Experimental Heat Transfer Coe±cient and Pressure Drop during Condensation of R-134a and R-410A in Horizontal Micro- n Tubes

Transcription

1 International Journal of Air-Conditioning and Refrigeration Vol. 26, No. 3 (2018) (16 pages) World Scienti c Publishing Company DOI: /S Experimental Heat Transfer Coe±cient and Pressure Drop during Condensation of R-134a and R-410A in Horizontal Micro- n Tubes Sanjeev Singh* and Rajeev Kukreja Department of Mechanical Engineering Dr. B. R. Ambedkar National Institute of Technology Jalandhar G. T. Road By Pass, Jalandhar , Punjab, India *singhs.me.14@nitj.ac.in; snrn_1976@rediffmail.com kukrejar@nitj.ac.in Received 2 February 2018 Accepted 28 May 2018 Published 18 July 2018 Condensation heat transfer coe±cients and pressure drops of HFC refrigerants R-134a and R-410A have been investigated experimentally in smooth and micro- n tubes (helix angles 18 and 15 ) of outer diameter 9.52 mm at mass uxes from 200 to 600 kg/m 2 s, vapor qualities between 0.1 and 0.9 and at saturation temperatures of 35 C and 40 C. Results showed that the heat transfer coe±cients of R-134a and R-410A inside micro- n tubes were and times higher and frictional pressure drops were and times higher than those of smooth tubes. These experimental results are compared with the existing heat transfer and frictional pressure drop correlations proposed by di erent researchers. The comparison showed fairly good agreement with these existing correlations within 30%. A new correlation has also been proposed for predicting heat transfer coe±cient in micro- n tubes. The oil concentrations measured for refrigerants R-134a and R-410A varied in the range of %, respectively. Keywords: Heat transfer coe±cient; frictional pressure drop; condensation; micro- n tubes; HFC refrigerants. Nomenclature C p : Speci c heat capacity at constant pressure [J/kg K] G : Mass ux [kg/m 2 s] D : Tube inner diameter [m] P : Pressure drop [N/m 2 ] h : Heat transfer coe±cient [W/m 2 K] m : Mass ow rate [kg/s] p : Pressure [bar] T : Temperature [ C] x : Vapor quality [dimensionless] R x : Heat transfer area enhancement factor Re eq : Equivalent Reynolds number [dimensionless] Nu : Nusselt number [dimensionless] Pr : Prandtl number [dimensionless] Q : Heat transfer rate [W] q : Heat ux [kw/m 2 ] N : Number of ns N p : Number of data points A : Heat transfer surface area [m 2 ] * Corresponding author

