Improved Compressor Exit Diffuser for an Industrial Gas Turbine

[+] Author and Article Information
U. Orth, H. Ebbing

MAN Turbomaschinen AG GHH BORSIG, Oberhausen, Germany

H. Krain, A. Weber, B. Hoffmann

German Aerospace Center (DLR), Cologne, Germany

J. Turbomach 124(1), 19-26 (Feb 01, 2001) (8 pages) doi:10.1115/1.1413476 History: Received February 01, 2001
Copyright © 2002 by ASME
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Meridional contours of rotors Ro0 and Ro5
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Inlet boundary conditions for impeller calculations
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Total/total isentropic efficiencies calculated for rotors Ro0 and Ro5
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Radial diffuser inverse boundary layer calculation
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Two different blade shapes for radial diffuser
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Velocity vectors at 50 percent span Original (tandem blading, top) and final design (bottom)
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Inlet distributions for original tandem arrangement and new radial blading at reference plane I2 of Fig. 10. Three-dimensional result without axial blading.
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Q3D-MISES optimization for M1=0.20 and AVDR=0.90. Isentropic profile Mach number distribution for α1=40 deg and simulated total pressure losses.
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S2m grid topology (each second grid line skipped) and reference planes, S1 grid inside of radial blading and modeling of clearance
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Geometry variation: Flow path and positioning of axial blading
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Static pressure contours and velocity vectors inside of axial blade passage at 50 percent chord. Comparison original versus final design (bottom).
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Exit swirl angle distribution α in reference plane E2 (NRO: axial part without blading)
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Mean total pressure distribution in meridional plane (conservative averaging); (p2: mean static pressure at exit plane)
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Spanwise development of circumferentially averaged swirl angle and total pressure along complete diffuser flow path
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Diffuser geometry: Comparison original (top) versus final design. Right: axial blading.
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Comparison of original and new diffuser region
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Total compressor pressure ratio versus mass flow
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Total compressor efficiency versus mass flow
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Centrifugal stage pressure ratio versus mass flow
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Centrifugal stage efficiency versus mass flow




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