# INTRODUCTION late heat exchangers (PHEs) were introduced in the 1930s and were almost exclusively used as liquid/liquid heat exchangers in the food industries because of their ease of cleaning. Over the years, the development of the PHE has generally continued towards larger capacity, as well as higher working temperature and pressure. Recently, a gasket sealing was replaced by a brazed material, and each thermal plate was formed with a series of corrugations (herringbone or chevron). These greatly increased the pressure and the temperature capabilities. The corrugated pattern on the thermal plate induces a highly turbulent fluid flow. The high turbulence in the PHE leads to an enhanced heat transfer, to a low fouling rate, and to a reduced heat transfer area. Therefore, PHEs can be used as alternatives to shelland-tube heat exchangers. Due to ozone depletion, the refrigerant R22 is being replaced by R410A (a binary mixture of R32 and R125, mass fraction 50 %/50 %). R410A approximates an azeotropic behavior since it can be regarded as a pure substance because of the negligible temperature gliding. The heat transfer and the pressure drop characteristics in PHEs are related to the hydraulic diameter, the increased heat transfer area, the number of the flow channels, and the profile of the corrugation waviness, such as the inclination angle, the corrugation amplitude, and the corrugation wavelength. These geometric factors influence the separation, the boundary layer, and the vortex or swirl flow generation. However, earlier experimental and numerical works were restricted to a single-phase flow. Since the advent of a Brazed PHE (BPHE) in the 1990s, studies of the condensation and/or evaporation heat transfer have focused on their applications in refrigerating and air conditioning systems, but only a few studies have been done. Much work is needed to understand the features of the two-phase flow in the BPHEs with alternative refrigerants. Xiaoyang et al., [1] experimented with the two-phase flow distribution in stacked PHEs at both vertical upward and downward flow orientations. They Author : Professor, Higher College of Technology, Oman. indicated that non-uniform distributions were found and that the flow distribution was strongly affected by the total inlet flow rate, the vapor quality, the flow channel orientation, and the geometry of the inlet port Holger [2].Theoretically predicted the performance of chevrontype PHEs under single-phase conditions and recommended the correlations for the friction factors and heat transfer coefficients as functions of the corrugation chevron angles. Lee et al., [3] investigated the characteristics of the evaporation heat transfer and pressure drop in BPHEs with R404A and R407C. Kedzierski [4] reported the effect of inclination on the performance of a BPHE using R22 in both the condenser and the evaporator. Several single-phase correlations for heat transfer coefficients and friction factors have been proposed, but few correlations for the two-phase flow have been proposed. Yan et al., [5] suggested a correlation of condensation with a chevron angle of 30 for R134a. Yan et al., reported that the mass flux, the vapor quality, and the condensation pressure affected the heat transfer coefficients and the pressure drops. Hieh and Lin [6] developed the correlations for evaporation with a chevron angle of 30 for R410A. The main objective of this work was to experimentally investigate the heat transfer coefficients and the pressure drops during condensation of R410A inside BPHEs. Three BPHEs with different chevron angles of 45, 35, and 20 were used. The results were then compared to those of R22. The geometric effects of the plate on the heat transfer and the pressure drop were investigated by varying the mass flux, the quality, and the condensation temperature. From the results, the geometric effects, especially the chevron angle, must be considered to develop the correlations for the Nusselt number and the friction factor. Correlations for the Nusselt number and the friction factor with the geometric parameters are suggested in this study. Experiments to measure the condensation heat transfer coefficient and the pressure drop in brazed plate heat exchangers (BPHEs) were performed with the refrigerants R410A and R22. Brazed plate heat exchangers with different chevron angles of 45°, 35°, and 20° were used. Varying the mass flux, the condensation temperature, and the vapor quality of the refrigerant, we measured the condensation heat transfer coefficient and the pressure drops. Both the heat transfer coefficient and the pressure drop increased proportionally with the