2 S. Singh & R. Kukreja p cr : Critical pressure [bar] Re : Reynolds number Re l : Reynolds number of liquid phase ð Gð1 xþd l Þ Re v : Reynolds number of vapor phase ð GDx v Þ Fr : Froude number B o : Bond number e : Percent deviation e r : Average deviation e ab : Mean absolute deviation S d : Standard deviation F.H. : Fin height HFC : Hydro- uoro-carbon Greek symbols " : Void fraction [dimensionless] : Dynamic viscosity [N s/m 2 ] : Density [kg/m 3 ] : Spiral or helix angle ( ) : Fins angle ( ) Subscripts s : saturation temperature g : Gas phase l : Liquid phase exp : Experimental w : Water MF : Micro- n tube i : Inlet o : Outlet eq : Equivalent Pre : Predicted BWT : Bottom wall thickness 1. Introduction Micro- n tubes (Fig. 1) are used in modern air conditioning and refrigeration systems and heat pump applications. These 1,2 small diameter tubes have better rates of heat transfer than conventionalsized condenser and evaporator tubes with round copper tubes and aluminum or copper n that have been the standard in the HVAC industry for many years. Micro- n tubes facilitate high heat transfer compared to smooth tubes. These tubes can withstand higher pressures required by the new generation of environment-friendly refrigerants such as R-134a and R-410A. Micro- n tubes involve lower material costs because these require low refrigerant inventory, smaller ns and coil materials. Micro- n Fig. 1. Micro- n tube. tubes are used in the design of smaller and lighter high-e±ciency air conditioners and refrigerators. The performance bene ts of copper micro- n tubes include high heat exchange properties, long-term durability, resistance to corrosion, low maintenance cost and antimicrobial properties. The choice of refrigerant for a particular application is based on its favorable thermodynamic, transport properties, noncorrosive nature and safety (nontoxic and non ammable). Although many uids can act as refrigerants, but CFCs became the most popular refrigerants due to their excellent thermodynamic and transport properties. However, CFCs were replaced with hydrochloro uorocarbons (HCFCs) due to Montreal Protocol and the most common HCFC is \R-22." HCFCs are used in most of the air conditioners available in the market even today. But the HCFCs are just marginally better than CFCs as they still contain chlorine which is unsafe for the environment. To remove chlorine from the refrigerant, manufacturers produced new generation refrigerants called hydro urocarbons (HFCs). Although these refrigerants have global warming potential (GWP), but still these are better than HCFCs. The most commonly used HFCs in refrigerator and air conditioners are R-134a and R-410A. R- 410A is a highly-e±cient nearly-azeotropic mixture (50% R-32 and 50% R-125) and will not damage the ozone layer of the atmosphere. Recently, air conditioning systems like window and unitary air conditioners are being topped with R-410A refrigerant charge. The objective of the present work has been to examine the advantages of micro- n tubes over smooth tubes in terms of higher heat transfer coef- cient (HTC) and frictional pressure drop during condensation of R-134a and R-410A (HFCs) in the large range of mass ux of kg/m 2 s and propose the heat transfer coe±cient correlation for micro- n tubes

3 The results of this study will create essential addition to understand the mechanism of heat transfer coe±cient and pressure drop during condensation of R-134a and R-410A in micro- n tubes. The proposed correlation can also be used by researchers for examining and developing the condensers and evaporators in air conditioning industries and heat pump systems. Sapali and Patil 3 investigated the heat transfer coe±cient during condensation of R-134a and R- 404A in smooth tubes of 8.56-mm inner diameter (ID) and micro- n tubes of 8.96-mm ID with saturation temperature ranging from 35 Cto60 C and mass uxes between 90 and 800 kg/m 2 s. They concluded that the condensation heat transfer coe±cients of R-134a and R-404A for the micro- n tubes were and times greater than that for smooth tubes. Cavallini et al. 4 performed the condensation heat transfer of R-410A inside mm outer diameter (OD) micro- n tube at mass ux ranging from 100 to 800 kg/m 2 s and have compared with the equivalent inner diameter plain tube. They reported that the condensation heat transfer enhancement factor (EF) and pressure drop for R-410A were, respectively, higher and times higher at 400 and 800 kg/m 2 s mass uxes than those with most common HFC and HCFC refrigerants. Nozu et al. 5 evaluated condensation heat transfer and pressure drop of R-11 in horizontal micro- n tubes. They proposed that static pressure gradients in the micro- n tubes were up to 70% higher than that of smooth tube. Kim and Shin 6 reported the experimental heat transfer of R-22 and R-410A in 9.52-mm outer diameter horizontal micro- n tubes for mass ow rates of kg/h, constant heat ux of 11 kw/m 2 and at saturation temperature of 45 C. They proposed that the condensation coe±cients of R-22 and R-410A for micro- n tubes were and times greater those in smooth tubes. Wang and Honda 7 investigated the ow condensation experiments of R-11, R-123, R-134a, R-22 and R-410A inside the horizontal micro- n tubes of mm inner diameter. They concluded that root-mean-square (RMS) errors of heat transfer coe±cients in the micro- n tubes of whole refrigerants decrease in the sequence of the correlations determined by Luu and Bergles, 8 Cavallini et al., 9 Kedzierski and Goncalves, 10 Yu and koyama 11 and Wang et al. 12 Goto et al. 13 experimentally calculated the heat transfer coe±cient for the condensation of R-410A and R-22 inside Condensation in Horizontal Tubes with Micro- ns internally grooved horizontal tubes. Dobson and Chato 14 performed condensation experiments using R-12, R-22, R-134a, R-410A and R-33/125 (60%/ 40%) in horizontal smooth tubes. They suggested two correlations for heat transfer coe±cients in strati ed ow regime and in annular ow regime. After that Cavallini et al. 15 proposed a new computational method to determine the heat transfer coef- cient in micro- n tube. This method subdivided the ow regimes into T -independent, related to vapor shear force, and T -dependent, related to gravitational force. 2. Experimental Test Facility The experimental test rig used in the present work is schematically shown in Fig. 2. It consists of three loops: refrigerant loop, cooling water loop and hot water loop. In the refrigerant loop the refrigerant ows through the evaporator and superheated vapor comes out and then partly condenses in the precondenser to accomplish the desired quality of vapor, it is condensed further using the cold water owing in the annulus. Figure 3 shows the test condenser which is a counter- ow tube-in-tube condenser, with the refrigerant entering the tube side of the test condenser at a known mass ux and vapor quality. The thermocouples (T-type) are brazed in the outer copper tube wall in the middle of the tube; four thermocouples are brazed to make a cross shape in the circumference and they are situated at six positions. The refrigerant mass ow rate is measured by using Coriolis mass ow meter installed between post-condenser and evaporator. Refrigerant temperatures and pressures at the inlet and outlet of the pre-condenser, test condenser, post-condenser and evaporator are measured by using thermocouple wires and pressure sensors installed along the refrigerant ow direction in tube wall and the pressure di erence between the inlet and outlet of test section is determined by di erential pressure transmitter (DPT). The two-phase ow mixture leaves the test condenser and goes to post-condenser which is a shelland-tube-type heat exchanger wherein refrigerant ows in shell and cooling water through the copper tubes. This fully condensed liquid refrigerant ows to the evaporator through needle expansion valve used as expansion device by open-type reciprocating compressor and it is compatible with refrigerants

