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Numerical investigation of thermal and hydraulic performances of a condenser coil with oblique-shaped tubes
Accepted Manuscript
Numerical Investigation of Thermal and Hydraulic Performances of a Condenser Coil with Oblique-Shaped Tubes Kent Loong Khoo , Nuttawut Lewpiriyawong , Chuan Sun , Poh Seng Lee PII: DOI: Reference:
S0140-7007(18)30466-3 https://doi.org/10.1016/j.ijrefrig.2018.11.022 JIJR 4182
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
14 May 2018 13 November 2018 18 November 2018
Please cite this article as: Kent Loong Khoo , Nuttawut Lewpiriyawong , Chuan Sun , Poh Seng Lee , Numerical Investigation of Thermal and Hydraulic Performances of a Condenser Coil with Oblique-Shaped Tubes, International Journal of Refrigeration (2018), doi: https://doi.org/10.1016/j.ijrefrig.2018.11.022
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Highlights A 3-fin and 3-tube model is used to study the performance of oblique tube design
Up to 14% higher air-side heat transfer amount at same inlet velocity is achieved
Up to 14% lower air-side pressure drop at same inlet velocity is achieved
Oblique tube has smaller recirculation zone at the back compared to circular tube
Optimal parameters for oblique tube and plain fin design are presented
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Numerical Investigation of Thermal and Hydraulic Performances of a Condenser Coil with ObliqueShaped Tubes Kent Loong Khoo, Nuttawut Lewpiriyawong, Chuan Sun, Poh Seng Lee* Department of Mechanical Engineering, National University of Singapore, 9 Engineering Drive 1, Singapore, 117576 Tel: +65-65164187, *Email: [email protected] ABSTRACT
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High efficiency condenser coil providing higher heat removal capacity with lower air-side pressure drop is desirable. However, coil designs with different fin patterns for higher heat transfer incur large air-side pressure drop and thus consume more energy. In this study, the thermal and hydraulic performances of the condenser coil design with plain fin, corrugated fin, and novel oblique-shaped tube are evaluated numerically with a 3-fin and 3-tube model in comparison with a conventional circular tube coil. This novel oblique-shaped tube design features smaller air recirculation zone behind the tube so as to increase effective heat transfer area and to lower air-side pressure drop. The numerical results show that, for the plain fin configuration, the shape of the oblique tube is important in reducing the recirculation zone on the fin, thus increasing effective heat transfer area and heat transfer amount up to 14% and reducing the air-side pressure drop up to 14% at the same inlet air velocity as compared with the circular tube coil. With corrugated fin guiding and promoting air speed, causing more flow separation zones, it is detrimental to the thermal-hydraulic performance of the oblique tube coil. The discussion of choice of materials for tube and fin, such as copper, brass, and aluminium, is provided with possible manufacturing processes required for realizing potential high performance coil design. Parametric optimisation on the best performing plain fin oblique tube design shows that, for the design of condenser coil with oblique tube, there is no optimal FPI, but it is recommended to keep the ratio of tube pitch to tube frontal length/diameter to about 3 to 3.5.
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Subscripts i inlet
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Greek symbols ρ density, (kg/m3)
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NOMENCLATURE A area, (m2) cp heat capacity, (J/kg·K) LMTD log mean temperature difference, C or K m mass flow rate, (kg/s) P fan power, W p pressure, Pa Q heat transfer amount, W T temperature, C or K U overall heat transfer coefficient, (W/m2·K) u,v velocity, (m/s)
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KEY WORDS: fin and tube heat exchanger, oblique tube, corrugated fin, FTHX optimisation
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INTRODUCTION Refrigeration and air conditioning appliances are the main energy usage for Singapore households due to hot and humid weather. More than 50% of the electricity consumption of a typical household is for this usage [1]. Vapour compression cycle, which comprises of the compressor, expansion valve, evaporator, and condenser, is the main working process for such appliances in providing cooling and dehumidification for the thermal comfort of home occupants. Besides the compressor efficiency, the overall efficiency of the system is governed by the refrigerant thermodynamic conditions in the evaporator and condenser, which in turn are controlled by user requirements and outdoor environment respectively. Invertor-driven compressor technologies have been the primary efficiency driver in recent years. However, it often comes with a cost premium where the manufacturers often size a larger compressor than necessary to take advantage of the invertor control for better part load performance. Emerging cooling and dehumidification alternatives such as membrane technology, passive displacement ventilation, and radiant cooling are mainly still in testing phase to suit the hot and humid weather condition in Singapore. Improved condenser coil design could be an economical solution for price-sensitive household consumers, as heat exchangers usually only constitute a small fraction of the total manufacturing cost of the unit, and their performance will determine the required refrigerant thermodynamic condition for effective heat exchange with the environment. Finned tube heat exchangers (FTHX) are the common heat exchangers for household appliances due to low cost and large fin area for a given volume. An alternative to FTHX is by using mini-channel heat exchanger for air conditioning (AC) system. Although AC system using mini2
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channel heat exchanger has higher overall efficiency with lesser refrigerant charged [2-3], the mini-channel heat exchangers is still not widely adopted. This study will focus on addressing two main limitations for FTHX, namely (1) low air-side heat transfer coefficient compared to the high two-phase heat transfer coefficient in the tube, and (2) high thermal contact resistance due to imperfect contact between the mechanically joined fin and tube, as well as oxide and dust contamination over time. Various passive methods on fin designs, such as plate fin, wavy fin, louver fin, high performance compounded fin, and vortex generator, have been documented in literatures [4-7] and in handbook [8] to improve the air-side heat transfer coefficient. These fin designs are used to perturb the fluid boundary layer, to enhance fluid mixing by turbulent flow, and to increase air residence time for improved heat transfer. For example, complex interactions between the interrupted fin surface, namely wavy fin and compounded fin, and the tube are observed by flow visualisation in enhancing the heat transfer performance by boundary layer restarting, wake management, and vortices generation to achieve turbulent flow [9]. Heat transfer coefficient is increased by up to 24% and 45.5% for wavy fin and compounded fin respectively as compared to plate fin. However, the pressure drop is also increased by up to 31.9% and 63.1% respectively. A design of radiantly positioned winglets as vortex generators to enhance fluid mixing on a FTHX even has over 100% heat transfer improvement but over 140% increase in friction factor [10]. While heat transfer performance is improved significantly, higher friction loss associated with changing the main flow direction, perturbing the flow boundary layer, or disturbing the flow to create vortex, is also observed. This is detrimental to the overall system efficiency. To reduce the air-side pressure drop penalty for high heat transfer performance, oval obstacles instead of winglets are designed to reduce wake zones while still directing the main flow and perturbing the boundary layer to achieve high heat transfer [11], and a novel guiding channel and fusiform fin configurations to direct main flow to the back of the circular tubes [12] had shown promising for high performance FTHX. Enhancement of heat transfer by these fin designs, despite the higher pressure penalty, could be beneficial for applications where the overall energy efficiency of the system is not the main design objective. Changing the tube shape to be more streamline is another design direction in reducing the pressure drop penalty for FTHX. Smaller wake region and thus lower pressure drop of about 20 – 30% is observed in studies using oval or elliptic tube design with same tube perimeter [13-14] instead of circular tube. The form drag and friction drag of the oval tube design is lower due to smaller tube frontal area and lower local flow acceleration effect. Nevertheless, heat transfer rate of these finned oval tube designs is just slightly higher than the finned circular tube. A two-row plain finned oblique-shaped tube coil design is studied after referring to an oblique-finned microchannel heat sink design which has two times higher heat transfer at 50% lower pressure drop compared to a straight finned microchannel heat sink [15]. The two-row plain finned oblique-shaped tube coil design has 50% lower air-side pressure drop at the same air flow velocity and heat transfer when compared to the plain finned circular tube coil [16]. When compared to elliptic tube, the pressure drop is lower by up to 25% too. Oblique shape could induce a secondary flow at the back due to a slight pressure differential resulting from different flow path length between the two sides of the oblique shape. The wake zone behind the tube is reduced and the heat transfer at that region is improved. Conventional manufacturing method for FTHX is by tube mechanical expansion to join the fins and tubes. The contact between the fins and tube is thus less than perfect and the thermal contact resistance is hard to be determined due to different tube expansion ratio that happened during the process. It is however reported that the thermal contact resistance is about 15 – 25% of the total thermal resistance for FTHX with 7mm tube diameter [17]. Hence, this thermal contact resistance is significant and will affect the thermal performance of FTHX. To minimise the thermal contact resistance, integrated molding technology has been used in study to create a non-contact thermal resistance aluminium heat exchanger for the evaporator and condenser for household AC [18]. Both the refrigerating capacity and EER of the AC prototype are found to be higher by up to 23%. In addition, controlled atmospheric