Jet Impingement Cooling

Transcription

1 Jet Impingement Cooling Colin Glynn*, Tadhg O Donovan and Darina B. Murray* *CTVR (Centre for Telecommunications Value-Chain-Driven Research), Department of Mechanical and Manufacturing Engineering, Trinity College, Dublin Abstract Convective heat transfer to impinging jets is known to yield high local and area averaged heat transfer coefficients. Impingement jets are of particular interest in the cooling of electronic components where advancement relies on the ability to dissipate extremely large heat fluxes. The current research is concerned with the measurement and comparison of heat transfer to submerged and confined air and water jets. The present paper reports on surface heat transfer measurements for a range of test parameters. These parameters include jet diameter (d from 0.5mm to 1.5mm), Reynolds number (Re from 1000 to 20000) and jet to target spacing (H from 0.5d to 6d). The results are presented in the form of local and full-field heat transfer coefficients for the range of conditions indicated. Introduction Jet impingement is an attractive cooling mechanism due to the capability of achieving high heat transfer rates. This cooling method has been used in a wide range of industrial applications such as annealing of metals, cooling of gas turbine blades, cooling in grinding processes [1] and cooling of photovoltaic cells [2]. Jet impingement has also become a viable candidate for high-powered electronic and photonic thermal management solutions and numerous jet impingement studies have been aimed directly at electronics cooling [3-8]. A fluid jet issuing into a region containing the same fluid is characterised as a submerged jet while a fluid jet issuing into a different, less dense, fluid is characterised as a free-surface jet. Womac et al. [9, 10] have shown that higher heat transfer coefficients result from submerged jet conditions than from free-surface jet conditions for Re An impinging jet is said to be confined or semiconfined if the radial spread is confined in a narrow channel, usually between the impingement surface and the orifice plate. The presence of a confining top wall in jet impingement causes lower heat transfer coefficients, thought to be caused by the recirculation of fluid heated by the target plate [11, 12]. Huang et al. [13] suggest that confinement promotes a more uniform heat transfer distribution for the area enclosed by a non-dimensional radial distance from the stagnation point (r/d) of 5. The key parameters determining the heat transfer characteristics of a single impinging jet are the Reynolds number, Prandtl number, jet diameter and jet-to-target spacing. Nozzle geometry can also have a sigficant influence on heat transfer. Numerous studies have been conducted to investigate the influence of each of these parameters. For all studies performed it has been shown that, for a constant jet diameter, heat transfer increases with increasing Reynolds number, with Nu proportional to Re 0.5 to 0.8 [14-20]. It has also been shown that, for a constant Reynolds number, decreasing the jet diameter yields higher stagnation and average heat transfer coefficients [9, 14, and 21]. This can be attributed to the higher jet velocities created by the smaller nozzles. It has been shown that the jet-to-target spacing has a much greater influence on heat transfer for submerged jets than for free-surface jets [9]. Many studies have shown little change in stagnation and average heat transfer for H/d < 4, then a decrease in heat transfer as H/d increases beyond this point [9, 22]. The relative consistency of heat transfer for H/d < 4 in the above studies can be 1

