AUTOMOTIVE COMPACT SURPERCHARGE-AIR INTERCOOLER NUMERICAL ANALYSIS G. Starace, E. Carluccio, D. Laforgia Università degli studi di Lecce, Dipartimento di Ingegneria dell Innovazione via per Arnesano, 73100, LECCE Abstract A compact aluminium intercooler for supercharge air for use in cars and trucks was numerically studied moving from an overall approach to a deep insight in finnings. The influence of different input values of cooling and supercharge air flowrates was observed, investigating microscopic effects on the flow field, on heat transfer local convective coefficients and macroscopic effects on performance variations in terms of heat transfer and pressure losses. Triangular wavy finnings, trapezoidal offset strip finnings (squared and rounded, single or coupled) and inside finned tubes efficiencies were calculated with reference to the different input conditions, useful for possible correlation equations among adimensional parameters for future design and sizing activities. 1. Introduction New-concept compact heat exchangers represent sometimes the only solution in terms of dimensions for applications on vehicles. They are unfortunately hard to be optimized as in most cases wrong design directions result in non-efficient heat transfer processes and in high pressure losses. The most advanced instruments have indeed to be used, as CFD, in order to give detailed predictions on heat transfer enhancment techniques and overall performances. The aim of this work is comparing possible different solutions in the finning choice of an aluminium intercooler used to cool the supercharge air to engines (Figure 1). A variety of increased heat transfer surfaces have been studied: plain, wavy, offset strip, perforated and louvered fins (Carluccio et al., 2005 and Roshsenow et al., 1998). Offset strip fins are very widely used as they have a high degree of surface compactness, and heat transfer enhancement is obtained as a result of the periodic starting and development of laminar boundary layers over uninterrupted channels formed by the fins and their dissipation in the fin wakes (Manglik and Bergles, 1995]. The numerical analysis was carried out investigating differences in terms of heat transfer, local heat transfer convective coefficient distributions and pressure losses, among finnings at different Figure 1 A supercharged engine scheme (left) and the intercooler under investigation (right)
Figure 2 Finnings crossed by fresh air hydrodynamic regimes,. As in a previous work of the same authors (Carluccio et al., 2005), the analysis was developed through three steps of increasing complexity. After a simple model built to evaluate the consistent boundary conditions of temperature, a small scale analysis of the fluid dynamic phenomena inside the intercooler was carried out, taking into account geometrical periodical units both at ambient air and at compressed air sides. 2. The finning geometries At ambient air side, two different geometries were accounted for: wavy fin (WF) with triangular cross section, as in Figure 2 (the upper part) single channel rounded offset turbulence enhancers (S-ROTE), as in Figure 2 (the lower part) At compressed air side, two different geometries were accounted for: double channel offset strip fin with squared turbulence enhancers (D-SOTE), as in Figure 3 (the upper part) extruded internally finned tube (EIFT), as in Figure 3 (the lower part) Figure 3 Finnings crossed by compressed air
3. Simulation Conditions The cases were run considering an aluminium (ρ=27.9[kg/ m 3 ]; c=871[j/(kg K)]; λ=202.4[w/(mk)]) supercharge air intercooler, designed for a 50[kW] heat transfer, operating in steady state, not dissipating heat towards surrounding environment. The heat capacities were supposed constant at an average temperature to reduce computational costs. The separating wall surface area was of 2.04[m 2 ]. The mean temperature of the plates was calculated with the same method described in a previous work of the same authors (Carluccio et al.,2005) and was of 345.53[K]. The calculated flowrates and inlet conditions were the following: 0,5[kg/s] for compressed air (to be cooled) coming from the turbocharger at a temperature of 160 C; 2.08[m 3 /s]for ambient air (30 C) supplied by the axial fan provided with the intercooler with density of 1.225[kg/m 3 ], specific heat at constant pressure of 1006.4[J/(kg K)], thermal conductivity of 0.024[W/ m K] and viscosity of 1.7894 E-5[kg/(m s)]. For each considered geometries, a half and a double values of flowrate were simulated with respect to the design one (Table 1). The adopted turbulence model was the Standard k-ε, with wall condition set up at Enhanced Wall Treatment, particularly appropriate for this simulation of wall effects on thermal and fluid dynamic boundary layer; the grid accuracy was evaluated through the Roache method as in a previous work of the same authors (Carluccio et al., 2005). In order to consider an average behaviour of the finning (the one supposed in the middle of the intercooler), a first simulation was carried out to get the right exit velocity distribution to be set as boundary condition at the air inlet in the next simulations. In this way, a completely developed flow in terms of local velocities and turbulence