Experimental Investigation of Heat Transfer in Laser Sintered and Wire Mesh Heat Exchangers

Transcription

1 Experimental Investigation of Heat Transfer in Laser Sintered and Wire Mesh Heat Exchangers by Reza Rezaey A thesis submitted in conformity with the requirements for the degree of Doctor of Philosophy Department of Mechanical and Industrial Engineering University of Toronto Copyright by Reza Rezaey 2017

2 Experimental Investigation of Heat Transfer in Laser Sintered and Wire-Mesh Heat Exchangers Reza Rezaey Doctor of Philosophy Department of Mechanical and Industrial Engineering University of Toronto 2017 Abstract In this thesis, an experimental investigation of fluid flow and heat transfer through open cell porous wire mesh and laser-sintered heat exchangers is presented. The thesis consists of two main sections that describe how to create a compact heat exchanger that uses open-cell porous structures. In the first part of the thesis a new method of building compact heat exchangers using direct metal laser sintering (DMLS), a technology which enables heat exchangers with a predetermined, fully controlled internal geometry to be built was investigated. Laser-sintering was used to fabricate stainless steel heat exchanger channels filled with struts arranged to form either cubic, round-strut tetradecahedral or thin-strut tetradecahedral cells. The objective was to demonstrate that the effect of adding internal struts is not simply to increase surface area, but that cell geometry has a significant effect on both heat transfer and fluid flow. This section also describes the importance of the connection between the porous structures, which is used to improve the performance of the heat exchanger, to the main body of the heat exchanger. It was possible to design internal geometries that maximize heat transfer while minimizing weight and frictional losses. ii

3 In the second part of the thesis, a simple method of increasing the heat transfer surface area has been developed by using a twin wire-arc thermal spray system to generate a dense, high strength coating that bonds porous structures, like wire mesh and perforated sheets, to the plain tube heat exchanger s outside surfaces. The porous structure and the main body of the heat exchanger must be well bonded together to minimize thermal resistance. The extended surfaces of the wire mesh and perforated sheet enhanced the heat transfer performance of the tube heat exchangers. Finding the right balance between pore density and number of screens of the porous structures is crucial for maximizing the heat transfer performance of the heat exchangers. iii

4 Acknowledgments I would like to thank my wonderful supervisor, Professor Sanjeev Chandra, for his time and nonstop support during the course of this research. It was an honour for me to work under his supervision and guidance during the last four years. I also want to thank Professor Javad Mostaghimi, Doctor Larry Pershin and Professor Thomas Coyle at the Center for Advanced Coating Technologies (CACT), at the University of Toronto, for providing valuable guidance in every aspect of this research. Also, my special thanks go to my lab mates and friends at CACT, Mehrdad Taheri, Saeid Salavati, Bharath Krishnan, Christiane Mubikayi, and all other lab mates. I want to thank my colleague Mr. Felix Loosmann and his supervisor Professor Cameron Tropea at Technische Universirat Dramstadt for the fabrication of laser-sintered prototypes. Last but not least, I would like to thank my father, Masieh, for his careful guidance, my mother, Nasrin, for her invaluable support and my brother, Mojtaba, for always being so supportive and encouraging. Finally, I would like to appreciate the endless patience and constant support of my beloved wife Newsha. Your encouragements in the toughest times, positive attitude and beautiful smile gave me the strength to finish my studies. iv

5 Table of Contents Acknowledgments... iv Table of Contents...v List of Tables... viii List of Figures... ix List of Appendices... xvii Chapter 1 Introduction Introduction Literature Review Objectives Organization of Thesis...7 DMLS Heat Exchangers Introduction Geometric Characteristics Fabrication of DMLS Heat Exchangers Fabrication of Heat Exchanger Channels Material Properties...19 Conduction Heat Transfer in DMLS Heat Exchangers Test Samples, Experimental Apparatus and Results Heat Transfer Characteristics Theory Analytical Models Analysis and Discussion Conclusion...46 v

6 Convection Heat Transfer in DMLS Heat Exchangers Experimental Apparatus Hydraulic Characteristics Heat Transfer Characteristics Conclusion...70 Wire-Arc Thermal Sprayed Heat Exchangers Introduction Geometric Characteristic Fabrication of Wire-Arc Thermal Sprayed Heat Exchangers...75 Preliminary Investigation of Flow over Perforated Sheet and Wire Mesh Fins Introduction Fabrication of Wire-Arc Thermal Sprayed Fins Experimental Apparatus and Methods Results and Discussion Plain Tube Perforated Sheet Fins Wire Mesh Fins Heat Transfer Characterization Conclusion Water-to-Air Wire Mesh Heat Exchangers Introduction Fabricated Heat Exchangers Wire Mesh Fabrication Process Experimental Apparatus and Methods vi

7 7.4 Pressure Drop Through Wire Mesh Screens Results and Discussion Heat Transfer Characterization Non-Dimensional Parameters Empirical Fin Model Correlation A Model for Prediction of Heat Exchanger Temperature Rise Conclusion Air-To-Air Wire Mesh Heat exchangers Introduction Heat Exchanger Design Manufacturing of the Heat Exchanger Experimental Apparatus Heat Transfer Calculation Results and Discussion Heat Transfer Characterization Conclusion Summary Laser Sintered Heat Exchangers Wire Mesh Heat Exchangers References Appendices vii

8 List of Tables Table 4-1: Structural comparison between the cubic, round-strut tetradecahedral and thin-strut tetradecahedral heat exchanger channels Table 5-1: Porosity, oxide content, and adhesion strength of the coatings sprayed under different conditions [6] Table 5-2: Wire-arc thermal spray parameters for deposition of stainless steel coating [6] Table 6-1: Perforated sheet specifications Table 6-2: Wire mesh fin specifications Table 6-3: Summary of the porous structures used in the study Table 6-4: Comparison between the variation of NuD and Anon-perf Table 7-1: Parameters of the wire mesh heat exchangers Table 7-2: Parameters of the wire mesh heat exchangers at a water mass flow rate of Kg/s Table 7-3: Parameters of the wire mesh heat exchangers at a water mass flow rate of Kg/s Table 8-1: Cold air velocities inside the wind tunnel viii

9 List of Figures Figure 2-1: Unit cell geometry (a) cubic, and (b) tetradecahedral Figure 2-2 : Unit cell geometry of the thin-strut tetradecahedral geometry Figure 2-3: Unit cell geometry (a) cubic, (b) round-strut tetradecahedral and (c) thin-strut tetradecahedral Figure 2-4: Schematic of DLMS manufacturing procedure [38] Figure 2-5: Channel section (a) cubic, (b) round-strut tetradecahedral, and (c) thin-strut tetradecahedral Figure 2-6: Heat exchanger channels (a) end view of cubic channel (b) end view of round-strut tetradecahedral channel, and (c) thin-strut tetradecahedral Figure 2-7: Round-strut tetradecahedral channel with side face removed Figure 2-8: Complete assembly of the heat exchanger with four sections welded together Figure 2-9: SEM images of (a) channel wall surface, and (b) cross section of the channel wall. 20 Figure 2-10: Connection between the strut and the channel wall of the heat exchanger Figure 2-11: EDS analysis of the coating micro structure at the connection point between the struts and the wall Figure 3-1: Tetradecahedral structure (a) one partially removed wall, and (b) without walls Figure 3-2: Schematic overview over the four different experimental setups Figure 3-3: Experimental apparatus to determine the thermal conductivity of laser-sintered stainless steel (a) a block manufactured with common methods on the left side and a block sintered using DMLS on the right side, and (b) a schematic of the experimental setup Figure 3-4: Temperature distribution over sample length for the first set of experiments ix

10 Figure 3-5: Experimental apparatus, which is used for the third and fourth set of experiments, to measure the temperature distribution for different heat fluxes with cooling at one side of the samples and heating on the opposite site, respectively Figure 3-6: Comparison of experimental results for the outer surface temperature distribution over relative location between cubic and round-strut tetradecahedral sample at a three different heat fluxes. Heating from one side, cooling from the other, results Figure 3-7: Comparison between heat transfer in samples (a) with zero heat loss to surrounding, and (b) with heat loss to the surroundings Figure 3-8: Schematics for sample segmentations for heat loss calculation Figure 3-9: Temperature distribution over sample length for the third set of experiments Figure 3-10: Thermal conductivity of the laser-sintered stainless steel block over applied heat flux for the first set of experiments. Data sheet values are taken from the EOS Stainless Steel 17-4 data sheet [37] Figure 3-11: Comparison of the effective thermal conductivity over applied heat flux for the cubic and round-strut tetradecahedral heat exchangers, and the round-strut tetradecahedral structure without walls Figure 3-12: Heat transfer direction through (a) Parallel Model, and (b) Series Model Figure 3-13: Comparison of the effective thermal conductivity for different porosities between predictions of various analytical models Figure 3-14: Comparison of the effective thermal conductivity for different porosities between predictions of various analytical models and the experimental results Figure 4-1: Schematic of experimental apparatus Figure 4-2: Variation of experimentally measured pressure gradient with average fluid velocity in channels with cubic, round-strut tetradecahedral and thin-strut tetradecahedral cells x

11 Figure 4-3: Friction factor variation with Reynolds number for channels with cubic, round-strut tetradecahedral and thin-strut tetradecahedral Figure 4-4: Increase in air temperature from the inlet to the outlet of a) cubic b) round-strut tetradecahedral, and c) thin-strut tetradecahedral heat exchangers with air flow rates varying from 10 to 90 L/min for applied heat flux in the range of 3.2 to 0.8 kw/m Figure 4-5: Rate of heat transfer to air flowing through cubic, round-strut tetradecahedral and thin-strut tetradecahedral heat exchangers with varying air flow rate and total heater power of 0.8 and 2.3 kw/m 2. The horizontal lines mark the total heater power of 0.8 and 2.3 kw/m Figure 4-6: Temperature variation across exit of cubic heat exchanger for constant applied heat flux of 2.3 kw/m 2 and air flow rate (a) 20 L/min, (b) 40 L/min (c) 60 L/min, and (d) 80 L/min. Temperature scales are in C Figure 4-7: Temperature variation across exit of round-strut tetradecahedral heat exchanger for constant applied heat flux of 2.3 kw/m 2 and air flow rate (a) 20 L/min, (b) 40 L/min, (c) 60 L/min, and (d) 80 L/min. Temperature scales are in C Figure 4-8: Temperature variation across exit of thin-strut tetradecahedral heat exchanger for constant applied heat flux of 2.3 kw/m 2 and air flow rate (a) 20 L/min, (b) 40 L/min, (c) 60 L/min, and (d) 80 L/min. Temperature scales are in C Figure 4-9: Variation of heat exchanger efficiency for cubic, round-strut tetradecahedral and thin-strut tetradecahedral channels with increasing Peclet number Figure 4-10: Measured wall temperature and calculated air temperature variation along the length of (a) the cubic, (b) the round-strut tetradecahedral, and (c) the thin-strut tetradecahedral heat exchanger for an applied heat flux of 2.3 kw/m 2 and air flow rates of 20 and 80 L/min Figure 4-11: Local heat transfer coefficient variation along the length of (a) the cubic, (b) the round-strut tetradecahedral, (c) the thin-strut tetradecahedral and (d) a hollow channel for 2.3 kw/m 2 heat flux xi

12 Figure 4-12: Average Nusselt number (NuH) as a function of Reynolds number (ReH) for roundstrut, cubic structure, thin-strut tetradecahedral and empty channels Figure 5-1: Unsuccessful welding of tube to the wire mesh Figure 5-2: Thermal skin deposition using wire-arc spray technique Figure 5-3: Woven copper wire mesh screens of (a) 10 PPI, and (b) 40 PPI Figure 5-4: Backscattered electron SEM images of stainless coatings deposited at spray distances of (a) 100 mm, (b) 150 mm, and (c) 200 mm [6] Figure 5-5: SEM micrograph of coated joint [6] Figure 5-6: SEM image of gap in the wire-tube joint filled by the coating material [6] Figure 6-1: Heat Transfer performance charts of different heat dissipation media [14] Figure 6-2: Heat transfer performance charts [43] Figure 6-3: Fabricated fins after thermal spray coating of aluminum on (a) perforated sheet, and (b) wire mesh Figure 6-4: Schematic diagram of the experimental setup Figure 6-5: Fabricated fins after sprayed using high emissivity black paint on (a) flat plate, and (b) perforated sheet (Ø= in (4.75 mm)) Figure 6-6: Temperature variation of the pipe at 15 V and 20 V (corresponding to surface heat fluxes of 1.3 kw/m 2 and 2.3 kw/m 2 ) applied voltage for different air velocities Figure 6-7: Comparison between the variation of (NuD) with (ReD) for experimental and theoretical model for flow over a cylinder Figure 6-8: IR map of the temperature distribution of the fins (a) Flat plate, and (b) Perforated sheet (Ø= in (4.75 mm)) xii

13 Figure 6-9: Comparison of the temperature profile at 55 V (17.7 kw/m 2 ) with a 10 m/s flow between the flat plate and perforated sheet (Ø= in (4.75 mm)) Figure 6-10: Comparison between the measured surface temperature and predicted theoretical model Figure 6-11: The temperature profile at 60 V (21.1 kw/m 2 ) applied voltage and for three air velocities for perforated sheet (Ø= in (4.75 mm)) Figure 6-12: The temperature profile at 10 m/s air velocity and three different applied voltages for the perforated sheet (Ø= in (4.75 mm)) Figure 6-13: Perforated sheet tested at 55V (17.7 kw/m 2 ) with a 10 m/s flow (a) Ø= in (4.76 mm), (b) Ø= in (3.17 mm), and (c) Ø= in (1.59 mm) Figure 6-14: Temperature profile of the perforated fins at 55 V (17.7 kw/m 2 ) applied voltage and 10 m/s air velocity Figure 6-15: Experimental temperature distribution at 55V (17.7 kw/m 2 ) applied power with a 10 m/s air velocity (a) 10 PPI, (b) 14 PPI, and (c) 20 PPI Figure 6-16: Temperature profile for different wire mesh at 55 V (17.7 kw/m 2 ) applied power and a 10 m/s air velocity Figure 6-17: Temperature profile of 14 PPI at 60 V (21.1 kw/m 2 ) applied power for different air velocities Figure 6-18: Temperature profile comparison between the wire mesh and perforated sheets at 55V (17.7 Kw/m 2 ) applied power with a 10 m/s air velocity Figure 6-19: Variation of Nusselt number (NuD) as a function of Reynolds number (ReD) based on tube outer diameter (OD) for wire mesh and perforated sheet Figure 6-20: Performance chart of the fabricated fins at a constant (ReD) of Figure 6-21: Nusselt number (NuH) variation as a function of Reynolds number (ReH) xiii

14 Figure 6-22: Comparison between the fin efficiency (ɳ) and effectiveness (ɛ) of the fins Figure 7-1: Sample of heat exchangers (a) single screen 5 PPI wire mesh, (b) single screens 10 PPI wire mesh, and (c) single screens 20 PPI wire mesh Figure 7-2: Sample heat exchangers (a) single screens 5 PPI wire mesh, and (b) double screens 5 PPI wire mesh Figure 7-3: Schematic representation of the experimental setup Figure 7-4: Schematic representation of the hot air chamber Figure 7-5: Variation of experimentally measured pressure gradient with average fluid velocity in channels for 20 PPI and 10 PPI wire mesh screen Figure 7-6: Temperature rise of water flowing through the tubes (a) heat exchangers with one wire mesh screen, and (b) heat exchangers with two wire mesh screens Figure 7-7: Variation of average air temperature at section 3 for different PPI wire mesh heat exchangers Figure 7-8: Variation of the Overall heat transfer coefficient across different pore densities Figure 7-9: Nusselt number variation (Nua,D) across different pore densities at a constant water mass flow rate of Kg/s Figure 7-10: Nusselt number (NuH) variation as a function of Reynolds number (ReH) Figure 7-11: IR camera surface temperature variation across heat exchangers for (a) 5 PPI, (b) 10 PPI and, (c) 20 PPI Figure 7-12: Schematic of eleven transverse and one longitudinal wire between two tubes Figure 7-13: Wire surface temperature variation along the length of one longitudinal and eleven transverse wires, measured experimentally using IR camera, for the 5 PPI wire mesh heat exchanger. The x and y axis are shown Figure xiv

15 Figure 7-14: Wire surface temperature variation along the length of a longitudinal and six transverse wires (T1, T2, T3, T4, T5, and T6 as shown in Figure 7-12) for the 5 PPI wire mesh heat exchanger. Temperatures were measured experimentally using IR camera Figure 7-15: Location of longitudinal and transverse wires of wire mesh screens Figure 7-16: Comparison between the measured surface temperature using IR camera and predicted empirical model Figure 7-17: Heat transfer energy balance for the fabricated heat exchangers Figure 7-18: Schematic of 3 heat exchangers connected in series Figure 7-19: Extended surface area ratio (RA) variation as a function of NTU Figure 8-1: Full assembly of a heat exchanger on top of the gas flare Figure 8-2: Assembly process for the heat exchanger Figure 8-3: Fabricated bare tube section of the main heat exchanger Figure 8-4: Fabricated section of the main heat exchanger, with one wire mesh screen attached on front and back side of the tubes Figure 8-5: Wire mesh section after thermal skin deposition of stainless steel using wire-arc Figure 8-6: Thermal sprayed surface of the wire mesh and the tube Figure 8-7: Mechanical bonding of 4 PPI wire mesh to the stainless steel tube [6] Figure 8-8: Front view of the fabricated heat exchanger before welding the manifolds Figure 8-9: Back view of the fabricated heat exchanger after the final assembly Figure 8-10: Schematic representation of the experimental setup Figure 8-11: Temperature drop for different hot air flow rates at a constant cold air velocity of 5.4 m/s xv

16 Figure 8-12: Heat transfer enchantment of the wire mesh sections compare to the plain tube Figure 8-13: Nusselt number (NuH) variation as a function of Reynolds number (ReH) Figure 9-1: Nusselt number (NuD) variation as a function of Reynolds number (ReH) xvi

17 List of Appendices Appendix A: Matlab Code for the Empirical Fin Model Appendix B: Heat Exchanger Assembly for the Hot Gas Incinerator Appendix C: Step-by-Step Fabrication Process of the Heat Exchanger Appendix D: Location of the Thermocouples on the Surface of the Heat Exchanger Appendix E: Shows a Schematic of the Fan, the Fan Performance and the Electrical Heater xvii

18 1 Chapter 1 Introduction 1.1 Introduction Heat exchangers have been used for many years to transfer heat between different fluid streams. Depending on the application, their performance can be improved by adding solid fins with different geometries on their heat-conducting surface to increase the surface area between the fluid media. There have been many attempts to optimize the shape of fins but their heat transfer and hydraulic performance is limited by the total surface area that can be obtained in a given volume. Open-cell porous structures such as metallic foams and wire mesh have a large surface area to volume ratio, and have been studied extensively for heat exchanger applications. However, one of the problems in making compact heat exchangers from open-cell porous structures is that the porous material and the main body of the heat exchanger must be well bonded together to minimize thermal resistance, which can be a difficult task. In recent years, new methods of building compact heat exchangers from porous metal foams, using technologies such as thermal spray coating, have been investigated. Thermal spray coating offers a convenient method of bonding porous materials to metal sheets and tubes, which can be used to make novel heat exchanger designs. Direct metal laser sintering (DMLS) is a rapid manufacturing technology that can be used for both prototyping and mass production, which offers the possibility of making structures with a predetermined, fully controlled internal geometry. This thesis will explore the application of both of these methods to the fabrication of heat exchangers.

