Cfd As a Design Tool for a Concentric Heat Exchanger
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HEFAT2012 9th International Conference on Heat Transfer, Fluid Mechanics and Thermodynamics 16 – 18 July 2012 Malta CFD AS A DESIGN TOOL FOR A CONCENTRIC HEAT EXCHANGER J.P. Oosterhuis1,*, S. Bühler1, D. Wilcox2, T.H. van der Meer1 *Author for correspondence 1 Department of Thermal Engineering, University of Twente, P.O. Box 217, 7500AE, Enschede, The Netherlands, E-mail: [email protected] 2 Bosch Thermotechnology, P.O. Box 3, 7400AA, Deventer, The Netherlands A common approach to design a heat exchanger is the use ABSTRACT of a one-dimensional model such as the P-NTU, ε-NTU or A concentric gas-to-gas heat exchanger is designed for LMTD method [3][4]. However, these methods do not take into application as a recuperator in the domestic boiler industry. The account additional effects such as axial conduction and non- recuperator recovers heat from the exhaust gases of a uniform heat transfer coefficients. Although extension of these combustion process to preheat the ingoing gaseous fuel mixture 1D models to include such effects is generally possible, resulting in increased fuel efficiency. This applied study shows detailed temperature and velocity profiles in flow cross section the use of computational fluid dynamics (CFD) simulations as cannot be obtained. In cases where this information is required an efficient design tool for heat exchanger design. An to make proper design decisions, computational fluid dynamics experimental setup is developed and the simulation results are (CFD) comes into focus as a relevant design tool. validated. In the recuperator design presented in this study, good flow temperature uniformity in the radial direction is of high INTRODUCTION importance in order to prevent undesired auto-ignition of the To increase the fuel efficiency of gas-fired domestic boiler fuel mixture. Therefore a CFD model is proposed which can be appliances, the principle of recuperation is investigated. By used as a realistic design tool capturing detailed information recovering heat from the exhaust gases of the combustion about heat exchange and flow temperature uniformity. process, the ingoing fuel mixture is preheated. This preheating results in an increased adiabatic flame temperature and hence in an increase in fuel efficiency [1]. The heat transfer from the exhaust gases to the ingoing fuel mixture is usually achieved by the application of a compact gas-to-gas heat exchanger (recuperator). Implementing recuperation in practical applications was for long a niche because of high costs in relation to the gain in efficiency which made them not economically feasible [2]. However, rising energy prices and emissions legislation have changed this fact, resulting in numerous recuperator research activities being carried out in recent years. In this study a recuperator design is proposed for a specific type of gas-fired domestic boiler appliance. Based on the Figure 1 Core profile of designed concentric heat exchanger, results of preliminary design iterations, in combination with = 200 mm, = Ø 88 mm. manufacturability and the current system lay-out, it was 퐿푐표푟푒 퐷표푢푡 decided to focus on a counter-flow concentric tube heat NOMENCLATURE exchanger design. In this heat exchanger type the exhaust gases [°] Angular location around core [m2] Flow cross-sectional area flow through a tube in the center, while the fuel mixture 훼 [kJ/kg∙K] Specific heat capacity at constant pressure counter-flows through the surrounding ring shown in the 퐴푐 푝 [m] Hydraulic diameter, = 퐶 4퐴푐 geometry in Figure 1. In this way the fuel mixture acts as an [m] Core outside diameter 퐷ℎ 퐷ℎ 푃 insulator to thermally insulate the exhaust gases from the [-] Relative solution error of numerical simulation 표푢푡 interior of the appliance. 퐷 [W/m∙K] Thermal conductivity 휀푟푒푙 푘 110 [m] Core flow length the headers equals those at the corresponding inlet and outlet of [kg/s] Mass flow rate 푐표푟푒 the recuperator core. Uniform outflow from the headers to the 퐿 [kg/m∙s] Dynamic viscosity 푚̇ [-] Number of nodes or number of flow channels core is assumed in the recuperator core simulations and the 휇 [Pa] Relative static pressure influence of heat transfer on the flow field is neglected in the 푁 [Pa] Absolute reference pressure 푟푒푙 header simulations. The outer shell is assumed to be well 푝̅ [m] Perimeter insulated and hence heat loss through the outer core walls is 푝푟푒푓 [-] Reynolds number, = neglected. All CFD simulations are performed using 푃 푣�퐷ℎ휌 [-] Effective mesh refinement level ANSYS CFX v13.0 with a high resolution advection scheme 푅푒 3 푅푒 휇 [kg/m ] Density eff and a convergence criterion for momentum, mass and energy 푟 [m/s] Absolute global flow non-uniformity 1 10 휌 [-] Relative local flow non-uniformity in flow channel transport of . For a detailed description of the −6 푆 [K] Temperature governing Navier-Stokes equations solved in ANSYS CFX, the 푖 푆 [K] Temperature difference 푖 reader is referred푅푀푆 to≤ the⋅ documentation [5]. 