An Investigation of R417a as a Drop-in Alternative for R22 in a Residential Heat Pump

Transcription

1 An Investigation of R417a as a Drop-in Alternative for in a Residential Heat Pump Zhiming Gao *, Viung C Mei, Fang C Chen, John Tomlinson Oak Ridge National Laboratory, Oak Ridge, TN Heat pumps are efficient units for providing heating and cooling in both residential and commercial buildings., however, is still being widely used in most heat pumps produced in the United States, which has been scheduled to phase-out by the end of the year 22. At present, the majority alternatives of consist of hydrofluorocarbon (HFC) refrigerants only. Unfortunately, these HFC replacements require synthetic oil and expensive modification. Recently, R417a, a new alternative of, is presented and does not require synthetic oil because R417a contains 3.4 wt % of R6 (butane). R6 is an excellent oil carrier. The emphasis of this paper is an attempt to document the findings of thermodynamic cycle, heat transfer, and experimental study on R417a serving as a drop-in working fluid in a two-ton split heat pump system. I. Introduction 22 is ozone-depleting dichlorofluoromethane refrigerant, and is still widely served as working fluid in the Rmajority of heat pumps produced in the United States. The Montreal Protocol, however, specifies that the production of hydrochlorofluorocarbons (HCFCs) in developed countries will be prohibited in 22, and the manufacture of equipment using will not be allowed after the year 2. Therefore, the heating, ventilation, and air conditioning industry has been screening out the replacement of for many years. The effort has primarily focused on exploring hydrofluorocarbon (HFC) refrigerant mixtures that possess thermodynamic properties close to. Two such HFC blends, and R4a, have been identified as possible alternatives so far. Each of them has individual advantages and shortcomings. Table 1 lists their properties information in brief. In brief, the two blends meet the necessary criteria to be classified A1/A1, which is the category of lowest risk in terms of toxicity and flammability, both for the formulated composition and for the worst-case leak scenario defined in the ASHRAE standard 34 [1]. Because the HFC refrigerants do not mix well with mineral oil, synthetic oil is required as lubricant in the equipment using them. The operating behavior of R47c and R4a in a heat pump system has been well investigated [2, 3, 4, 5, 6]. These results demonstrate that there is no appreciable improvement in the performance of R47c compared with, although its physical properties are similar to. Two major factors penalizing the system performance of R47c are temperature glide and poor transport properties. Also, as a ternary zeotropic mixture; R47c could lead to the variation of its components if a leak occurs. This poses refrigerant recycling and system recharging problems. Unlike R47c, R4a looks discouraging at first because of its poor theoretical performance, low critical temperature, and high pressure [5]. Specifically, the pressure level of R4a is significantly higher than that of (e.g., the saturated vapor pressure for R4a at 4ºC is almost 6% higher than that of ). Therefore, the application of R4a requires a thorough redesign of the system. But the advantages of using R4a have also been identified -- high heat transfer coefficient and low pressure drop. These advantages allow the use of smaller compressors and pipes in the refrigerating systems. The multi-year Alternative Refrigerant Evaluation Program (AREP) showed R4a optimized systems could achieve 5% higher COPs than systems; R47c optimized systems, on the other hand, have COPs about 5% lower than those of [6]. Recently, R417a, a new alternative of, is presented and does not require synthetic oil because R417a contains 3.4 wt% of R6 (butane). Since R417a is a relatively new blend being considered as an alternative to, data on the detailed properties of R417a are still very limited. However, the available information shows that * ORAU Postdoctoral researcher. Senior researcher, ESTD, Oak Ridge National Laboratory, PO Box 28, MS67. Retired senior researcher, ESTD, Oak Ridge National Laboratory, PO Box 28, MS67. Retired senior researcher, ESTD, Oak Ridge National Laboratory, PO Box 28, MS67. 1

