Mitsubishi Heavy Industries Technical Review Vol. 56 No. 2 (July 2019)
1Flow Field Analysis of a Turbocharger Centrifugal
Compressor Under Pulsating Conditions

ISAO TOMITA*1YOSHIHIRO HAYASHI*2

RICARDO F. MARTINEZ-BOTAS*3
MARIA E. BARRERA MEDRANO*3

There is growing awareness of global environmental conservation and the regulations onfuel consumption and exhaust emissions of automobile engines have been strengthened. In recentyears, especially, due to the spread of fuel efficiency improving technologies such as downsizing,
miller cycle and super lean burn, a wider operating range centrifugal compressor forturbochargers which achieves a high-pressure ratio in a wide operating range has been in highdemand. On the other hand, since a compressor is operated under pressure pulsation caused by air
intake of an engine, its internal flow exhibits a complicated unsteadiness, which has made itdifficult to grasp the flow phenomena toward enhancement of the operating range. Therefore, we
conducted a flow measurement using a pulsation test device which replicates an actual engine, andby applying an unsteady numerical analysis with the test result as the boundary condition, weconducted the analysis of the detailed flow structure in the pulsating flow which had not beenrevealed so far. The result showed that in the pulsating flow, the Helmholtz resonance frequency ofa pipe had a large effect on the operating range and the change of the flow pattern due to theappearance or disappearance of full stall at the impeller tip was predominant.

|1. Introduction
In recent years, there has been growing awareness of global environmental conservation, andthe regulations on exhaust emissions and fuel consumption of automobile engines have beenstrengthened. For turbochargers, improvement of the performance toward achievement of lower
fuel consumption of an engine has been strongly demanded. Figure 1 shows the system layout of
an automobile turbocharger. The engine exhaust gas energy rotates the turbine, which drives thecentrifugal compressor on the same shaft, and the compressed air is fed into the engine via thedownstream intercooler. At this time, the flow field of the centrifugal compressor and the turbine isexposed to pressure pulsation caused by air intake of the engine shown in Figure2. Therefore, for
improvement of the performance of a turbocharger, it is important to consider the pulsating flow
during development.
In the past development of a turbocharger considering the pulsating flow, MHI focused
attention on the flow under pulsation of turbine scroll to improve the efficiency, which wasdescribed in the MHI Technical Review Vol. 50 No. 1. This report introduces an example of thedevelopment focusing attention on the performance of a compressor under pulsation.
In the recent development trend for centrifugal compressors, fuel efficiency improving
technologies such as downsizing, miller cycle and super lean burn have widely spread and a wider
operating range of centrifugal compressors which achieve a stable supercharging pressure in a wide
operating range has been essential. Figure 3 shows the typical operating points of a centrifugal
compressor for a turbocharger. A compressor is used at various operating points according to the*1

Chief Staff Manager , Fluid Dynamics Research Department, Research & Innovation Center*2

HeadOffice Global Research & Innovation, Mitsubishi Heavy Industries Europe*3

Department of Mechanical Engineering,Imperial College London
University Technology Center, MHIET,Mitsubishi Heavy Industries Technical Review Vol. 56 No. 2 (July 2019)2vehicle operation state, and the operating range is restricted to the range from the operation limit on
the small flow rate side due to surging to the operation limit on the large flow rate side due tochoking. Surging, in particular, which restricts the operating range in the vicinity of the maximum
torque point of an engine, is an unstable flow phenomenon in which a reverse flow due to thereverse pressure gradient of a compressor is involved. In order to achieve a wider operating range,
it is necessary to accurately grasp the flow structure of surging in an actual engine.
As previously described, however, a turbocharger is exposed to pressure pulsation associatedwith air intake of an engine. Therefore, the internal flow is a complicated unsteady flow whichlargely deviates from a steady flow without pulsation. Concerning surging of centrifugal
compressors, many studies 1)2), irrespective of experimental or numerical analysis, have been
conducted conventionally, and a part of the surging is being revealed. But there are few examples
of detailed analysis of surging under the operating conditions involving pressure pulsation as
described above, and no sufficient findings have been obtained yet.

