Heat Transfer Characteristics for Condensation of R134a in a Vertical Smooth Tube

Experimental Heat Transfer A Journal of Thermal Energy Generation, Transport, Storage, and Conversion ISSN: 0891-6152 (Print) 1521-0480 (Online) Journal homepage: http://www.tandfonline.com/loi/ueht20 Heat Transfer Characteristics for Condensation of R134a in a Vertical Smooth Tube G. Arslan & N. Eskin To cite this article: G. Arslan & N. Eskin (2015) Heat Transfer Characteristics for Condensation of R134a in a Vertical Smooth Tube, Experimental Heat Transfer, 28:5, 430-445, DOI: 10.1080/08916152.2014.926430 To link to this article: http://dx.doi.org/10.1080/08916152.2014.926430 Accepted author version posted online: 11 Jun 2014. Submit your article to this journal Article views: 108 View related articles View Crossmark data Full Terms & Conditions of access and use can be found at http://www.tandfonline.com/action/journalinformation?journalcode=ueht20 Download by: [Mersin Universitesi] Date: 12 January 2016, At: 01:25

Experimental Heat Transfer, 28:430 445, 2015 Copyright q Taylor & Francis Group, LLC ISSN: 0891-6152 print/1521-0480 online DOI: 10.1080/08916152.2014.926430 HEAT TRANSFER CHARACTERISTICS FOR CONDENSATION OF R134A IN A VERTICAL SMOOTH TUBE G. Arslan 1 and N. Eskin 2 1 Mechanical Engineering Department, Engineering Faculty, Mersin University, Çiftlikköy Campus, Mersin, Turkey 2 Mechanical Engineering Faculty, Istanbul Technical University, Istanbul, Turkey In this study, condensation of pure refrigerant R134a vapor inside a smooth vertical tube was experimentally investigated. The test section was made of a copper tube with inside diameter of 7.52 mm and length of 1 m. Experimental tests were conducted for mass fluxes in the range of 20 175 kg/m 2 s with saturation pressure ranging between 5.8 and 7 bar. The effects of mass flux, saturation pressure, and temperature difference between the refrigerant and tube inner wall (DT) on the heat transfer performance were analyzed through experimental data. Obtained results showed that average condensation heat transfer coefficient decreases with increasing saturation pressure or temperature difference (DT). In addition, for the same temperature difference (DT), heat can be removed from the refrigerant at a higher rate at relatively low pressure values. Under the same operating conditions, it was shown that average condensation heat transfer coefficient increases as mass flux increases. Finally, the most widely used heat transfer coefficient correlations for condensation inside smooth tubes were analyzed through the experimental data. The best fit was obtained with Akers et al. s (1959) correlation with an absolute mean deviation of 22.6%. Keywords INTRODUCTION R134a, condensation, heat transfer coefficient, correlation Condensation inside tubes is an important multiphase phenomenon, which is generally observed in refrigeration and air-conditioning systems. By attributing to the International Institute of Refrigeration s report [1], Parise and Marques [2] stated that approximately 15% of the world s electricity is consumed in refrigeration and airconditioning applications. For that reason, the accurate design of these systems is important for better performance with less energy usage. Moreover, determination of operating limits for safety conditions is required. To meet these requirements, detailed theoretical and experimental knowledge about condensation is necessary. During condensation inside tubes, different flow patterns occur one after the other, and the condensation heat transfer rate changes depending on those flow patterns. Mathematical Received 9 October 2013; accepted 14 May 2014. Address correspondence to Dr. Gökhan Arslan, Mersin University, Mechanical Engineering Department, 33343, Mersin, Turkey. Tel.: þ 903243610001, #7448; Fax: þ 903243610258. E-mail: garslan@mersin.edu.tr. 430

