Research Papers: Heat Exchangers

Numerical Study of Heat Transfer Enhancement of Roll-to-Roll Microchannel Heat Exchangers

[+] Author and Article Information
Heng Wang, Lakshmi Balasubramaniam, Samuel D. Marshall, Rerngchai Arayanarakool, Poh Seng Lee

Department of Mechanical Engineering,
National University of Singapore,
Singapore 119077

Xin Jin

School of Mechanical Engineering,
Beijing Institute of Technology,
Beijing 100081, China

Peter C. Y. Chen

Department of Mechanical Engineering,
National University of Singapore,
Singapore 119077
e-mail: mpechenp@nus.edu.sg

1Corresponding author.

Contributed by the Heat Transfer Division of ASME for publication in the JOURNAL OF HEAT TRANSFER. Manuscript received September 14, 2017; final manuscript received December 12, 2017; published online March 21, 2018. Assoc. Editor: Guihua Tang.

J. Heat Transfer 140(6), 061801 (Mar 21, 2018) (8 pages) Paper No: HT-17-1540; doi: 10.1115/1.4038910 History: Received September 14, 2017; Revised December 12, 2017

The heat transfer performance of two roll-to-roll microchannel heat exchangers with square cross section and side length ranging from 0.2 mm to 0.5 mm were investigated via numerical studies. In order to assess the heat transfer enhancement, equivalent straight channel heat exchangers were also researched numerically as comparisons. For the roll-to-roll devices, numerical studies demonstrated that there were two reasons for heat transfer enhancement. First, when the average Dean number of the fluid was greater than approximately 10, Dean vortices started to form within the roll-to-roll microchannels, enhancing the convective heat transfer between channels. Second, the compact roll-to-roll structure of the heat exchangers increased the area of heat transfer compared with straight microchannel equivalents, and thus promoted the conductive heat transfer. Numerical simulations noted both higher Nusselt numbers and higher thermal performance factors (TPF) for roll-to-roll microchannel heat exchangers compared with equivalent straight channels and were employed to optimize both the microchannel cross section dimensions and the wall thickness between channels. In addition, the swirling strength and the heat transfer area were also calculated to characterize the convective and conductive heat transfer, respectively, allowing for a comparison between two roll-to-roll microchannel heat exchanger designs.

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Grahic Jump Location
Fig. 1

Basic structures of two microchannel heat exchanger designs and two corresponding straight channel heat exchangers proposed in this study: (a) design A, (b) design B, (c) straight channel configuration of design A, and (d) straight channel configuration of design B. Arrows indicate the flow directions of cold and hot fluids.

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Fig. 2

Geometric depiction of the roll-to-roll microchannel heat exchanger designs, including the top view, section view, and detailed view of the microchannels and the whole heat exchangers: (a) design A and (b) design B

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Fig. 3

Influence of the average Dean number De¯ for design A when Ls = 0.3 mm and Tw = 0.2 mm on (a) the average Nusselt number Nu¯ and (b) the TPF. Dashed and solid lines refer to roll-to-roll and straight microchannel heat exchangers, respectively.

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Fig. 4

Tangential velocity flow patterns across the cross section of the upper microchannel of cold fluid for values of De¯ at 5, 10, 20, 60, and 100, where the line indicates the exact position of the cross section

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Fig. 5

The Nu¯ percentage increase (a) and the TPF percentage increase (b) for design A in comparison with an equivalent straight channel against Reynolds Number when Tw = 2 mm and Ls = 0.2 mm (dotted), 0.3 mm (dashed), and 0.4 mm (solid)

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Fig. 6

The Nu¯ percentage increase (a) and the TPF percentage increase (b) for design B in comparison with an equivalent straight channel against Reynolds number when Tw = 0.2 mm and Ls = 0.3 mm (dotted), 0.4 mm (dashed), and 0.5 mm (solid)

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Fig. 7

The average swirling strength across the cross section of the cold fluid for design A (a) when Ls = 0.2 mm (dotted), 0.3 mm (dashed), and 0.4 mm (solid) and design B (b) when Ls = 0.3 mm (dotted), 0.4 mm (dashed), and 0.5 mm (solid) against Reynolds number when Tw = 0.2 mm

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Fig. 8

The Nu¯ percentage increase ((a) for design A, (c) for design B) and the TPF percentage increase ((b) for design A, (d) for design B) in comparison with straight channels for both designs, when the square side length Ls = 0.3 mm for design A and 0.4 mm for design B. Dotted, dashed, and solid lines refer to wall thickness 0.1 mm, 0.2 mm, and 0.3 mm, respectively.

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Fig. 9

The average Nusselt Number percentage increase (a) and the TPF percentage increase (b) for design A and design B. Dashed and solid lines represent design A and design B, respectively. For design A, Tw = 0.2 mm and Ls = 0.3 mm. For design B, Tw = 0.2 mm and Ls = 0.4 mm.




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