4 S. Singh & R. Kukreja Fig. 2. Schematic diagram of experimental test rig. Fig. 3. Schematic representation of test section

5 Condensation in Horizontal Tubes with Micro- ns Table 1. Experimental test conditions. Table 3. Estimated uncertainties. Operating parameters Range Parameters Accuracy Refrigerants R-134a, R-410A Mass ux, G (kg/m 2 s) Saturation temperatures, T s ð C) 35 and 40 Cooling water temperature, T w ð C) Vapor quality Outer diameter (mm) 9.52 Heat ux (kw/m 2 Þ Refrigerant temperature 0:1 C Cooling water temperature 0:5 C Di erential pressure 0:13 kpa Refrigerant ow rate 0:2% Water ow rate 5% Average heat transfer coe±cient (W/m 2 K) 6.18% R-134a and R-410A. The refrigerant ow can be controlled by needle-type expansion valve placed at the inlet of evaporator. The experimental test conditions used in this work are tabulated in Table 1. In the cooling water loop, temperature of the cooling water owing in the pre-condenser and the test condenser is governed by two secondary cooling water circuits. A resistance heater supplied hot water while cold water is supplied by a chiller. Thermostatic control of the water is accomplished by adjusting the electrical power supplied to the heater by dimmer stat. Cold water ow rates are measured using acrylic rotameters and temperatures by resistance temperature detector (RTD). In the hot water loop, evaporator is tube-coiland-tank-type heat exchanger, in which refrigerant owing in the tube side is vaporized and then water is heated by resistance heater submersed in the tank. Hot water temperature is measured by RTD. A data acquisition system Make Agilent, with three armature multiplexer cards, has been connected to computer to acquire the raw data. The various important dimensions of smooth and micro- n tubes and other geometrical parameters used during the experimentation are tabulated in Table 2. Table 2. Parameters Horizontal smooth and micro- n tube dimensions. Smooth tube Micro- n tube-1 (MF-1) Micro- n tube-2 (MF-2) OD (mm) Bottom wall thickness (BWT) (mm) Fins height, (FH), (mm) Fins angle (Þ; ð Þ Helix angle (Þ; ð Þ No. of ns, N Length of tube (m) Fig. 4. Geometry of micro- n tube. The thermophysical properties of refrigerants R-134a and R-410A used in this study are calculated from REFPROP version The uncertainty analysis for heat transfer coe±cients has been performed by using the method proposed by Mo at 17 (Table 3). Figure 4 shows the geometry of micro- n tubes. 3. Data Reduction 3.1. Local heat transfer coe±cient The heat transfer rates Q in the test sections were calculated by the mass ow rate and temperature di erence of cooling water at the inlet and outlet of test condenser: Q ¼ m w C pw T; ð1þ where C pw is the speci c heat of cooling water, m w is the cooling water mass ow rate and T is the temperature di erence of cooling water at the inlet and outlet of test condenser. The local condensation heat transfer coe±cient for two-phase ow in horizontal tubes is obtained from Newton's law of cooling represented in