brazing method commonly used to produce all aluminium car radiator is also able to achieve better joining quality between the extended surface and fluid channels. A near perfect joining, which is achievable with current available manufacturing techniques, could significantly improve the capacity of FTHX and the efficiency of the AC as compared to current mechanical expansion joining method. However, since it is difficult to perfectly join copper tube and aluminium fin used in FTHX by brazing technique, as well as high copper price, studies on alternative tube materials like aluminium-copper composed tubes [19] and copper-clad aluminium tube [20] have been investigated to substitute copper tube for household AC. As the thermal conductivity of aluminium is only half that of copper, AC system with such composite tubes has slightly lower capacity and efficiency. Nevertheless, for AC system of less than 3kW capacity in the study, a reduction of material cost by up to 63% could be achieved. Thus it would be of interest to the industry to study the performance of FTHX using different material combination. In current numerical study, single-row oblique tube coil is designed for application in household AC. The manufacturability and mechanical strength of the oblique tube design is discussed. A FTHX model with 3-fin and 3-tube domain is proposed for the study of FTHX with non-symmetrical tube shape, and the subsequent numerical methodology is laid out. The performance of the plain fin oblique tube design is then compared to that of circular tube to study the effect of the tube shape. Moreover, corrugated fin, which is the baseline design in household AC, will be integrated with the new oblique tube design for comparison to the corrugated fin circular tube coil of the condenser coil. The overall performance factor for all the four FTHX designs is discussed and compared. The use of different choice of materials, such as copper, aluminium, and brass, will be discussed and performance of each material combination is generated for future design options. Lastly, parametric optimisation is performed on the fin thickness, fin pitch, and tube pitch for the plain fin oblique tube design. METHODOLOGIES 3
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Design and Manufacturability of Oblique-shaped Tube for Condenser Coil By measuring a common commercial 2.5kW single split household AC with corrugated fin circular tube (CFCT) condenser coil, the dimensions of the FTHX models in this study are established as shown in Fig. 1 and Table 1. The overall fin size, fin thickness, and fin gap are the same for each design. The design of the oblique-shaped tube (oblique tube) is based on the optimal parameters, namely 30 oblique angle and achievable flow alignment angle for the particular fin dimension [16]. The internal tube area for the oblique tube is kept almost the same as the circular tube to represent the condition of the same refrigerant flow rate and speed in order to compare the thermal-hydraulic performance of each design. For practical fabrication purpose, the internal and external corner radii of the oblique tube are rounded to at least 0.5mm. A 0.6mm thick internal rib is added to strengthen the oblique tube for household AC application using R410A refrigerant where the operating condensation pressure is about 22 – 28 bar. Von Mises stress analysis is then carried out using the commercial software, COMSOL Multiphysics 5.1 to confirm the mechanical strength of the oblique tube design. At an applied internal pressure of 30 bar, the maximum von Mises stress for the oblique tube with internal rib design is 82 MPa. As such, aluminium with yield strength of 241 MPa could be used for oblique tube extrusion. In order to compare with commercial FTHX using CFCT coil, the dimension of the corrugated fin design is measured and shown in Fig. 1c too. The performance of three FTHX designs, namely plain fin circular tube (PFCT), plain fin oblique tube (PFOT), and corrugated fin oblique tube (CFOT), are studied and compared with the baseline CFCT design. (b)
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Figure 1: Model and dimension (in mm) of (a) plain fin circular tube, (b) plain fin oblique tube, (c) corrugated fin
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Table 1: Geometrical dimension of the model for computation Item
Dimension (mm2)
copper
Circular tube internal area
22.1
Oblique tube material
aluminium
Oblique tube internal area
24.6
Fin material
aluminium
Fin area occupied by circular tube
36.5
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Circular tube material
Dimension (mm)
0.15
Fin area occupied by oblique tube
54.2
Fin gap
1.228
Heat transfer area (plain fin circular tube)
793
Fin size (width x depth)
22.1 x 19
Heat transfer area (plain fin oblique tube)
776
Number of tube row
1
Heat transfer area (corrugated fin circular tube)
799
Longitudinal tube pitch
22.1
Heat transfer area (corrugated fin oblique tube)
782
Circular tube size
6.82
Oblique tube size
10.1 x 4.71
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Fin thickness
Computational Domain and Numerical Methods A FTHX model with 3-fin span-wise length (66.3mm) and 3 tubes is introduced to study the thermal-hydraulic performance of the condenser coil with corrugated fin and plain fin for both circular tube and oblique tube. The use of 3-fin model eliminates the effect of adjacent finned tubes on the performance of the centre finned tube, where the results are extracted for performance 4
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comparison. The model is not further expanded to more tubes since the influence of the rest of the further finned tubes is expected to be less or insignificant. The computation domain is extended half fin depth (9.5mm) before the model to eliminate any non-uniformity of the flow and one fin depth (19mm) after the model to avoid backflow from the outlet. The periodical repetition of the finned tubes in the condenser coil is represented by setting the upper and lower surfaces of the computation domain as periodic boundary condition, while the sides of the domain are set as slip stationary wall [21] to eliminate the wall shear effect on the flow. Ideally, the sides of the domain should be set as periodic boundary as well, but there is mesh matching limitation of the ANSYS Workbench 15.0 software, preventing such setting. Moreover, the CFOT design is non-symmetric, thus the symmetry setting is not valid too. Nevertheless, it is shown in Table 2 that, with slip stationary wall, or symmetry boundary condition, the results of the centre finned tube domain for the PFCT model are the same (less than 1% difference). This is because both boundary conditions actually have the same mathematical equation, which is zero velocity gradient at the boundary. As such, slip stationary wall boundary condition is adopted for 3-fin and 3-tube model with non-symmetrical tube like oblique tube. To simulate the negative suction of the draw-through fan of the household AC, pressure inlet and pressure outlet with a target mass flow rate is used. Inlet air temperature is set at 35C as per Singapore outdoor temperature on a hot day. Air properties are temperature-dependent. The inner wall of the tubes is set as constant wall temperature of 45C, which is the constant operating condensation temperature of R410A refrigerant at 28bar. The 3D computation domain and boundary conditions of the PFOT model are shown in Fig. 2.
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Table 2: Comparison of thermal-hydraulic performance of the plain fin circular tube design with slip stationary wall and symmetry boundary conditions for the 3-fin and 3-tube model. The difference in heat transfer amount and air-side pressure drop is less than 1%. Slip Stationary Wall Q (W)
1.0
0.266
1.5
0.330
2.0
0.377
2.5
0.416
3.0
0.450
Q (W)
Δp (Pa)
4.89
0.266
4.89
8.13
0.330
8.12
11.9
0.377
12.0
16.6
0.416
16.6
21.9
0.450
21.9
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slip stationary wall
Symmetry
Δp (Pa)
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Air Velocity (m/s)
post domain (1 fin depth)
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pressure outlet
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upper and lower periodic
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constant temperature at inner tube wall
slip stationary wall
Figure 2: 3D computation domain and boundary conditions of the 3-fin and 3-tube model for the plain fin oblique tube design Different grid sizes are used for the fin domain, and extended flow domains as shown in Fig. 3. Due to flow separation and resulting rapid changes in air flow velocity near the tubes, fully turbulent flow is assumed between the fins. The conjugate heat transfer between the finned tube model and a turbulent flow governed by the realizable k-ε turbulence model is solved using the commercial CFD software, ANSYS Fluent 15.0. The numerical study of the thermal-hydraulic performance of the various finned tube models is carried out under the following assumptions: 5
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i. Steady-state fluid flow and heat transfer ii. Negligible viscous dissipation, natural convection and radiation heat transfer Iteration convergence is considered to be achieved when the set target mass flow rate corresponding to inlet air velocity of 1 – 3m/s at the inlet of the domain is reached and constant, with the following residual convergence criteria satisfied: i. Continuity: < 10-4 ii. Velocity components: < 10-6 iii. Energy: < 10-8 (b)
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Figure 3: Grid systems showing the centre domain of the 3-fin and 3-tube model for (a) the plain fin circular tube design, and (b) the plain fin oblique tube design. The element size is much smaller in the fin domain as compared to the pre- and post-domain.
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Data Reduction, Grid Independence Assessment, Validation of Results Since the mass flow rate and the overall fin size are the same for both the circular tube and oblique tube designs, as well as the objective of the study is to compare the practical total heat transfer amount and air-side pressure drop of the designs for household AC with a fixed size coil, parameters used for direct comparison of the thermal and hydraulic performance of different finned tube designs are the air-side pressure drop (Δp), theoretical fan power (P), and heat transfer amount (Q) as per Eq. (1) - (3). In addition, the log mean temperature difference (LMTD) and overall heat transfer coefficient (U) are evaluated to understand the effect of the oblique tube and corrugated fin on the air-side heat transfer performance as per Eq. (4) – (5).
