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Research Papers

Enhancement of Centrifugal Compressor Operating Range by Control of Inlet Recirculation With Inlet Fins

[+] Author and Article Information
Hideaki Tamaki

Corporate Research and Development,
IHI Corporation,
1, Shin-Nakahara-Cho,
Isogo-Ku, Yokohama 235-8501, Japan
e-mail: hideaki_tamaki@ihi.co.jp

Masaru Unno

Numerical Engineering Department,
IHI Corporation,
Yokohama 235-8501, Japan

Ryuuta Tanaka, Yohei Ishizu

Research and Development Division,
Aero-Engine and Space Operations,
IHI Corporation,
Tokyo 190-1297, Japan

Satoshi Yamaguchi

Vehicular Turbocharger Operation,
IHI Corporation,
Yokohama 235-8501, Japan

Contributed by the International Gas Turbine Institute (IGTI) of ASME for publication in the JOURNAL OF TURBOMACHINERY. Manuscript received November 25, 2015; final manuscript received March 1, 2016; published online May 3, 2016. Editor: Kenneth C. Hall.

J. Turbomach 138(10), 101010 (May 03, 2016) (12 pages) Paper No: TURBO-15-1278; doi: 10.1115/1.4033187 History: Received November 25, 2015; Revised March 01, 2016

The operating points of a turbocharger compressor tend to approach or cross its surge line while an engine is accelerating, particularly under low-engine speed conditions, hence the need for an acceptable surge margin under low compressor-speed conditions. A method shifting the stability limit on a compressor low-speed line toward a lower flow rate is expected and inlet recirculation is often observed in a centrifugal compressor with a vaneless diffuser near a surge and under a low compressor-speed condition. First, examples of inlet recirculation were introduced in this paper, whereupon the effect of inlet recirculation on compressor characteristic was discussed by 1D consideration and the potential shown for growth of inlet recirculation to destabilize compressor operations. Accordingly, this study focused on suppressing the effect of inlet recirculation on compressor characteristics using small fins mounted in a compressor-inlet pipe, and examining whether they enhance the compressor operating range under low-speed conditions. Small fins are known as inlet fins in this paper. According to test results, they showed great promise in enhancing the compressor operating range during inlet recirculation. Besides, attempts were also made to investigate the qualitative effect of inlet fins on flow fields using computational fluid dynamics (CFD) and the disadvantages of inlet fins were also discussed.

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Figures

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Fig. 1

Schematic view of conventional recirculation device

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Fig. 2

Static pressure at positions in Fig. 3

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Fig. 3

Schematic of compressor

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Fig. 4

Axial velocity distribution at I in Fig. 3: (a) 0.45 m3/s and (b) 0.39 m3/s

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Fig. 5

Compressor pressure ratio and total temperature just before the bell-mouth

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Fig. 6

Streamline traversing near the leading edge of the full blade suction surface (left), streamline through the full blade tip clearance (right): (a) 0.70 m3/s, (b) 0.52 m3/s, (c) 0.45 m3/s, and (d) 0.37 m3/s

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Fig. 7

Axial flow velocity of the tip-clearance flow: (a) 0.70 m3/s, (b) 0.52 m3/s, (c) 0.45 m3/s, and (d) 0.37 m3/s

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Fig. 8

Circumferentially averaged streamline: (a) 0.70 m3/s, (b)0.52 m3/s, (c) 0.45 m3/s, and (d) 0.37 m3/s

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Fig. 9

Compressor characteristics and total temperature rise in the inlet pipe

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Fig. 10

Schematic of the test rig

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Fig. 11

Flow angle distribution in the inlet pipe: (a) operating point A and (b) operating point B

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Fig. 12

Illustration of meridional flow in the impeller with inlet recirculation and the 1D model for total pressure

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Fig. 13

Compressor characteristics

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Fig. 14

Static pressure at the impeller leading edge and exit at Mu = 0.82 and 1.01 and positions of the static pressure holes. (a) Static pressure at the impeller leading edge and exit and (b) positions of static pressure holes.

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Fig. 15

Oil-flow visualization: (a) m/md = 0.81, (b) m/md = 0.55, and (c) m/md = 0.35

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Fig. 16

Pictures of inlet fins (arrow showing the impeller rotational direction)

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Fig. 17

Compressor characteristics with and without inlet fins

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Fig. 18

Compressor efficiency with and without inlet fins

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Fig. 19

Static pressure at impeller leading edge and exit with and without inlet fins: (a) Mu = 0.82 and (b) Mu = 1.01

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Fig. 20

Calculated and measured static pressure: (a) static pressure at impeller exit and (b) static pressure at the diffuser exit

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Fig. 21

Spanwise axial and circumferential velocity distribution near the leading edge at m/md = 0.2

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Fig. 22

Spanwise total pressure distribution near the impeller leading edge at m/md = 0.2

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Fig. 23

Streamline with axial velocity at m/md = 0.2: (a) without inlet fins and (b) with inlet fins

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Fig. 24

Streamline with axial velocity at m/md = 0.35: (a) without inlet fins and (b) with inlet fins

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Fig. 25

Spanwise axial and circumferential velocity distribution near the leading edge at m/md = 0.35

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Fig. 26

Calculated impeller efficiency

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Fig. 27

Relative flow angle distribution at the leading edge: (a)m/md = 0.35 and (b) m/md = 0.20

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Fig. 28

Spanwise total temperature at the leading edge

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Fig. 30

Schematic of compressor systems

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Fig. 31

Compressor characteristics: (a) compressor A and (b) compressor B

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