Stability Improvement of High-Pressure-Ratio Turbocharger …scholar.harvard.edu/files/jhuenteler/files/yang_et_al... · 2015-02-20 · STABILITY IMPROVEMENT OF HIGH-PRESSURE-RATIO

Proceedings of ASME Turbo Expo 2010 Power for Land Sea and Air

GT2010

June 14ndash18 2010 Glasgow UK

GT2010-22581

STABILITY IMPROVEMENT OF HIGH-PRESSURE-RATIO TURBOCHARGER CENTRIFUGAL COMPRESSOR BY ASYMMETRIC FLOW CONTROLmdashmdashPART I NON-AXISYMMETRIC FLOW

IN CENTRIFUGAL COMPRESSOR

YANG Mingyang ZHENG Xinqian ZHANG Yangjun

State Key Laboratory of Automotive Safety and Energy

Tsinghua University Beijing 100084 China

Email zhengxqtsinghuaeducn

BAMBA Takahiro TAMAKI Hideaki

IHI Corporation Tokyo Japan

HUENTELER Joern

RWTH Aachen University Germany

LI Zhigang

Beijing Special Vehicle Institute Beijing

100072 China

ABSTRACT

This is the Part I of a two-part paper documenting the

development of a novel asymmetric flow control method to

improve the stability of a high-pressure-ratio turbocharger

centrifugal compressor Part I focuses on the

non-axisymmetric flow in a centrifugal compressor induced

by the non-axisymmetric geometry of the volute while Part

II describes the development of asymmetric flow control

method to avoid the stall on the basis of the characteristic of

non-axisymmetric flow To understand the asymmetries

experimental measurements and corresponding numerical

simulation were carried out The static pressure was

measured by probes at different circumferential and

stream-wise positions to gain insights about the

asymmetries The experiment results show that there is an

evident non-axisymmetric flow pattern throughout the

compressor due to the asymmetric geometry of overhung

volute The static pressure field in the diffuser is distorted at

approximately 90deg in rotational direction of the volute

tongue throughout the diffuser The magnitude of this

distortion varies slightly with the rotational speeds The

magnitude of the static pressure distortion in the impeller is

a function of the rotational speed There is a significant

phase shift between the static pressure distributions at the

leading edge of the splitter blades and the impeller outlet

The numerical steady state simulation neglects the

mentioned unsteady effects found in the experiments and

can not predict the phase shift but a detailed asymmetric

flow field structure are obviously obtained

NOMENCLATURE

AR ratio of volute throat area to its radius

C sound speed

pC specific heat at constant pressure

sD magnitude of static pressure distortion

(P

PPDs

max )

f impeller rotational frequency

H

head coefficient (22

12 )(

U

TTCH p

)

Copyright copy 2010 by ASME 1

Proceedings of ASME Turbo Expo 2010 Power for Land Sea and Air GT2010

June 14-18 2010 Glasgow UK

GT2010-22581

Downloaded 29 Oct 2011 to 12913220921 Redistribution subject to ASME license or copyright see httpwwwasmeorgtermsTerms_Usecfm

L characteristic length in strouhal number

P static pressure

p average static pressure

RPM rotation per minute

St strouhal number (

C

fLSt )

t time (

C

Lt )

T total temperature

U impeller rotational velocity

W relative flow velocity

Subscripts

1 impeller inlet

2 impeller outlet

max maximum

1 INTRODUCTION The history of turbocharging is almost as old as that of

the internal combustion engine A turbocharger consists of a

compressor and a turbine The compressor is driven by the

turbine extracting energy from exhaust gases Compared to a

naturally aspirated engine the benefits of a turbocharged

engine are increased power lower fuel consumption and

reduced emissions [1 2]

High-pressure-ratio turbocharging technology is the

developing trend of turbocharged internal combustion

engine for the following reasons 1) significant downsizing

to mitigate CO2 emission and reduce fuel consumption [3]

2) Satisfying rigid future emission regulation ie NOx

treatment by engine control means of high rates of exhaust

gas recirculation (EGR) [4] and 3) the facilitation of high

altitude operation [5] However high pressure ratio causes

the flow in the compressor to be transonic Hence the stable

flow range is narrowed as the stall incidence decreases with

increased relative mach number at the inlet of the impeller

[6] Therefore map width enhancement is a major issue for

state-of-the-art high-pressure-ratio compressor design and

development


A turbocharger centrifugal compressor comprises an

impeller a diffuser and a volute While the former two are

periodically symmetric in circumferential direction the

volute is asymmetric due to its gas-collection function It is

usually designed as spiral-collection overhung housing that

collects the air from diffuser and passed it to the pipe

system It has been recognized that the improvement of

centrifugal compressor performance requires a good

understanding of the flow mechanisms inside the volute

[7-9] especially the interaction among

volute-diffuser-impeller [10 11] The volute is mostly

designed in a way to shape a uniform circumferential static

pressure distribution both in volute and diffuser However

the volute acts as a diffuser at lower than design flow rate

and acts as a nozzle at higher than design flow rate

respectively A number of authors have researched this

subject It has already been confirmed that the asymmetrical

configuration has a significant impact on the flow field in

the diffuser as well as in the impeller [12 13] This

circumferential asymmetry has been recognized and

intensive experimental investigations of the flow within the

volute and the propagation of the distortion into upstream

components were carried out for subsonic compressor units

[14 15]