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where pinlet, poutlet is the mean inlet and mean outlet air pressure (Pa), mair is the mass flow rate of air (kg/s), air is the mean air density (kg/m3), cp air is the mean specific heat capacity of air (J/kg·K), Tair inlet, Tair outlet is the mean inlet and mean outlet air temperature (K), Twall,condensation is the constant refrigerant condensation temperature (45C) at the internal wall of the tube, and Aheat transfer is the total heat transfer area (m2). Grid independence assessment was conducted by comparing U and Δp for the PFCT model and the PFOT model with three grid systems of increasing element numbers as shown in Table 3. The difference in U and Δp between the grid systems with approximately 3.3 million and 10 million elements is less than 0.3% for the PFCT model, while for the PFOT model, the difference in U and Δp between the grid systems with approximately 2.9 million and 9.5 million elements is less than 1.8%. As such, it is considered practically acceptable to adopt the grid system with approximately 2.9 to 3.3 million elements for current study. To validate the numerical method and results in this study, a mechanically expanded test coil with PFOT design is fabricated and tested in a recirculating wind tunnel with air flow and temperature control. A residential air conditioner is used in this setup to provide the refrigerant flow at saturation temperature into the test coil. Measurement of air-side pressure drop across the test coil by a differential pressure transducer at different inlet air velocity is used to validate the air-side pressure drop results from the numerical model in this paper. The Q obtained from the experiment results could not be used for validation as there is an unaccountable contact resistance between the fins and tubes of the mechanically expanded test coil. Thus the deviation would be large compared to the Q of the finned tube with zero contact resistance from numerical results. Nevertheless, the numerical and experimental Δp results of the PFOT model could be compared and are shown in Fig. 4. The maximum deviation between the results is less than 2.5% for inlet air velocity of 1 to 3m/s. As the flow between the fins is primarily characterised by the flow 6
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separation caused by the shape of the tube, the use of the realizable k-ε turbulence model and enhanced wall treatment for flow separation [21] is validated by the experiment results for inlet air velocity of 1 to 3m/s. Table 3: Numerical Results of Different Element Numbers for Plain Fin Circular Tube Model and Plain Fin Oblique Tube Model for Grid Independence Assessment U (W/m2·K)
Δp (Pa)
68.89 67.99 68.19
12.09 11.98 11.96
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Total Element Number Plain Fin Circular Tube Model 1,654,820 3,357,189 10,068,912 Plain Fin Oblique Tube Model 1,344,056 2,948,110 9,529,168
88.30 82.97 83.11
11.70 10.88 11.07
(Simulation) Pressure Drop_Plain Fin Oblique Tube (Experiment) Pressure Drop_Plain Fin Oblique Tube
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Air Velocity (m/s)
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Figure 4: Validation of the air-side pressure drop (Δp) of the plain fin oblique tube model from numerical results with experiment results. The maximum deviation is less than 2.5% for inlet air velocity of 1 to 3m/s
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Thermal and Hydraulic Performances of PFOT Design To investigate the effect of the oblique tube on the thermal-hydraulic performance of FTHX, the Q and Δp of the PFCT and PFOT are compared in Fig. 5a. For oblique tube with same internal area as the circular tube, the Q is increased by 5.6% to 14.4% while the Δp is reduced by 2.7% to 14.4% at the same inlet air velocity. Fin designs typically have higher heat transfer as well as higher pressure drop penalty [9-10]. Using oblique tube in FTHX, the Q could be improved even with lower Δp. By inspecting the temperature contour and velocity streamline at the centre domain of the computation domain of both the PFCT and PFOT in Fig. 5b, smaller recirculation zone behind the tube (lower form drag), and less abrupt change of direction of the air flow streamline are observed for the PFOT design, thus lower Δp. For effective heat transfer between the FTHX with the air flow, higher temperature gradient and higher air flow velocity at most of the fin area are desired. Although the average air flow velocity in the PFOT domain is lower than the PFCT case, the Q is still higher as there is more fin area with higher air flow velocity for effective heat transfer. For example, at inlet air velocity of 2m/s, the percentage of air flow region with velocity lower than 0.5m/s for PFOT is just 3.5% while it is 10.7% for the PFCT, resulting in higher Q of the former.
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Heat Transfer Amount, Q (W)
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Air velocity (m/s)
Figure 5: (a) Heat transfer amount (Q) and air-side pressure drop (Δp) of plain fin circular tube and plain fin oblique tube at inlet air velocity of 1 to 3m/s, (b) centre plane temperature contour and velocity streamline of plain fin circular tube and plain fin oblique tube at 2m/s
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Discussion on Thermal and Hydraulic Performances of CFOT Design The CFCT commonly applied for household AC is the baseline for comparison with the plain fin and CFOT designs. The thermal-hydraulic performance of all four FTHX designs is as shown in Fig. 6a. The corrugated fin is found to improve the Q of the circular tube by 1.3% to 6.7% at a higher Δp of 9.9% to 15.4%. Improvement in heat transfer performance with increase in pressure drop penalty is expected for CFCT as per other fin designs in the literatures. However, for the CFOT design, the Q is decreased slightly and pressure drop is increased up to 10.7% as compared to PFOT design. The corrugated fin design is detrimental to the overall performance of the coil with oblique tube. Referring to the cross-sectional plane velocity contour in Fig. 6b, air flow is redirected according to the contour of the corrugated fin, increasing the maximum air flow velocity by 5.1% and 5.8% for the CFCT and CFOT as compared to the plain fin designs respectively. The increase in air velocity is supposed to increase the overall heat transfer of the FTHX. More air flow separation zones are also present, of which 12.8% of the air flow region in the CFCT and 7.8% in CFOT design having air velocity lower than 0.5m/s. The combined effects of higher air flow velocity but smaller effective heat transfer area due to flow separation make the corrugated fin design beneficial for circular tube but detrimental for oblique tube. By comparing the four designs in Fig. 6a, PFOT has the best overall performance with about 7% higher heat transfer, 11.5% to 25.8% lower Δp than the baseline CFCT design in condenser coil of household AC. While the total Q and Δp are the practical parameters for comparison of different FTHX designs, the overall heat transfer coefficient, U which is normalised by total heat transfer area and LMTD, is an important parameter indicating the heat transfer rate of the design and is shown in Fig. 7a. U for the PFOT is about 12% higher than the CFCT. For the application of household AC, the heat transfer capacity of the design should be compared at the same fan power as shown in Fig. 7b. Compared with the CFCT used in the condenser coil of household AC and at the same fan power, the heat transfer capacity of the PFOT design is about 12% higher. Higher coil capacity achieved with the same fan power could reduce the refrigerant condensation temperature and pressure, resulting in lower compressor lift and power consumption, and higher COP of the AC eventually.
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Heat Transfer Amount, Q (W)
0.55
15
0.25
10
0.20
5
0.15
Air Pressure Drop, p (Pa)
Q_Plain Fin Circular Tube Q_Plain Fin Oblique Tube Q_Corrugated Fin Circular Tube Q_Corrugated Fin Oblique Tube Pressure Drop_Plain Fin Circular Tube Pressure Drop_Plain Fin Oblique Tube Pressure Drop_Corrugated Fin Circular Tube Pressure Drop_Corrugated Fin Oblique Tube
(a)
0
1.0
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2.0
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Air velocity (m/s)
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1.5
2.0
2.5
Q_Plain Fin Circular Tube Q_Plain Fin Oblique Tube Q_Corrugated Fin Circular Tube Q_Corrugated Fin Oblique Tube
(b) 0.55 Heat Transfer Amount, Q (W)
95
U_Plain Fin Circular Tube U_Plain Fin Oblique Tube U_Corrugated Fin Circular Tube U_Corrugated Fin Oblique Tube
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Heat Transfer Coefficient, U (W/m2K)
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Figure 6: (a) Heat transfer amount (Q) and air-side pressure drop (Δp) of plain fin circular tube, plain fin oblique tube, corrugated fin circular tube, and corrugated fin oblique tube at inlet air velocity of 1 to 3m/s, (b) cross-sectional plane velocity contour of the corrugated fin circular tube
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20
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Fan Power, P x10-4 (W)
Air Velocity, ui (m/s)
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Figure 7: (a) Overall heat transfer coefficient (U) against inlet air velocity of 1 to 3m/s for four FTHX designs, (b) heat transfer amount (Q) against the theoretical fan power for four FTHX designs Choice of Materials and Manufacturability for PFOT Condenser Coil Typical FTHX for household AC is made of copper circular tube, aluminium fin, and joined by tube mechanical expansion technique. As the maximum von Mises stress for oblique tube design is more than the yield strength of copper, aluminium could be used for both the tube and fin to withstand the high refrigerant pressure. Tube mechanical expansion technique as well as controlled atmospheric brazing technique applied in making all-aluminium car radiator is feasible for manufacturing allaluminium PFOT condenser coil. In addition, brass oblique tube and copper fin is another material design option as the mechanical strength of brass is of similar order to aluminium, and there is brazing technique joining brass tube and copper fin, commonly used for high performance heat exchanger for cooling of diesel engine generator. The mechanical and thermal properties of copper, aluminium and brass for FTHX are tabulated in Table 4. The heat transfer performance of the PFOT design with Cu-tube-Al-fin, Al-tube-Al-fin, and Br-tube-Cu-fin is shown in Fig. 8. The influence of the tube material is negligible as seen from the results of Cu-tube-Al-fin and Al-tube-Al-fin cases due to small tube thickness. The use of copper fin improves the 9
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Q up to 6.3% as compared to Al-tube-Al-fin case, and up to 14% higher than the CFCT case. Al-tube-Al-fin design, which already has higher performance, lighter and cheaper in material cost, could be used to replace the existing condenser coil in household AC. When high performance heat exchanger is needed, Br-tube-Cu-fin design could be used. Table 4: Mechanical and Thermal Properties of Three Common Materials for FTHX Aluminium (Al)
Brass (Br)
8,978
2,719
8,580
381
871
380
387.6
202.4
111
70
241
200
Q_Corrugated Fin Circular Tube Q_Plain Fin Oblique Tube_Al Tube Al Fin Q_Plain Fin Oblique Tube_Cu Tube Al Fin Q_Plain Fin Oblique Tube_Br Tube Cu Fin
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1.5
2.0
2.5
3.0
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Air Velocity (m/s)
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Figure 8: Heat transfer amount (Q) for plain fin oblique tube design with different choice of materials at inlet air velocity of 1 to 3m/s Parametric Optimisation of PFOT Condenser Coil Design
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Effect of Fin Thickness In the market of FTHX for household AC, common aluminium fin thickness used is 0.1mm and 0.15mm. Apparently, FTHX with 0.1mm thick fin will save about 50% fin material and about 50% lighter as compared to FTHX with 0.15mm fin. The downside of a thinner fin is that lower strength FTHX will be more susceptible to damage during handling and maintenance of the FTHX. The thermal performances of the PFOT design with 0.1mm and 0.15mm fin thickness are plotted at the same theoretical fan power as shown in Fig. 9. The overall fin domain is maintained at 1.4mm for both cases. As such, the design with 0.1mm fin thickness will have a fin-to-fin gap of 1.3mm, and for that with 0.15mm fin thickness, the fin-to-fin gap is 1.25mm. For the same inlet air velocity, a smaller fin gap leads to higher air flow velocity across the fin surface, resulting in higher Q and higher Δp. Nevertheless, the increase in skin friction due to higher air velocity is not significant, thus when compared at the same fan power, the Q of PFOT design with 0.15mm fin thickness is about 6% higher than the 0.1mm fin thickness case. A separate study is also conducted by keeping the same fin-to-fin gap for both the 0.1mm and 0.15mm fin designs. It is found that the difference in thermal and hydraulic performance is negligible due to similar air flow velocity across the fin surface and similar total heat transfer area for both designs. As the heat transfer improvement is not significant and the motivation in the industry to reduce material cost and weight of the FTHX [19-20], 0.1mm fin thickness for the PFOT are be adopted for subsequent parametric optimisation study.
