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Experimental study of evaporation heat transfer of R-134a inside a corrugated tube with different tube inclinations
International Communications in Heat and Mass Transfer 55 (2014) 8–14
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International Communications in Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ichmt
Experimental study of evaporation heat transfer of R-134a inside a corrugated tube with different tube inclinations☆ M.A. Akhavan-Behabadi ⁎, M. Esmailpour School of Mechanical Engineering, College of Engineering, University of Tehran, Tehran, Iran
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Available online 26 March 2014 Keywords: Evaporation Heat transfer R-134a Corrugated tube Inclination Two phase flow
a b s t r a c t Experimental heat transfer studies during evaporation of R-134a inside a corrugated tube have been carried out. The corrugated tube has been provided with different tube inclination angles of the direction of fluid flow from horizontal, α. The experiments were performed for seven different tube inclinations, α, in a range of −90° to + 90° and four mass velocities of 46, 81, 110 and 136 kg m−2 s−1 for each tube inclination angle during evaporation of R-134a. Data analysis demonstrate that the tube inclination angle, α, affects the boiling heat transfer coefficient in a significant manner. The effect of tube inclination angle, α, on heat transfer coefficient, h, is more prominent at low vapor quality and mass velocity. In the low vapor quality region, the heat transfer coefficient, h, for the + 90° inclined tube is about 62% more than that of the −90° inclined tube. The results also showed that at all mass velocities, the highest average heat transfer coefficient were achieved for α = +90°. An empirical correlation has also been developed to predict the heat transfer coefficient during flow boiling inside a corrugated tube with different tube inclinations. © 2014 Elsevier Ltd. All rights reserved.
1. Introduction As the energy sources are limited and for their conservation, appropriate design and optimization of condensers is very important. Therefore, the design of efficient evaporators has always been significant to equipment designers. In this regard, the use of augmentative techniques, either active or passive, to increase the convective heat transfer rates on tube side has been studied for quite some time [1–3]. Evaporator is an important and widely used equipment of refrigeration and air-conditioning systems which determines the performance of the whole system. These heat exchangers are widely used in industrial application such as thermal power plants, chemical industries, cooling and air conditioning systems. Accordingly different methods have been proposed for increasing heat transfer in evaporators. One of the passive techniques to enhance heat transfer coefficient is the application of corrugated tubes. Corrugated tubes have corrugation on their surface that can increase heat transfer through creating rotational flow and limiting the growth of thermal boundary layer. This enhancement in heat transfer is also accompanied by slight increase in flow friction factor [4,5]. These kinds of tubes are better heat transfer devices compared to finned tubes, sand-grain textures, wire-coil inserted tubes and transverse rib roughened tubes due to the following reasons: (a) ease of ☆ Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (M.A. Akhavan-Behabadi).
http://dx.doi.org/10.1016/j.icheatmasstransfer.2014.03.003 0735-1933/© 2014 Elsevier Ltd. All rights reserved.
construction, (b) limited fouling and (c) considerable increase in heat transfer, while having a slight impact on pressure drop [6]. There have been many studies concerning heat transfer and pressure drop of single-phase flow in corrugated tube. As for two-phase flows, Targanski and Cieslinski [7] investigated heat transfer and pressure drop during evaporation of refrigerants inside corrugated tubes and reported the increase in heat transfer and pressure drop as compared with smooth tube. Laohalertdecha and Wongwises [8] investigated the effect of corrugation pitch on pressure drop and heat transfer during condensation of R-134a flowing horizontal corrugated tubes. Laohalertdecha and Wongwises [5] studied heat transfer and pressure drop of R-134a flow under evaporation condition inside a horizontal corrugated tube and reported that the ratio of heat transfer enhancement to pressure drop enhancement was 1.2. Aroonrat and Wongwises [9] investigated heat transfer and pressure drop during evaporation of refrigerants inside corrugated tubes and reported the increase in heat transfer and pressure drop as compared with smooth tube. Withers and Young [10] addressed vapor condensation in horizontal smooth and corrugated tubes applied as power plant condensers and showed an increase of 50% in heat transfer as compared with smooth tube. In another study, Laohalertdecha and Wongwises [11] examined the effects of changing geometrical parameters of corrugated tubes (corrugation depths and corrugation pitches) on pressure drop and heat transfer of R-134a condensation flow. Akhavan-behabadi and Khoeini [12] investigated heat transfer characteristics during evaporation of R-134a inside corrugated tube with different tube inclinations and found that the tube inclination angle, α, affects the boiling heat transfer coefficient in
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
Nomenclature CPf Di Do e G h h I k kf L Nu p Prf Q Ref t Ts Twi Two V x Xtt α β η μf μg ρf ρg
specific heat of liquid refrigerant, J kg−1 K−1 internal diameter of test-section, m outside diameter of test-section, m depth of corrugation, m mass velocity, refrigerant mass flow rate per unit crosssectional area, kg m−2 s−1 heat transfer coefficient, W m−2 K average heat transfer coefficient, W m−2 K current intensity of electric heater, A thermal conductivity of tube material, W m−1 K−1 thermal conductivity of liquid phase, W m−1 K−1 average saturation temperature of refrigerant, m Nusselt number, defined in Eq. (7) pitch of corrugation, m Prandtl number of liquid refrigerant, defined in Eq. (7) heat transfer rate from electric heater in the test section, W Reynolds number of liquid refrigerant, defined in Eq. (5) test-section tube wall thickness, m average saturation temperature of refrigerant, K average inside test-section wall temperature, K average outside test-section wall temperature, K voltage of electric heater, V vapor quality Lockhart–Martinelli parameter, defined in Eq. (7) test-section tube inclination angle, ° helix angle, ° heat loss factor of the facility liquid phase dynamic viscosity, Pa s vapor phase dynamic viscosity, Pa s liquid phase density, kg m−3 vapor phase density, kg m−3
accuracy of 2 W). Also a cylindrical brass block was located between heater and corrugated tube to ensure that the heat flux is uniform along the channel. In order to change the inclination angle of the testevaporator, the connections to this evaporator were made by flexible pressure hoses (11). During the experiments, a short range of vapor quality could be achieved in a single test run. Therefore, to cover the entire ranges of vapor quality a pre-evaporator (7) was installed before the test-evaporator. By regulating the voltage to the electrical coil around the pre-evaporator, the vapor qualities at the inlet of test-evaporator was controlled. It was ensured that the refrigerant coming from the test-evaporator is superheated vapor before it entered the compressor. This was attained by installing an after-evaporator (9) downstream of the test-evaporator. All the three evaporators were thermally insulated. An accumulator (10) was also provided before the suction of compressor. The refrigerant mass flow rate was measured by a rotameter (3) installed downstream of condenser. It was ensured that the complete liquid refrigerant enters the rotameter by providing a suitable condenser (2). The rotameter was calibrated prior to installation and the accuracy of it is within ± 1% of the measured value. Refrigerant enters rotameter (3) and enters pre-evaporator (7) after passing through receiver (4), filter drier (5) and expansion valve (6). The outside wall temperature of the test-evaporator tube was measured at six axial locations. At each location, the temperature of tube was measured at top and bottom positions (when corrugated tube is in horizontal position). t ws ¼
2. Experimental facility and procedure The schematic diagram of the test apparatus has been shown in Fig. 1. In fact, the experimental setup was a well-instrumented vapor compression refrigeration system. The test apparatus consisted of a test-evaporator (8) having a test-section length of 1100 mm. The testsection was a copper corrugated tube. The geometrical parameters of this tube are shown in Fig. 2. This tube was heated uniformly by a flexible electrical heating tape of 3.5 kW capacity (with a calibrated
tt þ tb 2
ð1Þ
Thus, the average outside tube wall temperature of the test evaporator, two, was calculated as the arithmetic mean of outside tube wall temperatures at six axial stations: 6 X
t wo ¼ a significant manner. They also found that the highest condensation heat transfer coefficient was obtained for α = + 30° which was 1.41 times greater than the least one obtained for α = − 90°. The results in these studies showed the increase of heat transfer and friction coefficient in such tubes as compared with smooth tube. A review of the existing literature reveals that although vast studies have been done on heat transfer enhancement in these tubes [4,5], the focus of almost all of the studies is on two phase refrigerant flow only in horizontal and vertical (downward flow) tubes. It is obvious that the mechanism of evaporation flow heat transfer in corrugated tubes is so dependent on the patterns of two-phase flow. The heat transfer coefficient generally keeps changing as the flow pattern changes along an evaporator tube. On the other hand, the flow regime is influenced by interfacial shear stress, buoyancy, surface tension and gravitational force. Thus there is a great necessity to consider and study the effect of magnitude and direction of gravity field on evaporation flow inside corrugated tube. Therefore, due to the lack of studies for evaporation flow inside corrugated tubes with different inclinations, this study is aimed for evaporation heat transfer of R-134a flow inside corrugated tube with different inclinations including upward and downward evaporation flow.