The work of Sorokes et al (1998) confirmed that the

pressure non-uniformity extended upstream of the impeller

implying that the impeller was subjected to varying inlet and

exit conditions deteriorating performance and especially

stability as the circumferential variations drove the

compressor into surge at higher flow rate as without the

volute [14] A three-dimensional unsteady analysis of the

flow in the impeller with circumferential distortion of the

outlet static pressure was investigated by numerical method

by Fatsis et al (1997) [16] The perturbation was thus

propagated upstream from the impeller outlet and influence

the incidence at the blade leading edges as well as other

flow parameters Gu et al (2001) [10 11] found that the

performance parameters of the single impeller passage differ

because of the asymmetric flow at the outlet of the impeller

There was almost no phase shift between the distortion in

the diffuser and impeller according to their results and it

was considered that unsteady effects of the volute-impeller

interaction can be neglected when the Strouhal number is

small enough

Little detailed measurement was carried out in the

impeller to investigate the asymmetric flow Further to the

authorsrsquo knowledge few research work focused on the

impact of the volute on the flow field in a

high-pressure-ratio turbocharger centrifugal compressor

The purpose of this two-part paper is first to understand the

asymmetry of flow field due to the asymmetric geometry of

the volute and subsequently to develop a novel asymmetric

flow control method to widen the stable flow range of a

turbocharger centrifugal compressor with


high-pressure-ratio the narrowing flow range of which is of

utmost importance for its application In Part I the

non-axisymmetric flow characteristics in the

high-pressure-ratio turbocharger centrifugal compressor are

investigated by using the experimental and numerical

means the results of which are the basis for the work

presented in Part II

2 EXPERIMENT FACILITIES The measurements were performed on the turbocharger

testing facility Key geometry and operational characteristics

of the investigated centrifugal compressor are listed in Table

1 The volute was designed overhung with an elliptic

cross-section as usually

Table 1 Centrifugal compressor geometry

Design rotational speed N 65000RPM

Exit blade angle (meridional) -38deg

Impeller outlet blade tip velocity U 511 ms

Diffuser exit diameter D2 228 mm

Impeller diameter D5 150 mm

2

5D

D 152

AR 2623 mm

Blade number 99 mainsplitter blades


In order to investigate the static pressure distribution in

the compressor 41 measurement points were positioned in

both meridional and streamwise direction Figure 1(a-c)

shows the locations of the probes in the compressor In

streamwise probes were located in the shroud in the vicinity

of the leading edge of the main blades and splitter blades

respectively as shown in Fig 1(a) Two other streamwise

positions were located at the inlet as well as outlet of

vaneless diffuser respectively as shown in Fig 1(b) For

each streamwise position in impeller and diffuser eight

static pressure probes were mounted in uniform

circumferential distribution Nine pressure probes were

mounted in the wall of the volute as shown in Fig 1(c) The

distribution of static pressure in circumferential direction

throughout the compressor was obtained by the arranged

probes with an error of less than plusmn02

(a) Measurement in the impeller

(b) Measurement in the diffuser

(c) Measurement in the volute

Fig 1 Measurement positions in the investigated

compressor

3 NUMERICAL METHODS In order to investigate the detailed flow field in the

compressor numerical simulation was employed The used

solver applies a CFD code based on a 3-D steady

compressible finite volume scheme to solve the

Reynolds-averaged Navier-Stokes equations in conservative


formulation Central scheme was used for spatial

discretization while forth-order Runge-Kutta scheme for

temporal discretization The Spalart-Allmaras 1-equation

model was applied for turbulence closure [17] The S-A

model features numeration accuracy for the calculation of

viscous boundary layer turbulent flow separated flows of

small or medium scale and free shear turbulent flow except

for jet [18]