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Figure 9: Heat transfer amount (Q) for plain fin oblique tube design with 0.1mm and 0.15mm fin thickness at the same theoretical fan power
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Effect of Fin Pitch Common fin per inch (FPI) for FTHX in the industry ranges from 15 to 21. The effect of FPI on the Q per inch of fin at the same inlet air velocity for the PFOT design is plotted in Fig. 10a. Higher FPI leads to higher heat transfer as there is more fin area for heat transfer. Referring to Fig. 10b, at the same fan power, the Q of a coil with higher FPI is significantly higher than a coil with a lower FPI but at a higher air flow velocity. For example, coil with 21 FPI at 2m/s consumes the same amount of fan power, but has over 30% higher Q per inch of fin than a coil with 15 FPI at 2.5m/s. This shows the effect of fin area per inch of coil is a more important factor in thermal performance of the coil than air flow velocity. For high performance coil, high FPI and small air flow velocity should be used to achieve high thermal performance without significant air-side pressure drop penalty. Nevertheless, the material cost and weight of the FTHX with high FPI will be higher too. There is no optimum FPI for PFOT design. FPI for PFOT design should be selected considering the thermal requirement of the applications as well as the material cost for the coil.
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Figure 10: (a) Heat transfer amount (Q) per inch of fin for plain fin oblique tube coil with 15 to 21 FPI at inlet air velocity of 1 to 3m/s, (b) Heat transfer amount (Q) per inch of fin for plain fin oblique tube coil with 15 to 21 FPI against fan power Effect of Tube Pitch PFOT with 18 FPI as per baseline coil is chosen for the study the effect of tube pitch on the thermal-hydraulic performance of the coil. The range of tube pitch is from 11.6mm to 22.1mm, which is equivalent to about 2.5 to 4.5 times the frontal length of the oblique tube. The thermal-hydraulic performance is calculated for per fin length of 527mm as per actual condenser coil size because each computational model is built with different tube pitch, thus different fin size and area. The results are shown in Fig. 11a and 11b. The effect of tube pitch on the total Q per fin length of 527mm is mostly minute as the total fin area is similar. Smaller tube pitch means there are more tubes per fin, but since the tube heat transfer area is small compared to the fin area, the drop in Q is gradual when the tube pitch increases as shown in Fig. 11a. When compared at the same fan power in Fig. 11b, an exponentially decreasing trend showing the flattening of Q per fin when the fan power increases is observed at 14mm and 16.3mm tube pitch, which are equivalent to 3 to 3.5 times of the oblique tube frontal length. Further increase in fan power input does not result in higher heat transfer performance. Condenser with 16.3mm tube pitch has fewer tubes than that of 14mm for similar amount of heat transfer. Thus the optimal tube pitch for oblique tube design is 16.3mm, which is equivalent to about 3.5 11
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times of the tube frontal length. This optimal tube pitch to tube frontal length ratio is actually close to the 3.2 ratio for the baseline corrugated fin circular tube design with a tube pitch of 22.1mm and tube frontal diameter of 6.82mm.
Heat Transfer Amount per Fin (W)
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Figure 11: (a) Heat transfer amount (Q) per fin of 527mm for plain fin oblique tube coil with 11.6 to 22.1mm tube pitch at inlet air velocity of 1 to 3m/s, (b) Heat transfer amount (Q) per fin of 527mm for plain fin oblique tube coil with 11.6 to 22.1mm tube pitch against fan power
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CONCLUSIONS In this paper, the characteristics of the PFOT and CFOT coil are studied and compared to the baseline circular tube coil. Choice of materials and parameter optimisations of the PFOT design are also made. The conclusions are summarized as follows: 1. The oblique tube has a smaller frontal area, resulting in smaller recirculation zone behind the tube and less abrupt change of air flow direction, thus increases the effective heat transfer area and heat transfer amount up to 14% and lower air-side pressure drop up to 14%. 2. The corrugated fin design is detrimental to the overall performance of the coil with oblique tube due to more induced air flow separation zones and disruption of the smooth air flow across the oblique tube. This results in lower heat transfer and higher airside pressure drop up to 10%. 3. Under steady-state operation condition, the effect of tube material on the coil thermal performance is insignificant. Copper fin is recommended for high performance heat exchanger. 4. The selection of fin thickness and FPI for PFOT coil depends on the heat dissipation requirement, the weight of the coil, and overall material cost constraint for the application. It is recommended to keep the ratio of the tube pitch to tube frontal length/diameter to 3 to 3.5 for maximum heat transfer at given air flow velocity.
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ACKNOWLEDGMENTS The authors would like to acknowledge the funding support from the National Research Foundation (NRF) Singapore under the Energy Innovation Research Programme (EIRP) Funding Scheme (NRF2013EWT-EIRP004-038) managed on behalf by the Building and Construction Authority (BCA) of Singapore.