CONDENSATION OF R134A IN A VERTICAL SMOOTH TUBE 431 NOMENCLATURE A heat transfer area, pd i L (m 2 ) c p specific heat capacity (Jkg 21 K 21 ) D i tube inner side diameter (m) D o tube outside diameter (m) G mass flux (kg/m 2 s) h heat transfer coefficient (W/m 2 K) h fg latent heat (jkg 21 ) k thermal conductivity (W/mK) L length of the test section (m) _m mass flow rate (kg/s) P pressure (Pa) _Q heat removal rate (W) T temperature (8C) x vapor quality Greek Letters Dx vapor quality change DT temperature difference (T sat 2 T w,i ) 1 mean deviation Subscripts abs absolute avg average i inlet L liquid o outlet r refrigerant, R134a s sensible sat saturated st standard T total wall tube wall wt water models developed to define condensation heat transfer are restricted with certain flow patterns. For that reason, researchers have focused on experimental studies. Condensation heat transfer performance has been investigated under different operating conditions. Based on the experimental data, new heat transfer coefficient correlations have been developed or the consistency of existing correlations investigated. In the literature, there have been extensive experimental studies dealing with condensation of refrigerants inside smooth horizontal and vertical tubes. Most of these studies have involved condensation inside smooth tubes. A detailed experimental study about condensation heat transfer and pressure drop of refrigerants was reported by Cavallini et al. [3]. On the basis of their experimental data, they hypothesized that heat transfer coefficient increased with increasing vapor quality or mass flux. At low mass flux (G ¼ 100 kg/m 2 s), the effect of the temperature difference between the refrigerant and the tube wall was very important for the heat transfer coefficient. As temperature difference decreases, heat transfer coefficient increases. Finally, it was stated that at low saturation pressures, the obtained heat transfer coefficient was higher than that obtained at high saturation pressure. The results of Jung et al. [4] were in parallel with the above-mentioned study. Obtained experimental data were compared with heat transfer coefficient correlations. It was stated that Dobson and Chato s [5] correlation under-predicted the experimental data in the low quality and mass flux regime, while it over-predicted in the high quality and mass flux regime. For that reason, Dobson and Chato s [5] correlation was modified. Park et al. [6] obtained experimental data for heat transfer coefficients and pressure drop while condensing R22, propylene, propane, dimethyl ethyl (DME), and isobutene inside a horizontal plain tube. Flow condensation heat transfer coefficients increased as the quality and mass flux increased for all refrigerants tested. The correlation of Jung et al. [4] showed the best prediction capability with a mean deviation of 11.5%. Agra and Teke [7] experimentally investigated condensation of R600a refrigerant inside a horizontal smooth copper tube; their results showed that the condensation heat transfer coefficient drops with a reduction in vapor quality.

432 G. ARSLAN AND N. ESKIN More recently, there has been considerable interest in studies of condensation of refrigerants inside vertical tubes. Most of the experiments were about water vapor condensation. Al-Shammari et al. [8] experimentally investigated water vapor condensation inside a vertical tube. Heat flux distribution along the test tube was determined. Experiment results showed that heat flux decreases at the bottom section of the tube. Increasing liquid film thickness along the test tube due to condensation led to an increase in thermal resistance of the film. Oh and Revankar [9] set up an experimental rig to investigate complete condensation of water vapor inside a vertical tube; experimental results showed that condensation heat transfer coefficient decreases depending on increasing system pressure. In addition, increasing temperature difference between saturated vapor and tube wall led to a reduction in condensation heat transfer coefficient. The obtained results showed a 15% deviation when compared with classical Nusselt [10] analysis. Kim and No [11] experimentally studied high-pressure steam condensation heat transfer in a large-diameter condenser tube. A new turbulent annular film condensation model was developed. The local heat fluxes and heat transfer coefficients were analyzed as a function of system pressure. It was stated that heat fluxes increase with an increase of system pressure. However, the heat transfer coefficients are within a range of 4,000 7,000 W/m 2 K, regardless of the system pressure. Condensation of R134a inside a vertical smooth tube was studied experimentally by Dalkilic et al. [12], and it was found that condensation heat transfer coefficient decreases along the test tube. Similar to previously mentioned researchers results, the condensation heat transfer coefficient decreased depending on the increasing saturation pressure. Meyer et al. [13] experimentally investigated heat transfer during condensation of R134a in an inclined smooth tube, where inclination angles ranging from 2908 (vertical downward) to þ 908 (vertical upward). Obtained results showed that the condensation heat transfer coefficient is strictly dependent on the saturation temperature and inclination angle. Experiments conducted at different mass fluxes (200, 300, and 400 kg/m 2 s) and saturation temperatures (308C 508C) showed that increasing saturation temperature led to a decrease in condensation heat transfer coefficient. Similarly, it was stated that inclination angle was also effective on condensation heat transfer coefficient. When inclination angle was changed, observed flow patterns also changed. That changing led to an increase or decrease in film thickness, and that affects the thermal resistance in liquid film thickness. Researchers have analyzed experimental data obtained at different inclination angles and different mass fluxes. At low mass flux, the highest heat transfer coefficient occurred at inclination angle values between 2308 and 2158. It was stated that stratifying effects of gravity was the main reason of the increase in condensation heat transfer coefficient. Cavallini et al. [14] reviewed the results in the open literature about in-tube and bundle condensation. For condensation inside plain round tubes, they emphasized that flow regimes strongly influence the heat and momentum transfer process. Flow regime maps and transition criteria between flow patterns were discussed in detail. It was stated that very few models published in literature were general and covered the entire map of two-phase regimes. Researchers have discussed the condensation heat transfer model developed by Cavallini et al. [15], which covers all flow regimes in detail. Heat transfer coefficients calculated with that model for condensation of R134a, R22, R410A, and R32 were compared against the correlation predictions of Shah [16], Haraguchi et al. [17], and Dobson and Chato [5]. It was noted that correlation developed by Haraguchi et al. [17] could not be applied at all operating conditions. Dobson and Chato s [5] correlation was