2 R417a has physical properties fairly similar to (see Table 1). The unique feature of R417a is compatible with traditional mineral oils or alkyl benzene lubricants because the 3.4% R6 (butane) in the mixture is an excellent oil carrier. This advantage makes R417a ideal for serving as a true drop-in replacement in existing equipment without any modification. However, there is also very limited data available so far on the cycle performance of R417a []. Aprea and his coauthor [7, 8] investigated the experimental performance of 417a in a vapor compression plant, which served as water chiller and heat pump. R417a was also analyzed by Spatz and Richard [9] for its oil return characteristics. The information is still not enough to reveal or judge the situation of R417a as a drop-in replacement of, especially in residential heat pump systems. R47c R4a R417a Composition CHClF 2 R32/R125/R134A (23%/25%/52%) R32/R125 (5%/5%) ODP* [R11] =1.55 GWP** (a) [CO 2 =1] Glide ( C) at bar Critical temperature ( C) Critical pressure (bar) Bubble pressure at 5 C (bar) Lubricant MO, AB, POE POE POE MO, AB, POE * ODP ozone depleting potential ** GWP global warming potential Table 1. Properties of and alternatives. R125/R134A/R6 (46.6%/5%/3.4%) The effort of the paper is to investigate the application of R417a in a residential heat pump system designed for without any modification. The lubricant is still traditional mineral oil. The detailed work includes the analysis of theoretical and experimental study on R417a performance as a drop-in refrigerant in heat pump system. The theoretical analysis includes both thermodynamic cycle and heat transfer by using the State of Equation and available heat transfer correlations. In the experimental study, both heating-mode and cooling-mode operation were tested in a nominal 2-ton split heat pump system. These tests were conducted in a two-room environmental chamber. The concern regarding the oil return characteristics of R417a was not tested. The results of R417a were further compared with those for for a baseline data sample. Comparisons among R47c, R4a and R417a were also briefly discussed by using experimental conclusion available from the published literatures. II. Thermodynamic Analysis To thoroughly understand the performance of R417a, the analysis of its properties and its theoretical cycle performance is necessary. The calculations were accomplished with the aid of REFPROP, version 6.1, which used a Helmholtz equation (Kamie) and a modified Benedict-Webb-Rubin (MBWR) equation. Figure 1 is a pressure-enthalpy diagram for R417a, which provides general information concerning its properties. The saturate pressure-temperature curves of R417a, together with, R47c, and R4a, are plotted together in Figure 2 that shows the pressure level of R417a to be about % less than that of and significantly less than that of R47c and R4a. The pressure values of, R47c, R4a and R417a are 97 Kpa (159 psi), 1248 Kpa (181 psi), 1734 Kpa (252 psi), and 995 Kpa (144 psi), respectively, at a 3 K (54 R) bubble point. Hence, R417a is the only one of three alternatives, whose pressure level is lower than at the same temperature. As a non-azeotropic refrigerant mixture, R417a also indicates a temperature glide, as found in Fig. 1. However, the glide phenomenon of R417a is not as pronounced as R47c; for example, the glide of R417a at 1.32 Kpa (14.7 psi) is only 4.7 C (4.5 F), while that of R47c is 7.1 C (44.8 F) at the same operating conditions. Fig. 3 illustrates the comparison of theoretical cycle performance among R417a,, R47c, and R4a. The current data are calculated based on the American Refrigerant Institute (ARI) specified operating conditions: 7.2 C (45. F) evaporating temperature, 54.4 C (129.9 F) condensing temperature, 11.1 C (52 F) superheat, and 8.33 C (47. F) subcooling. The evaporating and condensing temperatures are defined as the arithmetic mean of dew point and bubble point. From the view of Figure 3, the theoretical COP of R417a is slightly less than that of. The COP of R417a is almost equivalent to that of R47c. However, the capacity of R417a is quite less than and other alternatives (only around 8% of that of ). As for other parameters, such as pressure ratio and discharge temperature, there is no remarkable difference (less than 5%) between R417a and. Therefore, from the perspective of both properties and theoretical cycle performance, R417a is suitable to serve as a replacement for. 2

3 Figure 1. Pressure enthalpy diagram for R417a. Figure 2. Saturated-pressure curve of, R47c, R4a and R417a. Figure 3. Comparison of theoretical cycle performance for, R47c, R4a, and R417a (Qv is volumetric capacity, Pr is pressure ratio, and Tp is compressor discharge temperature). 3

