Simulation Methodologies — noise from axial cooling fans

Deliverable no.

E. D16

Dissemination level

Public

Work Package

WP E.4 Cooling and Auxiliary systems

Author(s)

Christophe MALICZAK, Pascal BOUVET (VIBRATEC), Serguei Timouchev (KTH)

Task Responsible

Mats Åbom (KTH)

Status (F: final, D: draft)

D2 (31 Aug 2006)

File Name

SILENCE_E.D16_KTH_VTC_D1.doc

Project Start Date and Duration

01 February 2005, 36 months

TABLE OF CONTENTS

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__________________________________________________________________________ SILENCE report: E.D16 - Simulation 3

methodologies

1 Introduction The SILENCE project is dedicated to the reduction of the noise emitted by transportation device. All types of sources are considered. The Work Package E4 deals with the noise generated by cooling and auxiliary systems for railway vehicles. These systems are a major source of noise at standstill and during the starting process. For instance: • • • • • • • • •

Traction motor cooling system Diesel engine cooling system Converter cooling system Brake resistor cooling system Transformer cooling system Compressors Pumps HVAC (heat ventilation air conditioning unit) Machine room ventilation

The diesel engine cooling fans are proven to be the main acoustic source at standstill for a DMU train. Figure 1: Microphone array measurement at standstill on DMU train (source: BOMBARDIER).

Diesel engine cooling fans

Diesel engine

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methodologies

1.1

Modelling of the cooling fan noise

Two types of calculation models will be tried during the project, see sections 1.2 and 1.3. The first step of the project is to implement the fan models based on, e.g., the geometrical description of the blades and guide vanes. In the second step the models will be validated by mounting a single fan unit in well defined test configuration, see Figure 2, and making a detailed mapping of the inflow as well as the radiated noise. The last step consists in using the analytical tool to propose some solutions to reduce the noise emitted by the cooling unit. The present report covers the first step, i.e. a description of the two methods including the underlying theory and some numerical examples. In these examples a preliminary flow measurement of the free fan has been carried out, but a full validation of the methods will follow in a subsequent report. Figure 2: Axial fan mounted in the wall between two rooms at KTH.

R >2 d

d

1.2

The rotating dipole model

VIBRATEC has developed a fan noise model based on rotating dipoles created by the unsteady blade forces. This model requires that the inflow distribution to fan unit is specified. Only the harmonic part of the acoustic sources is considered.

1.3

The 3D Acoustic-Vortex method

KTH has in co-operation will Prof. Serguei Timouchev tested a hybrid approach meaning that the acoustic source data will be extracted from a computation of the usteady flow. Also this model focuses on the discrete noise emitted at the blade passing frequencies of the fan.

1.4

Content of the report

Chapter 2 is devoted to the description of the fan unit and of the parameters of a single fan. Chapter 3-4 describes the background of the VIBRATEC model. It also deals with the important input parameters of the model and gives an example of noise maps using artificial input data. Chapter 5-6 describes a CFD based model developed by Prof. Serguei Timouchev and his co-workers. It starts with an introduction in Chapter 5 which also presents the underlying theoretical ideas to split the acoustical and flow field. Then in Chapter 6 the results of the first simulations are presented. __________________________________________________________________________ SILENCE report: E.D16 - Simulation 5

methodologies

2 General considerations on the cooling unit 2.1

Description of the cooling unit

The cooling unit consists of two fans mounted in a metallic structure located on the roof of the train. The air enters the unit by the flanks (blue arrows) and goes out through the fans in the vertical direction (red arrows). Picture 1: View of the fan unit.

Picture 2 : Single fan upper view.