CONDENSATION OF R134A IN A VERTICAL SMOOTH TUBE 433 unsatisfactory for high-pressure fluids. Shah s [16] correlation over-predicted data relative to high-pressure fluids. Valladares [18] summarized the condensation heat transfer coefficient correlations widely used in literature. Experimental data of several researchers published in literature were compared with different correlations, and under different operating conditions, Dobson and Chato s [5] and Cavallini et al. s [3] correlations have given the best results. In this study, condensation of R134a vapor inside a smooth vertical tube is investigated experimentally. A literature search shows that most of the experiments about condensation of refrigerants inside tubes are valid for horizontal tube flow at high mass fluxes, and condensation of refrigerants inside vertical tubes has not been studied extensively. For condensation inside vertical tubes, experimental studies focused on water vapor flow. Based on these facts, the main purpose of this study is to clarify the parameters that affect the refrigerant R134a condensation heat transfer inside a vertical tube. These parameters are determined as mass flux (20 175 kg/m 2 s), saturation pressure (5.8 7 bar), and temperature difference between saturated vapor and tube inner wall (1 7 K). The range of the vapor quality change in the test section is between 0.1 and 0.4. Moreover, experimental data are compared with condensation heat transfer coefficient correlations. EXPERIMENTAL STUDY A test facility was designed and constructed in which condensation heat transfer characteristics of refrigerant R134a vapor flowing downward inside a smooth tube could be determined. A schematic drawing of the system is shown in Figure 1. The test facility consisted of two leak-proof and closed loops. One was for R134a refrigerant, and the other was for heating and cooling water. The equipment in the refrigerant loop is also given in Figure 1. During the experiments, the storage tank (point 1) was cooled to control the system pressure and liquid level in the tank. Liquid refrigerant was circulated by using a variable-speed pump (point 2). A flow meter (point 3) was mounted at the exit of the pump to measure the mass flow rate. After that, liquid refrigerant was directed to the evaporator (point 4); this evaporator was a plate-type heat exchanger. Hot water was used to evaporate the liquid refrigerant. Refrigerant vapor was directed to the test section (point 5), and condensation process occurred here. A liquid vapor separator (point 7) was mounted at the exit of the test section. A scaled glass tube (point 8) was mounted at the liquid-side exit of the separator. During the experiments, valves at the inlet end exit of the scaled tube were used to collect the separated liquid in a scaled tube to determine the vapor quality of the refrigerant at the exit of the test tube. First, volumetric flow rate of the liquid collected in the scaled tube was measured by using a chronometer. Then the mass flow rate of the liquid was determined by using the density of the refrigerant, which was calculated according to the measured pressure and temperature data. In that way, vapor quality at the exit of the test section was determined. Separated vapor was cooled in a condenser (point 9), and the loop was completed at the storage tank. The other loop in the experimental facility was used for supplying hot water to the evaporator and cold water to the test section and condenser. The equipment in that cycle consisted of heating/cooling baths, a turbine-type mass flow meter, and temperature probes. The temperature of water in heating/cooling baths could be controlled sensitively; in that way, the temperature difference of the saturated vapor and tube wall could be controlled effectively. All sensors used for measuring temperature, pressure, and mass