2 explained by the jet impingement taking place within the potential core with its nearly uniform velocity, while the decrease in heat transfer at higher H/d values is attributed to complete degradation of the potential core prior to impingement. Baughn and Shimizu [23] tested for H/d between 2 and 14 for a large air jet diameter (25mm) and Reynolds number of and found the maximum stagnation heat transfer coefficient to be at H/d = 6. However, Lytle and Webb [20] show increasing heat transfer with decreasing H/d from 6 to 0.1. Several researchers [14, 18, 20, 23, and 24] have shown the presence of secondary peaks for low non-dimensional jet-to-target spacings, with some studies showing that the secondary peaks increase in magnitude relative to stagnation heat transfer with increasing Reynolds number and decreasing H/d, eventually becoming global maxima. Lytle and Webb [20] show that the position of the secondary peak moves toward the stagnation point with decreasing H/d. The minimum position of the secondary peak measured by [20] was at a non-dimensional radial distance of r/d 1.2. All secondary peaks in the other above studies occurred at 1.5 r/d 2. Secondary peaks in this region can be explained by transition to turbulence within the wall jet [12]. Decreasing H/d increases the turbulence levels close to the impingement surface. Huber and Viskanta [25] found two secondary peaks for H/d = 0.25 and a nozzle-diameter of 6.35mm. These peaks occurred at approximately r/d = 0.7 and 1.5. The inner peak was attributed in [25] to the fluid accelerating out of the stagnation region, thus thinning the local boundary layer, and to the influence of the shear-layer generated turbulence around the circumference of the jet. The current research investigates the local heat transfer coefficients for single, axisymmetric, submerged and confined impinging air and water jets. Jet diameters of 0.5mm, 1mm and 1.5mm are investigated for impinging jets with Reynolds numbers from 1000 to and non-dimensional jet-to-target spacings (H/d) from 1 to 4. Experimental Set-up Experiments were conducted for both air and water jets impinging on a uniform wall flux target surface. The same impingement configuration was employed for both the water and air jets; however different fluid flow test loops supplied the jets. A schematic of the test loops used for each working fluid is shown in figure 1. Air is supplied from the building compressed air supply. The flow rate through the test system was controlled using a pressure monitored lever valve. Downstream of the control valve an Alicat Scientic Inc. Precision Gas Flow Meter recorded both the air flow rate in standard litres per minute (SLPM) and temperature on entry to the jet impingement test system. The water test loop is a closed system. A large tank (0.025 m 3 ) filled with de-ionised water acts as a reservoir. Water from the reservoir is circulated through the flow loop using a 0.75kW 2900rpm positive displacement pump. The flow rate is controlled using an in-series gate valve combined with a parallel bypass valve. The flow rate is recorded from a glass tube flow meter, which itself can act as a valve, allowing finer flow rate control. The fluid then flows through the jet impingement test section and is returned to the water reservoir. The reservoir water temperature remains constant throughout the small number of tests performed in a single test set due to the large tank volume (0.025 m 3 ) compared with the very small flow rates (maximum of 2.36 x 10-5 m 3 /s). This is verified by the constant monitoring and recording of the reservoir water temperature. 2

3 Figure 1: Schematic of Test Loop with Air (left) and Water (right) as the Working Fluid The common jet impingement test section used is shown in figure 2. The fluid enters the test system into a chamber of 50mm diameter. An orifice plate is attached at the exit of this chamber, creating the desired impinging jet. Three orifice plates are used in the current study, containing single nozzles of 0.5mm, 1mm and 1.5mm in diameter respectively. The orifice plates are of 5mm constant thickness and all nozzles are sharp-edged. A single jet issues from the orifice plate and impinges normally onto an electrically heated stainless steel foil (target surface). Figure 2: Jet Impingement Test Section The ratio of jet diameter to orifice plate diameter and the small jet-to-target spacings used (1 H/d 4) ensured confinement. Four fluid exits of 10mm diameter are positioned in the four corners of the test chamber ceiling. Optical access is provided through four windows on each side of the test chamber. This is to enable future PIV fluid-flow measurements and for flow visualisation. The stainless steel foil used as the heating element has a thickness of 25µm. The foil is electrically heated by clamping at each end to one of the large sides of two opposite copper bus bar electrodes (50mm x 10mm x 10mm). Each clamp is between the bus bar and a copper plate (50mm x 10mm x 2mm), minimising contact resistance. The distance between bus bars is 60mm and the width of the foil is 50mm. External electrical connection is obtained via three copper rods that pass through each bus bar and extend to the exterior of the test rig (a total of six rods). The clamps also function as a tensioning device that tensions the foil over a 20mm diameter aperture. The electrical supply is from a DC Genesys 6V, 200A power source. Either the voltage or the current can be varied with such a power source, with the other following naturally due to system resistance. Both voltage and current were monitored using a National Instruments data acquisition system. The thermal boundary condition on the impingement surface was that of a uniform heat flux. 3

4 Results are presented in the form of a local heat transfer coefficient defined in equation 1, q h = (1) T s T j The heat flux, q, is determined from the product of the current and the voltage drop across the bus bars. All the power can be considered to be dissipated in the area between the bus bars due to negligible contact resistance between the power supply and the foil heater. T s represents the local surface temperature of the heater and is measured on the underside of the heated foil. Lateral conduction in the foil is considered negligible. Therefore, the temperature on the underside of the foil is representative of the temperature on the impingement side of the foil. Full-field temperature measurements are made using a FLIR ThermaCAM A40 infrared camera. The camera, with a 50µm lens, has 240 x 320 pixels and has a field of view of 11mm x 15mm. Results are presented as single profiles passing through the stagnation point. The underside of the foil was sprayed with a thin layer of black gloss paint of known emissivity (ε = 0.92). Reynolds number is defined in equation 2, 4. ρ. Q Re = (2) π. d. µ Results and Discussion A full field distribution of the heat transfer coefficient from the surface to an impinging air jet is presented in figure 3. The two dimensional plane measures 10 diameters square and the stagnation point heat transfer coefficient can be identified as the maximum that occurs at the geometric centre of the impinging jet. The near axisymmetry of the full field distribution about the geometric centre allows future results to be presented as heat transfer profiles through the geometric centre from -5d to +5d. Figure 3: Full Field Heat Transfer Coefficient Distribution; d = 0.5mm, Re = 5000; H/d = 1 (Air Jet) The results presented in figure 4 are heat transfer profiles to an impinging air jet for a range of Reynolds numbers and distance of the jet from the impingement surface. It is apparent that, even at relatively low Reynolds numbers, high heat transfer coefficients can be achieved by small diameter nozzles. As expected the maximum heat transfer coefficient is recorded at the stagnation point, and decreases with increasing radial distance for the range of parameters reported in this figure. For each of the two Reynolds numbers presented, both the peak and the area averaged heat transfer coefficient decreases as the distance between the jet exit and the impingement surface increases. For a jet diameter of 0.5mm the maximum Reynolds number tested was 5000 due to the large pressure drop across the jet orifice at such small diameters. 4