19 2 1.2 Literature Review Heat exchangers are ubiquitous in industry, used wherever energy is to be transferred from a high temperature fluid stream to another at lower temperature. There is an enormous body of literature dealing with analysis of heat exchangers, but typically a designer wants to minimize both the size of the heat exchanger and the work required to pump fluid through it. One method of reducing the external dimensions of a heat exchanger is to increase the internal surface area wetted by the fluid across which heat transfer occurs. Louvered fins, wire mesh, and other open-cell structures all serve to increase the surface-area-to-volume ratio [1-4]. Heat transfer is further enhanced by turbulence, induced by the complex flow path through small passages [5]. The choice of porous structures placed in the interior of heat exchangers to enhance heat transfer is usually limited by what can be readily fabricated. Wire mesh has therefore been a favorite option [6], since it is available in a wide variety of sizes and materials. Metal foams have attracted much attention in recent years, as they are now being manufactured in commercial quantities and have been shown to enhance heat transfer significantly [5, 7]. Salavati et al [7] fabricated open pore metallic foam core sandwich structures prepared by thermal spraying of a coating on the foams that can be used as high efficiency heat exchangers due to their high surface area to volume ratio and consequent high heat transfer. Lu et al. [8] reviewed the thermal characteristics of metallic sandwich structures with truss and prismatic cores used to cool the wall of a heated channel. They combined data showing the influence of topology on the Nusselt number, Reynolds number and friction factor. Sypeck [9, 10] studied metallic sandwich structures with truss cores and fabricated structures from perforated aluminum alloy sheets, connecting the outer wall to the wrought metals by brazing in a vacuum

20 3 furnace. In an investigation by Boomsma et al. [5], metal sheets were brazed to the surfaces of metal foams to create heat exchangers. Salimi Jazi et al. [11] and Tsolas [12] fabricated heat exchangers by using a wire-arc spray method to deposit an Inconel 625 skin on copper and nickel foams and measured the convection heat transfer rate. However, these porous structures are not specifically designed to maximize heat transfer or to minimize pressure losses they are used because they are readily available. Khayargoli et al. [13] investigated the effect of the microstructure of nickel and nickel-chromium alloys metal foams on flow parameters. They found that the permeability increases as the pore size increases which was due to increases in drag forces on the flowing fluid. Tian et al. [14] studied fluid flow and heat-transfer during forced convection through cellular copper lattice structures. To find the maximum heat transfer performance of the woven copper mesh they tested several configurations. They discovered that unlike open-cell metal foams and packed beds, the friction factor of the bonded wire screen, apart from being a function of porosity, is also a function of orientation. They concluded that wire-screen mesh competes favorably with the best available heat dissipation media. The overall thermal efficiency index of the copper textiles-based media (mesh) was found to be approximately 3 times higher than that of copper foam due to the high pressure drop of the copper foam. Assaad et al. [15] created a new class of heat exchangers using wire mesh. They stacked and sintered stainless steel woven wire mesh together and created wire mesh bricks. They machined the bricks and cut them into thin wafers that could be combined to create porous structures. In order to contain the working fluid inside this porous structure they deposited metal coatings on the outer surface of the wafers using pulsed gas dynamic spraying (PGDS). They claimed, based on

21 4 their burst and tensile tests, that the fabricated compact wire mesh heat exchanger could withstand internal pressure as high as 19.1 MPa. Joen et al. [16] fabricated a single row heat exchanger consisting of aluminum metal foam covered aluminum tubes. They placed their samples inside a wind tunnel and tested various parameters including Reynolds number, tube spacing, foam height and the type of foam. They discovered that increasing the foam height reduces the exterior convection resistance while increasing the pressure drop. They also tested brazed and unbrazed samples which proved the importance of bonding and concluded that more research is needed to develop efficient and cost-effective connection (brazing) techniques to better connect the tube to the foam to provide solid metallic bonds. Another factor impacting the performance of heat exchangers is the effective thermal conductivity of both the heat exchanger structure and the fluid that flows through it. In analysis it is often convenient to consider the solid and fluid as being one composite material in order to derive the effective thermal conductivity of a heat exchanger. A low effective thermal conductivity in the flow direction of the forced convection heat exchanger is desired, so that the applied heat is transported mainly by convection and not by conduction. Researchers have developed analytical models to predict the effective thermal conductivity of composite materials [4, 17, 18]. Zhao, Lu, Hodson and Jackson [19] examined the temperature dependence of effective thermal conductivity of steel alloy foams for temperatures between K, under both vacuum and atmospheric condition. They discovered that the transport of heat is dominated by thermal radiation and effective thermal conductivity increase at high temperatures. They also compared the effective thermal conductivity calculated at pressure varying from atmospheric to vacuum conditions and established the importance of natural convection since the effective thermal conductivity at atmospheric pressure was twice that in a vacuum.

22 5 Paek et al. [20] experimentally investigated thermo physical properties of different porosity aluminum foams. They measured the effective thermal conductivity and the permeability of the foam and found that effective thermal conductivity increases as the porosity decrease. Also, at a fixed porosity, as the surface area in a given volume increases, flow resistance and pressure drop increase due to a decrease of permeability. They correlated the friction factor with the permeability based Reynolds number. Open-cell porous materials used for heat transfer purposes must to be bonded to the external shell of the heat exchanger in a manner that minimizes thermal resistance. Methods such as cladding, welding, brazing, diffusion bonding and thermal spray coating have been used to connect an opencell structure to the body of the heat exchanger containing the flowing fluid [7-12], but these all add to the complexity of manufacturing. In recent years, rapid manufacturing techniques have given engineers the ability to make extremely complicated structures using additive techniques in which three-dimensional objects are made in a single step directly from computer-based designs. This offers the possibility of making heat exchangers with any arbitrary internal shapes: it may be possible to optimize the shape of passages for fluid flow to maximize heat transfer while reducing pressure losses.

23 6 1.3 Objectives This thesis investigates new methods of building compact heat exchangers, using either direct metal laser sintering (DMLS) to make channels with internal structures, or thermal spray coating to bond wire mesh to the outside surface of tubes. This thesis aims to investigate the heat transfer through open-cell spray coated and laser-sintered heat exchangers. The specific objectives to be achieved are: Fabricate channels with internal open-cell geometries using DMLS technology. Study conduction heat transfer through DMLS porous structures. Experimentally investigate the impact of internal cell geometry on pressure drop and forced convection heat transfer to air flowing through DMLS heat exchangers. Fabricate heat exchangers using thermal spraying to bond wire mesh screens or perforated metal sheets to the outer surface of the tubes. Model and compare the heat transfer enhancement for different wire mesh and perforated sheet sizes, varying their pore density, geometry and orientation. Fabricate an industrial size wire mesh heat exchanger and compare its performance to a conventional plain tube heat exchanger.

24 7 1.4 Organization of Thesis The first chapter starts with a general introduction to this research followed by a literature review of porous heat exchangers. Chapter 2 introduces the direct metal laser sintering (DMLS) fabrication process and the geometry of the porous heat exchangers that were studied. It describes in detail the manufacturing process and the material properties of the heat exchangers. Chapter 3 explains the theory behind conduction heat transfer in porous structures. In this chapter the effect of conduction for different DMLS heat exchangers was studied, and compared to analytical models. Chapter 4 analyzes convection heat transfer for different DMLS heat exchangers. Convection heat transfer coefficient; are calculated and Nusselt number correlations developed. The effect of pore geometry on hydraulic and heat transfer performance is discussed. Three different geometries were analyzed to maximize heat transfer while minimizing pressure drop. Chapter 5 describes the fabrication process of wire mesh heat exchangers. The wire-arc thermal spray process was used to provide an intimate bond between wire mesh and tubes to form water-air heat exchangers. The mechanical and material properties of the thermally sprayed wire mesh heat exchangers are described in this chapter. Chapter 6 describes laboratory experiments that contributed to the understanding of the heat transfer characteristics of perforated sheets and wire mesh sheets bonded to heated tubes that were tested inside a wind tunnel at different air velocities. The temperature distribution across the mesh or sheet was measured using an infrared camera.

25 8 Chapter 7 describes the laboratory-scale water-to-air heat exchangers that were fabricated using wire-arc thermal spray coating. Different pore densities of wire mesh were examined and the temperature rise of water flowing through the tubes, while hot air passed over them, was measured. Nusselt number correlations were developed for each heat exchanger. Chapter 8 describes the process of fabricating a large thermally sprayed air-to-air heat exchanger suitable for high temperature applications. The heat transfer enhancement due to addition of a wire mesh was measured experimentally.

26 9 DMLS Heat Exchangers 2.1 Introduction In the present study laser-sintering was used to fabricate stainless steel heat exchanger channels filled with thin struts arranged to form either cubic or tetradecahedral. When a given space is filled with identically shaped cells of equal volume, tetradecahedral cells (which have 14 faces, 6 square and 8 hexagonal) are known to have the least surface area separating them, according to the wellknown Kelvin Conjecture [21]. Foams made by blowing gas into a liquid, contain tetradecahedral bubbles, since surface tension minimizes their internal surface area. Heat exchanger channels with tetradecahedral structures, therefore, have a much lower surface area than those with cubic cells. Heat transfer and frictional drag forces increase approximately linearly with the area of contact between a liquid and solid surface, all else remaining constant. The tetradecahedral structure (similar to a metal foam) would be expected to have lower heat transfer efficiency than the cubic structure (which resembles a wire mesh) if the shape of the voids does not change fluid flow significantly. In the present study, the question is addressed whether heat transfer and drag force varied proportionally to the wetted area, or whether the cellular tetradecahedral structures altered the fluid flow in such a manner to have a significant effect on the heat exchanger efficiency. If the latter is true, it may be possible to design internal geometries that maximize heat transfer while minimizing weight and frictional losses. The heat exchangers examined in this study were square cross-section channels with either cubic or tetradecahedral inner structure, with a uniform heat flux applied to the outer channel walls. The

27 10 increase of flow temperature was used to calculate friction and convective heat transfer coefficients at varying airflow rates. The results for channels containing either cubic or tetradecahedral cells were compared with those for a hollow channel. The effect of varying the strut shape for tetradecahedral cells was studied. The objective was to demonstrate that the effect of adding internal struts is not simply to increase surface area, but that cell geometry has a significant effect on both heat transfer and fluid flow. Laser-sintered prototypes were fabricated by my colleague Mr. Felix Loosmann and his supervisor Professor Cameron Tropea at Technische Universirat Dramstadt in Germany.

28 Geometric Characteristics A porous structure is characterized by several parameters, including porosity, pore density, pore size and strut diameter. Porosity (ε) is defined as the void volume in the porous sample divided by its total volume. As the porosity of a sample increases, the amount of solid material of that sample decreases. Pore density is determined by counting the number of pores crossed by a randomly drawn line and measured in pores per inch (PPI). The dimensions of a pore are specified by the pore size (dp), which defines the size of a unit cell, and strut diameter (df). (a) (b) Figure 2-1: Unit cell geometry (a) cubic, and (b) tetradecahedral. Figure 2-1 shows the two unit cell geometries that were used to produce the 10 PPI cubic (Figure 2-1a) and tetradecahedral (Figure 2-1b) heat exchanger prototypes respectively. Both geometries have a strut diameter (df) of 1 mm. In order to construct a 10 PPI cubic heat exchanger, the distance between the centerlines of two adjacent struts was set to 2.54 mm (0.1 in). Another way of constructing a tetradecahedral structure is to subtract a sphere from a 14-sided block of material.

29 12 The result of that subtraction was a tetradecahedral structure with a variation of the strut diameter. In addition, the struts were not cylindrical in shape and the shape of the unit cell was much closer to shapes found, for example, in alumina metal foams (Figure 2-2). Figure 2-2 : Unit cell geometry of the thin-strut tetradecahedral geometry. (a) (b) (c) Figure 2-3: Unit cell geometry (a) cubic, (b) round-strut tetradecahedral and (c) thin-strut tetradecahedral.

30 13 Figure 2-3 shows the unit cell geometries that were used to produce heat exchanger channels. They will be referred to as cubic (Figure 2-3a), round-strut tetradecahedral (Figure 2-3b) and thin-strut tetradecahedral (Figure 2-3c) heat exchangers. The thin-strut tetradecahedral geometry has a nonuniform strut diameter (df), roughly triangular, that resembles those found in metal foams. The mass of the thin-strut tetradecahedral structures is significantly lower than that of round-strut tetradecahedral structure.

31 Fabrication of DMLS Heat Exchangers In the present study, a new method of building compact heat exchangers, using direct metal laser sintering (DMLS), was investigated. This technology can be used for both prototyping and mass production and enables heat exchangers with a predetermined, fully controlled internal geometry to be built. In DMLS a 3D CAD model was created and imported into the laser-sintering machine. Figure 2-4: Schematic of DLMS manufacturing procedure [38].

32 15 The fabrication process, as shown in Figure 2-4, starts by first preheating the building chamber, after which the recoater blade moves metal powder from the dispensing platform onto the building platform. Next, a laser beam melts the powder at the places where solid sections are desired by the CAD model, before the recoater blade moves a new material layer onto the building platform. Once all the layers are finished and the building chamber cooled slowly to room temperature to minimize internal stresses, parts are removed. The heat produced by the laser beam to melt the material powder can be transported out of the part faster if support structures are used to act as a heat sink Fabrication of Heat Exchanger Channels Three stainless steel heat exchangers were manufactured containing either cubic (Figure 2-5a), round-strut tetradecahedral (Figure 2-5b) or thin-strut tetradecahedral (Figure 2-5c) cells. Heat exchangers were manufactured in several sections to avoid any warping or bending, which can occur if the part is too long. The dimensions of the three differed slightly to get an integral number of cells across the channel width in each case. For the purpose of comparison, two hollow heat exchanger channels were also fabricated, one from a solid stainless channel and the other lasersintered. Both hollow channels had the same dimensions, with 25.4 mm square cross-sections and 1.7 mm thick walls.

33 16 (a) (b) (c) Figure 2-5: Channel section (a) cubic, (b) round-strut tetradecahedral, and (c) thin-strut tetradecahedral.

34 17 The cubic, round-strut tetradecahedral and thin-strut tetradecahedral channel sections were fabricated using direct laser-sintering system (Model EOSINT M270, EOS GmbH, Krailling, Germany) and EOS Stainless Steel 17-4 powder (Model SS_17-4_M270, EOS GmbH, Krailling, Germany). (a) (b) (c) Figure 2-6: Heat exchanger channels (a) end view of cubic channel (b) end view of round-strut tetradecahedral channel, and (c) thin-strut tetradecahedral.

35 18 Figure 2-6 shows end views of completed cubic (Figure 2-6a), round-strut tetradecahedral (Figure 2-6b) and thin-strut tetradecahedral channel (Figure 2-6c) sections. One section of the cubic channel weighed 151 g, while the round-strut tetradecahedral channel section weighed 103 g, and the thin strut tetradecahedral channel was significantly lighter, weighing only 71 g. Using this method, the mass of the tetradecahedral structure was minimized resulting in a mass-optimized structure. The results for the thin-strut tetradecahedral cells was compared with the conventional cubic and tetradecahedral channels. The objective of investigating the novel tetradecahedral structure was to optimize the shape of internal struts to minimize the mass while transporting the same amount of heat. Figure 2-7 shows the internal structure of the round-strut tetradecahedral channel with one wall removed. As can be seen from the figure, the struts are uniform throughout the structure and are in complete contact with the wall of the heat exchanger. Figure 2-7: Round-strut tetradecahedral channel with side face removed.

36 19 To assemble the four sections and form a complete channel, the offsets at the ends of each section were machined off and sections were welded together (Figure 2-8). Two stainless steel flanges were manufactured and welded to both ends. In order to measure the pressure drop across the porous structure two stainless steel tubes were connected as wall taps to the first and last channel sections, respectively. The cubic channel was approximately 295 mm long, the round-strut tetradecahedral 299 mm long and the thin-strut tetradecahedral 292 mm long. Figure 2-8: Complete assembly of the heat exchanger with four sections welded together Material Properties The surfaces and cross sections of the channels were examined using scanning electron microscopy (SEM) and energy dispersive x-ray spectroscopy (EDS) (TM3000, Hitachi High-Technologies Canada Incorporated, Toronto, ON, Canada) to analyze the porosity, oxide content, roughness and the material composition. The inner surface of the channel was rough, as shown by the SEM micrograph in Figure 2-9a, due to the DMLS fabrication process that sinters powder particles. The rough surface of the porous structure and the wall surface may enhance near wall flow turbulence [22].

37 20 An average oxide content of 4% was measured at the outer surface of the heat exchanger wall using EDS. Figure 2-9b shows a cross-section through the heat exchanger wall, which was found having negligible porosity and to be impervious to gas penetration. (a) (b) Figure 2-9: SEM images of (a) channel wall surface, and (b) cross section of the channel wall.

38 21 The location where the struts were connected to the wall of the heat exchanger were also analyzed, as shown in Figure In heat exchangers fabricated using DMLS the geometry is predetermined and the designer have full control over the connection and internal geometry of the heat exchanger. Using DMLS, the wall and the struts were built as one solid structure which results in a superior connection between all of the struts and the wall of the heat exchanger, with no thermal resistance at the interface of strut and the wall. Figure 2-10: Connection between the strut and the channel wall of the heat exchanger. In order to check the uniformity of the material composition and the connection at the strut and the wall connection point, the square area on Figure 2-10 was chosen and four areas (P1, P2, P3, and P4) were analyzed as shown in Figure The composition of the channel walls was analyzed at several points using EDS, which confirmed that the composition of the steel

39 22 corresponded to that provided by the manufacturer, the main alloying elements being Cr ( wt%), Ni (3-5 wt%), and Cu (3-5 wt%) with traces of Mn, Si, Mo and Nb. Figure 2-11: EDS analysis of the coating micro structure at the connection point between the struts and the wall.

40 23 Conduction Heat Transfer in DMLS Heat Exchangers 3.1 Test Samples, Experimental Apparatus and Results The geometry of the struts was uniform throughout the length of each channel section, as shown in Figure 3-1, where the wall of a round-strut tetradecahedral section (Figure 3-1a) was removed to check the uniformity of the geometry and to investigate the effective thermal conductivity of the round-strut tetradecahedral structure without the influence of outer walls. (a)

41 24 (b) Figure 3-1: Tetradecahedral structure (a) one partially removed wall, and (b) without walls. Three sets of experiments were conducted (Figure 3-2); the first set aims to determine the thermal conductivity of a laser-sintered stainless steel solid block (Figure 3-2a). The second set to determine the effective thermal conductivity of cubic and round-strut tetradecahedral heat exchangers (Figure 3-2b) and the last set to investigate the conduction heat transfer in the roundstrut tetradecahedral structure without the surrounding walls (Figure 3-2c).

42 25 (a) (b) (c) Figure 3-2: Schematic overview over the four different experimental setups. Figure 3-3 shows the experimental apparatus that is used to measure the thermal conductivity of the laser-sintered solid material. A solid block of stainless steel (28 mm x 28 mm x 50 mm) with known material parameters (k = 16 W/mK, ρ = 8000 kg/m 3 ) is connected to a laser-sintered stainless steel block of identical dimensions resulting in a test sample of the dimensions (28 mm x 28 mm x 100 mm). High thermal conductive paste (Omegatherm 201, Omega Company, Stamford, CT) was used to minimize the thermal resistance between the solid blocks. Eight K-type thermocouples with junction diameters of 0.6 mm were fixed onto the blocks, four on each block with a spacing of 10 mm. Thermal conductive paste was also applied to ensure a good thermal connection between the thermocouples and the block surface.

43 26 (a) (b) Figure 3-3: Experimental apparatus to determine the thermal conductivity of laser-sintered stainless steel (a) a block manufactured with common methods on the left side and a block sintered using DMLS on the right side, and (b) a schematic of the experimental setup. A copper heater 9.5 mm x 26.3 mm x 26.3 mm in dimension, which consisted of three holes containing three high-temperature cartridges heater (3614K34, McMASTER-CARR), was attached to the bottom square section of the stainless steel block to apply a constant heat flux varying between 10.6 kw/m 2 to 27.2 kw/m 2. A copper cooling jacket was attached to the top square section of the laser-sintered block to cool the surface and increased the temperature difference between the top and the bottom of the test section. The apparatus was surrounded by a 50 mm thick layer of aluminum silicate insulation (Zircar ceramics, AXHTM) with an average thermal conductivity of 0.08 W/mK.

44 Surface Temperature, (ºC) 27 Figure 3-4 shows the temperature distribution along the solid stainless steel blocks for the first set of experimental investigations. The measured temperature of stainless steel block from 0 mm to 50 mm and laser-sintered stainless steel block from 50 mm to 100 mm. The measured temperatures show a linear trend for each applied heat flux. The conduction of heat is considered to be linear and dominated by a one dimensional heat conduction from the heater block to the cooler block. Hence, k s = (q"-q" loss )/m can be used to calculate the thermal conductivity of the laser-sintered stainless steel block, with q" being the heat flux at 50 mm, q" loss being the heat flux to the surrounding and m being the slope of the measured temperature distribution of the laser-sintered stainless steel block kw/m kw/m kw/m Location, (cm) Figure 3-4: Temperature distribution over sample length for the first set of experiments.

45 28 The second set of experiments was designed to determine the effective thermal conductivity of the fabricated cubic and round-strut tetradecahedral heat exchangers. The experimental apparatus (Figure 3-5) fabricated to run, these experiments consists of a water cooling jacket on one side and a block heater on the other side of the sample. Seven K-type thermocouples with junction diameters of 0.6 mm were fixed on the top outer surface of the channel with high thermal conductivity paste, with the first one positioned at z = 18 mm and with a 37 mm spacing between thermocouples, to measure the local wall temperature. All seven thermocouples were connected to a National Instruments data acquisition (DAQ) system and recorded in a computer equipped with Lab View Signal Express v.3.0 (National Instrument Corporation, Austin, TX). The same copper heating unit, which was used in the first set of experiments, was used for this experiment. The apparatus is surrounded by a 50 mm thick layer of aluminum silicate insulation (Zircar ceramics, AXHTM).