434 G. ARSLAN AND N. ESKIN Figure 1. Schematic of experimental set-up. flow rate were connected to a data logger. Before collecting experimental data, the test facility was allowed to run for 2 3 h to allow time for conditions to become constant. Approximately 15 min were required to take a complete set of readings. Test Section The test section was built as a double-pipe heat exchanger with a counter-flow arrangement. It was mounted vertically with refrigerant flowing downward. The outside diameter and wall thickness of the inner cupper tube were 9.52 and 1 mm, respectively. The inner diameter of the outside cupper tube was 16 mm. The gap between the tubes where cooling water circulated was approximately 3.2 mm. A detailed schematic representation of the test section is given in Figure 2.

CONDENSATION OF R134A IN A VERTICAL SMOOTH TUBE 435 In the test section, refrigerant and cooling water temperatures (T) at the inlet and exit, tube wall temperatures, refrigerant pressure (P) at the inlet and exit, cooling water mass flow rate (M), and differential pressure of the refrigerant are measured. Sight glasses were mounted at the inlet and exit of the test section for visual observation. As already mentioned, test section was 1 m long, and the surface temperatures were measured at six different points. At each point, two thermocouples were soldered into the sockets placed outside of the inner tube. The depth of the socket was 0.3 mm, so soldered thermocouples were put at the same level as possible as the tube outside surface; this was an attempt to minimize the effect of water flow on wall temperature measurement. For temperature calibrations, a reference resistance temperature detector (RTD) with an accuracy of ^0.038C was used. Calibration data for temperature probes were collected between 08C and 328C at 12 different points. Pressure calibration data were collected between 0 and 15 bars at 24 different points. The standard error propagation method was used to determine the uncertainty of the calculated quantities. The uncertainties of the apparatus, instrument, and calculated quantities are given in Table 1. Data Reduction Figure 2. Test section. Experiments were conducted at the saturation pressure about 5.8 7 bar, mass flux about 20 175 kg/m 2 s, temperature difference between saturated vapor and wall (DT) about 1 7 K in the test section. At the inlet and exit of the test tube, K-type thermocouples were mounted to measure the temperature change of the cooling water; a turbine-type mass flow meter was used to measure mass flow rate. The total heat removal rate determined by measuring the mass flow rate and temperature increase of cooling water flowing inside the annulus of the test section is given by Eq. (1): _Q T ¼ _m wt c p;wt T wt;o 2 T wt;i : ð1þ During the experiments, an evaporator was used to obtain saturated vapor at the inlet of the test tube. Due to the fluctuations in system pressure and the uncertainty of the measuring

436 G. ARSLAN AND N. ESKIN Table 1. Experimental uncertainties Measured parameters ^ RTD 0.058C Thermocouple 0.058C Pressure 0.5% Coriolis flowmeter 0.25% Turbine type flowmeter 1% _Q (W) 10 20% Dx (Eq. (3)) 8 20% h (W/m 2 K) 7 18% devices, it was difficult to obtain and maintain the saturated vapor flow conditions at the inlet of the test section. For that reason, experiments in which refrigerant was superheated with a temperature difference of 1 28C above saturation temperature at the inlet of the test section were selected in this study. Webb [19] analyzed the effect of superheat on condensation heat transfer; that study emphasized that at a high degree of superheat, forced convection heat transfer was also important, as condensation heat transfer and the composite heat transfer coefficient must be defined in this situation. Li et al. [20] used a preheater to get an inlet condition of x ¼ 1 in their experimental study. They controlled the vapor superheating in the range of 2 38C and assumed that sensitive heat was negligible in heat transfer. According to those facts, the ratio of sensible heat in total heat transfer was investigated. The basic condition for the formation of condensate is that tube inner wall temperature must be below the vapor saturation temperature. Since the surface temperature was lower than the saturation temperature for all experiments, condensate formation started at the inlet of the test section. Some portion of the total heat removal rate was used to convert superheated vapor to saturated vapor, which was defined as sensible heat, and the remaining portion was defined as latent heat. Sensible heat was calculated using Eq. (2): _Q s ¼ _m r c p;r T r;i 2 T r;sat : ð2þ Mass flow rate of the refrigerant was measured by using a coriolis-type flow meter. At the inlet and exit of the test tube, temperatures of the refrigerant were measured by using RTDs. At the same points, gauge pressure transmitters were mounted to measure the saturation pressure. For all experiments, the sensible heat removal rate calculated by using Eq. (2) was lower than 5% of the total heat removal rate on average, and it was assumed that the sensible heat removal rate was negligible on heat transfer. According to that, the change in vapor quality of the refrigerant in test section was calculated by using Eq. (3): Dx ¼ _ Q T _m r h fg : ð3þ The experimental set-up was designed and constructed to measure the vapor quality at the exit of the test section. In that way, vapor quality change obtained from the heat balance equation could be checked with experimental data. Due to the nature of experimental study, temperature and pressure at the inlet and exit of the test section were measured with a certain uncertainty. Saturation temperature was determined according to the measured pressures at

