study on the simulation method of ...lin491/pub/glin-16-icef.pdfcompressor, the simulation method...

Corresponding author: [email protected] 1 Copyright © 2016 by ASME

Draft Proceedings of the ASME 2016 Internal Combustion Fall Technical Conference

ICEF2016

Oct 9-12, 2016, Greenville, SC, USA

ICEF2016-9420

STUDY ON THE SIMULATION METHOD OF COMPRESSOR AREODYNAMIC NOISE

BASED ON CFD AND IBEM

Liu Chen College of Power and

Engineering, Harbin Engineering University

Harbin, Heilongjiang, China

Cao Yipeng* College of Power and



Guang Lin School of Mechanical

Engineering, Department of Mathematics, Purdue University

West Lafayette, Indiana, United States

Sun Wenjian College of Power and



Zhang wenping College of Power and



Ming pingjian College of Power and



ABSTRACT Turbocharger compressor aerodynamic noise has been one

of the major noise sources of diesel engine. It is necessary to

study the characteristics of turbocharger fluid flow and radiation

noise for its effective noise control. In this paper, a new for

predicting compressor aerodynamic noise is presented, which

combined the calculated fluid dynamic (CFD) and indirect

boundary element method (IBEM. The unsteady viscous flow in

compressor was simulated based on the finite volume method.

In addition, the periodic pressure fluctuation of the rotor inlet

and blades were used to compressor radiation noise field

simulation by indirect boundary element method (IBEM). In

order to prove the feasibility of numerical simulation, the

acoustics experimental device for compressor aerodynamic

noise experiment was built and the sound pressure of

turbocharger were tested. The trend of simulation results and

amplitude level in blade passing frequency (BPF) coincide with

the experiment results. It indicates that the coupling method is

more effective and accurate in turbocharger noise prediction.

Key words: compressor; CFD; IBEM; aerodynamic noise

NOMENCLATURE

1 2 3, ,L L L the length of the reference body

V the inner area of the closed wall

V the outer area of the closed wall

Z compressor main blades number c sound velocity

d measuring distance

f sound wave frequency

BPFf blade passing frequency

k turbulent kinetic energy

0k sound wave number

n rotation speed

n the wall direction vector p the time harmonic external sound source

( , ,z)q x y volume velocity per unit volume density

0 static medium density fluid dynamic viscosity

angular frequency turbulent dissipation rate

a the boundary of indirect boundary element

region

INTRODUCTION Turbochargers are of great use in turbocharging diesel

engine of vehicles and power stations. The turbocharger

pressure ratio has been increasing due to the requirement of

higher engine power output, which causes the load of

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2 Copyright © 2016 by ASME

turbocharger components become higher as well as sound

pressure level of radiated noise of turbocharger. The excessive

noise has been one of the most important problems, which not

only increases the environmental pollution, but does harm to the

health of operators. Thus, it is necessary to study the noise

characteristics of turbocharger for the noise control.

Experimental and analytical methods have been presented

to study the characteristics of turbocharger noise and the effects

of structure parameters on the sound pressure level [1]. In

experimental study, the testing bench will be designed, built and

the accurate results can be obtained. Joachim once reported on

an experimental and numerical investigation aimed at

understanding the mechanisms of rotating instabilities in a low

speed axial flow compressor [2]. Raitor and Neise completed an

experimental study to explore the dominant sound generation of

the spectral components governing the overall noise level of

centrifugal compressors [3]. J. Galindo researched the influence

of tip clearance on flow behavior and noise generation of

centrifugal compressors in near-surge conditions, a correlative

experiment was completed to verify his numerical analysis

[4]. However, the experimental method will meet some

problems at the same time. The testing data contains large

amount information which is not related with the analysis and

need filter. The test is very tough when there are several

different working conditions, research objects, which may cause

time-consuming and laborious shortcomings, and so on. The

most important part of experimental method is that the

characteristic research of turbocharger is after the design

process.

The analytical method can be adopted for analyzing

different research objects, even the object that experimental

research can’t be carried out. The calculation method needs to

be carefully analyzed and selected, so that the result of the

calculation is sufficiently accurate. Sitki Uslu carried out

simulation of noise generation due to blade row interaction in a

high speed compressor [5]. Chu, S. and Dong, R. has done a lot

of work about unsteady flow, pressure fluctuation and noise

associated with the blade tongue interaction in a centrifugal

pump [6, 7]. The effects of structure parameters of compressor

on noise have been studied by various methods. Jeon calculated

the effects of design parameters on performance and noise of a

centrifugal fan [8], he also studied on the noise reduction

method of a centrifugal compressor [9]. Langthjem studied

flow-induced noise in a two-dimensional centrifugal pump [10].

