Experimental analysis of the effect of cold fluid inlet temperature on the thermal performance of a

Abstract

Keywords: Heat exchanger; louvered fin; mini channel; cold fluid

Introduction

One of the important application in compact heat exchanger design is the extended surfaces with multi-louvered fins. At first it is commonly used in the automotive industry to reduce the weight and the volume of the heat exchangers. Nowadays, louvered fin geometries are extensively used in the area of electronic devices, charge air coolers, evaporator and condensers to reduce the weight and size, also to save energy. Louvered fins provide more surface area relatively to the plain fins. The louvers create a series of thin boundary layer and interrupt the air flow. Therefore, the air side thermal resistance of the heat exchanger reduces and overall thermal performance 583

Research Article friction factor. In addition, correlations for Colburn j-factor and Fanning friction factor f were developed for the considered geometries. Li and Wang [4] performed an experimental study on the air-side heat transfer and pressure drop characteristics of the heat exchangers with multi-louvered fins and flat tubes. Experiments were conducted for heat exchangers with different numbers of louver regions at the air-side Reynolds numbers of 400–1600. The air-side thermal performance data were analyzed by using the effectiveness-NTU method. Colburn-j factor and Fanning friction factor f were presented as a function of Reynolds number. It is found that the j/f1/3 ratio decreased with the increasing of Reynolds numbers and increased with the increasing number of louver regions.

decreased louver angle patterns by numerically. 3D numerical analysis of the heat transfer and the fluid flow were carried out. The results indicated that the successively variable louver angle patterns could effectively enhance the heat transfer performance. Malapure et al. [14] performed three-dimensional simulations of a single and double row tubes with louvered fins. Effects of the louver pitch, louver angle, fin pitch, tube pitch, and Reynolds numbers on the thermal-hydraulic performance of the air side were investigated. The computed Stanton numbers and friction factors were found to be in good agreement with the experimental data at low Reynolds number. In addition, both the Stanton number and the friction factor increased with the decrease in fin pitch.

Lyman et al. [5] studied on the large-scale louver models with varied fin pitch and louver angle experimentally. A method was presented for evaluating the heat transfer coefficients using various reference temperatures to define the convective heat transfer coefficients. The results showed that the thermal field surrounding a particular louver is the major effect on the heat transfer from that louver. Park and Jacobi [6] studied the airside thermal-hydraulic performance of flat-tube aluminum heat exchangers experimentally. The heat transfer and pressure drop were measured at frontal air velocities from 0.5 m/s to 2.8 m/s for dry and wet surface conditions. Parametric effects on the heat transfer and the friction factor were investigated for both dry and wet conditions. It was found that the louver spacing was a significant design parameter under wet conditions. Park and Jacobi [7] developed an air-side data analysis method for a flat-tube louvered-fin heat exchangers under partially wet conditions. Park and Jacobi [8-9] generated correlations for the Colburn-j and friction factor by using the largest database in the literature for the flat-tube louvered-fin heat exchangers.

Perrotin and Clodic [15] presented the results of 2D and 3D CFD models of compact louvered heat exchangers for the determination of heat transfer and pressure drop characteristics. They compared the 2D and 3D steady simulations with the experimental results and the correlations of the literature. It was found that the difference between the heat transfer coefficients obtained from the 2D simulations and the experimental data is up to 80%. In addition, the heat transfer coefficient calculated with the 3D models was much closer to the experimental data. Tafti and Cui [16] performed three-dimensional simulations of the louver–tube junction geometries to investigate the effect on the friction and the heat transfer characteristics. Three Reynolds numbers based on the bulk velocity and louver pitch were calculated. According to the three-dimensional results the flow acceleration had a large impact on louver heat transfer locally. Comparisons with correlations derived from experiments showed that the computational modeling of a small subsystem can be used reliably to extract the performance data for the full heat exchanger.