9
i¼1
t ws 6:
ð2Þ
The refrigerant temperature at the inlet and outlet of the testevaporator was also measured. All the temperature measurements were done by T-type thermocouples with a calibrated accuracy of 0.1 °C. The arrangements were also made for the measurement of refrigerant pressure at inlet and outlet of the condenser and evaporators. A total of 224 test runs with four different refrigerant mass velocities of 46, 81, 110 and 136 kg m−2 s−1 were performed for seven different tube inclinations from α = −90° to α = +90° (with intervals of 30°). The range of operating parameters is given in Table 1. Here, it should be noted that generally nucleate boiling is a dominant heat transfer mechanism at low vapor quality. In this region, the heat transfer coefficient is dependent on heat flux. Therefore we ran tests for all tube inclination angles under identical operating conditions to study the effect of gravitational force on heat transfer rate. The vapor quality at the inlet of pre-evaporator was computed by considering the isoenthalpic expansion of refrigerant in the needle valve. An energy balance technique was used to get the vapor quality of the inlet and the exit of test-evaporator (heat gained by R-134a is equal to the heat rejected by heater). The vapor quality of test-evaporator was obtained by taking the average of vapor qualities at the inlet and outlet. The heat transfer coefficient, h, was calculated on the basis of heat gained by R-134a in test-evaporator and the temperature difference between inside wall of test-evaporator and boiling refrigerant. Therefore, for each test run, the heat transfer coefficient was computed by using the following equation:
h¼
−1 πDi LðT wo −T s Þ Di Do þ ln : Q 2k Di
ð3Þ
10
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
Fig. 1. Schematic diagram of the experimental setup.
The outside tube wall temperature, Two, was taken as the arithmetic mean of outside tube wall temperatures at top and bottom positions. The heating rate by the electric heater, Q, can be calculated from current intensity and voltage readings as follows: Q ¼ ηVI
ð4Þ
where the factor η is experimentally determined (typically around 0.95) and accounts for the heat lost to the environment. The thermo physical properties of R-134a were taken from ASHRAE Handbook of Fundamentals [13]. The uncertainty analysis of experimental results has been carried out by the method proposed in [14]. Schultz and Cole [14] suggested the following method of analyzing the effect of uncertainty in each variable on the uncertainty of the results: " Uh ¼
n X ∂h i¼1
∂V i
2 #1=2 ð5Þ
UVi
where UR is the estimate of uncertainty in the calculated value of the desired variable R, due to the independent uncertainty U V i in the primary measurement of n number of variables, Vi, affecting the result. Applying Eq. (5), the uncertainty in the heat transfer coefficient calculation can be written as ( Uh ¼
2
Uq ðt wi −t s Þ
þ
q:U t s
ðt wi −t s Þ2
2
þ
−q:U t wi
ðt wi −t s Þ2
2 )1=2 ð6Þ
It was found that the uncertainty in the determination of flow boiling heat transfer coefficient is within 9.2% for all the test runs. The schematic diagram of the experimental facility is shown in Fig. 1. The flow loop consisted of a test section, steam chamber, cooler, reservoir, gear pump, flow measuring apparatus, thermocouples and flow controlling system. The test section itself was a copper tube surrounded by a steam chamber to keep the temperature of the tube wall nearly constant. The fluid leaving the test section entered the flow measuring apparatus, and then is cooled down partially and restored in the reservoir. The gear pump drove the flow into a bypass valve which in turn divided the pumped flow into two separate flows; one passing through the test section and the other one bypassed directly to the reservoir. The experimental apparatus was designed to measure heat transfer as well as pressure drop characteristics of working fluids over the length of the test section. Five 1200 mm long round copper tubes were used. The outer diameter and wall thickness of those tubes were 15.87 mm and 0.75 mm, respectively. Four tubes were flattened into oblong shapes with internal heights of 13.4 mm, 11.7 mm, 10.6 mm and 8.6 mm and one was used as a round tube. Fig. 2 shows the crosssectional area of the applied test sections and Table 1 indicates the dimensions and corresponding hydraulic diameters. 3. Results and discussion First of all, the experimental heat transfer coefficients have been compared with those of Ref. [5] for the evaporation inside a horizontal corrugated tube. In Fig. 3, such a comparison has been made taking experimental heat transfer coefficient as abscissa and predicted heat
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
11
Fig. 3. Comparison of experimental heat transfer coefficients with those predicted by Laohalertdecha and Wongwises [5]. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
di=8.3mm do=9.5mm p=8mm Wt=0.6mm β=75 e=1.5mm Fig. 2. Geometry of corrugated tube. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
transfer coefficient as ordinate. From Fig. 3, it is observed that the correlation of Laohalertdecha and Wongwises [5] predicts the experimental heat-transfer coefficient, from the present investigation within an error range of −13.76% to + 14.11%. The agreement of the experimental results with those predicted by the correlation suggested by Laohalertdecha and Wongwises [5] establishes the integrity of the experimental apparatus. Dependence of heat transfer coefficient to vapor quality is derived of distribution of two boiling mechanism (nucleate and convective) and boiling flow pattern. The variation of evaporation heat transfer coefficient, h, with vapor quality for seven different tube inclination angles (α = − 90° to α = + 90°) at the mass velocity of 136 kg m− 2 s−1, has been shown in Fig. 4. From this figure, it is found that for all the tube inclination angles, the heat transfer coefficient increases with the increase in vapor quality until a vapor quality near 70–85% when it begins to decrease because of dryout. This is because the thickness of liquid film on inner wall of the tube decreases during evaporation in the flow direction and the void fraction increases, and the density of the liquid–vapor mixture decreases which would result in lower thermal resistance. As a result, higher evaporation heat transfer coefficients are observed as vapor quality increases until dryout occurs. The change in the orientation of
heat transfer coefficient with vapor quality might be because of change in flow pattern and different boiling mechanism (nucleate and convective). Looking at Fig. 4, one can conclude that the trend of variation in heat transfer coefficients versus vapor quality for each inclination is different from the others. This could be explained by the instability of heat transfer due to different boiling mechanism (nucleate and convection), phase change in the flow direction and also the creation of different flow patterns in each inclination due to corrugation on the tube [12]. The interaction between nucleate boiling and liquid convection influences heat transfer coefficient in flow boiling. Heat transfer is a strong function of the nucleate boiling mechanism especially at high heat fluxes and evaporating pressures but at low flux and evaporation pressure the convection is the main heat transfer mechanism [15]. In the present study, the convective contribution to heat transfer predominates at experimental tests. In Fig. 4, it has been shown that the heat transfer coefficient increases with vapor quality. Continues enhancing in heat transfer coefficient proceeds until the liquid film vanishes, leaving the tube-wall partially or totally dry. Indeed, as the flow vaporize, the void fraction increases, and the density of the liquid–vapor mixture decreases. As a result, the flow accelerates enhancing convective transport from the
Table 1 List of operating parameters. Working fluid
R-134a
Refrigerant mass velocity Average evaporating temperature heat flux Inlet vapor quality Exit vapor quality
46–136 kg m−2 s−1 −26 to −2 °C 4.56 to 9.13 kw/m2 0.2–0.9 0.3–1.0
