regulation (EC) No 842/2006. Official Journal of the European Union. Forster, P., V. Ramaswamy, P. Artaxo, T. Berntsen, R. Betts, D.W. Fahey, J. Haywood, J. Lean, D.C. Lowe, G. Myhre, J. Nganga, R. Prinn, G. Raga, M. Schulz and R. Van Dorland, 2007: Changes in Atmospheric Constituents and in Radiative Forcing. In: Climate Change 2007: The Physical Science Basis. Contribution of Working Group I to the Fourth Assessment Report of the Intergovernmental Panel on Climate Change [Solomon, S., D. Qin, M. Manning, Z. Chen, M. Marquis, K.B. Averyt, M.Tignor and H.L. Miller (eds.)]. Cambridge University Press, Cambridge, United Kingdom and New York, NY, USA. Gabrielii, C., Vamling, L., 1997. Replacement of R22 in tube-and-shell condensers: experiments and simulations. Int. J. Refrig. 20, 165-178. Jung, D., Cho, Y., Park, K., 2004. Flow condensation heat transfer coefficients of R22, R134a, R407C, and R410A inside plain and microfin tubes. Int. J. Refrig. 27, 25-32. Kandlikar, S., 2007. A Roadmap for Implementing Minichannels in Refrigeration and Air-Conditioning Systems-Current Status and Future Directions. Heat Transfer Eng. 28, 973-985.
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Kim, M., Bullard, C., 2003. Performance evaluation of a window room air conditioner with microchannel condensers. J. Energy Res. Technol. 124, 47-55. Lee, J., Bae, S., Bang, K., Kim, M., 2002. Experimental and numerical research on condenser performance for R-22 and R-407C refrigerants. Int. J. Refrig. 25, 372-382. Lemmon, E., Huber, M., McLinden, M., 2013. Reference Fluid Thermodynamic and Transport Properties - REFPROP Ver. 9.1. 2013, National Institute of Standards and Technology, Boulder, CO, USA. Li H.Z., Hrnjak, P., 2015a. Quantification of liquid refrigerant distribution in parallel flow microchannel heat exchanger using infrared thermography. Appl. Therm. Eng. 78, 410-418. Li H.Z., Hrnjak, P., 2015b. An experimentally validated model for microchannel heat exchanger incorporating lubricant effect. Int. J. Refrig. 59, 259-268. Liu, Z., Li, X., Wang, H., Peng, W., 2008. Performance comparison of air source heat pump with R407C and R22 under frosting and defrosting. Energy Convers. Manage. 49, 32-239. Moffat, R., 1988. Describing the Uncertainties in Experimental Results. Exp. Therm. Fluid Sci. 1, 3-17. Mota-Babiloni, A., Navarro-Esbrí, J., Barragán-Cervera, Á., Molés, F., Peris B., 2015. Analysis based on EU Regulation No 517/2014 of new HFC/HFO mixtures as alternatives of high GWP refrigerants in refrigeration and HVAC systems. Int. J. Refrig. 52, 21-31.
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Park, C., Hrnjak P., 2008. Experimental and numerical study on microchannel and roundtube condensers in a R410A residential air-conditioning system. Int. J. Refrig. 31,822831. Qi, Z., 2014. Advances on air conditioning and heat pump system in electric vehicles – A review. Renew. Sustain. Energy Rev. 38, 754–764. Qi, Z., Zhao, Y., Chen, J., 2010. Performance enhancement study of mobile air conditioning system using microchannel heat exchangers. Int. J. Refrig. 33, 301-312. Sencan, A., Selbas, R., Kizilkan, O., Kalogirou, S., 2006. Thermodynamic analysis of subcooling and superheating effects of alternative refrigerants for vapor compression refrigeration cycles. Int. J. Energy Res. 30, 323-347. Spatz, M., Motta, S., 2004. An evaluation of options for replacing HCFC-22 in medium temperature refrigeration systems. Int. J. Refrig. 27, 475-483. Wilson M., Newell, T., Chato, J., Ferreira, C., 2003. Refrigerant charge, pressure drop, and condensation heat transfer in flattened tubes. Int. J. Refrig. 26, 442-451. Yang, Z., Wu, X., 2013. Retrofits and options for the alternatives to HCFC-22. Energy 59, 1-21. Zou, Y., Tuo, H.F., Hrnjak P., 2014. Modeling refrigerant maldistribution in microchannel heat exchangers with vertical headers based on experimentally developed distribution results. Appl. Therm. Eng. 64, 172-181. Zhang, H., Li, J., Liu, N., Wang, B., 2012. Experimental investigation of condensation heat transfer and pressure drop of R22, R410A and R407C in mini-tubes. Int. J. Heat Mass Transfer 55, 3522-3532.