All passages of the impeller the vaneless diffuser and

the volute were modeled in the simulation The

computational domain in the region upstream of the impeller

was extended as shown in Fig 2(a) The whole domain

consisted of approximately 6500000 nodes The grid

exhibits a minimal grid quality as defined by measures of

orthogonality (especially at the block boundaries minimum

10deg) relative grid spacing in boundary layers (expansion

ratiole3) and grid skewness (aspect ratiole5000) To mesh

the overhung volute a butterfly mesh was applied in order

to augment orthogonality of boundary layer cells In Fig

2(a) the mesh of the passage projected on rotating solid

walls is shown with important entities marked In Fig 2(b)

a cross-section of the volute exit mesh is displayed in order

to point out the butterfly mesh approach Figure 2(c) shows

the volute mesh and passage numbering

Total temperature and total pressure together with the

velocity direction were imposed as inlet boundary

conditions static pressure was set as outlet boundary

condition No-slip and impermeability conditions were

imposed on the solid walls The interface between impeller

and volute was modeled as frozen rotor as angular

variations of flow quantities exceed a negligible level and

are object of this research work Thus the rotating system

was calculated in relative coordinates yet the flow

quantities are transferred locally without varying the relative

position of impeller and volute To quantify the effect of the

relative position on the simulation results three different

positions of the impeller were modeled and analyzed The

periodicity of the rotor is 40 degrees thus the rotor was

moved from angle φ to φ+133deg and φ+2666deg respectively

The performance results of the three different conditions are

shown in Fig 3 The results are quite similar to each other

which means that neglecting the influence of the relative

position of impeller and volute is reasonable

(a) passage mesh projected on solid walls

(b) butterfly meshing approach in volute cross-section

(c) volute mesh and passage numbering

Fig 2 Created mesh in detail



Fig 3 Influence of the relative position of the

impeller and volute on performance

4 RESULTS AND ANALISIS 41 Experimental results

Testing points and compressor performance

The static pressure distribution at different flow rates

and speed lines were measured on the testing facilities Four

speed lines ie 45000RPM (69N) 55000RPM (85N)

65000RPM (100N) 68000RPM (105N) will be

discussed (N is the design rotational speed) Three operating

points on each line were chosen for analysis which

represent the highest efficiency condition highest flow rate

condition (near choke) and lowest mass condition (near

surge) The points are shown in Fig 4

Fig 4 Measured compressor performance

The static pressure distribution in the volute

The static pressure distribution in the volute at design

rotational speed is shown in Fig 5 The detailed measuring

points can be referred in Fig 1(c) It can be seen from the

figure that the circumferential static pressure variations are

quite different for the three operating conditions Usually a

volute is designed as a tube in which the velocity of the air

remains constant at the design point or the point of highest

efficiency Therefore the static pressure distribution at the

highest efficiency point is relatively flatter than other

operating conditions as shown in the Fig 5 However as

the compressor approaches the surge condition the flow is

decelerated in the volute and it acts as a diffuser because of

the reduced flow rate As a result the static pressure in the

volute increases from the tongue to the compressor outlet

On the other hand when the compressor approaches the

choke point the static pressure in the volute decreases due

to the function of the volute as a nozzle Hillewaert et al

[12] and Hagelstein et al [19] reported similar results in

the volute (diffuser exit) It can be noticed from the figure

that there is a fluctuation downstream of the volute tongue

in the distribution at three conditions which is caused by the

complex disturbance of the tongue

Fig 5 Static pressure distributions in the volute at

design rotational speed N

The static pressure distribution in the diffuser

The static pressure distribution in the diffuser is shown

in Figs 6and 7 for two rotational speeds 45000RPM and

65000RPM respectively At both operation conditions near

surge and near choke similar trends are present in the

diffuser for both rotational speeds Near surge there is a

remarkable region of static pressure distortion with a

minimum at approximately 90deg downstream of the tongue at

the diffuser outlet and inlet (as shown in Fig 6) which is

coincident with the results reported in Ref [19] It was

supposed to be caused by the perturbation at volute tongue

The static pressure distortion is quite different for the

condition near choke as shown in Fig 7 where the static

pressure reaches the maximum value at around 90deg while it

falls to the minimum value around 90deg The fluctuation

around 90deg is caused by the pressure rise across the tongue

region



Fig 6 Static pressure distribution in the diffuser

near surge


near choke

A non-dimensional parameter Ds was used to evaluate

the magnitude of the distortion in different conditions

P

PPDs

max (1)