Figure 1 System layout ofturbocharger

Figure 2 Pressure pulsation of theturbocharger engine

Figure 3 Operating points of turbocharger
compressor for automobile|2. Experiment and flow analysis methods
2.1

Compressor under study

Figure 4 shows the turbocharger compressor used in the current study. This compressor is acentrifugal compressor used for a turbocharger and uses an open-type backswept impeller whichhas 10 full blades and 10 splitter blades s. After air is given kinetic energy by the impeller, its speed
is reduced and its pressure is increased by the downstream vaneless diffuser, and the air is supplied
to the engine through the scroll connecting to the air supply manifold. This compressor has noadditional device such as casing treatment*i. Figure 5 shows the relationship between the flow rate
and the pressure ratio for each rotational speed of the test compressor (performance characteristic
diagram). In this study, the flow near surge point at a pressure ratio of 1.7 is analyzed. This
operating point corresponds to the engine low end torque region where the turbocharger operates at
low flows, and thus, compressor surge margin plays a key role.

Mitsubishi Heavy Industries Technical Review Vol. 56 No. 2 (July 2019)3

Figure 4 Appearance and schematic diagram of test
compressor

2.2

Figure 5 Pressure and flow ratecharacteristic diagram

Measurement method for surge under pressure pulsation3)

To analyze the effect of the pressure pulsation waveform on the surge characteristic, MHI
has conducted a joint research with Imperial College London, being this institution one of the mostexperienced in measuring turbocharger performance. Imperial College team has developed apulsating flow device able to reproduce a wide variety of pulsating conditions similar to thoseencountered in engine operation. 3) Figure 6 shows the compressor facility including the bespoke
pulsation flow device. This pulsating flow device is composed of a rotary valve locateddownstream of the compressor outlet, which generates different pulsing conditions – amplitude andfrequency of the pulse can be set in the pulse generator by changing the rotating speed and/or thedisc size/shape, respectively.
Table 1 shows a summary of the test conditions evaluated during the experimental campaign
carries out at Imperial College London with the above-described pulsation flow device. In this test,
the pressure waveform at the engine torque point of a general passenger car was replicated by thepulsating flow device. The selected engine condition used as reference in this campaign
corresponds to a sinusoidal wave form and pulse frequency of 66.7 Hz (equivalent to an engine
speed of 2,000 rpm) in the pulsating flow device. The experimental campaign considered thecompressor performance measurement at different outlet pulsating frequencies and amplitudes. Thesurge flow rate at each test condition was measured and the effect of the pressure pulsation
waveform on the surge characteristic was analyzed.

Figure 6Pressure pulsation test device

Mitsubishi Heavy Industries Technical Review Vol. 56 No. 2 (July 2019)
4

Table 1 Pressure pulsation test condition
Rotational Speed of Wheel
Frequency of Pulsation
Amplitude of Pulsation
Downstream Pipe Volume

2.3

40000rpm
20～100Hz
66.7Hz for referenceApprox. 5% from mean pressure(Near surge point)
0.1661m3

Unsteady flow analysis method

In order to analyze the detailed internal flow in the above test result, unsteady flow fieldanalysis was conducted using Computational Fluid Dynamics (CFD). The numerical analysis
model and the analysis condition used in this study are shown in Figure 7 and Table 2. The
ANSYS-CFX analysis code was used and the SST--kω turbulence model was used. The analyticalarea imitated the test device in the foregoing chapter, the hemispherical domain which corresponds
to the open field upstream of the impeller inlet pipe was provided, the outlet boundary was set atthe position corresponding to the rotational valve of the test device, and analysis was conducted at
the rotational speed equivalent to the test condition. For the computational grid, impeller anddiffuser are created by a structural grid and an upstream/downstream pipe including scroll iscreated by an unstructured grid. The total number of cells is about 13 million. The static pressure
fluctuation obtained by the measurement result in the foregoing chapter was added as the outlet
boundary condition, and the time history data of unsteady flow for four cycles of pressure pulsation
waveform was obtained.