Fig. 4. Variation of heat transfer coefficient with vapor quality at refrigerant mass velocity of 136 kg m−2 s−1. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
12
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
heated wall of the tube. In this region, because of the low thermal conductivity of the vapor, the heat transfer coefficient decreases. The maximum heat transfer coefficient occurs at a vapor quality which depends upon mass velocity and the tube inclination. With increase in mass velocity and increase in tube inclination angle (from − 90 to + 90), the dryout occurs in lower vapor quality thus the maximum heat transfer coefficient occurs earlier. for example, for mass velocity of 136 kg m−2 s−1, maximum heat transfer coefficient occurs at a lower vapor quality comparably with mass velocity of 46 kg m−2 s−1. For the mass velocity of 136 kg m−2 s−1, dryout occurs at x = 78–80% in the tube with inclination angle of −90° and at x = 70–71% for the tube with inclination angle of +90°. For the mass velocity of 46 kg m−2 s−1, dryout occurs at x = 79% and x = 85% for the tubes with inclination angle of +90° and −90°, respectively. The tube inclination angle influences the heat transfer coefficient, h, in a significant manner. The tube with +90° inclination angle (vertical upward flow) turns out to be the best performing tube at low vapor quality region. The tube having inclination angle of − 90° (vertical downward flow) has the lowest heat transfer coefficient, h. AkhavanBehabadi et al. [16] also reported similar results on heat transfer of evaporating flow in inclined microfin tubes. The performance of the tube with + 90° inclination angle is much superior to that of tube with −90° inclination angle in the low vapor quality region where the heat transfer coefficient, h, for +90° inclined tube is about 39.9% more than that of the −90° inclined tube. However, the average heat transfer coefficient, h, for the tube with +90° inclination is 30% more than that for the tube with −90° inclination angle for the entire ranges of vapor qualities. For a given mass velocity, the average heat transfer coefficient, h, is defined as the arithmetic mean of heat transfer coefficients, h, calculated at different vapor qualities. In the low vapor quality region and at the lowest mass velocity, the heat transfer coefficient, h, for +90° inclined tube is about 62% more than that of the − 90° inclined tube. In comparison to the performance of the horizontal tube, the heat transfer coefficient, h, for the tube having inclination angle of + 90° is 20.2% more at lower vapor quality, while the average heat transfer coefficient, h, is 15.7% more than that for a horizontal tube, which lies within the uncertainty of 9.2% in experimental data. The variation of heat transfer coefficient with vapor quality for a corrugated tube with different inclinations is shown in Figs. 5, 6 and 7 for the mass velocity of 46 kg m−2 s−1, 81 kg m−2 s−1 and 110 kg m−2 s−1 respectively. The tube with α = +90° has still a better heat transfer performance. Moreover, the increases of 36.9%, 17.2%, and 13.4% in heat transfer coefficient at low vapor qualities and increase of 20.8%, 12.8% and 9.1% in average heat transfer of the test evaporator with α = + 90° as compared to the test evaporator with α = − 90° are observed for
mass velocities of 81, 110 and 136 [kg m−2 s−1], respectively. In these figures, it is observed that the change in tube inclination has considerable effect on evaporation heat transfer coefficient. In fact, at low vapor quality, the vapor phase velocity is reduced, resulting in a decrease in interfacial shear stresses and inertia force as well. Therefore, at low vapor quality, the gravitational force has considerable effect on two-phase flow, and this is the major reason for different rates of heat transfer coefficient for different inclinations of corrugated tube. In other words, interfacial shear stress is dominant at high vapor velocity and high vapor quality, while at low vapor qualities, the vapor phase motion is slowed down, and this would result in a decrease in interfacial shear stress and consequently lower inertia forces near the wall. Therefore, the effect of gravity force on the evaporation of R134a flow at low vapor qualities is more considerable as it can be seen in Figs. 4, 5, 6 and 7 for different inclinations of test evaporator. At low velocities where buoyancy is dominant, the effect of orientation is very pronounced. The random variation in heat transfer coefficient, h, is also visible in these figures. This is due to the instability of heat transfer, in which both the phases are flowing together during phase change. In addition, the presence of the different flow regimes inside the tube with different tube inclinations causes these random variations. At low vapor quality, the increase in evaporation heat transfer coefficient in tube with inclination angle of α = +90° in comparison with other inclination angles is probably due to the better flow of liquid inside grooves and over tube inside wall because of favorable direction of gravitational force. Indeed, during upward vertical flow, α = +90°,
Fig. 5. Variation of heat transfer coefficient with vapor quality at refrigerant mass velocity of 46 kg m−2 s−1. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
Fig. 7. Variation of heat transfer coefficient with vapor quality at refrigerant mass velocity of 110 kg m−2 s−1. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
Fig. 6. Variation of heat transfer coefficient with vapor quality at refrigerant mass velocity of 81 kg m−2 s−1. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
the buoyancy force and fluid flow are unidirectional and as a result the flow accelerates enhancing convective transport from the heated wall of the tube [16]. During upward vertical flow of vapor, due to high interfacial turbulence, higher heat transfer coefficient is observed in comparison to that for downward vertical flow, α = − 90°, in low vapor qualities. Furthermore, vertical tube does not suffer from the adverse effect of flow stratification like a horizontal tube. In the dryout region, at high vapor quality, upward vertical boiling flow has the lower heat transfer coefficient than that for the horizontal flow. In other words, when dryout happens, the wetted perimeter of the tube is the most effective region for heat transfer. At high vapor quality in horizontal tube (also tubes with inclinations close to horizontal), the liquid tends to flow in lower side of the tube, and as a result, the flow boiling heat transfer in the flooded part of the tube takes place while mist flow in the form of entrained droplets occurs during upward vertical flow [17]. At vapor quality of 0.95 and refrigerant mass velocities of 46 and 81 kg m− 2 s− 1, the heat transfer coefficients of horizontal tube are 26.3% and 11.2% more than those of − 90° inclined tube, respectively. From Figs. 4 to 7, it is also noted that the vertical tube with inclination angle of −90° has the lowest heat transfer coefficient. Wongwises et al. [9] also reported the same observations in their study of orientation effects for R-134a flow boiling in vertical downward flow inside a corrugated tube. In the downward vertical flow, since the interfacial shear stress and the gravity are acting in the same direction, the flow pattern will remain in annular status [12]. The phenomenon of annular flow causes to form a thick layer of liquid film around the periphery of tube as a result the heat transfer coefficient, h, is reduced. Since the gravitational force and vapor shear stress are in same direction, the interfacial turbulence is the lowest [18]. In Fig. 8, the variation of heat-transfer coefficient with vapor quality at different mass velocities of R-134a is shown for the tube having inclination angle of +90°. The experimental data clearly show that the heat transfer coefficients increase with increasing the refrigerant mass velocity. Indeed, increasing the refrigerant mass velocity increase the fluid velocity, thus enhancing convective boiling. The increase in mass velocity changes the flow pattern from stratified flow to annular flow. In convective boiling, when fluid velocity increases, the heat transfer coefficient increase. In higher mass velocity, effects of convection are much higher and more turbulence occurs, as a result heat transfer coefficient increase. With the increase in mass velocity, flow pattern changes from stratified flow to annular flow in a much lower vapor quality, and the increase in vapor quality in annular flow causes more increase in heat transfer relative to stratified flow. With the increase in mass velocity, maximum heat transfer coefficient will occur in much lower vapor quality because with the increase
