Mitsubishi Heavy Industries Technical Review Vol. 49 No. 1 (March 2012)
68

Development of a Wide-Range Centrifugal Compressor
for Automotive Turbochargers

SEIICHI IBARAKI*1

ISAO TOMITA*2

MOTOKI EBISU*3

TAKASHI SHIRAISHI*4

The use of turbochargers has been increasing in response to strengthened automotive
exhaust emission and fuel consumption regulations for global environmental protection. The
variable geometry (VG) turbocharger has recently become especially popular, creating a demand
for a centrifugal compressor that provides wide and stable operation. The internal flow of acentrifugal compressor is very complicated with three dimensional and unsteady flow phenomena,and the analysis of flow phenomena and expansion of the operational range are difficult problems.
Mitsubishi Heavy Industries, Ltd. (MHI) has developed a centrifugal compressor with anoperational range wider than that of conventional units, based on experimental and numericalanalysis of the complicated flow phenomena, and control of the tip leakage vortices at the impeller
blades.

|1. Introduction
With the ongoing movement toward global environmental protection, regulations related tothe exhaust emissions and fuel consumption of automobiles are being strengthened. To cope withthese requirements, turbochargers provide an effective tool for improving fuel consumption andreducing carbon dioxide emissions, by reducing engine weight and friction loss. Since aturbocharger supplies compressed air to an engine, it can reduce the engine displacement relative toa naturally aspirated engine. Variable geometry (VG) turbochargers, which can control the boostpressure according to engine operating conditions, are becoming increasingly popular, creating ademand for a centrifugal compressor with a wide and stable operational range.
The structure of an automotive turbocharger is illustrated in Figure 1. The turbine, which isrotated by exhaust gas, drives the coaxial compressor impeller. Air is compressed and accelerated
by the impeller, and its pressure increases while decelerating in the diffuser. This compressed air issupplied to the engine through the scroll.
When the flow rate of a centrifugal compressor is reduced, surging occurs. This produces anoverall system pulsation and leads to an operational limit. To expand the operational range of acentrifugal compressor, it is necessary to lower the flow rate limit at which surging occurs.
However, flow phenomena are complex and unsteady near the surging range, and the detailed flow
structure in the impeller has been unexplained in the past.1 In this report, we clarify the unsteadyflow phenomena near the surging range by measuring pressure fluctuations and using unsteadynumerical analysis,2,3 and introduce the new wide-range centrifugal compressor, which wasdeveloped on the basis of this analysis.

*1

Nagasaki Research & Development Center, Technology & Innovation Headquarters*2

Manager, Nagasaki Research & Development Center, Technology & Innovation Headquarters*3Manager, Turbocharger Engineering Department, General Machinery & Special Vehicles*4Turbocharger Engineering Department, General Machinery & Special Vehicles

Mitsubishi Heavy Industries Technical Review Vol. 49 No. 1 (March 2012)
69

Figure 1

Automotive turbocharger|2. Experiments and the flow analysis technique
2.1Study subject

Table 1 lists the impellers and the main specifications of a conventional compressor and thenew compressor. These compressors are used in automotive turbochargers. The compressors haveopen-type backswept impeller, in which the blade exit is slanted opposite to the direction ofrotation. In addition to five full blades, an equal number of splitter blades are installed in theconventional impeller to increase the pressure ratio. The new type of impeller is equipped with four
full blades and four splitter blades. The performance characteristics, shown in Figure 2, indicate
the relationship between the flow rate and the pressure ratio at each rotational speed. The
operational range of the centrifugal compressor extends from a maximum flow limit (at whichchoking occurs) to a minimum flow limit (at which surging occurs). The abscissa represents the air
flow rate normalized by a 180,000 rpm choke flow, and the efficiency depicted by the contour lines
is normalized by the maximum efficiency of the conventional compressor. For the conventional
compressor, the surge flow limit increases at pressure ratios of 2.2 and above, and the operational
range is thereby reduced. On the other hand, a wide operational range is maintained at pressure
ratios of 2.2 and above with the new compressor. An unsteady upward sloping of the 160,000 rpm
line is seen in the low flow rate area of the conventional compressor. For the new compressor,
however, all areas have stable downward-sloping curves, and the operational range at a pressure
ratio of 2.5 is increased by 130%. This report describes the phenomena occurring at 160,000 rpm,
where there are significant differences between the characteristics of the conventional and newcompressors.
Table 1Main specifications
ConventionalNewcompressor compressor

