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Polymeric hollow fiber heat exchanger as an automotive radiator
Applied Thermal Engineering 108 (2016) 798–803
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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Polymeric hollow fiber heat exchanger as an automotive radiator Ivo Krásny´ a,⇑, Ilya Astrouski b, Miroslav Raudensky´ b a b
Tomas Bata University in Zlin, Faculty of Technology, Inst. Foodstuff Technology, nam. T.G. Masaryka 275, Zlin, Czech Republic Brno University of Technology, Faculty of Mechanical Engineering, Heat Transfer and Fluid Flow Laboratory, Technicka 2896/2, 616 69 Brno, Czech Republic
h i g h l i g h t s Polymeric hollow fiber heat exchanger as an automotive radiator is proposed. The mechanism of heat transfer (HT) relies on diameter of polymeric hollow fiber. Grimson equation is sufficient for approximate prediction of the heat transfers.
a r t i c l e
i n f o
Article history: Received 1 April 2016 Revised 28 June 2016 Accepted 28 July 2016 Available online 29 July 2016 Keywords: Polymeric hollow fiber Heat exchanger Heat transfer Calorimeter
a b s t r a c t Nowadays, different automotive parts (tubing, covers, manifolds, etc.) are made of plastics because of their superior characteristics, low weight, chemical resistance, reasonable price and several other aspects. Manufacturing technologies are already well-established and the application of plastics is proven. Following this trend, the production of compact and light all-plastic radiators seems reasonable. Two plastic heat exchangers were manufactured based on polypropylene tubes of diameter 0.6 and 0.8 mm (so-called fibers) and tested. The heat transfer performance and pressure drops were studied with hot (60 °C) ethyleneglycol-water brine flowing inside the fibers and air (20 °C) outside because these conditions are conventional for car radiator operation. It was observed that heat transfer rates (up to 10.2 kW), overall heat transfer coefficients (up to 335 W/m2 K), and pressure drops are competitive to conventional aluminium finned-tube radiators. Moreover, influence of fiber diameter was studied. It was observed that air-side convective coefficients rise with a decrease of fiber diameter. Air-side pressure drops of plastic prototypes were slightly higher than of aluminium radiator but it is expected that additional optimization will eliminate this drawback. Experimentally obtained air-side heat transfer coefficients were compared with the theoretical prediction using the Grimson equation and the Churchill and Bernstein approach. It was found that the Grimson equation is sufficient for approximate prediction of the outer HTCs and can be used for engineering calculations. Further work will concentrate on optimizing and developing a polymeric hollow fiber heat exchanger with reduced size, weight and optimized performance and pressure drops. Ó 2016 Elsevier Ltd. All rights reserved.
Contents 0. 1. 2.
3. 4.
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Theoretical background . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Experimental details . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1. Samples of heat exchangers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.2. Experimental setup and thermal performance test procedure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Results and discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.1. Comparison of PHFHEs and conventional finned-tube heat exchangers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Conclusions. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Acknowledgements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
⇑ Corresponding author. E-mail addresses: [email protected] (I. Krásny´), [email protected] (I. Astrouski), [email protected] (M. Raudensky´). http://dx.doi.org/10.1016/j.applthermaleng.2016.07.181 1359-4311/Ó 2016 Elsevier Ltd. All rights reserved.
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Nomenclature A D Gz h L N Nu Pr Q R Re T
heat transfer area or area of cross-section, m2 diameter, m Graetz number, DRePr/L, dimensionless heat-transfer coefficient, W m2 K1 fiber length, m number of fibers, dimensionless Nusselt number, hD/k, dimensionless Prandtl number, Cpl/k, dimensionless rate of heat transfer, W thermal resistance, m2 K W1 Reynolds number, Du/l, dimensionless temperature, K
0. Introduction Polymeric material and their processing technology [1] have been of growing importance over the last decades. They are applicable across all branches including chemical [2], pharmaceutical [3], and automotive industries [4,5]. From the point of view of polymeric material properties, development of new polymeric material has led to modified polymers being used as alternative automotive materials [6]. These new materials should be more ecologically and economically competitive. For example, investigations of glass-fiber reinforced polyamide composites applicable in the automotive industry were reported by Rudzinski et al. [7], Njuguna et al. [8], Teixeira et al. [9], Thomason [10], Mouti et al. [11]. The polymeric hollow fiber heat exchanger [12] is a novel technology with the potential not only to significantly improve the current products but to enable entirely new applications and markets. The specific objectives leading to production are characterized as: design, testing and technology. This new approach to the enhanced functionality of polymeric fiber materials used in heat transfer surfaces already has the potential in its current development stage to be competitive in large segments of commercial heat exchangers concerning cost [13], weight, mass scalability, recyclability, resistance to corrosion [14] and low fouling. This technology by itself applied in the car-radiator market has the potential to save globally over one hundred thousand tons of aluminium annually. The main goal of this paper is to study thermal performance and compare it with existing aluminium radiators with the aim to develop completely new product on high innovative level. 1. Theoretical background An evaluation was conducted of potentially replacing metallic heat exchangers by considering thermal-hydraulic performance, mechanical strength, size, weight and material cost which was investigated by Park and Jacobi [13]. It was stated that a polymer-tube-bundle heat exchanger could be a potential replacement for a conventional metallic plate-fin-and-tube heat exchanger for water-to-air applications. The lower thermal conductivity of polymeric material was overcome by employing a larger number of thin-walled small-diameter tubes. The use of hollow metallic particles for thermal conductivity enhancement and lightening of a filled polymer was studied by Garnier and Boudenne [15]. As mentioned earlier [12], the overall concept of heat transfer is based on heat transfer features of micro-objects [16], where the intensity of convective heat flow through a given material grows with the