46 29 Figure 3-5: Experimental apparatus, which is used for the third and fourth set of experiments, to measure the temperature distribution for different heat fluxes with cooling at one side of the samples and heating on the opposite site, respectively. Figure 3-6 shows the measured temperatures on the outer surface of round-strut tetradecahedral and the cubic heat exchanger, respectively. The cubic and tetradecahedral samples are heated on one end and cooled at the other, see Figure 3-5 for more information about the experimental apparatus. The experiment was conducted for three different heat fluxes 9.8 kw/m 2, 15.3 kw/m 2 and 22.1 kw/m 2. At location 0, the temperature of the heater block is shown and the temperature at location 1 is the temperature of the cooling block. The temperatures measured for the cubic and

47 Temperature,T ( C) 30 round-strut tetradecahedral heat exchanger are similar. The measured temperatures are not linearly distributed Round-Strut Tetradecahedral, 22.1 kw/m2 Cubic, 22.1 kw/m2 Round-Strut Tetradecahedral, 15.3 kw/m2 Cubic, 15.3 kw/m2 Round-Strut Tetradecahedral, 9.8 kw/m2 Cubic, 9.8 kw/m Relative location, X Figure 3-6: Comparison of experimental results for the outer surface temperature distribution over relative location between cubic and round-strut tetradecahedral sample at a three different heat fluxes. Heating from one side, cooling from the other, results. If the heat transfer for samples shown in Figure 3-6 is modeled as 1-D conduction and the heat " loss to the surroundings is zero (q loss = 0) as shown in Figure 3-7a for a material with a constant thermal conductivity then

48 31 " " q x=0 = q x=1 -k T x x = 0 T = -k x x = 1 (3-1) T x x = 0 = T x x = 1 which would result in a constant slope line of temperature and relative location. (a) (b) Figure 3-7: Comparison between heat transfer in samples (a) with zero heat loss to surrounding, and (b) with heat loss to the surroundings.

49 32 If the heat transfer is modeled as 1-D conduction and the heat loss to the surrounding is larger than " zero (q loss > 0) as shown in Figure 3-7b for a material with a constant thermal conductivity then " " q x=0 > q x=1 -k T x x = 0 T > -k x x = 1 abs ( T T ) > abs ( x x = 0 x x = 1 ) (3-2) T x x = 0 < T x x = 1 which results in a sharper rate of change of temperature and relative location in the beginning of the sample (x = 0) compare to the end of the sample (x = 1). To estimate the heat loss to the surrounding; the sample was divided into 3 segments as shown in Figure 3-8. The amount of heat loss in each segment is calculated individually. The total heat loss from the sample can be obtained from the summation of heat loss in each segment. The grey line demonstrate the rate of change of temperature with respect to location. Temperature points (P1, P2, and P3) which are at the center of each segment were measured using thermocouples. Conservation of energy for each segment yields X = X = X = q loss 1 + q loss 2 + q loss 3 = q loss,total X = X = X = q x=0.125 = q loss 1 + q x=0.375 (3-3) q x=0.375 = q loss 2 + q x=0.625

50 33 q x=0.625 = q loss 3 + q X=0.875 Substituting Fourier's law of conduction to the equation (3-3) k A Cross ( T x x = T x x = ) = q loss 1 k A Cross ( T x x = T x x = ) = q loss 2 (3-4) { k A Cross ( T x x = T x X = ) = q loss 3 Figure 3-8: Schematics for sample segmentations for heat loss calculation.

51 34 The temperature measurement for the cubic structure at 22.1 kw/m 2 (Figure 3-6) were substituted into Equation (3-4) k A Cross ( 413 C 240 C) = q loss 1 = 0.34 W 47% of Total Loss k A Cross ( 240 C 134 C) = q loss 2 = 0.21 W 29% of Total Loss (3-5) { k A Cross ( 134 C 49.0 C) = q loss 3 = 0.17 W 24% of Total Loss The heat loss in each segment is not equal. The loss is reduced when moving from the hot side of the segment (x = 0.125) to the cold side of the segment (x = 0.875) because the average temperature of the segment reduces. The heat loss of each segment should be proportional to the temperature differential between surface of the sample and the ambient. The total heat loss was approximately 5% of the total heat transfer to the heat exchanger. The same experimental setup as shown by Figure 3-5, was used to measure the effective thermal conductivity of the tetradecahedral structure (Figure 3-1b) by replacing the heat exchanger channel with the round-strut tetradecahedral structure Figure 3-9. With the outer surface walls filling around 15% of the cross sectional area, experimental investigation of the inner structure is necessary to fully understand the heat conduction mechanism within laser-sintered stainless steel heat exchangers. These temperature measurements were used to calculate the effective thermal conductivity of the round-strut tetradecahedral structure.

52 Temperature, T ( C) kw/m kw/m kw/m Relative Location, X Figure 3-9: Temperature distribution over sample length for the third set of experiments. Figure 3-10 shows the calculated thermal conductivity of the laser-sintered stainless steel block for three different heat fluxes. Furthermore, thermal conductivity values provided by EOS GmbH [37] for a temperature of K and for a temperature of K are shown by dotted lines. The calculated thermal conductivity values are in good agreement with the values provided by EOS GmbH [37]. Only a small variation of the thermal conductivity with different applied heat fluxes is observed. A thermal conductivity value of 14 W/mK is used for the solid in analytical predictions of the effective thermal conductivity of the round-strut tetradecahedral and cubic samples.

53 Thermal Conductivity, (W/mK) Laser Sintered Block Applied Heat Flux, (kw/m 2 ) Figure 3-10: Thermal conductivity of the laser-sintered stainless steel block over applied heat flux for the first set of experiments. Data sheet values are taken from the EOS Stainless Steel 17-4 data sheet [37]. The results of the effective thermal conductivity for all three applied heat fluxes and both heat exchangers are shown in Figure The round-strut tetradecahedral heat exchanger provides less conductive heat transfer than the cubic one due to the difference in porosity and the effective thermal conductivity being inversely proportional to the temperature. The calculated effective thermal conductivities are small in comparison to the conductivity of the solid material, which is 14 W/mK for laser-sintered stainless steel (Figure 3-10). The calculated effective thermal conductivities for different heat fluxes do not vary significantly and the differences are within the error range of the experimental error.

54 Effective Thermal Conductivity, (W/mK) 37 Figure 3-11 shows the huge impact of the walls on the effective thermal conductivity for the roundstrut tetradecahedral structure. The experimentally derived effective thermal conductivity values for the round-strut tetradecahedral inner structure is about a third of the effective thermal conductivity values, which are based on the experimental results for the round-strut tetradecahedral sample with outer walls. The porosity values for both structures without walls are different from the calculated porosity considering with walls, 0.89 and 0.75 for tetradecahedral and cubic structure respectively. 6 5 Cubic Round-Strut Tetradecahedral Round-Strut Tetradecahedral Structure Applied Heat Flux, (kw/m 2 ) Figure 3-11: Comparison of the effective thermal conductivity over applied heat flux for the cubic and round-strut tetradecahedral heat exchangers, and the round-strut tetradecahedral structure without walls.

55 Heat Transfer Characteristics Theory Important factors influencing the thermal performance of porous structures are the porosity, pore density, pore size, and fiber diameter of the open-cell porous media. Porosity (ε) is defined as the void volume in the porous sample divided by its total volume. As the porosity of a sample increases, the amount of solid material of that sample decreases, which decreases the strength of the sample. Pore density is measured in pores per inch (PPI) by counting the number of pores per linear inch. The dimensions of a pore are specified by the pore size (dp), which defines the size of a unit cell, and fiber diameter (df). Conduction of heat is described by the equation: T t (α T) = q C p ρ (3-6) where α is the thermal diffusivity, c p is the specific heat capacity and ρ is the density of the material. All experiments were conducted at steady state, and material parameters were assumed constant and isotropic. Assuming heat conduction is one-dimensional, Equation (3-6) simplifies to T L dt dx = q" k (3-7) where the heat flux q " = q, q is the applied heat and A = A solid + A fluid is the total conducting area A of the channel. L is the length of the channel and T is the corresponding temperature difference. Equation (3-7) is applicable for a block of material, but not for a composite material, such as the laser-sintered porous prototypes. In the case where heat exchanger channels are filled with air, heat

56 39 is conducted almost exclusively through the laser-sintered stainless steel, yielding Equation (3-8), which is used to calculate the effective thermal conductivity in the present study. k eff = q"l T = q"a solid L q = (1 ɛ) A solid T A solid L T (3-8) Analytical Models Analytical models are used to predict the effective thermal conductivity of multiphase or composite materials. In addition, analytical models allow the modeling of heat transfer and flow through heat exchangers, which simplifies numerical investigations. All analytical models, which are presented in the following, assume a certain spatial distribution of a fluid phase and solid phase within a given sample. In addition, the porosity and thermal conductivity of each material participating in the composite sample is taken into account. The participating materials in the present study are air (k f = W/mK) and stainless steel (k s =14 W/mK). The series model shown in Figure 3-12b predicts that the effective thermal conductivity k eff = 1 ɛ + (1 ɛ)/k k s f (3-9) is the harmonic average of the thermal conductivities of the solid and gas phases, weighted by the porosity. It assumes that both materials are oriented horizontally to the direction of the temperature gradient, e.g. the heat flux with fluid and solid phases alternating. This combination leads to a material that has neither a direct solid path nor a direct fluid path from the hot side to the cold side of the first sample. The Series Model is regarded as being the lower bound of the thermal conductivity and is dominated by the thermal conductivity of the fluid phase. The Serial Model

57 40 was first introduced by Reuss [29] in the field of elasticity and transferred to heat conduction by Egli [28], Wiener [32]. In contrast, the Parallel Model as shown in Figure 3-12a [28, 32, 33] (arithmetic mean weighted by porosity) k eff = ɛk f + (1 ɛ)k s (3-10) assumes a material distribution, which consists of equally-sized layers that are oriented vertically to the direction of the temperature gradient. Direct solid paths of minimal length exist between hot and cold side of the test sample. Hence, the Parallel Model is considered as being the upper bound of the possible effective thermal conductivity and it is dominated by the thermal conductivity of the solid phase. Both these models define the bounds of possible effective thermal conductivity values. (a) (b) Figure 3-12: Heat transfer direction through (a) Parallel Model, and (b) Series Model. The Effective Medium Theory (EMT) (1 ɛ) k s k eff k s + 2k eff + ɛ k f k eff k f + 2k eff = 0 (3-11)

58 41 assumes a random distribution of both phases within the sample and was introduced by Bruggeman [35]. A different attempt to model the effective thermal conductivity is to regard the distribution of one phase as being regularly shaped. The Maxwell-Eucken Model [36] k eff = k f 2k f + k s 2(k f k s )(1 ɛ) 2k f + k s + (k f k s )(1 ɛ) (3-12) is such a model which assumes that the solid phase is spherical and that the solid phase is covered by the fluid phase, which is continuous. In the given form, it is only valid for high porosity values. Leach [17] formulated two models with the Cubic Series Parallel Model (CSP) k eff = k s (1 ɛ 2 k s ɛ 2 3 3) + k f + (k s k f )ɛ 1 3 (3-13) being the lower bound like the Serial Model and the Cubic Parallel Serial Model (CPS) k eff = k s k s (k s k f )ɛ 2 3 k s (k s k f )(ɛ 2 3 ɛ) (3-14) being the upper bound like the Parallel Model. Both Cubic Cell Models (CCM) assume that the fluid phase fills cubical shaped cells of solid. The main difference between the two Cubic Cell Models is the treatment of cell corners. The Cubic Cell Models are expected to predict the effective thermal conductivity of the Cubic heat exchanger well, as the unit cell is quite similar. Boomsma and Poulikakos [34] introduced an analytical model based on the assumption that the solid phase consists of tetradecahedral shaped cells and is filled with the fluid phase. The Tetrahedral Unit Cell (TUC) model

59 42 k eff = 2 2(R A + R B + R C + R D ) (3-15) where, R A = 4d (2e 2 + Πd(1 3))k s + (4 2e 2 Πd(1 e))k f R B = (e 2d) 2 (e 2d)e 2 k s + (2e 4d (e 2d)e 2 )k f R C = ( 2 2e) 2 2Πd 2 (1 2e 2)k s + 2( 2 2e Πd 2 (1 2 2e)) 2e R D = e 2 k s + (4 e 2 )k f d = 2(2 (5 8 ) e3 2 2ɛ Π(3 4e 2 e) e =0.339 is similar to the Cubic Cell models, but uses a tetradecahedral as a unit cell. The TUC Model is sensitive to the value of е. Consequently, the TUC Model is only applicable for porosities within the range 50% to 98%, and the value for e, which is suggested by Boomsma and Poulikakos[34] and is used by this paper to compare the prediction of the effective thermal conductivity with other results. The TUC Model uses the same unit cell as the tetradecahedral heat exchanger and is expected to estimate the effective thermal conductivity of the tetradecahedral heat exchanger adequately. All of the above mentioned analytical models are used by this study to predict the

60 Effective Thermal Conductivity, (W/mK) 43 effective thermal conductivity of both heat exchangers, and the predictions are compared to the experimental results Analysis and Discussion Serial Model Parallel Model Maxwell-Eucken Model Tetrahedral Unit Cell Model Cubic Series Parallel Model Cubic Parallel Serial Model Porosity Figure 3-13: Comparison of the effective thermal conductivity for different porosities between predictions of various analytical models. Figure 3-13 summarizes the predictions of various analytical models. The Parallel Model Equation (3-10) marks the upper bound, whereas the Serial Model Equation (3-9) is the lower bound of the possible effective thermal conductivity values. The Tetradecahedral Unit Cell model (TUC), the

61 44 Effective Medium Theory model (EMT) and the Maxwell-Eucken model (MEM) are only valid for porosity values higher than 0.5. The reason for that limitation are model assumptions. For smaller values of porosity, the TUC is not porous any longer, while the EMT and MEM are descending from solid being covered by fluid, towards, fluid being covered by solid. Small to no difference is observed for the MEM and the EMT, despite the difference in material distribution, for porosity values above Both models were not designed for the prediction of the effective thermal conductivity of open porous media, as presented by this study. On the other hand, the Cubic Cell models, Cubic Parallel Series model (CPS) and Cubic Series Parallel model (CSP) are based on a regular structure, which is patterned in space. The three lasersintered heat exchangers are designed in a similar manner. Hence, it is expected that CSP and CPS are able to predict the effective thermal conductivity of the presented heat exchangers. In addition, the TUC is expected to be accurate for porosity values above 0.6.

62 Effective Thermal Conductivity, (W/mK) Tetrahedral Unit Cell Model Cubic Series Parallel Model Cubic Parallel Serial Model Round-Strut Tetradecahedral Structure Porosity Figure 3-14: Comparison of the effective thermal conductivity for different porosities between predictions of various analytical models and the experimental results. The effective thermal conductivity for the round-strut tetradecahedral inner structure is derived from experimental results and shown in Figure The Tetrahedral Unit Cell (TUC) model Equation (3-15) is not able to predict the effective thermal conductivity values, while the Cubic Series Parallel Model (CSP) Equation (3-13) and Cubic Parallel Serial Model (CPS) Equation (3-14), are able to predict them for porosity values less than 0.5. Researchers use porosity values above 0.7 in the field of forced convection heat exchangers. Smaller porosity values lead to a very high loss of pressure, while the additional surface area does not offer any benefits for heat transfer [39].

63 Conclusion An investigation on the effective thermal conductivity for two different laser-sintered channels and a round-strut tetradecahedral structure has been conducted. Both channels have an ideal bonding between the wall and the inner structure and are dense. Furthermore, it was shown that the investigated samples almost conduct no heat with the effective thermal conductivity being small compared to the thermal conductivity of the solid inside the samples, with the round-strut tetradecahedral sample conducting even less heat. However, the effective thermal conductivity that was derived from the experimental results of the channels is not characteristic of the inner structures. The conduction heat transfer in the round-strut tetradecahedral structure without the surrounding walls was also investigated. The results were compared to predictions of the effective thermal conductivity by analytical models. Cubic Series Parallel Model (CSP) and Cubic Parallel Serial Model (CPS) were able to predict the change of the effective thermal conductivity with a variation of the fiber diameter porosity values, while the Tetrahedral Unit Cell (TUC) Model can be used to predict the effective thermal conductivity for porosity values above 0.6. The measured effective thermal conductivity of the round-strut tetradecahedral structure was in a good agreement with Cubic Parallel Serial Model (CPS).

64 47 Convection Heat Transfer in DMLS Heat Exchangers 4.1 Experimental Apparatus The experimental apparatus fabricated to test heat exchangers consists of a compressed air supply and instruments to control and measure air pressure, flow rate, and temperature (Figure 4-1). The laboratory air supply provides a maximum flow rate of 1.6 x 10-3 m 3 /s (100 L/min at standard temperature and pressure) at a pressure of 620 kpa (90 psi). The compressed air pressure was set by a pressure regulator (Model R , Ingersoll-Rand plc, Dublin, Ireland) and a mass flow controller (Model FMA5542, Omega Company, Stamford, CT) regulated the airflow rate through the heat exchanger channel. The inlet (Tin) and outlet (Tout) temperatures were measured by type- K thermocouple probes (Model TJ36-CASS-032-G-6, Omega Company, Stamford, CT), which were located in T-junctions placed before and after the diverging and converging inlet and outlet manifolds (Figure 4-1) that were also manufactured using DMLS. Air pressures were read at the mid-point of the first and fourth channel section (at z = 37 mm and 221 mm, where z is the distance measured from the beginning of the channel). The pressure drop was measured using a digital manometer (Model HHP-103, Omega Company, Stamford, CT) set to a maximum range of 498 Pa with an accuracy of 0.2% of full scale. Eight K-type thermocouples with junction diameters of 0.6 mm were fixed on the top outer surface of the channel, with the first one positioned at z = 18 mm and 37 mm spacing between thermocouples. To ensure good contact between the thermocouples and the surface, a high thermal conductivity paste (Omegatherm 201, Omega Company, Stamford, CT) was applied. All 10 thermocouples were connected to a National Instruments Data Acquisition (DAQ) system and

65 48 recorded in a computer equipped with Lab View Signal Express v.3.0 (National Instrument Corporation, Austin, TX). A 12.7 mm diameter x 2439 mm long rope heater (Model FGH , Omega Company, Stamford, CT) was wrapped uniformly around the heat exchanger and surrounded by a fiberglass mat insulation (Micro-Flex, John Manville Corporation, Denver, CO) with an average thermal conductivity of W/mK. Figure 4-1: Schematic of experimental apparatus. A high temperature infrared camera (IR) (FLIR SC5000, FLIR Systems Inc., Wilsonville, OR) was used to observe temperature variations across the cross-sections of the heat exchangers while they were operating. To take infrared images, the converging exit section of the heat exchanger was removed and the IR camera positioned in front of it. In order to ensure uniform emissivity over the strut surfaces, they were coated with black, high-temperature thermally conductive paint with an emissivity of A constant heat flux of 2.3 kw/m 2 was applied to cubic, round-strut tetradecahedral and thin-strut tetradecahedral heat exchangers while airflow rates were varied from 20 L/min to 80 L/min.

66 Hydraulic Characteristics Figure 4-2 presents the pressure drop measurements at different flow velocities through cubic, round-strut tetradecahedral and thin-strut tetradecahedral cell heat exchangers. It can be seen that the pumping power required is highest for the cubic cells while the thin-strut tetradecahedral resulted in the lowest pressure drop. The mean inlet air temperature was kept constant at 21ºC. Leong and Jin [23] compared the pressure drop through metal foams, which are usually modeled as having tetradecahedral pores, to the pressure drop through wire-screens [24, 25], which have pores with square cross-section. They also found that frictional losses were much higher for the wire screens than for the foams. They attributed the difference to the different structures, with the inter-connected cells of the foam providing less resistance to flow than the wire screens.

67 Pressure Gradient, ΔP/m (Pa/m) Cubic Round-Strut Tetradecahedral 4000 Thin-Strut Tetradecahedral Fluid Velocity, u (m/s) Figure 4-2: Variation of experimentally measured pressure gradient with average fluid velocity in channels with cubic, round-strut tetradecahedral and thin-strut tetradecahedral cells. The pressure gradient for fluid flow through a porous medium is often expressed using Darcy s law [26]. ΔP L = µ K u + ρc F K u2 (4-1) where P is the pressure difference across the length of the channel, µ and ρ the dynamic viscosity and density, and u the average velocity, determined by dividing the volume flow rate of air by the

68 51 open cross-sectional area of the channel. The permeability (K) and Forchheimer coefficient (C F ) are properties of the porous media that are determined experimentally. Equation (4-1) was non-dimensionalize by defining the Darcy friction factor f f = 2ΔP L H (4-2) ρu 2 where H is the internal height of the square cross-section channel which is the hydraulic diameter of the square channel. Substituting Equation (4-1) in Equation (4-2) fda 1/2 2 = 1 Re K + C F (4-3) with Da = K H 2 and Re K = ρu K µ. The Reynolds number Re K is based on the characteristic length scale K and Da is the Darcy number [26]. Values of K and CF for the three structures, listed in Table 1, were determined by using a least squares fit of Equation (4-1) to the data in Figure 4-2. The permeability of the thin-strut tetradecahedral was higher than the permeability of the cubic and round-strut tetradecahedral structures. Both conventional cubic and round-strut tetradecahedral heat exchangers had the same strut diameter (1 mm), but since the porosity of the conventional round-strut tetradecahedral structure was higher it had a higher permeability (K). The Forchheimer coefficient (C F ) is a measure of the total resistance to flow due to fluid drag, and since the tetradecahedral channel had a surface area that was only half that of the cubic structure its CF value was correspondingly smaller. The thin-strut tetradecahedral channel had a surface area that was smaller than that of the cubic and round-strut tetradecahedral structure and its CF value was correspondingly smaller.

