Numerical study of heat pipe application in heat recovery systems *Song Lin, John Broadbent, Ryan McGlen Thermacore Europe, 12 Wansbeck Business Park Ashington, Northumberland NE3 QW, UK E-mail: song.lin@thermacore.com Abstract Heat pipes, as two-phase heat transfer devices with extremely high effective thermal conductivity. They can be cylindrical or planar in structure. Heat pipes can be embedded in a metal cooling plate, which is attached to the heat source, and can also be assembled with fin stack for fluid heat transfer. Due to the high heat transport capacity, heat exchangers with heat pipes have become much smaller than traditional heat exchangers in handling high heat fluxes. With the working fluid in a heat pipe, heat can be absorbed on the evaporator region and transported to the condenser region where the vapour condenses releasing the heat to the cooling media. Heat pipe technology has found its increasing applications in enhancing the thermal performance of heat exchangers in microelectronics, energy and other industrial sectors. Utilisation of heat pipe fin stack in the drying cycle of domestic appliances for heat recovery may lead to a significant energy saving in domestic sector. However, the design of the heat pipe heat exchanger will meet number of challenges. This paper presents a design method by using CFD simulation of the dehumidification process with heat pipe heat exchangers. The strategies of simulating the process with heat pipes are presented. The calculated results show that the method can be further used to optimise the design of the heat pipe fin stack. The study suggests that the CFD modelling is able to predict thermal performance of the dehumidification solution with heat pipe heat exchangers. Key words Heat pipe technology, heat pipe fin assemblies, CFD simulation, high thermal performance, heat recovery, dehumidification and drying process, dishwasher dying system 1. Introduction The increasing demand for energy efficiency in domestic appliances (such as dishwasher, air conditioner, durable drier or fridge/freezer) and industrial devices is the main drive for continuously introducing and/or improving heat recovery systems in these appliances and devices. Heat transfer efficiency in such systems is the primary factor for efficient performance of the whole systems [1]. Heat pipes, as a high-efficient heat transfer element, are widely used in electronics cooling industry and energy efficiency sector. They can be embedded with Aluminium heat sinks to enhance cooling efficiency and/or compactness of cooling devices [2]. Heat pipes are also widely used in energy recovery systems in domestic and industrial applications, such as in some domestic appliances for improving the efficiency of the drying cycles. However, the design of the heat 1
recovery systems with heat pipe units is the key to provide a heat exchanger system to work as efficient as expected. Without correct design of such systems, heat pipes are not able to transport enough heat or function as an extremely poor thermal conductor in the systems. Computational Fluid Dynamics is a powerful tool for fluid dynamics and thermal design in industrial applications, as well as in academic research activities. Based on the current capabilities of main CFD packages suitable for industries (such as FLOTHERM, ICEPAK, FLUENT and CFX) and the nature of industrial applications, understanding the physics of the processes, introducing adequate simplifications and establishing an appropriate model are essential factors for obtaining reasonable results and correct thermal design [3] [] [5]. In detail, this paper aims to provide a thermal model for simulating the performance of a heat pipe system for recovering waste heat in the drying cycle in a domestic appliance. In this model, the details of two-phase flow and heat transfer are not simulated, but the overall performance of condensing and evaporating can be predicted. This is a typical case of CFD simulations for industrial applications where time-consuming and uncertain simulations of detailed two-phase flow and heat transfer should be avoided. The basic principle of the drying cycle system is that in a conduit, warm and humid air with 100% relative humidity is sucked in at the inlet, passes through the fin stack of the heat pipe unit, and further goes through the additional condenser. In the condensation section of the conduit, condensate is obtained on the fin surfaces and additional condenser surfaces and drained from the conduit. The humid air continuously flows through the other stack of the heat pipe unit and is warmed up by the heat transported from the upstream warmer humid air through the heat pipe unit. An additional heater is used to further heat up the air to certain temperature and humidity before the air flow returns into the drying process. The concept of the system is illustrated in Figure 1. The modelling strategy and results are discussed in the following sections. A similar system of the same concept used in a dishwasher drying system is shown in Figure 2. Heat pipe condenser fin stack Heat pipe(s) Additional condenser Dry air out Condensate outlet Additional heater Conduit Heat pipe evaporator fin stack Humid air in Figure 1 Heat recovery system using heat pipe solution in drying cycle 2