Qi et al. [10] focused on the geometrical factors of the louvered fins including flow depth, ratio of fin pitch and fin thickness, tube pitch, number of louvers and angle of louver. Fifteen samples were used from the experimental data to analyze the heat transfer and the fluid flow characteristics by using the Taguchi method. The results showed that the contribution ratios to the overall performance of the flow depth, the ratio of the fin pitch and fin thickness and the number of the louvers are 31.57%, 21.53% and 20.34%, respectively. Chang and Wang [11] developed a generalized heat transfer correlation for louvered-fin geometry. This data bank consisted of 91 samples of louvered fin heat exchangers with different geometrical parameters. It was shown that 89.3% of the corrugated louver fin data are correlated within ± 15% with a mean deviation of 7.55%. Atkinson et al. [12] performed a detailed evaluation of 2D and 3D numerical simulations of flow and heat transfer over the louvered fins. Two 2D models were used, both of which incorporate the effects of tube surface area and fin resistance on the overall heat transfer rate. It was found that all the models gave accurate predictions of the pressure losses, but only 3D models are in good agreement with the experimental observations in terms of overall heat transfer. Hsieh and Jang [13] investigated the successively increased or

Uğurlubilek et al. [17-18] investigated the effect of louver angle on the heat transfer and the pressure drop characteristics of mini channel flat-tube with louvered fin heat exchanger. Numerical simulations were performed for different louver angles at constant wall temperature boundary condition. The result showed that the pressure drop increases with the increasing of the louver angle which create more resistance to the flow. The pressure drop took its lowest value on which the geometry has the smallest louver angle of 20°. The rate of heat transfer took its highest value for the geometry having a louver angle of 26°. Furthermore, it was also seen that the relation between the louver angle and the heat transfer was not linear. Akyüz [19] conducted a numerical investigation for the effects of fin pitch and fin height on the thermo-hydraulic performance of an air-cooled, flat-tube heat exchanger. The thermohydraulic performance of the heat exchanger was evaluated using the performance factor j/f1/3. Fin pitches of 1.50, 2.00 and

Experimental Setup

Test Apparatus In this study, the thermal performances of two mini channel flat-tube heat exchangers with louvered fins have been studied experimentally at three different (15°C, 24°, 33°C) inlet temperature of the cold fluid (air). A plexi-glass wind-tunnel which has a 1m×1m cross-section and 2m long is located in an insulated test room as shown in Fig. 1. The dimensions of constant temperature room are 4.5m×3.5m×2.3m. The inlet condition of the air-side of the heat exchanger is maintained constant by controlling the temperature of the test room. The steady-state values are used to calculate the thermal performance of the heat exchangers. The experimental data are collected with a data-acquisition system. The experiments are repeated three times to get the average results. The temperature control of the test room is obtained by using an electric resistance heater and a refrigeration system using R-22 on the top of the perforated chrome ceiling. The wind-tunnel system is designed to suck the test room air over the louvered fin side of the heat exchangers by a centrifugal fan as shown in Fig. 1. The heat exchanger height is less than that of the tunnel inlet dimensions. Therefore, the bypass flow is eliminated by a thin layer of foam. Firstly, room air is sucked by the fan and air flow is forced to pass through the louvered fins.

Table 1. Summary of the uncertainty analysis Parameters Uncertainty Air inlet temperature 0.4% Air outlet temperature 0.4% Mass flow rate of air 0.2% Water inlet temperature 0.4% Water inlet temperature 0.4% Mass flow rate of water 0.15% The Types of Test Heat Exchanger In this study, two mini channel flat-tube heat exchangers with multi louvered fins are tested. Heat exchangers have identical frontal area of 160mm×160mm. As shown in Fig. 2 and Fig. 3, test heat exchangers are called Type-I and Type-II. Mini channel flat-tube is serpentine shaped and Type-I and Type-II have 9 and 7 tube passes, respectively. Type-I has one intermediate plate and two rows of louvered fins between the serpentine flat tubes. Type-II has two intermediate plates and three rows of louvered fins between the serpentine flat tubes. Table 2. Geometric properties of the test heat exchangers Property Type-I Type-II