mass flux and the vapor quality VII ( A ) and inversely with the condensation temperature and the chevron angle. Correlations of the Nusselt number and the friction factor with the geometric parameters are suggested for the tested BPHEs. In an effort to study and optimize the design of a plate heat exchanger comprising of corrugated walls with herringbone design, a CFD code is employed. Due to the difficulties induced by the geometry and flow complexity, an approach through a simplified model was followed as a first step. This simple model, comprised of only one corrugated plate and a flat plate, was constructed and simulated. The Reynolds numbers examined are 400, 900, 1000, 1150, 1250 and 1400. The SST turbulence model was preferred over other flow models for the simulation. The case where hot water (60oC) is in contact with a constant-temperature wall (20oC) was also simulated and the heat transfer rate was calculated. The results for the simplified model, presented in terms of velocity, shear stress and heat transfer coefficients, strongly encourage the simulation of one channel of the typical plate heat exchanger, i.e. the one that comprises of two corrugated plates with herringbone design having their crests nearly in contact. Preliminary results of this latter work, currently in progress, comply with visual observations. In recent years, compact heat exchangers with corrugated plates are being rapidly adopted by food and chemical process industries, replacing conventional shell-and-tube exchangers. Compact heat exchangers consist of plates embossed with some form of corrugated surface pattern, usually the chevron (herringbone) geometry [1].The plates are assembled being abutting, with their corrugations forming narrow passages. This type of equipment offers high thermal effectiveness and close temperature approach, while allowing ease of inspection and cleaning [1], [2]. In order to be able to evaluate its performance, methods to predict the heat transfer coefficient and pressure drop must be developed. In this direction, CFD is considered an efficient tool for momentum and heat transfer rate estimation in this type of heat exchangers. The type of flow in such narrow passages, which is associated with the choice of the most appropriate flow model for CFD simulation, is still an open issue in the literature. Due to the relatively high pressure drop, compared to shell-and-tube heat exchangers for equivalent flow rates, the Reynolds numbers used in this type of equipment must be lower so as the resulting pressure drops would be generally acceptable [1]. Moreover, when this equipment is used as a reflux condenser, the limit imposed by the onset of flooding reduces the maximum Reynolds number to a value less than 2000 [3]. Ciofalo et al. [4], in a comprehensive review article concerning modeling heat transfer in narrow flow passages, state that, for the Reynolds number range of 1,500-3,000, transitional flow is expected, a kind of flow among the most difficult to simulate by conventional turbulence models. On the other hand, Shah & Wanniarachchi [1] declare that, for the Reynolds number range 100-1500, there is evidence that the flow is already turbulent, a statement that is also supported by Vlasogiannis et al. [5], whose experiments in a plate heat exchanger verify that the flow is turbulent for Re>650. Lioumbas et al. [6], who studied experimentally the flow in narrow passages during counter-current gas-liquid flow, suggest that the flow exhibits the basic features of turbulent flow even for the relatively low gas Reynolds numbers tested (500 2 m/s) and water flow rates (jf > 0.025 m/s). Thirdly, the liquid continuous pattern with a gas pocket or a gas bubble at the high water flow rates (jf >0.1 m/s) and low air flow rates (jg < 1 m/s).According to the flow regime map proposed by Vlasogiannis et al., the expected flow pattern in this experimental study is the gas continuous flow pattern with liquid pockets. However, their flow regime map has a significant limitation for use since many important features, such as the phase-change, the heating or cooling conditions, the densities or specific volumes of the working fluids, the geometries of the PHEs, etc., were not considered in detail. According to the flow regime map proposed by Crawford et al. [11], which was developed for vertical downward flow in a round tube, all experimental flow patterns are located in the intermittent flow regime, but this flow regime cannot represent the correct flow regime in a BPHE due to the different geometries. Where ? h local X local is the local heat transfer coefficient at the local vapor quality. The experimental