DT U
temperature difference, K overall heat-transfer coefficient, W m2
Subscripts f frontal i inner o outer ov overall t tube T3 convection boundary condition (the third type) w wall
square of the object’s size. However, the heat transfer, thermal stability, pressure limits [17] and toughness of a hollow polymeric material is still a major problem in the development sphere in the automotive industry. Thus, engineers are focusing their scientific interests to improve these properties. The basic parameter of convective heat transfer is the Nusselt number. It is viewed as the dimensionless convection heat transfer coefficient characterized as the ratio of convective to conductive heat transfer through the fluid layer [18]. Lamilar flow is characterized for similar magnitudes of convection and conduction. The Nusselt number is independent of fluid flow velocity and is close to 4 for a circular tube. Zarkadas and Sirkar [19] provided an analytical solution for laminar flow inside circular tubes with boundary conditions described by HTC and ambient temperature. The solution is based on the asymptotic Nusselt number calculated by Hickman’s approach and the incremental heat transfer number calculated by Hsu’s approach. The incremental heat transfer number evaluated as a function of the external resistance Nuw and the length of the thermal developing region expressed by the Graetz number Gz [19]. These values were used to construct a simple relationship for the mean Nusselt number of the tube, which is only a function of the external resistance and the hydrodynamic conditions prevailing in the tube [19]:
NuT3
0 1 48 þ Nu 0:06487 4Gz w A ¼ 11 59 þ @0:0499 Nuwþ5:37935 p 1 þ 220 Nuw 1 þ exp
ð1Þ
2:17887
Considering flow on the outer surface of the hollow fiber bunch, the Nusselt number for a given geometry can be expressed in terms of the Reynolds number and the Prandtl number. We decided to evaluate two models for predicting the outer HTC. The first model is for a single tube and the second is for a bunch of tubes with a cross-flow of air. The average Nusselt number for a cross-flow over a single tube can be evaluated using the equation by Churchill and Bernstein [20]:
" 5=8 #4=5 0:62Re1=2 Pr1=3 Re Nud ¼ 0:3 þ h i1=4 1 þ 282:000 1 þ ð0:4=PrÞ2=3
ð2Þ
For a bunch of tubes, depending on fiber width and depth pitches, the Grimson approach can be recommended [21]:
Nu ¼ CRem Pr 1=3
ð3Þ
Experimentally determined constants C and exponents m represent a geometric description of the tube-bundle arrangement
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2. Experimental details 2.1. Samples of heat exchangers Polymeric hollow fiber heat exchanger prototypes were prepared in the laboratory from polypropylene (PP) with thermal conductivity k = 0.18 W/m K [22] by extrusion. This material is suitable to withstand operation temperatures up to 80 °C [23]. These tiny fibers can withstand an inner pressure of 50 bar at room temperature without any damage. Extrusion technology can make fibers of much smaller outer diameters (about 0.2 mm). Selecting a smaller diameter results in increasing the number of shorter fibers which must be used for the same performance and pressure drop. It is in principal difficult to prepare a prototype for testing using thousands of tiny fibers. The main core is compounded with a total amount of 14 layers per 140 fibers each connected together to obtain the heat transfer capacity needed (Fig. 1). The fibers are spaced in the heat exchanger surface and regularly separated. Table 1 represents parameters of prepared modules: number of tubes/fibers N, fiber length L, outer and inner diameters (Do and Di), tube pitches in width and depth (b1 and b2), overall heat transfer area based on outer surface (Ao) and frontal flow area (Af).
Fig. 1. Polymeric hollow fiber heat exchanger.
Table 1 Geometrical characteristics of tested modules P3, P4. Module
N
L (mm)
Do (mm)
Di (mm)
b1 (mm)
b2 (mm)
Ao (m2)
Af (m2)
2.2. Experimental setup and thermal performance test procedure
P3 P4
1960 1904
250 220
0.6 0.8
0.48 0.64
1.8 1.8
2 2
0.92 1.05
0.060 0.053
Cooling performance tests and pressure drop tests have been performed on an industry certified calorimetric circuit RAIV having a measurement accuracy of 5% according to automotive standards. Characterization of polymeric hollow fiber heat properties was implemented in 50/50% water glycol coolant solution for coolant inlet temperatures 60 °C, 75 °C and 90 °C. Air was used as the outside cooling media. During normal operation, radiators work with air velocities approximately 2–10 m/s so a similar range of velocities was chosen for the tests. Tested heat exchanger is fixed (see Fig. 2, position 7). The coolant (50/50% water-glycol solution) (Fig. 2, position 8) is circulated in the fibers and transfers heat to the air flowing around the fibers, controlled by the air speed fan (Fig. 2, position 1). Outer air is driven by another air speed fan (Fig. 2, position 2). Initial inlet temperature of air can be set by electrical heating (Fig. 2, position 3) or external cooling. The heat transfer rate and pressure drops are determined for the coolant side and the air side based on the temperature change and pressure change. Actual inlet/outlet temperatures and heat rejection were recorded at a frequency of 1 Hz and were obtained by the system sensors (Fig. 2, position 5–6) for both sides. Mass flow rates were calculated with calorimeter software according to measured volumetric flow rate values of the liquid coolant media and the air side using flow meters.