Number offull blade
Number ofsplitter blade
Inlet outer
diameterOutletdiameter
Diffuser outlet
diameter

5454

39.5mm

37.1mm

50.0mm

49.6mm

71.5mm

71.5mmFigure 2 Performance characteristics of centrifugalcompressors

Mitsubishi Heavy Industries Technical Review Vol. 49 No. 1 (March 2012)
702.2

Technique of measuring pressure fluctuations

To investigate the unsteady phenomena, high-response pressure transducers were installed atpositions 3 mm upstream from the impeller and 1.1 times the radius at the periphery of the diffuser,
as shown in Figure 3. The transducers were installed at two positions along the circumference, andunstable phenomena were analyzed via the time difference between the disturbances detected in thecompressor. The blade passing frequency (BPF: speed N  number of blades Z (Hz)) of thepressure fluctuations caused by the passage of full blades and splitter blades is 26.7 kHz in theconventional compressor and 21.6 kHz in the new compressor. Hence, the sampling frequency forthe pressure fluctuation measurements was set at 1 MHz, and 35 or more points between the fullblades and splitter blades were measured. Near the surging (where the flow phenomena becomeunstable), the rotating stall phenomenon (in which a separated flow area rotates at a speed lower
than that of the impeller) is known to occur. Therefore, to detect this, frequency fluctuations of 3 N(Hz) or lower were selected for a low-pass filter.

2.3

Numerical analysis technique

Computational fluid dynamics (CFD) was conducted to reproduce the unsteady phenomena,
and was applied to the entire compressor, including the impeller, diffuser, and scroll. A total of
3.13 million cells (2.47 million for the impeller, 0.41 million for the diffuser, and 0.25 million forthe scroll) were used for the computational grid, illustrated in Figure 4. For the CFD analysis, theANSYS CFX three-dimensional viscous flow analysis code was used. The k-ε model was used forthe turbulence model.

Figure 3

Pressure fluctuation measuring positions

Figure 4

Computational grid

|3. Clarification and improvement of the stall phenomenon
3.1

Result of pressure fluctuation measurement

Figure 5 shows the results of the pressure fluctuation measurements in the conventionalcompressor.
Figure 5 (a) shows the results at the peak pressure point. The pressure wave fluctuation of theimpeller was plotted over a 0.25 ms interval, corresponding to 0.7 of the impeller rotation. The
pressure fluctuation caused by blade passing was clearly detected, and the pressure wave of eachblade had the same shape. The waveform in the low-pass filter was plotted over 5 ms,corresponding to 13 rotations of the impeller. At this operating point, low-frequency phenomena
such as rotating stall were not detected. In the pressure fluctuation wave of the diffuser, nosignificant phenomena other than the passing of full blades and splitter blades were observed.
Figure 5 (b) shows the pressure fluctuation waves at the small flow rate point. The low-pass filter
wave of the impeller exhibited low-frequency instability. Moreover, at the diffuser, pressure
fluctuations corresponding to full blades and splitter blade passage were difficult to distinguish, andunstable phenomena were assumed to affect the diffuser.
Figure 6 shows detailed analysis results from the impeller low-pass filter at the small floworate point. Transducers 1 and 2 were separated by a 30 angle, and the time interval between thedetection of a low-frequency disturbance at transducer 1 and the detection of the disturbance atootransducer 2 was 0.05 ms. The disturbance therefore required 0.05 ms  360 / 30 = 0.60 ms toocover 360 . In the actual measurements, a similar disturbance was detected by transducer 1 after
0.60 ms. The time required for one complete impeller rotation was 0.375 ms. Thus, the rotational
speed of the disturbance was approximately 63% of the impeller speed. Moreover, threeMitsubishi Heavy Industries Technical Review Vol. 49 No. 1 (March 2012)71

disturbances were simultaneously detected. This phenomenon is the same as a rotating stall withseparated flow areas on the impeller blades, and each disturbance was presumed to be a rotatingstall cell. However, it differed from a normal rotating stall. Some very unstable behaviors, such as
temporal changes from two to four rotating stall cells, were observed.
The pressure fluctuation measurement results for the new compressor are shown in Figure 7.
At the peak efficiency point, shown in Figure 7(a), a low-frequency fluctuation was not observed,
unlike the case of the conventional compressor. At the small flow rate point, a slight low-frequency
fluctuation was observed in the low-pass filter waveform. However, the fluctuation was contained
within several rotations, and no significant phenomenon was evident. Also, full blades and splitterblades could be distinguished at the diffuser, and the flow phenomena at the blades were assumed
to be uniform.