was investigated, dispelling doubts about the region influenced by flow inlet. This approach was common to all finnings taken into consideration. Table 1 Fluid dynamic and geometrical parameters set in the simulations Finning typology Parameter D-SOTE WF EIFT S-ROTE Pr 0.744 0.744 0.744 0.744 D h [mm] 2.5 4.01 3.2 3.82 Design valueof V i [m/s] 22 18 26 18 Half value of V i [m/s] 11 9 13 9 Double value of V i [m/s] 44 36 52 36 Design value of Re 7535 4941 11159 4711 Half valueof Re 3767 2501 5579 2356 Double value ofre 15071 9883 22318 9422 Wet surface / Separating surface 10.16 9.11 7.53 9.12 Inlet free section / Total cross section 0.75 0.89 0.64 0.86 Double Channel Turbulence Enhancers A 3D periodic unit was simulated, made of two mirror coupled trapezoidal squared offset strip fins crossed by compressed air. Computational constraints, together with the consideration that a fully developed flow had to be taken into account for feasible results, induced to simulate a five turbulence enhancers sequence; their width was their minimum periodic dimension, (5.5[mm]) and the sheet metal thickness was of 0.2[mm]. Each fin is shifted of 1[mm] with respect to the previous one. The modelled flow duct was contained between two walls (of half-thickness each), which separated the flow from the ambient air and from the mirrored geometry. The boundary temperature were set to the previously calculated mean value at the separation between the two fluids.
Figure 4 - Velocity (left) and temperature (right) flow field in planes parallel and normal to the flow across the Double Channel Turbulence Enhancers At the side sheet, a geometrical, fluid dynamic and thermal symmetry condition was imposed. Figures hereafter reported, are depicted with reference to a mirrored geometry and so to a double calculation domain. Wavy Finning This finning was simulated being crossed by ambient air. The second and more complex numerical schematization of the AS of the HX was done as following: a periodic unit was cut from the device and this included two complete wavy triangular flow channels (whose longitudinal dimension was the same as the entire intercooler) bounded on the top and at the bottom by two aluminium walls of half thickness at the average temperature calculated at the first step. Extruded Internally Finned Tube This geometry was compared with the Double Channel Turbulence Enhancers. The different dimensions of the two geometry lead to a completely different compressed air inlet velocities for the same flowrate. The domain volume was chosen to be representative of the completely developed thermal and fluid dynamic fields. The upper and lower walls were set to the average temperature between the two fluid. On the side rounded wall an adiabatic condition was set up and on the other side a periodicity Figure 5 - Velocity (left) and temperature (right) flow field in planes parallel and normal to the flow in the Wavy Finning
condition was imposed. Inlet velocity and temperature profiles were set after a purposely run simulation to cancel borders effects by sharpen front edges of the intercooler. Rounded Offset Strip Fins The schematization was carried out to be compared with the wavy fin crossed by ambient air. The finning solutions are alternative in the overall intercooler dimensions. For this reason the inlet flowrates were the same, as well as the temperature setup. The solid cross section of these fins corresponds to that of the wavy fin. This means that the inlet velocities are the same. This finning typology has the following characteristics to be carefully examined when adopting it: the high heat transfer efficiency due to the wide surface and to the high level of turbulence generation, has as weak point the presence of continuous obstacles to the flow that causes high pressure losses. This geometry has to be used in particular environment. 3. Results In the following figures the images of velocity and temperature fields were reported, on the median planes parallel and normal to the flow, by which modelled geometries were cut. On those planes the boundary layer structures can be observed for the compared finnings. Different inlet mean velocities come from different finnings cross dimensions at fixed volumetric flowrates both for fresh air and compressed air channels. Overall dimensions of the modelled volumes were chosen keeping an acceptable compromise between computational times and significance of the represented flow field. The volume has to be thought localized in a median position along the path of the compressed air in the HX. As said above, border effects were reduced with preliminary runs, to find proper inlet velocity profiles to be set as inlet boundary conditions. The complete fluid path along the HX was, instead, modelled for the fresh air side. For all the geometries it was found that no qualitative modifications occurred, even in presence of very different inlet flowrates; the increase of the turbulence level with velocity was instead apparent. Double Channel Turbulence Enhancers Along the consecutive single turbulence enhancers, the compressed air was slowed down in correspondence of the solid walls, and a consequent acceleration in the centre of the HX channel could be observed. The flow showed