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7-9 October 2009, Leuven, Belgium Heat transfer enhancement due to pulsating flow in a microchannel heat sink T. Persoons * , T. Saenen, R. Donose, M. Baelmans Katholieke Universiteit Leuven, Department of Mechanical Engineering (TME) Celestijnenlaan 300A, P.O. box 2421, 3001 Leuven, Belgium Abstract – Heat sinks with liquid forced convection in microchannels are targeted for cooling microelectronic devices with a high dissipated power density. Given the inherent stability problems associated with two-phase microchannel heat transfer, this paper investigates experimentally the potential for enhancing single-phase convection cooling rates by applying pulsating flow. To this end, a pulsator device is developed which allows independent continuous control of pulsation amplitude and frequency. For a single microchannel geometry and a range of parameters (steady and pulsating Reynolds number, Womersley number), experimental results are presented for the overall heat transfer enhancement compared to the steady flow case. Enhancement factors up to 40% are observed for the investigated parameter range (50 < Re < 400, 35 < Re p < 225, 2 < Wo < 17). Key words – Pulsating flow; single-phase heat transfer enhancement; boundary layer redevelopment I. INTRODUCTION Heat exchangers with microchannels are targeted for high heat flux applications such as microelectronics cooling [1]. Research attention is divided between operation in single and two-phase flow. Due to the high boiling heat transfer rates, two-phase systems are considered the most promising technique for high-end microelectronics cooling [1]. Single phase flow in microchannels remains an active research area [2], also since micro scale two-phase flow is characterised by stability problems. In a microchannel (with hydraulic diameter of a few 100 μm) the Reynolds number typically ranges from 100 to 1000. In these conditions, nucleate boiling is the dominant heat transfer mode. This regime is characterised by a high wall superheat which causes rapid evaporation and bubble growth after nucleation. The sudden volume expansion disturbs the microchannel flow and may cause flow reversal [3]. Different regimes have been identified in two-phase microchannel flow using water and other working media [4,5]. Depending on the heat and mass flux conditions, irregular transitions occur between single-phase liquid and two-phase liquid/vapour flow in various modes (e.g. bubble, slug, annular flow) and pure vapour flow (i.e. dry-out), causing sudden excessive peak wall temperatures [4]. These studies [3-5] illustrate the stability problems in twophase flow in microchannels, even when applying an ideal uniform heating load. In reality, high-end microelectronics are characterised by strongly non-uniform heating, which aggravates the instability and the risk for periodic dry-out and related damage to the electronics [6]. Although some researchers are striving to stabilise twophase operation using flow restrictions and artificial nucleation sites [7], alternatives are also being investigated to enhance the cooling performance of single-phase systems. When superimposing pulsation on a steady channel flow, the hydrodynamic and thermal boundary layers are affected, which in turn affects the overall convective heat transfer rate. Some analytical studies of laminar pulsating flow show a frequency-dependent influence on the heat transfer compared to steady flow, yet overall the effect on the average heat transfer rate is found to be negligible [8,9]. However, some numerical and experimental studies found enhancement factors of up to 11% for laminar and 9% for turbulent pulsating flow [10,11] in smooth channels. Some recent experimental heat and mass transfer studies using pulsating flow in channels with cross-stream ribbed walls report enhancement factors of 100% up to 250% compared to steady flow [12,13]. The enhancement is more pronounced in laminar compared to turbulent flow, and increases with Prandtl number [14]. Impinging jets are another configuration where the effect of flow pulsation on the heat transfer enhancement has been investigated. Using synthetic jets (zero net mass flux), heat transfer rates comparable to steady impinging jets have been obtained [15-17]. Given the encouraging findings in similar applications, this paper aims to determine experimentally the potential for heat transfer enhancement using pulsating flow in a microchannel heat sink in single-phase operation. This study uses a single rectangular channel to serve as a reference case for subsequent studies using pulsating flow in parallel microchannels. II. EXPERIMENTAL APPROACH Microchannel heat sink and flow loop This reference case heat sink contains a single rectangular channel milled in an aluminium base (Fig. 1). The channel is H = 1 mm deep, W = 16 mm wide and L = 32 mm long. The channel is covered by an aluminium plate with fluidic connections on either side (4 mm internal diameter). The heat sink is bonded with thermal paste (10 W/(mK)) to a copper block with embedded cartridge heater (up to 40 W/cm 2 ). Based on a thermocouple measurement on the block and heat sink cover, the channel wall temperature T w is estimated using a lumped resistance model. * Corresponding author: tel +32 16 322546, fax +32 16 322985, email: tim.persoons@mech.kuleuven.be Present address: Mechanical Engineering dept., Parsons Building, Trinity College, Dublin 2, Ireland (tim.persoons@tcd.ie) ©EDA Publishing/THERMINIC 2009 163 ISBN: 978-2-35500-010-2