The inlet and exit temperatures across the air side of the heat exchangers are measured by T-type thermocouple grid. Both the inlet and the outlet temperature grids consist of four thermocouples in an evenly spaced array. Each thermocouple value is recorded with a data-acquisition system, and their average values are used as the cold fluid inlet and outlet temperature. After the air flow passes through the tested heat exchangers, it passes through a three layered screen set, nozzle set and again a three layered screen set, respectively. The screen sets are used to get uniform flow at the inlet and the exit of the nozzle set. At the inlet and the exit of the nozzle set, the pressure drop of air is measured for each surface of the windtunnel by digital manometers, and their average values are recorded. The mass flow rate of the air is measured in terms of the pressure drop across the nozzle set according to ASHRAE Standard 41.2. In this work, the pressure drop at the nozzle is observed as 20 Pa. The mass flow rate of air is 0.047 kg/s according to this pressure drop value. The mass flow rate and the inlet temperature of hot fluid (water) are regulated by a water circulator. The water is heated up to the temperature of 42ᵒC and pumped to the mini channel flat-tube heat exchanger. The mass flow rate of the water is 0.025 kg/s. The water inlet

Fh [mm] Fp [mm] Lh [mm L[o] Lp [mm] Fd [mm] Pt [mm] Tp [mm] a [mm] b [mm] A [m2] Number of tube pass [-] Number of fin row [-] Total length of flat-tube [mm]

4. In Fig. 4, cross- section of A-A shows that the louvered fin

geometry is the same for both Type-I and Type-II. However, the air side heat transfer areas are different, due to the louvered fin 585

Research Article configuration between the serpentine flat tubes. Table 2 shows the similarities and differences of the heat exchangers in terms of geometrical properties. The notations used by Kays and

London [21] are followed throughout the figures, tables and calculations.

Fig. 1. Constant temperature test room, wind-tunnel and test apparatus

Fig. 4. Definition of the geometrical terminology of the heat exchangers (a) Type-I (b) Type-II (c) Cross-sectional view of A-A for both models

Variation of the temperature is taken into consideration on the cold side and the specific heat of the air is calculated as follows [22].

Data reduction Average heat transfer rate from the heat exchanger is given in Eq. 1. Q

 Tc,in  Tc,out   8.31447     1.337  10 3 c p,c    3.653    2  28.97    

 Tc,in  Tc,out  T T   3.294  10 6   c,in c,out   1.913  10 9      2 2     4   Tc,in  Tc,out   0.2763  10 12    (4)   2   

The heat transfer rate to the cold fluid can be calculated from the temperature increase at the air side as shown in Eq. 2. Similarly, heat transfer from the hot fluid can be calculated from the temperature decrease at the water side as shown in Eq. 3.

The overall thermal conductance of the heat exchangers is calculated as follows.

Q (5) Tm where ΔTm is the logarithmic mean temperature difference given as; UA 

where Cc and Ch are the heat capacities of the air and the water as Cc=mccp,c and Ch=mhcp,h, respectively. Specific heat of the water is assumed as 4.178 kJ/kg°C, due to the small temperature difference at the hot side. 587

effect can cause misinterpretations. The overall thermal conductance first increases with the increasing of the inlet temperature of the cold fluid from 15oC to 24oC, then decreases with the increasing of the inlet temperature of the cold fluid from 24oC to 33oC for both heat exchangers. Due to the combined effects of the geometrical and the operational parameters, such results are frequently observed also in the literature [6-10].

Number of transfer units (NTU) is calculated according to the following equation. NTU 

Table 3. Experimental results Tc,i Tc,o Test [ᵒC] [ᵒC] 1 14.34 22.18 2 13.79 21.75 3 15.56 22.97 Avg. 14.56 22.30

The effectiveness method can be used to compare the heat exchangers by using the equations for the unmixed fluid [23],

Tc,i [ᵒC] 14.83 14.66 15.00 14.83 24.11 23.55 23.73 23.80 32.66 32.83 32.81 32.76

Tc,o [ᵒC] 20.60 20.50 20.84 20.65 27.67 27.24 27.32 27.41 34.37 34.45 34.44 34.42

