Experimental investigation on the effect of aluminum foam on natural convection in horizontal channel heated below

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1 Experimental investigation on the effect of aluminum foam on natural convection in horizontal channel heated below ORONZIO MANCA, SERGIO NARDINI, BERNARDO BUONOMO, LORENZO MARINELLI, CLAUDIO MONTANIERO Dipartimento di Ingegneria Industriale e dell'informazione Seconda Università degli Studi di Napoli Via Roma 29, Aversa (CE) 81031, Italy ITALY oronzio.manca@unina2.it, sergio.nardini@unina2.it, bernardo.buonomo@unina2.it, lorenzo.marinelli@unina2.it, claudio.montaniero@gmail.com Abstract: - Natural convection of air in horizontal parallel plates without and with aluminum foam is experimentally and numerically investigated. The lower wall is heated at uniform heat flux and the upper parallel plate is not thermally insulated to the external ambient. The investigation is carried out by means of a visualization technique and heated wall temperature measurements. The investigated aluminum foam had 10 and 30 Pores per Inches (PPI). The considered configurations are compared and it is found that the presence of the aluminum foam determines a better thermal heat transfer with respect to the clean channel with a low emissivity whereas higher heated wall temperature values with respect to the clean channel with the heated lower plate at high emissivity. Flow visualization is carried out to detect the development and the shape of the main flow and the transitions along the channel. Results of the visualization allow the description of secondary motions inside the channel. Key-Words: - Natural convection, Aluminum foams, Horizontal channel, partially filled with porous foam 1 Introduction In this work natural convection in a horizontal channel partially filled with a porous medium and the lower wall heated at uniform heat flux is studied experimentally. Natural convection between two horizontal parallel plates gets a great attention for its importance in practical applications in various modern systems such as drying processes, fire safety research, electronic cooling, chemical vapor deposition, safety in nuclear reactors and solar energy systems as underlined, for example, in [1,2]. Flow motion is strongly affected by the location of the heated surface installed on the top or bottom wall of horizontal parallel plates [1-6]. Buoyancy force due to the heating of the lower cavity wall induces secondary flows hence the local heat transfer increases [2-4, 6]. In any cases, the main flow presents a "C" loop behavior very close to the flow inside open cavities and open ended cavities, as also observed in [7]. The main disadvantage in natural convection is related to low heat exchanged between fluid and walls and several techniques can be employed to enhance the heat transfer such as to enlarge the exchange surfaces. Porous media with high thermal conductivity, metal foams, have emerged as an effective method of heat transfer enhancement due to their large surface area to volume ratio and to intense mixing of the flow. In the recent past, metal foams have been applied to increase the heat transfer and several investigations have been carried out [8]. Natural convection in high porosity metal foams heated from below was studied numerically and experimentally in [9, 10]. A metal foam heat sink with rectangular section was placed in a large Plexiglas housing and a two dimensional model was assumed both for the experimental runs and for the numerical simulations [9]. Aluminum foam samples of different pore sizes (5 40 PPI) and porosities ( ) were employed to illustrate the effects of metal foam geometry on heat transfer. Significant enhancements in heat transfer were found from the use of metal foams. Metal foam disk was considered in [10] to evaluate the effect of natural convection on the effective thermal conductivity. Moreover, porous medium Rayleigh number cannot characterize natural convection alone when the Darcy number is relatively large. A numerical investigation on natural convection for the idealized ISBN:

2 regular metal blocks by using lattice Boltzmann method was accomplished in [11] to reveal the detailed flow and thermal behaviors at the pore level. Results showed that the heat transfer is enhanced with the increase of the pore density and it is weakened with the increase of the porosity. Experimental investigations on natural convection heat transfer in superposed metal foams with internal heat sources were reported in [12, 13]. Numerical results of laminar fully developed natural convection in an inclined channel partially filled with metal foam were studied numerically in [14]. An experimental investigation of air natural convection in horizontally-positioned copper metallic foams with open cells was performed in [15]. It was found that the porosity influence on the heat transfer performance is more remarkable when the pore density is higher and natural convection in the copper foam reduces its thermal resistance and enhances its heat transfer performance. The natural convection on metallic foam-sintered plate at different inclination angles was experimentally studied in [16]. The sintered foam surface with lower porosity and pore density was recommended for heat transfer enhancement. Studies on partially opened cavities filled with porous media are reported in [16-19]. It seems that natural convection in horizontal channel partially filled with metal foams has scarcely been investigated. Investigations on partially filled horizontal channels with porous media are reported in [12, 20]. In this work natural convection in a horizontal channel partially filled with a porous medium and the lower wall heated at uniform heat flux is experimentally studied. The investigation is carried out by means of a visualization technique and heated wall temperature measurements. 2 Experimental apparatus The investigated channel was made of two principal horizontal walls and two vertical side walls. The lower wall consisted of two 400x530 mm 2 sandwiched phenolic fiberboard plates. The upper wall was a plate of glass 4.0 mm thick. The side walls were made of Plexiglas rectangular rods, machined to an accuracy of ±0.03 mm. The plate spacing was measured to an accuracy of ±0.25 mm by a dial-gauge equipped caliper. The cavity has a length L=400 mm, width W=450 mm and spacing b=40 mm and was open to the ambient along the edges of dimensions Wxb, Fig.1. The cavity was partially filled by a porous medium lain on the lower wall. The porous medium was aluminum foam and it was placed over the heated lower wall. The porous plate had a thickness equal to 20 mm whereas the length and the width were the same of the channel. The investigated aluminum foam had 10 and 30 Pores per Inches (PPI) with porosity equal to 0.97 and 0.92 respectively. The lower wall was made of two plates. The plate facing the channel was 3.2 mm thick whereas the rear plate was 1.6 mm thick. Its back surface was coated with a 17.5 m thick copper layer, which was the heater. It was an electrical resistance obtained cutting the copper layer in a serpentine shape. Its runs were 19.6 mm wide with a gap of nearly 0.5 mm between each one, giving the heater a total length of 9.0 m. Its expected electrical resistance was 0.50 the wall was heated by passing a direct electrical current through the heater. In order to reduce conductive heat losses, a 150 mm Polystyrene block was affixed to the rear face of the lower wall. Comparisons with experimental data for a channel without porous medium were carried out for two values of the hemispherical surface emissivity of the lower wall, ε r = 0.05 or ε r =0.95. The lower emissivity, ε r = 0.05, which minimized radiation effects on heat transfer, was obtained by coating the plate facing the channel with a 35 m thick nickel plated copper Fig. 1, View of the test section layer. The highest one was obtained by spraying onto the wall surface a black varnish. The emissivity values were measured by means of radiometric direct measurements by an infrared camera FLIR SC3000. The narrow gaps between the runs, together with the relatively high thickness (4.8 mm) of the resulting low-conductive fiberglass were suitable to maintain a nearly uniform heat flux at the plate surface. Direct electrical current through the heaters was accomplished by using a Hewlett-Packard 6260B dc power supply. The electrical power supplied was ISBN:

3 evaluated by measuring the voltage drop across the plate and the current passing through it. An HP- 3465A digital multi-meter measured the voltage drop, while the current was calculated by the measured voltage drop across a reference resistance. To avoid electrical contact resistances, thick copper bars soldered both to the electric supply wire and to the ends of heater were bolted together. The dissipated heat flux was evaluated to an accuracy of ±2%. The entire apparatus was located within a room, sealed to eliminate extraneous air currents. Wall temperatures were measured by twelve 0.50 mm OD ungrounded iron-constantan thermocouples embedded in each fiberboard plate and in contact with the outer layer and located at longitudinal stations. Furthermore fifteen thermocouples were affixed to the rear surface of the plates and embedded in the Polystyrene to enable the evaluation of conductive heat losses. The ambient air temperature was measured by a shielded thermocouple placed near the leading edge of the channel. An Isotech