results indicate that the averaged heat transfer coefficients vary proportionally with the mass flux and inversely with the chevron angles and the condensation temperature. The small chevron angle forms narrow pitches to the flow direction, creating more abrupt changes in the velocity and the flow direction, thus increasing the effective contact length and time in a BHPE. The zigzag flow increases the heat transfer, and the turbulence created by the shape of the plate pattern is also important in addition to the turbulence created by the high flow rates. Increasing the mass flux at a given condensation temperature showed that the differences in the averaged heat transfer coefficients were significantly enlarged with decreasing chevron angle. This indicates that a PHE with the small chevron angle is more effective at a large mass flux (Gc > 25 kg/m 2 s) than at a small mass flux. The averaged heat transfer coefficient of R410A decreases with increasing condensation temperature. The vapor velocity is a more influential factor than the liquid film thickness for the heat transfer. Vapor bubbles in the flow enhance the disturbance in the bubble wake as a turbulence promoter, and the turbulence induced by the vapor bubbles increases with the vapor velocity. Also, since the specific volume of the vapor increases with decreasing condensation temperature, the vapor velocity increases for a fixed mass flux and quality. The vapor velocity at 20°C is faster than that at 30°C. The rates of the averaged heat transfer coefficients between condensation temperatures of 20°C and 30°C increased 5 % for a chevron angle of 45°, 9 % for 35°, and 16 % for 20°. These results show that different chevron angles lead partly to different flow pattern. Thus, we may conclude that the flow regime map should be modified by geometric considerations. The heat transfer coefficients in the high-quality region (fast velocity region) are larger than those in the low-quality region (slow velocity region). As mentioned above, this happens because the vapor velocity is the dominant effect on the heat transfer mechanism. Increasing the vapor quality at the same mass flux induces a faster bubble velocity, which increases the turbulence level and the convection heat transfer coefficient. The difference of heat transfer coefficients between the low-quality region and the high-quality region becomes larger with decreasing chevron angle. The PHE with a low chevron angle shows a better heat transfer performance in the high-quality region (i.e., the high vapor velocity region). The frictional pressure loss in a BPHE is obtained by subtracting the acceleration pressure loss, the static head loss, and the port pressure loss from the total pressure loss. kW/m 2 . The frictional pressure drops in the BPHEs increase with increasing mass flux and quality and decreasing condensation temperature and chevron angle. This trend is similar to that of the condensation heat transfer. As mentioned above, since the vapor velocity is much faster than the liquid velocity during the two-phase flow in the tube, the vapor velocity is the dominant influence on the pressure drop, as well as the heat transfer. A high vapor velocity also tends to increase the turbulence of the flow. From Figures 3, 4 The ratios of R410A to R22 for the condensation heat transfer coefficients and pressure drops at a condensation temperature of 30°C are shown in the Figure -7. The ratios for the heat transfer coefficients are relatively constant in the range of 1 -1.1, regardless of the mass flux, while the ratios for the pressure drops decrease with increasing mass flux, except for the data at a chevron angle of 20° in the present experimental range. For a chevron angle of 20°, the heat transfer ratios of R410A to R22 are about 1.1, and the pressure drop ratios about 0.8, which is a 10 % higher heat transfer and a 20 % lower pressure drop.The smaller specific volume of the vapor of R410A relative to that of R22 makes the vapor velocity slower and yields a small pressure drop under the same conditions of the mass flux. While the two fluids have almost equal values of their latent heats, the liquid-phase thermal conductivity of R410A is larger than that of R22. The higher thermal conductivity for R410A helps to produce better heat transfer even if a reduction in the specific volume occurs. Also, a BPHE with a small chevron angle is known to have more effective performance from the ratios when replacing R22 with R410A. Based on the experimental data, the following correlations for Nu and f during condensation for the tested BPHEs are established: Where G e1 , G e2 , G e3 , and G e4 are non-dimensional geometric parameters that involve the corrugation pitch, the equivalent diameter, and the chevron angle. Re Eq is the equivalent Reynolds number, and G Eq the equivalent mass flux: where G c is the channel mass flux. The suggested correlations for the Nusselt number and the friction factor can be applied in the range of Re Eq from 300 to 4000. 