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a. Single slab
b. Sketch of the tested condenser
Fig. 1 Slab and condenser assembly
Fig. 2 Schematic diagram of calorimeter test room
a. Condenser with fan
b. Connection with wind tunnel
Fig. 3 Modifications on wind tunnel 13 Page 13 of 17
Refrigerant Side Heat Rejection(kW)
45.0
40.0
R407C R410A R22
+5% -5%
35.0
30.0
25.0
20.0
15.0 15.0
20.0
25.0
30.0
35.0
40.0
45.0
Air Side Heat Rejection (kW)
Fig. 4 Energy balance between air and refrigerant side
35.0
Refri. Side Pressure Drop (kPa)
43.00
Heat Rejection (kW)
41.00 39.00
37.00 35.00
33.00 31.00
R407C
29.00 8500
9500
R410A
10500
R22
11500
Air Volume Flow Rate (m3 hr-1)
a. Heat rejection
12500
30.0
25.0
20.0
15.0
R407C
10.0 600.0
650.0
700.0
R410A
R22
750.0
800.0
Refri. Mass Flow Rate (kg hr-1)
b. Refrigerant side pressure drop
Fig. 5 MCPF condenser performance at full load conditions
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30.0
Refri. Side Pressure Drop (kPa)
Heat Rejection (kW)
33.00 31.00 29.00 27.00 25.00 23.00 21.00 8500
R407C
9500
R410A
10500
R22
11500
25.0
20.0 15.0 10.0 R407C
5.0 0.0 400.0
12500
Air Volume Flow Rate (m3 hr-1)
450.0
500.0
550.0
R410A
600.0
R22
650.0
700.0
Refri. Mass Flow Rate (kg hr-1)
a. Heat rejection
b. Refrigerant side pressure drop
22.00
16.0
21.00
14.0
Refri. Side Pressure Drop (kPa)
Heat Rejection (kW)
Fig. 6 MCPF condenser performance at 75% full load conditions
20.00 19.00
18.00 17.00
16.00
R407C
15.00 8500
9500
R410A
10500
R22
11500
Air Volume Flow Rate (m3 hr-1)
a. Heat rejection
12500
12.0 10.0 8.0 6.0 4.0 R407C
R410A
R22
2.0 0.0 300.0
320.0
340.0
360.0
380.0
400.0
Refri. Mass Flow Rate (kg hr-1)
b. Refrigerant side pressure drop
Fig. 7 MCPF condenser performance at 50% full load conditions
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Air Side Pressure Drop (Pa)
55.0
50.0
R407C R410A R22
45.0
40.0
35.0
30.0 8500
9500
10500
11500
12500
Air Volume Flow Rate (m3 hr-1)
Fig. 8 Air side pressure drop under different volume flow rates
Table 1. The detailed geometries of the tested condenser Parameters (Unit)
Size
Core width (mm)
1042
Core height (mm)
987
Core depth (mm)
26
Minichannel tube(mm×mm)
26.0×2.0 with 12 ports
Port size (mm×mm)
1.2×1.75
Fin width (mm)
26.0
Fin height (mm)
8.52
Fin pitch (mm)
1.4
Fin thickness (mm)
0.1
Louver length (mm)
7.1
Louver pitch (mm)
1.1
Louver angle (o)
27
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Passes Configuration
1st pass:62; 2nd pass:31
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