where is the maximum value of the static pressure

(absolute pressure) and

maxP

P is the circumferential area-averaged pressure for each operating point


Figure 8 shows the magnitude of the pressure distortion

in the diffuser near surge It can be seen that the magnitude

of the distortion at the diffuser inlet and outlet changes

slightly with the rotational speeds near surge For all the

rotational speeds the magnitude of the distortion at diffuser

inlet is much larger than at diffuser outlet For the design

rotational speed Ds is 0171 at diffuser inlet which is

almost 24 times the magnitude of the distortion at the

diffuser outlet A simple propagation of the pressure field

inside the volute is not sufficient to explain this

phenomenon because one would expect this kind of

disturbance to decay along the diffuser This indicates that

the interaction of impeller and diffuser further contributes to

the distorted flow field in the diffuser near surge Further

research should be performed on this phenomenon

Fig 8 Magnitude of static pressure distortion in

the diffuser near surge

The pressure distribution in the impeller

The static pressure distribution was further measured at

the shroud near the leading edge of the main and splitter

blades Figure 9 shows the static pressure near surge at the

leading edge of the splitter blades for four rotational speeds

The static pressure varies significantly in circumferential

direction at all rotational speeds viewed yet with quite

different patterns of disturbance For the lower rotational

speedss ie 45000RPM and 55000RPM there are two

peaks in circumference The static pressure has a minimum

around 90deg and 270deg For the higher speed the magnitude of

the distortion increases At 65000RPM and 68000RPM the

distortion has a main wave length of 360deg and remarkable

amplitude The static pressure reaches a minimum value at

about 220deg for the former and 207deg for latter The magnitude

of the fluctuation increased sharply as the rotational speed

increases It shall be noted that the patterns of the distortion

are significantly different from those measured in the

diffuser for the same speed and operation condition as

shown in Fig 6 where there are no great differences in

pattern between low and high rotational speeds although it

is understand that the distortion in the impeller is due to the

distortion at the inlet of the diffuser


Fig 9 Static pressure distribution at leading edge

of the splitter blades near surge

Figure 10 shows the static pressure distribution near

surge at the leading edge of the main blades Because that

the distortion is significantly suppressed by the upstream

uniform flow the fluctuation of the pressure is much slighter

than that near the leading edge of the splitter blades

However the distortion still exists at the inlet of the main

blades That means the distortion due to the asymmetric

geometry of overhung volute can be propagated upstream

from the diffuser outlet to the impeller inlet The incidence

as well as other flow parameters at the impeller inlet is

correspondingly variant in circumferential direction So

impeller is subjected to varying inlet flow conditions

although the upstream flow is uniform Comparing the

results shown in Figs 9 and 10 the static pressure at leading

edge of the splitter blades is higher than that at leading edge

of the main blades which indicates that the flow is diffused

in the inducer near surge

Fig 10 Static pressure distribution at the leading

edge of the main blades near surge


Figures 11 and 12 show the static pressure near choke at

the leading edge of the splitter and main blades

respectively The static pressure at the lead edge of the

splitter blades for the four rotational speeds peaks between

180deg and 225deg Similar to the near surge condition the

fluctuations of static pressure at the leading edge of the main

blades are is much slighter than that near the leading edge of

the splitter blades The relative flow velocity in the passage

is generally larger for higher rotational speeds Therefore

the static pressure for the higher speed is lower than that for

low rotational speed the similar phenomenon of which is

shown in Figs 10-12 Comparing the results shown in Figs

11 and 12 the static pressure at leading edge of the splitter

blades is lower than that at leading edge of the main blades

which indicates that the flow is accelerated in the inducer

near choke


edge of the splitter blades near choke


edge of the main blades near choke

The distortion phase shift in the impeller

Figure 13 shows the magnitude of static pressure

distortion at the leading edge of the splitter blades near

surge The magnitude of the distortion increases sharply as

the rotational speeds increases from 40000RPM to

68000RPM The unsteady interaction of the impeller and the

volute is characterized by the acoustic Strouhal number [11]

The number can be defined as follows

C

fLSt (2)

St represents the temporal relation between the rotation of


the impeller and the perturbation wave propagation in the

impeller For the compressor investigated in this paper lsquofrsquo is

defined as the rotational frequency lsquoLrsquo is defined as the

average length of the impeller passage in streamwise

direction and lsquoCrsquo is defined as sonic velocity evaluated with

for convenience the inlet temperature When the periodicity

of the perturbation and the wave propagation in the passages

are the same resonance happens St is about 02 at

45000RPM and 03 at 68000RPM Therefore St will

approach the case of resonance ie St = 05 with increasing

rotor speed This results in an increasing amplitude of

distortion inside the passages as the rotational speed

increases as it was shown in Fig 13 A high pressure ratio

compressor is likely to have a large St due to its high

rotational speed Therefore compared with the flow in a low

pressure ratio compressor the flow distortion in a

high-pressure-ratio compressor impeller is expected to be

much more severe and of utmost importance for

consideration of flow stability and compressor performance

Fig 13 Magnitude of static pressure distortion at

the leading edge of the splitter blades near surge

Figure 14 shows the static distribution at the diffuser

inlet and the leading edge of the splitter blades near surge at

design rotational speed N It can be seen that there is a phase

shift of about 110deg in the distortion from the impeller outlet

(diffuser inlet) to the leading edge of the splitter blades

Numerical simulation of a low-pressure-ratio compressor

was performed before to investigate the interaction between

the impeller and volute [19] and it was found that flow

distortion at the outlet and inside of the impeller is present in

the passages upstream of the tongue The blade wrap angle

was responsible for a part of the phase shift between the

trailing and leading edge but it is obviously not large

enough to be the single cause of the phenomenon As

discussed before St of the compressor is larger than 01

which means that unsteady effects need to be considered

The pressure perturbation propagates at the local sound

speed in the relative frame Therefore when it propagates

upstream in the passage ie from outlet to inlet the relative

velocity is C-W upstream in the rotating frame and C+W

when it is reflected and travels downstream (W is relative

flow velocity) Considering the back and forth propagation

of the perturbation the average velocity of propagation can

be evaluated by the sound speed C Therefore the time

which it takes for the distortion wave to travel from the

trailing edge to the splitter leading edge is evaluated as

follows

C

Lt (3)