Fig. 2. The schematic diagram of the test rig: 1- air speed fan, 2- air convector, 3- air inlet electrical heating, 4- air inlet cooling, 5- air mass flow measuring, 6- air temperature measuring, 7- heat exchanger, 8- coolant liquid medium.
Table 2 Summarized experimental data: air speed (m/s), mass flow m (kg/s), input and output temperatures Tin/out (°C), pressure drops dP (Pa), heat rejection Q (kW) for cooling air and liquid coolant sides. Air speed (m/s)
Cooling air side
Liquid coolant side
ms (kg/s)
Ts,in (°C)
Ts,out (°C)
dT,s (°C)
dP,s (Pa)
Qs (kW)
mt (kg/s)
Tt,in (°C)
Tt,out (°C)
dTt (°C)
dP,t (kPa)
Qt (kW)
0.6 mm 1.02 1.97 4.01 10.00
0.071 0.137 0.279 0.695
20.5 20.3 20.2 20.2
53.6 50.4 44.5 34.7
33.1 30.1 24.2 14.6
40 77 218 1108
2.35 4.13 6.78 10.2
0.290 0.289 0.290 0.290
60.2 60.3 60.0 59.7
56.5 55.5 53.2 49.3
3.7 4.8 6.9 10.4
50.81 51.20 52.36 55.64
3.70 4.86 6.89 10.4
0.8 mm 1.00 2.02 4.03 10.02
0.069 0.140 0.280 0.697
20.1 19.9 20.1 20.4
54.2 50.9 42.7 34.7
34.1 31.0 22.6 14.3
47 96 275 1387
2.37 4.36 6.37 10.0
0.289 0.289 0.289 0.289
60.1 60.4 59.9 60.2
56.8 55.9 53.6 50.3
3.3 4.5 6.4 9.9
39.5 39.8 40.5 42.7
3.30 4.55 6.38 9.88
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It was revealed in this study that the diameter of polymeric hollow fiber has significant influence on heat transfer coefficients on both the inside and outside surfaces of fibers. It can be stated that decreasing values of outer diameter results in higher values of heat transfer coefficients at the air side, which corresponds to our results summarized in Table 2. As given in Table 2, the increasing values of air speed in the frame of 1.0–10 m/s was reflected in increasing values of heat rejection in the frame of 2.37–10.01 kW for fibers with an outer diameter of 0.8 mm and of
600
W/m2 K
3. Results and discussion
801
500 400 300
ho
0.6mm
200
0.8mm
100 800
aluminium core
0
Experimental
0
2
4
Grimson
600
W/m2 K
6
10
8
Air velociy (m/s) Fig. 5. Outer (air-side) HTC vs. air velocity for an aluminium core of a car radiator and a core made of fiber fabric (PHFHE module).
Churchill and Bernstein
ho
400 Table 4 Comparison of the heat exchange core of a conventional aluminium car radiator and the fiber core of the PHFHE module.
200
Low temperature radiator made of PHFHE module, total fiber number is aluminium alloy 1960 (14 layers 140 each), outer/inner Core size is 290 292 25 mm, fiber diameter is 0.6/0.48 mm, length 1 5 mm fins on 1.5 25 mm 240 mm. Core size is flat tubes, 1.9 m2 surface 240 240 40 mm, 0.89 m2 surface Weight of empty core (without coolant and headers), full core weight and weight ratio 933 g, 1502 g, 100% 101 g, 396 g, 26%
0
0
100
200
300
400
Reair
600
Experimental
500
Grimson
0.8mm
Churchill and Bernstein
0.6mm
400
1500
Pressure drop (Pa)
ho
W/m2 K
Fig. 3. Dependence of outer (air-side) heat transfer coefficient on Air Reynolds number for a heat exchanger with an outer diameter of 0.6 mm.
300 200
1000
aluminium core
500
100 0
0
100
200
300
400
500
0
600
0
2
Reair
4
6
8
10
Cooling air input (m/s)
Fig. 4. Dependence of outer (air-side) heat transfer coefficient on Air Reynolds number for a heat exchanger with an outer diameter of 0.8 mm.
Fig. 6. Effect of cooling air input on pressure drops for an aluminium core of a car radiator and a core made of fiber fabric (PHFHE module).