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REFERENCES [1] Energy Efficiency Programme Office (E2PO) (2012). Study to Characterise Energy Consumption in Singapore Households [online]. Available at: http://www.e2singapore.gov.sg/Households/Saving_Energy_At_Home/Energy_Consumption_Survey.aspx [Accessed 03 March 2017] [2] C. Y. Park and P. Hrnjak, “Experimental and Numerical Study on Microchannel and Round-tube Condensers in a R410A Residential Air-conditioning System,” International Journal of Refrigeration, vol. 31, pp. 822-831, 2008. [3] T. Dixit and I. Ghosh, “Review of Micro- and Mini-channel Heat Sinks and Heat Exchangers for Single Phase Fluids,” Renewable and Sustainable Energy Reviews, vol. 41, pp. 1298-1311, 2015. [4] T. A. Tahseen et al., “An Overview on Thermal and Fluid Flow Characteristics in a Plain Plate Finned and Un-finned Tube Banks Heat Exchanger,” Renewable and Sustainable Energy Reviews, vol. 43, pp. 363-380, 2015. [5] P. Date and V. W. Khond, “Heat Transfer Enhancement in Fin and Tube Heat Exchanger – A Review,” ARPN Journal of Engineering and Applied Sciences, vol. 8, no. 3, pp. 241-245, March 2013. [6] C. C. Wang, “A Survey of Recent Patents of Fin-and-tube Heat Exchangers from 2001 to 2009,” International Journal of AirConditioning and Refrigeration, vol. 18, no. 1, pp. 1-13, 2010. [7] S. A. E. S. Ahmed et al., “Flow and Heat Transfer Enhancement in Tube Heat Exchangers,” Heat Mass Transfer, vol. 51, pp. 1607-1630, 2015. [8] R. L. Webb and N. H. Kim, Principles of Enhanced Heat Transfer, 2nd Edition, Boca Raton, Taylor & Francis, 2005. 12
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[9] M. Y. Wen and C. Y. Ho, “Heat-Transfer Enhancement in Fin-and-Tube Heat Exchanger with Improved Fin Design,” Applied Thermal Engineering, 29 (5), pp. 1050-1057, 2009. [10] M. J. Li, W. J. Zhou, J. F. Zhang, J. F. Fan, Y. L. He, and W. Q. Tao, “Heat Transfer and Pressure Performance of a Plain Fin with Radiantly Arranged Winglets around Each Tube in Fin-and-tube Heat Transfer Surface,” International Journal of Heat and Mass Transfer, 70, pp. 734-744, 2014. [11] N. Lewpiriyawong, C. Sun, K. L. Khoo, P. S. Lee, and S. K. Chou, “Enhanced Air-Side Performance of Finned Tube Heat Exchanger with Oval Obstacles for Residential Air-Conditioning Systems – Numerical Approach,” 15th IEEE Intersociety Conference on Thermal and Thermomechanical Phenomena in Electronic Systems (ITherm). Las Vegas, NV: IEEE, pp. 633-638, 2016. [12] C. Sun, N. Lewpiriyawong, K. L. Khoo, P. S. Lee, and S. K. Chou, (2016), “Numerical Modelling and Thermal Optimization of Finned Tube Heat Exchanger with Guiding Channel and Fusiform Configurations,” 15th IEEE Intersociety Conference on Thermal and Thermomechanical Phenomena in Electronic Systems (ITherm). Las Vegas, NV: IEEE, pp. 1099-1106, 2016. [13] H. Han et al., “A Numerical Study on Compact Enhanced Fin-and-tube Heat Exchangers with Oval and Circular Tube Configurations,” International Journal of Heat and Mass Transfer, vol. 65, pp. 686-695, 2013. [14] J. Y. Jang and J. Y. Yang, “Experimental and 3-D Numerical Analysis of the Thermal-Hydraulic Characteristics of Elliptic Finned-Tube Heat Exchangers,” Heat Transfer Engineering, vol. 19, no. 4, pp. 55-67, 1998. [15] Y. J. Lee et al., “Enhanced Thermal Transport in Microchannel using Oblique Fins,” Journal of Heat Transfer, vol. 134, no. 10, 101901, Oct. 2012. [16] K. L. Khoo, N. Lewpiriyawong, C. Sun, P. S. Lee, and S. K. Chou, “Numerical Investigation of the Thermal-Hydraulic Performance of Finned Oblique-Shaped Tube Heat Exchanger,” 15th IEEE Intersociety Conference on Thermal and Thermomechanical Phenomena in Electronic Systems (ITherm). Las Vegas, NV: IEEE, pp. 625-632, 2016. [17] J. Jeong, C. N. Kim, and B. Youn, (2006), “A Study on the Thermal Conductance in Fin-Tube Heat Exchangers with 7mm Tube,” International Journal of Heat and Mass Transfer, 49, pp. 1547-1555, 2006. [18] R. Jia, Y. Wang, H. Wang, and J. Rong, “Optimization Design and Test on Non-Contact Thermal Resistance Heat Exchanger of Air-Condition,” Journal of Huazhong University of Science and Technology (Natural Science Edition), 42, pp. 12-16, 2014. [19] S. Wang, S. Cheng, H. Yu, Z. Rao, and Z. Liu, “Experimental Investigation of Al-Cu Composed Tube-Fin Heat Exchangers for Air Conditioner,” Experimental Thermal and Fluid Science, 51, pp. 264-270, 2013. [20] T. Jin, G. Li, Y. Cao, R. Xu, S. Shao, and B. Yang, “Experimental Research on Applying the Copper-Clad Aluminium Tube as Connecting Tubes of Air Conditioners,” Energy and Buildings, 97, pp. 1-5, 2015. [21] Fluent, A.N.S.Y.S., 15.0 Documentation, Ansys Inc., 2013.
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Numerical Investigation of Thermal and Hydraulic Performances of a Condenser Coil with Oblique-Shaped Tubes Kent Loong Khoo , Nuttawut Lewpiriyawong , Chuan Sun , Poh Seng Lee PII: DOI: Reference:
S0140-7007(18)30466-3 https://doi.org/10.1016/j.ijrefrig.2018.11.022 JIJR 4182
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
14 May 2018 13 November 2018 18 November 2018
Please cite this article as: Kent Loong Khoo , Nuttawut Lewpiriyawong , Chuan Sun , Poh Seng Lee , Numerical Investigation of Thermal and Hydraulic Performances of a Condenser Coil with Oblique-Shaped Tubes, International Journal of Refrigeration (2018), doi: https://doi.org/10.1016/j.ijrefrig.2018.11.022
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Highlights A 3-fin and 3-tube model is used to study the performance of oblique tube design
Up to 14% higher air-side heat transfer amount at same inlet velocity is achieved
Up to 14% lower air-side pressure drop at same inlet velocity is achieved
Oblique tube has smaller recirculation zone at the back compared to circular tube
Optimal parameters for oblique tube and plain fin design are presented
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Numerical Investigation of Thermal and Hydraulic Performances of a Condenser Coil with ObliqueShaped Tubes Kent Loong Khoo, Nuttawut Lewpiriyawong, Chuan Sun, Poh Seng Lee* Department of Mechanical Engineering, National University of Singapore, 9 Engineering Drive 1, Singapore, 117576 Tel: +65-65164187, *Email: [email protected] ABSTRACT
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High efficiency condenser coil providing higher heat removal capacity with lower air-side pressure drop is desirable. However, coil designs with different fin patterns for higher heat transfer incur large air-side pressure drop and thus consume more energy. In this study, the thermal and hydraulic performances of the condenser coil design with plain fin, corrugated fin, and novel oblique-shaped tube are evaluated numerically with a 3-fin and 3-tube model in comparison with a conventional circular tube coil. This novel oblique-shaped tube design features smaller air recirculation zone behind the tube so as to increase effective heat transfer area and to lower air-side pressure drop. The numerical results show that, for the plain fin configuration, the shape of the oblique tube is important in reducing the recirculation zone on the fin, thus increasing effective heat transfer area and heat transfer amount up to 14% and reducing the air-side pressure drop up to 14% at the same inlet air velocity as compared with the circular tube coil. With corrugated fin guiding and promoting air speed, causing more flow separation zones, it is detrimental to the thermal-hydraulic performance of the oblique tube coil. The discussion of choice of materials for tube and fin, such as copper, brass, and aluminium, is provided with possible manufacturing processes required for realizing potential high performance coil design. Parametric optimisation on the best performing plain fin oblique tube design shows that, for the design of condenser coil with oblique tube, there is no optimal FPI, but it is recommended to keep the ratio of tube pitch to tube frontal length/diameter to about 3 to 3.5.
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Subscripts i inlet
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Greek symbols ρ density, (kg/m3)
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NOMENCLATURE A area, (m2) cp heat capacity, (J/kg·K) LMTD log mean temperature difference, C or K m mass flow rate, (kg/s) P fan power, W p pressure, Pa Q heat transfer amount, W T temperature, C or K U overall heat transfer coefficient, (W/m2·K) u,v velocity, (m/s)
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KEY WORDS: fin and tube heat exchanger, oblique tube, corrugated fin, FTHX optimisation
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INTRODUCTION Refrigeration and air conditioning appliances are the main energy usage for Singapore households due to hot and humid weather. More than 50% of the electricity consumption of a typical household is for this usage [1]. Vapour compression cycle, which comprises of the compressor, expansion valve, evaporator, and condenser, is the main working process for such appliances in providing cooling and dehumidification for the thermal comfort of home occupants. Besides the compressor efficiency, the overall efficiency of the system is governed by the refrigerant thermodynamic conditions in the evaporator and condenser, which in turn are controlled by user requirements and outdoor environment respectively. Invertor-driven compressor technologies have been the primary efficiency driver in recent years. However, it often comes with a cost premium where the manufacturers often size a larger compressor than necessary to take advantage of the invertor control for better part load performance. Emerging cooling and dehumidification alternatives such as membrane technology, passive displacement ventilation, and radiant cooling are mainly still in testing phase to suit the hot and humid weather condition in Singapore. Improved condenser coil design could be an economical solution for price-sensitive household consumers, as heat exchangers usually only constitute a small fraction of the total manufacturing cost of the unit, and their performance will determine the required refrigerant thermodynamic condition for effective heat exchange with the environment. Finned tube heat exchangers (FTHX) are the common heat exchangers for household appliances due to low cost and large fin area for a given volume. An alternative to FTHX is by using mini-channel heat exchanger for air conditioning (AC) system. Although AC system using mini2