4 III. Evaluating Pressure Drop and Heat Transfer To further evaluate the application of R417a in heat pump systems, it is also important to figure out its pressure drop and heat transfer performance. Therefore, several semi-empirical correlations are employed to compare the heat transfer coefficient and pressure drop of, R47c, R4a, and R417a in both micro-fin and smooth tubes. The model published by Choi, et. al. [11] was employed to evaluate the pressure drop of both evaporation and condensation. The validation of the model showed an average absolute residual between measurements and predictions not to exceed 17.6% for various data sets including pure and mixing refrigerants. The condensation model for micro-fin tube came from Cavallinni, et. al. [12], which showed a mean absolute percentage deviation of about 17% for 396 experimental data including pure and mixing refrigerants. The condensation model for smooth tube came from Thome, et. al. [14], which pointed out that 85% of heat transfer coefficients for 185 points were within ±2%. The evaporation model for micro-fin tube came from Cavallinni, et. al. [13], which illustrated a mean absolute deviation of 14% for 643 experimental data including pure and mixing refrigerants. The evaporation model for smooth tube came from Thome, et. al. [15], which indicated a mean absolute deviation of about 2% for pure and mixing refrigerants. Condensation Pressure Drop in Micro Tube (KPa) (a) Condensation at micro-fin tube Condensation Pressure Drop in Smooth Tube (KPa) (b) Condensation at smooth tube Evaporation Pressure Drop in Micro Tube (KPa) (c) Evaporation at micro-fin tube Evaporation Pressure Drop in Smooth Tube (KPa) (d) Evaporation at smooth tube Figure 4 Comparison of pressure drop for various alternatives as a function of mass flux. 4

5 12 5 Condensation Heat Transfer Coefficient in Microtube (Kw/m 2 K) Condensation Heat Transfer Coefficient in smooth tube (Kw/m 2 K) (a) Condensation at micro-fin tube (b) Condensation at smooth tube Evaporation Heat Transfer Coefficient in Microtube (Kw/m 2 K) Evaporation Heat Transfer Coefficient in Smooth Tube (Kw/m 2 K) (c) Evaporation at micro-fin tube (d) Evaporation at smooth tube Figure 5 Comparison of heat transfer coefficient for various alternatives as a function of mass flux. To reflect the real system operating condition, an actual operating condition obtaining from the experimental data was chosen to conduct the simulation. The refrigerant flows for condensation are between 2kg/m 2 s and 5kg/m 2 s. The refrigerant flows for evaporation are between 2kg/m 2 s and 45kg/m 2 s. The tube diameters are 8mm for evaporation and 9.3mm for condensation in the case of smooth tube. In the case of micro-fin tube, the fintip diameters of micro-fin tube are 7.6mm for evaporation and 8.9mm for condensation. Meanwhile, the fin height is.2mm. In the simulation, heat flux rates are assumed as 15.6kw/m 2 for 46 o C condensation and 12kw/m 2 for o C evaporation. Also it need to note that the prediction here is conducted at vapor quality = This is because heat transfer correlations show low accuracy at vapor quality <.1 or >.9. The detailed result is plotted in figures 4 and 5. Figure 4 shows the pressure drop of, R47c, R4a, and R417a as a function of mass flux rate. It is found that the pressure drop of R417a is less than and R47c at both evaporation and condensation process. But it s obviously higher than R4a. R47c is pretty close to. Based on the result plotted in Figure 4, the ranking of low pressure drop for and its alternatives is listed as R4a, R417a, and R47c. Figure 5 shows the average heat transfer for and its alternatives. The average heat transfer is conducted at vapor quality between.1.9. Whenever it is evaporation or condensation at both micro-fin and smooth tube, R417a is the worst at heat transfer performance among, R47c, R4a, and R417a. The heat transfer performance of R47c is slightly higher than R417a. From the view of the figures, the ranking of better heat transfer for and its alternatives is R4a,, R47c, and R417a. Since the ranking of the other three refrigerants except R417a satisfy the result available from the published data (Wijaya, 1995), the foregoing prediction is assumed reasonably accurate. Therefore, R417a should be worse heat transfer coefficient than. Also it is indicated in figures 4 and 5 that micro-fin tube could cause higher pressure drop and heat transfer performance simultaneously for each refrigerant. 5