an evident trace of the shift of one enhancer with respect to the consecutive one; in particular the turbulence intensity reached its maximum values in the trailing edges. The thermal evolution was similar: the interactions between the flow and the wall determined the compressed air cooling by convection. The effect of the turbulence enhancers walls was apparent on the compressed air temperature when they were overcome, and so a region with better cooling efficiency could be identified (Figure 4). Wavy Finning The fresh air flow field inside the heat exchanged revealed itself qualitatively identical for each case studied. The figure 5 of the fresh air channel showed a flow characterized by velocity peaks in correspondence with finnings convexities and by stagnation regions close to the concavities. Most of the flow close to concave edges led to a higher local thermal gradient between the wall and the undisturbed flow, and so to a more effective heat transfer. Extruded Internally Finned Tube The third modelled geometry, was supplied by compressed air, and represented an alternative to the double channel turbulence enhancers. For this reason the same inlet and operating conditions were set up for the simulations (Tab.1). The boundary layer developed itself along walls, with evident slow downs in the upper and lower regions between separating walls and fins. The deceleration caused by the presence of these regions produced a velocity increase in the centre of the channels,
Figure 6 Velocity (left) and temperature (right) flow field in planes parallel and normal to the flow in the Extruded Internally Finned Tube and made the fin tip the most effective point from the heat transfer point of view. The side channel had a different shape and cross section and so its behaviour was different from internal channels. For their limited weight these channels effects on global heat transfer could be neglected. Observing the temperature field, it can be seen that the cooling effect was less intense in upper and lower regions where flow was slowed down. In the region were the two fins face each other a high thermal gradient was present from the tip to the channel centre and this demonstrated the high contribute of the fins to the total heat transfer (Figure 6). Rounded Offset Strip Fins In this geometry the high heat transfer area, the high number of flow interruptions and the shifting of the enhancers, caused a high turbulence level, and an effective heat transfer. These elements were, anyway, the same that led to high pressure drops. The solid cross section of the fins was the same as the wavy fin, an this brought to equal mean inlet velocities. As with the double channel turbulence enhancers, the inlet compressed air experienced a slowdown in the vicinity of the solid walls, and a consequent acceleration could be observed in the centre of the channel; the effect of the shifting of the single enhancers had effects on the flow evolution in terms of turbulence level: this reached the highest values in the trailing edge of the enhancers. From the thermal point of view, similar conclusions can be made: localized regions of high thermal effectiveness could be identified after each enhancer vertical walls. (Figure 7). For each modelled geometry an optimization analysis in terms of heat transfer and pressure drop Figure 7 Velocity (left) and temperature (right) flow field in planes parallel and normal to the flow in the Rounded Offset Strip Finnings
j 0,006 0,005 0,004 0,003 0,002 0,001 0,000 3000 7000 11000 15000 19000 23000 Re f 0,10 0,08 0,06 0,04 0,02 0,00 j factor forr finned tube j factor for double channel turbulence enhancers f factor for finned tube f factor for double channel turbulence enhancers Figure 8 Performance comparison of double turbulence enhancer finning and extruded internally finned tube for compressed air was performed with three different of the inlet flowrate values: design value, half and double flowrates. Compressed Air Channels Comparing the performance of the geometries, it can be observed that the overall heat transfer increases as the inlet flowrates became higher. This happened for both the geometries and the consequence was that the total heat transfer per unit area increased. For the double channel turbulence enhancers, the curve rate (that represents the heat transfer) showed a better effectiveness in the range of velocities lower than the design value, with respect to the high values. This determines the presence of an optimal point for heat transfer performance. In the case of the extruded internally finned tube, instead, the heat transfer increased proportional to inlet velocity. The analysis of the data of the pressure losses confirmed for both geometries that the design value 60 50 40 v inlet 11 m/s v inlet 22 m/s v inlet 44 m/s 60 50 40 Inlet Centre Outlet v inlet 26 m/s Nu 30 30 20 20 10 10 0 0,3 0,4 0,5 0,6 0,7 0,8 0,9 0 0 0,1 0,2 0,3 0,4 0,5 0,6 0,7 0,8 x / Normalized Cross Dimension Figure 10 Trend of Nu for the three compressed air flowrates in the double turbulence enhancer channel at half path(left), and for the design compressed air flowrate at the inlet, in the center and at the outlet of the extrude internally finned tube (right)