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channel heat exchanger has higher overall efficiency with lesser refrigerant charged [2-3], the mini-channel heat exchangers is still not widely adopted. This study will focus on addressing two main limitations for FTHX, namely (1) low air-side heat transfer coefficient compared to the high two-phase heat transfer coefficient in the tube, and (2) high thermal contact resistance due to imperfect contact between the mechanically joined fin and tube, as well as oxide and dust contamination over time. Various passive methods on fin designs, such as plate fin, wavy fin, louver fin, high performance compounded fin, and vortex generator, have been documented in literatures [4-7] and in handbook [8] to improve the air-side heat transfer coefficient. These fin designs are used to perturb the fluid boundary layer, to enhance fluid mixing by turbulent flow, and to increase air residence time for improved heat transfer. For example, complex interactions between the interrupted fin surface, namely wavy fin and compounded fin, and the tube are observed by flow visualisation in enhancing the heat transfer performance by boundary layer restarting, wake management, and vortices generation to achieve turbulent flow [9]. Heat transfer coefficient is increased by up to 24% and 45.5% for wavy fin and compounded fin respectively as compared to plate fin. However, the pressure drop is also increased by up to 31.9% and 63.1% respectively. A design of radiantly positioned winglets as vortex generators to enhance fluid mixing on a FTHX even has over 100% heat transfer improvement but over 140% increase in friction factor [10]. While heat transfer performance is improved significantly, higher friction loss associated with changing the main flow direction, perturbing the flow boundary layer, or disturbing the flow to create vortex, is also observed. This is detrimental to the overall system efficiency. To reduce the air-side pressure drop penalty for high heat transfer performance, oval obstacles instead of winglets are designed to reduce wake zones while still directing the main flow and perturbing the boundary layer to achieve high heat transfer [11], and a novel guiding channel and fusiform fin configurations to direct main flow to the back of the circular tubes [12] had shown promising for high performance FTHX. Enhancement of heat transfer by these fin designs, despite the higher pressure penalty, could be beneficial for applications where the overall energy efficiency of the system is not the main design objective. Changing the tube shape to be more streamline is another design direction in reducing the pressure drop penalty for FTHX. Smaller wake region and thus lower pressure drop of about 20 – 30% is observed in studies using oval or elliptic tube design with same tube perimeter [13-14] instead of circular tube. The form drag and friction drag of the oval tube design is lower due to smaller tube frontal area and lower local flow acceleration effect. Nevertheless, heat transfer rate of these finned oval tube designs is just slightly higher than the finned circular tube. A two-row plain finned oblique-shaped tube coil design is studied after referring to an oblique-finned microchannel heat sink design which has two times higher heat transfer at 50% lower pressure drop compared to a straight finned microchannel heat sink [15]. The two-row plain finned oblique-shaped tube coil design has 50% lower air-side pressure drop at the same air flow velocity and heat transfer when compared to the plain finned circular tube coil [16]. When compared to elliptic tube, the pressure drop is lower by up to 25% too. Oblique shape could induce a secondary flow at the back due to a slight pressure differential resulting from different flow path length between the two sides of the oblique shape. The wake zone behind the tube is reduced and the heat transfer at that region is improved. Conventional manufacturing method for FTHX is by tube mechanical expansion to join the fins and tubes. The contact between the fins and tube is thus less than perfect and the thermal contact resistance is hard to be determined due to different tube expansion ratio that happened during the process. It is however reported that the thermal contact resistance is about 15 – 25% of the total thermal resistance for FTHX with 7mm tube diameter [17]. Hence, this thermal contact resistance is significant and will affect the thermal performance of FTHX. To minimise the thermal contact resistance, integrated molding technology has been used in study to create a non-contact thermal resistance aluminium heat exchanger for the evaporator and condenser for household AC [18]. Both the refrigerating capacity and EER of the AC prototype are found to be higher by up to 23%. In addition, controlled atmospheric brazing method commonly used to produce all aluminium car radiator is also able to achieve better joining quality between the extended surface and fluid channels. A near perfect joining, which is achievable with current available manufacturing techniques, could significantly improve the capacity of FTHX and the efficiency of the AC as compared to current mechanical expansion joining method. However, since it is difficult to perfectly join copper tube and aluminium fin used in FTHX by brazing technique, as well as high copper price, studies on alternative tube materials like aluminium-copper composed tubes [19] and copper-clad aluminium tube [20] have been investigated to substitute copper tube for household AC. As the thermal conductivity of aluminium is only half that of copper, AC system with such composite tubes has slightly lower capacity and efficiency. Nevertheless, for AC system of less than 3kW capacity in the study, a reduction of material cost by up to 63% could be achieved. Thus it would be of interest to the industry to study the performance of FTHX using different material combination. In current numerical study, single-row oblique tube coil is designed for application in household AC. The manufacturability and mechanical strength of the oblique tube design is discussed. A FTHX model with 3-fin and 3-tube domain is proposed for the study of FTHX with non-symmetrical tube shape, and the subsequent numerical methodology is laid out. The performance of the plain fin oblique tube design is then compared to that of circular tube to study the effect of the tube shape. Moreover, corrugated fin, which is the baseline design in household AC, will be integrated with the new oblique tube design for comparison to the corrugated fin circular tube coil of the condenser coil. The overall performance factor for all the four FTHX designs is discussed and compared. The use of different choice of materials, such as copper, aluminium, and brass, will be discussed and performance of each material combination is generated for future design options. Lastly, parametric optimisation is performed on the fin thickness, fin pitch, and tube pitch for the plain fin oblique tube design. METHODOLOGIES 3
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Design and Manufacturability of Oblique-shaped Tube for Condenser Coil By measuring a common commercial 2.5kW single split household AC with corrugated fin circular tube (CFCT) condenser coil, the dimensions of the FTHX models in this study are established as shown in Fig. 1 and Table 1. The overall fin size, fin thickness, and fin gap are the same for each design. The design of the oblique-shaped tube (oblique tube) is based on the optimal parameters, namely 30 oblique angle and achievable flow alignment angle for the particular fin dimension [16]. The internal tube area for the oblique tube is kept almost the same as the circular tube to represent the condition of the same refrigerant flow rate and speed in order to compare the thermal-hydraulic performance of each design. For practical fabrication purpose, the internal and external corner radii of the oblique tube are rounded to at least 0.5mm. A 0.6mm thick internal rib is added to strengthen the oblique tube for household AC application using R410A refrigerant where the operating condensation pressure is about 22 – 28 bar. Von Mises stress analysis is then carried out using the commercial software, COMSOL Multiphysics 5.1 to confirm the mechanical strength of the oblique tube design. At an applied internal pressure of 30 bar, the maximum von Mises stress for the oblique tube with internal rib design is 82 MPa. As such, aluminium with yield strength of 241 MPa could be used for oblique tube extrusion. In order to compare with commercial FTHX using CFCT coil, the dimension of the corrugated fin design is measured and shown in Fig. 1c too. The performance of three FTHX designs, namely plain fin circular tube (PFCT), plain fin oblique tube (PFOT), and corrugated fin oblique tube (CFOT), are studied and compared with the baseline CFCT design. (b)
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Figure 1: Model and dimension (in mm) of (a) plain fin circular tube, (b) plain fin oblique tube, (c) corrugated fin
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Table 1: Geometrical dimension of the model for computation Item
Dimension (mm2)
copper
Circular tube internal area
22.1
Oblique tube material
aluminium
Oblique tube internal area
24.6
Fin material
aluminium
Fin area occupied by circular tube
36.5
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Circular tube material
Dimension (mm)
0.15
Fin area occupied by oblique tube
54.2
Fin gap
1.228
Heat transfer area (plain fin circular tube)
793
Fin size (width x depth)
22.1 x 19
Heat transfer area (plain fin oblique tube)
776
Number of tube row
1
Heat transfer area (corrugated fin circular tube)
799
Longitudinal tube pitch
22.1
Heat transfer area (corrugated fin oblique tube)
782
Circular tube size
6.82
Oblique tube size
10.1 x 4.71
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Fin thickness
Computational Domain and Numerical Methods A FTHX model with 3-fin span-wise length (66.3mm) and 3 tubes is introduced to study the thermal-hydraulic performance of the condenser coil with corrugated fin and plain fin for both circular tube and oblique tube. The use of 3-fin model eliminates the effect of adjacent finned tubes on the performance of the centre finned tube, where the results are extracted for performance 4
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comparison. The model is not further expanded to more tubes since the influence of the rest of the further finned tubes is expected to be less or insignificant. The computation domain is extended half fin depth (9.5mm) before the model to eliminate any non-uniformity of the flow and one fin depth (19mm) after the model to avoid backflow from the outlet. The periodical repetition of the finned tubes in the condenser coil is represented by setting the upper and lower surfaces of the computation domain as periodic boundary condition, while the sides of the domain are set as slip stationary wall [21] to eliminate the wall shear effect on the flow. Ideally, the sides of the domain should be set as periodic boundary as well, but there is mesh matching limitation of the ANSYS Workbench 15.0 software, preventing such setting. Moreover, the CFOT design is non-symmetric, thus the symmetry setting is not valid too. Nevertheless, it is shown in Table 2 that, with slip stationary wall, or symmetry boundary condition, the results of the centre finned tube domain for the PFCT model are the same (less than 1% difference). This is because both boundary conditions actually have the same mathematical equation, which is zero velocity gradient at the boundary. As such, slip stationary wall boundary condition is adopted for 3-fin and 3-tube model with non-symmetrical tube like oblique tube. To simulate the negative suction of the draw-through fan of the household AC, pressure inlet and pressure outlet with a target mass flow rate is used. Inlet air temperature is set at 35C as per Singapore outdoor temperature on a hot day. Air properties are temperature-dependent. The inner wall of the tubes is set as constant wall temperature of 45C, which is the constant operating condensation temperature of R410A refrigerant at 28bar. The 3D computation domain and boundary conditions of the PFOT model are shown in Fig. 2.