69 52 Table 4-1: Structural comparison between the cubic, round-strut tetradecahedral and thin-strut tetradecahedral heat exchanger channels. Cubic Round-Strut Tetradecahedral Thin-Strut Tetradecahedral Internal surface area (m 2 ) Mass per section (g) Porosity Permeability K (m 2 ) 1.16 x x x 10-7 Forchheimer coefficient C F In order to predict the pressure drop for the fabricated heat exchangers, Equation (4-3) was fitted to the experimental data shown in Figure 4-2. Cubic: fda 1/2 2 = 1 Re K (4-4) Round-Strut Tetradecahedral: fda 1/2 2 = 1 Re K (4-5) Thin-Strut Tetradecahedral: fda 1/2 2 = 1 Re K (4-6) These relations are illustrated graphically in Figure 4-3, together with the experimental data.

70 Cubic Round-Strut Tetradecahedral Thin-Strut Tetradecahedral Eq. (4-4) Eq. (4-5) Eq. (4-6) fda Figure 4-3: Friction factor variation with Reynolds number for channels with cubic, roundstrut tetradecahedral and thin-strut tetradecahedral. Re K

71 Heat Transfer Characteristics Internal structures in a heat exchanger channel promote convection, but they also increase the cross-sectional area for axial heat conduction and the surface area for radiation. Heat supplied to the walls of the heat exchanger is transferred by conduction and radiation both radially towards the center of the channel and axially along the length of the tube. At low airflow rates a significant portion of the total heat applied may be lost by heat conduction to the piping at the ends of the heat exchanger channel instead of being transferred to the air flowing through it. The heat exchangers were tested at nine different air flow rates ranging from 8.3x10-5 to 1.3x10-3 m 3 /s (10 to 90 L/min) and at four different heater voltages ranging from 30 to 60 V (giving uniform wall heat flux varying from 0.8 to 3.2 kw/m 2 ). The inlet air mean temperature and pressure were 21ºC and 130 kpa. Experiments were performed at steady state and readings were taken when the thermocouple outputs had stabilized. To measure heat transfer to the air passing through the heat exchanger the temperature rise from the inlet to the outlet of the heat exchangers was measured as a function of flow rate as shown in Figure 4-4 for cubic (Figure 4-4a), round-strut tetradecahedral (Figure 4-4b) and thin-strut tetradecahedral (Figure 4-4c) heat exchangers. The temperature difference decreased with increasing flow rate, with the cubic heat exchanger producing a slightly greater temperature rise for the same applied heat flux and air flow rate. The cubic structure has an internal surface area roughly twice that of the round-strut tetradecahedral heat exchanger which would lead us to expect higher absolute heat transfer rate. The cubic structure has an internal surface area three times of the thin-strut tetradecahedral heat exchanger, which leads to the expectation of a higher absolute heat transfer rate. The thin-strut tetradecahedral heat exchanger produced the lowest temperature rise (Figure 4-4c).

72 Temperature Rise, ΔT ( C) Temperature Rise, ΔT ( C) Temperature Rise, ΔT ( C) kw/m² 2.3 kw/m² 1.5 kw/m² 0.8 kw/m² kw/m² 2.3 kw/m² 1.5 kw/m² 0.8 kw/m² Air Volumetric Flow Rate, ὺ (L/min) Air Volumetric Flow Rate, ὺ (L/min) (a) (b) 3.2 kw/m² 2.3 kw/m² 1.5 kw/m² 0.8 kw/m² Air Volumetric Flow Rate, ὺ (L/min) (c) Figure 4-4: Increase in air temperature from the inlet to the outlet of a) cubic b) round-strut tetradecahedral, and c) thin-strut tetradecahedral heat exchangers with air flow rates varying from 10 to 90 L/min for applied heat flux in the range of 3.2 to 0.8 kw/m 2.

73 56 To calculate the efficiency of the heat exchangers the increase in enthalpy of air passing through them was calculated: Q = m c p,a (T in T out ), (4-7) where m is the mass flow rate of air, c p,a its specific heat and Tin and Tout the air inlet temperature and outlet temperature of the heat exchanger respectively. Figure 4-5 shows values of Q as a function of air flow rate for the case of both 2.3 kw/m 2 and 0.8 kw/m 2 heat flux for all three heat exchangers. The total power input to the heat exchangers in all three cases is indicated by the horizontal lines. Three heat exchangers transferred approximately the same amount of energy from the heaters to the air, in spite of their different structures. As the airflow rate increased more of the heat supplied by the heater was transported out of the channels by the air. Heat not transferred to the air was either lost to the surroundings through the insulation or conducted to the tubes connected to the end of the heat exchangers.

74 Heat Transfer Rate,Q (W) W/m2 Round-Strut 3281 W/m2 Thin-Strut 3281 W/m2 820 W/m2 Round-Strut 820 W/m2 Thin-Strut 820 W/m W/m2 820 W/m Air Volumetric Flow Rate, ὺ (L/min) Figure 4-5: Rate of heat transfer to air flowing through cubic, round-strut tetradecahedral and thinstrut tetradecahedral heat exchangers with varying air flow rate and total heater power of 0.8 and 2.3 kw/m 2. The horizontal lines mark the total heater power of 0.8 and 2.3 kw/m 2. To observe the effect of the internal structure of the heat exchangers on conduction heat transfer, infrared images of the internal struts were taken at the outlet of the channel, with end-caps removed. Figure 4-6 shows sample images for flow rates varying from L/min in the cubic heat exchanger with an applied wall heat flux of 2.3 kw/m 2. The temperatures at the center and edge of the channel cross-section are indicated. At the highest flow rate of 80 L/min (Figure 10d) the edges of the heat exchanger tube were at the highest temperature (62 C), while at the center the temperature was a little lower, about 60 C.

75 58 Figure 4-6: Temperature variation across exit of cubic heat exchanger for constant applied heat flux of 2.3 kw/m 2 and air flow rate (a) 20 L/min, (b) 40 L/min (c) 60 L/min, and (d) 80 L/min. Temperature scales are in C. The temperature drop across the cross-section of the heat exchanger demonstrated that heat was being conducted to the interior of the cubic mesh. Under these conditions the heat exchanger transferred over 90% of the heater power to the air (see Figure 4-5). The temperature was lower at the top of the channel than at the bottom as a result of natural convection increasing the flow near the upper surface and decreasing it along the lower surface. As the air flow rate decreased to 60

76 59 L/min (Figure 4-6c) the temperature became more uniform across the channel, indicating little heat transfer. At the lowest flow rate, 20 L/min (see Figure 4-6a), the temperature gradient was reversed: the highest temperature of 139 C was at the center of the grid and decreased to the walls of the channel that were at approximately 129 C. Heat was no longer being transmitted from the walls to the air at the end of the channel, which explains why the air transported out less than 50% of the heater power (see Figure 4-5). The temperature distribution was symmetrical about the center of the channel, showing that convection heat transfer was small compared to conduction. The round-strut tetradecahedral heat exchanger exhibited similar behavior, as shown in Figure 4-7. At a low flow rate of 20 L/min the highest temperature was at the center of the channel and decreased towards the walls (Figure 4-7a). Increasing the flow rate to 80 L/min reversed the temperature gradient (Figure 4-7c). At the higher flow rates the center of the round-strut tetradecahedral channel is significantly cooler than that of the cubic channel (compare Figure 4-6d with Figure 4-7d). The thin-strut tetradecahedral heat exchanger also exhibited similar behavior, as shown in Figure 4-8. At a low flow rate of 20 L/min the highest temperature was not at the center of the channel but it was lower than the walls (Figure 4-8a). Increasing the flow rate to 80 L/min reversed the temperature gradient (Figure 4-8c). At the higher flow rates the center of the thin-strut tetradecahedral channel is significantly cooler than that of the cubic channel (compare Figure 4-6d with Figure 4-8d) but close to that of round-strut tetradecahedral (compare Figure 4-7d with Figure 4-8d). As can be seen from Figure 4-8, while increasing the flow rate to 80 L/min the center of the channel of the thin-strut tetradecahedral (Figure 4-8) became cooler faster than round-strut tetradecahedral (Figure 4-7) and that of the cubic channel (Figure 4-6).

77 60 Figure 4-7: Temperature variation across exit of round-strut tetradecahedral heat exchanger for constant applied heat flux of 2.3 kw/m 2 and air flow rate (a) 20 L/min, (b) 40 L/min, (c) 60 L/min, and (d) 80 L/min. Temperature scales are in C.

78 61 Figure 4-8: Temperature variation across exit of thin-strut tetradecahedral heat exchanger for constant applied heat flux of 2.3 kw/m 2 and air flow rate (a) 20 L/min, (b) 40 L/min, (c) 60 L/min, and (d) 80 L/min. Temperature scales are in C. Heat is transferred into the channel by conduction along the metal struts and carried out by convection through the air. When an air stream is heated through a characteristic temperature difference T the heat transfer rate due to convection is given by:

79 62 Q conv = m c p,a T (4-8) The rate of heat conduction along the struts over the same temperature difference is given by Q cond = k s A s T H (4-9) where ks is the thermal conductivity of the solid, As the cross-sectional area of conduction and the channel height H is taken as a characteristic length. The ratio of the convective to conduction heat flux gives a dimensionless Peclet number: Pe = Q conv Q cond = m c p,ah k s A s (4-10) The solid area across any cross-section is related to the porosity ε by the relation: ε = 1 A s A t (4-11) where At is the total cross-sectional area of the channel. For a channel with a square cross-section, A t = H 2. Substituting this in Equation (4-10) gives Pe = m c p,a (1 ε)h (4-12) Assuming that for stainless steel k s = 16 W/mK, that H =25 mm, and with the porosity values given in Table 4-1, Pe for all three heat exchanger channels were calculated. Figure 4-9 shows the variation of heat exchanger efficiency, defined as the fraction of the total heater power transferred to the air flowing through the heat exchanger, with Peclet number. The efficiency is a function of Pe alone, irrespective of the heat flux applied. At a given flow rate Pe is smaller for the cubic channel, since it has lower porosity (ε). Convection is a more dominant heat transfer mechanism

80 63 for the thin-strut tetradecahedral, and round-strut tetradecahedral than the cubic channel, compared to conduction, since it has greater porosity, explaining why its interior is cooler at the same airflow rate (compare Figure 4-6d to Figure 4-7d). At a flow rate of 50 L/min Pe is 10. Convection is an order of magnitude greater than conduction at flow rates above this value, and carries heat away from the interior faster than it can be conducted in, so that the interior remains cooler than the walls as seen in Figure 4-6d and Figure 4-7d. At lower flow rates heat cannot be transported away by air more rapidly than it is conducted in, and the interior of the channel starts to overheat (see Figure 4-6a, and Figure 4-7a). The heat exchanger efficiency, therefore, decreases at low flow rates (Pe < 10).

81 Actual Heat Exchanger Efficiency, η (%) Cubic Round-Strut Tetradecahedral Thin-Strut Tetradecahedral Peclet Number, Pe Figure 4-9: Variation of heat exchanger efficiency for cubic, round-strut tetradecahedral and thin-strut tetradecahedral channels with increasing Peclet number. Since the heater was wrapped uniformly around the channel, a uniform heat flux was applied to the heat exchanger walls. For a constant heat flux an energy balance gives that the air temperature increases linearly with position along the length of a heated channel [27], so the value of the local air temperature at a given axial distance from the inlet was calculated by interpolating between the inlet and outlet air temperature. Figure 4-10 shows measured surface temperatures and calculated air temperatures for cubic (Figure 4-10a), round-strut tetradecahedral (Figure 4-10b) and thin-strut tetradecahedral (Figure 4-10b) channels at two different flow rates (20 L/min and 80 L/min) and an applied heat flux of 2.3 kw/m 2.

82 Temperature, ( C) Temperature, T ( C) Temperature, T ( C) Wall Air L/min Wall Air L/min Axial Distance from Inlet, x (mm) Axial Distance from Inlet, x (mm) (a) (b) Wall Air L/min Axial Distance from Inlet, x (mm) (c) Figure 4-10: Measured wall temperature and calculated air temperature variation along the length of (a) the cubic, (b) the round-strut tetradecahedral, and (c) the thin-strut tetradecahedral heat exchanger for an applied heat flux of 2.3 kw/m 2 and air flow rates of 20 and 80 L/min.

83 Heat transfer Coefficient, h (W/m 2 K) Heat transfer Coefficient, h (W/m 2 K) Heat transfer Coefficient, h (W/m 2 K) Heat transfer Coefficient, h (W/m 2 K) L/min L/min L/min 40 L/min L/min 40 L/min L/min L/min Axial Distance from Inlet, x (mm) Axial Distance from Inlet, x (mm) (a) 80 L/min 60 L/min 40 L/min 20 L/min (b) 80 L/min 60 L/min 40 L/min 20 L/min Axial Distance from Inlet, x (mm) Axial Distance from Inlet, x (mm) (c) (d) Figure 4-11: Local heat transfer coefficient variation along the length of (a) the cubic, (b) the round-strut tetradecahedral, (c) the thin-strut tetradecahedral and (d) a hollow channel for 2.3 kw/m 2 heat flux.

84 67 A local convective heat transfer coefficient h was defined for a section of the channel by using an energy balance Q conv = ha sf (T s T f ) (4-13) where Asf is the internal surface area of the channel section (the area wetted by the fluid), Ts is the local wall temperature, measured by thermocouples attached to the channel, and Tf the calculated bulk mean fluid temperature. Figure 4-11 shows calculated values of local heat transfer coefficients at various axial positions for cubic, round-strut and thin-strut tetradecahedral heat exchangers and a hollow channel. For the cubic channel, (Figure 4-11a) heat transfer coefficients increased with flow rate, reaching a value of approximately 20 W/m 2 K at a flow rate of 80 L/min. The value of h did not change very much with axial position. For the round-strut tetradecahedral channels (Figure 4-11b) heat transfer coefficients were higher, reaching over 50 W/m 2 K. Thinstrut tetradecahedral resulted in similar heat transfer coefficients than round-strut tetradecahedral. The maximum values were closest to the inlet, and then decreased sharply to reach a constant value at approximately 125 mm (about 5 times channel height H), indicating fully developed flow [25]. Average heat transfer coefficients h for all three channels were calculated as follows: h = Δx L h(x) (4-14) where L is the length of the channel and Δx is the spacing between thermocouples, which is approximately 37 mm. Equation (4-15) non-dimensionalized the average heat transfer coefficient, using the channel height H as a length scale Nu H = h H k (4-15)

85 68 Nusselt numbers were plotted as a function of flow Reynolds number, defined as Re H = ρhu μ (4-16) The variation of Nu H with Re H is presented in Figure All the data for each channel collapsed onto a single line. The results for cubic, round-strut and thin-strut tetradecahedral channels were compared with the results for both an empty steel channel and laser-sintered empty channel. Since the flow inside the empty channel was a developing flow, the fully developed Nusselt number equation could not be used for the comparison in this case. Nu H increased by factors of 4 and 2 for round-strut tetradecahedral and cubic structure respectively compared to the empty channels, whose heat transfer did not differ significantly. The enhancement of Nu H was higher for the round and thin-strut tetradecahedral than the cubic structure, even though it exhibited a lower pressure drop (Figure 4-2). The thin-strut tetradecahedral heat structure resulted in a similar enhancement of Nu H than the conventional round-strut tetradecahedral structure. The thin-strut tetradecahedral structure resulted in the lowest pressure drop compare to the other two structure, a competitive enhancement of Nu H while having the lowest weight. A detailed study of flow through the different internal structures will be required to understand why their heat transfer properties differ.

86 69 40 Round-Strut Tetradecahedral Thin-Strut Tetradecahedral Cubic 30 Empty Channel Empty Channel Laser Sintered Nusselt number, Nu H Reynolds number, Re H Figure 4-12: Average Nusselt number (NuH) as a function of Reynolds number (ReH) for roundstrut, cubic structure, thin-strut tetradecahedral and empty channels.

87 Conclusion Laser-sintering was used to fabricate heat exchanger channels with complex internal structures. Three channels were built, two containing cubic and round-strut tetradecahedral cells with identical strut diameters and one with thin-strut tetradecahedral cells. The round-strut tetradecahedral channel had 56% of the surface area and 68% of the weight of the cubic channel, yet gave higher permeability, lower friction factor and lower pressure drop. The round-strut and thin-strut tetradecahedral channels also gave much higher local heat transfer coefficients than the cubic channel, with a 100% higher Nusselt number. The thin-strut tetradecahedral channel yielded higher permeability, lower friction factor and lower pressure drop compared to the cubic and round-strut tetradecahedral channels. The thin-strut tetradecahedral structure had the lowest mass per section of 71 g compare to 151 g and 103, and highest porosity of 0.84 compared to 0.64 and 0.77 of cubic and round-strut tetradecahedral structures respectively. All three structures had much higher NuH than empty channels. Heat transfer in the investigated channels was by both conduction and convection. In was found that the Peclet number must be large (>10) for convection to be sufficiently rapid to carry away heat conducted to the interior. At lower values of Pe the interior struts become hotter than the walls of the channel. Heat was no longer conducted in and heat exchanger efficiency decreases. Convection was a more dominant heat transfer mechanism for the thin-strut tetradecahedral, and round-strut tetradecahedral than the cubic channel, compared to conduction, due to their higher porosity.

88 71 Wire-Arc Thermal Sprayed Heat Exchangers 5.1 Introduction Porous structures such as metal foam and wire mesh can act as fins to enhance heat transfer in heat exchangers due to their light weight, large surface area to volume ratio, high strength to weight ratio. Wire mesh may not have as large a surface area to volume ratio as other porous structures like metallic foams but are available in a much wider variety of materials and are also more cost effective. In this study wire mesh was used as the porous structure for the fabrication of stainless steel tube heat exchangers since it is available in materials that are much more resistant to oxidation. The porous structures must be in good thermal contact with the surface of the tubes, through which fluid passes, to reduce thermal resistance and ensure high heat transfer. Brazing and welding have been in use for many years to join metallic substrates together. Brazing is expensive since it requires a vacuum furnace to heat the substrate. For successful welding both parts should melt at the same time which did not happen. The wire mesh melts and evaporates much faster than the tube Figure 5-1.

89 72 Figure 5-1: Unsuccessful welding of tube to the wire mesh. Better welded connections can be achieved using larger wire mesh diameters but there are only commercially available with very small pore density, 1 PPI which was not suitable for this study. There is a need for more efficient and economical method of connecting the wire mesh to the tube since heat transfer depends on a good bond between the two. Wire-arc spray coating is a technique to deposit metallic coatings on substrates as shown in Figure 5-2. In this technique two electrically conductive wires are fed into a spray gun, where they generate an arc by applying a voltage between their tips. The arc melts both wires and a jet of compressed air is used to atomize the liquid metal and accelerate molten particles toward the substrate to be coated. The accelerated particles solidify after hitting the substrate and form a coating. Other thermal spray techniques such as high velocity oxy-fuel spraying and plasma spraying are available commercially to create dense coatings.

90 73 Figure 5-2: Thermal skin deposition using wire-arc spray technique. In this study wire-arc spray was used to provide a bond between the wire mesh and the tube due to its low cost and ability to produce thick coatings. A wire-arc spraying process was used due to its low operational cost and high efficiency compared to other thermal spraying processes, which made it a good candidate for mass production of heat exchangers.

91 Geometric Characteristic Metallic wire mesh (Figure 5-3) are porous structures consisting of an array of metals forming square, rectangular or circular pore patterns. Wire mesh screens are manufactured in a variety of pore sizes, wire diameters and wire types and are categorized based on the type of connection between wires as welded, woven, crimped or molded. Tube heat exchangers were modified using wire mesh structures in order to enhance the heat transfer performance of the heat exchanger by increasing its external surface area. Wire mesh enhances the heat transfer of the heat exchanger in a manner similar to solid plate fins, but due to their porosity, the pressure drop across them is significantly lower. The ideal wire mesh PPI should have a high ratio of surface area to occupied volume without reducing air penetration. Additionally, the size of the pore should be large enough to permit thermal spray particles to penetrate and provide sufficient mechanical bonding between the mesh and the tube, reducing the thermal resistance and ensuring high thermal contact. (a) (b) Figure 5-3: Woven copper wire mesh screens of (a) 10 PPI, and (b) 40 PPI.