Table 3 Heat exchanger heat performances with precisions according to the Student’s t-distribution. Module
Ret
Reair
Q (kW)
Uo (W/m2 K)
e
NTU
0.6 mm 0.8 mm
277.25 ± 3.04 209.5 ± 2.0
161.0 ± 45.2 215.5 ± 60.2
6.47 ± 0.85 6.03 ± 0.83
335 ± 25 228.5 ± 17.2
0.80 ± 0.12 0.73 ± 0.10
1.34 ± 0.26 1.23 ± 0.45
802
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Fig. 7. Detailed view of heat transfer surface of currently widely used heat transfer surface of car radiators (left) and PHFHE’s surface (right).
2.35–10.19 kW for fibers with an outer diameter of 0.6 mm respectively. Next, our experimental results will be examined against the theory presented above in this paper. Fig. 3 shows a graph of the outer (air-side) HTC versus air Reynolds number for the module with fibers 0.6 mm. A large discrepancy exists between the experimental data and the theoretical prediction by the Churchill and Bernstein model. This model gives significantly underestimated results for the whole range of Reynolds numbers. Agreement between the theoretical prediction by the Grimson model and experimental data is better but not good enough to be used in engineering calculations. Fig. 4 shows the same data for the module with 0.8 mm fibers. The data has the same tendency: values theoretically predicted by the Churchill and Bernstein model differ significantly from experimental data (underestimated). On the other hand, the prediction by the Grimson model is closer to experimentally obtained data. As reported in Table 3, the heat exchanger was characterized by a set of thermal performance property values: tube Re,t and air-side Reair, Reynolds numbers, heat transfer rates Q, overall heat transfer coefficients U0, heat exchanger effectiveness e and the number of transfer units (NTU). The values were averaged from results obtained from a set of 4 measurements. The results revealed higher values of heat transfer rates for the 0.6 mm tube module: 6.47 kW having a measurement accuracy of ±13.18% with an overall heat transfer coefficient of 335 W/m2 K ± 7%. This effect was also observed in our previous paper [24] confirming the fact that small tube devices reach high heat transfer coefficients. As reported by Oliet et al. [25] heat transfer and performance of a heat exchanger is affected by air and coolant mass flow rates. The effect of air velocity was also discussed. Fig. 5 shows heat transfer coefficient (HTC) dependence on air velocity. A clean linear relationship was found in the range from 1.0 to 4.03 m/s tested in our study. This conclusion was supported by outcomes of Bowang et al. [26]. 3.1. Comparison of PHFHEs and conventional finned-tube heat exchangers Tested PHFHE modules were developed to be used as liquidto-gas heat transfer surfaces in different systems. For example, they can be used to replace existing car radiator heat transfer surfaces. Fig. 7 shows the heat transfer surface of currently widely used heat transfer surfaces of car radiators and PHFHE. As standard car radiator is made of aluminium flat tubes with aluminium fins soldered to their surface. This design can be considered as very modern and high-efficient; in fact this technology is very close to the most effective existing solution of a liquid-togas heat transfer surface. Thus, we assume that if a hollow fiber surface will be competitive to such a structure – it will be thermally competitive to the majority of other existing metal liquid-to-gas heat transfer surfaces. Table 4 presents detailed
information about a car radiator and the PHFHE module. We used experimentally obtained results of pressure drops and OHTCs of the PHFHE module to calculate surface-to-air HTC. To estimate parameters of the aluminium core of conventional radiators we used thermal performance data of low temperature radiators (automotive part of Audi Q7). Figs. 5 and 6 present graphs of pressure drops and air-side HTCs ho of both heat exchanger cores depending on air velocity. We can conclude that the core made of hollow fibers is very competitive from the point of view of heat transfer. The pressure drops are slightly higher for a fiber core but we expect that this drawback can be eliminated by additional optimization of selecting fiber diameter, length and number. An interesting comparison is given in Table 4 where weight is compared for metal and plastic heat transfer surfaces. The plastic heat transfer surface (manifold is not considered) is four times lighter than the aluminium product. 4. Conclusions Prototypes of polymeric hollow fiber heat exchangers with fiber outer diameters of 0.6 mm and 0.8 mm were studied. It was found that these devices are able to achieve thermal performances comparable to metal finned tube heat exchangers which can be successfully used as a replacement for metal exchangers in the automotive industry. It was found that these devices can achieve high values of overall heat transfer coefficients (up to 335 W/m2 K) and efficiencies (up to 0.8). It can be concluded that the low thermal conductivity of polypropylene is not an obstacle to using hollow fibers for gas-to-liquid heat transfer. Moreover, it was found that air-side convective coefficients rise with a decrease of tube (fiber) diameter. Heat transferred by modules made of 0.6 mm and 0.8 mm fibers was similar (up to 10.4 kW) but for smaller fibers this heat was transferred using a smaller surface. Air-side pressure drop characteristics of both examples were similar too. The heat exchanger made of 0.8 mm fibers had an approximately 20% higher pressure drop. Experimental results were compared with theoretical prediction by the Grimson equation and the Churchill and Bernstein approach. It was found that the Grimson equation is sufficient for predicting outer HTC and can be used for engineering calculations. Furthermore, new concepts of polymeric hollow fiber heat exchangers need to be designed and developed effectively in order to increase the longevity, chemical stability and mechanical properties. Acknowledgements The theoretical work leading to these results has received funding from the MEYS under the National Sustainability Programme I (Project LO1202).
