Heat Transfer Characteristics and Performance of a Spirally Coiled Heat Exchanger under Sensible Cooling Conditions

810 Heat Transfer Characteristics and Performance of a Spirally Coiled Heat Exchanger under Sensible Cooling Conditions Somchai WONGWISES and Paisarn NAPHON In the present study, new experimental data on the heat transfer characteristics and the performance of a spirally coiled heat exchanger under sensible cooling conditions is presented. The spiral-coil heat exchanger consists of a steel shell and a spirally coiled tube unit. The spiral-coil unit consists of six layers of concentric spirally coiled tubes. Each tube is fabricated by bending a 9.27 mm diameter straight copper tube into a spiral-coil of five turns. The innermost and outermost diameters of each spiral-coil are 67.7 and 227.6 mm, respectively. Air and water are used as working fluids in shell side and tube side, respectively. A mathematical model based on the conservation of energy is developed to determine the heat transfer characteristics. There is a reasonable agreement between the results obtained from the experiment and those obtained from the model and a good agreement for the high air mass flow rate region. The results obtained from the parametric study are also discussed. Key Words: Heat Transfer Characteristics, Spirally Coiled Heat Exchanger, Effectiveness 1. Introduction Due to their high heat transfer coefficient and smaller space requirement compared with straight tubes, curvedtubes are the most widely used tubes in several heat transfer applications, for example, heat recovery processes, air conditioning and refrigeration systems, chemical reactors, and food and dairy processes. Helical and spiral coils are well known types of curved-tubes which have been used in a wide variety of applications. However, most studies for curved-tubes have been carried out with the helically coiled tube. Dravid et al. (1) numerically investigated the effect of secondary flow on the laminar flow heat transfer in helically coiled tubes. Numerical results were compared with measured data in the range in which they overlapped. Kalb et al. (2) concentrated on the heat transfer to gases flowing in a uniform wall-temperature helical coil at the entrance-region with transition from turbulent to laminar flow. The second paper by Kalb et al. (3) studied the fully developed force-convection heat transfer for viscous flow of a constant-properties Newtonian fluid in curved circular tubes with uniform wall-temperature. Received 31st May, 2004 (No. 04-5084) Fluid Mechanics, Thermal Engineering and Multiphase Flow Research Lab. (FUTURE), Department of Mechanical Engineering, King Mongkut s University of Technology Thonburi, Bangmod, Bangkok 10140, Thailand. E-mail: somchai.won@kmutt.ac.th Series B, Vol. 48, No. 4, 2005 Rahul et al. (4) obtained experimental results for estimating the heat transfer coefficient for coiled tube surfaces in cross-flow air. The effects of the Reynolds number and pitch of the coiled tube on the heat transfer coefficient were investigated. Xin et al. (5) studied the effects of the Prandtl numbers and geometric parameters on the local and average convective heat transfer characteristics in helical-pipes. The experiments were performed with three different working fluids: air, water, and ethylene-glycol. Xin et al. (6) experimentally investigated the single-phase and two-phase flow pressure drop in annular helicoidal pipes. Kang (7) studied the condensation heat transfer of R-134a flowing through a long helicoidal pipe. Ju et al. (8) investigated the performance of small bending radius helical-coil pipe. The formulas for the Reynolds number of single-phase flow structure transition, and single-phase and two-phase flow friction factor were obtained. Ali (9) proposed the pressure drop correlations for fluid flows through regular helical coil tubes. Generalized pressure drop correlations were developed in terms of the Euler number, Reynolds number, and the geometrical group. Guo (10) investigated the frictional pressure drops of single-phase water and steam-water two-phase flows in helical coils. Two helically coiled tubes were employed as test sections and their four different helix axial inclinations were examined. The heat transfer characteristics in spiral-coil heat exchangers has received comparatively little attention in lit- JSME International Journal