6 S. Singh & R. Kukreja Eq. (2): h ¼ Q AðT sat T wi Þ : ð2þ Vapor quality at the inlet and outlet of test condenser has been calculated from enthalpy balance of pre-condenser and test condenser. The change of vapor quality at test condenser is from 0.07 to Then, mean vapor quality in test condenser is calculated by 3.2. Pressure drop x m ¼ x i þ x o 2 : ð3þ Total pressure drop or pressure penalty in twophase ow in horizontal tubes is expressed as the sum of three components namely frictional, gravity and momentum components: dp dz total ¼ dp dz fric For horizontal tubes, dp dz dp dz gravity ¼ 0: gravity dp dz : mom ð4þ ð5þ Total pressure drop is measured by di erential pressure transmitter connected between the inlet and exit of the test condenser. Finally, two-phase local frictional pressure drop can be obtained by subtracting the calculated momentum pressure drop from total measured pressure drop. Momentum pressure is de ned as 2 3 ð1 xþ 2 p mom ¼ G 2 l ð1 "Þ þ x2 g " out ð1 xþ2 l ð1 "Þ þ x2 5 : ð6þ g " While the void fraction " is calculated using Smith correlation 18 as follows: 2 8 vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi9 3 1 x l " ¼ 1 þ ð1 xþ >< u þ 0:4ð1 xþ g >= t g 6 0:4 þ 0:6 7 : 4 x l x þ 0:4ð1 xþ 5 >: >; in ð7þ 4. Experimental Results and Discussion 4.1. Heat transfer coe±cient Experimental tests have been conducted during condensation of R-134a and R-410A with mass ux range between 200 and 600 kg/m 2 s and at saturation temperatures of 35 C and 40 C in smooth and micro- n (MF-1 and MF-2) tubes. The experimental data of HTC are compared between smooth and micro- n (MF-1 and MF-2) tubes under the same operating conditions during condensation of R-134a and R-410A as shown in Figs. 5 and 6. These gures show that the values of HTC of R-134a and R-410A in micro- n tubes (MF-1 & MF-2) are higher than those in smooth tube and decrease when the saturation temperature increases. The values of HTC of R-134a are higher than those of R-410A for the mass ux range from 200 to 600 kg/m 2 s and saturation temperatures of 35 C and 40 C, it is times for R-134a and times for R-410A. In micro- n tubes liquid density and surface tension play an important role on condensation heat transfer coe±cient due to the internal grooves. Heat transfer coe±cient is in uenced highly by the liquid lm thickness on the n tip. For R-134a the liquid phase density is 1.16 times higher than R-410A causing the liquid at the upper portion of the tube to fall down freely and decrease of liquid lm thickness on the n tip resulting in increase of heat transfer coe±cient of R-134a than R-410A. The surface tension also in uenced the heat transfer coe±cient. Due to large value of surface tension liquid layer is uniformly distributed on the tube surface because of thinning e ect resulting in increase of heat transfer coe±cient. 19 The condensation HTC in micro- n tube (MF-1) is greater than that of micro- n tube (MF-2) because of higher helix angle and n height which increase the turbulence and heat transfer surface area Heat transfer enhancement factor Another method of comparing the performance of micro- n tube with smooth tube is heat transfer EF. It is de ned as the ratio of HTC of micro- n tube to the HTC of smooth tube at the same inner diameter of tube, mass ux and saturation temperature. Figure 7 shows the enhancement factors for micro- n tubes MF-1 and MF-2 at saturation temperatures of 35 C and 40 C during the condensation of refrigerants R-134a and R-410A