where t is time For the investigated compressor t is about

018 ms During this short time the rotation angle of

impeller is about 70deg at the design rotational speed N Thus

the influence of the rotation of the impeller on the phase

shift is much larger than effect of the wrap angle which is

about 165deg Therefore the propagation of the disturbance

wave and the meanwhile rotation of the impeller are the two

main reasons for the phase shift of the distortion from the

outlet upstream into the impeller

Fig 14 Static pressure distribution at the diffuser

inlet and the leading edge of the splitter blades

near surge at N

42 Flow Simulation

As discussed above the Strouhal number of the

compressor at design point is about 03 Therefore the

unsteady effect of propagation of the disturbance wave was

significant in the impeller because of its rotation However

the steady state numerical method could still be used to

investigate the circumferential asymmetric flow field in the

compressor which is caused by the volute although the

phase shift at the outlet and inside of the impeller will

inevitably not be predicted correctly The steady state

simulation can be considered as the case in which the



velocity of the disturbance propagation is so large that the

time needed to propagate can be neglected

The numerical simulation has to be validated by

experimental results therefore the compressor performance

is compared in Fig 15 The flow rate was normalized by the

simulated choke flow rate and the head coefficient H was

defined as follows

22

12 )(

U

TTCH p

(4)

where is specific heat at constant pressure and

are the total temperature at the inlet and outlet of the

impeller respectively is the impeller rotational velocity

at the impeller outlet Figure 15 shows that the head

coefficient by simulation is 79 higher than that by

experiment Several factors caused the discrepancy between

the simulation and experiment which were discussed in Ref

[20] Overall good agreement is obtained between the

experimental and the simulation results indicating the

reliability of the simulation setup

pC 1T

2T

2U

Fig 15 Comparison of performance of experiment

and numerical simulation at N

Figure 16 shows the comparison of static pressure

distribution in circumferential direction at the outlet and

inlet of the diffuser between simulation and experiment

both near surge condition at 68000RPM A static pressure

distortion can be seen at approximately 90deg circumferential

angle for both simulation and experiment The predicted and

measured phase of distortion are in accordance with each

other which further indicates that the used numerical

simulation is credible to predict the asymmetric

characteristics of the flow field

Fig 16 Static pressure distributions comparison

in diffuser near surge at N

Figure 17 compares the circumferential distribution of

static pressure by simulation and experiment at the leading

edge of the splitter blades near surge at the design rotational

speed N In the simulation results a pressure distortion

exists at around 100deg which is approximately coincident

with the distortion phase at the impeller outlet The similar

phenomenon was reported in Ref [11 12] However it can

be seen that there is a large phase shift (about 110deg) between

simulation and experiment results The That means steady

state simulation can not predict the phase shift of the

distortion St is zero in steady state simulation However the

actual St is up to 03 at design rotational speed N and

unsteady effects can not be neglected

Fig 17 Static pressure distributions in

circumferential direction near the leading edge of

the splitter blades near surge at N

Although steady state simulation can not predict the

phase shift in the impeller the shapes of the static pressure

distribution by simulation and experiment are similar The

numerical steady simulation neglects the mentioned Copyright copy 2010 by ASME 9


unsteady effects found in the experiments but facilitates a

detailed analysis of the asymmetric flow field in the

impeller

Figure 18 shows the flow rate and the inlet flow angle

as well as the outlet static pressure distribution of each

passage near surge at design rotational speed N The x-axis

is circumferential degrees and passage number which can

be referred in Fig 1(c) The flow rate varies from passage to

passage The inlet flow angle increases as the flow rate