811 erature. The most productive studies have been continuously carried out by Ho et al. (11) (13). The relevant correlations of the tube-side and air-side heat transfer coefficients reported in literature were used in the simulation to determine the thermal performance of the spiral-coil heat exchanger under cooling and dehumidifying conditions. The mathematical model was validated by comparing with the experimental data. Due to the lack of the heat transfer coefficients correlations obtained directly from the spirally coiled configuration, Naphon and Wongwises (14) proposed a correlation for the average in-tube heat transfer coefficient for a spirally coiled heat exchanger under dehumidifying conditions. Recently, their second and third paper of Naphon and Wongwises (15), (16), the mathematical model was proposed to determine the performance and heat transfer characteristics of spirally coiled finned tube heat exchangers under wet-surface and dry-surface conditions, respectively. The numerical results were compared with very few data showed in papers published in open literature. As described above, information in open literature on spirally coiled heat exchangers is still limited, especially, experimental data. The objective of this paper is to study, both experimentally and theoretically, the heat transfer characteristics and performance of a spirally coiled heat exchanger under sensible cooling conditions. Experiments are conducted to obtain the heat transfer characteristics and performance for verifying the mathematical model. The effects of various relevant parameters on the model prediction are also investigated by comparing with the existing measured data. Nomenclatures A :area C p : specific heat, kj/(kg K) d : diameter of tube, m De : Dean number, Re i (d i /d c ) 0.5 G max : mass flux based on minimum free flow area, kg/m 2 s h : heat transfer coefficient, W/m 2 K J :Colburnjfactor j : number of segments k : thermal conductivity, W/mK M : mass flow rate per coil, kg/s M : mass flow rate per coil per length, kg/sm m : total mass flow rate, kg/s n : number of turns of coil Nu : Nusselt number Pr : Prandtl number Q : heat transfer rate, W r : radius of tube, m R min : minimum coil radius, m Re : Reynolds number Re i : Reynolds number based on inner diameter of tube Re o : Reynolds number based on outer diameter of tube R n : average curvature radius of coil, m t : tube thickness, m T : temperature, C α : radius change, m/radian θ : angular position, radian ε :effectiveness Subscripts a :air ave :average c :coil i :inside in :inlet max : maximum min : minimum o : outside out : outlet s : surface tot :total w : water 2. Experimental Apparatus and Method A schematic diagram of the experimental apparatus is shown in Fig. 1. The test loop consists of a test section, refrigerant loop, chilled water loop, hot air loop and data acquisition system. Water and air are used as working fluids. The test section is a spirally coiled heat exchanger consisting of a shell and spirally coiled tube unit as shown in Fig. 3. The test section and the connections of the piping system are designed such that parts can be changed or repaired easily. The open-loop wind tunnel is fabricated from zinc, with an inner diameter of 300 mm and a length of 12 m and is well-insulated by an insulator with a thickness of 6.4 mm. Air flow in the open wind tunnel is discharged by a centrifugal blower into the channel and is passed through a straightener, heater, guide vane, test section, and then discharged to the atmosphere. Air velocity is measured by a hot wire anemometer and controlled through the use of a variable speed drive on the blower motor. Hot air flows into the center core and then flows across the spiral coils, radially outwards to the inner wall of the shell before leaving the heat exchanger at the air outlet section as shown in Fig. 2. The inlet and outlet-air temperatures are measured by eight type-t copper-constantan thermocouples. Humidity transmitters are employed to measure the inlet and outlet-air relative humidities. The close-loop of chilled water consists of a 0.3 m 3 storage tank, an electric heater controlled by adjusting the voltage, a stirrer, and a cooling coil immerged inside a storage tank. R22 is used as the refrigerant for chilling the water. After the temperature of the water is adjusted to achieve the desired level, the chilled water is pumped JSME International Journal Series B, Vol. 48, No. 4, 2005