92 Fabrication of Wire-Arc Thermal Sprayed Heat Exchangers A twin wire-arc spraying system (ValuArc, Sulzer Metco Inc., Westbury, NY) was used to spray Alloy Metcoloy 2 wires on the tube-wire mesh joints to create a metallic bond between tubes and wire mesh at ambient conditions. Optimized parameters, from the work of Rezaey et al. [6], pertaining to wire mesh heat exchangers using wire-arc spraying were employed. They analyzed the in-flight characteristics of the molten droplets measured using the DPV-2000 spray monitoring system (DPV particle diagnostic monitoring system, Tecnar Automation Ltd., St-Bruno, Quebec, Canada) to analyze in-flight characteristics of the molten droplets and effect of spraying distance on their size, temperature, and velocity when they hit the substrate surfaces. Results of these experiments were correlated with coating properties. It is important to minimize the amount of porosity in the thermal sprayed coatings since it has a significant effect on their mechanical and physical properties and would reduce the thermal conductivity of the layer between the tubes and wire mesh. Porosity also reduces the tensile strength of the thermally sprayed coatings, which would result in reduction in the resistance of the coating to thermal stresses at the time of heat exchanger start up and shut down. Backscattered electron images of coatings deposited at spraying distances of 100,150, and 200 mm are shown in Figure 5-4. The light gray metal splats, the intermediate contrast oxide regions at the splat boundaries, and the black pores can be readily discerned. The gray phase in the back scattered electron SEM micrographs shown in Figure 5-4 was identified as oxides by EDS analysis. SEM micrographs of the coating between the wire and tube were analyzed to find the porosity and oxide content and the results are shown in Table 5-1.

93 76 The optimum spraying distance of 0.15 m (6 in) was used to minimize the porosity and oxide content, while maintaining sufficient adhesion strength to obtain satisfactory heat conduction, and mechanical connections at the interface of the tube and the wire mesh. Table 5-1: Porosity, oxide content, and adhesion strength of the coatings sprayed under different conditions [6]. Sample Porosity* (%) Oxide Content* (%) Adhesion Strength** (MPa) 1 (100mm) 4 ± 1 5 ± ± 1 2 (150mm) 2 ± ± 1 24 ± 2 3 (200mm) 6 ± 1 8 ± 1 20 ± 1 * Standard deviation calculated based on 8 SEM image analysis for each sample. ** Standard deviation calculated based on 5 measurements for each sample.

94 77 (a) (b) (c) Figure 5-4: Backscattered electron SEM images of stainless coatings deposited at spray distances of (a) 100 mm, (b) 150 mm, and (c) 200 mm [6].

95 78 In-flight characteristics of the molten droplets were also measured using the DPV-2000 system to investigate the effect of spraying distance on their size, temperature, and velocity when they hit the substrate surfaces. Wire-arc spraying parameters selected for fabricating heat exchangers are listed in Table 5-2. Table 5-2: Wire-arc thermal spray parameters for deposition of stainless steel coating [6]. Gun ValuArc Wire Feed Rate (m/min) 7 Voltage (V) 31 Inlet Pressure (psi) 85 Air Flow Rate (SCFM) 60 Spraying Distance (m) 0.15 Using these parameters, a superior mechanical was achieved between the wire and tube, as shown in Figure 5-5. The gap between the wire and the tube, shown at higher magnification in Figure 5-6, was completely filled by coating material, forming a good path for heat conduction.

96 79 Figure 5-5: SEM micrograph of coated joint [6]. Figure 5-6: SEM image of gap in the wire-tube joint filled by the coating material [6].

97 80 Preliminary Investigation of Flow over Perforated Sheet and Wire Mesh Fins 6.1 Introduction Fins have been conventionally used to enhance heat transfer. The use of porous materials has been proposed as a way to increase the surface area of the heat exchanger that at the same time reduces the weight of the system and increases the efficiency of heat exchange. There is an enormous body of literature dealing with analysis of heat exchangers but few researchers have investigated the heat transfer performance of wire mesh as a heat transfer enhancer for heat exchanger applications. Armour and Cannon [40] studies fluid flow through woven screens measuring the pressure drop across different types of woven metal screens and developing a general correlation for pressure drop. Varshney and Saini [41] used wire mesh screen for solar air heated applications where they packed the air duct with wire mesh screens to enhance heat transfer. They concluded that heat transfer enhancement depends strongly on the geometrical parameters of the wire mesh matrix. Li et al. [42] used wire mesh at the inlet of a channel to create turbulent flow and enhance heat transfer. They looked into different types of wire mesh and varied the Reynolds while measuring the heat transfer enhancement. The effective heat transfer enhancement was also compared to the low pressure loss due to the presence of wire mesh. The heat transfer enhancement was attributed to the presence of wire mesh as a turbulence generator.

98 81 Figure 6-1: Heat Transfer performance charts of different heat dissipation media [14]. Tian et al. [14] studied fluid flow and heat-transfer during forced convection through cellular copper lattice structures. Heat was applied the bottom of the test samples by a heating pad. To find the maximum heat transfer performance of the woven copper mesh they tested several configurations. They discovered that unlike open-cell metal foams and packed beds, the friction factor of the bonded wire screen, apart from being a function of porosity, is also a function of orientation. They concluded that wire-screen mesh can compete with the available heat dissipation media.

99 82 Venugopal, Balaji and Venkateshan [43] experimentally studied the pressure drop and heat transfer in a vertical duct filled with metallic porous structures. The test section consisted of two identical porous structures on both side of a plate-heater. A large increase in average Nusselt number, by a factor of up to 4.52, was observed with a material with 0.85 porosity. The porous media model they developed shows a similar thermo-hydrodynamic performance to that seen in metallic foams. Figure 6-2: Heat transfer performance charts [43].

100 83 Kurian, Balaji and Venkateshan [44] studied the heat transfer enhancement due to packed wire mesh screens. They filled a horizontal channel with wire screens to create a porous block. The results were comparable with heat transfer enhancement of metallic foam structures. Previous researchers have used wire mesh screens to produce a block of porous structure in order to study its pressure drop and heat transfer performance. In this study, single wire mesh screens were used to enhance heat transfer. A simple method of increasing the heat transfer surface area has been developed by using a twin wire-arc thermal spray system to generate a dense, high strength coating that bonds perforated sheet and wire mesh to the outside surfaces of fins which was explained in the previous chapter. In this study, as the first step toward understanding the performance of these porous structures, the heat transfer from perforated sheet; and wire mesh was experimentally investigated. Experiments were done in which electrically heated fins were placed inside a wind tunnel with the air velocity varying between 0 to 20 m/s. To understand the heat transfer enhancement, the temperature distribution of the porous structures was measured using a high temperature infrared camera for different applied voltages and at different velocities. Several fin designs were fabricated and tested using 0.06, 0.12 and 0.18 in (1.52, 3.05 and 4.57 mm) perforated sheet hole diameter, and 10, 14 and 20 PPI wire mesh to understand the heat transfer enhancement due to convection in each case. It was possible to produce significant increases in the heat transfer from the plain tube by connecting porous screens to the outer surface of the tubes.

101 Fabrication of Wire-Arc Thermal Sprayed Fins For this study, one flat plate, three perforated aluminum sheets with 0.06, 0.12 and 0.18 in (1.52, 3.05 and 4.57 mm) hole diameter and three aluminum wire mesh sheets with pore density of 10, 14 and 20 PPI were used as shown in Table 6-1 and Table 6-2. The ideal wire mesh pore size or hole diameter for perforated sheets should correspond to a high ratio of surface area to volume, without reducing air penetration. Additionally, the size of the pore should be large enough to permit thermal spray particles to penetrate, and provide sufficient mechanical bonding between the mesh and the tube, reducing thermal resistance and ensuring good thermal contact. The screen dimensions were identical for both wire mesh and perforated sheet, 76 mm x 76 mm. As the hole diameter decreased from 0.18 in (4.57 mm) Perfo (a) to 0.06 in (1.54 mm) Perfo (c), the number of the holes increased which resulted in the reduction of the open area from 51% to 30%. Table 6-1: Perforated sheet specifications. Fin Opening Size Open Area Center-to-Center Spacing Number of Holes Porosity Surface Area of the Fin Perfo (a) 0.18 in (4.57 mm) 51 % 0.25 in (6.35 mm) in 2 (5690 mm 2 ) Perfo (b) 0.12 in (3.05 mm) 40 % 0.19 in (4.83 mm) in 2 (6968 mm 2 ) Perfo (c) 0.06 in (1.54 mm) 30 % 0.11 in (2.79 mm) in 2 (8129 mm 2 )

102 85 Wire mesh is available in different pore densities, specified as pores per inch (PPI), and experiments were conducted with 10 PPI, 14 PPI and 20 PPI wire mesh (Table 6-2). As the opening size decreased from in (1.9 mm) to in (0.86 mm), which resulted in the increase in the number of pores per inch of the wire mesh, the percentage of the open area decreased from 56% to 46%. Table 6-2: Wire mesh fin specifications. PPI Opening Size Open Area Wire Diameter Porosity Surface Area of the Fin in (1.90 mm) 56% in (0.63 mm) in 2 (4561 mm 2 ) in (1.29 mm) 51% 0.02 in (0.51 mm) in 2 (5110 mm 2 ) in (0.86 mm) 46% in (0.41 mm) in 2 (5837 mm 2 ) Aluminum was sprayed using wire-arc spray system on the sample shown in Table 6-3 where aluminum perforated and wire mesh screens were fastened to a 9.5 mm outer diameter, 6.4 mm inner diameter and 83.8 mm long aluminum tube. A dense coating was deposited on samples as shown in Figure 48. The area over which the porous structures were in contact with the tube were covered by the sprayed aluminum coating, with strong mechanical bonding between the coating, tube and the wire mesh.

103 86 Table 6-3: Summary of the porous structures used in the study. Flat plate Perforated Sheet Wire Mesh Ø= in (4.75 mm) 10 PPI Ø= in (3.17 mm) 14 PPI Ø= in (1.59 mm) 20 PPI

104 87 (a) (b) Figure 6-3: Fabricated fins after thermal spray coating of aluminum on (a) perforated sheet, and (b) wire mesh.

105 Experimental Apparatus and Methods An existing horizontal bench-mounted wind tunnel (Plint & Partners TE-93) with 127 mm x 127 mm square cross-section was retrofitted with a custom made working section to allow for modular placement of a cartridge heater and specimen within the forced air stream as shown in Figure 6-4. The working section itself is fully customizable with interchangeable polycarbonate and aluminum panels. Aluminum honeycomb (0.5 in (12.7 mm) diameter) flow strengthener sheets were placed both upstream and downstream of the working section to enhance the homogeneity of the flow inside the wind tunnel. A sliding orifice plate at the discharge of the fan allows the wind tunnel flow rate to be throttled between 0 m/s to 30 m/s; flow velocities up to 20 m/s can be measured at any vertical position via a top side port with hot-wire anemometer (Model HHF42, Omega Company, Stamford, CT). Figure 6-4: Schematic diagram of the experimental setup.

106 89 Air temperature was measured with a K-type thermocouples with junction diameters of 0.6 mm placed in the air stream and air velocity was recorded with a hot-wire anemometer (Model HHF42, Omega Company, Stamford, CT) with a range of 0 to 20 m/s and a resolution of 0.1 m/s placed inside the testing section as shown in Figure 6-4. To avoid disturbing incoming air flow during experiments, the anemometer was removed and the port sealed once the wind tunnel ramped up and the air velocity had come to steady state. To measure the ambient air temperature, two thermocouples were placed in the approaching air stream, with one at the intake silencer baffle of the wind tunnel and another at the frontal area of the test section. Measurements from these two thermocouples were monitored to maintain a constant incoming air temperature of approximately 20 C during the experiments. All fins were tested in parallel to the incoming air flow configuration. Three-inch long electrical heaters (3618K403, High-Temperature Cartridge Heaters, McMaster- Carr) were used to provide a constant heat input to the fins during the experiments. To measure the applied heat flux; the first step was to calculate the derated wattage provided by the heater using the following equation given by the manufacurer Operating Voltage ( ) 2 Wattage at rated Voltage = Derated Wattage Rated Voltage (6-1) where operation voltage is the voltage provided by the power supply, rated voltage was given at 120 V by the supplier and wattage at rated voltage was also given at 200 W by the manufacturer. The tube surface temperature was measured using five type K thermocouples attached with 0.5 in (12.7 mm) spacing to the surface of the tube. The temperature distributions of the surface of the perforated sheet and wire mesh were acquired via an infrared (IR) camera. An IR camera was positioned in front of the wind tunnel upstream the fin as shown in Figure 6-4. All fins were

107 90 sprayed with high emissivity black paint (Figure 6-5), rated for high temperature use, to increase the emissivity of the surface to 0.95 and make it uniform. (a) (b) Figure 6-5: Fabricated fins after sprayed using high emissivity black paint on (a) flat plate, and (b) perforated sheet (Ø= in (4.75 mm)).

108 Temperature, ( C) Results and Discussion In this section the results for the plain tubes are presented first, followed by the results of the perforated sheet and wire mesh fins and a comparison between the two Plain Tube The performances of the plain tubes was investigated for three different air velocities of 3.7 m/s, 10 m/s, 15 m/s and at two different applied heater voltage of 15 and 20 V (corresponding to surface heat fluxes of 1.3 kw/m 2 and 2.3 kw/m 2 ). Figure 6-6 demonstrates the tube s surface temperature at different velocities. By increasing the velocity the heat transfer coefficient increased and therefore surface temperature dropped V 15V Velocity, (m/s) Figure 6-6: Temperature variation of the pipe at 15 V and 20 V (corresponding to surface heat fluxes of 1.3 kw/m 2 and 2.3 kw/m 2 ) applied voltage for different air velocities.

109 92 The heat transfer from the plain tube (Q Tube) was calculated and compared to the widely acceptable theoretical mode [45] for flow over a cylinder Nu = 0.683Re Pr 1/3 for Re = 40 to 4000 Nu = 0.193Re Pr 1/3 for Re = 4000 to 40,000 (6-2) The convective heat transfer coefficient, h a, was calculated using Equations (6-3) and Equation (6-4) h a = Q tube A ( T T ) (6-3) Nu D = hd k (6-4) where A is the surface area of the tube, T is the surface temperature, T is the air temperature, D is the diameter of the tube and k is the thermal conductivity of air at the film temperature of Tf = (T + T )/2. The Reynolds number was calculated using the following equation Re D = VD ν (6-5) where V is the velocity of air and ν is the kinematic viscosity of air. The experimental and theoretical models for flow over the plain tube are plotted in Figure 6-7 as the variation of Nu D with Re D based on tube diameter as the length scale.

110 Theoretical Model Experimental Results (15 V) Experimental Results (20 V) 50 Nusselt Number, Nu D Reynolds Number, Re D Figure 6-7: Comparison between the variation of (NuD) with (ReD) for experimental and theoretical model for flow over a cylinder. Assuming the heat loss to the surrounding is zero (q loss = 0) from the conservation of energy q heater = q tube (6-6) V 2 R = h ideal A ( T T )

111 94 The ideal heat transfer coefficient, assuming no heat losses due to conduction through the base of the heater, is h ideal = V 2 R A ( T T ) (6-7) In the experiment the heat loss to the surrounding is not zero (q loss > 0) therefore q heater = q tube + q loss V 2 R = h exp A ( T T ) + q loss (6-8) h exp = V 2 R q loss A ( T T ) Therefore, the measured heat transfer coefficient from the experiment (h exp ) is smaller than that expected in the ideal case when there are no losses. The results presented in Equation (6-8) confirms that the fin exhibits loss and therefore calculated Nusselt numbers are lower than the theoretical value. The percentage of q loss to the heat input varies between 8 to 12%. Considering a black surface radiation and a constant surface temperature of 350K, the radiative heat transfer is less than 1% of the total heat transfer. The radiative heat transfer is negligible due to low surface temperature Perforated Sheet Fins The performance of the perforated sheet fins was investigated for three different air velocities of 3.7 m/s, 10 m/s, 15 m/s and at three different applied heater voltages of 55, 60 and 65 V

112 95 (corresponding to surface heat flux of 17.7, 21.1 and 24.7 kw/m 2 based on the outer surface area of the tube). To ensure that steady state was reached during the experiment, the wind tunnel was operated for 30 minutes for each applied voltage. The experiments were performed at steady state, and the readings were taken when the thermocouple outputs were stabilized. The initial step was to compare the performance of the flat plate to the perforated sheet and analyze the temperature distribution along the surface of both structures. Figure 6-8 demonstrates the comparison between the flat plate and the perforated sheet for a 55 V heater voltage and at 10 m/s air velocity. Air velocity and the voltage applied to the heater were kept constant while fin temperature distribution was analyzed. For the flat plate, due to its non-porosity, the air could not pass through the plate and was forced to pass through the 1 in (25.4 mm) gap between the edge of the fin and the test section wall. Analysis of the temperature distribution, presented in Figure 6-8, demonstrated a greater heat transfer from the perforated sheet than the flat plate. Flat plate fin resulted in a higher surface temperature compare to the perforated sheet. This could be explained since the flat plate has a nonporous structure which results in an air velocity drop and consequently lower convection heat transfer rate.

113 96 (a) (b) Figure 6-8: IR map of the temperature distribution of the fins (a) Flat plate, and (b) Perforated sheet (Ø= in (4.75 mm)).

114 97 Figure 6-9: Comparison of the temperature profile at 55 V (17.7 kw/m 2 ) with a 10 m/s flow between the flat plate and perforated sheet (Ø= in (4.75 mm)). To better understand the surface temperature distribution along both fins, the temperature distribution was mapped as a function of fin length as shown in Figure 6-9. The x axis represents the distance along either the flat plate or the perforated sheet, which were 3 in (76.2 mm) wide and were bonded at their centers so that the fin length was 1.5 in (38.1 mm). The perforated sheet gave much higher heat transfer and consequently lower fin tip temperature of 26 C compare to 41 C for the flat plate. The perforated sheet also exhibited a maximum temperature of 69 C on the surface of the tube compared to 74 C for the solid plate. The sudden drop in temperature near the centerline of the solid plate shows that bonding with the tube was not

115 98 perfect. As seen in Table 6-3, holes were drilled in the flat plate to allow the thermal spray coating to penetrate and connect it to the tube. Better bonding may improve heat transfer from the tube. The temperature variation along a fin with an insulated tip is given by [45] T(x) T T b T = cosh m(l x) cosh ml (6-9) m = hp ka c where x is the distance from the fin base, A c the cross sectional area of the fin, h the convection heat transfer coefficient, p the perimeter, T b the temperature of the fin base and L the length of the fin. Figure 6-10 shows a comparison between the experimental temperature measurements and the theoretical temperature variation in which the heat transfer coefficient was adjusted to get the best agreement between the two. The convection heat transfer coefficients which best fitted the theoretical model to the experimental results were h =115 W/m 2 K for the solid plate and h =170 W/m 2 K for the perforated sheet. The model deviated from the experimental measurements near the base where the plates were attached to the tube because the contact was not perfect. The higher heat transfer coefficient of the perforated sheet was a result of the penetration of air through the holes in it.

116 Temperature, ( C) Flat Plate Theoretical [45] Flat Plate Experiment Perforated Sheet Experiment Perforated Sheet Theoretical [45] Length, (inch) Figure 6-10: Comparison between the measured surface temperature and predicted theoretical model. Figure 6-11 Shows the temperature profile at 60 V for the perforated sheet (Ø= in (4.75 mm)) at different air velocities.

117 Temperature, ( C) m/s 10 m/s 15 m/s Length, (inch) Figure 6-11: The temperature profile at 60 V (21.1 kw/m 2 ) applied voltage and for three air velocities for perforated sheet (Ø= in (4.75 mm)). The maximum temperature of 115 C, which was in the middle of the fin, decreases to 66 C as the velocity was increased from 3.7 to 15 m/s. The heat transfer coefficient varied from 61 to 224 W/m 2 K for 3.7 to 15 m/s respectively. To investigate the effect of change in heat flux on the surface temperature, the air velocity was kept constant at 10 m/s while the heat flux was varied from 17.7 kw/m 2 to 24.7 kw/m 2 (corresponding to voltages from 55 V to 65 V) as shown in Figure As the applied voltage increased from 55 V to 65 V, the maximum temperature increased from 69 C to 84 C. Some fluctuations are present in the experimental temperature profile due to the presence of holes on the perforated sheet.