increases both peak in the second passage and fall to the

minimal value in passage 5 and 6 The variation of the flow

rate in the passages can be explained to some extent by the

pressure distribution at the outlet of the impeller since the

passage with larger adverse pressure gradient exhibits a

smaller flow rate However there is a phase shift between

the flow rate characteristic and the outlet static pressure

distribution Flow rate and inlet flow angle falls to the

lowest at passage 6 where the outlet static pressure does not

reach its maximum value

Fig 18 Flow parameters distribution in impeller

passages near surge at N

Figure 19 shows the static pressure distribution near the

tip of the impeller near surge at design rotational speed N It

can be seen that the pressure distribution in the impeller

passage is quite different because of the difference in the

structure of the supersonictransonic zone at the leading

edge of the main blades It is the flow rate and incidence that

changes the velocity distribution and therewith the flow

acceleration in each passage

Fig 19 Static pressure distribution near the tip of

the impeller blades near surge at N

Figure 20 shows the relative Mach number near the tip

of the impeller near surge at design rotational speed N The

largest part of the suction surface is covered by low

momentum flow in passage 5 and 6 in which the flow rate

is the least among passages The basic reason for impeller

stall is the large incidence which will cause blade surface

flow separation strong leakage flow and other kinds of

deteriorated flow in the passage According to the flow

angle distribution investigation in passages it can be

inferred that stall firstly happens in passages 5 and 6 It can

be concluded that the passages with flow distortion will be

more likely to stall at higher pressure ratio than at lower

pressure ratio because of the larger amplitude of distortion

Obviously a treatment for stability improvement

concentrating in this zone will be more efficient

Fig 20 Relative Mach number distribution near

the tip of the impeller blades near surge at N

5 CONCLUSIONS AND REMARKS

The non-axisymmetric flow in a high-pressure-ratio

turbocharger centrifugal compressor was studied The static




pressure distribution along the streamwise in the compressor

was measured and analyzed Conclusions could be drawn

from the experiments as follows

1 There is an evident non-axisymmetric flow pattern

throughout the compressor due to the asymmetric

geometry of overhung volute The static pressure field

in the diffuser is distorted at approximately 90deg in

rotational direction of the volute tongue throughout the

diffuser The magnitude of this distortion varies slightly

with the rotational speeds

2 The magnitude of the static pressure distortion in the

impeller is a function of the rotational speed The

magnitude increases sharply as the Strouhal number

approaches 05 which reflects the ratio of time needed

for wave propagation and between two excitations ie

the case where resonance happens in the passage

3 There is a significant phase shift between the static

pressure distribution at the leading edge of the splitter

blades and the impeller outlet This shift is to some

extend caused by the blade wrap angle but is mainly

due to the rotation in the timeframe that the pressure

wave needs to travel upstream in the impeller passage


mentioned unsteady effects found in the experiments but

facilitates a detailed analysis of the asymmetric flow field in

the impeller and allows to identify the passages where

inducer stall incepts first It can be concluded as follows

1 The static pressure distribution at the outlet of impeller

has a significant impact on the performance parameters

of the passages including the flow rate As the

magnitude of these fluctuations increases the stability

of the passage flow will be reduced

Based on the detailed study of the non-axisymmetric

flow field induced by the volute Part II of this two-part

paper will develop an asymmetric flow control method to

widen the stable flow range of the high-pressure-ratio

turbocharger centrifugal compressor

ACKNOWLEDGMENT

This research is supported by National Natural Science

Foundation of China (Grant No 50806040) and Specialized

Research Fund for the Doctoral Program of Higher

Education) (Grant No 200800031075)