instrument mod.938 ice point, with 50 channels and an accuracy of ±0.03 C, was used as a reference for thermocouple junctions. Their voltages were recorded to 1 V by a National Instruments SCXI module data acquisition system and a personal computer was used for the data collection and reduction. The acquired data were recorded and processed through the LabView 6.0 program. Calibration of the temperature measuring system showed an estimated precision of the thermocouple-readout system of ±0.1 C. Smoke for flow visualization was generated by burning incense sticks in a steel tube, connected to a compressor. The smoke was injected through a glass heat exchanger to reduce its temperature which, measured by a thermocouple, turned out to be close to that of the ambient air entering the cavity. Finally, the smoke was sent into a plenum and driven to the test section through a small slot situated under the leading edge of the bottom plate along the plate width. A sketch of the experimental smoke generation arrangement is reported in Fig. 2. Particularly, the longitudinal view of the arrangement for the air temperature measurements and of the visualization set-up is shown. Preliminary tests were carried out to determine the plenum location that does not interfere with the air flow at the inlet section. The visualization was made possible by means of a laser sheet, generated by a He-Ne laser source. The laser sheet was produced by passing the laser beam through a cylindrical lens to enlarge the beam as needed, Fig. 3. Small Plenum Plates Incense sticks Heat exchanger Compressed air Fig. 2, Sketch of the smoke generation arrangement adjustments were allowed by means of a micrometer screw system, in order to take photographs at different locations along the z axis. The still camera was a programmable Nikon D Data reduction The Rayleigh number is defined as: 4 g qcb Ra Gr Pr Pr 2 (1) k and qc is the average convective heat flux. L 1 qc qc( x) dx L (2) 0 The thermo-physical properties were evaluated at the reference temperature: T r TW To (3) 2 where T o is the ambient temperature and T W is the average temperature of the heated wall L 1 TW TW( x) dx (4) L 0 The local and average Nusselt numbers are defined, respectively, as h( x )b q(x) c b Nu( x ) k T (x) T k w 0 (5) The local convective heat flux was evaluated as follows ISBN:

4 Light sheet x y z Camera Cylindrical lens Laser source Fig. 3, Sketch of the visualization arrangement q ( x) q ( x) q ( x) q ( x) (6) c k r where q x) is the local heat flux due to Ohmic dissipation, which is assumed to be uniform, q k (x) denotes the local conduction heat losses from the plate and q r (x) is the local radiative heat flux from the plate. The local convective heat flux, q c (x), is not uniform because of radiation and conduction. For each run, the terms q k (x) were calculated by a numerical procedure, a three-dimensional distribution of the temperature being assumed in the polystyrene. The predicted temperatures in significant configurations of the system had been previously compared with those measured by thermocouples embedded in the polystyrene insulation and the agreement was very good, the maximum deviation being 3%. The q r (x) terms were calculated for each temperature distribution of the wall, ambient temperature and channel spacing, dividing each plate into sixteen strips along the length. The uncertainty in the calculated quantities was determined according to the standard single sample analysis recommended by Moffat [21]. The uncertainty of the channel Grashof number, local Nusselt number was 7% and 6%, respectively. 3 Model description and numerical procedure A preliminary simplified numerical simulation is accomplished in a two-dimensional, steady state and incompressible flow. The domain is made of a principal channel, partially filled with an aluminum foam plate. The principal channel is made up of a uniformly heated horizontal wall at uniform heat flux and a parallel glass plate located above which allows heat transfer with the external ambient. Two external reservoirs side the channel, eleven times the channel height, b, simulates the natural convection toward the external ambient. Viscous dissipation, heat generation and pressure work are all assumed to have negligible effect on the velocity and temperature fields, therefore they are neglected. All the thermophysical properties of the fluid and the solid matrix of the porous medium are assumed constant except for the variation in density of the air with temperature (Boussinesq approximation) giving rise to the buoyancy forces. The thermophysical properties of the fluid and the solid matrix of the porous medium are evaluated at the ambient temperature, T 0, which is equal to 300 K in all