Fig. 8. Effect of mass velocity on heat transfer coefficient at inclination angle of +90°. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
13
in mass velocity, dryout will occur sooner. The average heat transfer coefficient, h, at the mass velocity of 136 kg m−2 s− 1 is nearly 128% more than that for the mass velocity of 46 kg m−2 s−1. Finally based on the present experimental results, the following correlation, Eq. (7), has been developed to predict the heat transfer coefficient at different vapor qualities, mass velocities and tube inclinations. −3
Nu ¼ 3:79 10
1:02
Re f
0:11
Fα
0:92 Pr f X tt
ð7Þ
Where, hD GDð1−xÞ ; Re f ¼ ; Pr f kf μf !0:1 0:9 0:5 μ f cpf ρg 1−x μf ¼ ; Xtt ¼ x kf ρf μg
Nu ¼
0:6
F α ¼ 1 þ 0:25ð1 þ xÞ
F α ¼ 1−0:6x
0:97
Sinðα Þ
for x ≤ 0:7
Cos α−10 for x N 0:7
ð8Þ
ð8aÞ
ð8bÞ
The correlation given in Eq. (7) predicts the experimental data for all angles of inclination of corrugated tube in an error band of ± 10% as shown in Fig. 9. 4. Conclusion The following conclusions have been drawn from the present investigation: 1. For all tube inclination angles, the heat transfer coefficient increases with the increase in vapor quality until a vapor quality near 70–85% when it begins to decrease because of dryout. 2. The tube inclination angle, α, affects the boiling heat transfer coefficient in a significant manner. The effect of inclination angle is more prominent at low vapor qualities and mass velocities. 3. At low vapor qualities, the highest heat transfer coefficient is attained at inclination angle of + 90° (vertical upward flow), and at high vapor qualities, the highest heat transfer coefficient occurs when the corrugated tube is horizontal. The vertical tube with inclination angle of −90° has the lowest heat transfer coefficient for the entire
Fig. 9. Comparison of experimental heat transfer coefficient with those predicted by proposed correlation.
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M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
range of vapor quality. The corrugated tube having inclination angle of + 90° outperforms the tube with − 90° inclination in a range of 7–62%. 4. A correlation has been developed, Eq. (7), which predicts the evaporation heat transfer coefficients of R-134a inside a corrugated tube with different tube inclinations within an error range of ±10%. Acknowledgment The authors would like to express their thanks to School of Mechanical Engineering, College of Engineering, University of Tehran, for financial support of the present experimental work. References [1] K.N. Agrawal, H.K. Varma, Experimental study of heat transfer augmentation versus pumping power in a horizontal R-12 evaporator, Int. J. Refrig. 14 (1991) 273–281. [2] M.A. Akhavan-Behabadi, Ravi Kumar, A. Mohammadpour, Effect of twisted tape insert on heat transfer and pressure drop in horizontal evaporator for the flow of R-134a, Int. J. Refrig. 32 (9) (2009) 922–930. [3] S. Wongwises, M. Polsongkram, Evaporation heat transfer and pressure drop of HFC-134a in a helically coiled concentric tube-in-tube heat exchanger, Int. J. Heat Mass Transf. 49 (2005) 658–670. [4] S. Lohalertdecha, A.S. Dalkilic, S. Wongwises, Correlation for evaporation heat transfer coefficient and two-phase friction factor for R-134a flowing through horizontal corrugated tubes, Int. J. Heat Mass Transf. 38 (2011) 1406–1413. [5] S. Lohalertdecha, S. Wongwises, An experimental study into the evaporation heat transfer and flow characteristics of R-134a refrigerant flowing through corrugated tubes, Int. J. Heat Mass Transf. 34 (2011) 280–291.
[6] V.D. Zimparov, N.L. Vulchanov, L.B. Delov, Heat transfer and friction characteristic of spirally corrugated tube for power plants condensors-1. Experimental investigation and performance evaluation, Int. J. Heat Mass Transf. 34 (1991) 2187–2197. [7] T. Targanski, J.T. Cieslinski, Evaporation of R407C/oil mixtures inside corrugated and micro-fin tubes, Appl. Therm. Eng. 27 (2007) 2226–2232. [8] S. Laohalertdecha, S. Wongwises, The effects of corrugation pitch onthe condensation heat transfer coefficient and pressure drop of R-134a inside horizontal corrugated tube, Int. J. Heat Mass Transfer 53 (2010) 2924–2931. [9] K. Aroonrat, S. Wongwises, Evaporation heat transfer and friction characteristics of R-134a flowing downward in a vertical corrugated tube, Exp. Thermal Fluid Sci. 35 (2011) 20–28. [10] J.G. Withers, E.H. Yong, Steam condensation on vertical rows of horizontal corrugated and plain tube, Ind. Eng. Chem. Process Des. Dev. 10 (1) (1971) 19–30. [11] S. Laohalertdecha, S. Wongwises, Condensation heat transfer and flow characteristics of R-134a flowing through corrugated tubes, Int. J Heat Mass Transfer 54 (2011) 2673–2682. [12] D. Khoeini, M.A. Akhavan-Behabadi, A. Saboonchi, Experimental study of condensation heat transfer of R-134a flow in corrugated tubes with different inclination, Int. Commun. Heat and Mass Transfer (2011) 1001–1006. [13] American society of heating refrigerating and air-conditioning engineers handbook, Fundamentals, 2005. 17 (Chapter 20). [14] R.R. Schultz, R. Cole, Uncertainty analysis in boiling nucleation, Am. Inst. Chem. Eng. Symp. Ser. 75 (189) (1979) 32–38. [15] A. Greco, Convective boiling of pure and mixed refrigerants, Int. J. Heat Mass Transfer 51 (2008) 896–909. [16] M.A. Akhavan-Behabadi, S.G. Mohseni, S.M. Razavinasab, Evaporation heat transfer of R-134a inside a microfin tube with different tube inclinations, Exp. Thermal Fluid Sci. 35 (2010) 996–1001. [17] J.G. Collier, J.R. Thome, Convective Boiling and Condensation, 3rd edition Oxford University Press, Oxford, 1994. 13–15. [18] M.A. Akhavan-Behabadi, Ravi Kumar, S.G. Mohseni, Condensation heat transfer of R134a inside a microfin tube with different tube inclinations, Int. J. Heat Mass Transfer 50 (2007) 4864–4871
Contents lists available at ScienceDirect
International Communications in Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ichmt
Experimental study of evaporation heat transfer of R-134a inside a corrugated tube with different tube inclinations☆ M.A. Akhavan-Behabadi ⁎, M. Esmailpour School of Mechanical Engineering, College of Engineering, University of Tehran, Tehran, Iran
a r t i c l e
i n f o
Available online 26 March 2014 Keywords: Evaporation Heat transfer R-134a Corrugated tube Inclination Two phase flow
a b s t r a c t Experimental heat transfer studies during evaporation of R-134a inside a corrugated tube have been carried out. The corrugated tube has been provided with different tube inclination angles of the direction of fluid flow from horizontal, α. The experiments were performed for seven different tube inclinations, α, in a range of −90° to + 90° and four mass velocities of 46, 81, 110 and 136 kg m−2 s−1 for each tube inclination angle during evaporation of R-134a. Data analysis demonstrate that the tube inclination angle, α, affects the boiling heat transfer coefficient in a significant manner. The effect of tube inclination angle, α, on heat transfer coefficient, h, is more prominent at low vapor quality and mass velocity. In the low vapor quality region, the heat transfer coefficient, h, for the + 90° inclined tube is about 62% more than that of the −90° inclined tube. The results also showed that at all mass velocities, the highest average heat transfer coefficient were achieved for α = +90°. An empirical correlation has also been developed to predict the heat transfer coefficient during flow boiling inside a corrugated tube with different tube inclinations. © 2014 Elsevier Ltd. All rights reserved.