REFERENCES

[1] Baines N C 2005 ldquoFundamentals of

Turbochargersrdquo Concepts ETI Inc

[2] Hawley J G Wallace1 F J Cox1 A et al 1999

ldquoVariable Geometry Turbocharging for Lower

Emissions and Improved Torque Characteristicsrdquo

Proceedings of the Institution of Mechanical

Engineering Part D Journal of Automobile

Engineering 213 (2) pp 145-159

[3] Clenci A C Descombes G Podevin P and Hara

V 2007 ldquoSome Aspects Concerning the Combination

of Downsizing with Turbocharging Variable

Compression Ratio and Variable Intake Valve Liftrdquo



Engineering 221 (10) pp 1287-1294

[4] Maiboom A Tauzia X and Heacuteteta J F 2008

ldquoExperimental Study of Various Effects of Exhaust

Gas Recirculation (EGR) on Combustion and

Emissions of an Automotive Direct Injection Diesel

Enginerdquo Energy 33 (1) pp 22-34

[5] Rodgers C 2001 ldquoTurbocharging a High Altitude

UAV CI Enginerdquo AIAA Paper No 2001-3970

[6] Krain H Karpinski G and Beversdorff M 2001

ldquoFlow Analysis in a Transonic Centrifugal Compressor

Rotor Using 3-Component Laser Velocimetryrdquo ASME

paper No 2001-GT-0315

[7] Ayder E and Van den Braembussche R A 1991

ldquoExperimental Study of the Swirling Flow in the

Internal Volute of a Centrifugal Compressorrdquo ASME

Paper No 91-GT-7


ldquoNumerical Analysis of the Three-Dimensional

Swirling Flow in Centrifugal Compressor Volutesrdquo

Journal of Turbomachinery 116 (2) pp 462-468

[9] Ayder E Van den Braembussche R A and Brasz J

J 1993 ldquoExperimental and Theoretical Analysis of

the Flow in a Centrifugal Compressor Voluterdquo Journal

of Turbomachinery 115 (3) pp 582-589

[10] Gu F H Engeda A Cave M and Liberti J L D

2001 ldquoA Numerical Investigation on the

VoluteDiffuser Interaction Due to the Axial Distortion

at the Impeller Exitrdquo Journal of Fluids Engineering

123 (3) pp 475-483

[11] Gu F H and Engeda A 2001 ldquoA Numerical

Investigation on the VoluteImpeller Steady-State

Interaction due to Circumferential Distortionrdquo ASME

Paper No 2001-GT-0328

[12] Hillewaert K and Van den Braembussche R A

1999 ldquoNumerical Simulation of Impeller-Volute



Interaction in Centrifugal Compressorsrdquo Journal of

Turbomachinery 121 (3) pp 603-608

[13] Reunanen A Pitkanen H Siikonen T et al 2000

ldquoComputational and Experimental Comparison of

Different Volute Geometries in a Radial Compressorrdquo

ASME Paper No 2000-GT-469

[14] Sorokes J Borer C and Koch J 1998

ldquoInvestigation of the Circumferential Static Pressure

Non-Uniformity Caused by a Centrifugal Compressor

Discharge Voluterdquo ASME Paper No 98-GT-326

[15] Hagelstein D Van den Braembussche R Keiper R

and Rautenberg M 1997 ldquoExperimental

Investigation of the Circumferential Pressure

Distortion in Centrifugal Compressor Stagesrdquo ASME

Paper No 97-GT-50

[16] Fatsis A Pierret S and Van den Braembussche R A

1997 ldquoThree Dimensional Unsteady Flow and Forces

in Centrifugal Impellers with Circumferential

Distortion of the Outlet Static Pressurerdquo Journal of


[17] Spalart P R and Allmaras S R 1992 ldquoA One

Equation Turbulence Model for Aerodynamic

Flowsrdquo AIAA Paper No 92-0439

[18] Guo G D Zhang Y J Xu J Z et al 2008

ldquoNumerical Simulation of a Transonic Centrifugal

Compressor Blades Tip Clearance Flow of Vehicle

Turbochargerrdquo ASME Paper No GT2008-50957

[19] Hagelstein D Hillewaert K and Van den

Braembussche R A 2000 ldquoExperimental and

Numerical Investigation of the Flow in a Centrifugal

Compressor Voluterdquo Journal of Turbomachinery 122

(1) pp 22-31

[20] Guo G D Zhang Y J Zheng X Q et al 2008

ldquoVortex Flows in the Blade Tip Region of a Transonic

Centrifugal Compressorrdquo Journal of Tsinghua

University (Sci amp Tech) 48 (11) pp 1804-1808 (in

Chinese)


L characteristic length in strouhal number

P static pressure

p average static pressure

RPM rotation per minute

St strouhal number (

C

fLSt )

t time (

C

Lt )

T total temperature

U impeller rotational velocity

W relative flow velocity

Subscripts

1 impeller inlet

2 impeller outlet

max maximum

1 INTRODUCTION The history of turbocharging is almost as old as that of

the internal combustion engine A turbocharger consists of a

compressor and a turbine The compressor is driven by the

turbine extracting energy from exhaust gases Compared to a

naturally aspirated engine the benefits of a turbocharged

engine are increased power lower fuel consumption and

reduced emissions [1 2]

High-pressure-ratio turbocharging technology is the

developing trend of turbocharged internal combustion

engine for the following reasons 1) significant downsizing

to mitigate CO2 emission and reduce fuel consumption [3]

2) Satisfying rigid future emission regulation ie NOx

treatment by engine control means of high rates of exhaust

gas recirculation (EGR) [4] and 3) the facilitation of high

altitude operation [5] However high pressure ratio causes

the flow in the compressor to be transonic Hence the stable

flow range is narrowed as the stall incidence decreases with

increased relative mach number at the inlet of the impeller

[6] Therefore map width enhancement is a major issue for

state-of-the-art high-pressure-ratio compressor design and

development


A turbocharger centrifugal compressor comprises an

impeller a diffuser and a volute While the former two are

periodically symmetric in circumferential direction the

volute is asymmetric due to its gas-collection function It is

usually designed as spiral-collection overhung housing that

collects the air from diffuser and passed it to the pipe

system It has been recognized that the improvement of

centrifugal compressor performance requires a good

understanding of the flow mechanisms inside the volute

[7-9] especially the interaction among

volute-diffuser-impeller [10 11] The volute is mostly

designed in a way to shape a uniform circumferential static

pressure distribution both in volute and diffuser However

the volute acts as a diffuser at lower than design flow rate

and acts as a nozzle at higher than design flow rate

respectively A number of authors have researched this

subject It has already been confirmed that the asymmetrical

configuration has a significant impact on the flow field in

the diffuser as well as in the impeller [12 13] This

circumferential asymmetry has been recognized and

intensive experimental investigations of the flow within the

volute and the propagation of the distortion into upstream

components were carried out for subsonic compressor units

[14 15]