cases. In the porous medium region, the flow model, known as the Brinkman-Forchheimer-extended Darcy model, is used in the governing equations. It is assumed that fluid and solid phase of porous medium are in local thermal equilibrium. Governing equations for the considered problem are not given in the present paper but the reader can refer to [22]. The geometrical parameters in the simulations are the same of the experimental section along its longitudinal section. Fig. 4, Simplified 2D model 4 Results and Discussion 4.1 Experimental results In this paper, results are reported for a plate distance b=40 mm, Ohmic heat fluxes q =120 W/m 2 and horizontal channel with a metal foam plate with a thickness equal to 20 mm. The corresponding Rayleigh number values are 8.50x10 5. Temperature profiles for 120 W/m 2 are given, in Fig. 5, for the channel without and with metal foam plate. In the first case two emissivities are considered whereas with foams two PPI values are examined. It is observed that the value of temperature profiles are the lowest for the channel without aluminum foam with emissivity equal to ε r =0.95 whereas the highest values are attained in the channel without foam with ε r =0.05. ISBN:

5 T w - Tamb [ C] no foam no foam 10 ppi 30 ppi x [cm] Fig. 5, Wall temperature profiles for channel without foam, with two emissivity values 0.05 and 0.95, and with aluminum foam, with two PPI values 10 and 30 for a Ohmic heat flux equal to 120 W/m 2 mm, 100 mm and 200 mm, is reported in Fig. 7. At first examined transversal section, in Fig. 7a, some weak secondary flows on the heated lower plate are observed and a flow configuration similar to a thermal is in the central zone of the section. This secondary flow goes toward the upper plate and penetrating in the upper boundary layer it can determine a more chaotic flow along the upper plate. At x=100 mm, in Fig. 7b, secondary motions are developing and mushroom structures are present in the section along the heated plate. The boundary layer thickness on the unheated upper plate is less than the one in the transversal section at x=50 mm, showed in Fig. 7a. Mushrooms and some lateral cell vortices are observed in the transversal section at x=200 mm (Fig. 7c). In this section, secondary motions fill all the section and the main fluid flow has only a velocity component along the y-direction and it can be indicated as "inversion section" PPI 10PPI no foam no foam a) x=50 mm Nu x 10 b) x=100 mm x/b Fig. 6, Local Nusselt numbers profiles for channel without foam, with two emissivity values 0.05 and 0.95, and with aluminum foam, with two PPI values 10 and 30 for a Ohmic heat flux equal to 120 W/m2 Local Nusselt numbers are calculated along the x axis and are shown in Fig. 6, for the channel with and without metal foam plate. The same cases like the previous temperature profiles are considered. The lowest Nusselt numbers correspond to emissivity =0.05 and without foam. Growing the foam pores por inch increases convective heat transfer. The presence of aluminum foam leads to a reduction of radiative effect and, consequently, a wall temperature increase as greater as the PPI is. Flow visualization for the channel without foam and r =0.95, with Ohmic heat flux equal to 120 W/m 2, in the transversal section, yz plane at x=50 c) x=200 mm Fig. 7, Flow visualization in the yz plane at x=200 mm, for channel without foam The fluid flow in the channel with aluminum foam presents in all considered transversal sections vortex cells with longitudinal axis very similar to Rayleigh-Benard cells. In the section at x=50 mm, in Fig. 8a, the cells are developing and some structures seem to have a structure very close to mushrooms. At the other two sections the cells are completely developed, in Fig. 8b and 8c. In this case the penetration along the longitudinal section, at z=0 mm, is about equal to the distance between the upper surface of the porous plate and the upper wall of the channel, as showed in Fig. 8d. Moreover, along this longitudinal section vortex cell structures are noted. ISBN:

6 a) x=50 mm b) x=100 mm a) Velocity field, without foam c) x=200 mm d) z=0 mm Fig. 8, Flow visualization for channel with aluminum foam with 10 PPI: a), b) and c) transversal sections and d) longitudinal section. b) Velocity field,10 ppi foam 4.2 Numerical simulation The numerical model is solved using the FLUENT code [23]. The segregated solution method is chosen to solve the governing equations, which are implicitly linearized with respect to the dependent variable equation. The second-order upwind scheme is chosen for the unsteady energy and momentum equations. The semi implicit method for pressure-linked equations (SIMPLE) scheme is chosen to couple pressure and velocity. Computation starts with zero values of velocities and with pressure and temperature values equal to the ambient ones. The convergence criteria of 10 6 for the residuals of velocity components and of 10 8 for the residuals of the energy are assumed. The physical domain under investigation consists in horizontal channel with a metal foam plate with a thickness equal to 20 mm. Experimental analysis is compared with numerical simulation. In the velocity field for the case without porous medium the phenomenon of back flow occurs as it can observed in the Fig. 9a. The fluid is unable to across completely the longitudinal section of the channel but, as noted in the photos, it turns in the middle transversal section and comes back toward the external ambient. In the configurations with the foam of 10 PPI and 30 PPI, the phenomenon of back flow does not exist or is limited very close to the external openings and the fluid velocity presents several vortex cells because it warms and rises upward than it cools and then falls c) Velocity field,30 ppi foam Fig. 9, Velocity field back down, as shown in Fig. 9b and in Fig. 9c. It is also confirmed by the flow visualization given in Fig. 8d. The temperature fields confirm what has been said before. In the Fig. 10a, the isotherms show the presence of back flow inside the open cavity. This kind of phenomenon does not exist with the aluminum foam as indicated in the Fig. 10b and Fig. 10c. The presence of the porous medium determines a small penetration of the main flow inside the channel and the back flow happens very close to the openings whereas inside the channel vortex cell structures are noted and they are very similar to the Rayleigh-Benard cells. 5 Conclusion An experimental investigation on natural convection in air in horizontal channels, without and with aluminum foam, with lower wall heated at uniform heat flux was accomplished by means of wall temperature measurements and flow visualizations. ISBN:

7 a) Temperature field, for channel without foam Gr Grashof number, Eq.(1), dimensionless h heat transfer coefficient, Wm -2 K -1 k thermal conductivity, Wm -1 K -1 L length of the wall, m Pr Prandtl number, dimensionless q heat flux, Wm -2 Ra Rayleigh number, Eq. (1), dimensionless T temperature, K x coordinate along the length of the plates, m y coordinate along the height of the plates, m z coordinate along the width of the plates W width of the plate, m Greek symbols volumetric coefficient of expansion, K -1 kinematic viscosity, m 2 s -1 b) Temperature field, for channel with 10 ppi foam Subscripts c convective k conductive o ambient air r radiative, reference w wall Ohmic dissipation c) Temperature field, for channel with 30 ppi foam Fig. 10, Temperature field A comparison between the two channel configurations was performed and the presence of the aluminum foam determined a better thermal heat transfer with respect to the clean channel with a low emissivity (ε r = 0.05) whereas presented higher wall temperature values with respect to the clean channel with the heated lower plate with high emissivity (ε r = 0.95). The flow development and the shape of flow transitions along the channel were visualized. Flow visualizations allowed descriptions of secondary motions inside the channel. The presence of the porous plate determined a small penetration of the main flow and vortex cell structures, very similar to the Rayleigh-Benard cells, both in the longitudinal and transversal sections. The comparison between the experimental flow visualization with the results of a simplified twodimensional model showed a good agreement. Nomenclature: a thermal diffusivity, m s -1 b plate spacing, m g acceleration due to gravity, ms -2 References: [1] O. Manca and S. Nardini, Experimental Investigation of Radiation Effects on Natural Convection in Horizontal Channels Heated From Above, J. Heat Transfer, vol. 131, 2009, art. no [2] W.-S. Fu, W.-H. Wang and S.-H. Huang, An Investigation of Natural Convection of Three Dimensional Horizontal Parallel Plates from a Steady to an Unsteady Situation by a CUDA Computation Platform, Int. J. Heat Mass Transfer, vol. 55, 2012, pp [3] O. Manca, B. Morrone and S. Nardini, Experimental analysis of thermal instability in natural convection between horizontal parallel plates uniformly heated, J. Heat Transfer, vol. 122, 2000, pp [4] H. Yang, Z. Zhu and J. Gilleard, Numerical Simulation of Thermal Fluid Instability between Two Horizontal Parallel Plates, Int. J. Heat Mass Transfer, vol. 44, 2001, pp [5] O. Manca and S. Nardini, Experimental investigation on natural convection in horizontal channels with the upper wall at uniform heat flux, Int. J. Heat Mass Transfer, vol. 50, 2007, pp ISBN:

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