6 IV. Experimental Test and Result Theoretical analysis often assumes an ideal process and ignores the effect of heat transfer, pressure drop, and real equipment size. As a result, theoretical results might be quite different from actual operating data. Thus, in order to well compare the performance of and R417a, both and R417a were tested under the same testing conditions. All test conditions are according to ARI test standard 2/24. Testing condition and installation The testing equipment is an Amana residential heat pump optimized for operation. The heat pump had a nominal 8.8-kW (2.5-ton) cooling capacity. The refrigerant charge was 3.9 kg (9 oz), and the lubricant is mineral oil. A thermostatic expansion valve (TVX) is employed as the throttling device. The indoor coil consists of four circuits, each with eight thermocouples at every U bend. The outdoor coil consists of six circuits with two thermocouples installed at the inlet and outlet of each circuit. Also thermocouples are placed at the refrigerant entrance and exit of other components part of the heat pump. Pressure transducers are installed to measure the pressure at compressor suction and discharge, liquid line, and evaporator inlet. A turbine flow meter is installed at the exit of the condenser. Compressor and fan power are measured continuously with a watt transducer. All data are collected through HP-VEE at 3-second intervals. During the test, the effect of indoor airflow on the performance of and R417a is studied in detail. The test is conducted in a two-room environmental chamber, where includes two rooms: one room simulating indoor environment and the other simulating outdoor environment. The whole testing included both cooling and heating modes. In the cooling-mode test, the indoor environmental dry-bulb temperature is maintained at C (8 F) and relative humidity is 52%. Outdoor ambient dry bulb temperatures are set separately at 27.78, 29.44, 32.22, 35., and C (82, 85, 9, 95 and F), and outdoor relative humidity is less than 5%. In the heating-mode test, the indoor dry bulb temperature is maintained at C (7 F), and the room relative humidity is set at 5%. The outdoor dry bulb temperatures are set separately at.56, 1.67, 2.78, 5.56, and 8.33 C (33, 35, 37, 42, and 47 F), and outdoor room relative humidity is higher than 75%. Figure 6 indicates the schematic information for the testing system and sensor distribution. Figure 6 Schematic of testing system and sensor map at cooling mode. Testing result The testing result in cooling mode is plotted in figure 7. Figure 7(a) shows the COP for R417a as a function of outdoor ambient temperature. The COP curve is illustrated as a linear function of outdoor environmental 6

7 temperature. On the other hand, the COP also increases with the augment of the indoor airflow rate. Compared to, the COP of R417a is 11 15% less than the former at the identical operating condition. This COP lose of R417a becomes a slight extension when the indoor airflow is low. For example, the COP of R417C is about % lower than that of if the indoor airflow is between 22.7 m and 17. m (8 and 6 cfm); however, it extends 15% lose at 12.7 m (45 cfm) indoor airflow. Figure 7(a) also points out that the cooling operation behavior of R417a is inferior to that of in spite of the variation of evaporating temperature. The cooling capacity of R417a is nearly 15 2% less than, as seen in figure 7(b). The lower capacity means the system need longer running time to satisfy the same house load. It is also reflected indoor air supplying temperature. The indoor air outlet temperature of R417a is higher than that of. The difference is about 1.6 to ~2.2 o C on the basis of data plotted in figure 7(c). Therefore, in order to provide the same cooling capacity, R417a has to low the evaporating temperature, which could lead to a poorer cooling COP. However, Fig. 7(d) shows that the discharge temperature of R417a is tremendously less than. This attribute makes R417a ideal to serve as a drop-in replacement in compressors designed for because the chance for compressor burnout is well reduced by using R417a as working fluid. Square: 22.7 m Diamond: 17. m Triangle: 12.7 m Solid line: Dash line: Square: 22.7 m Diamond: 17. m Triangle: 12.7 m Solid line: Dash line: (a) COP Cooling capacity Indoor Air Outlet Temperature ( o C) 2. SSquare: m 3 Diamond: Diamond: 2.27m m / m in Triangle: m 18. Triangle: Solid line: 1.7m /min Solid Dash line: line: Outdoor Environmental Temperature ( o C) Discharge Temperature ( o C) 75. SSquare: mm 7. DDiamond: 2.27m m/min Triangle: 12.7 Triangle: 1.7m 3 m / min Solid line: Solid line: 65. Dash line: Outdoor Environmental Temperature (of) (c) Indoor air supplying temperature (d) Discharge temperature Figure 7 Experimental cycle performance of and R417a as a function of outdoor air temperature and indoor airflow rate cooling mode. 7