1. Introduction As the energy sources are limited and for their conservation, appropriate design and optimization of condensers is very important. Therefore, the design of efficient evaporators has always been significant to equipment designers. In this regard, the use of augmentative techniques, either active or passive, to increase the convective heat transfer rates on tube side has been studied for quite some time [1–3]. Evaporator is an important and widely used equipment of refrigeration and air-conditioning systems which determines the performance of the whole system. These heat exchangers are widely used in industrial application such as thermal power plants, chemical industries, cooling and air conditioning systems. Accordingly different methods have been proposed for increasing heat transfer in evaporators. One of the passive techniques to enhance heat transfer coefficient is the application of corrugated tubes. Corrugated tubes have corrugation on their surface that can increase heat transfer through creating rotational flow and limiting the growth of thermal boundary layer. This enhancement in heat transfer is also accompanied by slight increase in flow friction factor [4,5]. These kinds of tubes are better heat transfer devices compared to finned tubes, sand-grain textures, wire-coil inserted tubes and transverse rib roughened tubes due to the following reasons: (a) ease of ☆ Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (M.A. Akhavan-Behabadi).
http://dx.doi.org/10.1016/j.icheatmasstransfer.2014.03.003 0735-1933/© 2014 Elsevier Ltd. All rights reserved.
construction, (b) limited fouling and (c) considerable increase in heat transfer, while having a slight impact on pressure drop [6]. There have been many studies concerning heat transfer and pressure drop of single-phase flow in corrugated tube. As for two-phase flows, Targanski and Cieslinski [7] investigated heat transfer and pressure drop during evaporation of refrigerants inside corrugated tubes and reported the increase in heat transfer and pressure drop as compared with smooth tube. Laohalertdecha and Wongwises [8] investigated the effect of corrugation pitch on pressure drop and heat transfer during condensation of R-134a flowing horizontal corrugated tubes. Laohalertdecha and Wongwises [5] studied heat transfer and pressure drop of R-134a flow under evaporation condition inside a horizontal corrugated tube and reported that the ratio of heat transfer enhancement to pressure drop enhancement was 1.2. Aroonrat and Wongwises [9] investigated heat transfer and pressure drop during evaporation of refrigerants inside corrugated tubes and reported the increase in heat transfer and pressure drop as compared with smooth tube. Withers and Young [10] addressed vapor condensation in horizontal smooth and corrugated tubes applied as power plant condensers and showed an increase of 50% in heat transfer as compared with smooth tube. In another study, Laohalertdecha and Wongwises [11] examined the effects of changing geometrical parameters of corrugated tubes (corrugation depths and corrugation pitches) on pressure drop and heat transfer of R-134a condensation flow. Akhavan-behabadi and Khoeini [12] investigated heat transfer characteristics during evaporation of R-134a inside corrugated tube with different tube inclinations and found that the tube inclination angle, α, affects the boiling heat transfer coefficient in
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
Nomenclature CPf Di Do e G h h I k kf L Nu p Prf Q Ref t Ts Twi Two V x Xtt α β η μf μg ρf ρg
specific heat of liquid refrigerant, J kg−1 K−1 internal diameter of test-section, m outside diameter of test-section, m depth of corrugation, m mass velocity, refrigerant mass flow rate per unit crosssectional area, kg m−2 s−1 heat transfer coefficient, W m−2 K average heat transfer coefficient, W m−2 K current intensity of electric heater, A thermal conductivity of tube material, W m−1 K−1 thermal conductivity of liquid phase, W m−1 K−1 average saturation temperature of refrigerant, m Nusselt number, defined in Eq. (7) pitch of corrugation, m Prandtl number of liquid refrigerant, defined in Eq. (7) heat transfer rate from electric heater in the test section, W Reynolds number of liquid refrigerant, defined in Eq. (5) test-section tube wall thickness, m average saturation temperature of refrigerant, K average inside test-section wall temperature, K average outside test-section wall temperature, K voltage of electric heater, V vapor quality Lockhart–Martinelli parameter, defined in Eq. (7) test-section tube inclination angle, ° helix angle, ° heat loss factor of the facility liquid phase dynamic viscosity, Pa s vapor phase dynamic viscosity, Pa s liquid phase density, kg m−3 vapor phase density, kg m−3
accuracy of 2 W). Also a cylindrical brass block was located between heater and corrugated tube to ensure that the heat flux is uniform along the channel. In order to change the inclination angle of the testevaporator, the connections to this evaporator were made by flexible pressure hoses (11). During the experiments, a short range of vapor quality could be achieved in a single test run. Therefore, to cover the entire ranges of vapor quality a pre-evaporator (7) was installed before the test-evaporator. By regulating the voltage to the electrical coil around the pre-evaporator, the vapor qualities at the inlet of test-evaporator was controlled. It was ensured that the refrigerant coming from the test-evaporator is superheated vapor before it entered the compressor. This was attained by installing an after-evaporator (9) downstream of the test-evaporator. All the three evaporators were thermally insulated. An accumulator (10) was also provided before the suction of compressor. The refrigerant mass flow rate was measured by a rotameter (3) installed downstream of condenser. It was ensured that the complete liquid refrigerant enters the rotameter by providing a suitable condenser (2). The rotameter was calibrated prior to installation and the accuracy of it is within ± 1% of the measured value. Refrigerant enters rotameter (3) and enters pre-evaporator (7) after passing through receiver (4), filter drier (5) and expansion valve (6). The outside wall temperature of the test-evaporator tube was measured at six axial locations. At each location, the temperature of tube was measured at top and bottom positions (when corrugated tube is in horizontal position). t ws ¼
2. Experimental facility and procedure The schematic diagram of the test apparatus has been shown in Fig. 1. In fact, the experimental setup was a well-instrumented vapor compression refrigeration system. The test apparatus consisted of a test-evaporator (8) having a test-section length of 1100 mm. The testsection was a copper corrugated tube. The geometrical parameters of this tube are shown in Fig. 2. This tube was heated uniformly by a flexible electrical heating tape of 3.5 kW capacity (with a calibrated
tt þ tb 2
ð1Þ
Thus, the average outside tube wall temperature of the test evaporator, two, was calculated as the arithmetic mean of outside tube wall temperatures at six axial stations: 6 X
t wo ¼ a significant manner. They also found that the highest condensation heat transfer coefficient was obtained for α = + 30° which was 1.41 times greater than the least one obtained for α = − 90°. The results in these studies showed the increase of heat transfer and friction coefficient in such tubes as compared with smooth tube. A review of the existing literature reveals that although vast studies have been done on heat transfer enhancement in these tubes [4,5], the focus of almost all of the studies is on two phase refrigerant flow only in horizontal and vertical (downward flow) tubes. It is obvious that the mechanism of evaporation flow heat transfer in corrugated tubes is so dependent on the patterns of two-phase flow. The heat transfer coefficient generally keeps changing as the flow pattern changes along an evaporator tube. On the other hand, the flow regime is influenced by interfacial shear stress, buoyancy, surface tension and gravitational force. Thus there is a great necessity to consider and study the effect of magnitude and direction of gravity field on evaporation flow inside corrugated tube. Therefore, due to the lack of studies for evaporation flow inside corrugated tubes with different inclinations, this study is aimed for evaporation heat transfer of R-134a flow inside corrugated tube with different inclinations including upward and downward evaporation flow.