23) is 10 %. V. # STUDY OF A SIMPLIFIED GEOMETRY In an effort to simulate the flow configuration, a simple channel was designed and constructed in order to conduct experiments and obtain formation on the flow pattern prevailing inside the furrows of the conduit. The flow configuration, apart from affecting the local momentum and heat transfer rates of a plate heat exchanger, suggests the appropriate flow model for the CFD simulation. A module of a plate heat exchanger is a single pass of the exchanger, consisting of only two plates. The simple channel examined is a single pass made of Plexiglas (Figure 9). It is formed by only one corrugated plate comprised of fourteen equal sized and uniformly spaced corrugations as well as a flat plate and it is used for pressure drop measurements and flow visualization. Details of the plate geometry are presented in Table 2. This model was chosen in an attempt to simplify the complexity of the original plate heat exchanger and to reduce the computational demands. The geometry studied in the CFD simulations (similar to the test section) is shown in Figure 10. The Reynolds numbers examined are 400, 900, 1000, 1150, 1250 and 1400, which are based on the distance between the plates at the entrance (d=10mm), the mean flow velocity and the properties of water at 60oC. In addition to isothermal flow, heat transfer simulations are carried out for the same Reynolds numbers, where hot water (60oC) is cooled in contact with a constanttemperature wall (20oC). The latter case is realized in condensers and evaporators. Additionally, it is assumed that heat is transferred only through the corrugated plate, while the rest of the walls are considered adiabatic. # VII ( A ) 2011 December A commercial CFD code, namely the CFX ® 5.6 code developed by AEA Technology, was employed to explore its potential for computing detailed characteristics of this kind of flow. In general, the models used in CFD codes give reasonably good results for single-phase flow systems. The first step in obtaining a solution is the division of the physical domain into a solution mesh, in which the set of equations is discretised. The grid size used is selected by performing a grid dependence study, since the accuracy of the solution greatly depends on the number and the size of the cells. The resulting mesh was also inspected for inappropriate generated cells (e.g. tetrahedral cells with sharp angles) and fixed, leading to a total number of 870,000 elements. The SST model was employed in the calculations for the reasons explained in the previous chapter. The mean velocity of the liquid phase was applied as boundary condition at the channel entrance (i.e. Dirichlet BC on the inlet velocity) and no slip conditions on the channel walls. A constant temperature boundary condition was applied only on the corrugated wall, whereas the rest of the walls are considered adiabatic. Calculations were performed on a SGI O2 R10000 workstation with a 195MHz processor and 448Mb RAM. The CFX ®5.6 code uses a finite volume method on a non-orthogonal body-fitted multi-block grid. In the present calculations, the SIMPLEC algorithm is used for pressure-velocity coupling and the QUICK scheme for discretisation of the momentum equations [31], [32]. The results of the present study suggest that fluid flow is mainly directed inside the furrows and follows them (Figure 11a). This type of flow behavior is also described by Focke & Knibbe[7], who made visual observations of the flow between two superposed corrugated plates (Figure 11b). They confirm that the fluid, after entering a furrow, mostly follows it until it reaches the side wall, where it is reflected and enters the anti-symmetrical furrow of the plate above, a behavior similar to the one predicted by the CFD simulation. It seems that, in both cases, most of the flow passes through the furrows, where enhanced heat transfer characteristics are expected. The comparison of the values of the above Nusselt numbers shows that they do not differ more that 1%; therefore, the smooth part of the corrugated plate does not seem to influence the overall heat transfer. Figure 13b shows a typical local Nusselt number distribution over the corrugated wall for Re=900. All the Reynolds numbers studied exhibit similar distributions. It is