Blower Heat pipes working against gravity M K Q Washing chamber with humid air Gravity Condensate outlet Heater Figure 2 Concept of heat pipe solution of heat recovery in dishwasher drying cycle 2. Modelling strategy In the drying system with a heat pipe solution concept, a part of the heat from the condensation section of the conduit can be transported through the heat pipe unit into the downstream air flow in the heating section of the conduit. Thus a part of heat released from the condensate can be used to heat up the down-stream air. A typical operation condition is that the humid air goes into the conduit at an inlet temperature 5 C, relative humidity 100%, and volume flow rate litre/s. The drying cycle can be simplified as dehumidification process in which a moist air stream is cooled at constant mixture pressure to temperature below its dew point temperature, a certain amount of condensate of the water vapour would present. The dehumidification process is schematically shown in Figure 3. Cooling coil Heating coil T 1 φ 1 ω 1 Condensing T 2 φ 2 ω 2 T 3 φ 3 ω 3 2 φ=100% 1 3 ω Heating φ3 Condensate, m w T T: Temperature φ: Relative humidity ω: Mass ratio of vapour to air a) Schematic process b) Psychrometric chart indication Figure 3 Dehumidification process 3
In this work, a CFD package, FLOTHERM, has been used to simulate the drying cycle. Since FLOTHERM can only deal with single-phase flows, a strategy to simulate the two-phase process by using FLOTHERM in conjunction with EXCEL calculations of the equilibrium dehumidification is presented as follows: 1) The conduit is divided into two sections, condensing and heating sections. 2) In the condensing section, the mass flow rate of the condensate is calculated by applying conservation of mass separately for the dry air and water in the condensing section and using the governing equations. 3) Based on the temperature and relative humidity at the inlet and outlet of the condensing section, the average density, thermal conductivity, viscosity and specific heat of the mixture flow, and the condensate mass rate at the outlet of the section are calculated by using the mass balance and equilibrium equations. ) The average properties are used in the single-phase flow in the condensing section in the FLOTHERM simulation. 5) In the heating section, the temperature and relative humidity at the heating inlet will be the same as those at the condensing outlet. However the flow rate for the heating section is the total flow rate for the condensing section subtracted by the condensate flow rate. With the initial temperature and relative humidity at the inlet and outlet of the heating section, the mass conservation equations are used to calculate the average properties of the mixture flow. ) The average properties are also used for the single-phase fluid in the heating section in the FLOTHERM model. The steps to complete the simulation for a drying cycle are: 1) Initial guess for outlet temperatures of condensing and heating sections 2) Calculation of fluid properties and parameters for FLOTHERM simulation 3) Input the properties and parameters in FLOTHERM model for simulation ) FLOTHERM-calculated results compared with the initial guess values. If the errors are less than 1%, the final results are obtained, otherwise replace the initial temperature value with new values and carry out Step 2) until the final results are obtained. In FLOTHERM, the typical parameters used to construct the model, are: 1) Calculation domain defined 2) Condensing section and heating section defined in separate conduits 3) Two flow regions defined for condensing section and heating section, respectively ) Forced convection and turbulent flow in the condensing and heating conduits 5) Heat sink smart parts representing fin stacks on heat pipes, auxiliary condenser and heater, and Aluminium patched to all the heat sink parts ) Cuboids representing heat pipes with the thermal conductivity, density and specific heat obtained from Thermacore proprietary database 7) Collapsed cuboids representing interface between fin stack and heat pipes ) The global ambient temperature defined as 25 C in this case 3. Results and discussions The simulation results shown in this paper only reflect the overall performance of the system at different operating conditions. The details of the thermal characteristics and design parameters of the system can not be publicized in this paper, due to the Non-Disclosure Agreement restrictions. The operating parameters used in this paper are: the inlet temperature of the humid air 35 ~ 50 C, inlet flow rate ~ litre/s, inlet relative humidity 100%, and auxiliary heating power 100W.