Moreland studied the housing effects on centrifugal blower

noise [11]. Chehhat Abdelmadjid researched the effect of volute

geometry on the turbulent Air flow through the turbocharger

compressor [12]. Hyosung Sun had done a lot of work in

analysis and optimization of aerodynamic noise in a centrifugal

compressor, the simulation method (it is named traditional

method in this paper) can calculate the sound pressure level of

BPF noise and its harmonics [13, 14].

It is widely known that the turbocharger noise level is

dominated not only BPF noise spectrum, but continuous

broadband spectrum. An experimental result was selected as

example and its spectral components are decomposed, the

sound pressure level (SPL) of each component was showed in

Table 1.

Table 1. Spectral component decomposition

Spectral

component

BPF

spectrum

Continuous

broadband

spectrum

Full

spectrum

SPL(dB) 98.25 99.27 101.8

It can be found that the BPF spectral component can’t

represent the total noise level accurately, which means the

traditional BPF method is not comprehensive enough, more

effective methods should be proposed in the compressor noise

prediction. As mentioned above, the coupling method

combining CFD and IBEM is presented in compressor

aerodynamic noise prediction. On the basis of present research,

the prediction range was expanded from fixed frequency points

to the whole frequency band of interest. The feasibility of

numerical simulation method was verified by compressor

aerodynamic noise experiment.

1. NUMERICAL SIMULATION

1.1 Computational model The turbocharger aerodynamic noise can be divided into

two parts, the compressor noise and turbine noise. Because the

turbine is connected with exhaust pipeline which is always

connected with exhaust muffler, the turbine noise can’t radiate

to the air directly. The compressor noise is the main noise

source of turbocharger. In this paper, JTH150 turbocharger is

selected as the research object, there are 8 main blades and 8

split blades, as well as 11 diffuser blades. The three-

dimensional compressor model is shown in Fig.1 (a). The

calculation model of the compressor includes the following four

parts: inlet section, rotor section, stator section and volute

section. The single channel and periodic boundary are used in

rotor and stator flow calculation and the main calculation field

is divided into hexahedral mesh. Unstructured tetrahedral

meshes are used in the inlet section and the flow passage of the

volute section. The computational model used in numerical

simulation is shown as Fig.1 (b) and (c).

(a) The 3-D model


(b) Numerical simulation model

(c) Numerical simulation mesh

Figure 1. JTH150 compressor model

1.2 Flow field analysis Computational fluid dynamics has been quite mature in the

unsteady flow field analysis [15-17]. In this paper, in order to

adapt to the complicated flow field of the compressor and

ensure the accuracy of the flow field calculation result, the

implicit RNG k double equations turbulent model were

used. The governing equations and the parameters in the

equation are shown in Eq.(1)-(3).

i

k eff k

i j j

k ku kG

t x x x

(1)

2

1

2

i

eff k

i j j

u CG C

t x x x k k

(2)

2

0

1 1 3

1 2

1/2

0

0.0845, 1.39

1 /

1

1.42, =1.68

= 2

1

2

4.377, =0.012

eff t

t

k

ij ij

ji

ij

j i

kC

C

C C

C C

kE E

uuE

x x

(3)

In order to compare with the experimental results later, the

experimental boundary condition is used directly for the

numerical simulation, shown in Table 2.

Table 2. The calculating condition

Working

condition

Rotation

speed(rpm)

Mass

flow(kg/s)

Pressure

ratio

1 36913 0.69 1.67

The method presented in this paper should calculate the

pressure characteristic of compressor flow fluid, so the steady

flow field data would be calculated firstly, and then the data will

be used as the initial flow field to calculate the unsteady flow

field. The pressure fluctuation data on the rotor inlet surface

was extracted, and then the data can be used as sound source

information for further numerical calculation of aerodynamic

noise of compressor.

Figure 2-4 were the compressor pressure、temperature

and density contour, respectively. Figure 2 showed that static

pressure gradually increased from the compressor inlet to outlet.

At the same time, gas temperature in the compressor also

increased significantly, as shown in Fig.3; in Fig.4, it could be

seen that gas density distribution inside the compressor showed

a significant increasing trend from the compressor inlet to

outlet. These figures told that compressor could increase the gas

pressure and gas density in the process of centrifugal

compressor, which indicated the compressor performance

ability.