of velocity represented the limit for the compressed air to have an acceptable pressure drop per unit length. The double turbulence enhancers channel showed a better thermal efficiency as the heat transfer was twice the one realized with the extruded internally finned tube; this behaviour could be observed for all the studied flowrates. Concerning the pressure drops, it can be shown that, for high velocity values, the double channel became the worst and this greatly lowered its attractiveness (Figure 8). Fresh air channel Comparing the two geometries performance at the fresh air side, it can be observed that in both cases the heat transfer increased in the same way as the volume flowrate, even if the rounded offset strip fins showed a higher heat transfer intensity than the wavy finning. The overall pressure drop along the HX increased, instead, as a function of the square velocity of the air flow. That was relevantly higher in the offset strip fins than in the triangular wavy finning (Figure 9). Local convective coefficients A deep analysis of the thermo-fluid dynamic behaviour in the studied cases led to the investigation of the local convective heat transfer coefficient h in correspondence of the wall separating the hot and the cold fluids. Three different locations of investigation were chosen: the inlet inside the space were finnings were already present, the centre and the outlet of the channels. This was done for each flowrate and each geometry. The comparison among results is shown in figures no.10 and 11 in terms of the Nusselt number plotted as a function of an opportune normalized geometric parameter. In figures 10 (left) and 11 (left) the increase of absolute value of the local convective heat transfer coefficient with the flowrate (and so with velocities) can be observed, remaining its trend very similar with different geometries. The local values obtained tend gradually to decrease going from the channel inlet to the exit, and this is caused by a decrease of the ΔT between the fluid and the cold wall; in particular in Figure 10 (right), concerning the extruded internally finned tube and crossed by compressed air, the data related to the extreme side of the channel were deleted, as they were not significant with respect to the total volume. The maximum values are those correspondent to the centre of the channel, characterized by a lower fluid dynamic resistance. In Figure 11 (left), the differences among the flow behaviour at the three different inlet velocity values can be observed for fresh air. The heating trend can be easily seen in Figure11, referred to j 0,028 0,024 0,020 0,016 0,012 0,008 0,004 f 0,14 0,12 0,10 0,08 0,06 0,000 0,04 2000 4000 6000 8000 10000 Re j factor for wavy fin j factor for rounded offset strip fin f factor for rounded offset strip fin f factor for wavy fin Figure 9 Performance comparison between rounded offset strip fin and wavy finning for fresh air
140 120 100 v inlet 9 m/s v inlet 18 m/s v inlet 36 m/s 140 120 100 Inlet Centre Outlet v inlet 18 m/s Nu 80 60 40 20 80 60 40 20 0 0,2 0,4 0,6 0,8 1 0 0 0,1 0,2 0,3 0,4 0,5 0,6 0,7 0,8 0,9 1 x / Normalized Cross Dimension Figure 11 Trend of Nu for the three fresh air flowrates in wavy finning at half path (left), Trend of Nu for the design fresh air flowrate at the inlet, in the centre and at the outlet of the rounded offset strip fin (right) the compressed air temperature variation at the centre of the duct along the flow. Increasing the flowrate, the temperature difference between the exit and the inlet is obviously lower, but the overall heat transfer per surface unit is higher. 4. Conclusions In this work different finning geometries were examined and compared to get information about their behaviour when used in an automotive compact supercharge-air intercooler. The double channel turbulence enhancers compared with the extruded internally finned tube, both crossed by compressed air to be cooled and redirected to the engine, has shown a better behaviour. The former allows to obtain an almost double thermal efficiency in terms of heat transfer with respect to the latter, for each analyzed flowrate value. Pressure losses show a completely opposite trend. The same differences can be found when comparing the rounded offset strip fins with the wavy finning. Here the former has a double heat transfer at the same flowrate values and more intense pressure losses due to frequent flow interruptions. The analysis of local convective heat transfer coefficient can help in understanding new design strategies to maximize compact heat exchangers heat transfer. References Carluccio E., Starace G., Ficarella A., Laforgia D., Numerical analysis of a cross-flow compact heat exchanger for vehicle applications Applied Thermal Engineering Vol. 25, pages 1995 2013, 2005 Durbin P. A., Medic G., Seo J.-M., Eaton J. K., Song S. - Rough Wall Modification of Two- Layer k-ε - Journal of Fluids Engineering, Vol. 123, pages 16-21, March 2001. Jang Y. J., Chen H. C., Han J. C. - Computation of Flow and Heat Transfer in Two-Pass Channels with 60 Deg Ribs - Journal of Heat Transfer, Vol. 123 No. 3, pages 563 575, June 2001. Kays W. M., London A. L. Compact Heat Exchangers - (Mc-Graw Hill) 3rd edition (1984).
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