7 Condensation in Horizontal Tubes with Micro- ns The EF for R-134a varies from minimum (1.2) at high mass ux to maximum (1.9) at low mass ux, due to increase of mass ux liquid phase of refrigerant ows over the ns and this process reduces the n e ect. Therefore the heat transfer surface area enhancement is predominant in high mass ux region. Though the EF for R-410A is minimum (1.3) at low mass ux and maximum (1.65) at high mass ux, due to low values of surface tension of R-410A the liquid phase of refrigerant ows over the ns in low mass ux region also. Therefore turbulence is predominant for HTC enhancement in low mass ux region. Therefore enhancement factor increases with mass ux E ect of pressure ratio on heat transfer coe±cient Figure 8 shows the e ects of pressure ratios (ratio of saturation pressure to critical pressure) of refrigerants R-134a and R-410A in smooth and micro- n (MF-1 and MF-2) tubes of same inner (c) Fig. 5. Comparison of heat transfer coe±cients of R-134a and R-410A in smooth and micro- n (MF-1) tubes at saturation temperatures of 35 C and 40 C. (d) diameter and working under similar conditions. It shows clearly the bene t of micro- n tube over the smooth tube because the heat transfer coe±cient in micro- n tube is higher than in smooth tube. This is due to the lower thickness of liquid condensate lm in micro- n tubes. Also Fig. 8 showsthatthe heat transfer coe±cient of R-134a is higher than that of R-410A, this is due to higher liquid density of R-134a than R-410A Proposed correlation for micro- n tubes The correlation is developed by using multiple linear regression technique 21 with 280 data points. The correlation is shown in Eq. (8), its structure is clearly developed by using geometrical characteristics of micro- n tubes and also the thermodynamic and thermophysical properties of HFC refrigerants R-134a and R-410A. In the newly developed correlation equivalent Reynolds number signi es the relative predominance of the inertia forces to the

8 S. Singh & R. Kukreja viscous forces occurring during the ow of refrigerants while Prandtl number provides the relative effectiveness of the momentum and energy transport by di usion. The involvement of pressure ratio or (c) Fig. 6. Comparison of heat transfer coe±cients of R-134a and R-410A in smooth and micro- n (MF-2) tubes at saturation temperatures of 35 C and 40 C. (d) reduced pressure makes the correlation convenient for large range of saturation temperatures. For the micro- n tubes, heat transfer area enhancement factor (R x )isacommandingparameterthanthe Fig. 7. Heat transfer enhancement factors for R-134a and R-410A in micro- n tubes at saturation temperatures of 35 C and 40 C

9 Condensation in Horizontal Tubes with Micro- ns dimensionless numbers. The product of Froude and Bond numbers indicates the e ect of related signi cance of inertia and surface tension forces. Figure 9 shows the comparison between the newly developed correlation data and experimentally calculated HTC data. This gure expressed that the newly developed correlation data correlate nicely with experimentally calculated HTC data of R-134a and R-410A for two di erent micro- n tubes (MF-1 and MF-2) within 20% range with coe±cient of correlation of 0.968, mean absolute deviation of 8.14% and standard deviation of 9.6%: Nu ¼ 0:0629 Re 0:58346 eq Pr 1:41522 p 0:0615 l p cr R 0:5521 x ðfr B o Þ 0:01314 ; ð8þ (c) Fig. 8. E ects of pressure ratios on heat transfer coe±cient during condensation of R-134a and R-410A in smooth and micro- n (MF-1 and MF-2) tubes. Fig. 9. Comparison of experimental heat transfer coe±cient with newly developed correlation in micro- n tube