Th,i [ᵒC] 42.30 42.28 42.29 42.29 42.79 42.83 42.83 42.82 42.39 42.41 42.41 42.40

Th,o [ᵒC] 39.32 39.25 39.22 39.26 40.31 40.16 39.93 40.14 41.03 41.05 41.03 41.03

The exit temperatures of the cold and the hot fluids are measured for three different cold side inlet temperatures. The measurements are given in Table 3. The two types are compared by using the most important parameters of the heat exchanger; namely the overall thermal conductance, the number transfer units, and the effectiveness. First the logarithmic mean temperature difference ΔTm, and the average heat transfer rate Q are calculated by using Eq. 5 and Eq. 1, respectively. They are presented in Table 4. Then the overall thermal conductance UA and the number of transfer units NTU are obtained via the Eq. 5 and Eq. 7, respectively. Table 4 includes these parameters as well as the effectiveness ε, obtained by Eq. 8 for both types calculated at three different inlet conditions.

Results And Discussion

The variation of the thermal conductance of both types can be observed in Fig. 5. At 24oC condition, the highest overall thermal conductance of 18.05 W/oC is obtained for Type-I. At the same time, it is indicated that, the overall thermal conductance of Type-I is higher at any inlet temperature of the air. The heat transfer area designates the physical size of a heat exchangers. Although the outer frames of the heat exchangers are the same, the configurations provide different heat transfer areas.

Fig. 6 is prepared for the comparison of NTU values of two types. NTU indicates the thermal size of the heat exchanger and provides a compound measure of the total heat transfer area A, the overall heat transfer coefficient U and the minimum heat capacity rate Cmin. Since U is not constant, the definition of NTU should be considered as;

As it is seen in Table 2, Type-I and Type-II specimens have the total heat transfer area of 0.383 m 2 and 0.316 m2, respectively. In the definition of the overall thermal conductance, the area A is included. It can understand that the greater heat transfer area provides greater overall thermal conductance, but an assumption of a monotonic parametric

In this study, Cmin is approximately constant. Therefore, the variations of NTU have similar trends with respect to the UA curves. Type-I has higher NTU than Type-II at any inlet temperature of the air as seen in Fig. 6. 588

Table 4. System parameters Tc,i ΔTm Test [ᵒC] [ᵒC] 1 14.34 22.92 2 13.79 23.27 3 15.56 22.12 Avg. 14.56 22.77 1 24.22 13.87 2 24.46 13.66 3 24.70 13.35 Avg. 24.46 13.62 1 31.74 8.03 2 32.73 7.30 3 32.59 7.46 Avg. 32.35 7.60 Tc,i ΔTm Test [ᵒC] [ᵒC] 1 14.83 23.06 2 14.66 23.15 3 15.00 22.81 Avg. 14.83 23.01 1 24.11 15.65 2 23.55 16.10 3 23.73 15.85 Avg. 23.80 15.87 1 32.66 8.20 2 32.83 8.09 3 32.81 8.09 Avg. 32.76 8.13

Q [W] 341.73 353.51 330.97 342.07 239.72 246.38 251.29 245.80 115.71 105.07 107.62 109.47 Q [W] 292.22 297.12 298.42 295.92 213.99 227.13 236.29 225.80 111.85 109.57 110.75 110.72

Qmax [W] 1346.62 1369.81 1297.85 1338.09 837.58 821.92 809.10 822.86 467.01 425.30 442.37 444.89 Qmax [W] 1301.58 1308.66 1293.05 1301.09 885.38 913.79 905.26 901.48 461.33 454.23 455.17 456.91

UA [W/ᵒC] 14.91 15.19 14.95 15.02 17.28 18.04 18.83 18.05 14.39 14.36 14.41 14.38 UA [W/ᵒC] 12.67 12.82 13.08 12.86 13.64 14.10 14.89 14.21 13.61 13.51 13.68 13.60

Ch [W/ᵒC] 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50 Cmax [W/ᵒC] 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50 104.50