As the main parts of centrifugal compressor, the rotor

blade would transform the compressor mechanical energy into

gas pressure energy and kinetic energy. Under the centrifugal

force action, the gas rotates at high-speed with rotor, which

causes the gas pressure and density increasing, and the gas

kinetic engine also increases accordingly. Figure 5 showed that

the compressor flow speed is increased and reached its

maximum value at the outlet surface of rotor.

In Fig. 2, the static pressure on the inlet surface of the

diffuser is large smaller than the outlet surface’s pressure, and it

could also be seen that the static pressure on the surface of the

diffuser is significantly increased. With the increasing of the

volute flow channel cross-sectional area, the pressure amplitude

on volute surface is slowly increasing. The Fig.3 shows that gas

temperature has an increasing trend in vane diffuser and volute,

the peak value appears in volute.

Figure 2. Pressure contour


Figure 3. Temperature contour

Figure 4. Density contour

Figure 5. Velocity streamline

The main function of the flow passage is to convert the

kinetic energy of the air flow into the pressure energy, the

streamline form of gas in the flow passage of the volute is in a

spiral shape. The air flow velocity in the vane diffuser is

decreasing continuously, but the static pressure is increased

gradually.

1.3 Sound field analysis The boundary element method (BEM) is not based on the

compact sound source hypothesis, but uses the "singularity"

processing method, therefore, the boundary element method

does not need to distinguish the acoustic far field and near field.

Theoretically, it can solve the sound field of any position [18-

20]. The specific idea of BEM for solving the sound field is as

follows; the BEM method transforms the differential equation

of the computational acoustic field domain to the discrete

integral equation of the domain boundary, and uses the

weighted method to calculate the integral value on the

boundary.

In this paper, the Helmholtz equation is adopted in the

BEM method to solve the acoustic field. The steady sound

pressure in the acoustic region is generated by the contribution

of the time harmonic external sound source p , the

inhomogeneous Helmholtz equation is:

2 2

0 0, , , , , ,p x y z k p x y z j q x y z (4)

/ 2 /k c f c (5)

Figure 6. Indirect boundary element region

Two fluid regions are defined as shown in Fig.6. V is the

inner region of the closed wall, and V is the outer area of the

closed wall. Based on the front definition of the wall direction

vector n , the direction of the non-boundary area is positive.

The pressure and pressure gradient along the positive direction

of the wall are ap r and ap r n , on the contrary, the

pressure and pressure gradient along the wall of the wall

are ap r and ap r n .

A general indirect boundary integral formula is used, which

can be applied to the external sound field with open boundary

or the combination of both internal and external fields. The

formula can be written as:

,

,

a

a

a a a a

G r rp r r r G r r d r

n

(6)

In the Eq.(6), the single layer potential is the difference of the

pressure gradient on both sides of the boundarya :

a a

a

p r

n

p

n

rr

(7)

The double layer potential is the difference of the pressure on

both sides of the boundarya :

a a ar p r p r (8)

The single layer potential can be regarded as a monopole

sound source distribution of the boundary interface, and the

double layer potential can be regarded as the dipole source

distribution of the boundary interface.

The boundary element model of the inlet pipe and the rotor

of the compressor is shown in Fig.7 (a), you could have a more

direct view of the boundary element mesh from another


perspective in Fig.7 (b). In order to consider the acoustic

transmission without consider the acoustic reflection, rotor

outlet surface is set to a non-reflecting boundary.

(a) Mesh schematic diagram

(b) Section view

Figure 7. BEM mesh

Figure 8. Field point mesh

The field point mesh is established as a box shaped

envelope, shown in Fig.8. Nine sound pressure reference points

are selected on the surface or the edge of envelope. The sound

source information calculated by CFD is imported into the

sound pressure calculation process. The prediction process

based on CFD/BEM is shown in Fig.9.

Flow field simulation(time domain)

Sound field simulation

(frequency domain)

Compressor steady flow field simulation

Compressor transient flow field simulation

Compressor dipole sound source (CGNS)

Sound source information import and data transfer

Acoustic boundary definition

Sound field calculation and

analysis

Figure 9. Noise prediction process

The calculation formula of compressor blade passing

frequency is shown below:

60

BPF

nZf (9)

The compressor blade passing frequency and its first two

order harmonics are 4921.7Hz、9843.5Hz and 14765.2Hz,

shown in Table 3.