9
i¼1
t ws 6:
ð2Þ
The refrigerant temperature at the inlet and outlet of the testevaporator was also measured. All the temperature measurements were done by T-type thermocouples with a calibrated accuracy of 0.1 °C. The arrangements were also made for the measurement of refrigerant pressure at inlet and outlet of the condenser and evaporators. A total of 224 test runs with four different refrigerant mass velocities of 46, 81, 110 and 136 kg m−2 s−1 were performed for seven different tube inclinations from α = −90° to α = +90° (with intervals of 30°). The range of operating parameters is given in Table 1. Here, it should be noted that generally nucleate boiling is a dominant heat transfer mechanism at low vapor quality. In this region, the heat transfer coefficient is dependent on heat flux. Therefore we ran tests for all tube inclination angles under identical operating conditions to study the effect of gravitational force on heat transfer rate. The vapor quality at the inlet of pre-evaporator was computed by considering the isoenthalpic expansion of refrigerant in the needle valve. An energy balance technique was used to get the vapor quality of the inlet and the exit of test-evaporator (heat gained by R-134a is equal to the heat rejected by heater). The vapor quality of test-evaporator was obtained by taking the average of vapor qualities at the inlet and outlet. The heat transfer coefficient, h, was calculated on the basis of heat gained by R-134a in test-evaporator and the temperature difference between inside wall of test-evaporator and boiling refrigerant. Therefore, for each test run, the heat transfer coefficient was computed by using the following equation:
h¼
−1 πDi LðT wo −T s Þ Di Do þ ln : Q 2k Di
ð3Þ
10
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
Fig. 1. Schematic diagram of the experimental setup.
The outside tube wall temperature, Two, was taken as the arithmetic mean of outside tube wall temperatures at top and bottom positions. The heating rate by the electric heater, Q, can be calculated from current intensity and voltage readings as follows: Q ¼ ηVI
ð4Þ
where the factor η is experimentally determined (typically around 0.95) and accounts for the heat lost to the environment. The thermo physical properties of R-134a were taken from ASHRAE Handbook of Fundamentals [13]. The uncertainty analysis of experimental results has been carried out by the method proposed in [14]. Schultz and Cole [14] suggested the following method of analyzing the effect of uncertainty in each variable on the uncertainty of the results: " Uh ¼
n X ∂h i¼1
∂V i
2 #1=2 ð5Þ
UVi
where UR is the estimate of uncertainty in the calculated value of the desired variable R, due to the independent uncertainty U V i in the primary measurement of n number of variables, Vi, affecting the result. Applying Eq. (5), the uncertainty in the heat transfer coefficient calculation can be written as ( Uh ¼
2
Uq ðt wi −t s Þ
þ
q:U t s
ðt wi −t s Þ2
2
þ
−q:U t wi
ðt wi −t s Þ2
2 )1=2 ð6Þ
It was found that the uncertainty in the determination of flow boiling heat transfer coefficient is within 9.2% for all the test runs. The schematic diagram of the experimental facility is shown in Fig. 1. The flow loop consisted of a test section, steam chamber, cooler, reservoir, gear pump, flow measuring apparatus, thermocouples and flow controlling system. The test section itself was a copper tube surrounded by a steam chamber to keep the temperature of the tube wall nearly constant. The fluid leaving the test section entered the flow measuring apparatus, and then is cooled down partially and restored in the reservoir. The gear pump drove the flow into a bypass valve which in turn divided the pumped flow into two separate flows; one passing through the test section and the other one bypassed directly to the reservoir. The experimental apparatus was designed to measure heat transfer as well as pressure drop characteristics of working fluids over the length of the test section. Five 1200 mm long round copper tubes were used. The outer diameter and wall thickness of those tubes were 15.87 mm and 0.75 mm, respectively. Four tubes were flattened into oblong shapes with internal heights of 13.4 mm, 11.7 mm, 10.6 mm and 8.6 mm and one was used as a round tube. Fig. 2 shows the crosssectional area of the applied test sections and Table 1 indicates the dimensions and corresponding hydraulic diameters. 3. Results and discussion First of all, the experimental heat transfer coefficients have been compared with those of Ref. [5] for the evaporation inside a horizontal corrugated tube. In Fig. 3, such a comparison has been made taking experimental heat transfer coefficient as abscissa and predicted heat
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
11
Fig. 3. Comparison of experimental heat transfer coefficients with those predicted by Laohalertdecha and Wongwises [5]. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
di=8.3mm do=9.5mm p=8mm Wt=0.6mm β=75 e=1.5mm Fig. 2. Geometry of corrugated tube. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
transfer coefficient as ordinate. From Fig. 3, it is observed that the correlation of Laohalertdecha and Wongwises [5] predicts the experimental heat-transfer coefficient, from the present investigation within an error range of −13.76% to + 14.11%. The agreement of the experimental results with those predicted by the correlation suggested by Laohalertdecha and Wongwises [5] establishes the integrity of the experimental apparatus. Dependence of heat transfer coefficient to vapor quality is derived of distribution of two boiling mechanism (nucleate and convective) and boiling flow pattern. The variation of evaporation heat transfer coefficient, h, with vapor quality for seven different tube inclination angles (α = − 90° to α = + 90°) at the mass velocity of 136 kg m− 2 s−1, has been shown in Fig. 4. From this figure, it is found that for all the tube inclination angles, the heat transfer coefficient increases with the increase in vapor quality until a vapor quality near 70–85% when it begins to decrease because of dryout. This is because the thickness of liquid film on inner wall of the tube decreases during evaporation in the flow direction and the void fraction increases, and the density of the liquid–vapor mixture decreases which would result in lower thermal resistance. As a result, higher evaporation heat transfer coefficients are observed as vapor quality increases until dryout occurs. The change in the orientation of
heat transfer coefficient with vapor quality might be because of change in flow pattern and different boiling mechanism (nucleate and convective). Looking at Fig. 4, one can conclude that the trend of variation in heat transfer coefficients versus vapor quality for each inclination is different from the others. This could be explained by the instability of heat transfer due to different boiling mechanism (nucleate and convection), phase change in the flow direction and also the creation of different flow patterns in each inclination due to corrugation on the tube [12]. The interaction between nucleate boiling and liquid convection influences heat transfer coefficient in flow boiling. Heat transfer is a strong function of the nucleate boiling mechanism especially at high heat fluxes and evaporating pressures but at low flux and evaporation pressure the convection is the main heat transfer mechanism [15]. In the present study, the convective contribution to heat transfer predominates at experimental tests. In Fig. 4, it has been shown that the heat transfer coefficient increases with vapor quality. Continues enhancing in heat transfer coefficient proceeds until the liquid film vanishes, leaving the tube-wall partially or totally dry. Indeed, as the flow vaporize, the void fraction increases, and the density of the liquid–vapor mixture decreases. As a result, the flow accelerates enhancing convective transport from the
Table 1 List of operating parameters. Working fluid
R-134a
Refrigerant mass velocity Average evaporating temperature heat flux Inlet vapor quality Exit vapor quality
46–136 kg m−2 s−1 −26 to −2 °C 4.56 to 9.13 kw/m2 0.2–0.9 0.3–1.0
Fig. 4. Variation of heat transfer coefficient with vapor quality at refrigerant mass velocity of 136 kg m−2 s−1. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
12
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