10 S. Singh & R. Kukreja where R x ¼ 1 cos eN 1 sin 2 D cos þ 15: Heat transfer area enhancement factor or n geometry parameter is a function of the ratio of inside surface area of micro- n tube to the inside area of the smooth tube having the same n tip diameter. The equivalent Reynolds number is given by Re eq ¼ Re l þ 0:5 g l Re v : 4.5. Pressure drop l The total pressure drops during condensation of refrigerants between the inlet and outlet of test section were measured by using di erential pressure transmitter. The frictional pressure gradients as shown in Figs. 10 and 11 are calculated by subtracting the momentum pressure gradient from the g total measured pressure gradient. Figures 10 and 11 show the frictional pressure gradient comparisons between the micro- n tubes (MF-1 and MF-2) and smooth tubes of equivalent inner diameter during condensation of R-134a and R-410A at saturation temperatures of 35 C and 40 C. The trend is an increase of frictional pressure drop with the increase of vapor quality and mass ux and decrease of saturation temperature. Because at higher vapor quality frictional pressure drop is higher due to the large friction developed by the high velocity of vapor due to high shear force, although at low vapor quality ow is in uenced by gravity force. This indicates that the e ect of ns is very low or negligible on the friction at low vapor quality values. The frictional pressure drops in two micro- n tubes (MF-1 and MF-2) are higher than that in smooth tube caused by ns increasing the friction. Also the frictional pressure drop during the condensation of R-134a is higher than that of R-410A; this is caused by the higher density of liquid phase of R-134a than R-410A (1.16 times of R-410A). (c) (d) Fig. 10. Comparison of pressure drops of R-134a and R-410A in smooth and micro- n (MF-1) tubes at saturation temperatures of 35 C and 40 C

11 Condensation in Horizontal Tubes with Micro- ns 4.6. Pressure drop penalty factor The performance of micro- n tubes has been analyzed by penalty factor (PF). The penalty factor is de ned as the ratio of frictional pressure drop in micro- n tubes to that of smooth tubes of similar inner diameter and operating conditions. (c) Fig. 11. Comparison of pressure drops of R-134a and R-410A in smooth and micro- n (MF-2) tubes at saturation temperatures of 35 C and 40 C. (d) Figure 12 shows the e ects of mass uxes of refrigerants R-134a and R-410A on penalty factor at saturation temperatures of 35 C and 40 C in micro- n tubes (MF-1 and MF-2), respectively. In Fig. 12 penalty factor is the local value at mass ux and saturation temperature. The points shown in this gure vary with di erent geometries of tubes and Fig. 12. Penalty factor versus mass ux in micro- n tubes (MF-1 and MF-2)

12 S. Singh & R. Kukreja mass ux and each tube shows similar trends with variation of mass ux. The penalty factor increased with mass ux in the range of kg/m 2 sand decreased after reaching the optimum value at 300 kg/m 2 s. This is due to the turbulence increasedbymicro- ntubes.inthelowermass ux range of kg/m 2 s micro- n increases the frictionallossbutinhighermass uxrangeof kg/m 2 s turbulence is very high, so that the contribution of turbulence generated by micro- n tubes decreases with mass ux. 27 The experimentally measured penalty factor for R-410A is in the range of and for R-134a the range is , respectively. 5. Correlations Comparison 5.1. Heat transfer coe±cient The comparisons between experimental Nusselt number obtained from the present work with predicted Nusselt number obtained from the correlations proposed by Yu and Koyama, 11 Kedzierski and Goncalves, 10 Cavallini et al. 22 and Sapali and Patil 3 for R-134a and R-410A are shown in Figs (d). From the gures, the experimental Nusselt values show good agreement with the predicted values with deviations of about 30% but most of the data points are gathered in the range of 20% excluding some points of R-134a and R-410A. (c) (d) Fig. 13. (d) Predicted versus experimental Nusselt numbers for R-134a and R-410A in micro- n tubes (MF-1 and MF-2): correlations by Cavallini et al. 22 Sapali and Patil, 3 Kedzierski and Goncalves 10 and Yu and Koyama