푇 [var] Target variable in numerical simulation Δ푇 [s] Residence time Core Model Configuration 휃 [m/s] Flow velocity in flow channel 휏 [m/s] Average flow velocity over flow cross-section The core model has been subdivided into two fluid domains 푖 푣 [-] Scaled dimension over flow channel푖 width and one solid domain. Rotational periodicity has been applied 푣 ̅ [m] Axial (flow) direction and only 1/3 of the actual geometry is modeled. On the inlets of 휉 the fluid domains, a fixed mass flow rate is specified resulting Subscripts푧 1 Exhaust gas domain in a uniform flow inlet velocity boundary condition. Moreover 2 Gas/air-mixture domain the inlet temperature is specified. On the outlets of the fluid Location at inlet domains an average relative static pressure boundary condition Location at outlet is specified of = 0 Pa. Reference pressure is set to 푖푖I Indicates refined mesh 표표표II Indicates coarser mesh = 1 10 Pa. All other boundaries of the fluid domains are 푝̅푟푒푙 specified by the5 interface with the solid domain inquiring a no 푝푟푒푓 ⋅ DESIGN PARAMETERS slip wall condition and a continuous heat flux through the wall. Based on the boiler appliance under investigation, a number As the expected maximum Reynolds numbers are in the range of design parameters are defined. The exhaust gas enters the of = 300 ~ 1500, the fluid flow is modeled laminar. The heat exchanger at , = 1032 K with a density of boundary conditions of the solid domain are fully defined by 푅푒 , = 0.330 kg/m and a specific heat capacity the interface with the corresponding fluid domains and an 푇1 푖푛 adiabatic boundary condition on the outer walls. A schematic , = 1.283 kJ/kg3 K. The gas/air-mixture enters the heat 휌1 푖푛 representation of the model configuration is shown in Figure 2. exchanger푝 1 푖푛 in counter-flow configuration shown in Figure 2, at 퐶 ⋅ , = 293 K with a density of , = 1.136 kg/m and a specific heat capacity , = 1.055 kJ/kg K. The mass3 flow 푇2 푖푛 휌2 푖푛 rate of both the exhaust 푝gas2 푖푛 and gas/air-mixture stream equals = 0.001336 kg/s. The퐶 desired temperature⋅ increase for the gas/air-mixture = 288 ~ 400 K aims at producing the maximum푚̇ increase in efficiency. Higher values are limited due 2 to possible auto-ignitionΔ푇 of the gas/air-mixture, which has to be prevented in order to obtain safe operation. The core flow length is fixed by the system configuration to = 200 mm and the core material is chosen to be aluminum having a Figure 2 Outline of recuperator core model. 푐표푟푒 thermal conductivity of = 237 W/m K. Due퐿 to the system interfaces for the exhaust gas the headers require a 90° bend to Header Model Configuration allow for a straight recuperator푘 core. ⋅ Due to the relative low pressure drop in the core, the inlet and outlet header geometry influence each other’s performance SIMULATION METHOD and must be taken into account together. As the effect of the To benefit from the (rotational) periodic geometry of the flow field on the temperature distribution is already known heat exchanger core compared to the heat exchanger headers, from the core simulations, only cold flow simulations are used the simulation and design of core and headers has been for the header design. This allows decoupling of the two separated. In first instance the core is designed to meet the different fluid domains; moreover the solid domain has become required temperature specifications. Secondly the headers are redundant. Hence a simulation model consisting of header – designed based on maintaining flow uniformity to reduce the core – header is constructed. risk on hot spots and possible auto-ignition of the gas/air- Regarding the boundary conditions, the same inlet and mixture. outlet boundary conditions are used as described in the core Four major assumptions are required in order to provide a model configuration: uniform mass flow rate at the inlet and a fully defined simulation model. No heat transfer in the headers pressure boundary condition at the outlet. A constant fluid is assumed, i.e. the flow temperatures at the inlet and outlet of temperature has been specified as the average temperature of 111 the concerning fluid predicted by the core simulation. Along hence the flow is considered thermally fully developed. This the remaining boundaries of the fluid domain, a no-slip wall quadratic velocity and temperature profile is already well condition is defined. defined with only 6 nodes and no significant deviations are Due to larger hydraulic diameters in the headers, turbulent observed. This result highly reduces computational costs. flow can be expected. In the gas/air-mixture domain the small Applying this mesh resolution to the total core mesh results in a flow channels and corresponding influence of the walls it was mesh with 3.90 10 nodes.