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Table 2: Comparison of thermal-hydraulic performance of the plain fin circular tube design with slip stationary wall and symmetry boundary conditions for the 3-fin and 3-tube model. The difference in heat transfer amount and air-side pressure drop is less than 1%. Slip Stationary Wall Q (W)
1.0
0.266
1.5
0.330
2.0
0.377
2.5
0.416
3.0
0.450
Q (W)
Δp (Pa)
4.89
0.266
4.89
8.13
0.330
8.12
11.9
0.377
12.0
16.6
0.416
16.6
21.9
0.450
21.9
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slip stationary wall
Symmetry
Δp (Pa)
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Air Velocity (m/s)
post domain (1 fin depth)
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pressure outlet
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upper and lower periodic
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pre domain (1/2 fin depth) pressure inlet
constant temperature at inner tube wall
slip stationary wall
Figure 2: 3D computation domain and boundary conditions of the 3-fin and 3-tube model for the plain fin oblique tube design Different grid sizes are used for the fin domain, and extended flow domains as shown in Fig. 3. Due to flow separation and resulting rapid changes in air flow velocity near the tubes, fully turbulent flow is assumed between the fins. The conjugate heat transfer between the finned tube model and a turbulent flow governed by the realizable k-ε turbulence model is solved using the commercial CFD software, ANSYS Fluent 15.0. The numerical study of the thermal-hydraulic performance of the various finned tube models is carried out under the following assumptions: 5
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i. Steady-state fluid flow and heat transfer ii. Negligible viscous dissipation, natural convection and radiation heat transfer Iteration convergence is considered to be achieved when the set target mass flow rate corresponding to inlet air velocity of 1 – 3m/s at the inlet of the domain is reached and constant, with the following residual convergence criteria satisfied: i. Continuity: < 10-4 ii. Velocity components: < 10-6 iii. Energy: < 10-8 (b)
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Figure 3: Grid systems showing the centre domain of the 3-fin and 3-tube model for (a) the plain fin circular tube design, and (b) the plain fin oblique tube design. The element size is much smaller in the fin domain as compared to the pre- and post-domain.
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Data Reduction, Grid Independence Assessment, Validation of Results Since the mass flow rate and the overall fin size are the same for both the circular tube and oblique tube designs, as well as the objective of the study is to compare the practical total heat transfer amount and air-side pressure drop of the designs for household AC with a fixed size coil, parameters used for direct comparison of the thermal and hydraulic performance of different finned tube designs are the air-side pressure drop (Δp), theoretical fan power (P), and heat transfer amount (Q) as per Eq. (1) - (3). In addition, the log mean temperature difference (LMTD) and overall heat transfer coefficient (U) are evaluated to understand the effect of the oblique tube and corrugated fin on the air-side heat transfer performance as per Eq. (4) – (5).
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where pinlet, poutlet is the mean inlet and mean outlet air pressure (Pa), mair is the mass flow rate of air (kg/s), air is the mean air density (kg/m3), cp air is the mean specific heat capacity of air (J/kg·K), Tair inlet, Tair outlet is the mean inlet and mean outlet air temperature (K), Twall,condensation is the constant refrigerant condensation temperature (45C) at the internal wall of the tube, and Aheat transfer is the total heat transfer area (m2). Grid independence assessment was conducted by comparing U and Δp for the PFCT model and the PFOT model with three grid systems of increasing element numbers as shown in Table 3. The difference in U and Δp between the grid systems with approximately 3.3 million and 10 million elements is less than 0.3% for the PFCT model, while for the PFOT model, the difference in U and Δp between the grid systems with approximately 2.9 million and 9.5 million elements is less than 1.8%. As such, it is considered practically acceptable to adopt the grid system with approximately 2.9 to 3.3 million elements for current study. To validate the numerical method and results in this study, a mechanically expanded test coil with PFOT design is fabricated and tested in a recirculating wind tunnel with air flow and temperature control. A residential air conditioner is used in this setup to provide the refrigerant flow at saturation temperature into the test coil. Measurement of air-side pressure drop across the test coil by a differential pressure transducer at different inlet air velocity is used to validate the air-side pressure drop results from the numerical model in this paper. The Q obtained from the experiment results could not be used for validation as there is an unaccountable contact resistance between the fins and tubes of the mechanically expanded test coil. Thus the deviation would be large compared to the Q of the finned tube with zero contact resistance from numerical results. Nevertheless, the numerical and experimental Δp results of the PFOT model could be compared and are shown in Fig. 4. The maximum deviation between the results is less than 2.5% for inlet air velocity of 1 to 3m/s. As the flow between the fins is primarily characterised by the flow 6
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separation caused by the shape of the tube, the use of the realizable k-ε turbulence model and enhanced wall treatment for flow separation [21] is validated by the experiment results for inlet air velocity of 1 to 3m/s. Table 3: Numerical Results of Different Element Numbers for Plain Fin Circular Tube Model and Plain Fin Oblique Tube Model for Grid Independence Assessment U (W/m2·K)
Δp (Pa)
68.89 67.99 68.19
12.09 11.98 11.96
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Total Element Number Plain Fin Circular Tube Model 1,654,820 3,357,189 10,068,912 Plain Fin Oblique Tube Model 1,344,056 2,948,110 9,529,168
88.30 82.97 83.11
11.70 10.88 11.07
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Figure 4: Validation of the air-side pressure drop (Δp) of the plain fin oblique tube model from numerical results with experiment results. The maximum deviation is less than 2.5% for inlet air velocity of 1 to 3m/s
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Thermal and Hydraulic Performances of PFOT Design To investigate the effect of the oblique tube on the thermal-hydraulic performance of FTHX, the Q and Δp of the PFCT and PFOT are compared in Fig. 5a. For oblique tube with same internal area as the circular tube, the Q is increased by 5.6% to 14.4% while the Δp is reduced by 2.7% to 14.4% at the same inlet air velocity. Fin designs typically have higher heat transfer as well as higher pressure drop penalty [9-10]. Using oblique tube in FTHX, the Q could be improved even with lower Δp. By inspecting the temperature contour and velocity streamline at the centre domain of the computation domain of both the PFCT and PFOT in Fig. 5b, smaller recirculation zone behind the tube (lower form drag), and less abrupt change of direction of the air flow streamline are observed for the PFOT design, thus lower Δp. For effective heat transfer between the FTHX with the air flow, higher temperature gradient and higher air flow velocity at most of the fin area are desired. Although the average air flow velocity in the PFOT domain is lower than the PFCT case, the Q is still higher as there is more fin area with higher air flow velocity for effective heat transfer. For example, at inlet air velocity of 2m/s, the percentage of air flow region with velocity lower than 0.5m/s for PFOT is just 3.5% while it is 10.7% for the PFCT, resulting in higher Q of the former.
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Q_Plain FIn Circular Tube Q_Plain FIn Oblique Tube Pressure Drop_Plain FIn Circular Tube Pressure Drop_Plain FIn Oblique Tube
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Figure 5: (a) Heat transfer amount (Q) and air-side pressure drop (Δp) of plain fin circular tube and plain fin oblique tube at inlet air velocity of 1 to 3m/s, (b) centre plane temperature contour and velocity streamline of plain fin circular tube and plain fin oblique tube at 2m/s
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Discussion on Thermal and Hydraulic Performances of CFOT Design The CFCT commonly applied for household AC is the baseline for comparison with the plain fin and CFOT designs. The thermal-hydraulic performance of all four FTHX designs is as shown in Fig. 6a. The corrugated fin is found to improve the Q of the circular tube by 1.3% to 6.7% at a higher Δp of 9.9% to 15.4%. Improvement in heat transfer performance with increase in pressure drop penalty is expected for CFCT as per other fin designs in the literatures. However, for the CFOT design, the Q is decreased slightly and pressure drop is increased up to 10.7% as compared to PFOT design. The corrugated fin design is detrimental to the overall performance of the coil with oblique tube. Referring to the cross-sectional plane velocity contour in Fig. 6b, air flow is redirected according to the contour of the corrugated fin, increasing the maximum air flow velocity by 5.1% and 5.8% for the CFCT and CFOT as compared to the plain fin designs respectively. The increase in air velocity is supposed to increase the overall heat transfer of the FTHX. More air flow separation zones are also present, of which 12.8% of the air flow region in the CFCT and 7.8% in CFOT design having air velocity lower than 0.5m/s. The combined effects of higher air flow velocity but smaller effective heat transfer area due to flow separation make the corrugated fin design beneficial for circular tube but detrimental for oblique tube. By comparing the four designs in Fig. 6a, PFOT has the best overall performance with about 7% higher heat transfer, 11.5% to 25.8% lower Δp than the baseline CFCT design in condenser coil of household AC. While the total Q and Δp are the practical parameters for comparison of different FTHX designs, the overall heat transfer coefficient, U which is normalised by total heat transfer area and LMTD, is an important parameter indicating the heat transfer rate of the design and is shown in Fig. 7a. U for the PFOT is about 12% higher than the CFCT. For the application of household AC, the heat transfer capacity of the design should be compared at the same fan power as shown in Fig. 7b. Compared with the CFCT used in the condenser coil of household AC and at the same fan power, the heat transfer capacity of the PFOT design is about 12% higher. Higher coil capacity achieved with the same fan power could reduce the refrigerant condensation temperature and pressure, resulting in lower compressor lift and power consumption, and higher COP of the AC eventually.