8 COP Square: Square: m m Diamond: 17. Diamond: 2.27m 3 m Triangle: 12.7 m /min Triangle: Solid line: 1.7m Solid Dash line: Heating Capacity (KW) SSquare: m 3 /min Diamond: 2.27m 17. m 3 Triangle: 1.7m 12.7 m 3 Solid Solid line: line: Dash line: Outdoor Environmental Temperature ( o C) Outdoor Environmental Temperature (oc) (a) COP (b) Cooling capacity Indoor Air Outlet Temperature ( o C) S Square: m 3 DDiamond: 2.27m 17. m 3 TTriangle: 1.7m 12.7 m Solid Solid line: line: Dash line: Outdoor Environmental Temperature ( o C) Discharge Temperature ( o C) Square: 3.3 m Square: 22.7 m 3 Diamond: 2.27m /min Diamond: 17. m Triangle: 1.7m 12.7 m 3 Solid line: Outdoor Environmental Temperature ( o C) (c) Indoor air supplying temperature (d) Discharge temperature Figure 8 Experimental cycle performance of and R417a as a function of outdoor air temperature and indoor airflow rate heating mode. In the heating-mode test, the operating behaviors of R417a are plotted in Figures 8(a) 8(d). Figure 8(a) shows the heating COP of R417a is close to that of at a high outdoor ambient temperature, but less at a low outdoor ambient temperature. The heating capacity of R417a, however, is still 12 2% less than that of, as shown in Fig. 8(b). Compared to the cooling-mode performance, R417a shows some improvement in COP and capacity in heating-mode operation. Figure 8(c) shows supply air temperature for R417a in the heating mode is 2 to ~ 4.5 o C (3.6 to 8.1 F) less than that of. Thus, to provide the same comfort level as, R417a has to raise the condensing temperature, which also leads to low heating COP. Figure 8(d) illustrates that the discharge temperature of R417a is still significantly less than that of at the same operating conditions for heating-mode operation. Generally, the performance of R417a shown in heating mode less than, as is very similar to what happens at cooling mode. Based on the foregoing result, it indicates that although the theoretical COP of R417a is just 2-3% less than, the testing COP of R417a is actually 12% less than. The conclusion is consistent with other testing data (Spatz and Richard, 22; Aprea and his coauthors, 24a, 24b). The reason why it causes so much different from theoretical analysis might be due to the poor heat transfer performance of R417a, as indicated in figures 4 and 5. The heat transfer coefficient of R417a is around 3% less than. The heat transfer coefficient of R417a is also worse than R47c and R4a. As to R4a, the past research proved the optimized R4a systems could have a 2-7% higher system COP than that of (Sami et al., 1997; Henderson et al., 21; Hundy and Pham, 21). 8