´ et al. / Applied Thermal Engineering 108 (2016) 798–803 I. Krásny
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Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Polymeric hollow fiber heat exchanger as an automotive radiator Ivo Krásny´ a,⇑, Ilya Astrouski b, Miroslav Raudensky´ b a b
Tomas Bata University in Zlin, Faculty of Technology, Inst. Foodstuff Technology, nam. T.G. Masaryka 275, Zlin, Czech Republic Brno University of Technology, Faculty of Mechanical Engineering, Heat Transfer and Fluid Flow Laboratory, Technicka 2896/2, 616 69 Brno, Czech Republic
h i g h l i g h t s Polymeric hollow fiber heat exchanger as an automotive radiator is proposed. The mechanism of heat transfer (HT) relies on diameter of polymeric hollow fiber. Grimson equation is sufficient for approximate prediction of the heat transfers.
a r t i c l e
i n f o
Article history: Received 1 April 2016 Revised 28 June 2016 Accepted 28 July 2016 Available online 29 July 2016 Keywords: Polymeric hollow fiber Heat exchanger Heat transfer Calorimeter
a b s t r a c t Nowadays, different automotive parts (tubing, covers, manifolds, etc.) are made of plastics because of their superior characteristics, low weight, chemical resistance, reasonable price and several other aspects. Manufacturing technologies are already well-established and the application of plastics is proven. Following this trend, the production of compact and light all-plastic radiators seems reasonable. Two plastic heat exchangers were manufactured based on polypropylene tubes of diameter 0.6 and 0.8 mm (so-called fibers) and tested. The heat transfer performance and pressure drops were studied with hot (60 °C) ethyleneglycol-water brine flowing inside the fibers and air (20 °C) outside because these conditions are conventional for car radiator operation. It was observed that heat transfer rates (up to 10.2 kW), overall heat transfer coefficients (up to 335 W/m2 K), and pressure drops are competitive to conventional aluminium finned-tube radiators. Moreover, influence of fiber diameter was studied. It was observed that air-side convective coefficients rise with a decrease of fiber diameter. Air-side pressure drops of plastic prototypes were slightly higher than of aluminium radiator but it is expected that additional optimization will eliminate this drawback. Experimentally obtained air-side heat transfer coefficients were compared with the theoretical prediction using the Grimson equation and the Churchill and Bernstein approach. It was found that the Grimson equation is sufficient for approximate prediction of the outer HTCs and can be used for engineering calculations. Further work will concentrate on optimizing and developing a polymeric hollow fiber heat exchanger with reduced size, weight and optimized performance and pressure drops. Ó 2016 Elsevier Ltd. All rights reserved.
Contents 0. 1. 2.
3. 4.
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Theoretical background . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Experimental details . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1. Samples of heat exchangers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.2. Experimental setup and thermal performance test procedure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Results and discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.1. Comparison of PHFHEs and conventional finned-tube heat exchangers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Conclusions. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Acknowledgements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
⇑ Corresponding author. E-mail addresses: [email protected] (I. Krásny´), [email protected] (I. Astrouski), [email protected] (M. Raudensky´). http://dx.doi.org/10.1016/j.applthermaleng.2016.07.181 1359-4311/Ó 2016 Elsevier Ltd. All rights reserved.
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Nomenclature A D Gz h L N Nu Pr Q R Re T
heat transfer area or area of cross-section, m2 diameter, m Graetz number, DRePr/L, dimensionless heat-transfer coefficient, W m2 K1 fiber length, m number of fibers, dimensionless Nusselt number, hD/k, dimensionless Prandtl number, Cpl/k, dimensionless rate of heat transfer, W thermal resistance, m2 K W1 Reynolds number, Du/l, dimensionless temperature, K
0. Introduction Polymeric material and their processing technology [1] have been of growing importance over the last decades. They are applicable across all branches including chemical [2], pharmaceutical [3], and automotive industries [4,5]. From the point of view of polymeric material properties, development of new polymeric material has led to modified polymers being used as alternative automotive materials [6]. These new materials should be more ecologically and economically competitive. For example, investigations of glass-fiber reinforced polyamide composites applicable in the automotive industry were reported by Rudzinski et al. [7], Njuguna et al. [8], Teixeira et al. [9], Thomason [10], Mouti et al. [11]. The polymeric hollow fiber heat exchanger [12] is a novel technology with the potential not only to significantly improve the current products but to enable entirely new applications and markets. The specific objectives leading to production are characterized as: design, testing and technology. This new approach to the enhanced functionality of polymeric fiber materials used in heat transfer surfaces already has the potential in its current development stage to be competitive in large segments of commercial heat exchangers concerning cost [13], weight, mass scalability, recyclability, resistance to corrosion [14] and low fouling. This technology by itself applied in the car-radiator market has the potential to save globally over one hundred thousand tons of aluminium annually. The main goal of this paper is to study thermal performance and compare it with existing aluminium radiators with the aim to develop completely new product on high innovative level. 1. Theoretical background An evaluation was conducted of potentially replacing metallic heat exchangers by considering thermal-hydraulic performance, mechanical strength, size, weight and material cost which was investigated by Park and Jacobi [13]. It was stated that a polymer-tube-bundle heat exchanger could be a potential replacement for a conventional metallic plate-fin-and-tube heat exchanger for water-to-air applications. The lower thermal conductivity of polymeric material was overcome by employing a larger number of thin-walled small-diameter tubes. The use of hollow metallic particles for thermal conductivity enhancement and lightening of a filled polymer was studied by Garnier and Boudenne [15]. As mentioned earlier [12], the overall concept of heat transfer is based on heat transfer features of micro-objects [16], where the intensity of convective heat flow through a given material grows with the