CONDENSATION OF R134A IN A VERTICAL SMOOTH TUBE 437 the inlet and exit of the test section and compared with the measured temperature at those points. Since pure vapor without droplets was observed at sight glass section, and uncertainties of the temperature and pressure were known, the state of the vapor at the inlet was determined. Agreement of the measured temperature and pressure was checked by using the measured vapor quality at the exit of the test section. For all experiments, an inlet condition of x ¼ 1 was maintained, and average vapor quality was calculated by using Eq. (4): x avg ¼ 1 2 Dx 2 : ð4þ As previously mentioned, the test section was 1 m in length and outside wall temperature was measured at six different points by thermocouples. By using these experimental data, inner wall temperature of the test section was calculated by using Eq. (5): _Q T ln D o D i T wall;i ¼ T wall;o þ : ð5þ 2pLk tube Finally, the average condensation heat transfer coefficient was calculated by using the average values of thermocouples mounted on the tube surface and the saturation temperature of the refrigerant as given in Eq. (6): Q T =A h avg ¼ : ð6þ T sat 2 T wall;i;avg To validate the accuracy of the measurements in the test section, refrigerant vapor flowing at low mass flux (G ¼ 12.5 kg/m 2 s) was completely condensed (Dx ¼ 1) in the test section at a saturation pressure of 5.9 bar. An energy balance between the heat loss of the refrigerant and the heat gain of the water in the test section was checked, and the average deviation between the two was less than 2.0%. As stated before, vapor quality of the refrigerant at the exit of the test section was measured by using a liquid vapor separator and scaled tube. Complete condensation at this situation was also validated by measuring the exit vapor quality. Based on this application, inspection of the heat balances for all experiments showed very fair agreement between the rate of heat transfer to the cooling water and heat transfer from the refrigerant. The average deviation was less than 6%, which is satisfactory for this type of study. For all experiments, the uncertainty of condensation heat transfer coefficient was calculated as ^7 18%. As the heat removal rate increased, uncertainties of calculated quantities decreased. Finally, thermo-physical properties of the refrigerant and water were evaluated by using NIST Standard Reference Database 23. RESULTS AND DISCUSSION Condensation heat transfer characteristics of R134a vapor flowing inside a vertical smooth tube was investigated experimentally. Thirty-nine experimental runs were made with refrigerant R134a for saturation pressures between 5.8 and 7 bar and mass fluxes between 20 and 175 kg/m 2 s. According to the measured variables, such as temperature,

438 G. ARSLAN AND N. ESKIN Figure 3. Comparison of measured and calculated vapor quality change in test section. pressure, and mass flow rate, parameters that affect the condensation heat transfer coefficient were determined as saturation pressure, mass flux, temperature difference between saturated vapor, and tube inner wall (DT) of the test section. For all experiments, inlet vapor quality was unity, and the range of the vapor quality change in the test section was between 0.1 and 0.4. According to Eq. (4), the range of the average vapor quality was between 0.95 and 0.8. Vapor quality change in the test section was measured at the exit of the test section. In Figure 3, calculated and measured vapor quality change data are compared. It is found that all calculated vapor quality change data are in ^25% deviation band when compared with the experimental data. The range of uncertainty of the calculated vapor quality change is between 8 and 20% according to the standard error propagation method. By using that method, vapor quality change obtained from the heat balance equation could be checked with experimental data. Effect of Temperature Difference (DT), Mass Flux, and Saturation Pressure Figure 4 represents the heat transfer data obtained from this experimental study. These data points clearly illustrate the general effects of temperature difference (DT) on the condensation heat transfer coefficient. As the temperature difference increased, the condensation heat transfer coefficient showed a tendency to decrease. The slope of the curve at a high temperature difference became flat. That means condensation heat transfer is independent from temperature difference after a certain limit. It can be seen from this figure that heat transfer data are collected between certain temperature difference values. At similar temperature difference values, higher condensation heat transfer coefficients were obtained. This situation states that parameters different from the temperature difference are also effective on the condensation heat transfer coefficient. To clarify the parameters that affect the condensation heat transfer

