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2nd Quadrant Centrifugal Compressor Performance: Part I

[+] Author Affiliations
Elisabetta Belardini, Dante Tommaso Rubino, Libero Tapinassi, Marco Pelella

GE Oil & Gas, Florence, Italy

Paper No. GT2016-57117, pp. V009T24A016; 10 pages
doi:10.1115/GT2016-57117
From:
  • ASME Turbo Expo 2016: Turbomachinery Technical Conference and Exposition
  • Volume 9: Oil and Gas Applications; Supercritical CO2 Power Cycles; Wind Energy
  • Seoul, South Korea, June 13–17, 2016
  • Conference Sponsors: International Gas Turbine Institute
  • ISBN: 978-0-7918-4987-3
  • Copyright © 2016 by ASME

abstract

Performance curves in 2nd quadrant are important to size protection equipment of both compressor and surrounding system. With reverse flow also the level and frequency of pressure fluctuations in different operating points is important to estimate blade loading and possible presence of excitation frequencies. The capability of performance predictive tools (either CFD or correlations based methods) as also mechanical design criteria are generally poor in 2nd quadrant and suffer for the scarcity and inadequacy of validation data.

The second quadrant branch for a centrifugal compressor has been experimentally tested after the standard characterization in direct flow. A test arrangement has been designed, with a booster compressor connected in parallel with the tested stage, forcing the flow to be stable in reverse flow. The compressor characteristics have been measured with static and dynamic instrumentation. Present experience showed that when machine is operated in the stable reverse flow condition, pressure fluctuations and vibration are higher with respect to the values measured in nominal direct flow operating conditions. The increase is in the order of 10–20% of the corresponding value in direct flow. The same can be stated also for axial thrust and secondary flows that increase when the gas flows in reverse direction but the increase is in the order of 10–15%. Thus in 2nd quadrant, compressor equipment (in particular impeller blades) and all other system devices experience unusual loading levels but the additional loads are not big enough to cause relevant damaging if sustained for limited time periods. This result may allow simplifying the design of system layout: in particular, if during ESD no surging cycle is expected but only a reverse flow sustained steadily by the system (which is actually the most typical experience), the additional ASV’s, such as hot or cold gas by-pass valves, may be reduced in size or eventually removed optimizing the plant BOP.

Copyright © 2016 by ASME
Topics: Compressors

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