The overall condensate rate, ratio of heat transported through heat pipe unit and auxiliary condenser, relative humidity at outlet, and overall pressure drop through the conduit, are plotted against the operating conditions, as shown in Figure. It can be seen that the performance of the heat pipe unit is influenced by the operating conditions. With increasing inlet flow rate and temperature of the humid air, the condensate rate can be increased co-ordinately, due to the enhancement of heat transfer rate in the system with higher flow velocity and temperature. However, the ratio of heat transported through the heat pipe unit and auxiliary condenser decreases with increasing inlet flow rate. This could indicate that the auxiliary heat exchanger can work more efficiently with higher flow rate and there is demand to improve the design of heat pipe unit to achieve higher performance. The relative humidity at the conduit outlet can be determined by the inlet flow conditions and auxiliary heating power. With the same heating power as simulated in this case, the higher flow rate of the saturated humid air will lead to a higher relative humidity at the outlet, although the heat transfer coefficient at the fin stack heat exchangers can increase to certain extent. The simulation results have shown the correct tendency in the system. Similarly, with higher inlet temperature of the saturated humid air, more vapour mass is contained in the air flow, thus higher relative humidity of the air flow at the outlet can be expected with the same auxiliary heating input. The pressure drop is also an indication of energy consumption in the system, since the air flow is provided through a fan or blower. Figure d) show that the overall pressure drop is almost independent of the inlet temperature but is a function of the flow rate. A simulation without using heat pipe unit has also been simulated at the same operating conditions. The comparison shows that the system performance can be improved by 20 ~ 30%, in terms of condensate rate, but the pressure drop through the system can increase by 25 ~ 35 %. 3.5E-0 0.5 Total mass flow rate of condensate, kg/s 3.0E-0 2.5E-0 2.0E-0 1.5E-0 1.0E-0 5.0E-05 Inlet flow rate, liter/s Q (heat pipe unit) /Q (auxiliary condenser) 0.0 0.35 0.30 0.25 0.20 0.15 0.10 0.05 Inlet flow rate, litre/s 0.0E+00 30 35 0 5 50 55 0.00 30 35 0 5 50 55 a) Overall condensate rate b) Ratio of heat transported through heat pipe unit and auxiliary condenser 5
0.0 50.0 0.0 00.0 300.0 Inlet flow rate, litre/s RH out, % 30.0 20.0 DP overall, Pa 200.0 10.0 Inlet flow rate, litre/s 0.0 30 35 0 5 50 55 c) Relative humidity at outlet d) Overall pressure drop through the conduit 100.0 Figure Results obtained from FLOTHERM simulations 0.0 30 35 0 5 50 55. Conclusions Based on the simulation results, it can be concluded that: 1) The simulation methodology can provide reasonable results to predict the system performance. The method is relatively simple and very suitable for industrial application. Soon the results will be compared with the test results, and modification can be carried out, with reference to the test results. 2) The predicated results show that using heat pipe solution can lead to significant improvement of dehumidification process. 3) The performance of the drying cycle system is significantly influence by the operating conditions at the inlet. In general, higher inlet temperature and flow rate of the saturated humid air can lead to higher heat transfer performance in the system, but more additional heating is needed to achieve similar relative humidity at the outlet. ) The predicted results also indicate that the heat pipe unit can be further optimised to achieve similar or better heat transfer performance than the auxiliary condenser. However the auxiliary condenser is necessary to maintain the temperature difference between the two ends of the heat pipe unit. References: 1. Sauciuc, I.; Akbarzadeh, A.; Johnson, P.; Characteristics of two-phase closed thermosyphons for medium temperature heat recovery applications; Heat Recovery Systems and CHP, Vol.15, Issue 7, October 1995, pp 31-0 2. Pastukhov, V. G.; Maidanik, Yu. F.; Vershinin, C. V.; Korukov, M. A.; Miniature loop heat pipes for electronics cooling; Applied Thermal Engineering, Vol.23, Issue 9, June 2003, pp 1125-1135 3. Aitchison, J.M.; A numerical model of the dry spinning process; Proceedings of International Conference on Computational Fluid Dynamics 2002, Sydney, 15-19, July, 2002.. Cruchaga, M.; Celentano, D.; Numerical Analysis of Thermally Coupled Flow Problems with Interfaces and Phase-change Effects; International Journal of Computational Fluid Dynamics, Vol. 1, No., 2002, pp 27-22 5. Basara, B.; S. Jakirlic, S.; A new hybrid turbulence modelling strategy for industrial CFD; International Journal for Numerical Methods in Fluids; Vol.2, Issue 1, 2003, pp 9-11