CONDENSATION OF R134A IN A VERTICAL SMOOTH TUBE 439 Figure 4. Test results of refrigerant R134a vapor condensation. coefficient, experimental data were classified according to temperature difference DT (8C), saturation pressure P sat (bar), mass flux G (kg/m 2 s), and average vapor quality x avg. In Figure 5, experimental data points were classified according to similar temperature difference values, and the effect of mass flux on the condensation heat transfer coefficient was analyzed. It was found that at the same mass flux, higher average condensation heat transfer coefficients are obtained at low temperature differences. Adding to that, increasing mass flux led to an increase in the condensation heat transfer coefficient. Figure 5. Relation between mass flux and average condensation heat transfer coefficient.

440 G. ARSLAN AND N. ESKIN In Figure 6, experimental data with similar mass flux and saturation pressure were collected, and the effect of temperature difference on the condensation heat transfer coefficient were analyzed. It is found that at the same temperature difference, higher average condensation heat transfer coefficients are obtained at high mass fluxes. At constant mass flux, the average condensation heat transfer coefficient decreased with increasing temperature differences. For condensation inside vertical smooth tube, the dominated flow pattern is an annular flow. At the center of the tube, vapor core flows downward and liquid film forms inside the tube wall periphery. Despite this, the thickness of the liquid film is very small when compared with tube diameter, and heat removal rate is strictly related with film thickness. During the experiments, it was observed that heat removal rate is directly proportional to the temperature difference. When temperature difference increased, heat removal rate, condensation rate, and liquid film thickness increased. Especially at low mass flux, the main heat transfer mechanism is conduction through the film thickness. Thermal resistance increases with increasing liquid film thickness, and as a result, that process led to a reduction in average condensation heat transfer coefficient. At the same temperature difference and saturation pressure, increasing mass flux results in an increase in average condensation heat transfer coefficient. At this situation, average vapor qualities of the experiments are similar, so condensation rate and vapor velocity is higher for the experiments conducted at high mass fluxes. It is known that at high vapor velocities, interfacial stress and mass transfer effects on heat transfer are remarkable. High interfacial shear stress makes the liquid film thinner; for that reason, an increase in average condensation heat transfer coefficient is observed. Another parameter that affects condensation heat transfer inside tubes is the saturation pressure. In Figure 7, the relation between the average condensation heat Figure 6. Relation between temperature difference and average condensation heat transfer coefficient.

CONDENSATION OF R134A IN A VERTICAL SMOOTH TUBE 441 Figure 7. Effect of saturation pressure on condensation heat transfer coefficient. transfer coefficient and temperature difference for different saturation pressures is given. Since thermo-physical properties of R134a, such as thermal conductivity, viscosity, etc., are inversely proportional to the saturation pressure, a higher average condensation heat transfer coefficient is obtained for the experiments conducted at low saturation pressures. Adding to that, Dalkilic et al. [12] compared the average condensation heat transfer coefficients for the mass flux of 29 kg/m 2 s at saturation pressures of 7.7 9 bar. The represented heat transfer coefficient data range was between 1,100 800 W/m 2 K, and average vapor quality change was given as 0.9. That result verifies trends of experimental data obtained in the current study. Correlations Several correlations from the literature were selected for comparison with the experimental data. These correlations were developed by Jung et al. [4], Akers et al. [21], Shah [16], Traviss et al. [22], Chen et al. [23], Haraguchi et al. [17], Tandon et al. [24], Jaster and Kosky [25], Dobson and Chato [5], Fujii [26], Rosson and Meyers [27], Cavallini and Zecchin [28], and Cavallini et al. [29]. Some of those correlations were developed for the condensation inside horizontal tube. Since the dominated flow pattern is an annular flow for high flow rates in horizontal tubes and for all flow rates in vertical tubes, the general success of the correlations in a vertical flow arrangement was investigated in that study. Correlations developed for estimating local heat transfer coefficient were also compared with experimental data. Here, linear vapor quality change along the test section was assumed, and local heat transfer coefficient was calculated by using correlations. Finally, the average value of the heat transfer coefficient was obtained by using the local data. For each experimental dataset, the obtained condensation heat transfer coefficient was compared with one obtained from correlations, and results are given in Table 2. When