Table 3. BPF and its harmonics

BPF 1st harmonic 2nd harmonic

Frequency(Hz) 4921.7 9843.5 14765.2

The compressor sound pressure level will be calculated by

the simulation method presented in this paper. In the process of

numerical calculation, the sound field is predicted by the

surface dipole noise model, because the contribution of the

dipole is much larger than the monopole and the quadrupole.

The sound pressure spectrum of all reference points are shown

in Fig.10. It can be found that the sound pressure spectrum

presents a typical characteristic of the impeller machinery

aerodynamic noise. The dominating source mechanisms of the

compressor aerodynamic noise were BPF and its harmonics,

just like the experiment results listed in previous literatures. The

measuring point 1 which is located near the compressor inlet

also has the largest total sound pressure level. The measuring

point that is closer to the inlet of the compressor has the greater

sound pressure level. Because the computational time and

computer hardware limits, the noise spectrum of numerical

calculation seems rougher.

Non-reflecting

Boundary


0 2500 5000 7500 10000 12500 15000 17500 20000 22500

0

20

40

60

80

100

SP

L/d

B

f /Hz

P1

P2

P3

P4

P5

P6

P7

P8

P9

4829.1Hz9286.8Hz

13744.4Hz

Figure 10. Numerical simulation sound pressure level

The sound pressure distribution at inlet of turbocharger is

shown in the Fig.11. The pressure distribution of compressor

nozzle showed different patterns at each formant of Fig. 10.

(a) 995Hz

(b) 3096Hz

(c) BPF

(d) 1st harmonic

(e) 2nd harmonic

Figure 11. Pressure distribution of compressor nozzle

It can be found in Fig.11 that the sound pressure

distribution are significantly different at different frequencies,

which is decided by the sound wavelength especially in low

frequency. Compared with the wavelength, the size of inlet pipe

is smaller, and the sound wave diffraction phenomenon will

occur at orifice. The pipe diameter in this paper is 0.141m, so

the corresponding frequency which is equal to the acoustic

wavelength is as follows,

2432c

f Hz

(10)

In Eq. (10), 343 /c m s , 0.141m . In fig. 11 (a), the

sound wave propagate in the form of spherical wave. When the

frequency is higher than 2432Hz, the acoustic wavelength will

be smaller than orifice size, orifice acoustic diffraction

phenomenon gradually disappeared and sound wave reflection

phenomenon occurred at pipe orifice.

2. EXPERIMENTAL STUDY

2.1 Experiment facility For testing the validity of simulation method presented in

this paper, the turbocharger noise experiment bench was

designed and built, shown in Fig. 12. The gas provided by the

high-pressure air compressor will be burn fully in the burner.

The high temperature, high pressure gas obtained in the burner

would be used to simulate the diesel exhaust, and the gas is

leaded into the turbine and drive the turbine rotating. At the

same time, the coaxial compressor was driven to rotate by

turbine. There is a front exhaust value on the pipeline used to

adjust the amount of gas entering the turbine to control the

rotational speed of the turbocharger. The opening degree of the

electric control valve can be changed by the control device to


change the gas intake of the turbine easily. The experimental

pressure ratio adjustment can be achieved by changing the

compressor outlet back-pressure through the butterfly regulating

value [3, 21]. Pressure, temperature and vibration sensors are

arranged in the turbine and compressor connecting pipe, the

operating parameters of the compressor are measured and then

used to guide the numerical calculation.

Figure 12. Experiment facility

Distribution details of the turbocharger and pipes used for

experimental measurements are shown in Fig 13. The

turbocharger is fixed on the bracket, turbine outlet is connected

with the exhaust muffler to reduce the exhaust noise,

compressor inlet is open to air directly.

Figure 13. Turbocharger detail diagram

The noise measuring points are distributed as shown in Fig

14(a). The turbocharger is simplified as a cube and its size is

0.4 0.4 0.4m m m . In Fig.14 (a), the negative X axis direction

was the air-inlet direction of the turbocharger. Measurement

points 1-5 were the middle points on each surface, and points 6-

9 were edge points respectively. To get the coordinates of each

monitoring point, we computed them according to the length

shown in Fig.14 (b) and Fig.14 (c). The distance of reference

points are shown as Eq.(10) and Eq. (11).

1 2 3 0.4L L L m (10)

1.0d m (11)

The other distance parameters in the measurement surface

are calculated by the following formula.