Figure 7 Numerical analysis model
Table 2 Unsteady numerical analysis condition
Flow Solver

ANSYS-CFX 14.5

Governing Equation

3D-RANS

Numerical Scheme

2nd-Order

Turbulence Model

SST k-omega
Total Pressure 1atm
Total Temperature 293.15k
Static Pressure (Ref. Fig.8)

Inlet Boundary Condition
Outlet Boundary Condition|3. Clarification and improvement of the flow phenomenon under
discharging pulsation
3.1

Measurement result for surge characteristic under pulsation4) 5)

In order to analyze the effect of the pressure pulsation waveform, the surge flow rate wasmeasured with different pulsation frequencies and pulsation amplitudes. In this study, the pulsation
for the frequency of 66.7 Hz (equivalent to 2,000 rpm) corresponding to the maximum torque pointof a general 4-cylinder engine was set as the reference waveform (Figure 8), the pulsation
frequency f was changed in the range of 20 Hz to 100 Hz, and the surge flow rate in each case wasMitsubishi Heavy Industries Technical Review Vol. 56 No. 2 (July 2019)5obtained.
Figure 9 shows the comparison of the pressure flow characteristic between the steady flow
without addition of pressure pulsation and the pulsating flow with the reference pressure waveform
added. When the pressure pulsation is added to the compressor, the inertia force associated withincrease or decrease of the flow rate acts on the fluid in the flow path, which suppresses a reverse
flow associated with surging, and the surge flow rate is reduced compared to the steady flow. 5) 6)
According to the test result with this compressor, the surge flow rate SP2 in the pulsation flow isshifted to the small flow rate side by 8.0% relative to the surge flow rate SP1 in the steady flow,
and this indicates that the pressure pulsation suppresses surging in the compressor.
Figure 10 shows the pressure flow characteristic of the test compressor for different pressure
pulsation frequencies. The largest surge flow rate is at the pulsation frequency of 20 Hz, where theoperating range is narrow. This corresponds to the resonance frequency of the downstream pipe ofthe test compressor, and the Helmholtz resonance frequency*2 is calculated from the cross-sectional
area and volume/length of the pipe is around 23 Hz. The test result shows that at 20 Hz, which isclosest to the Helmholtz resonance frequency, the surge flow rate has an extreme value. As thefrequency increases from the above condition, the surge flow rate reduces, and the operating range
is the largest at 43.3 Hz close to the frequency which is twice as high as the Helmholtz resonance.
Accordingly, the surging characteristic under pulsation has a strong correlation with the resonancefrequency of the downstream pipe, and it is considered that an appropriate design of the pipe shapeof an actual engine can increase the stable operating range on the small flow rate side.

Figure 8

Reference pressure waveform

Figure 9

Relationship between presence or absence of pressure pulsation and surge flow rate

Mitsubishi Heavy Industries Technical Review Vol. 56 No. 2 (July 2019)
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Figure 10 Relationship between pulsation frequency and surge flow rate

3.2

Result of unsteady flow analysis

In order to analyze the detailed internal flow under pressure pulsation, flow analysis wasconducted under the conditions equivalent to that of the test result for the above reference
waveform. Figure 11 shows the comparison of the pressure flow characteristic between thenumerical analysis result and the experimental result. In Figure 10, the characteristic curve shows
good agreement, and it is considered that the analysis result adequately reproduced an actual flow.
Figure 12(a) and (b) show the fluctuations of the velocity vector angle (flow angle) over time at thecompressor impeller inlet and the diffuser outlet. In Figure 11, at the impeller inlet (a), the flow
angle suddenly increases and decreases at several points corresponding to the pressure pulsation
cycle, while at the diffuser outlet (b), the fluctuations synchronized with pulsation are small. Thismeasurement was conducted at the operating point of a low-pressure ratio of around 1.7, where thespeed reduction rate of the diffuser is smaller compared to that at other operating points.
Accordingly, it is considered that the fluctuations synchronized with pulsation at the impeller rather
than at the diffuser were shown.

Figure 11

Pressure and flow rate characteristic (numerical analysis result and experimental result)Mitsubishi Heavy Industries Technical Review Vol. 56 No. 2 (July 2019)
7Figure 12

Fluctuations of flow angle at impeller inlet and outlet

In order to visualize the change in the internal flow of the impeller mentioned above, thevelocity contour in the reverse flow region in the vicinity of the impeller tip at each unsteadyanalysis time (designated as Points A to D) is shown in Figure 13. The size of the reverse flow
range at each of Points A to D corresponds to the intake air flow rate shown in the same figure andthe reverse flow develops in the upstream side from the impeller inlet. So, it is considered that theabove-mentioned fluctuations of the flow angle at the impeller inlet indicate the development andreduction of the reverse flow range. Especially, at Point C, which is the minimum flow point, afull-stall flow field is observed where the reverse flow region covers the area all around theimpeller inlet and the pulsation causes a strong unsteady flow in which appearance anddisappearance of full stall are repeated.