7-9 October 2009, Leuven, Belgium (a) (b) (c) (b) (a) Fig. 1. Microchannel heat sink ((1) aluminium cover plate with fluidic connections, (2) aluminium channel plate, (3) insulated heater block). Pulsator Microchannel heat sink p,T p,T Fig. 3. Pulsator for generating pulsating flow component with adjustable amplitude and frequency; (a) manifolds with check valves, (b) two pumping chambers with opposite phase, (c) piezoceramic actuator disk. Cooler Mass flow meter Pump Fig. 2. Microchannel heat sink test facility with pulsating flow device. Inlet and outlets with pressure and temperature taps are connected to a flow loop, driven by a variable speed gear pump (Fig. 2). A Coriolis flow meter (Bronkhorst CORI- FLOW, 50 kg/h) is used as mass flow measurement. The water temperature is kept constant during testing using a 2 litre reservoir at atmospheric pressure and a fan-driven platefin heat exchanger as cooling section. Pulsating flow generator A pulsating velocity component of continuously adjustable amplitude and frequency is generated using an inline pulsator device. The pulsator consists of a miniature membrane pump with two pumping chambers in opposite phase (Fig. 3), driven by a piezoelectric bending disk actuator. Fast acting check valves ensure the oscillating flow component is restricted to the heat sink test section (see Fig. 2). Its operating range covers frequencies from 0 to over 1000 Hz, with maximum displacement of 10 mm 3 per stroke. At 50 Hz, this yields a pulsating velocity amplitude up to U p = 0.2 m/s (Re p ≅ 500) in this microchannel heat sink geometry. The pulsator does not yet contain a fast-response sensor to monitor the displaced volume or pressure. As such, the resulting pulsating velocity amplitude is estimated based on the applied actuator voltage, neglecting the dynamics of the inertia of accelerating fluid and pump shaft, membrane stiffness and viscous damping. A displacement sensor will be incorporated shortly however for the results in this paper, the values of Re p are likely to exhibit some overestimation at higher frequency. As such, the frequency range has been limited between 0 and 40 Hz. The Womersley number, a dimensionless frequency in pulsating channel flow, is defined as Wo = ½D h (2πf/ν) 1/2 , where f is the pulsation frequency and ν is the kinematic viscosity. The objective of the study is to match the Womersley number range of other studies [9-11], with a typical maximum of Wo = 10. Here, a frequency range of 0 to 40 Hz corresponds to a range in Wo between 0 and 17. III. DISCUSSION OF HEAT TRANSFER RESULTS The heat transfer is quantified by the heat transfer coefficient h defined as q/ΔT lm where q is the surface heat flux. The logarithmic mean temperature difference ΔT lm is defined as ( Tw −To) −( Tw −Ti) Δ Tlm = (1) log (( Tw −To) ( Tw −Ti) ) where T i and T o are the inlet and outlet mean fluid temperature respectively. The wall surface temperature T w is estimated based on a lumped resistance model as described above. The Nusselt number is defined as Nu = hD h /k, where k is the thermal conductivity of water at the mean fluid temperature. Steady flow Experimental heat transfer results for steady flow are presented in Fig. 4. In the Reynolds number range under investigation (50 < Re < 500), the heat sink length L (32 mm) is shorter than the thermal entrance length (L < 0.04 D h Re Pr). As such, Fig. 4 also presents some existing correlations established for thermally developing flow between parallel plates [18-20]. In Fig. 4, the labels Nu q and Nu T indicate whether the correlation is valid for constant wall heat flux or temperature respectively. The heat transfer results for this heat sink are considerably higher than the parallel plate correlations. This can be attributed to the collector geometry (see Fig. 1), which does not provide a smooth transition for the flow into the channel. Furthermore, the heat is supplied to only one side of the channel, although via conduction within the heat sink walls a fraction of the heat can also be transferred into the flow from the non-heated side. ©EDA Publishing/THERMINIC 2009 164 ISBN: 978-2-35500-010-2
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