Cc [W/ᵒC] 47.383 47.382 47.384 47.383 47.401 47.400 47.399 47.400 47.412 47.414 47.414 47.414 Cmin [W/ᵒC] 47.382 47.381 47.382 47.382 47.397 47.396 47.396 47.396 47.413 47.414 47.414 47.414

Fig. 6. Variation of the number of transfer units with respect to the inlet temperature of air

Fig. 5. Variation of the overall thermal conductance with respect to the inlet temperature of air

Research Article The effectiveness of a heat exchanger is important to learn the ratio of the actual heat transfer rate to the possible maximum heat transfer rate which depends on the overall effect of the operating conditions and design parameters. Fig. 7 presents the effectiveness changing with respect to the inlet temperature of the air. The maximum possible heat transfer changes with the inlet conditions although the flow rates stay constant. In this study, the actual heat transfer rate is considered as the average of the heat transfer rates of both the cold and hot sides.

temperatures. The normalized number of transfer unit values are parallel to that of the thermal conductance due to the definition. As a design parameter NTU values also indicate the Type-I having larger heat transfer area. Note that the perfect heat exchanger requires higher effectiveness. When the  values are compared, the Type-I becomes prominent in all cases.

Conclusion

The effect of the cold fluid inlet temperature is investigated experimentally using two different mini channel flat-tube heat exchangers with multi-louvered fins. Even though the heat exchangers have identical size, due to the different louvered fin row configuration they end up with different surface areas. The thermal performance of the heat exchangers is compared by using both LMTD and effectiveness-NTU method.

Fig. 7. Variation of effectiveness with respect to the inlet temperature of air Additionally, the heat transfer rate inevitably includes the inlet temperature of the cold fluid. The actual and the maximum heat transfer rates under these conditions give rise to the change in the effectiveness seen in Fig. 7. The final evaluation of the effectiveness of two types is that the heat exchanger of the Type-I has higher effectiveness with respect to the effectiveness of Type-II due to the change of the geometry.

The heat exchanger named Type-I has higher thermal performance in terms of the overall thermal conductance, the number of transfer units and the effectiveness at any inlet temperature of the air. This may be a consequence of the total length of the serpentine flat-tube, despite the number of fin rows of the Type-I has less than the Type-II. Especially, at an inlet temperature of 15oC and 24oC, the thermal performance of the Type-I is greater than the Type-II. The difference is about 10% and 20% at an inlet temperature of 15oC and 24oC, respectively. Therefore, it is recommended from heat transfer performance point of view. At an inlet temperature of 33oC, the difference between the thermal performances of the heat exchangers decreases about to 4%. Two types should be evaluated by economic constraints at higher inlet temperatures of air. Since the number of fins, the length of the flat tubes and the number of intermediate plates between the fin rows are different, the cost of both types will be different.

To get a clear comparison between the heat exchangers, it is quite valuable to use the normalized values of parameters like the overall thermal conductance ( UA ), number of transfer units ( NTU ), or effectiveness (  ) for different structures of the heat exchangers from the designer’s point of view. Normalized values ( UA, NTU ,  ) are obtained such that the average value of the selected parameter divided by the maximum value of the related results of both types. They are presented in Fig. 8.

Acknowledgment

The highest normalized thermal conductance is obtained in the Type-I for the inlet temperature of 24oC of the air. At any temperature of the cold fluid, the difference between the types is remarkable and the Type-I is advantageous. The difference between the UA values of both types at the higher inlet temperature is less with respect to the differences at lowest inlet

The research leading to these results has been performed under Santez project number of 00865-STZ.2011-1. The authors would like to thank the Ministry of Science, Industry and Technology and RD Department of Arçelik A.Ş. Eskişehir Refrigerator Plant. 590

Nomenclature

2D 3D A a b Cc Ch Cr cp,c cp,h Fd Fh Fp hc Lh Lp Lα mc mh NTU P Pt Q Qc Qh t Tc,in Tc,out Td Th,in Th,out Tp ΔTm UA