442 G. ARSLAN AND N. ESKIN Table 2. Comparison of correlations with experimental data 1 abs 1 1 st 1 Akers et al. [21] 22.6 216.7 24.6 2 Cavallini et al. [29] 25.8 26.2 31.3 2 Shah [16] 26.0 1.5 30.8 3 Traviss et al. [22] 26.5 24.0 31.8 4 Jung et al. [4] 27.1 211.2 33.3 5 Dobson and Chato [5] 27.5 27.5 9.5 6 Chen et al. [23] 28.0 23.2 23.1 7 Haraguchi et al. [17] 28.5 8.9 31.2 8 Tandon et al. [24] 30.8 230.8 15.9 9 Cavallini and Zecchin [28] 32.4 215.1 37.7 10 Jaster and Kosky [25] 34.2 253.8 16.6 11 Rosson and Meyers [27] 54.9 254.8 23.8 12 Fujii [26] 59.7 259.7 24.2 all experiments are taken into account, Akers et al. s [21], Cavallini et al. s [29], and Shah s [16] correlations estimated experimental condensation heat transfer coefficient with an absolute mean deviation of 22.6, 25.8, and 26%, respectively. In Figure 8, results obtained from most successful correlations (Akers et al. [21], Cavallini et al. [29], Shah [16], Traviss et al. [22], Jung et al. [4], Dobson and Chato [5], and Chen et al. [23]) are given. Most of the correlations mentioned here were developed for an annular flow inside a horizontal tube at high mass flux. For condensation inside a vertical tube, the dominated Figure 8. Comparison of the heat transfer coefficient correlations with experimental data.

CONDENSATION OF R134A IN A VERTICAL SMOOTH TUBE 443 flow regime is an annular flow, and it is independent of the mass flux. Interfacial shear stress and gravitational forces are in the same direction. For that reason, nearly all of the data were predicted by the correlations within ^30% except Fujii s [26] and Rosson and Meyers s [27] correlations. Akers et al. [21] developed their correlation based on the equivalent liquid flow approach. The main idea of this approach is that when the two-phase flow is replaced by an equivalent all liquid flow, an equivalent Reynolds number can be obtained and a singlephase condensation heat transfer coefficient expression will predict the condensation heat transfer coefficient. This correlation is more successful than others because vapor quality change in the test section is low and the equivalent liquid vapor flow approach can be applied. In Figure 8, the obtained results are classified according to the saturation pressure. For the high-pressure region, it is seen that the general behavior of the correlations differ, and the most successful correlations are determined as Dobson and Chato s [5] and Chen et al. s [23] correlations. CONCLUSION There are numerous theoretical and experimental studies in the literature about intube condensation. Since the theory of condensation is complicated, experimental studies gain importance to determine heat transfer characteristics of condensation inside tubes. Under different operating conditions, parameters that effect condensation heat transfer inside tubes have been investigated. In this study, condensation of R134a vapor inside a smooth vertical tube was investigated experimentally. According to the obtained results, a decrease in the average condensation heat transfer coefficient is observed when the temperature difference between the saturated vapor and tube inner wall increased. After a certain temperature difference value, the condensation heat transfer coefficient is not affected from the change of temperature difference. Another result obtained from this experimental study is that average condensation heat transfer coefficient increases with increasing mass flux at the same saturation pressure and temperature difference. Moreover, saturation pressure is very effective on the average condensation heat transfer coefficient. At low saturation pressures, a higher average condensation heat transfer coefficient is obtained. Finally, heat transfer coefficient correlations are compared with experimental data, and best result is obtained from Akers et al. s [21] correlation with an absolute mean deviation of 22.6%. ACKNOWLEDGMENTS Experiments were conducted in Istanbul Technical University, Mechanical Engineering Faculty, Heat and Mass Transfer Laboratory, for which the authors are grateful. FUNDING This study was financially supported by Turkish Scientific and Technical Research Foundation (TUBITAK; project number MAG 108M262) and Istanbul Technical University, Science and Technology Institute.

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