Figure 5

Results of pressure fluctuation measurements (conventional compressor)

Figure 6

Structure of a rotating stall (conventional impeller)

Figure 7

Results of pressure fluctuation measurements (new compressor)

Mitsubishi Heavy Industries Technical Review Vol. 49 No. 1 (March 2012)72

3.2

Results of the unsteady flow analysis

The unsteady flow analysis was conducted under the same operational conditions as thepressure fluctuation measurements. Figure 8 compares the calculated and measured results for thepressure ratio. Although the absolute values showed slight differences, the pressure-flow
characteristics of the conventional and new compressor were well simulated, and the actual flow
was reproduced. Figure 9 shows the three-dimensional vortex structure and stream lines near theblade. The center of the vortex is colored in accordance with the normalized helicity, and +1/-1
indicates a clockwise/anticlockwise longitudinal vortex. The figure shows the generation of a
tornado-shaped vortex on the blades of the conventional compressor. Its movement between blades
is as follows. The mainstream flow avoids the vortex on the suction surface of the splitter blade,and the attack angle (the difference between the flow angle and the blade angle) becomes large at t
= 0.0 ms. Flow separation is observed on the next blade at t = 0.049 ms, and the separated flow
rises as a tornado-shaped vortex at t = 0.090 ms. Each vortex moves over the impeller in this way,while the number of vortices changes temporally, as the measurements indicate. This type of vortex
is the main generator of a rotating stall cell, and it is thought that the circumferential expansion ofthese cells with decreasing flow rate is what triggers surging.

Figure 8 Verification ofnumerical analysis

Figure 9 Internal flow structure(small flow rate point in a conventional compressor)

The internal flow structure of the new compressor is shown in Figure 10. In the newcompressor, the leakage flow between the blade tip and the blade casing forms a tip leakage vortex.
This type of vortex cannot maintain a longitudinal shape via its own rotation and the adverse
pressure gradient in the impeller. Hence, vortex breakdown occurs, and the vortex dissipates in aspiral flow. The dissipated leakage vortex forms a uniform stall layer in the circumferential
direction at the blade tip ends. Figure 11 shows an image of the flow structure at the small flowrate point. In the conventional compressor, a rotating stall cell is generated by the tornado-shaped
vortex and its accompanying stall area. In the new compressor, the flow is stabilized by creating auniform thin stall area at the tip of the blades, and this is accomplished by controlling the local tip
leakage vortices. The stall area created by the dissipated leakage flow becomes uniformly thick as
the flow rate decreases, and contributes to flow stabilization near the surging range. On the otherhand, reduction of efficiency is controlled by eliminating the stall area at the high flow rate
operating point.

Mitsubishi Heavy Industries Technical Review Vol. 49 No. 1 (March 2012)73

Figure 10 Internal flow structure(small flow rate point in a newcompressor)

Figure 11

Images of flow structure|4. Tasks and future prospects
The operating range of the compressor is expanded by controlling the blade-tip leakage
vortices of the impeller. However, factors leading to the occurrence of surging still requireclarification, and quantitative prediction of surging is beset by many problems, due to the extreme
complexity of the internal flow of a centrifugal compressor. Improved accuracy of numericalanalysis will be necessary in the future, and unsteady flow experiments should be conducted tovalidate the unsteady numerical analysis. Moreover, we would like to further improve theperformance of the compressor by optimizing the operating range expansion in terms of devices
such as the casing treatment, by realizing a variable compressor, and in other ways.

|5. Conclusion
A wide-range centrifugal compressor was developed, based on pressure fluctuation
measurements and unsteady numerical analysis, in order to expand the operational ranges ofturbochargers to cope with more stringent exhaust emission and fuel consumption regulations. The
new compressor will contribute to improved engine torque by increasing the boost pressure foracceleration at the operational point of a small flow rate. New types of superchargers (such as
two-stage turbochargers4 and electric superchargers5,6) are currently being developed for practicalapplication. Improved performance of the compressor (which is one of the fundamental elements)
will affect the performance of all types of superchargers. We would like to continue ourcommitment to improved environmental protection and driving pleasure via improved turbocharger
performance.

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