1

2

La d (12)

2

2

Lb d (13)

3c L d (14)

2

ch (15)

(a)

(b)

(c)

Figure 14. Measurement points distribution

2.2 Experimental results The sound pressure spectrum of each measurement point is

shown in Fig.15. It shows that the BPF and its harmonics

dominate the compressor aerodynamic noise. Its first two order

harmonic calculated by (9) are 4921.7Hz、 9843.5Hz and


14765.2Hz, accordingly. The corresponding frequencies of the

experimental measurements are 4912.0Hz 、 9824.0Hz and

14736.0Hz. Two sets of data are compared in Table.4 and the

results are nearly same. The slight error might be the sensor

error or the instability of the compressor operation.

Table 4. Frequency comparison


Theoretical

frequency(Hz) 4921.7 9843.5 14765.2

Experimental

frequency(Hz) 4912.0 9824.0 14736.0

error 0.20% 0.20% 0.20%

0.0 5.0x103

1.0x104

1.5x104

2.0x104

20

40

60

80

100

SP

L /dB

A

P1

(a) Measuring point1

(b) Nine measuring points

Figure 15. Experimental sound pressure level

Fig. 15 shows that the trend of all reference points is same.

The measuring point 1 which located at the compressor inlet

side has the largest sound pressure level. The measuring points

that are closer to the inlet of the compressor have the greater

sound pressure level.

2.3 Validity of calculation method All the information used in CFD and IBEM calculations

were got from the experiment measurement. The rotation speed

of experiment condition used in this paper is 36913 rpm. The

numerical calculation of centrifugal compressor aerodynamic

noise was completed with the pressure fluctuation data on the

rotor inlet surface and blade surface that got from flow field

analysis.

In Fig.16, three method result are plotted together, which

are traditional method (dots), CFD/BEM method (solid black

line) and experimental method (solid red line).

The sound pressure level of the BPF and its first two order

harmonic frequencies that are calculated with the pressure

fluctuation data of the blade surface is shown as the three dots.

The sound pressure level on the fundamental frequency and the

harmonic frequencies agree very well with the experiment

result. This method is much less time-consuming, in the

industrial design and the preliminary study of noise control, it

can play a very important role.

The fundamental frequency of experiment and calculation

by presented method in this paper are respectively 4912.96Hz

and 4829.12Hz, and the corresponding sound pressure value are

96.32dB and 96.54dB, indicating that the numerical calculation

result are predicted with sufficient accuracy. Table.5 show a

more detailed comparison. The errors are all calculated with

numerical calculation results and experimental results.

The results calculated by presented method agree well with

the experimental results, which are shown in Fig.16 and Tab. 5.

The trend of sound pressure curve is nearly envelop line of

experimental curve. The difference of peak value at blade

passing frequency is same. The error is not too much under 1st

harmonic frequency. With the calculating frequency is

increasing, the error of two method will be bigger.

0.0 5.0x103

1.0x104

1.5x104

2.0x104

20

40

60

80

100

SP

L /dB

f /Hz

tranditional simulation

continuous spectrum simulation

experimental measurement

Figure 16. Comparison of calculation and experimental results

Table 5 Results comparison


Theoretical

frequency(Hz) 4921.7 9843.5 14765.2

Calculation

frequency(Hz) 4829.1 9286.8 13774.4

Experiment

frequency(Hz) 4912.0 9824.0 14736.0

0.0 5.0x103

1.0x104

1.5x104

2.0x104

40

50

60

70

80

90

100

SP

L/d

B

f /Hz

P1

P2

P3

P4

P5

P6

P7

P8

P9

4912Hz9824Hz

14736Hz


Frequency

error 1.69% 5.47% 6.53%

Calculation

SPL(dB) 96.5 92.9 86.3

Experiment

SPL(dB) 96.3 93.6 80.5

SPL error 0.21% 0.75% 7.21%

3 CONCLUSION The coupling method combining CFD and IBEM in

compressor aerodynamic noise prediction is presented in this

paper. The experimental bench is built and the validity of the

method is tested.

(1) The calculation results for sound pressure in the whole

frequency band are agree very well with the experiment

result. The trend of curve is nearly same.

(2) The sound pressure near the inlet of turbocharger is higher

than other position. As the distance from the inlet

increasing, the sound pressure is smaller. Some devices like

intake muffler can be the effective way to control the

turbocharger noise.

(3) Comparing with the existing method, the simulation

frequency band expands from fixed BPF points to the

frequency range. More information such as fluid

characteristic, all band sound pressure and sound

distribution can be obtained in the same time, which can be

used for noise prediction and noise control.

ACKNOWLEDGMENTS The authors acknowledge the support for this research

work by Research Institute of Power Engineering Technology of

Harbin Engineering University.

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