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Research Papers: Flows in Complex Systems

Numerical Analysis of the Fluid Flow in the First Stage of a Two-Stage Centrifugal Pump With a Vaned Diffuser

[+] Author and Article Information
H. Stel, C. O. R. Negrão

Federal University of Technology - Paraná,
UTFPR, 80230-901 Curitiba, PR, Brazil

G. D. L. Amaral

UO-BC/PETROBRAS,
Macaé, RJ, 27913 350 Brazil

S. Chiva

Universitat Jaume I,
Castellón de la Plana, E 12071 Spain

V. Estevam

E&P-ENGP/PETROBRAS,
Rio de Janeiro, RJ, 20035-900 Brazil

R. E. M. Morales

Federal University of Technology - Paraná,
UTFPR, 80230-901 Curitiba, PR, Brazil
e-mail: rmorales@utfpr.edu.br

1Corresponding author.

Contributed by the Fluids Engineering Division of ASME for publication in the JOURNAL OF FLUIDS ENGINEERING. Manuscript received July 16, 2012; final manuscript received October 23, 2012; published online May 15, 2013. Assoc. Editor: Chunill Hah.

J. Fluids Eng 135(7), 071104 (May 15, 2013) (9 pages) Paper No: FE-12-1331; doi: 10.1115/1.4023956 History: Received July 16, 2012; Revised October 23, 2012

This work presents a numerical investigation of the fluid flow in the first stage of a two-stage centrifugal pump with a vaned diffuser. A computational fluid dynamics (CFD) analysis is performed by using the ANSYS-CFX software for a wide range of volumetric flow rates and also for different rotor speeds. The numerical results are validated against measured values of pressure rise through the impeller and diffuser of the first stage with an overall good agreement. Nevertheless, not only the best efficiency point evaluated numerically is overestimated in comparison with the measured two-stage pump values but also the computed hydraulic efficiency of the first stage. Investigation of the flow pattern for different flow rates reveals that the flow becomes badly oriented for part-load conditions. In such cases, significant levels of turbulence and blade orientation effects are observed, mainly in the diffuser. In spite of different flow rates or rotor speeds, the flow pattern is quite similar if the flow dimensionless coefficient is kept constant, showing that classical similarity rules can be applied in this case. By using such rules, it was also possible to derive a single equation for the pump head to represent the whole operational range of the pump.

Copyright © 2013 by ASME
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References

Figures

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Fig. 6

Experimental setup

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Fig. 5

Influence of the initial condition. (a) Time evolution of pressure at the leading edge of a diffuser vane; and (b) azimuthal pressure calculation at the impeller-diffuser interface after half impeller revolution.

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Fig. 4

Computational grid

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Fig. 3

Schematics of the subdomains and boundary conditions adopted in the numerical simulations

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Fig. 11

Hydraulic efficiency for the rotor speed of 1150 rpm. Comparison of the numerical results for the single-stage model, experimental data and manufacturer's curve for the two-stage pump.

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Fig. 2

Schematics of the pump under analysis. (a) Section view of the two-stage pump; and (b) top view of the first stage (without the shroud and the pump casing).

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Fig. 1

Pictures of the two-stage centrifugal pump under analysis: (a) Full mounted; (b) first stage and return channel; (c) first impeller without the hub; and (d) diffuser

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Fig. 7

Positions of the pressure taps in the two-stage pump

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Fig. 8

Pictures of the test rig. (a) Overview of the setup and (b) pump with a Plexiglas cover to show the pressure taps and the internal parts.

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Fig. 9

Comparison of the numerical and experimental head curves for the first stage of the pump

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Fig. 10

Comparison of the numerical and experimental pressure rises through the first impeller (a) and the diffuser (b)

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Fig. 12

2D streamline paths and velocity magnitude contours at a plane located in the middle height of the impeller-diffuser interface for four selected flow rates and a rotor speed of 1150 rpm: (a) 20 m3/h, (b) 35 m3/h, (c) 45 m3/h, and (d) 55 m3/h

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Fig. 13

2D streamline paths and velocity magnitude contours at a plane located in the middle height of the impeller-diffuser interface for four rotor speeds and a fixed flow coefficient (φ=0.132)

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Fig. 15

Volume average of Ti and Ui inside the diffuser as a function of the flow coefficient for the four different rotor speeds. (Dashed lines are fourth-order polynomial fittings for each data set.)

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Fig. 16

Numerical results of the head coefficient ψ versus flow coefficient φ for four different rotor speeds

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Fig. 17

Numerical head values compared with the fitted curve

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Fig. 14

Contour plots of the instantaneous turbulence and unsteady intensity at a plane located in the middle height of the impeller-diffuser interface for a part-load condition and the best efficiency flow coefficient: (a) Ti for ϕ = 0.059, (b) Ui for ϕ = 0.059, (c) Ti for ϕ = 0.132, and (d) Ui for ϕ = 0.132

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