DT U
temperature difference, K overall heat-transfer coefficient, W m2
Subscripts f frontal i inner o outer ov overall t tube T3 convection boundary condition (the third type) w wall
square of the object’s size. However, the heat transfer, thermal stability, pressure limits [17] and toughness of a hollow polymeric material is still a major problem in the development sphere in the automotive industry. Thus, engineers are focusing their scientific interests to improve these properties. The basic parameter of convective heat transfer is the Nusselt number. It is viewed as the dimensionless convection heat transfer coefficient characterized as the ratio of convective to conductive heat transfer through the fluid layer [18]. Lamilar flow is characterized for similar magnitudes of convection and conduction. The Nusselt number is independent of fluid flow velocity and is close to 4 for a circular tube. Zarkadas and Sirkar [19] provided an analytical solution for laminar flow inside circular tubes with boundary conditions described by HTC and ambient temperature. The solution is based on the asymptotic Nusselt number calculated by Hickman’s approach and the incremental heat transfer number calculated by Hsu’s approach. The incremental heat transfer number evaluated as a function of the external resistance Nuw and the length of the thermal developing region expressed by the Graetz number Gz [19]. These values were used to construct a simple relationship for the mean Nusselt number of the tube, which is only a function of the external resistance and the hydrodynamic conditions prevailing in the tube [19]:
NuT3
0 1 48 þ Nu 0:06487 4Gz w A ¼ 11 59 þ @0:0499 Nuwþ5:37935 p 1 þ 220 Nuw 1 þ exp
ð1Þ
2:17887
Considering flow on the outer surface of the hollow fiber bunch, the Nusselt number for a given geometry can be expressed in terms of the Reynolds number and the Prandtl number. We decided to evaluate two models for predicting the outer HTC. The first model is for a single tube and the second is for a bunch of tubes with a cross-flow of air. The average Nusselt number for a cross-flow over a single tube can be evaluated using the equation by Churchill and Bernstein [20]:
" 5=8 #4=5 0:62Re1=2 Pr1=3 Re Nud ¼ 0:3 þ h i1=4 1 þ 282:000 1 þ ð0:4=PrÞ2=3
ð2Þ
For a bunch of tubes, depending on fiber width and depth pitches, the Grimson approach can be recommended [21]:
Nu ¼ CRem Pr 1=3
ð3Þ
Experimentally determined constants C and exponents m represent a geometric description of the tube-bundle arrangement
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2. Experimental details 2.1. Samples of heat exchangers Polymeric hollow fiber heat exchanger prototypes were prepared in the laboratory from polypropylene (PP) with thermal conductivity k = 0.18 W/m K [22] by extrusion. This material is suitable to withstand operation temperatures up to 80 °C [23]. These tiny fibers can withstand an inner pressure of 50 bar at room temperature without any damage. Extrusion technology can make fibers of much smaller outer diameters (about 0.2 mm). Selecting a smaller diameter results in increasing the number of shorter fibers which must be used for the same performance and pressure drop. It is in principal difficult to prepare a prototype for testing using thousands of tiny fibers. The main core is compounded with a total amount of 14 layers per 140 fibers each connected together to obtain the heat transfer capacity needed (Fig. 1). The fibers are spaced in the heat exchanger surface and regularly separated. Table 1 represents parameters of prepared modules: number of tubes/fibers N, fiber length L, outer and inner diameters (Do and Di), tube pitches in width and depth (b1 and b2), overall heat transfer area based on outer surface (Ao) and frontal flow area (Af).
Fig. 1. Polymeric hollow fiber heat exchanger.
Table 1 Geometrical characteristics of tested modules P3, P4. Module
N
L (mm)
Do (mm)
Di (mm)
b1 (mm)
b2 (mm)
Ao (m2)
Af (m2)
2.2. Experimental setup and thermal performance test procedure
P3 P4
1960 1904
250 220
0.6 0.8
0.48 0.64
1.8 1.8
2 2
0.92 1.05
0.060 0.053
Cooling performance tests and pressure drop tests have been performed on an industry certified calorimetric circuit RAIV having a measurement accuracy of 5% according to automotive standards. Characterization of polymeric hollow fiber heat properties was implemented in 50/50% water glycol coolant solution for coolant inlet temperatures 60 °C, 75 °C and 90 °C. Air was used as the outside cooling media. During normal operation, radiators work with air velocities approximately 2–10 m/s so a similar range of velocities was chosen for the tests. Tested heat exchanger is fixed (see Fig. 2, position 7). The coolant (50/50% water-glycol solution) (Fig. 2, position 8) is circulated in the fibers and transfers heat to the air flowing around the fibers, controlled by the air speed fan (Fig. 2, position 1). Outer air is driven by another air speed fan (Fig. 2, position 2). Initial inlet temperature of air can be set by electrical heating (Fig. 2, position 3) or external cooling. The heat transfer rate and pressure drops are determined for the coolant side and the air side based on the temperature change and pressure change. Actual inlet/outlet temperatures and heat rejection were recorded at a frequency of 1 Hz and were obtained by the system sensors (Fig. 2, position 5–6) for both sides. Mass flow rates were calculated with calorimeter software according to measured volumetric flow rate values of the liquid coolant media and the air side using flow meters.