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Q_Plain Fin Circular Tube Q_Plain Fin Oblique Tube Q_Corrugated Fin Circular Tube Q_Corrugated Fin Oblique Tube Pressure Drop_Plain Fin Circular Tube Pressure Drop_Plain Fin Oblique Tube Pressure Drop_Corrugated Fin Circular Tube Pressure Drop_Corrugated Fin Oblique Tube
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Figure 6: (a) Heat transfer amount (Q) and air-side pressure drop (Δp) of plain fin circular tube, plain fin oblique tube, corrugated fin circular tube, and corrugated fin oblique tube at inlet air velocity of 1 to 3m/s, (b) cross-sectional plane velocity contour of the corrugated fin circular tube
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Figure 7: (a) Overall heat transfer coefficient (U) against inlet air velocity of 1 to 3m/s for four FTHX designs, (b) heat transfer amount (Q) against the theoretical fan power for four FTHX designs Choice of Materials and Manufacturability for PFOT Condenser Coil Typical FTHX for household AC is made of copper circular tube, aluminium fin, and joined by tube mechanical expansion technique. As the maximum von Mises stress for oblique tube design is more than the yield strength of copper, aluminium could be used for both the tube and fin to withstand the high refrigerant pressure. Tube mechanical expansion technique as well as controlled atmospheric brazing technique applied in making all-aluminium car radiator is feasible for manufacturing allaluminium PFOT condenser coil. In addition, brass oblique tube and copper fin is another material design option as the mechanical strength of brass is of similar order to aluminium, and there is brazing technique joining brass tube and copper fin, commonly used for high performance heat exchanger for cooling of diesel engine generator. The mechanical and thermal properties of copper, aluminium and brass for FTHX are tabulated in Table 4. The heat transfer performance of the PFOT design with Cu-tube-Al-fin, Al-tube-Al-fin, and Br-tube-Cu-fin is shown in Fig. 8. The influence of the tube material is negligible as seen from the results of Cu-tube-Al-fin and Al-tube-Al-fin cases due to small tube thickness. The use of copper fin improves the 9
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Q up to 6.3% as compared to Al-tube-Al-fin case, and up to 14% higher than the CFCT case. Al-tube-Al-fin design, which already has higher performance, lighter and cheaper in material cost, could be used to replace the existing condenser coil in household AC. When high performance heat exchanger is needed, Br-tube-Cu-fin design could be used. Table 4: Mechanical and Thermal Properties of Three Common Materials for FTHX Aluminium (Al)
Brass (Br)
8,978
2,719
8,580
381
871
380
387.6
202.4
111
70
241
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Figure 8: Heat transfer amount (Q) for plain fin oblique tube design with different choice of materials at inlet air velocity of 1 to 3m/s Parametric Optimisation of PFOT Condenser Coil Design
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Effect of Fin Thickness In the market of FTHX for household AC, common aluminium fin thickness used is 0.1mm and 0.15mm. Apparently, FTHX with 0.1mm thick fin will save about 50% fin material and about 50% lighter as compared to FTHX with 0.15mm fin. The downside of a thinner fin is that lower strength FTHX will be more susceptible to damage during handling and maintenance of the FTHX. The thermal performances of the PFOT design with 0.1mm and 0.15mm fin thickness are plotted at the same theoretical fan power as shown in Fig. 9. The overall fin domain is maintained at 1.4mm for both cases. As such, the design with 0.1mm fin thickness will have a fin-to-fin gap of 1.3mm, and for that with 0.15mm fin thickness, the fin-to-fin gap is 1.25mm. For the same inlet air velocity, a smaller fin gap leads to higher air flow velocity across the fin surface, resulting in higher Q and higher Δp. Nevertheless, the increase in skin friction due to higher air velocity is not significant, thus when compared at the same fan power, the Q of PFOT design with 0.15mm fin thickness is about 6% higher than the 0.1mm fin thickness case. A separate study is also conducted by keeping the same fin-to-fin gap for both the 0.1mm and 0.15mm fin designs. It is found that the difference in thermal and hydraulic performance is negligible due to similar air flow velocity across the fin surface and similar total heat transfer area for both designs. As the heat transfer improvement is not significant and the motivation in the industry to reduce material cost and weight of the FTHX [19-20], 0.1mm fin thickness for the PFOT are be adopted for subsequent parametric optimisation study.
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Figure 9: Heat transfer amount (Q) for plain fin oblique tube design with 0.1mm and 0.15mm fin thickness at the same theoretical fan power
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Effect of Fin Pitch Common fin per inch (FPI) for FTHX in the industry ranges from 15 to 21. The effect of FPI on the Q per inch of fin at the same inlet air velocity for the PFOT design is plotted in Fig. 10a. Higher FPI leads to higher heat transfer as there is more fin area for heat transfer. Referring to Fig. 10b, at the same fan power, the Q of a coil with higher FPI is significantly higher than a coil with a lower FPI but at a higher air flow velocity. For example, coil with 21 FPI at 2m/s consumes the same amount of fan power, but has over 30% higher Q per inch of fin than a coil with 15 FPI at 2.5m/s. This shows the effect of fin area per inch of coil is a more important factor in thermal performance of the coil than air flow velocity. For high performance coil, high FPI and small air flow velocity should be used to achieve high thermal performance without significant air-side pressure drop penalty. Nevertheless, the material cost and weight of the FTHX with high FPI will be higher too. There is no optimum FPI for PFOT design. FPI for PFOT design should be selected considering the thermal requirement of the applications as well as the material cost for the coil.
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Figure 10: (a) Heat transfer amount (Q) per inch of fin for plain fin oblique tube coil with 15 to 21 FPI at inlet air velocity of 1 to 3m/s, (b) Heat transfer amount (Q) per inch of fin for plain fin oblique tube coil with 15 to 21 FPI against fan power Effect of Tube Pitch PFOT with 18 FPI as per baseline coil is chosen for the study the effect of tube pitch on the thermal-hydraulic performance of the coil. The range of tube pitch is from 11.6mm to 22.1mm, which is equivalent to about 2.5 to 4.5 times the frontal length of the oblique tube. The thermal-hydraulic performance is calculated for per fin length of 527mm as per actual condenser coil size because each computational model is built with different tube pitch, thus different fin size and area. The results are shown in Fig. 11a and 11b. The effect of tube pitch on the total Q per fin length of 527mm is mostly minute as the total fin area is similar. Smaller tube pitch means there are more tubes per fin, but since the tube heat transfer area is small compared to the fin area, the drop in Q is gradual when the tube pitch increases as shown in Fig. 11a. When compared at the same fan power in Fig. 11b, an exponentially decreasing trend showing the flattening of Q per fin when the fan power increases is observed at 14mm and 16.3mm tube pitch, which are equivalent to 3 to 3.5 times of the oblique tube frontal length. Further increase in fan power input does not result in higher heat transfer performance. Condenser with 16.3mm tube pitch has fewer tubes than that of 14mm for similar amount of heat transfer. Thus the optimal tube pitch for oblique tube design is 16.3mm, which is equivalent to about 3.5 11
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times of the tube frontal length. This optimal tube pitch to tube frontal length ratio is actually close to the 3.2 ratio for the baseline corrugated fin circular tube design with a tube pitch of 22.1mm and tube frontal diameter of 6.82mm.
Heat Transfer Amount per Fin (W)
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Figure 11: (a) Heat transfer amount (Q) per fin of 527mm for plain fin oblique tube coil with 11.6 to 22.1mm tube pitch at inlet air velocity of 1 to 3m/s, (b) Heat transfer amount (Q) per fin of 527mm for plain fin oblique tube coil with 11.6 to 22.1mm tube pitch against fan power
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CONCLUSIONS In this paper, the characteristics of the PFOT and CFOT coil are studied and compared to the baseline circular tube coil. Choice of materials and parameter optimisations of the PFOT design are also made. The conclusions are summarized as follows: 1. The oblique tube has a smaller frontal area, resulting in smaller recirculation zone behind the tube and less abrupt change of air flow direction, thus increases the effective heat transfer area and heat transfer amount up to 14% and lower air-side pressure drop up to 14%. 2. The corrugated fin design is detrimental to the overall performance of the coil with oblique tube due to more induced air flow separation zones and disruption of the smooth air flow across the oblique tube. This results in lower heat transfer and higher airside pressure drop up to 10%. 3. Under steady-state operation condition, the effect of tube material on the coil thermal performance is insignificant. Copper fin is recommended for high performance heat exchanger. 4. The selection of fin thickness and FPI for PFOT coil depends on the heat dissipation requirement, the weight of the coil, and overall material cost constraint for the application. It is recommended to keep the ratio of the tube pitch to tube frontal length/diameter to 3 to 3.5 for maximum heat transfer at given air flow velocity.
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ACKNOWLEDGMENTS The authors would like to acknowledge the funding support from the National Research Foundation (NRF) Singapore under the Energy Innovation Research Programme (EIRP) Funding Scheme (NRF2013EWT-EIRP004-038) managed on behalf by the Building and Construction Authority (BCA) of Singapore.
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