812 Fig. 1 Schematic diagram of experimental apparatus Table 1 Dimensions of the spirally coiled heat exchanger Fig. 2 Schematic diagram of the spirally coiled heat exchanger Table 2 Experimental conditions Fig. 3 Schematic diagram of the shell and the spirally coiled unit out of the storage tank, and is passed through a filter, flow meter, test section, and returned to the storage tank. The flow rate of the water is controlled by adjusting the valve and measured by a flow meter with a range of 0 10 GPM. The spirally coiled heat exchanger consists of a steel shell with a spirally coiled tube unit. The spiral-coil unit consists of six layers of spirally coiled copper tubes. Each tube is constructed by bending a 9.3 mm diameter straight copper tube into a spiral-coil of five turns. The innermost and outermost diameters of each spiral-coil are 67.7 and 227.6 mm, respectively. The measuring positions are concentrated at the third layer of the spiral-coil unit from the uppermost layer. The thermocouples are installed to measure the chilled water temperatures at the inlet and outlet sections. The temperatures distribution of the chilled wa- Series B, Vol. 48, No. 4, 2005 ter and tube wall are measured in five positions with 1 mm diameter probes extending inside the tube and mounted on the tube wall surface in which the water flows. The dimensions of the heat exchanger are listed in Table 1. Experiments were conducted with various temperatures and flow rates of hot air and chilled water entering the test section. Inlet water temperatures were kept above the dew point of air to avoid condensation. In the experiments, the chilled-water flow rate was increased in small increments while the hot-air flow rate, inlet chilled water and hot-air temperatures were kept constant. The inlet hot air and chilled water temperatures were adjusted to achieve the desired level by using electric heaters controlled by temperature controllers. Before any data were recorded, the system was allowed to approach the steady JSME International Journal

813 Fig. 4 Schematic diagram of simulation approach and control volume of each segment Table 3 Uncertainty of measurement state. The range of experimental conditions in this study and uncertainty of the measurement are given in Tables 2 and 3, respectively. 3. Mathematical Modelling The basic physical equations used to describe the heat transfer characteristics are developed from the conservation equations of energy. The model is based on that of Ho et al. (11), and our previous work (15) with the following assumptions: -Flowsofairandwateraresteady. - There is no heat loss between the system and surrounding. - Air-side convective heat transfer coefficient of each section of a coil turn in horizontal plane is equal. - Tube-side convective heat transfer coefficient of each section of a coil turn in horizontal plane is equal. - Each complete coil turn is circular. - Thermal conductivity of the spirally coiled tube is constant. 3. 1 Air-side heat transfer By considering the control volume of each segment as shown in Fig. 4, the heat transferred from the hot air is determined from dq= M ac p,a R n (dθ)(t a,in T a,out ) (1) where M a is the air mass flow rate per coil per length, C p,a is the specific heat of the air, R n is the mean radius of each turn, dθ is the angular position, T a,in and T a,out are the inlet and outlet air temperatures, respectively. 3. 2 Water-side heat transfer The heat transferred to the chilled water is determined from dq= M w C p,w dt w (2) where C p,w is the specific heat of the water, and M w is the water mass flow rate per coil. The heat transfer rate in terms of overall heat transfer coefficient can be expressed as dq= U o A o R n (dθ)(t a,in T w ) (3) where 1 = 1 + A o + A ot (4) U o h o h i A i ka m where U o is the overall heat transfer coefficient based on the outside surface area, T w is the temperature of the water, h o is the air-side heat transfer coefficient, t is the tube wall thickness, k is the thermal conductivity, and A i is the inside surface area per length of tube for each turn. 3. 3 Energy balance By considering the energy balance over the control volume for each segment, we get M w C p,w (dt w ) = U o A o R n (dθ)(t a,in T w ) (5) where A o is the outside surface area per length of tube for each turn as follows: A o = A o,tot (6) By arranging Eq. (5), we get dt w = ( Ao,tot ) Uo R n (dθ) M w C p,w T a,in + ( Ao,tot ) Uo R n (dθ) M w C p,w T w (7) or ( ) Ao,tot Uo R min (dθ) dt w = R r T a,in M w C p,w ( ) Ao,tot Uo R min (dθ) + R r T w (8) M w C p,w On rearranging, we get dt w xr r T w + xr r T a,in = 0 (9) where x = U ( ) or min Ao,tot [dθ] (10) M w C p,w R r = R [ n = 1+ (2n 1)απ ] (11) R min R min JSME International Journal Series B, Vol. 48, No. 4, 2005