5 Figure 4: Local Heat Transfer Coefficient; d = 0.5mm; H/d = 2, 4; Re = 1000, 5000; (Air Jet) The results presented in figure 5 illustrate the effect of varying jet diameter on the heat transfer distribution for a Reynolds number of 5000 and H/d = 2. The increase in heat transfer with decreasing jet diameter is clear. This is due to the high air flow velocities involved for smaller diameter jets. An increase of over 60% is observed for a decrease in jet diameter from 1.5mm to 0.5mm. Figure 5: Local Heat Transfer Coeffiicient Distribution; d = 0.5mm, 1mm, 1.5mm; H/d = 2; Re = 5000; (Air Jet) It is worth noting that figure 5 displays the heat transfer coefficient distribution for the dimensionless radial distance, r/d. The maximum dimensional radial distance shown in figure 5 is 2.5mm for the jet diameter of 0.5mm compared to 7.5mm for the jet diameter of 1.5mm. The higher stagnation heat transfer coefficients combined with the smaller effective heat transfer area makes the use of arrays of small diameter impinging jets an attractive option. Figure 6 show the variation in the heat transfer coefficient distribution with dimensionless jet-totarget surface spacing (H/d) for a jet diameter of 1.5mm and Reynolds number of Flattening of the heat transfer distribution about the stagnation point and secondary peaks are present in all cases with the radial peaks most prominent for H/d = 1. The heat transfer coefficient is a local minimum at the stagnation point and increases to a radial peak at approximately 0.5 to 1d. This is thought to be due to thinning of the wall jet boundary layer as the wall jet air flow escapes from the lip of the impinging jet flow. At low H/d ( 2) the magnitude of the heat transfer coefficient is similar for the entirety of the distribution. This is attributed to the uniformity of the arrival jet velocity and turbulence intensity when the impingement surface is located within the potential core of the jet. Beyond the core of the jet (H/d = 4) the arrival velocity has decreased and therefore so too does the heat transfer coefficient. 5

6 Figure 6: Local Heat Transfer Distribution; d = 1.5mm, H/d = 1, 2, 4; Re = 20000; (Air Jet) Results have also been obtained for impinging water jets and figure 7 presents heat transfer distributions for a 1mm diameter jet and H/d = 1 (left) and 2 (right). Overall the magnitude of the heat transfer coefficients is far greater for the water jets than it is for air. Flattening of the heat transfer distributions is more obvious at the lower H/d = 1. At H/d = 1 and Re = the heat transfer coefficient is a local minimum at the stagnation point and two peaks are located at radial locations of approximately r/d = 0.5 and 2. These peaks are attributed to the thinning and the transition of the wall jet boundary layer respectively. Unusually, the peaks are not as pronounced when H/d = 1 and Re = This could be due to the confinement condition. Figure 7: Local Heat Transfer Coefficient Distribution; d = 1mm; (Water Jet) At the larger jet to impingement surface spacing of H/d = 2, the heat transfer distribution is more uniform and has a lesser magnitude. This indicates that the core length of the water jet is quite short, especially in comparison with the air jet results presented previously in figure 6. Figure 8 shows a comparison of the heat transfer distributions for both air and water for a Reynolds number of 10000, a diameter of 1mm and a range of jet to impingement surface spacings. It is once again apparent that the flattening of the heat transfer distributions to an air jet is more substantial and occurs up to a larger H/d for air than it does for water. While secondary peaks occur in the heat transfer distribution for a water jet with a Reynolds number of at H/d of 1, both the magnitude and the shape of the profile at H/d = 2 indicate that the jet is fully developed at this stage. 6