Fig. 2. The schematic diagram of the test rig: 1- air speed fan, 2- air convector, 3- air inlet electrical heating, 4- air inlet cooling, 5- air mass flow measuring, 6- air temperature measuring, 7- heat exchanger, 8- coolant liquid medium.
Table 2 Summarized experimental data: air speed (m/s), mass flow m (kg/s), input and output temperatures Tin/out (°C), pressure drops dP (Pa), heat rejection Q (kW) for cooling air and liquid coolant sides. Air speed (m/s)
Cooling air side
Liquid coolant side
ms (kg/s)
Ts,in (°C)
Ts,out (°C)
dT,s (°C)
dP,s (Pa)
Qs (kW)
mt (kg/s)
Tt,in (°C)
Tt,out (°C)
dTt (°C)
dP,t (kPa)
Qt (kW)
0.6 mm 1.02 1.97 4.01 10.00
0.071 0.137 0.279 0.695
20.5 20.3 20.2 20.2
53.6 50.4 44.5 34.7
33.1 30.1 24.2 14.6
40 77 218 1108
2.35 4.13 6.78 10.2
0.290 0.289 0.290 0.290
60.2 60.3 60.0 59.7
56.5 55.5 53.2 49.3
3.7 4.8 6.9 10.4
50.81 51.20 52.36 55.64
3.70 4.86 6.89 10.4
0.8 mm 1.00 2.02 4.03 10.02
0.069 0.140 0.280 0.697
20.1 19.9 20.1 20.4
54.2 50.9 42.7 34.7
34.1 31.0 22.6 14.3
47 96 275 1387
2.37 4.36 6.37 10.0
0.289 0.289 0.289 0.289
60.1 60.4 59.9 60.2
56.8 55.9 53.6 50.3
3.3 4.5 6.4 9.9
39.5 39.8 40.5 42.7
3.30 4.55 6.38 9.88
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It was revealed in this study that the diameter of polymeric hollow fiber has significant influence on heat transfer coefficients on both the inside and outside surfaces of fibers. It can be stated that decreasing values of outer diameter results in higher values of heat transfer coefficients at the air side, which corresponds to our results summarized in Table 2. As given in Table 2, the increasing values of air speed in the frame of 1.0–10 m/s was reflected in increasing values of heat rejection in the frame of 2.37–10.01 kW for fibers with an outer diameter of 0.8 mm and of
600
W/m2 K
3. Results and discussion
801
500 400 300
ho
0.6mm
200
0.8mm
100 800
aluminium core
0
Experimental
0
2
4
Grimson
600
W/m2 K
6
10
8
Air velociy (m/s) Fig. 5. Outer (air-side) HTC vs. air velocity for an aluminium core of a car radiator and a core made of fiber fabric (PHFHE module).
Churchill and Bernstein
ho
400 Table 4 Comparison of the heat exchange core of a conventional aluminium car radiator and the fiber core of the PHFHE module.
200
Low temperature radiator made of PHFHE module, total fiber number is aluminium alloy 1960 (14 layers 140 each), outer/inner Core size is 290 292 25 mm, fiber diameter is 0.6/0.48 mm, length 1 5 mm fins on 1.5 25 mm 240 mm. Core size is flat tubes, 1.9 m2 surface 240 240 40 mm, 0.89 m2 surface Weight of empty core (without coolant and headers), full core weight and weight ratio 933 g, 1502 g, 100% 101 g, 396 g, 26%
0
0
100
200
300
400
Reair
600
Experimental
500
Grimson
0.8mm
Churchill and Bernstein
0.6mm
400
1500
Pressure drop (Pa)
ho
W/m2 K
Fig. 3. Dependence of outer (air-side) heat transfer coefficient on Air Reynolds number for a heat exchanger with an outer diameter of 0.6 mm.
300 200
1000
aluminium core
500
100 0
0
100
200
300
400
500
0
600
0
2
Reair
4
6
8
10
Cooling air input (m/s)
Fig. 4. Dependence of outer (air-side) heat transfer coefficient on Air Reynolds number for a heat exchanger with an outer diameter of 0.8 mm.
Fig. 6. Effect of cooling air input on pressure drops for an aluminium core of a car radiator and a core made of fiber fabric (PHFHE module).