13 Condensation in Horizontal Tubes with Micro- ns Table 4. Average, mean and standard deviations obtained in the comparisions between predicted and experimental Nusselt number data. Yu and Koyama 11 Cavallini et al. 22 Kedzierski and Goncalves 10 Sapali and Patil 3 e r (%) e ab (%) S d (%) Figures 13 and 13 indicate that the experimental calculated data are distributed between 30% and þ30% of predicted data excluding some data of R-134a and R-410A in micro- n (MF-1) tube in Fig. 13. Similarly Figs. 13(c) and 13(d) show that the calculated data in MF-1 and MF-2 are distributed ranging from 50% to þ 20% and 50% to þ 50%, respectively. The positive and negative values of average deviation indicate underprediction and overprediction. From the comparisons between measured data and predicted data, it has been seen that the Sapali and Patil's correlation 3 shows good agreement with experimental data among all the four correlations. The correlation of Sapali and Patil 3 is developed by using multiple linear regression of individual data and take into account the in uence of saturation temperature, subcooling, superheating and thermophysical properties of refrigerants. Table 4 compiles the results of comparison between measured experimental HTC data and that existing HTC correlation data were computed in terms of average deviation, mean absolute deviation and standard deviation, de ned as follows: Percent deviation: e ¼ h Pre h exp h exp Average deviation: e r ¼ 1 N p X ðeþ; 100; Mean absolute deviation: e ab ¼ 1 N p X jej; ð9þ ð10þ ð11þ " P # ðe er Þ 2 1=2 Standard deviation:¼ S d : ð12þ N p Pressure drop The experimental pressure gradient data for the condensation of R-134a and R-410A in micro- n tubes (MF-1 and MF-2) have been compared against the existing correlations by Friedel, 23 Choi et al., 24 Kedzierski and Goncalves, 10 Gronnerud 25 and Haraguchi et al. 26 as shown in Figs (e). From Figs (e) and Table 5, correlations by Kedzierski and Goncalves 10 and Haraguchi et al. 26 overpredict and correlations by Friedel 23 Fig. 14. (e) Predicted versus experimental pressure gradients for R-134a and R-410A in micro- n tubes (MF-1 and MF-2): correlations by Friedel, 23 Choi et al., 24 Kedzierski and Goncalves, 10 Gronnerud 25 and Haraguchi et al

14 S. Singh & R. Kukreja Choi et al. 24 and Gronnerud 25 underpredict the experimental data. The correlation by Kedzierski and Goncalves 10 establishes good association with experimental data within 30% deviation, having (c) Fig. 14. Table 5. Average, mean and standard deviations obtained from the comparisions between predicted and experimental pressure drop data. Friedel 23 Gronnerud 25 et al. 26 Haraguchi Kedzierski and Goncalves 10 Choi et al. 24 e r (%) e ab (%) S d (%) (e) (Continued) absolute deviation of 5.54% and almost 91% of data were predicted between 20% deviations. 6. Conclusion (d) The following conclusions can be drawn from the analysis of the results presented in the previous section: (1) The condensation heat transfer coe±cients and pressure drops of R-134a and R-410A have been measured inside smooth tube and two micro- n tubes (MF-1 and MF-2) in the mass ux range of kg/m 2 s, vapor quality of and at saturation temperatures of 35 C and 40 C, respectively