noticeable that local Nusselt numbers attain their maximum value at the top of the corrugations. This confirms the strong effect of the corrugations, not only on the flow distribution, but also on the heat transfer rate. To the best of author's knowledge, experimental values of heat transfer and pressure drop are very limited in the open literature for the corrugated plate geometry, since these data are proprietary. Therefore, the data of Vlasogiannis et al. [16] were used to validate the simulation results. These data concern heat transfer coefficients measurements of both single (Re<1200) and two-phase flow in a plate heat exchanger with corrugated walls and a corrugation inclination angle of 60o. Heavner et al. [14] proposed a theoretical approach, supported by experimental data, to predict heat transfer coefficients of chevron-type plate heat exchangers. Figure14 presents the experimental friction factors, obtained from the Plexiglas test section of Figure 9, as well as the CFD predictions for the simple Where m and n constants with values 0.27 and 0.14 respectively. Heavner et al. [14] proposed a similar empirical correlation based on their experimental results on a single pass of a plate heat exchanger with 45? corrugation angle, but with two corrugated plates. In spite of the differences in geometry, it appears that the present results are in good agreement with the experimental data of Heavner et al. [14] (0.687 and 0.141 for the variables m and n, respectively). It must be noted that Focke et al. [15] , who also measured heat transfer coefficients in a corrugated plate heat exchanger having a partition of celluloid sheet between the two plates, reported that the overall heat transfer rate is the 65% of the corresponding value without the partition. Figure 15 shows that the mean j-Colburn factor values calculated using the overall Nusselt number are practically equal to the 65% of the values measured by Vlasogiannis et al. This holds true for all Reynolds numbers except the smallest one (Re=400). In the latter case the Nusselt number is greatly overpredicted by the CFD code. This is not unexpected, since the two-equation turbulence model is not capable to predict correctly the heat transfer characteristics for such low Reynolds number.The CFD results reveal that the corrugations enhance the heat transfer coefficient, whereas the pressure losses due to the augmentation of friction factor f are increased (Table 3), compared to a smooth-wall plate heat exchanger. Additionally, comparison of the normalized values of Nusselt number and the friction factor, with respect to the corresponding values for the smooth plate (fsm, Nusm), indicates that as the Reynolds number increases, heat transfer enhancement is slightly reduced, while the friction factor ratio, f/f , is increased. This is typical for plate heat exchangers with corrugations [16]. # Global # STUDY OF A HEAT EXCHANGER CHANNEL The results for the simplified geometry confirm the validity of the CFD code and strongly encourage the simulation of a module (pass) consisting of two corrugated plates of a compact heat exchanger (Figure16a). In order to quantitatively evaluate the results of this simulation, the experimental setup of Vlasogiannis et al. [16] was used as the design model (Figure 16b). Due to the increased computational demands, an AMD AthlonXP 1.7GHz workstation with 1GB RAM was used. The geometric characteristics of the new model are presented in Table 4. Preliminary results of the present study, which is still in progress, are shown in Figure 17. It is obvious that the herringbone design promotes a symmetric flow pattern (Figure 16b). Focusing on the left half of the channel (Figure 17a), a close-up of the flow streamlines (Figure17b) reveals a "peacock-tail" pattern as the liquid flows inside the furrows and over the corrugations. The same flow pattern, which is characteristic for this type of geometry, has also been observed by Paras et al. [14] in similar cross-corrugated geometries (Figure17c), where "dry areas" of ellipsoidal shape are formed around the points where the corrugations come into contact. The on the shape and the extent of these areas, which are considered undesirable, will be examined in the course of this study. The experimental data were taken at two different condensation temperatures of 20°C and 30°C in the range of mass flux of 14-34 kg/m 2 s with a heat flux of 4.7 -5.3 kW/m 2 . ? Both the heat transfer coefficient and the pressure drop increased proportionally with the mass flux and the vapor quality and inversely with the condensation temperature and the chevron angle. Those effects must be carefully