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Research Papers: Natural and Mixed Convection

An Experimental Study of Forced Heat Convection in Concentric and Eccentric Annular Channels

[+] Author and Article Information
N. Kline

Department of Mechanical Engineering,
University of Ottawa,
Ottawa, ON K1N 6N5, Canada
e-mail: nklin013@uotttawa.ca

S. Tavoularis

Professor
Department of Mechanical Engineering,
University of Ottawa,
Ottawa, ON K1N 6N5, Canada
e-mail: stavros.tavoularis@uottawa.ca

Contributed by the Heat Transfer Division of ASME for publication in the JOURNAL OF HEAT TRANSFER. Manuscript received July 8, 2014; final manuscript received July 21, 2015; published online August 18, 2015. Assoc. Editor: Gongnan Xie.

J. Heat Transfer 138(1), 012502 (Aug 18, 2015) (11 pages) Paper No: HT-14-1452; doi: 10.1115/1.4031160 History: Received July 08, 2014; Revised July 21, 2015

An experimental study of the effect of eccentricity on forced convective heat transfer was conducted for upward flows in vertical, open-ended annular channels with a diameter ratio of 0.61, a length to outer diameter ratio of 18:1, and both internal surfaces heated uniformly. Flows with Reynolds numbers Re = 5450, 10,000, and 27,500 and eccentricities varying from 0 to 0.9 were examined. These results are deemed to be mostly in the forced convection regime with some possible overlap with the mixed convection regime at the lowest Reynolds number considered. This work complements our previous work on natural and mixed convection using the same facility. The effect of eccentricity was not significant at lower eccentricities, but, in highly eccentric cases, the wall temperature in the narrow gap was much higher than in the wide gap and the average heat transfer coefficient was as low as one-fifth of the concentric value. For Re > 10,000, the average Nusselt number for the concentric case was nearly four times higher than the value predicted by the Dittus–Boelter correlation.

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References

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Figures

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Fig. 1

Diagram of the experimental apparatus

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Fig. 2

Sketch of annular duct cross section showing positions of thermocouples and foil gaps

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Fig. 3

Variation of the local wall temperature rise along the annulus for different eccentricities; ○, ⋄, and □ represent measurements from thermocouples S0i, S90i, and S180i, respectively, and +, ×,and * represent measurements from thermocouples S0o, S90o, and S180o, respectively; Re = 27,500

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Fig. 4

Azimuthal variations of the temperature rise for the inner and outer cylinders at z/H = 0.5; e = 0 (*), 0.1 (□), 0.3 ( × ), 0.5 (+), 0.7 (⋄), 0.8 (○), and 0.9 (Δ); Re = 27,500

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Fig. 5

Variations of the azimuthally averaged temperature rise along the annulus for various eccentricities and Reynolds numbers (symbol definitions as in Fig. 4)

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Fig. 6

Circumferentially averaged wall temperature rise at z/H = 0.5; present results for Re = 5450 (□), 10,000 (Δ) and 27,500 (⋄); mixed convection results by MCT for Re = 5700 (○), Re = 2800 (*), and 1500 (+); and natural convection results by CT for Re = 900 (×)

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Fig. 7

Ratio of Richardson number to the Richardson number for natural convection at the same eccentricity for e = 0 (○) and e = 0.9 (Δ) at midheight; black symbols correspond to present results, whereas gray symbols correspond to MCT results

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Fig. 8

Application of the Jackson and Hall criterion (dashed line) to evaluate the significance of buoyancy forces on the heat transfer coefficient for e = 0 (○) and e = 0.9 (Δ) at midheight; black symbols correspond to present results, gray symbols correspond to measurements by MCT, and open symbols correspond to natural convection measurements by CT

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Fig. 9

Azimuthal variation of the local Nusselt number normalized by Nu0  deg,con=258 for the inner cylinder and by Nu0  deg,con=273 for the outer cylinder; z/H = 0.5; e = 0 (*), 0.1 (□), 0.3 ( × ), 0.5 (+), 0.7 (⋄), 0.8 (○) and 0.9 (Δ); Re = 27,500; uncertainty bars have been shown for the e = 0 and 0.9 cases; these were the maximum and minimum uncertainties, respectively

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Fig. 10

Azimuthally averaged Nusselt number versus eccentricity at z/H = 0.5; Re = 27,500 (⋄), 10,000 (Δ), 5450 (□), 5700 (○; from MCT), 2800 (*; from MCT), 1500 ( + ; from MCT), and 900 (×; from CT); uncertainty bars have been included for the concentric cases as representative for all cases

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Fig. 11

Variation of the average Nusselt number with Reynolds number at z/H = 0.5; black symbols correspond to present results, gray symbols correspond to MCT results, and open symbols correspond to natural convection results by CT; circles denote concentric cases and triangles denote highly eccentric cases (e = 0.9); solid lines are fitted to present, MCT, and CT data and have power law sections at large Reynolds numbers; dashed line is the Dittus–Boelter correlation

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Fig. 12

Comparison of the average Nusselt number at z/H = 0.5 for the present concentric case with correlations from literature for annular channels with the present diameter ratio; black circles indicate the present concentric results; the thin solid line is fitted to the present results; the dashed line is the Dittus–Boelter correlation; the dash–dotted line is the Davis correlation; the dotted line is the Dirker et al. correlation (inner cylinder heated); and the thick solid line is the Gnielinski correlation (inner cylinder heated)

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