Figure 13

Distribution of reverse flow region inside impeller

Mitsubishi Heavy Industries Technical Review Vol. 56 No. 2 (July 2019)
8In order to analyze the three-dimensional flow structure for the impeller internal unsteadyflow described above, a visualization process for the vortex structure based on the singular pointtheory 7) was performed. Figure 14 shows the three-dimensional vortex structure at each of Points
A to D. The vortex structure is colored by nondimensional helicity (cosine of the velocity vector
and the vortex vector), and +1.0 (red) in the figure shows a reverse flow and -1.0 (blue) shows avertical vortex in the forward flow direction. In the same figure, the typical three-dimensionalvortex structure patterns for a centrifugal compressor impeller, such as leakage vortex at theimpeller tip and secondary vortex and horseshoe vortex at the impeller hub, can be observed. At
each operating point of Points A to D, the leakage flow vortex at the tip exhibits significant
changes such as inversed helicity by reverse flow, while the vortex structure at the hub maintains
almost the same pattern. Correspondingly, according to the impeller inlet velocity distribution in
the same figure, a significant change in the flow velocity occurs in the reverse flow region at thetip, but due to this effect, the fluctuations in the flow rate are cancelled at the tip, where the flow isalmost steady. Accordingly, in the pulsating flow, the flow change at the impeller tip ispredominant, and it is suggested that the surge margin can be further expanded through positive
control of the flow at the tip .

Figure 14

Three-dimensional vortex flow structure inside impeller|4. Challenges and future prospects
With this technology, the flow structure in a compressor under pressure pulsation waselucidated. The future challenges are further detailed measurement for verifying the analysis result
and improvement of the precision of numerical analysis by a more advanced numerical analysis
scheme/turbulence model. In the future, we would like to promote development of an aerodynamic
device adaptable to pulsating flow by reflecting the findings obtained by this technology in thedesign.

|5. Conclusion
In order to make detailed flow analysis in the vicinity of the surge point in a centrifugal
compressor for a turbocharger, we developed a pressure pulsation test device which imitated anactual engine, conducted unsteady measurement for the surge characteristic under pulsation, which
was conventionally difficult, and clarified unsteady internal flow. This technology facilitates
extraction of the flow that relates to expansion of the surge margin and will contribute to expansionof the operating range of a compressor. In the future, we will continue to contribute to improving
engine performance and environmental conservation through improving the performance ofturbochargers.

Mitsubishi Heavy Industries Technical Review Vol. 56 No. 2 (July 2019)9The test device used in this study was developed in joint research with Imperial CollegeLondon (U.K.), and we are grateful to Professor Ricardo Martinez Botas and Dr. Maria Esperanza
Medrano of the same college.
*i Recirculating casing treatment: Device for expanding the operating range by connecting the impeller internal
wall (casing) and the inlet pipe and forming a recirculation flow*ii Helmholtz resonance frequency: Calculated by the frequency, sound velocity, cross-sectional area ofopening, pipe length and pipe volume when resonance is produced in a pipe line having an opening

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Eckardt,D ”Detailed Flow Investigations Within a High Speed Centrifugal Compressor Impeller”, ASME
Journal of Fluids Engineering,Vol 98,pp.390-402(2) Ibaraki,S.,et al.: Design and Off-Design Flow Field of Transonic Centrifugal Compressor Impeller”,
ASME GT2009-27791(3) M. E. Barrera-Medrano,2017, “The effect of exit pressure pulsation on the performance and stability limitof a turbocharger centrifugal compressor,” Imperial College London.
(4) M. E. Barrera-Medrano and R Martinez-Botas,”On the effect of engine pulsation on the performance of acentrifugal compressor”, Institution of Mechanical Engineers Turbocharging Conference , London,2018(5) M. E. Barrera-Medrano, 2016,P. Newton, R. Martinez-Botas, S. Rajoo, I. Tomita, and S. Ibaraki, “Effectof Exit Pressure Pulsation on the Performance and Stability Limit of a Turbocharger CentrifugalCompressor,” J. Eng. Gas Turbines Power, vol. 139, no. 5, p. 052601.
(6) M Shu and M yang, “Unsteady Responses of The Impeller of a Centrifugal Compressor Exposed toPulsating Backpressure” ASME, GT2018-76851(7) Sawada, K., A convenient visualization method foridentifying vortex centers, Trans. Japan Soc. forAero.Space Sci., Vol.38 (1995), 102.