considered in the design of a BPHE due their opposing effects. ? A comparison of the data for R410A and R22 showed that the heat transfer coefficient for R410A was about 0 -10 % larger and the pressure drop about 2-21 % lower than those for R22. Therefore, R410A is a suitable alternative refrigerant for R22. ? Correlations for the Nusselt number and the friction factor with the geometric parameters were suggested for the tested BPHEs within 20 % (r.m.s. deviation: 10.9 %) for Nu and 15 % (r.m.s. deviation: 10 %) for f. Although compact heat exchangers with corrugated plates offer many advantages compared to conventional heat exchangers, their main drawback is the absence of a general design method. The variation of their basic geometric details (i.e. aspect ratio, shape and angle of the corrugations) produces various design configurations, but this variety, although it increases the ability of compact heat exchangers to adapt to different applications, renders it very difficult to generate an adequate 'database' covering all possible configurations. Thus, CFD simulation is promising in this respect, as it allows computation for various geometries, and study of the effect of various design configurations on heat transfer and flow characteristics. In an effort to investigate the complex flow and heat transfer inside this equipment, this work starts by simulating and studying a simplified channel and, after gaining adequate experience, it continues by the CFD simulation of a module of a compact heat exchanger consisting of two corrugated plates. The data acquired from former simulation is consistent with the single corrugated plate results and verifies the importance of corrugations on both flow distribution and heat transfer rate. To compensate for the limited experimental data concerning the flow and heat transfer characteristics, the results are validated by comparing the overall Nusselt numbers calculated for this simple channel to those of a commercial heat exchanger and are found to be in reasonably good agreement. In addition, the results of the simulation of a complete heat exchanger agree with the visual observations in similar geometries. Since the simulation is computationally intensive, it is necessary to employ a cluster of parallel workstations, in order to use finer grid and more appropriate CFD flow models. The results of this study, apart from enhancing our physical understanding of the flow inside compact heat exchangers, can also contribute to the formulation of design equations that could be appended to commercial process simulators. Additional experimental work is needed to validate and support CFD results, and towards this direction there is work in progress on visualization and measurements of pressure drop, local velocity profiles and heat transfer coefficients in this type of equipment. This page is intentionally left blank # VII ( A ) # APPENDIX ![a) Flow regime Before the behaviors of heat transfer are considered, it is necessary to predict what flow regime exists at a given set of operating conditions. The detailed flow regime map for the PHE has not been proposed yet because of the difficulty of flow visualization. Vlasogiannis et al., [10] suggested the criterion of a two-phase flow regime for a PHE in terms of superficial liquid (jf) and vapor velocities (jg) by using water and air under adiabatic conditions. They only simulated a mixture of water and air as a two-phase fluid. According to their work, the flow patterns in a PHE are significantly different from those inside the vertical round tubes. They detected 3 types of flow patterns. The first was a gas continuous pattern with a liquid pocket at flow water flow rates (jf < 0.025 m/s) over wide range of air flow rates.](image-2.png "") ![Figure-3 shows the effects of the refrigerant mass flux, the chevron angle, and the condensation temperature on the averaged heat transfer coefficient for R410A. The term "averaged heat transfer coefficient" means the average of the heat transfer coefficients calculated by varying the quality of the refrigerant from 0.15 to 0.9, and the coefficients were obtained from Eq. (19):](image-3.png "") 4![also shows the variation of the heat transfer coefficients with the condensation temperatures. Like Figure-3, the heat transfer coefficients decreased with increasing condensation temperature. Also, the variations of the heat transfer coefficients with the condensation temperature are larger in the high-quality region. From the experimental results in Figures, 3 and 4, lowering the chevron angle and the condensation temperature gives the desired heat transfer effect. c) Frictional pressure