Table 3 Heat exchanger heat performances with precisions according to the Student’s t-distribution. Module
Ret
Reair
Q (kW)
Uo (W/m2 K)
e
NTU
0.6 mm 0.8 mm
277.25 ± 3.04 209.5 ± 2.0
161.0 ± 45.2 215.5 ± 60.2
6.47 ± 0.85 6.03 ± 0.83
335 ± 25 228.5 ± 17.2
0.80 ± 0.12 0.73 ± 0.10
1.34 ± 0.26 1.23 ± 0.45
802
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Fig. 7. Detailed view of heat transfer surface of currently widely used heat transfer surface of car radiators (left) and PHFHE’s surface (right).
2.35–10.19 kW for fibers with an outer diameter of 0.6 mm respectively. Next, our experimental results will be examined against the theory presented above in this paper. Fig. 3 shows a graph of the outer (air-side) HTC versus air Reynolds number for the module with fibers 0.6 mm. A large discrepancy exists between the experimental data and the theoretical prediction by the Churchill and Bernstein model. This model gives significantly underestimated results for the whole range of Reynolds numbers. Agreement between the theoretical prediction by the Grimson model and experimental data is better but not good enough to be used in engineering calculations. Fig. 4 shows the same data for the module with 0.8 mm fibers. The data has the same tendency: values theoretically predicted by the Churchill and Bernstein model differ significantly from experimental data (underestimated). On the other hand, the prediction by the Grimson model is closer to experimentally obtained data. As reported in Table 3, the heat exchanger was characterized by a set of thermal performance property values: tube Re,t and air-side Reair, Reynolds numbers, heat transfer rates Q, overall heat transfer coefficients U0, heat exchanger effectiveness e and the number of transfer units (NTU). The values were averaged from results obtained from a set of 4 measurements. The results revealed higher values of heat transfer rates for the 0.6 mm tube module: 6.47 kW having a measurement accuracy of ±13.18% with an overall heat transfer coefficient of 335 W/m2 K ± 7%. This effect was also observed in our previous paper [24] confirming the fact that small tube devices reach high heat transfer coefficients. As reported by Oliet et al. [25] heat transfer and performance of a heat exchanger is affected by air and coolant mass flow rates. The effect of air velocity was also discussed. Fig. 5 shows heat transfer coefficient (HTC) dependence on air velocity. A clean linear relationship was found in the range from 1.0 to 4.03 m/s tested in our study. This conclusion was supported by outcomes of Bowang et al. [26]. 3.1. Comparison of PHFHEs and conventional finned-tube heat exchangers Tested PHFHE modules were developed to be used as liquidto-gas heat transfer surfaces in different systems. For example, they can be used to replace existing car radiator heat transfer surfaces. Fig. 7 shows the heat transfer surface of currently widely used heat transfer surfaces of car radiators and PHFHE. As standard car radiator is made of aluminium flat tubes with aluminium fins soldered to their surface. This design can be considered as very modern and high-efficient; in fact this technology is very close to the most effective existing solution of a liquid-togas heat transfer surface. Thus, we assume that if a hollow fiber surface will be competitive to such a structure – it will be thermally competitive to the majority of other existing metal liquid-to-gas heat transfer surfaces. Table 4 presents detailed
information about a car radiator and the PHFHE module. We used experimentally obtained results of pressure drops and OHTCs of the PHFHE module to calculate surface-to-air HTC. To estimate parameters of the aluminium core of conventional radiators we used thermal performance data of low temperature radiators (automotive part of Audi Q7). Figs. 5 and 6 present graphs of pressure drops and air-side HTCs ho of both heat exchanger cores depending on air velocity. We can conclude that the core made of hollow fibers is very competitive from the point of view of heat transfer. The pressure drops are slightly higher for a fiber core but we expect that this drawback can be eliminated by additional optimization of selecting fiber diameter, length and number. An interesting comparison is given in Table 4 where weight is compared for metal and plastic heat transfer surfaces. The plastic heat transfer surface (manifold is not considered) is four times lighter than the aluminium product. 4. Conclusions Prototypes of polymeric hollow fiber heat exchangers with fiber outer diameters of 0.6 mm and 0.8 mm were studied. It was found that these devices are able to achieve thermal performances comparable to metal finned tube heat exchangers which can be successfully used as a replacement for metal exchangers in the automotive industry. It was found that these devices can achieve high values of overall heat transfer coefficients (up to 335 W/m2 K) and efficiencies (up to 0.8). It can be concluded that the low thermal conductivity of polypropylene is not an obstacle to using hollow fibers for gas-to-liquid heat transfer. Moreover, it was found that air-side convective coefficients rise with a decrease of tube (fiber) diameter. Heat transferred by modules made of 0.6 mm and 0.8 mm fibers was similar (up to 10.4 kW) but for smaller fibers this heat was transferred using a smaller surface. Air-side pressure drop characteristics of both examples were similar too. The heat exchanger made of 0.8 mm fibers had an approximately 20% higher pressure drop. Experimental results were compared with theoretical prediction by the Grimson equation and the Churchill and Bernstein approach. It was found that the Grimson equation is sufficient for predicting outer HTC and can be used for engineering calculations. Furthermore, new concepts of polymeric hollow fiber heat exchangers need to be designed and developed effectively in order to increase the longevity, chemical stability and mechanical properties. Acknowledgements The theoretical work leading to these results has received funding from the MEYS under the National Sustainability Programme I (Project LO1202).
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