814 The energy balance over the control volume for each segment may be written in terms of the air mass flow rate as follows: U o ( Ao,tot ) (R n dθ) ( T a,in T w ) = M ac p,a (R n dθ) ( ) T a,in T a,out (12) ( ) Ao,tot U o T a,out = M ac T w p,a ( ) Ao,tot U o M ac T a,in +T a,in (13) p,a Substituting M a = M a /, and Eq. (11) into Eq. (13), then gives ( ) Ao,tot T w T a,out = 2π(R rr min )U o M a C p,a 2π(R rr min )U o M a C p,a ( Ao,tot ) T a,in +T a,in (14) or T a,out = N a R r T w N a R r T a,in +T a,in (15) where N a = 2πU ( ) or min Ao,tot (16) M a C p,a 4. Solution Method Each layer of the spirally coiled tube consists of numerous circular coils, with a mean radius as follows: R n = [R min +(2n 1)απ] (17) where R min is the minimum coil radius, n is the number of coil turns, and α is the radius change per radian. Each circular coil turn is divided into numerous segments as shown in Fig. 4. The calculation begins at a segment of the innermost coil turn and then is done segment by segment along the circular coil turn. In order to solve the model, relevant tube-side and air-side heat transfer coefficients are needed. The correlations proposed by Naphon and Wongwises (16) to predict the tube-side and air-sideheat transfer coefficients of the spirally coiled tube are as follows: - For the tube-side heat transfer coefficient Nu i = h id i k = 5.38De0.287 Pr 0.949 (18) for 300 De 2 200, Pr 5 - For the air-side heat transfer coefficient h o J = Pr 2/3 = 0.133Re 0.376 o (19) G max C p,a for Re o < 6 000 In addition, the spirally coiled heat exchanger configurations and properties of working fluids, as well as the operating conditions, are also needed. The iteration process is described as follows: - The outlet water temperature is assumed. Series B, Vol. 48, No. 4, 2005 - Equations (9) and (15) are solved simultaneously to obtain the water temperature, T w, and outlet air temperature, T a,out, at Segment 1. - The heat transfer rate, Q, is calculated. - The computation described above is next performed for Segment 2 and the remaining segments until the last segment. - The same computation is performed at the next circular coil. - The computation is terminated when the calculation at the last segment of the outermost coil turn is finished. - The calculated water temperature at the last segment of the outermost coil turn is compared with the inlet water temperature (initial condition). If the difference is within 10 6, the calculation is ended, and if not, another outletwater temperature value of the first section at the innermost coil turn is tried again and the computations are repeated until convergence is obtained. 5. Results and Discussion An overall energy balance was performed to estimate the extent of any heat losses or gains from the surroundings. In the present study, only the data that satisfied the energy balance conditions, Q w Q a /Q ave is less than 0.05, were used in the analysis. The total rate of heat transfer, Q ave was averaged from the air-side heat transfer rate, Q a, and the water-side heat transfer rate, Q w. Figure 5 shows the variation of the outlet air temperature with air mass flow rate at T w,in = 30.5 C, m w = 0.14 kg/s for the different inlet air temperatures of 55.5 and 60.5 C. For a given water mass flow rate, inlet air temperature and inlet water temperature, the outlet air temperature tends to increase with increasing air mass flow rate. In addition, at the same air mass flow rate, the outlet air temperature at 55.5 C is lower than at 60.5 C across the range of air mass flow rates. Verification of the present numerical simulation is done by comparison with experi- Fig. 5 Variation of the outlet-air temperature with air mass flow rate for different inlet-air temperatures JSME International Journal

815 Fig. 6 Variation of the outlet-air temperature with air mass flow rate for different water mass flow rates Fig. 8 Variation of the outlet-water temperatures with air mass flow rate for different water mass flow rates Fig. 7 Variation of the outlet-water temperatures with air mass flow rate for different inlet-air temperatures mental data. As shown in the figure, at low air mass flow rate region, the model slightly underpredicts the present measured data. The low flow rate of air, together with the temperature which is higher than the ambient air downstream, causes the measured outlet air temperatures to be higher than the calculated ones. However, this deviation diminishes as the air mass flow rate increases. Figures 6 shows the variation of the outlet-air temperature with air mass flow rate at T a,in = 60.5 C, T w,in = 30.5 C for the different water mass flow