loss](image-4.png "Figure- 4") ![Figure-5 shows the trend of the pressure drop along the mass flux, and Figure-6 shows the trend of the pressure drop along the quality at a mass flux of 34 kg/m 2 s and a heat flux of 4.7-5.3](image-5.png "") ![, 5 and 6, we may concluded that since the trends of the the condensation heat transfer and the pressure loss in BPHEs are similar, those effects must be carefully considered in the design of a BPHE.](image-6.png "") ![Figure-8(a) shows a comparison of the Nusselt number among the experimental data, the correlation proposed in this paper, and the correlation of Yan et al., [5]. The correlation of Yan et al., is and is obtained from one PHE with a chevron angle of 30° for R134a. Regardless of the BPHE types and refrigerants, most of the experimental data are within 20 % for the correlation proposed in this paper. of Nusselt number and friction factor for tested BPHEs data relatively well for : 20 and 35 within30 %, but over-predicted the data quite a bit for 45. This discrepancy results from the correlation of Yan et al., being developed for only a+30 PHE. Also,the correlation of Yan et al. for the Nusselt number only adopted the equivalent Reynolds number and Prandtl number without any geometric parameters. Because a BPHE has a strong geometric effect, the correlation with geometric parameters must be developed for general applications. The root-mean-square (r.m.s.) of the deviations is defined as ? The correlation of Y an et al.(5), matched the r.m.s. deviation for the correlation of Yan et al., [Eq. (29)] is 50.2 % and for Eq. (20), it is only 10.9 %. Figure-8(b) shows a comparison of the friction factor between the experimental data and the proposed correlation. Similar to the correlation of the Nusselt number, the correlation of the friction factor includes the equivalent Reynolds number and the geometric parameters. Regardless of the BPHE types and refrigerants, most of the experimental data are within 15 % of the correlation proposed in this paper; the r.m.s. deviation for Eq. (](image-7.png "") 1112![Figure 11.a Typical flow pattern for the: a) simple channel, CFD results, Re=900](image-8.png "Figure 11 Figure 12") ![Figure 13. : Typical results of the CFD simulation for Re=900; distributions of: (a) z-shear stress component](image-9.png "") 13![Figure 13. : Typical results of the CFD simulation for Re=900; distributions of: (b) local Nusselt number](image-10.png "Figure 13 .") 14![Figure 14. : Comparison of friction factor predictions (CFD) with experimental data](image-11.png "Figure 14 .") ![Figure 17. (a) : Module of a corrugated plate exchanger; (b) The CFD model and (c) Detail of the grid distribution over the corrugated wall.](image-12.png "") 18![Figure 18. (a) : Streamlines in the left half of the channel; (b) Close up of the flow pattern; (c) Photo of the flow in the cross-corrugated geometry [14] VII.](image-13.png "Figure 18 .") ![](image-14.png "") ![](image-15.png "") -.1ParametersUncertaintyTemperature±0.2 o CPressure±4.7KPaPressure drop±250PaWater flow rate±2%Refrigerant mass flux±0.5%Heat flux of Test Section±5.7%Vapour quality±0.03Heat transfer Coefficients of Water side ±10.1%Heat transfer Coefficients of Refrigerant ±9.1%VII( A )2011DecemberThe heat transfer coefficient of the refrigerant side (hr) was evaluated from the following equation: Nomenclature A b C p D f G Ge g h i j L c L h L v L w LMTD 18. mheat transfer area of plate [m 2 ] mean channel spacing [m] constant pressure specific heat [J/kg K] diameter [m] friction factor mass flux [kg/m 2 s] non-dimensional geometric parameter gravitational acceleration [m/s 2 ] heat transfer coefficient [W/m 2 K] enthalpy [J/kg] superficial velocity [m/s] distance between the end plates [m] distance between the ports [m] vertical length of the fluid path [m] horizontal length of the plates [m] log mean temperature difference [°C] mass flow rate [kg/s]fg fr g in lat m out p pre r s sat sens w Eq fdifference the liquid phase and the vapor phase friction vapour inlet latent mean outlet port pre-heater refrigerant static saturated sensible water equivalent liquidN cpnumber of channels for the refrigerantN datatotal number of dataN ttotal number of platesNuNusselt numberNu expNusselt number obtained from experimentNu predNusselt number obtained from correlationpplate pitch [m]p cocorrugation pitch [m]PrPrandtl number [v]Qheat transfer rate [W]qheat flux [W/m 2 ]ReReynolds numberTtemperature [°C]tplate thickness [m]Uoverall heat transfer coefficient [W/m 2 K]xVapour qualitySubscriptsaaccelerationcchannel December © 2011 Global Journals Inc. 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