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Research Papers

The Design Space Boundaries for High Flow Capacity Centrifugal Compressors

[+] Author and Article Information
Daniel Rusch

Compressor Development (Dept. ZTE),
ABB Turbo Systems Ltd.,
Bruggerstrasse 71a,
CH-5401 Baden, Switzerland
e-mail: Daniel.Rusch@ch.abb.com

Michael Casey

Institute of Thermal Turbomachinery (ITSM),
University of Stuttgart, Germany and PCA Engineers Limited,
Lincoln, England

Contributed by the International Gas Turbine Institute (IGTI) of ASME for publication in the JOURNAL OF TURBOMACHINERY. Manuscript received August 8, 2012; final manuscript received August 13, 2012; published online March 25, 2013. Editor: David Wisler.

J. Turbomach 135(3), 031035 (Mar 25, 2013) (11 pages) Paper No: TURBO-12-1170; doi: 10.1115/1.4007548 History: Received August 08, 2012; Revised August 13, 2012

A methodology has been derived allowing a fast preliminary assessment of the design of centrifugal compressors specified for high specific swallowing capacity. The method is based on one-dimensional (1D) design point values using classical turbomachinery analysis to determine the inlet geometry for the maximum mass flow function. The key results are then expressed in a series of diagrams which draw out the nature of the conflicting boundary conditions of the design. In particular, it is shown how the inlet casing relative Mach number causes the design flow coefficient to decrease with the total pressure ratio and determines the inlet eye diameter. Physically based boundaries of operation are added to the diagrams giving guidelines for the proper choice of specification values to the designer. In addition, links are given to some well-known impeller efficiency correlations, so that a preliminary estimate of the performance can be made. Comparisons are made with a range of compressor data which supports the approach. The derived methodology allows any given specifications to be checked rapidly for feasibility and development risk or can be used to define a challenging specification for the design of a new product.

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References

Figures

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Fig. 1

Dependence of stage adiabatic efficiency on the dimensionless specific speed (ωs = Ns), pressure ratio, and impeller inlet casing relative Mach number (Mw1 = M1s) given by Rodgers [2] for high pressure ratio stages

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Fig. 2

Dependence of compressor polytropic efficiency on the flow coefficient and tip-speed Mach number (Mu2 = Mu) derived by Casey and Robinson [3]

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Fig. 3

Dependency of isentropic efficiency on the polytropic efficiency and total pressure ratio according to Eq. (4)

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Fig. 4

Velocity triangle at impeller inlet casing

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Fig. 5

Modified mass flow function Φ' for a centrifugal compressor with zero inlet swirl and for two different values of the isentropic exponent based on Eq. (17)

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Fig. 6

Dependence of Mu2 on φt1/k and different values of γ according to Eq. (19)

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Fig. 7

Flow coefficient according to Eq. (21). The diameter ratio values are given for an example design at an inlet flow coefficient of φt1 = 0.09 in air.

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Fig. 8

Stage total pressure ratio versus swallowing capacity expressed as inlet flow coefficient, assuming constant efficiency

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Fig. 9

Stage total pressure ratio versus swallowing capacity expressed as specific speed, assuming constant efficiency

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Fig. 10

Stage total pressure ratio versus swallowing capacity expressed as inlet flow coefficient, but using the efficiency correlation according to Fig. 2

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Fig. 11

Tip-speed Mach number versus swallowing capacity expressed as inlet flow coefficient. The efficiency levels are given in terms of polytropic efficiency divided by the maximum value using the correlation according to Fig. 2.

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Fig. 12

Physical limits of the centrifugal compressor design space. The shading is explained in the text and indicates the regions where designs become more difficult.

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Fig. 13

1D design flow chart

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Fig. 14

Stage tip-speed Mach number versus swallowing capacity for a range of compressor stages in comparison to Fig. 11

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