rates of 0.11 and 0.14 kg/s. In general, the same explanation described above for Fig. 6 can be given. However, the effect of the water mass flow rate on the outlet-air temperature in the present experiment is quite low. Figure 7 shows the variation of the outlet water temperature with air mass flow rate at T w,in = 30.5 C, m w = 0.14 kg/s for the different inlet-air temperatures of 55.5 and 60.5 C. At a specific inlet-air temperature, inlet-water temperature and water mass flow rate, the increase of the outlet-water temperature caused by the increase of the heat transfer rate results in an increase of the outlet-air temper- ature (Fig. 5). As the outlet-air temperature increases, the temperature difference between inlet and outlet-air temperature decreases. Therefore, one way of keeping the heat transfer rate equal to the water side is by increasing the air mass flow rate. Therefore, it can be clearly seen that the outlet-water temperature increases with increasing air mass flow rate. At the same air mass flow rate, the outlet-water temperature at T a,in = 60.5 C tends to be higher than at T a,in = 55.5 C. The reason for this is similar to the one as described above. At a specific inlet-water temperature, water mass flow rate and air mass flow rate, as the outlet-water temperature increases, the heat transfer rate increase. This results in the increase of the outlet air temperature. Therefore, in order to keep the heat transfer rate equal to the water-side heat transfer rate, the inlet-air temperature must be increased. Considering Fig. 8, which shows the effect of water mass flow rate on the outlet- water temperature, it is clear that at the same inlet-air temperature, inlet-water temperature and air mass flow rate, the outlet-water temperature decreases with increasing water mass flow rate. This is because at a specific air mass flow rate, and inlet-air and water temperatures, the outlet-air temperature is almost independent of the water mass flow rate (Fig. 6). In other words, the heat transfer rate absorbed by the chilled water is mainly dependent on the mass flow rate and the outlet temperature of the water i.e. the lower water flow rate therefore gives the higher water- outlet temperature. Again, as shown in this figure, the results obtained from the present model are in reasonable agreement with the experimental data over the whole range of air mass flow rates. Figures 9 and 10 show the variation of the heat transfer rate with air mass flow rate at the same conditions as in Figs. 7 and 8. As expected, the heat transfer rate is directly proportional to the air mass flow rate. In addition, it can be noted that the inlet-air temperature strongly affects the heat transfer rate while the water mass flow rate JSME International Journal Series B, Vol. 48, No. 4, 2005

816 Fig. 9 Variation of the heat transfer rates with air mass flow rate for different inlet-air temperatures Fig. 12 Variation of the heat exchanger effectiveness with air mass flow rate for different water mass flow rates evaluate the performance of the spirally coiled heat exchanger is calculated from Fig. 10 Fig. 11 Variation of the heat transfer rates with air mass flow rate for different water mass flow rates Variation of the heat exchanger effectiveness with air mass flow rate for different inlet air temperatures has an insignificant effect. In general, the proposed model simulated the experimental data fairly well. Figures 11 and 12 show the variation of the heat exchanger effectiveness with air mass flow rate at the same conditions as in Figs. 9 and 10. The effectiveness used to ε = Q ave Q ave = (20) Q max (mc p ) min (T a,in T w,in ) It is found from both figures that the effectiveness decreases with increasing air mass flow rate. This phenomenon can be explained by Eq. (20). From the whole range of water and air mass flow rates, the capacity rate of the chilled water, (mc p ) w, is higher than that of the hot air, (mc p ) a. Therefore, the minimum capacity rate, ( mc p )min, in Eq. (20) is replaced by the capacity rate of the hot air. At a given air mass flow rate, the higher inlet air temperature and the higher water mass flow rate tend to lead to the increase in effectiveness. This is because the heat transfer rate from the hot air to the chilled water increases. The difference between effectiveness at different water mass flow rate is small because the difference of water mass flow rate is also small. It should be noted that if experimental error is accounted for, the difference of measured data at different water mass flow rate is