Research Papers

Simulation of Multistage Compressor at Off-Design Conditions

[+] Author and Article Information
Feng Wang

Osney Thermo-Fluids Laboratory,
Department of Engineering Science,
University of Oxford,
Oxford OX2 OES, UK
e-mail: feng.wang@eng.ox.ac.uk

Mauro Carnevale

Osney Thermo-Fluids Laboratory,
Department of Engineering Science,
University of Oxford,
Oxford OX2 OES, UK
e-mail: mauro.carnevale@eng.ox.ac.uk

Luca di Mare

Osney Thermo-Fluids Laboratory,
Department of Engineering Science,
University of Oxford,
Oxford OX2 OES, UK
e-mail: luca.dimare@eng.ox.ac.uk

Simon Gallimore

Rolls-Royce plc,
Derby DE21, UK
e-mail: Simon.Gallimore@Rolls-Royce.com

1Corresponding author.

Contributed by the International Gas Turbine Institute (IGTI) of ASME for publication in the JOURNAL OF TURBOMACHINERY. Manuscript received September 18, 2017; final manuscript received September 27, 2017; published online December 12, 2017. Editor: Kenneth Hall.

J. Turbomach 140(2), 021011 (Dec 12, 2017) (10 pages) Paper No: TURBO-17-1165; doi: 10.1115/1.4038317 History: Received September 18, 2017; Revised September 27, 2017

Computational fluid dynamics (CFD) has been widely used for compressor design, yet the prediction of performance and stage matching for multistage, high-speed machines remains challenging. This paper presents the authors' effort to improve the reliability of CFD in multistage compressor simulations. The endwall features (e.g., blade filet and shape of the platform edge) are meshed with minimal approximations. Turbulence models with linear and nonlinear eddy viscosity models are assessed. The nonlinear eddy viscosity model predicts a higher production of turbulent kinetic energy in the passages, especially close to the endwall region. This results in a more accurate prediction of the choked mass flow and the shape of total pressure profiles close to the hub. The nonlinear viscosity model generally shows an improvement on its linear counterparts based on the comparisons with the rig data. For geometrical details, truncated filet leads to thicker boundary layer on the filet and reduced mass flow and efficiency. Shroud cavities are found to be essential to predict the right blockage and the flow details close to the hub. At the part speed, the computations without the shroud cavities fail to predict the major flow features in the passage, and this leads to inaccurate predictions of mass flow and shapes of the compressor characteristic. The paper demonstrates that an accurate representation of the endwall geometry and an effective turbulence model, together with a good quality and sufficiently refined grid, result in a credible prediction of compressor matching and performance with steady-state mixing planes.

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Fig. 1

Schematic of the shrouded compressor

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Fig. 2

(a) Mesh on the blade-to-blade section, (b) close-up view of the mesh around the platform filet, (c) mesh for a stator variable vane with penny gaps, and (d) boundary layer mesh around the blade filet. The figures have been intentionally distorted.

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Fig. 3

Hybrid mesh for the shroud cavity under S4. The figure has been intentionally distorted.

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Fig. 4

Schematic of the compressor model. The figure has been intentionally distorted.

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Fig. 5

Grid sensitivity study of S3–S4 block

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Fig. 6

Radial flow angle in front of S3 leading edge with two sets of meshes. Left: radial flow angles within 20% span. Right: radial flow angles within 4% span.

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Fig. 7

Compressor map using different turbulence models and geometry details at the 100% speed. Data are normalized by their corresponding values at the design point of the 100% speed.

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Fig. 8

Turbulent kinetic energy on the midchord plane at the design and high working line conditions of the 100% speed. The figures have been intentionally distorted.

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Fig. 9

Stage performance for the front stage and one of the back stages

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Fig. 10

Compressor map computed by meshes with truncated and nontruncated filets at the 100% speed. Values are normalized by the choked mass flow of the experiment.

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Fig. 11

Comparison of truncated and nontruncated filet–I

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Fig. 12

Comparison of truncated and nontruncated filet–II. The figure has been intentionally distorted.

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Fig. 13

Comparison of total pressure radial profiles in front of S1 and S4 leading edges

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Fig. 14

Compressor map at the 85% speed. Data are normalized by their corresponding values at the design point of the 85% speed.

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Fig. 15

Relative total pressure coefficient at the midchord plane of S1 and R2 at the design point of the 85% speed. This figure has been intentionally distorted.

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Fig. 16

Radial profiles of total pressure in front of S2 leading edge at the design point of the 85% speed



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