heated wall of the tube. In this region, because of the low thermal conductivity of the vapor, the heat transfer coefficient decreases. The maximum heat transfer coefficient occurs at a vapor quality which depends upon mass velocity and the tube inclination. With increase in mass velocity and increase in tube inclination angle (from − 90 to + 90), the dryout occurs in lower vapor quality thus the maximum heat transfer coefficient occurs earlier. for example, for mass velocity of 136 kg m−2 s−1, maximum heat transfer coefficient occurs at a lower vapor quality comparably with mass velocity of 46 kg m−2 s−1. For the mass velocity of 136 kg m−2 s−1, dryout occurs at x = 78–80% in the tube with inclination angle of −90° and at x = 70–71% for the tube with inclination angle of +90°. For the mass velocity of 46 kg m−2 s−1, dryout occurs at x = 79% and x = 85% for the tubes with inclination angle of +90° and −90°, respectively. The tube inclination angle influences the heat transfer coefficient, h, in a significant manner. The tube with +90° inclination angle (vertical upward flow) turns out to be the best performing tube at low vapor quality region. The tube having inclination angle of − 90° (vertical downward flow) has the lowest heat transfer coefficient, h. AkhavanBehabadi et al. [16] also reported similar results on heat transfer of evaporating flow in inclined microfin tubes. The performance of the tube with + 90° inclination angle is much superior to that of tube with −90° inclination angle in the low vapor quality region where the heat transfer coefficient, h, for +90° inclined tube is about 39.9% more than that of the −90° inclined tube. However, the average heat transfer coefficient, h, for the tube with +90° inclination is 30% more than that for the tube with −90° inclination angle for the entire ranges of vapor qualities. For a given mass velocity, the average heat transfer coefficient, h, is defined as the arithmetic mean of heat transfer coefficients, h, calculated at different vapor qualities. In the low vapor quality region and at the lowest mass velocity, the heat transfer coefficient, h, for +90° inclined tube is about 62% more than that of the − 90° inclined tube. In comparison to the performance of the horizontal tube, the heat transfer coefficient, h, for the tube having inclination angle of + 90° is 20.2% more at lower vapor quality, while the average heat transfer coefficient, h, is 15.7% more than that for a horizontal tube, which lies within the uncertainty of 9.2% in experimental data. The variation of heat transfer coefficient with vapor quality for a corrugated tube with different inclinations is shown in Figs. 5, 6 and 7 for the mass velocity of 46 kg m−2 s−1, 81 kg m−2 s−1 and 110 kg m−2 s−1 respectively. The tube with α = +90° has still a better heat transfer performance. Moreover, the increases of 36.9%, 17.2%, and 13.4% in heat transfer coefficient at low vapor qualities and increase of 20.8%, 12.8% and 9.1% in average heat transfer of the test evaporator with α = + 90° as compared to the test evaporator with α = − 90° are observed for
mass velocities of 81, 110 and 136 [kg m−2 s−1], respectively. In these figures, it is observed that the change in tube inclination has considerable effect on evaporation heat transfer coefficient. In fact, at low vapor quality, the vapor phase velocity is reduced, resulting in a decrease in interfacial shear stresses and inertia force as well. Therefore, at low vapor quality, the gravitational force has considerable effect on two-phase flow, and this is the major reason for different rates of heat transfer coefficient for different inclinations of corrugated tube. In other words, interfacial shear stress is dominant at high vapor velocity and high vapor quality, while at low vapor qualities, the vapor phase motion is slowed down, and this would result in a decrease in interfacial shear stress and consequently lower inertia forces near the wall. Therefore, the effect of gravity force on the evaporation of R134a flow at low vapor qualities is more considerable as it can be seen in Figs. 4, 5, 6 and 7 for different inclinations of test evaporator. At low velocities where buoyancy is dominant, the effect of orientation is very pronounced. The random variation in heat transfer coefficient, h, is also visible in these figures. This is due to the instability of heat transfer, in which both the phases are flowing together during phase change. In addition, the presence of the different flow regimes inside the tube with different tube inclinations causes these random variations. At low vapor quality, the increase in evaporation heat transfer coefficient in tube with inclination angle of α = +90° in comparison with other inclination angles is probably due to the better flow of liquid inside grooves and over tube inside wall because of favorable direction of gravitational force. Indeed, during upward vertical flow, α = +90°,
Fig. 5. Variation of heat transfer coefficient with vapor quality at refrigerant mass velocity of 46 kg m−2 s−1. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
Fig. 7. Variation of heat transfer coefficient with vapor quality at refrigerant mass velocity of 110 kg m−2 s−1. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
Fig. 6. Variation of heat transfer coefficient with vapor quality at refrigerant mass velocity of 81 kg m−2 s−1. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
M.A. Akhavan-Behabadi, M. Esmailpour / International Communications in Heat and Mass Transfer 55 (2014) 8–14
the buoyancy force and fluid flow are unidirectional and as a result the flow accelerates enhancing convective transport from the heated wall of the tube [16]. During upward vertical flow of vapor, due to high interfacial turbulence, higher heat transfer coefficient is observed in comparison to that for downward vertical flow, α = − 90°, in low vapor qualities. Furthermore, vertical tube does not suffer from the adverse effect of flow stratification like a horizontal tube. In the dryout region, at high vapor quality, upward vertical boiling flow has the lower heat transfer coefficient than that for the horizontal flow. In other words, when dryout happens, the wetted perimeter of the tube is the most effective region for heat transfer. At high vapor quality in horizontal tube (also tubes with inclinations close to horizontal), the liquid tends to flow in lower side of the tube, and as a result, the flow boiling heat transfer in the flooded part of the tube takes place while mist flow in the form of entrained droplets occurs during upward vertical flow [17]. At vapor quality of 0.95 and refrigerant mass velocities of 46 and 81 kg m− 2 s− 1, the heat transfer coefficients of horizontal tube are 26.3% and 11.2% more than those of − 90° inclined tube, respectively. From Figs. 4 to 7, it is also noted that the vertical tube with inclination angle of −90° has the lowest heat transfer coefficient. Wongwises et al. [9] also reported the same observations in their study of orientation effects for R-134a flow boiling in vertical downward flow inside a corrugated tube. In the downward vertical flow, since the interfacial shear stress and the gravity are acting in the same direction, the flow pattern will remain in annular status [12]. The phenomenon of annular flow causes to form a thick layer of liquid film around the periphery of tube as a result the heat transfer coefficient, h, is reduced. Since the gravitational force and vapor shear stress are in same direction, the interfacial turbulence is the lowest [18]. In Fig. 8, the variation of heat-transfer coefficient with vapor quality at different mass velocities of R-134a is shown for the tube having inclination angle of +90°. The experimental data clearly show that the heat transfer coefficients increase with increasing the refrigerant mass velocity. Indeed, increasing the refrigerant mass velocity increase the fluid velocity, thus enhancing convective boiling. The increase in mass velocity changes the flow pattern from stratified flow to annular flow. In convective boiling, when fluid velocity increases, the heat transfer coefficient increase. In higher mass velocity, effects of convection are much higher and more turbulence occurs, as a result heat transfer coefficient increase. With the increase in mass velocity, flow pattern changes from stratified flow to annular flow in a much lower vapor quality, and the increase in vapor quality in annular flow causes more increase in heat transfer relative to stratified flow. With the increase in mass velocity, maximum heat transfer coefficient will occur in much lower vapor quality because with the increase
Fig. 8. Effect of mass velocity on heat transfer coefficient at inclination angle of +90°. (For interpretation of the references to colour in this figure, the reader is referred to the web version of this article.)
