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Research Papers: Multiphase Flows

Model of Flow in the Side Chambers of an Industrial Centrifugal Pump for Delivering Viscous Oil

[+] Author and Article Information
Wenguang Li

Department of Fluid Machinery,
Lanzhou University of Technology,
287 Langongping Road,
730050 Lanzhou, Gansu, PRC

Manuscript received March 29, 2012; final manuscript received February 12, 2013; published online March 29, 2013. Assoc. Editor: Chunill Hah.

J. Fluids Eng 135(5), 051201 (Mar 29, 2013) (22 pages) Paper No: FE-12-1161; doi: 10.1115/1.4023664 History: Received March 29, 2012; Revised February 12, 2013

A series of experiments has been conducted to identify the effects of both fluid viscosity and wear-rings gap on the performance of a low specific speed industrial centrifugal pump of type 65Y60 for transporting viscous oils by the author group. Unfortunately, the experimental results remained unexplained on a fluid dynamics base. To remedy this problem, a highly viscous oil flow model and computational method in the side chambers in that pump were proposed based on the existing theoretical and experimental results. The flow coupling between the chambers and the gaps of the wear-rings and/or the balance holes was realized. The model was validated by making use of the existing experiment data in the chamber between a rotating disk and the walls of a stationary cylindrical container. Then the flow model was applied into the two side chambers in that pump when the wear-rings clearances and liquid viscosity were changed. The results demonstrated that the flow model is sensitive to wear-rings gap, liquid viscosity, the roughness of the wet walls of the chambers, and leakage flow rate. For this pump, an enlarged clearance can improve the mechanical efficiency, but the increment in the efficiency is unable to compensate for the considerable drop in the volumetric efficiency, causing the gross efficiency not be improved, especially at a high viscosity.

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Figures

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Fig. 1

Cross-sectional drawing of industrial centrifugal pump of 65Y60 (a), and detailed wear-rings structure (b)

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Fig. 2

Simplified structure of impeller, volute, and wear-rings (a), control volume (b), control volume enlarged and developed in circumferential direction (c), and the volume zoomed in and projected to meridional plane (d)

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Fig. 3

Flow resistance coefficient across balance holes against corresponding experimental data

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Fig. 4

Comparison of local skin friction coefficient among a variety of formulas for boundary layer flow over a flat plate with semi-infinite length, 1-lamina flow, 2-transtion range from laminar to turbulent one, 3-hydraulic smooth range, 4-curve plotted by using the empirical formula in Ref. [56], 5-Nikuradse's curve for fully rough regime

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Fig. 5

Dimensionless fluid rotational angular velocity kf and pressure coefficient Cpf profile along radius in side chamber of enclosed rotating disk; (a) and (b) example 1, experimental data presented in Ref. [31], (c)–(f) example 2, observed data shown in case A of Ref. [18], dashed line denotes the pressure coefficient estimated by using condition kf = kf2

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Fig. 6

Effect of roughness of wet walls of rotating disk and cylindrical container on dimensionless fluid rotational angular velocity

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Fig. 7

Effect of radial velocity Vrf on pressure coefficient profile in example 2 under qf = −2.09 m3/h and tf = 7.64 mm conditions

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Fig. 8

Comparison of predicted pump gross efficiency with experimental one in terms of impeller Reynolds number for case A, B, and C at BEP

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Fig. 9

Predicted pump hydraulic (a)-(c), volumetric (d)-(f), and mechanical (g)-(i) efficiencies in terms of liquid viscosity at BEP for case A, B, and C

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Fig. 10

Leakage flow rate through wear-rings clearance as a function of viscosity for case A, B, and C

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Fig. 11

Total flow resistance coefficient and equivalent flow resistance coefficient in terms of liquid viscosity for case B

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Fig. 12

Dimensionless rotational angular velocity profiles radius in two side chambers at three viscosities 1, 93, and 237 mm2/s for case A; (a)–(c) front side chamber and (d)–(f) rear side chamber

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Fig. 13

Dimensionless rotational angular velocity profiles in two side chambers at three viscosities 1, 93, and 237 mm2/s for case B; (a)–(c) front side chamber and (d)–(f) rear side chamber

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Fig. 14

Dimensionless rotational angular velocity profiles in two side chambers at three viscosities 1, 93, and 237 mm2/s for case C; (a)–(c) front side chamber and (d)–(f) rear side chamber

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Fig. 15

Pressure coefficient profiles in two side chambers at three viscosities 1, 93, and 237 mm2/s for case A; (a)–(c) front side chamber and (d)–(f) rear side chamber

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Fig. 16

Pressure coefficient profiles in two side chambers at three viscosities 1, 93, and 237 mm2/s for case B; (a)–(c) front side chamber and (d)–(f) rear side chamber

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Fig. 17

Pressure coefficient profiles in two side chambers at three viscosities 1, 93, and 237 mm2/s for case C; (a)–(c) front side chamber and (d)–(f) rear side chamber

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Fig. 18

Local skin friction coefficients on wet walls of pump casing and impeller outside surfaces as a function of Reynolds number for case A, B, and C

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Fig. 19

Skin friction coefficients in front and rear wear-rings clearances (a) and flow resistance coefficient across balance holes (b) for case C in terms of Reynolds number

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