118 Temperature, ( C) V 60V 65V Length, (inch) Figure 6-12: The temperature profile at 10 m/s air velocity and three different applied voltages for the perforated sheet (Ø= in (4.75 mm)). Figure 6-13 shows the surface temperature distribution for the perforated sheet fins which was recorded using an IR camera. The fin with the highest percentage of open area Perfo (a) (Figure 6-13a) resulted in the hottest tube temperature, which means the lowest heat transfer from the surface of the perforated sheet to the surroundings. At this point it is clear that there should be an optimized opening size that maximizes the heat transfer without producing a large pressure drop across the fin. To better understand the surface temperature distribution along fins, the temperature distribution was mapped as a function of the fin s length as shown in Figure The heat transfer coefficient varied from 170 to 231 W/m 2 K for Ø= in to Ø= respectively. The fin with the smallest hole diameter resulted in a better heat transfer and consequently the lowest tube temperature of 57 C compare to 70 C of the perforated sheet with 0.18 in (4.57 mm) hole diameter.

119 102 (a) (b) (c) Figure 6-13: Perforated sheet tested at 55V (17.7 kw/m 2 ) with a 10 m/s flow (a) Ø= in (4.76 mm), (b) Ø= in (3.17 mm), and (c) Ø= in (1.59 mm).

120 Temperature, ( C) Perfo (a) Perfo (b) Perfo (c) Figure 6-14: Temperature profile of the perforated fins at 55 V (17.7 kw/m 2 ) applied voltage and 10 m/s air velocity. Length, (inch) Wire Mesh Fins Figure 6-15 shows the surface temperature distribution of the three wire mesh fins for a constant air velocity of 10 m/s and applied voltage of 55 V (17.7 kw/m 2 ). The 10 PPI wire mesh fin had the highest percentage of open area (56%), resulting in the lowest tube temperature compared to the 20 PPI wire mesh fin that has the lowest open area (46%). 20 PPI wire mesh resulted in the smallest heat transfer from the surface of the perforated sheet to the surroundings. The IR results

121 104 demonstrated that the worst connection exists for 20 PPI wire mesh since the thermally sprayed particles could not effectively penetrate between the pores (Figure 6-15). As the pore size increases from 20 PPI to 10 PPI, the quality of the connection increased between the wire mesh and the tube since the particles could penetrate between the pores to connect the mesh to the tube. To better understand the surface temperature distribution along three fins, the temperature distribution was mapped as a function of the fin s length as shown in Figure (a) (b) (c) Figure 6-15: Experimental temperature distribution at 55V (17.7 kw/m 2 ) applied power with a 10 m/s air velocity (a) 10 PPI, (b) 14 PPI, and (c) 20 PPI.

122 Temperature, ( C) 105 The fin with the largest pores, 10 PPI, resulted in better heat transfer rate and consequently lower tube temperature of 71 C compare to 82 C of the 20 PPI mesh. The maximum surface temperature seems to increase with the porosity. Smaller pores clog easily during the thermal spraying process and therefore prevent good bonding of tube with the fin (Figure 6-15c). In addition, as it is shown in Figure 6-16, the lowest temperature was almost the same for 10 & 20 PPI, which means the 20 PPI configuration transfers as much heat as 10 PPI due to its high wire mesh surface area. 14 PPI mesh resulted in the highest fin temperature of 23 C on the end of the fin PPI 14 PPI 20 PPI Figure 6-16: Temperature profile for different wire mesh at 55 V (17.7 kw/m 2 ) applied power and a 10 m/s air velocity. Length, (inch)

123 Temperature, ( C) PPI wire mesh sheet fin was further investigated by varying the air velocity while keeping the applied voltage constant, as shown in Figure The maximum temperature of 128 C, which is in the middle of the fin, decreases to 71 C as the air velocity was increased from 3.7 to 15 m/s m/s 10 m/s 15 m/s Length, (inch) Figure 6-17: Temperature profile of 14 PPI at 60 V (21.1 kw/m 2 ) applied power for different air velocities. To better compare the performance of wire mesh to perforated sheet the surface temperature distribution for the three perforated sheets and 10 PPI wire mesh was plotted in Figure The temperature distribution was mapped as a function of the fins. 10 PPI wire mesh resulted in the

124 Temperature, ( C) 107 highest surface temperature of 71 C but it had the lowest temperature profile across the measured line. The low surface temperature of the wire mesh is due to its high permeability compare to perforated sheets. Wire mesh is the best configuration as the temperature quickly drops and a reduced fin size of approximately 1.5 inch (38.1 mm) would be sufficient since the fin temperature reaches the ambient air temperature Perfo (a) Perfo (b) Perfo (c) Woven wire mesh Length, (inch) Figure 6-18: Temperature profile comparison between the wire mesh and perforated sheets at 55V (17.7 Kw/m 2 ) applied power with a 10 m/s air velocity.

125 Heat Transfer Characterization The total heat transfer from the tube and fins is a combination of heat transfer from the plain tube and from the porous structure connected to the tube. To calculate the enhancement due to the porous structure, the heat transfer from the tube was subtracted from the total heat transfer as shown in Equation (6-10). The tube temperature was applied to previously found relationship of plane tube and then the heat transfer from the tubes, Q tube, was calculated Q total = Q mesh + Q tube (6-10) Q mesh = Q perp + Q para (6-11) The heat transfer from the wire mesh Q mesh consistd of the heat transfer from the wires which were perpendicular to the tube (Q perp ) and wires which were parallel to the tube (Q para ) as shown in Equation (6-11). To calculate Q mesh the wires were considered as long pin fins using the following equation [45] Q perp = h a P f k f A f.c (T b T ) tanh(ml) (6-12) m = h a P f /k f A f.c Q para = h a A f.c ( T bn T ) (6-13) where P f and A f.c are the perimeter and cross-sectional area of the fin respectively, L is the length, k f the thermal conductivity and T bn is the base temperature of the wire in the n th row, which was measured using the IR camera. The wire mesh heat transfer is calculated by subtracting the heat

126 109 transfer of plain tube from the total heat transfer. The heat transfer coefficient was calculated by substituting Equation (6-12) and (6-13) to Equation (6-11). The heat transfer coefficient for the perforated sheet and the flat plate were also estimated with long fin equations. Figure 6-19 shows the Nusselt number (Nu D ) variation with the change of Reynolds number (Re D ) for both the 14 PPI wire mesh and 0.18 hole diameter perforated sheet. 14 PPI mesh fin resulted in a higher Nu D at different Re D than the perforated sheet PPI Mesh 0.18" Perforated 120 Nusselt Number, Nu D Reynolds Number, Re D Figure 6-19: Variation of Nusselt number (NuD) as a function of Reynolds number (ReD) based on tube outer diameter (OD) for wire mesh and perforated sheet.

127 Nusselt Number, Nu D PPI Mesh 14 PPI Mesh 20 PPI Mesh 0.06" Perforated 0.12" Perforated 0.18" Perforated Flat sheet 0 Figure 6-20: Performance chart of the fabricated fins at a constant (ReD) of All the fins were also compared at a constant Re D of 5290 as shown in Figure In general all wire mesh fins had higher Nu D than perforated sheet fins. The highest Nu D corresponds to the 10 PPI wire mesh which has the largest pore size. The Nu D reduced when wire mesh pore size was decreases to 20 PPI. The flat plate fin has the lowest Nu D at a given Re D. The perforated sheet structure creates large resistance to the air flow because the solid portion of the perforated sheet is perpendicular to the direction of the flow. This air blockage creates stagnation (low velocity) regions that have low heat transfer coefficient and reduce the overall heat transfer performance of perforated sheets compared to wire mesh screens.

128 111 By comparing the Nu D for three different hole size perforated sheets of 0.18, 0.12 and 0.06 with non-perforated surface area ratio of 49%, 60% and 70% (Table 6-1), It was concluded that 0.06 diameter hole perforated sheet has the highest Nu D. This value is the direct contribution of the non-perforated surface area (Table 6-1). In other words, the perforated sheet with 70% surface area ratio had the highest heat transfer area among those three perforated sheets, independent of the hole size. Values are tabulated as follows Table 6-4: Comparison between the variation of NuD and Anon-perf. Hole Diameter A non perf A total Nu D Nu D A non perf A total % % % The experimental data suggest that Nu D is proportional to the non-perforated surface area. Nu D A non perf A total (6-14) These results confirms that the higher heat transfer in the 0.06 hole diameter perforated sheet is due to its higher surface area.

129 PPI Wire Mesh 0.18" Perforared Sheet Venugopal, Balaji and Venkateshan (0.92 Porosity) [43] Venugopal, Balaji and Venkateshan (0.89 Porosity) [43] Venugopal, Balaji and Venkateshan (0.85 Porosity) [43] Li et al. (copper wire screens) [42] Nusselt Number, Nu H Reynolds Number, Re H Figure 6-21: Nusselt number (NuH) variation as a function of Reynolds number (ReH). Data found in literature was for a stack of wire meshes, as shown in Figure Li et al. [42] obtained Nu H based on channel height of 10 mm as characteristic length for copper wire screens and Venugopal, Balaji and Venkateshan [43] have also looked into stacks of perforated sheets and presented their results based on channel height. Calculated results were also converted to Nu H base on the hydraulic diameter of the wind tunnel for the sake of comparison. The values of Nu H found in the experiment were in agreement by the results of Li et al. [42] and Venugopal, Balaji and Venkateshan [43]. The difference between the calculated experimental results and the results found in the literature can be due to the existence of bypass flow around porous screens.

130 Fin Effectivness, ɛ 113 The performance of the fins was further analyzed based on their fin efficiency and effectiveness as shown in Figure Fin Efficiency (ɳ) and Effectivness (ɛ) were calculated using the following equations [45] Fin Efficiency, longfin (ɳ ) = Q fin Q fin,max = 1 ml (6-15) Fin Effectivness, longfin (ɛ) = Q fin = kp Q no,fin ha (6-16) PPI Mesh 14 PPI Mesh 0.18" Perforated 0.12" Perforated 10 PPI Mesh 0.06" Perforated Fin Efficiency, ɳ Figure 6-22: Comparison between the fin efficiency (ɳ) and effectiveness (ɛ) of the fins.

131 114 The results from Figure 6-22 indicate that wire mesh fins were more effective than perforated sheet fins. As can be seen from Figure 6-18, the temperature of the wire mesh dropped much faster and reached the ambient air temperature faster than the perforated sheet. This finding indicates that the same heat transfer could have been achieved with a shorter fin length. The high fin efficiency can also be explained using Figure The maximum surface temperature was much lower for the perforated sheet fin than the wire mesh. The wire mesh was not as efficient as perforated sheet which resulted in a lower overall heat transfer from the hot tube. By increasing wire or plate thickness m in Equation (6-15) decreases, therefore fin efficiency increases. The heat exchangers and experimental setup for the next chapter was designed to further enhance the understanding of heat transfer in wire mesh heat exchangers. The temperature distribution along each wire needs to be analyzed to better understand the enhancement due to the wires which were perpendicular to the primary wires, which were also connected to the body of the tube heat exchanger.

132 Conclusion In this chapter the heat transfer characteristics of the perforated and wire mesh porous structures have been investigated. Aluminum fins were fabricated by connecting aluminum wire mesh, perforated sheet and flat plate to aluminum tubes using wire-arc thermal spray coating. The results indicated the importance of the quality of contact between tubes and porous structure. The porous structures with high open area allowed good penetration of the coating material into the gap between wire and tube surface, and thus providing good adhesion and thermal conduction. The fins were tested inside a wind tunnel and their surface temperature was measured. Significant increase in heat transfer were achieved by attaching wire mesh or perforated sheet to the plain tube. All fins resulted in larger heat transfer rate than the flat plate. The performance of the wire mesh fins was affected by pore density that affects air penetration. The extended surfaces of the perforated sheet and wire mesh enhanced heat transfer from the tube to the surrounding air inside the wind tunnel. Wire mesh is the best configuration as the temperature quickly drops and a reduced fin size of approximately 1.5 in (38.1 mm) would be sufficient since the fin temperature reaches the ambient air temperature and there is no need for a 3 in. Wire mesh fins were more effective than perforated sheet fins but less efficient.

133 116 Water-to-Air Wire Mesh Heat Exchangers 7.1 Introduction The purpose of this study was to fabricate a high temperature gas to liquid wire mesh heat exchanger, and to measure heat transfer through the wire mesh. The wire mesh screens were bonded to the outer surface of tubes using thermal spraying. Experiments were done in which water cooled 5, 10, and 20 pores per inch (PPI) wire mesh heat exchangers were placed inside a chamber with an air temperature of 320 ± 20 C at the test section. To measure the heat transfer enhancement, compared to a plain tube heat exchanger, the temperature rise of the water between the inlet and outlet of the heat exchanger was measured for three different water flow rates, varying from 500 to 900 ml/min. A high temperature infrared camera was used to study the surface temperature, investigate the wire mesh fin efficiency and effectiveness, and to investigate the connection between the wire mesh and the tube.

134 Fabricated Heat Exchangers Wire Mesh Woven wire mesh screens with pore densities of 5, 10, and 20 PPI were used to make heat exchangers. The ideal wire mesh pore density should correspond to a high ratio of surface area to occupied volume, without reducing air penetration. Additionally, the size of the pores should be large enough to permit thermal spray particles to penetrate, and provide sufficient mechanical bonding between the mesh and the tube, reducing the thermal resistance and ensuring high thermal contact. The screen dimensions were identical for all tested samples with the dimensions of 152 mm x 203 mm (L m x W m ) in an attempt to investigate and compare the efficiency and effectiveness of the extended surface area of porous materials for the same occupied area Fabrication Process Six heat exchangers were fabricated by bonding stainless steel tubes and wire mesh screens using a thermal spray process. Each heat exchanger was composed of four tubes with 6.3 mm (0.25 in) diameter outer diameter, 178 mm (7 in) length, and 0.25 mm (0.01 in) wall thickness; with 57 mm (2.25 in) center to center tube spacing. The first three heat exchangers were fabricated by thermal spraying a 5, 10, or 20 PPI wire mesh screen on top of the tubes, respectively, as shown in Figure 7-1. A second set of heat exchangers was fabricated by connecting one wire mesh screen on top, and another on the bottom, of the tubes as shown in Figure 7-2b. The thermal contact resistance between the wire mesh and the tube surface is low since the wire mesh is simply placed on top of the tube to simplify the manufacturing process. The thermal contact resistance would have increased if the wire mesh was wrapped around the tube to increase the contact area between them.

135 118 (a) (b) (c) Figure 7-1: Sample of heat exchangers (a) single screen 5 PPI wire mesh, (b) single screens 10 PPI wire mesh, and (c) single screens 20 PPI wire mesh.

136 119 Stainless steel tubes were connecting to each other by six 90 degree elbow compression fittings in order to form a path for the coolant. The important parameters of the fabricated heat exchangers used in this study are summarized in Table 7-1. Table 7-1: Parameters of the wire mesh heat exchangers. Samples Pore Density (PPI) Pore Size (m²) Open Area (%) Number of Screens Wire Diameter (m) Mesh Surface Area, A m (m²) Total Surface Area, A Total (m²) Porosity Heat Ex 1 Heat Ex 2 Heat Ex 3 Heat Ex 4 Heat Ex 5 Heat Ex 6 Heat Ex 7 N/A N/A N/A 0 N/A N/A N/A 5 1.5x x x x x x x x x x x x Each wire of the screen mesh is modeled as a cylinder. The mesh surface area calculations were performed by accounting for the number of layers in the x and y directions, the wire diameters, and the dimensions of the mesh. The total surface area (A Total ) is the sum of the mesh surface area (A m ) and the tube surface area (A t ). The tube surface area is simply the surface area of the tube without the wire mesh screen. In order to ensure a uniform emissivity over the surface of all heat exchangers, they were painted with layers of black, high temperature, thermally conductive paint with emissivity of 0.95, as shown in Figure 7-2.

137 120 (a) (b) Figure 7-2: Sample heat exchangers (a) single screens 5 PPI wire mesh, and (b) double screens 5 PPI wire mesh.

138 Experimental Apparatus and Methods The experimental apparatus consisted of an open loop water system and a hot gas chamber. The coolant flow passing through the tube was maintained by a water circulation loop. A schematic representation of the experimental setup and the hot air chamber are shown in Figure 7-3 and Figure 7-4, respectively. The high temperature air chamber was designed to create a steady high temperature environment to test the fabricated heat exchangers. The channel was positioned vertically and an experimental rig was designed and fabricated to hold the structure. The test section, which was designed to hold the heat exchanger perpendicular to the direction of the hot gas flow, consisted of an inlet and outlet connection to bring water into and out of the heat exchanger. Figure 7-3: Schematic representation of the experimental setup.

139 122 The secondary chamber was designed with the same dimensions as the main test section to permit a uniform high temperature air flow to reach the heat exchanger, and to avoid recirculation of the outer air to the test section. Figure 7-4: Schematic representation of the hot air chamber. The heat exchanger is located in the test section (Section 3), perpendicular to the direction of the air flow. The hot air passing through the chamber was supplied by an electrical air heater (F076029, SKORPION AIR HEATERS, OSRAM SYLVANIA, Exeter, NH). The velocity flow field inside the wind tunnel was measured using a pitot tube (P06A Pitot Static Probe, FlowKinetics,

140 123 Texas) with an accuracy within ±0.24% of full scale velocity. Water flowed through a damper and pressure regulator to prevent the flow rate from fluctuating. A valve was placed before the flow meter to adjust the flow and the flow rate measured a flow meter, with a range of 0.1 to 1 L/min with an accuracy of 1% of full scale, before the water entered the heat exchanger. The water flow circulated through the heat exchanger while air was forced over it. Two type-k thermocouple probes were attached to the inlet and the outlet of the heat exchangers to measure the inlet (T i ) and outlet (T o ) water temperatures. An IR camera was used to record the heat exchanger surface temperature during the experiment. The average air temperature was measured using 15 Type-K thermocouples at section 3, located before the heat exchanger.

141 Pressure Drop Through Wire Mesh Screens The hydraulic performance of wire mesh screens alone, without any tubes, and the effect of spacing between screens on the air pressure drop were also investigated. The experimental apparatus used to test the wire mesh samples consists of a horizontal wind tunnel with a square chamber of 127 mm x 127 mm cross section. An aluminum flow straightener was used downstream of the experimental test section to achieve a uniform flow. The velocity flow field inside the channel was measure using a hot wire anemometer (Model HHF42, Omega Company, Stamford, CT) with a range of 0 to 20 m/s, and a resolution of 0.1 m/s. The pressure drop was measured using a digital manometer (Model HHP-103, Omega Company, Stamford, CT) set to a maximum range of 498 Pa (2 inh2o) with an accuracy of 0.2% of full scale at the beginning and end of the test sections. Two different pore densities of 20 PPI and 10 PPI stainless steel were chosen, while a fixed width and height of 127 mm was used for all the samples. The effect of providing distance between wire mesh screens on the pressure drop were also analyzed providing 13 mm and 25 mm (0.5 and 1 in) spacing between wire mesh screens. As the wire mesh pore size increases from 20 PPI to 10 PPI, air flow resistance through the wire mesh decreases, which results in higher air penetration through the mesh and increase in permeability as shown in Figure 7-5.The permeability and air flow resistance of the wire mesh plays an important role in hydraulic characteristic of wire mesh heat exchanger.

142 Pressure Drop, ΔP (Pa) PPI - 0.5" Space 20 PPI - 1" Space 10 PPI - 0.5" Space 10 PPI - 1" Space Velocity, V (m/s) Figure 7-5: Variation of experimentally measured pressure gradient with average fluid velocity in channels for 20 PPI and 10 PPI wire mesh screen.

143 Temperature Rise, ( ) Results and Discussion The performance of the fabricated stainless steel heat exchangers was investigated for three different water flow rates in the range of L/min and an air temperature of 320 ± 20 C. Fabricated wire mesh heat exchangers outperformed the bare tube heat exchanger, as seen in Figure 7-6, which demonstrates the effectiveness of the wire mesh as a heat transfer enhancer plain tube 5 PPI 10 PPI 20 PPI 10 8 (a) Water Volume Flow Rate, (L/min)

144 Temperature Rise, ( ) plain tube 2 sheets 20 PPI 2 sheets 10 PPI 2 sheets 5 PPI 10 8 (b) Water Volume Flow Rate, (L/min) Figure 7-6: Temperature rise of water flowing through the tubes (a) heat exchangers with one wire mesh screen, and (b) heat exchangers with two wire mesh screens. As seen in Figure 7-6a, single screen wire mesh heat exchangers resulted in an 84% higher temperature rise compared to the plain tube heat exchanger. When comparing the performance of the 2 screen wire mesh heat exchanger. 2 screen 10 PPI wire mesh outperformed the other heat exchangers (Figure 7-6b). For both 10 PPI and 5 PPI heat exchangers, having two wire mesh screens was better than one. Air penetrated more easily through the 5 and 10 PPI than the 20 PPI wire mesh, since the pore sizes were much larger. High pore density wire mesh caused extra resistance to air flow, which reduced forced convection heat transfer. For the case where a heat exchanger was built using two screen wire mesh (Figure 7-2b) the heat transfer performance increased by 130 %, 105 %, and 76 % for 10, 5 and 20 PPI wire mesh respectively, compared to the plain tube heat exchanger.