identical with the difference of simulation results. However at low air flow rate, the trend of curve obtained from the calculation is difference from that obtained from the experiment. The same reason as described for Fig. 5 can be given. The mathematical model is shown to fit the experimental data fairly well. A number of graphs can be drawn from the output of the simulation, but because of the space limitation, only typical results are shown. Figures 13 and 14 illustrate the variation of the calculated outlet water temperature with air mass flow rate for various inlet air temperatures and water mass flow rates, respectively. The variation of the outlet-water temperature with air mass flow rate is similar to that shown in Figs. 7 and 8. However, it can be clearly seen from Fig. 13 that as the inlet-air temperature decreases, the curve becomes flatter. As shown in Fig. 13, at any air mass flow rate, the outlet-water temperature increases relatively constantly with constantly increasing Series B, Vol. 48, No. 4, 2005 JSME International Journal

817 inlet-air temperature. However, the increase of the outletwater temperature for the low air mass flow rate region is lower than that for the high air mass flow rate region. As shown in Fig. 14, similar results as in Fig. 13 are obtained while the water mass flow rate increases. It can be seen from this figure that at the same air mass flow rate, the outlet water temperature is inversely proportional to the water mass flow rate. The same explanation as described for Fig. 8 can be given. At a specific water mass flow rate, the outlet-water temperature tends to increase with increasing air mass flow rate. However, the increase of the outlet-water temperature becomes relatively smaller as the water mass flow rate increases. Figure 15 illustrates the effect of inlet-air temperature on the heat transfer rate. It can be clearly seen from the figure that at a specific inlet-air temperature, the heat transfer rate increases with increasing air mass flow rate; however, the increase becomes relatively smaller as the inlet air temperature decreases. In other words, the in- crease of the heat transfer rate at higher inlet-air temperatures is higher than for lower ones at the same range of air mass flow rates. The trends of the heat transfer rate curves are similar to those of the outlet water temperature curves shown in Fig. 13. In Fig. 16 the effect of water mass flow rate on the heat transfer rate can be clearly seen at higher air mass flow rate i.e. the heat transfer rate is much higher for a higher water flow rate than that for a lower water flow rate. However, at very low air mass flow rates, there is almost no effect of water mass flow rate on the heat transfer rate. Figure 17 shows the variation of the effectiveness with air mass flow rate for various water mass flow rates. It should be noted that the effectiveness is inversely proportional to the air mass flow rate. The effectiveness decreases rapidly in the low air mass flow rate region and then decreases moderately as the air mass flow rate increases. For a specific air mass flow rate at constant inletair and water temperatures, the effectiveness increases with increasing water mass flow rate. Again, it can be clearly seen from Eq. (20) that the average heat transfer Fig. 13 Variation of the calculated outlet-water temperature with air mass flow rate for various inlet-air temperatures Fig. 15 Variation of calculated heat transfer rate with air mass flow rate for various inlet-air temperatures Fig. 14 Variation of the calculated outlet-water temperature with air mass flow rate for various water mass flow rates Fig. 16 Variation of the calculated heat transfer rate with air mass flow rate for various water mass flow rates JSME International Journal Series B, Vol. 48, No. 4, 2005

818 in some of the experimental work. References Fig. 17 Variation of calculated effectiveness with air mass flow rate for various water mass flow rates rate increases with increasing water mass flow rate, but the maximum heat transfer rate is kept constant. Therefore, the effectiveness also increases. However, the difference becomes relatively less as the air mass flow rate decreases. 