9 R4a, however, requires system redesign to cope with its inherent high discharge pressure. Obviously, R4a can t serve as a drop-in replacement of. Usually, R47c is considered primarily as the drop-in replacement of so far. However, besides the restrict requirement of synthetic oil, R47c s COP is also 5 17% lower than (Greco et al., 1997). Therefore, the COP of R417a is close to or slightly higher that R47c. But, the unique and pronounced advantage of R417a over R47c and R4a is the fact that it could serve as working fluid without any systems modification. Therefore, R417a should be considered as a true drop-in replacement with the certain penalty of energy efficiency. V. Conclusion The paper is in an effort to investigate the drop-in application of R417a in a residential heat pump system designed for. The detailed work includes the analysis of thermodynamic cycle, heat transfer, and experimental testing. In the experimental study, R417a is used as a drop-in replacement for air conditioning and heat pump applications without modifying equipment or switching to synthetic oil. Both heating-mode and cooling-mode operation were tested in a nominal 2-ton split heat pump system using and R417a. Based on the analysis of thermodynamic cycle, R417a can t achieve the better theoretical COP than while its capacity is quite lower than. The further heat transfer analysis also indicates that the heat transfer performance of R417a is worse than, R47c, and R4a. The testing results of R417a show that the cooling and heating COPs of R417a are 11 15% and 12% lower, respectively, than those of. The capacity of R417a is only 8 88% of that of. Therefore, it takes a longer running time for the R417a system to achieve the desired room temperature. However, the discharge temperature of R417a is much lower than that of. Compared with past research, it shows that the COP of R417a could be just equivalent to or slightly higher than that of R47c. But the unique advantage of R417a is that R417a does not require any equipment modification or a switch to synthetic oil if it is served as a drop-in replacement in the equipment designed for. The testing results illustrate there is no necessary to concern about compressor burnout if R417a is filled directly into equipment designed for. Therefore, R417a is a suitable candidate as a drop-in low-cost replacement with the certain penalty of energy efficiency for existing systems. If R417a is to serve as a long-term replacement for, a new optimized system for 417A is necessary, and further research is also needed on the operating behavior of R417a in such an optimized system. Acknowledgments The authors appreciate Esher Kweller (formally U.S. DOE) and Arun Vehra of the US Department of Energy, for the support of this project. This project was sponsored by the US Department of Energy. Oak Ridge National Laboratory is managed by UT-Battelle, LLC for the U.S. Department of Energy under contract DE-AC5- O725. References 1 Roberts, N.A., Use of R417a (ISCEON 59) in Refrigeration and Air Conditioning Applications, Technical Article, EuroCooling and Heating Centro Studi Galileo, 2. 2 Wijaya, Halim, Two-Phase Flow Heat transfer and Pressure Drop Characteristics of and R32/R125, ASHRAE Transactions, Vol. 1(2), Mei, V.C., Domitrovic, R., and Chen, F.C., Experimental Study of an R47c Drop-in Test on an Off-the-Shelf Air Conditioner with a CounteRCross-Flow Evaporator, ASHRAE Transactions, 4(2), 1998, pp Greco, A., Mastrullo, R., and Palombo, A., R47c as an Alternative to in Vapour Compression plant: An Experimental Study, Int. J. of Energy Research, 21(12), 1997, pp Henderson, P.C., Mongey, B., Hewitt, N.J., and McMullan, J.T., Replacing with a Hydrocarbon or Hydrofluorocarbon? Int. J. Energy Research, Vol.25, 21, pp Hundy, G.F., and Pham, H.M., Effect of Refrigerant Choice on Efficiency in Air Conditioning, Annual Conference of the Institute of Refrigeration, Nov. 1, 21, London, UK. 7 Aprea, C., Mastrullo, R., Renno, C., and Vanoli, G.P., An Analysis of the Performances of a Vapor Compression Plant Working Both as a Water Chiller and a Heat Pump Using and R417a, Applied thermal Engineering, Vol. 24, 24, pp

10 8 Aprea, C., Mastrullo, R., and Renno, C., An Evaluation of Substitutes Performances Regulating Continuously the Compressor Refrigeration Capacity, Applied thermal Engineering, Vol. 24, 24, pp Spatz, M.W. and Richard, R.C., Performance of Oil Return Characteristics of HFC/HC Blends, Paper R13-1 in Proceedings of Ninth International Refrigeration and Air Conditioning Conference at Purdue, 22. Sami, S.M., Song, B., and Poirier, B., Energy Efficiency Analysis of A new Ternary HFC Alternative, Int. J. Refrigeration, Vol. 21, 1997, pp Choi, J.Y., Kedzierski, M.A., and Domanski, P.A., A Generalized Pressure Drop Correlation for Evaporation and Condensation of Alternative Refrigerants in Smooth and Micro-Fin Tubes, NISTIR-6333, NIST, Cavallinni, A., Del Col, D., Doretti, L., Longo, G.A., and Rossetto, L., Heat Transfer and Pressure Drop during Condensation of Refrigerants Inside Horizontal Enhanced Tubes, Int. J. Refrigeration, Vol. 23, 2, pp Cavallinni, A., Del Col, D., Doretti, L., Longo, G.A., and Rossetto, L., Refrigerant vaporization inside enhanced tubes: a heat transfer model, Heat and Technology, Vol. 17, 1999, pp Thome, J.R., Hajal, J.El, and Cavallinni, A., Condensation in Horizontal Tubes 2: New Heat Transfer Model Based on Flow Regimes, Int. J. Heat and Mass Transfer, Vol. 46, 23, pp Thome, J.R., and Hajal, J.El, Two-Phase Flow Pattern Map for Evaporation in Horizontal Tubes Latest Version, 1ST International Conference on HEAT Transfer, Fluid Mechanics, and Thermodynamics, April 8-, 22, Kruger Park, South Africa.