145 128 Heat transfer was always enhanced since the extended wires were used to increase the overall surface area, and also increase the convective heat transfer. To better understand this variation in surface temperature, the average air temperature before the heat exchanger is shown in Figure 7-7. The average air temperature was measured by positioning sixteen K-type thermocouples with junction diameters of 0.6 mm in the air stream before the heat exchanger and averaging their temperatures at steady state. The air temperature increased as the pore density of the mesh increased from 5 PPI to 20 PPI. Double screen heat exchangers resulted in a higher temperature rise compared to the plain tube and single screen wire mesh heat exchangers, as shown in Figure 7-6. The air accumulates behind the heat exchangers because of the pressure drop due to the porous geometry of the wire mesh. The pressure drop test across wire mesh in the wind tunnel experiment further proved the variation in air penetration through these structures. In the case where heat exchangers were fabricated using only one wire mesh screen on average, the wire mesh surface temperatures were 300 C, 312 C and 327 C for 5, 10 and 20 PPI wire mesh, respectively.

146 Air Temperature, ( C) PPI, 1 Screen 5 PPI, 2 Screen 10 PPI, 2 Screen 20 PPI, 2 Screen PPI, 1 Screen 5 PPI, 1 Screen Plain Tube 260 Figure 7-7: Variation of average air temperature at section 3 for different PPI wire mesh heat exchangers.

147 Heat Transfer Characterization Non-Dimensional Parameters The following experiments were conducted by measuring the temperature rise of the water between the inlet and outlet of the heat exchanger for three different water flow rates, from 500 to 900 ml/min. In these experiments water cooled heat exchangers were placed inside a high temperature chamber with an air temperature of 320 ± 20 C and average air velocity of 1 m/s. The water and air inlet and outlet temperatures were measured by placing thermocouples at the corresponding inlet and outlet. The heat transfer rate of water as calculated using Q = m w C p.w T w (7-1) where C p is specific heat, m is mass flow rate and subscript w stands for water flow loop. The overall heat transfer coefficient U and log mean temperature T LMTD were calculated using the equations below Q = UA t T LMTD (7-2) T LMTD = T 1 T 2 ln ( T 1 T 2 ) (7-3) where A t is the tube outer surface area and T 1 and T 2 represent the temperature difference between two fluids at the two ends (inlet and outlet) of a heat exchanger. The heat transfer coefficient of air, h a, is found using the equation:

148 Overall Heat Transfer Coefficient,U (W/m 2 K) UA t = 1 h w A 1 + ln ( D 2 D 1 ) 2 k t L t + 1 h a A 2 (7-4) where D 1 and A 1 was the inner tube s diameter and area and D 2 and A 2 was the outer tube s diameter and area, L t is the length of the tube and k t is the thermal conductivity of the stainless steel tube PPI, 2 Screen PPI, 1 Screen 5 PPI, 2 Screen 10 PPI, 1 Screen 20 PPI, 1 Screen PPI, 2 Screen Figure 7-8: Variation of the Overall heat transfer coefficient across different pore densities. To study the heat transfer coefficient, a graph is plotted with wire mesh PPI on the X axis and overall heat transfer coefficient on the Y axis (Figure 7-8). For the case of 5 and 10 PPI wire mesh, double screen heat exchangers outperformed single screen heat exchangers. This relation turns reverse in 20 PPI heat exchanger, where single screen outperformed double screen wire mesh heat exchanger.

149 132 The heat transfer coefficient between the water and the inner surface of the tube can be calculated using standard correlations. The value of Nu w has a constant value of 4.36 for fully developed flow since the flow inside the tube is laminar Re w < 2300 [45]. Reynolds and Prandtl number for water flow in the tube can be calculated using the equations, Re w = V wd 1 ν w (7-5) Nu w = h wd 1 k w (7-6) Pr w = μ w C p.w k w (7-7) where V w is the velocity of water in the tube, μ w the dynamic viscosity, C p.w the specific heat, ν w the kinematic viscosity and k w the thermal conductivity of water. After calculating the heat transfer coefficient ha using the Equation (7-4), the Nusselt number on the outer surface of the tube was calculated using Nu a = h a D 2 k a (7-8) where k a is the thermal conductivity of air. Figure 7-9 shows the Nusselt number (Nu a,d ) variation with mesh size for a constant water mass flow rate of Kg/s. The Nusselt numbers for both single mesh heat exchangers (1 Screen) and double mesh heat exchangers (2 Screen) were plotted. The single mesh heat exchangers all had similar values of Nusselt number (Nu a,d ), approximately 12, indicating that the flow fields

150 133 were similar in all cases and not strongly affected by the mesh. It is apparent that the 10 PPI double mesh heat exchanger has the highest Nusselt number (Nu a,d ) followed by the 5 PPI double mesh. The 20 PPI double mesh heat exchanger has the lowest Nusselt number (Nu a,d ). Increasing the mesh surface area promotes heat transfer, but after a certain point blocking of the air flow leads to a decrease in heat transfer PPI, 2 Screen 15 Nu a,d PPI, 1 Screen 10 PPI, 1 Screen 20 PPI, 1 Screen 5 PPI, 2 Screen 20 PPI, 2 Screen 10 Figure 7-9: Nusselt number variation (Nua,D) across different pore densities at a constant water mass flow rate of Kg/s. Calculated results were also compared to Nusselt number (Nu H ) variation with the change of Reynolds number (Re H ) relations for stack wire mesh screens, H is the channel height, found in the literature as shown in Figure Calculated heat transfer coefficients were generally greater than valued reported by Li et al. [42] and Venugopal, Balaji and Venkateshan [43]. This can be

151 134 explained by the fact that in this study thermal sprayed heat exchangers were used to reduce interfacial thermal resistance. Nusselt Number, Nu H PPI, 2 Screen 5 PPI, 2 Screen 10 PPI, 1 Screen 5 PPI, 1 Screen 20 PPI, 1 Screen 20 PPI, 2 Screen Venugopal, Balaji and Venkateshan (0.89 Porosity) [43] Li et al. (copper wire screens) [42] Venugopal, Balaji and Venkateshan (0.92 Porosity) [43] Venugopal, Balaji and Venkateshan (0.85 Porosity) [43] Reynolds Number, Re H Figure 7-10: Nusselt number (NuH) variation as a function of Reynolds number (ReH).

152 Empirical Fin Model Correlation To observe the effect of the wire mesh on heat transfer, infrared (IR) images of the heat exchangers were captured and analyzed. Figure 7-11 shows sample IR images for 5, 10, and 20 PPI wire mesh heat exchangers inside the hot air chamber. The infrared images demonstrate the effectiveness of the wire mesh fin, and also the flaws in the tube-mesh connection. The IR results demonstrated that the worst connection exists for 20 PPI wire mesh, since the thermally sprayed particles could not effectively penetrate through the pores. As the pore size increases from 20 PPI to 5 PPI, the quality of the connection between the wire mesh and the tube increased since the particles could penetrate through the pores and connect the mesh to the tube.

153 136 (a) (b) (c) 90 ⁰C 200 ⁰C 250 ⁰C 290 ⁰C 315 ⁰C Figure 7-11: IR camera surface temperature variation across heat exchangers for (a) 5 PPI, (b) 10 PPI and, (c) 20 PPI.

154 137 Heat transfer to the tube is limited by to the thermal contact resistance between the wire mesh and the tube. The analysis begins with comparing the effectiveness of the wire mesh with a long fin and then finding contribution of the transverse wires as illustrated in Figure Y X Figure 7-12: Schematic of eleven transverse and one longitudinal wire between two tubes. The variation of temperature along the long fin with a constant cross section (Ac = constant) can be model using the following equation T fin tip = T T(x) = T + (T b T )e x hp KA c (7-9)

155 138 where A c is the cross-sectional area of the fin at location x, p is the perimeter of a fin, h is the convection heat transfer coefficient, k f is thermal conductivity of the fin, T is the temperature of the surrounding and T b is a fin base temperature. IR camera images of 5 PPI wire mesh (Figure 7-11a) were post processed and the experimental wire mesh and tube surface temperatures were measured. 5 PPI wire mesh was chosen since it was easier to analyze due to the thickness of its wires. To find the contribution of transverse wires to the overall heat transfer enhancement due to wire mesh screens, the surface temperature along a single longitudinal wire and eleven transverse wires perpendicular to it were analyzed. Eleven temperature probes were used to map the temperature distribution of the transverse wires and one to map the temperature of the longitudinal wire (Figure 7-12). The temperature variation along the longitudinal wire, located on the x-axis of schematic Figure 7-12, is plotted in Figure The transverse wires were located perpendicular to the longitudinal wires, with a fixed x coordinate as shown in Figure The y axis in Figure 7-12 was used to map the temperature distribution along the transverse wires (Figure 7-14). As can be seen from Figure 7-13, the temperature of the longitudinal wire increases slowly from where the wire is connected to one tube, until it reaches its maximum half way between the cold tubes. The temperature of the longitudinal wire eventually reached the temperature of the ambient temperature. It was found that the transverse wires were conducting heat towards the longitudinal wire, since the temperature of the longitudinal wire was lower at the point of overlap with the transverse wires, Figure 7-13.

156 Surface Temperature, ( C) Longitudinal Wire T1 T2 T3 T4 T5 T6 T7 T8 T9 T10 T Location Along x Axis, (mm) Figure 7-13: Wire surface temperature variation along the length of one longitudinal and eleven transverse wires, measured experimentally using IR camera, for the 5 PPI wire mesh heat exchanger. The x and y axis are shown Figure The y axis in Figure 7-12 is used to map the temperature distribution along the longitudinal and six transverse wires (T1, T2, T3, T4, T5, and T6), which is plotted in Figure The temperature distribution along the longitudinal wire is shown as a vertical line in Figure 7-14.

157 Surface Temperature, ( C) Longitudinal wire T6 T5 T4 T3 T2 T Location Along y Axis, (mm) Figure 7-14: Wire surface temperature variation along the length of a longitudinal and six transverse wires (T1, T2, T3, T4, T5, and T6 as shown in Figure 7-12) for the 5 PPI wire mesh heat exchanger. Temperatures were measured experimentally using IR camera. Figure 7-14 demonstrates the conductive heat transfer from transverse wires (T1, T2, T3, T4, T5, and T6 as shown in Figure 7-12) to the longitudinal fin. The temperature of the transverse wires were higher at their point of contact with the longitudinal wire. Since the presence of transverse wires results in a higher heat transfer rate, the conventional straight fin model will not predict the heat transfer from the wire mesh heat exchangers accurately.

158 141 Wire mesh screens as shown in Figure 7-15 were similar to longitudinal fin structures but the temperature distribution will differ due to the presence of transverse wires perpendicular to the longitudinal. The effect of transverse wires on the heat transfer was investigated in this study by analyzing the surface temperature distribution of the wire mesh screen. Figure 7-15: Location of longitudinal and transverse wires of wire mesh screens. If l is the length of a pore and d is a fin diameter then the longitudinal surface temperature variation of the first pore can be expressed as s = { 2 l + x l + d 0 x < l d d (x l) l d x < l (7-10) T(x) = T + (T b T )e ms (7-11)

159 142 m = hp KA c where s is the actual length of the wire as a function of x, and x is the location on the longitudinal wire. Substituting Equation (7-10) into Equation (7-11) results in the temperature variation for the longitudinal wire for the first pore T + (T b T )e mx T(x) = { T + (T b T )e m [2 l+ l+d d (x l)] 0 x < l d l d x < l (7-12) The temperature distribution of the second pore can also be analyzed by ms (x) T(x) = T + (T b T )e 2l + (x l) s = { l + d 4 l + (x 2l) d l x < 2l d 2l d x < 2l (7-13) The temperature distribution along the nth pore can be calculated using ms (x) T(x) = T + (T b T )e (n 1)l + x s = { l + d n l + (x nl) d (n 1)l x < n l d nl d x < n l (7-14) Surface measurements conducted using IR camera were used to validate Equation (7-14) as shown in Figure Appendix A shows the Matlab code for the empirical fin model.

160 Temperature, ( C) 143 Empirical Fin Model 300 Experimental Result Length, X (mm) Figure 7-16: Comparison between the measured surface temperature using IR camera and predicted empirical model.

161 A Model for Prediction of Heat Exchanger Temperature Rise The objective of this section was to obtain a model for predicting the performance of the tested heat exchangers for various inlet flow rates and surface area. In order to predict the outlet temperature of the heat exchangers the known inlet and outlet temperatures of the tested heat exchangers were used as a base line. Next a non-dimensional number was extracted which was used for obtaining (estimating) the unknown outlet temperature of the same heat exchanger with a different flow rate. Figure 7-17: Heat transfer energy balance for the fabricated heat exchangers. Figure 7-17 shows the energy balance, Equation (7-15), for a small section of a heat exchanger. From the conservation of energy m C p T x m C p (T x + T x dx) = h da (T s T ) (7-15) If T s = T fluid,inside = T, then

162 145 m C p T x dx = U P dx ( T T ), where da = P dx T U P + ( ) T = T x m C p U P m C p (7-16) If the form of T(x) = ke tx + C, then by substituting in to (7-16) the differential equation t ke tx + U P m C p (ke tx + C) = U P m C p T (7-17) e tx (k U P U P tk) + C = T m C p m C p U P m C p Therefore U P C = T, and t = m C p U P x m C T(x) = ke p + T (7-18) Since P = A surf, where A L surf is the overall area of the mesh and the tube and L is the length of the tube (4 x178 mm (7 in)), then we can introduce a number of transfer units (NTU ) based on water side of the heat exchanger NTU = U A surf m C p (7-19) Substituting Equation (7-19) into Equation (7-18) T(x) = ke NTU x L + T (7-20)

163 146 By applying the boundary conditions at x = 0 and x = L we get T in = ke NTU ( 0 L ) + T (7-21) Therefore k = T in T Also T out = (T in T ) e NTU ( L L ) + T (7-22) Therefore NTU = ln (T in T ) (T out T ) (7-23) Depending on the known variables, NTU can be used for calculating (predicting) the outlet temperature of the heat exchanger (T out ), as shown in Table 7-2, or if T out is known, equation (7-23) can be used to calculate NTU. By comparing NTU with the well know definition of NTU NTU = U A m C p (7-24) Which can be simplified to NTU = U A (T in T out ) m C p (T in T out ) ΔT LM ΔT LM (7-25) Since U A ΔT LM = m C p (T in T )

164 147 NTU = (T in T out ) ΔT LMTD (7-26) NTU = (T in T out ) (T in T out ) ln (T in T ) (T out T ) = ln (T in T ) (T out T ) (7-27) The equation (7-23) and equation (7-23) were identical therefore it was concluded that NTU = NTU. Table 7-2: Parameters of the wire mesh heat exchangers at a water mass flow rate of Kg/s. Samples Pore Density (PPI) Number of Screens T in ( C) T out ( C) T ( C) NTU Eq.(7-23) T out Eq.(7-22) Heat Ex 1 Heat Ex 2 Heat Ex 3 Heat Ex 4 Heat Ex 5 Heat Ex 6 Heat Ex 7 N/A T out = (T in T ) e T T out = (T in T ) e T T out = (T in T ) e T T out = (T in T ) e T T out = (T in T ) e T T out = (T in T ) e T T out = (T in T ) e T

165 148 For a cases with different flow rates (m ) a new NTU can be estimated using NTU new = NTU old ( m old ) (7-28) m new Table 7-3 compares the experimentally calculated non-dimensionalize number NTU, calculated using Equation (7-23) for a water mass flow rate of Kg/s to the predicated nondimensionalize number NTU new calculated using Equation (7-28). The calculate nondimensionalize numbers NTU were in a good agreement with the predicted non- dimensionalized number NTU new. Table 7-3: Parameters of the wire mesh heat exchangers at a water mass flow rate of Kg/s. Samples Pore Density (PPI) Number of Screens T in ( C) T out ( C) T, ( C) NTU Eq. (7-23) NTU old Table 7-2 NTU new Eq.(7-28) Heat Ex 1 N/A Heat Ex Heat Ex Heat Ex Heat Ex Heat Ex Heat Ex

166 149 We can also estimate an outlet temperature (T out ) of a larger heat exchanger using the calculated NTU from this chapter while considering the change of surface area as well. NTU new = NTU old ( m old ) ( A new ) (7-29) m new A old Figure 7-18: Schematic of 3 heat exchangers connected in series. If multiple heat exchangers with known NTU were connected in series (Figure 7-18), the exit temperature from each heat exchanger could be estimated as follow T out,1 = (T in,1 T ) e NTU + T T out,2 = (T out,1 T )e NTU + T (7-30) T out,3 = (T out,2 T ) e NTU + T

167 150 Extended surface area ratio was defined as R A = A surf A tube (7-31) Figure 7-19 shows the variation of NTU with different extended surface ratios. All of the cases were compared against the bare tube heat exchanger. The highest NTU was achieved for the 10 PPI single screen heat exchanger. In other words, while single screen of 10 PPI wire mesh didn't have the highest surface area for heat transfer but it was the most effective one by yielding the highest NTU. It can also be noticed that the double screen 5 PPI HEX outperformed the single screen. It is different from the pattern we observed for 10 and 20 PPI. The reason can be explained by the fact that 5 PPI screen has larger cells and air was able to sufficiently flow over the second screen. As a result the second layer of 5 PPI screen actually contributed to the heat transfer rather than creating air blockage. R A PPI, 2 Screen 10 PPI, 1 Screen 5 PPI, 2 Screen 5 PPI, 1 Screen 20 PPI, 1 Screen 20 PPI, 2 Screen Plain Tube NTU Figure 7-19: Extended surface area ratio (RA) variation as a function of NTU.

168 Conclusion Stainless steel heat exchangers were fabricated by connecting stainless steel wire mesh screens to stainless steel tubes using wire-arc thermal spray coating. The optimum spraying distance of 152 mm was used to achieve a porosity of 2 %, oxide content of 6.6 %, and adhesion strength of 24 MPa for the deposited stainless steel. Results indicated superior penetration of the coating material into the gaps between wire mesh and tube s outer surface, which provided strong adhesion and thermal conduction in 10 and 5 PPI wires mesh. Fabricated heat exchangers were tested inside a hot air chamber and heat transfer performance were also analyzed. The extended surface area of the wire mesh enhanced the heat transfer from the hot air to the cooling water running inside the heat exchanger. All fabricated heat exchangers resulted in a higher temperature rise than the plain tube, with the maximum of 130 % for 2 screens, 10 PPI wire mesh, compared to the plain tube heat exchanger. It was found that the performance of wire mesh heat exchangers depended on the pore density of the mesh which effects the air penetration through the heat exchanger. If the pore density was high (20 PPI), then adding the second screen to the heat exchanger resulted in the reduction of the overall performance the heat exchanger. Based on the surface temperature analysis of the wire mesh, it was concluded that fins cannot be simply modeled as a plain longitudinal fin, since the heat was also being conducted from the transverse wires which were perpendicular to the longitudinal wire. An empirical model was developed to predict the temperature variation of the wire mesh screens. Another model was also developed for predicting the performance of the tested heat exchangers for various inlet flow rates and surface area.

169 152 Air-To-Air Wire Mesh Heat exchangers 8.1 Introduction Gas flares are used to eliminate waste gases such as methane, which are not feasible to use or transport. In theory, the heat of combustion can be recovered from the combustion gases using heat exchangers for different commercial or industrial processes. By positioning a heat exchanger (HEX) on top of the hot gas stack, as shown on Figure 8-1, the heat of combustion can be captured. Figure 8-1: Full assembly of a heat exchanger on top of the gas flare.

170 153 It is difficult to manufacture heat exchangers that can withstand high combustion temperatures and have a high enough efficiency to make them commercially viable. Waste gases exit flares at temperatures of over 1000ºC, which exceeds the operation range of most high thermal conductivity materials such as copper and aluminum that are typically used to fabricate heat exchangers. New, high efficiency heat exchanger design can compensate for the low thermal conductivity of materials, such as stainless steel and Inconel that can withstand high temperatures. In the previous chapters, I analyzed the efficiency and effectiveness of the thermally sprayed wire mesh porous heat exchangers on a small scale. In this study a large scale wire mesh heat exchanger was built and compared to a plain tube heat exchanger.

171 Heat Exchanger Design The proposed design is a modular heat exchanger which can be easily modified for different operating conditions. The heat exchanger consists of four sections which were fabricated separately and connected together using bolts, as seen in Figure 8-2. Wire mesh is placed only on the top and the bottom of the tubes on the first and the third section. To analyze the performance of the wire mesh heat exchanger (first and third section) to the plain tube heat exchanger, the second and forth sections were built without adding wire mesh screen on the tubes. The heat exchanger assembly was originally designed to fit onto a gas flare incinerator by simply aligning the support section of the design to the existing flange of the incinerator. Appendix B shows a schematic of the heat exchanger on an incinerator.