15 (2) The condensation heat transfer coe±cients and pressure drops inside micro- n tubes (MF-1 and MF-2) have been compared with those of similar inner diameter smooth tube under similar working conditions to show the advantages of micro- n tubes. (3) The condensation heat transfer coe±cients of R- 134a and R-410A inside micro- n tubes (MF-1 and MF-2) were and times higher and similarly pressure drops were and times higher than those of smooth tubes. (4) When R-134a was compared with R-410A at similar working conditions, the condensation heat transfer coe±cients and pressure drops of R-410A were lesser than R-134a. (5) A new correlation has been developed for the measurement of heat transfer coe±cient (280 data points) in micro- n tubes: Nu ¼ 0:0629 Re 0:58346 eq Pr 1:41522 p 0:0615 l R 0:5521 x ðfr B o Þ 0:01314 : (6) The experimental data of heat transfer coe±cients and pressure drops were compared with existing correlations inside two di erent micro- n tubes. The correlation by Sapali and Patil 3 showed the best agreement with heat transfer coe±cient data (average deviation of 2.22%) and the correlation by Kedzierski and Goncalves 10 showed good association with pressure drop data (average deviation of 5.54%). Acknowledgments This research was supported by Dr. B. R. Ambedkar National Institute of Technology Jalandhar, Jalandhar, India, and Spirotech Heat Exchangers Company, Bhiwadi, Rajasthan (India). References p cr 1. N. Cotton, COOL TECHNOLOGY: Small copper tubes make a big impact on air-conditioner e±ciency, MachineDesign (2012), design.com/news/. 2. R. Weed and J. Hipchen, Bene ts of reduced diameter copper tubes in evaporators and condensers, Proc. ASHRAE Annu. Conf. (2011), pp S. N. Sapali and P. A. Patil, Heat transfer during condensation of HFC-134a and R-404A inside of a Condensation in Horizontal Tubes with Micro- ns horizontal smooth and micro- n tubes, Exp. Therm. Fluid Sci. 34 (2010) A. Cavallini, D. Del Col, S. Mancin, L. Rossetto and C. Zilio, Visualization of the heat transfer enhancement during condensation in a micro- n tube, Proc. ASME nd Joint U. S.-European Fluids Engineering Summer Meeting Collocated with 14th Int. Conf. Nuclear Engineering (2006). 5. S. Nozu, H. Katayama, H. Nakata and H. Honda, Condensation of a refrigerant CFC11 in horizontal micro n tubes (proposal of a correlation equation for frictional pressure gradient), Exp. Therm. Fluid Sci. 18 (1998) M. H. Kim and J. S. Shin, Condensation heat transfer of R-22 and R-410A in horizontal smooth and micro- n tubes, Int. J. Refrig. 28 (2005) H. S. Wang and H. Honda, Condensation of refrigerants in horizontal micro n tubes: Comparison of prediction methods for heat transfer, Int. J. Refrig. 26 (2003) M. Luu and A. E. Bergles, Enhancement of horizontal in-tube condensation of refrigerant-113, ASHRAE Trans. 86 (1980) A. Cavallini, L. Doretti, N. Klammsteiner, L. G. Longo and L. Rossetto, Condensation of new refrigerants inside smooth and enhanced tubes, Proc. 19th Int. Congr. Refrigeration, Vol. 4 (1995), pp M. A. Kedzierski and J. M. Goncalves, Horizontal convective condensation of alternative refrigerants with in a micro- n tube, Enhanc. Heat Transf. 6 (1999) J. Yu and S. Koyama, Condensation heat transfer of pure refrigerants in micro- n tubes, Proc. Int. Refrigeration and Air Conditioning Conf. (1998), pp H. S. Wang, H. Honda and S. Nozu, Modi ed theoretical models of lm condensation in horizontal micro n tubes, Int. J. Heat Mass Transf. 45 (2002) M. Goto, N. Inoue and N. Ishiwatari, Condensation and evaporation heat transfer of R410A inside internally grooved horizontal tubes, Int. J. Refrig. 24 (2001) M. K. Dobson and J. C. Chato, Condensation in smooth horizontal tubes, J. Heat Transf. 120 (1998) A. Cavallini, D. Del Col, S. Mancin and L. Rossetto, Condensation of pure and near-azeotropic refrigerants in micro n tubes: A new computational procedure, Int. J. Refrig. 32 (2009) NIST, NIST Standard Reference Database 23: NIST Thermodynamic properties of Refrigerants and Refrigerant Mixtures (REFPROP), version 7.0,

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