The work of Sorokes et al (1998) confirmed that the

pressure non-uniformity extended upstream of the impeller

implying that the impeller was subjected to varying inlet and

exit conditions deteriorating performance and especially

stability as the circumferential variations drove the

compressor into surge at higher flow rate as without the

volute [14] A three-dimensional unsteady analysis of the

flow in the impeller with circumferential distortion of the

outlet static pressure was investigated by numerical method

by Fatsis et al (1997) [16] The perturbation was thus

propagated upstream from the impeller outlet and influence

the incidence at the blade leading edges as well as other

flow parameters Gu et al (2001) [10 11] found that the

performance parameters of the single impeller passage differ

because of the asymmetric flow at the outlet of the impeller

There was almost no phase shift between the distortion in

the diffuser and impeller according to their results and it

was considered that unsteady effects of the volute-impeller

interaction can be neglected when the Strouhal number is

small enough

Little detailed measurement was carried out in the

impeller to investigate the asymmetric flow Further to the

authorsrsquo knowledge few research work focused on the

impact of the volute on the flow field in a

high-pressure-ratio turbocharger centrifugal compressor

The purpose of this two-part paper is first to understand the

asymmetry of flow field due to the asymmetric geometry of

the volute and subsequently to develop a novel asymmetric

flow control method to widen the stable flow range of a

turbocharger centrifugal compressor with




















2

5D

D 152

AR 2623 mm























compressor














for jet [18]

















































65000RPM (100N) 68000RPM (105N) will be
















































region




near surge


near choke



P

PPDs

max (1)



maxP

















































































near choke













C

fLSt (2)

















































follows

C

Lt (3)












near surge at N

42 Flow Simulation



















defined as follows

22

12 )(

U

TTCH p

(4)











pC 1T

2T

2U






































impeller

























































































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)




















2

5D

D 152

AR 2623 mm























compressor














for jet [18]

















































65000RPM (100N) 68000RPM (105N) will be
















































region




near surge


near choke



P

PPDs

max (1)



maxP

















































































near choke













C

fLSt (2)

















































follows

C

Lt (3)












near surge at N

42 Flow Simulation



















defined as follows

22

12 )(

U

TTCH p

(4)











pC 1T

2T

2U






































impeller

























































































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)









for jet [18]

















































65000RPM (100N) 68000RPM (105N) will be
















































region




near surge


near choke



P

PPDs

max (1)



maxP

















































































near choke













C

fLSt (2)

















































follows

C

Lt (3)












near surge at N

42 Flow Simulation



















defined as follows

22

12 )(

U

TTCH p

(4)











pC 1T

2T

2U






































impeller

























































































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)









65000RPM (100N) 68000RPM (105N) will be
















































region




near surge


near choke



P

PPDs

max (1)



maxP

















































































near choke













C

fLSt (2)

















































follows

C

Lt (3)












near surge at N

42 Flow Simulation



















defined as follows

22

12 )(

U

TTCH p

(4)











pC 1T

2T

2U






































impeller

























































































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)



near surge


near choke



P

PPDs

max (1)



maxP

















































































near choke













C

fLSt (2)

















































follows

C

Lt (3)












near surge at N

42 Flow Simulation



















defined as follows

22

12 )(

U

TTCH p

(4)











pC 1T

2T

2U






































impeller

























































































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)







































near choke













C

fLSt (2)

















































follows

C

Lt (3)












near surge at N

42 Flow Simulation



















defined as follows

22

12 )(

U

TTCH p

(4)











pC 1T

2T

2U






































impeller

























































































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)
















































follows

C

Lt (3)












near surge at N

42 Flow Simulation



















defined as follows

22

12 )(

U

TTCH p

(4)











pC 1T

2T

2U






































impeller

























































































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)








defined as follows

22

12 )(

U

TTCH p

(4)











pC 1T

2T

2U






































impeller

























































































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)




impeller

























































































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)








































ACKNOWLEDGMENT





REFERENCES















Engineering 221 (10) pp 1287-1294















Paper No 91-GT-7













123 (3) pp 475-483























Paper No 97-GT-50

















(1) pp 22-31





Chinese)

















Paper No 97-GT-50

















(1) pp 22-31





Chinese)


Documents

Stability Improvement of High-Pressure-Ratio Turbocharger …scholar.harvard.edu/files/jhuenteler/files/yang_et_al... · 2015-02-20 · STABILITY IMPROVEMENT OF HIGH-PRESSURE-RATIO