[5] Lyman A.C., Stephan R.A., Thole K.A., Zang L.W. and Memory S. B., 2002, Scaling of heat transfer coefficients along louvered fins, Experimental and Thermal Fluid Science 26, 547-563. [6] Park Y. and Jacobi A. M., 2009, The air-side thermalhydraulic performance of flat-tube heat exchangers with louvered, wavy, and plain fins under dry and wet conditions, J. of Heat Transfer 131, 061801-1-13. [7] Park Y. and Jacobi A. M., 2011, A simple air-side data analysis method for partially wet flat-tube heat exchangers, Heat Transfer Engineering 32:2, 133-140. [8] Park Y. and Jacobi A. M., 2009, Air-side heat transfer and friction correlations for flat-tube louver fin heat exchangers, J. of Heat Transfer 131, 021801-1-12. [9] Park Y. and Jacobi A. M., 2001, Air-side performance characteristics of round- and flat-tube heat exchangers: A literature review, analysis and comparison, Air Conditioning and Refrigeration Center ACRCCR-36. [10] Qui, Z., Chen, J., and Chen Z., 2006, Parametric study on the performance of a heat exchanger with a corrugated louvered fins, Applied Thermal Engineering 27, 539-544. [11] Chang Y., Wang C., 1997, A generalized heat transfer correlation for louvered fin geometry, Int. J. of Heat and Mass Transfer 40:3, 533-44. [12] Atkinson K.N., Drakulic R., Heikal M.R. and Cowell T.A., 1998, Two-and three-dimensional numerical models of flow and heat transfer over louvered fin arrays in compact heat exchangers, Int. J. of Heat and Mass Transfer 41, 4063-4080. [13] Hsieh C. T. and Jang J. Y., 2006, 3-D thermal-hydraulic analysis for louver fin heat exchangers with variable louver angle, Applied Thermal Engineering 26, 1629– 1639. [14] Malapure V. P., Mitra S. K. and Bhattracharya A., 2007, Numerical investigation of fluid flow and heat transfer over louvered fins in compact heat exchanger, Int. J. of Heat and Mass Transfer 46, 199–211. [15] Perrotin T.D. and Clodic D., 2004, Thermal-hydraulic CFD study in louvered fin-and-flat-tube heat exchangers, Int. J. of Refrigeration 27, 422–432. [16] Tafti D.K. and Cui J., 2003, Fin-tube junction effects on flow and heat transfer in flat tube multi-louvered heat exchangers, Int. J. of Heat and Mass Transfer 46, 2027– 2038. [17] Uğurlubilek N., Erbay L.B. and Doğan B., 2013, Numerical investigation of the pressure drop characteristics in a heat exchanger, Proceedings of the 19. ULIBTK, pp.386-391, 9-12 September, Samsun, Turkey. [18] Uğurlubilek N., Erbay L.B. and Doğan B., 2013, Numerical investigation of the heat transfer characteristics in a heat exchanger, Proceedings of the 19. ULIBTK, pp.380-385, 9-12 September, Samsun, Turkey. [19] S. Akyüz, Investigation of effect of fin height and fin pitch to performance on air cooled mini micro channel condensers, Master’s thesis, Institute of Science, University of Eskişehir Osmangazi, Turkey, 2013.

two-dimensional three-dimensional air side heat transfer area, m2 tube clearance for Type-I, mm tube clearance for Type-II, mm heat capacity rate of cold fluid, W/oC heat capacity rate of hot fluid, W/oC heat capacity ratio specific heat of cold fluid, J/(kg oC) specific heat of hot fluid, J/(kg oC) flow depth, mm fin height, mm fin pitch, mm heat transfer coefficient, W/(m2oC) louver height, mm louver pitch, mm louver angle, o mass flow rate of cold fluid, kg/s mass flow rate of hot fluid, kg/s number of transfer unit Pressure, Pa intermediate plate thickness, mm average heat transfer rate, W cold fluid heat transfer rate, W hot fluid heat transfer rate, W fin thickness, mm inlet temperature of cold fluid, oC outlet temperature of cold fluid, oC tube depth, mm inlet temperature of hot fluid, oC outlet temperature of hot fluid, oC tube pitch, mm logarithmic mean temperature difference, oC Overall thermal conductance, W/ oC

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