7 Figure 8: Local Heat Transfer Distribution; d = 1mm, Air Jet (left), Water Jet (right) Conclusions Heat transfer to single, axisymmetric, submerged and confined air and water jets for jet diameters of 0.5mm to 1.5mm, Reynolds numbers of 1000 to and dimensionless jet-to-target spacings of 1 to 4 was investigated. Local heat transfer coefficients were presented as a function of dimensionless radial distance from the stagnation point, r/d. It has been shown that the area averaged heat transfer increases with decreasing jet diameter and this is attributed to the higher jet velocities involved when smaller nozzles are used. For the air jets, secondary peaks were present at low jet-to-target spacings and high Reynolds numbers. The peaks became more pronounced with decreasing H/d and increasing Reynolds number. The water jets also exhibit secondary peaks, however these have only been observed at a low Reynolds number of and a low H/d of 1. Nomenclature d = jet or nozzle diameter, m h = heat transfer coefficient, W/m 2 K. H = distance between orifice and impingement plates, m q = heat flux, W/m 2 Q = volumetric flow rate, m 3 /s r = radial distance from the stagnation point, m Re = Reynolds number, dimensionless T = temperature, K ε = emissivity, dimensionless µ = dynamic viscosity, kg/ms ρ = density, kg/m 3 Subscripts s = impingement surface. j = impinging jet. 7

8 Acknowledgements This research is conducted within the CTVR (Centre for Telecommunications Value-Chain-Driven Research) and is supported by Science Foundation Ireland under its CSET programme. References 1. Babic, D., Murray, D. B., Torrance, A. A., Mist Jet Cooling of Grinding Processes, Int. J. Mach. Tools Manufact. 45, pp , Royne, A., and Dey, C., Experimental Study of a Jet Impingement Device for Cooling of Photovoltaic Cells Under High Concentration, Proceedings of Solar 2004: Life, the Universe and Renewables, Narumanchi, S. V. J., Amon, C. H., Murthy, J. Y., Influence of Pulsating Submerged Liquid Jets on Chip-Level Thermal Phenomena, Transactions of the ASME, Vol. 125, pp , Wang, E. N., Zhang, L., Jiang, L., Koo, J. M., Goodson, K. E., Kenny, T. W., Maveety, J. G., Sanchez, E. A., Micromachined Jet Arrays for Liquid Impingement Cooling of VLSI Chips, Solid-State Sensor, Actuator and Microsystems Workshop, Hilton Head Island, South Carolina, June 2-6, Fabbri, M., Jiang, S., Dhir, V. K., Experimental Investigation of Single-Phase Micro Jets Impingement Cooling for Electronic Applications, Proc. of the HT2003 ASME Summer Heat Transfer Conference, Las Vegas, Nevada, USA, July 21-23, Cheng, Y., Tay, A. A. O., Hong, X., An Experimental Study of Liquid Jet Impingement Cooling of Electronic Components With and Without Boiling, Int. Symp. on Electronic Materials and Packaging, IEEE, pp Kercher, D. S, Lee, J.-B., Brand, O., Allen, M. G., Glezer, A, Microjet Cooling Devices for Thermal Management of Electronics, IEEE Transactions on Components and Packaging Technologies, Vol. 26, No. 2, pp , Hollworth, B. R., Durbin, M., Impingement Cooling of Electronics, Journal of Heat Transfer, ASME, Vol. 114, pp , Womac, D. J., Aharoni, G., Ramadhyani, S., Incropera, F. P., Single Phase Liquid Jet Impingement Cooling of Small Heat Sources, Proceedings of the International Heat Transfer Conference, pp , Womac, D. J., Ramadhyani, S., Incropera, F. P., Correlating Equations for Impingement Cooling of Small Heat Sources with Single Circular Liquid Jets, Transactions of the ASME, Vol. 115, pp , Fitzgerald, J. A., Garimella, S. V., Flow Field Effects on Heat Transfer in Confined Jet Impingement, Transactions of the ASME, Vol. 119, pp , Fitzgerald, J. A., Garimella, S. V., A Study of the Flow Field of a Confined and Submerged Impinging Jet, International Journal of Heat and Mass Transfer, Vol. 41, Nos 8-9, pp , Huang, C. L., Shieh, Y. R., Hung, Y. H., Local Thermal Characteristics of a Confined Round Jet Impinging onto a Heated Disk, Electronic Packaging Technology Conference, IEEE, pp , Garimella, S. V. and Rice, R. A., Confined and Submerged Liquid Jet Impingement Heat Transfer, Journal of Heat Transfer, Vol. 117, pp ,

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