6. Conclusions The heat transfer characteristics and the performance of a spirally coiled heat exchanger under sensible cooling conditions are studied. The results obtained from the experiments are compared with those obtained from the developed model. The effects of the inlet conditions of the working fluids flowing through the spirally coiled heat exchanger are discussed. The following conclusions can be given: - The experimental data at low air mass flow rate is less reliable than those at high air mass flow rate, due to the characteristics of present experimental apparatus. - The agreement between the results obtained from the experiment and those obtained from the model is good agreement for the high air mass flow rate region. - The outlet temperatures of air and water increase with increasing air mass flow rate and inlet air temperature. - The outlet temperatures of air and water increase with decreasing water mass flow rate. - The heat transfer rate depends directly on mass flow rates of air and water, and inlet air temperature. -Theeffectiveness is inversely proportional to the air mass flow rate and directly proportional to the water mass flow rate and inlet air temperature. Acknowledgements The authors would like to express their appreciation to the Thailand Research Fund (TRF) for providing financial support for this study. The authors also wish to acknowledge Miss Supajaree Maroongruang, Mr. Anucha Kasikapast and Mr. Chanit Somphol, for their assistance Series B, Vol. 48, No. 4, 2005 ( 1 ) Dravid, A.N., Smith, K.A., Merrill, E.W. and Brian, P.L.T., Effect of Secondary Fluid on Laminar Flow Heat Transfer in Helically Coiled Tubes, AIChE Journal, Vol.17 (1971), pp.1114 1122. ( 2 ) Kalb, C.E. and Seader, J.D., Heat and Mass Transfer Phenomena for Viscous Flow in Curved Circular Tubes, International Journal of Heat Mass Transfer, Vol.15 (1972), pp.801 817. ( 3 ) Kalb, C.E. and Seader, J.D., Fully Developed Viscous- Flow Heat Transfer in Curved Circular Tubes with Uniform Wall Temperature, AIChE Journal, Vol.20 (1974), pp.340 346. ( 4 ) Rahul, S., Gupta, S.K. and Subbarao, P.M.V., An Experimental Study for Estimating Heat Transfer Coefficient from Coiled Tube Surfaces in Cross-Flow of Air, Proceedings of the Third ISHMT-ASME Heat and Mass Transfer Conference and Fourth National Heat and Mass Transfer Conference, India, December (1997), pp.381 385. ( 5 ) Xin, R.C. and Ebadian, M.A., The Effects of Prandtl Numbers on Local and Average Convective Heat Transfer Characteristics in Helical Pipes, Journal of Heat Transfer, Vol.119 (1997), pp.467 473. ( 6 ) Xin, R.C., Awwad, A., Dong, Z.F. and Ebadian, M.A., An Experimental Study of Single-Phase and Two- Phase Flow Pressure Drop in Annular Helicoidal Pipes, International Journal of Heat and Fluid Flow, Vol.18 (1997), pp.482 488. ( 7 ) Kang, H.J., Lin, C.X. and Ebadian, M.A., Condensation of R-134a Flowing Inside Helicoidal Pipe, International Journal of Heat and Mass Transfer, Vol.43 (2000), pp.2553 2564. ( 8 ) Ju, H., Huang, Z., Xu, Y., Duan, B. and Yu, Y., Hydraulic Performance of Small Bending Radius Helical Coil-Pipe, Journal of Nuclear Science and Technology, Vol.18 (2001), pp.826 831. ( 9 ) Ali, S., Pressure Drop Correlations for Flow through Regular Helical Coil Tubes, Fluid Dynamics Research, Vol.28 (2001), pp.295 310. (10) Guo, L., Feng, Z. and Chen, X., An Experimental Investigation of the Friction Pressure Drop of Steam- Water Two-Phase Flow in Helical Coils, International Journal of Heat and Mass Transfer, Vol.44 (2001), pp.2601 2610. (11) Ho, J.C., Wijeysundera, N.E., Rajasekar, S. and Chandratilleke, T.T., Performance of a Compact Spiral Coil Heat Exchange, Heat Recovery System & CHP, Vol.15 (1995), pp.457 468. (12) Ho, J.C., Wijeysundera, N.E. and Rajasekar, S., Study of a Compact Spiral-Coil Cooling and Dehumidifying Heat Exchanger Unit, Applied Thermal Engineering, Vol.16 (1996), pp.777 790. (13) Ho, J.C., Wijeysundera, N.E. and Rajasekar, S., An Unmixed-Air Flow Model of a Spiral Cooling Dehumidifying Heat Transfer, Applied Thermal Engineering, Vol.19 (1999), pp.865 883. (14) Naphon, P. and Wongwises, S., An Experimental Study JSME International Journal

819 on the in-tube Convective Heat Transfer Coefficients in Spiral-Coil Heat Exchanger, International Communications in Heat and Mass Transfer, Vol.29 (2002), pp.797 809. (15) Naphon, P. and Wongwises, S., Investigation of the Performance of a Spiral-Coil Finned Tube Heat Exchanger under Dehumidifying Conditions, Journal of Engineering Physics and Thermophysics, Vol.76 (2003), pp.71 79. (16) Naphon, P. and Wongwises, S., Experimental and Theoretical Investigation of the Heat Transfer Characteristics and Performance of a Spiral-Coil Heat Exchanger under Dry-Surface Conditions, 2nd International Conference on Heat Transfer, Fluid Mechanics, and Thermodynamics, 24-26 June, Victoria Falls, Zambia, (2003). JSME International Journal Series B, Vol. 48, No. 4, 2005