13
in mass velocity, dryout will occur sooner. The average heat transfer coefficient, h, at the mass velocity of 136 kg m−2 s− 1 is nearly 128% more than that for the mass velocity of 46 kg m−2 s−1. Finally based on the present experimental results, the following correlation, Eq. (7), has been developed to predict the heat transfer coefficient at different vapor qualities, mass velocities and tube inclinations. −3
Nu ¼ 3:79 10
1:02
Re f
0:11
Fα
0:92 Pr f X tt
ð7Þ
Where, hD GDð1−xÞ ; Re f ¼ ; Pr f kf μf !0:1 0:9 0:5 μ f cpf ρg 1−x μf ¼ ; Xtt ¼ x kf ρf μg
Nu ¼
0:6
F α ¼ 1 þ 0:25ð1 þ xÞ
F α ¼ 1−0:6x
0:97
Sinðα Þ
for x ≤ 0:7
Cos α−10 for x N 0:7
ð8Þ
ð8aÞ
ð8bÞ
The correlation given in Eq. (7) predicts the experimental data for all angles of inclination of corrugated tube in an error band of ± 10% as shown in Fig. 9. 4. Conclusion The following conclusions have been drawn from the present investigation: 1. For all tube inclination angles, the heat transfer coefficient increases with the increase in vapor quality until a vapor quality near 70–85% when it begins to decrease because of dryout. 2. The tube inclination angle, α, affects the boiling heat transfer coefficient in a significant manner. The effect of inclination angle is more prominent at low vapor qualities and mass velocities. 3. At low vapor qualities, the highest heat transfer coefficient is attained at inclination angle of + 90° (vertical upward flow), and at high vapor qualities, the highest heat transfer coefficient occurs when the corrugated tube is horizontal. The vertical tube with inclination angle of −90° has the lowest heat transfer coefficient for the entire
Fig. 9. Comparison of experimental heat transfer coefficient with those predicted by proposed correlation.
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range of vapor quality. The corrugated tube having inclination angle of + 90° outperforms the tube with − 90° inclination in a range of 7–62%. 4. A correlation has been developed, Eq. (7), which predicts the evaporation heat transfer coefficients of R-134a inside a corrugated tube with different tube inclinations within an error range of ±10%. Acknowledgment The authors would like to express their thanks to School of Mechanical Engineering, College of Engineering, University of Tehran, for financial support of the present experimental work. References [1] K.N. Agrawal, H.K. Varma, Experimental study of heat transfer augmentation versus pumping power in a horizontal R-12 evaporator, Int. J. Refrig. 14 (1991) 273–281. [2] M.A. Akhavan-Behabadi, Ravi Kumar, A. Mohammadpour, Effect of twisted tape insert on heat transfer and pressure drop in horizontal evaporator for the flow of R-134a, Int. J. Refrig. 32 (9) (2009) 922–930. [3] S. Wongwises, M. Polsongkram, Evaporation heat transfer and pressure drop of HFC-134a in a helically coiled concentric tube-in-tube heat exchanger, Int. J. Heat Mass Transf. 49 (2005) 658–670. [4] S. Lohalertdecha, A.S. Dalkilic, S. Wongwises, Correlation for evaporation heat transfer coefficient and two-phase friction factor for R-134a flowing through horizontal corrugated tubes, Int. J. Heat Mass Transf. 38 (2011) 1406–1413. [5] S. Lohalertdecha, S. Wongwises, An experimental study into the evaporation heat transfer and flow characteristics of R-134a refrigerant flowing through corrugated tubes, Int. J. Heat Mass Transf. 34 (2011) 280–291.
[6] V.D. Zimparov, N.L. Vulchanov, L.B. Delov, Heat transfer and friction characteristic of spirally corrugated tube for power plants condensors-1. Experimental investigation and performance evaluation, Int. J. Heat Mass Transf. 34 (1991) 2187–2197. [7] T. Targanski, J.T. Cieslinski, Evaporation of R407C/oil mixtures inside corrugated and micro-fin tubes, Appl. Therm. Eng. 27 (2007) 2226–2232. [8] S. Laohalertdecha, S. Wongwises, The effects of corrugation pitch onthe condensation heat transfer coefficient and pressure drop of R-134a inside horizontal corrugated tube, Int. J. Heat Mass Transfer 53 (2010) 2924–2931. [9] K. Aroonrat, S. Wongwises, Evaporation heat transfer and friction characteristics of R-134a flowing downward in a vertical corrugated tube, Exp. Thermal Fluid Sci. 35 (2011) 20–28. [10] J.G. Withers, E.H. Yong, Steam condensation on vertical rows of horizontal corrugated and plain tube, Ind. Eng. Chem. Process Des. Dev. 10 (1) (1971) 19–30. [11] S. Laohalertdecha, S. Wongwises, Condensation heat transfer and flow characteristics of R-134a flowing through corrugated tubes, Int. J Heat Mass Transfer 54 (2011) 2673–2682. [12] D. Khoeini, M.A. Akhavan-Behabadi, A. Saboonchi, Experimental study of condensation heat transfer of R-134a flow in corrugated tubes with different inclination, Int. Commun. Heat and Mass Transfer (2011) 1001–1006. [13] American society of heating refrigerating and air-conditioning engineers handbook, Fundamentals, 2005. 17 (Chapter 20). [14] R.R. Schultz, R. Cole, Uncertainty analysis in boiling nucleation, Am. Inst. Chem. Eng. Symp. Ser. 75 (189) (1979) 32–38. [15] A. Greco, Convective boiling of pure and mixed refrigerants, Int. J. Heat Mass Transfer 51 (2008) 896–909. [16] M.A. Akhavan-Behabadi, S.G. Mohseni, S.M. Razavinasab, Evaporation heat transfer of R-134a inside a microfin tube with different tube inclinations, Exp. Thermal Fluid Sci. 35 (2010) 996–1001. [17] J.G. Collier, J.R. Thome, Convective Boiling and Condensation, 3rd edition Oxford University Press, Oxford, 1994. 13–15. [18] M.A. Akhavan-Behabadi, Ravi Kumar, S.G. Mohseni, Condensation heat transfer of R134a inside a microfin tube with different tube inclinations, Int. J. Heat Mass Transfer 50 (2007) 4864–4871