172 Figure 8-2: Assembly process for the heat exchanger. 155

173 Manufacturing of the Heat Exchanger The heat exchanger has four sections that were manufactured seperately, and bolted together. Two of the sections consists of only stainless steel tubes, and the other two has 2 sheets of 5 PPI wire mesh attached to the tubes. Each section of the heat exchanger was 2 in (50.8 mm) high, and there were a total of 5 tubes in each section. The tubes have an outer diameter of 1 in (25.4 mm), an inner diameter of 0.87 in (22.1 mm), and were spaced 3.5 in (88.9 mm) apart (Figure 8-3). The wire mesh used in the heat exchanger was of 5 PPI, and the overall size of the wire mesh was 17 in (431.8 mm) by 17 in (431.8 mm). The overall surface area of the tube, for each section, inside the 17 in x 17 in channel, was mm 2. The surface area for sections with wire mesh was furthur enhanced by mm 2 due to the presence of 2 sheets of 5 PPI wire mesh. Figure 8-3: Fabricated bare tube section of the main heat exchanger.

174 157 Figure 8-4: Fabricated section of the main heat exchanger, with one wire mesh screen attached on front and back side of the tubes. Before a thermal sprayed skin was applied, both the tube and the wire mesh were sand blasted before and after they were fastened together. The spray coating for the heat exchanger was done at the CACT lab as shown in Figure 8-5 where the twin wire-arc spraying gun is visible. Figure 8-5 shows the heat exchanger after the spraying process in which wires were fully connected to the tube s surface along the length of the tube. A dense layer of stainless steel coating was sprayed on the point of contact between the mesh and the tube using wire-arc thermal spraying gun, as shown Figure 8-6. The quality and the mechanical bonding of the connection between the wire mesh and the tube could have been furthered improved if the wire mesh was more flexible. A flexible wire mesh that could further bend around the tube would enhance the mechanical connection between them (Figure 8-7).

175 158 Figure 8-5: Wire mesh section after thermal skin deposition of stainless steel using wire-arc. Figure 8-6: Thermal sprayed surface of the wire mesh and the tube.

176 159 Figure 8-7: Mechanical bonding of 4 PPI wire mesh to the stainless steel tube [6]. The stacked heat exchanger consisted of two separate water distributing tubes that supply water to the heat exchanger units, as shown in Figure 8-8. Inlet and outlet manifolds were provided seperately for bare tube and wire mesh sections of the heat exchanger. Appendix C shows the stepby-step fabrication process of the heat exchanger.. Figure 8-8: Front view of the fabricated heat exchanger before welding the manifolds.

177 160 Figure 8-9: Back view of the fabricated heat exchanger after the final assembly. The wire mesh heat exchanger units (section 1 & 3) shared the same inlet and outlet manifold.

178 Experimental Apparatus The experimental apparatus consisted of a compressed air supply, an electrical heater, a heat exchanger unit and a wind tunnel. The hot air flowed inside the tubes of the heat exchanger unit, while cold air flowed over them and through the porous structures. A schematic representation of the experimental setup is shown in Figure An air electrical heater supplied hot air that passed through the heat exchanger. The compressor fed the air to the mass flow meter (Model FMAA844A, Omega Company, Stamford, CT) which was then supplied to the electrical heater (F076029, SKORPION AIR HEATERS, OSRAM SYLVANIA, Exeter, NH). The hot air from the electrical heater then entered the inlet of the heat exchanger unit, which was placed inside the wind tunnel. The wind tunnel provided constant air flow, with the use of direct drive centrifugal inline fan (DSI-135ANE, Twin City Fan and Blower, Minneapolis, MN). Cold air was blown through an 18 in (457.2 mm) diameter duct (Blo-R-Vac flexible duct, McMASTER-CARR) and over heat exchanger. To reduce the heat loss from the heat exchanger, it was insulated with two layers of fiber glass insulation (Micro-Flex, John Manville Corporation, Denver, CO). The heat exchanger was painted with a black paint to provide a uniform emissivity thought the heat exchanger. Four K-type thermocouple probes measured the hot air temperature at the inlet (T i ) and outlet (T o ) of the heat exchanger, as shown in Figure Twelve thermocouples were attached to the surface of the heat exchanger to record the local surface temperature of the wire mesh and the tube surface temperature. Eight thermocouples were used to measure the average cold air temperature downstream of the heat exchanger, and two to read the cold air temperature upstream of the heat exchanger. A National Instruments Data Acquisition (DAQ) unit was used to record the thermocouple signals. The DAQ was connected directly to a computer which was equipped with Lab VIEW Signal Express v.3.0 (National Instruments Corporation, Austin, TX). The velocity

179 162 flow field inside the wind tunnel was measured using a hot wire anemometer (Model HHF42, Omega Company, Stamford, CT) with a range of 0 to 20 m/s and a resolution of 0.1 m/s. Appendix D shows the location of the thermocouples on the surface of the heat exchanger. Appendix E shows a schematic of the fan, the fan performance specification and the electrical heater. Figure 8-10: Schematic representation of the experimental setup.

180 Heat Transfer Calculation The objective of this experiment was to investigate the effect of wire mesh in increasing heat transfer compared to a bare tube heat exchanger. The comparison was made by measuring the temperature drop of hot air along the heat exchanger. Three hot air flow rates of 250, 350 and 450 L/min and three different cold air velocities of 3.7, 4.7 and 5.4 m/s, as shown in Table 8-1, were used. Experiments were done in which heat exchangers were placed inside a wind tunnel with variable speed fan. The rate of heat extracted from the hot air can be calculated from Q a = m a C p.a T a (8-1) where C p is specific heat, m is mass flow rate and the subscripts a stands for hot air flow loop. Table 8-1: Cold air velocities inside the wind tunnel. Cold Flow Average Velocity (m/s) Volume flow rate (cfm)

181 Temperature Drop, ( C) Results and Discussion The results for a constant cold air flow rate of 5.4 m/s in the wind tunnel, and for a hot air flow rate of 250 to 450 L/min are presented in Figure The results illustrate the temperature drop of hot air through the heat exchanger for various flow rates. The results indicate higher temperature drop when using wire mesh, compared to bare tube heat exchangers. In all of these cases, temperature drop was reduced when the flow rate was reduced Mesh Tube Hot Gas Flow Rate, (L/min) Figure 8-11: Temperature drop for different hot air flow rates at a constant cold air velocity of 5.4 m/s.

182 Percentage Heat Transfer Increase, (%) 165 Using Equation (8-1), the heat transfer from the hot was calculated. The increase in heat transfer by using wire mesh was then compared with a bare tube heat exchanger, as shown in Figure m/s 4.7 m/s 3.7 m/s Hot Gas Flow Rate, (L/min) Figure 8-12: Heat transfer enchantment of the wire mesh sections compare to the plain tube. The improvement in the heat transfer rate was in the range 5 to 15%. The maximum increase in heat transfer was experienced for a hot air flow rate of 450 L/min. It could also be observed that the improvement in the heat transfer was superior for the cold air with the low air velocity of 3.7 m/s than 5.4 m/s.

183 Heat Transfer Characterization The overall heat transfer coefficient U and log mean temperature T LMTD were calculated using equations below Q = UA t T LMTD (8-2) T LMTD = T 1 T 2 ln ( T 1 T 2 ) (8-3) where A t is the tube outer surface area and T 1 and T 2 represent the temperature difference between two fluids at the two ends (inlet and outlet) of a heat exchanger. The heat transfer coefficient of air h out is found using the equation: 1 UA t = 1 h in A 1 + ln ( D out D in ) 2 k t L t + 1 h out A out (8-4) where D in and A in are the inner tube diameter and area while D out and A out are the outer tube diameter and area, L t is the length of the tube and k t is the thermal conductivity of the stainless steel tube. Reynold and Nusselt number were calculated using the equation, Re D,out = V outd out ν (8-5) Nu D,out = h outd out k (8-6) where V out is the velocity of air, ν the kinematic viscosity, h out is the heat transfer coefficient of the cold air, and k the thermal conductivity of air.

184 167 Experimental results were compared to Nusselt number (Nu H ) variation with the change of Reynolds number (Re H ) relations for stack wire mesh screens in the literature as shown in Figure The calculated heat transfer coefficients were generally lower than valued reported by Li et al. [42] and greater than values reported by Venugopal, Balaji and Venkateshan [43] Venugopal, Balaji and Venkateshan (0.92 Porosity) [43] Venugopal, Balaji and Venkateshan (0.89 Porosity) [43] Nusselt Number, Nu H Venugopal, Balaji and Venkateshan (0.85 Porosity) [43] Li et al. (copper wire screens) [42] 5 PPI, 2 screen Reynolds Number, Re H Figure 8-13: Nusselt number (NuH) variation as a function of Reynolds number (ReH).

185 Conclusion Thermally sprayed air-to-air heat exchanger suitable for high temperature applications was fabricated. Fabricated heat exchanger was tested inside a wind tunnel and its heat transfer performance was analyzed. Wire mesh heat exchanger resulted in a higher temperature rise than the plain tube. The improvement in the heat transfer rate was in the range 5 to 15%. The maximum increase in heat transfer was experienced for a hot air flow rate of 450 L/min. It could also be observed that the improvement in the heat transfer was superior for the cold air with the low air velocity of 3.7 m/s than 5.4 m/s. The heat transfer enhancement due to addition of a wire mesh was measured experimentally. Experimental results were compared to Nusselt number (Nu H ) variation with the change of Reynolds number (Re H ) relations for stack wire mesh screens in the literature.

186 169 Summary 9.1 Laser Sintered Heat Exchangers Laser sintering process was an effective manufacturing methods for fabricating heat exchangers and the following objectives were achieved: Fabricated channels containing cubic and round-strut tetradecahedral cells with identical strut diameters and one with thin-strut tetradecahedral cells using DMLS technology. Enhanced the bonding and the contact area between the porous structure and the surface of the heat exchanger while controlling the uniformity of the porous structure. Experimentally investigated the impact of internal cell geometry on pressure drop, conduction and forced convection heat transfer through DMLS heat exchangers. Designed a thin-strut tetradecahedral geometry, which maximized the heat transfer while minimizing the weight and friction loss. 9.2 Wire Mesh Heat Exchangers In the second part of the thesis wire mesh heat exchangers were fabricated using thermal spraying process to bond wire mesh screens or perforated metal sheets to the outer surface of the tubes. The following objectives were achieved in the second part of the thesis: Developed a simple method of increasing the heat transfer surface area by using a twin wire-arc thermal spray system to generate a dense, high strength coating that bonds porous structures to the body of the heat exchanger.

187 170 Concluded that a porous structures with high open area allowed for superior penetration of the coating material into the gap between the wire mesh and the tube surface, and thus providing good adhesion and thermal conduction. Investigated the heat transfer enhancement for different pore density wire mesh and perforated sheets sizes. The experimental results indicated that a right balance between pore density and number of screens is crucial for maximizing the heat transfer performance of the porous heat exchangers. Enhanced the heat transfer performance of plain tube heat exchangers using various pore density wire mesh and perforated sheets. It was found that the performance of the heat exchangers depended on the air penetration between the porous structures. An empirical model was developed to predict the temperature variation of the wire mesh screens. Fabricated an industrial size wire mesh heat exchanger and compare its performance to a conventional plain tube heat exchanger. To summarize, the values of heat transfer enhancement achieved by using wire mesh as presented in chapters 6, 7 and 8, Nu D and Re D based on tube diameter are shown in Figure 9-1. By fitting a trend line of best fit to these data points a general empirical relation between Nu D and Re D is found as follow Nu D = Re D (9-1)

188 PPI Mesh (Ch.6) 0.18" Perforared (Ch.6) 5 PPI, 1 Screen (Ch.7) 5 PPI, 2 Screen (Ch.7) 10 PPI, 1 Screen (Ch.7) 10 PPI, 2 Screen (Ch.7) 5 PPI, 2 Screen (Ch.8) Log Nu D Log Nu D Figure 9-1: Nusselt number (NuD) variation as a function of Reynolds number (ReH).

189 172 References 1. L. Tianjian, Ultralight Porous Metals: from Fundamentals to Applications, Acta Mechanica Sinica, vol. 18(5), pp , J. P. Bonnet, F. Topin and L. Tardist, Flow Laws in Metal Foams: Compressibility and Pore Size Effects, Transport in Porous Media, vol. 73(2), pp , A. Bhattacharya and R. L. Mahajan, Metal Foam and Finned Metal Foam Heat Sinks for Electronics Cooling in Buoyancy-Induced Convection, ASME Journal of Electronic Packaging, vol. 32(128), pp , J. Banhart, Manufacture, Characterization and Application of Cellular Metals and Metal Foams, Progress in Material Science, vol. 46, pp , K. Boomsma, D. Poulikakos, and F. Zwick, Metal Foams as Compact High Performance Heat Exchangers, Mechanics of Materials, vol. 35, pp , R. Rezaey et al., Fabrication of Wire Mesh Heat Exchangers for Waste Heat Recovery Using Wire-Arc Spraying, Journal of Thermal Spray Technology, vol. 23, pp , S. Salavati et al., Development of High Density Twin Wire Arc Sprayed Coatings On Metallic Foam Substrates, Int. Thermal Spray Conference & Exposition, Busan, May 13-15, T. J. Lu, L. Valdevit, and A.G. Evans, Active Cooling by Metallic Sandwich Structures with Periodic Cores, Progress in Materials Science, vol. 50, pp , D. J., Sypeck and H. N. G. Wadley, Cellular Metal Truss Core Sandwich Structures, Special Issue, Advanced Engineering Materials, vol. 4(10), pp. 759, D. J. Sypeck, Wrought Aluminum Truss Core Sandwich Structures, Metall. Mat. Trans.

190 173 B, vol. 36B(1), pp , H. R. Salimi Jazi, J. Mostaghimi, S. Chandra, L. Pershin, and T. Coyle, Spray-Formed, Metal-Foam Heat Exchangers for High Temperature Applications, J. of Thermal Science and Engineering Applications, vol. 1(3), pp , N. Tsolas, S. Chandra, H. R. Salimi Jazi, J. Mostaghimi and L. Pershin, Thermal spray forming of high-efficiency, metal-foam heat exchanger tubes, Int. Thermal Spray Conference & Exposition, Singapore, May 3-5, P. Khayargoli, V. Loya, L. P. Lefebvre and M. Medraj, The Impact of Microstructure on the Permeability of Metal Foams, Proceedings of Canadian Society for Mechanical Engineering, London, pp , J. Tian et al., The Effects of Topology Upon Fluid-Flow and Heat-Transfer Within Cellular Copper Structures, International Journal of Heat and Mass Transfer, vol. 47 (14-16), pp , J. Assaad, A. Corbeil, P. F. Richard and B. Jodoin, Novel Stacked Wire Mesh Compact Heat Exchangers Produced Using Cold Spray, Journal of Thermal Spray Technology, vol. 20(6), pp , C. T. Joen et al., Thermo-Hydraulic Study of a Single Row Heat Exchanger Consisting of Metal Foam Covered Round Tubes, International Journal of Heat and Mass Transfer, vol. 53(15-16), pp , A. G. Leach, The thermal conductivity of foams. I. Models for heat conduction, Journal of Physics D: Applied Physics, vol. 26(5), pp. 733, R. C. Progelhof, J. L. Throne and R. R. Ruetsch, Methods for predicting the thermal conductivity of composite systems: A review, Polymer Engineering & Science, vol. 16(9), pp , 1976.

191 C. Y. Zhao, T. J. Lu, H. P. Hodson and J. D. Jackson, The Temperature Dependence of Effective Thermal Conductivity of Open-Celled Steel Alloy Foams, International Journal of Heat and Mass Transfer, vol. 367 (1-2), pp , J. W. Paek, B. H. Kang, S. Y. Kim and J. M. Hyun, Effective Thermal Conductivity and Permeability of Aluminum Foam Materials, International Journal of Thermophysics, vol. 21(2), pp , S. W. Thomson, On the division of space with minimum partitional area, Journal of Acta Mathematica, vol. 11(1-4), pp , E. Achenbach, The effect of surface roughness on the heat transfer from a circular cylinder to cross flow of air, Int. J. of Heat and Mass Transfer, vol. 20(4), pp , K. C. Leong and L.W. Jin, Characteristics of oscillating flow through a channel filled with open-cell metal foam, Int. J. of Heat and Fluid Flow, vol. 27(1), pp , T. S. Zhao and P. Cheng, Oscillatory pressure drops through a woven-screen packed column subjected to a cyclic flow, Cryogenics, vol. 36(5), pp , Y. L. Ju, Y. Jian and Y. Zhou, Experimental study of the oscillating flow characteristics for a regenerator in a pulse tube cryocooler, Cryogenics, vol. 38(6), pp , K. Vafai, Handbook of Porous Media, Taylor and Francis Group/CRC Press, Boca Raton, FL., Y. A. Cengel, Heat and Mass Transfer: A Practical Approach, McGraw-Hill, New York, NY, P. H. Egli, Thermoelectricity: Including the Proceedings of the Conference on Thermoelectricity Sept Wiley, New York, A. Reuss, Berechnung der Flie ssgrenze von Mischkristallen auf Grund der Plastizittsbedingung freinkristalle, ZAMM Journal of Applied Mathematics and

192 175 Mechanics / Zeitschrift fr Angewandte Mathematik und Mechanik, vol. 9(1), pp , N. Bianco, R. Capuano, W. K. Chiu, S. Cunsolo, V. Naso and M. Oliviero, Numerical Analysis of Conjugate Heat Transfer in Foams, Proceedings of COMSOL Conference, Milan, October 10-12, M. Iasiello, S. Cunsolo, M. Oliviero, W. M. Harris and N. Bianco, Numerical Analysis of Heat Transfer and Pressure Drop in Metal Foams for Different Morphological Models, Journal of Heat Transfer, vol. 136(11), pp , O. H. Wiener, Die Theorie des Mischkörpers für das Feld der stationären Strömung. 1. Abhandlung: Die Mittelwertsätze für Kraft, Polarisation und Energie, Leipzig : B. G. Teubner, W. Voigt, Lehrbuch der Kristallphysik, Leipzig, Berlin, B.G. Teubner, K. Boomsma and D. Poulikakos, On the effective thermal conductivity of a threedimensionally structured fluid-saturated metal foam, International Journal of Heat and Mass Transfer, vol. 44(4), pp , D. A. G. Bruggeman, Berechnung verschiedener physikalischer Konstanten von heterogenen Substanzen. I. Dielektrizittskonstanten und Leitfähigkeiten der Mischkörper aus isotropen Substanzen, Annalen der Physik, vol. 416(7), pp , J. C. Maxwell, A treatise on electricity and magnetism, Oxford: Clarendon Press, EOS GmbH, (2007). Material data sheet: Eos stainless steel Available: EOS, Additive Manufacturing, Laser-Sintering and industrial 3D printing - Benefits and Functional Principle, Available:

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194 177 Appendices Appendix A: Matlab Code for the Empirical Fin Model. clc clear nx=101; x=zeros(nx,1); s=zeros(nx,1); T=zeros(nx,1); Texp2=[ ]; Xexp2=[0

195 ]; Xexp2=Xexp2+1.5; l= ; d=0.0012; np=5; length=np*l; Tinf= ; Tb= ; h=30; h1=h/10; p=3.1415*d; A=3.1415*d^2/4; k=16; m=(h*p/k/a)^0.5; m1=(h1*p/k/a)^0.5; % pore length m % wire thickness "m" % number of pores % total length % Temp inf in Kelvin % Base Temp in Kelvin % Heat Transfer Coef % wire perimeter in m2 % wire cross section area m2 % thermal condcutivity of wire

196 179 for i=1:nx x(i)=length*(i-1)/(nx-1); for j=1:(np+1) if (x(i)>=(j-1)*l) && (x(i)<(j*l-d)) s(i)=x(i)+(j-1)*l; elseif (x(i)>=(j*l-d)) && (x(i)<(j*l)) s(i)=2*j*l+(l+d)/d*(x(i)-j*l); end end if x(i)<(l-d) T(i)=Tinf+(Tb-Tinf)*exp(-m1*s(i)); else T(i)=Tinf+(Tb-Tinf)*exp(-m*s(i)); end Torig(i)=Tinf+(Tb-Tinf)*exp(-m*x(i)); end C=T-273; Corig=Torig-273; xmm=1000*x; close all; figure plot(xmm,c) xlabel('length, x in mm') ylabel('temperature in C') hold on; plot(xexp2,texp2,'-o','color','black'); hold on; plot(xmm,corig,'-*','color','green'); legend('show');

197 Appendix B: Heat Exchanger Assembly for the Hot Gas Incinerator. 180

198 181

199 182

200 Appendix C: Step-by-Step Fabrication Process of the Heat Exchanger. 183

201 184

202 185

203 186

204 187

205 188

206 189 Appendix D: Location of the Thermocouples on the Surface of the Heat Exchanger.

207 190

208 191 Appendix E: Shows a Schematic of the Fan, the Fan Performance and the Electrical Heater.

209 192