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Research Papers

Three-Dimensional Computational Fluid Dynamics Prediction of Turbocharger Centrifugal Compression System Instabilities

[+] Author and Article Information
Rick Dehner

Mem. ASME
Department of Mechanical and Aerospace Engineering,
The Ohio State University,
201 W. 19th Avenue, Columbus, OH 43210
e-mail: dehner.10@osu.edu

Ahmet Selamet

Mem. ASME
Department of Mechanical and Aerospace Engineering,
The Ohio State University,
201 W. 19th Avenue, Columbus, OH 43210
e-mail: selamet.1@osu.edu

1Corresponding author.

Manuscript received September 20, 2017; final manuscript received January 30, 2019; published online March 6, 2019. Assoc. Editor: Nicole L. Key.

J. Turbomach 141(8), 081004 (Mar 06, 2019) (13 pages) Paper No: TURBO-17-1169; doi: 10.1115/1.4042728 History: Received September 20, 2017; Accepted January 31, 2019

The present work combines experimental measurements and unsteady, three-dimensional computational fluid dynamics predictions to gain further insight into the complex flow-field within an automotive turbocharger centrifugal compressor. Flow separation from the suction surface of the main impeller blades first occurs in the mid-flow range, resulting in local flow reversal near the periphery, with the severity increasing with decreasing flow rate. This flow reversal improves leading-edge incidence over the remainder of the annulus, due to (a) reduction of cross-sectional area of forward flow, which increases the axial velocity, and (b) prewhirl in the direction of impeller rotation, as a portion of the tangential velocity of the reversed flow is maintained when it mixes with the core flow and transitions to the forward direction. As the compressor operating point enters the region where the slope of the constant speed compressor characteristic (pressure ratio versus mass flow rate) becomes positive, rotating stall cells appear near the shroud side diffuser wall. The angular propagation speed of the diffuser rotating stall cells is approximately 20% of the shaft speed, generating pressure fluctuations near 20% and 50% of the shaft frequency, which were also experimentally observed. For the present compressor and rotational speed, flow losses associated with diffuser rotating stall are likely the key contributor to increasing the slope of the constant speed compressor performance curve to a positive value, promoting the conditions required for surge instabilities. The present mild surge predictions agree well with the measurements, reproducing the amplitude and period of compressor outlet pressure fluctuations.

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References

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Figures

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Fig. 1

Locations of instrumentation (identified in red) installed at the compressor inlet, along with locations utilized for CFD analyses (identified in green)

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Fig. 2

Confined computational domain for steady-state 3D CFD predictions

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Fig. 3

Full computational domain (turbocharger stand compression system) for 3D CFD predictions

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Fig. 4

Mesh of the rotating region for the 3D CFD model, including (a) shroud surface (wall) and interface meshes, and (b) inner surface (wall) mesh, along with volume mesh on planes passing through the axis of rotation and main blade leading edges

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Fig. 5

Comparison of compressor predictions from CFD with steady-state experimental data for (a) total-to-total pressure ratio versus corrected mass flow rate and (b) total-to-total isentropic efficiency versus corrected mass flow rate

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Fig. 6

Comparison of temperature elevation near the main blade leading edges relative to ambient at Ncor = 80 krpm for CFD predictions and experimental data

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Fig. 7

Predicted radial variation of static temperatures at Ncor = 80 krpm located 14.6 mm upstream of the main blade leading edge

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Fig. 8

Predicted radial variation of axial velocity at Ncor = 80 krpm located 14.6 mm upstream of the main blade leading edge

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Fig. 9

Predicted tangential velocity contour at Ncor = 80 krpm and m˙c,cor=21.2g/s on the plane located 14.6 mm upstream of the main blade leading edge

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Fig. 10

Predicted impeller velocity streamlines showing flow reversal from the inducer blade tips at Ncor = 80 krpm and m˙c,cor=21.2g/s

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Fig. 11

Predicted radial velocity contours at Ncor = 80 krpm on a radial plane near (0.1136 mm from) the hub side of the diffuser

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Fig. 12

Predicted compressor operating points during the perturbation from the stable region to the unstable region at Ncor = 80 krpm

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Fig. 13

Time-resolved mass flow rates during the transient CFD perturbation to the unstable operating regime

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Fig. 14

Time-resolved static pressure at the compressor duct absolute pressure transducer locations during the transient CFD perturbation to the unstable regime, for (a) compressor inlet and (b) compressor outlet

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Fig. 15

Total SPL in the 20–30 Hz range for mild surge detection from the compressor outlet duct experimental data at Ncor = 80 krpm

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Fig. 16

Comparison of measured and predicted low-pass filtered pressure at the compressor outlet duct absolute pressure transducer location during mild surge with a time-averaged corrected mass flow rate of 15.5 g/s and Ncor = 80 krpm

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Fig. 17

Predicted radial velocity contour on a radial plane located 0.1466 mm from the shroud side diffuser wall at t = 146.27 ms

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Fig. 18

Predicted radial velocity near the shroud side outlet of the diffuser as a function of angular position during the flow rate reduction transient

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Fig. 19

Predicted rotating stall cell propagation near the shroud side outlet of the diffuser as a function of angular position

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Fig. 20

Predicted radial velocity contour plots on a radial plane near the shroud side diffuser wall, showing the propagation of rotating stall cells in the diffuser

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Fig. 21

Frequency domain analysis of predicted pressure fluctuations near the outlet of the shroud side diffuser wall during rotating stall

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Fig. 22

Experimental SPL at the compressor inlet duct measurement location at Ncor = 80 krpm and time-averaged corrected mass flow rate of 15.5 g/s

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Fig. 23

Dimensionless wall distance on the impeller surface at t = 0.144 s

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