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2.2

Description of the fans

The cooling unit has two axial fans which are supposed to be identical (see Picture 2). Main parameters of the fan unit: • Maximum rotational speed: 2100 rpm (for the validation tests on a single unit the speeds 950 and 1900 RPM will be used). • Fan diameter: 720 mm • Cylindrical housing (duct) length : 200 mm Rotor : • Number of blades : 8 • Blade inner diameter : 320 mm • Blade outer diameter : 716 mm • Blade chord : 125 mm •

Blade angle (at blade root and relative to fan axis) : αR = 60 degrees

•

Blade pitch or twist angle ( root to tip) : αT

•

Blade thickness : 4 mm at the edges and 9 mm in the middle

- αR = 11 degrees

Picture 3 : Detailed view of a rotor blade

αR

Blade chord

αT

Blade span length

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methodologies

Picture 4 : Detailed views of a stator vane

Stator chord

βR

βT Stator : • Number of outlet guide vanes (stator) : 7 •

Stator angle (at root and relative to fan axis) : βR = 30 degrees

•

Stator pitch or twist angle (root to tip) : βT

• • •

Stator chord: 100 mm at root and linearly decreasing to 70 mm at tip. Stator thickness : 5 mm Separation rotor-stator : linearly increasing from 15 to 80 mm (root to tip)

- βR = 3 degrees

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Figure 3 : Cross view of the fan unit

Separation rotor-stator

60° Rotor

Stator 30o

The main acoustic sources are: • •

the interaction between the rotor and the velocity distortion of the upstream flow, the impingement of the impeller wakes on the stator vanes, at the outlet of the rotor.

__________________________________________________________________________ SILENCE report: E.D16 - Simulation 9

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3 The rotating dipole model 3.1

Theoretical background

The formulas and notations are the ones used by M. ROGER from Ecole Centrale de Lyon in his aeroacoustics courses (see Reference [4]). For both sources, the formula of the noise radiated by a rotating dipole in free field and far field conditions is used:

~p( x , ω ) =

iω 8π 2 c 0

∞ −∞

F .R′ R2

e i ω ( t + R′ / c0 ) dt ′

(3.1)

With: • • •

ω is the angular frequency c0 is the sound velocity R and R’ are geometrical parameters describing the position of the observer and the blade segment, R’ being a function of time (see Figure 4)

•

F (t ) is the local instantaneous lift force

Figure 4: Frame of reference used in the calculations, isolated compact blade segment, axial flow rotating blade technology (from [2]).

x = ( R,θ ,φ )

ez

R F (t ′)

γ

R′

θ

ey R0

M

ex

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Mathematical developments and the introduction of Bessel functions lead to the following expressions for the two sources: 1. for the interaction between the rotor and the velocity distortion of the upstream flow At the mth harmonic, the full acoustic pressure from discrete tones for the inlet source can be written (from [2]):

p mB ( x , t ) =

i mB 2 Ω 4π c 0 R

cos γ cos θ −

∞

F s e − i ( mB − s ) π / 2 J mB − s ( mB M sin θ )

s = −∞

mB − s sin γ mB M

e i ( mB − s ) {ϕ − Ω s ( t − R / c0 )}

(3.2)

Where: − − − −

R, θ, ϕ are the spherical coordinates of the observer point M is the local Mach number Ω is the rotational speed of the machine B is the number of blades of the rotor

−

Ωs =

− −

γ is the stagger angle of the rotor Fs is the loading harmonic of order s on the blade

mB Ω mB − s

2. for the impingement of the impeller wakes on the stator vanes, at the outlet of the rotor The discrete tone noise for the rotor-stator source can be written (from [2]):

i m BV Ω p mB ( x , t ) = FmB 4π c 0 R cos γ ′ cos θ +

∞

e − i ( m B − s V ) (α 0 + π / 2 ) J mB − sV ( mB M sin θ )

s = −∞

mB − sV sin γ ′ e i ( m B − s V ) [ϕ − Ω sV ( t − R / c0 ) ] mB M

(3.3)

Where: −

FmB

is the loading harmonic of order (mB) on the vanes

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−

γ ′ is the stagger angle, defined in the opposite direction with respect to the stagger angle of the rotor γ

− −

α0 is the angular location of the 1st stator vane in reference to the first axis. V is the number of vanes of the stator

These expressions are valid for compact blade segments. But in the study case, the blade is not compact anymore from the acoustic point of view from the 3rd harmonic. Indeed, for a given harmonic n of the BPF, the wavelength m is given by the following formula:

λm = With: • • • •

c0 2 π c0 = f n mBω

(3.4)

c0, the sound velocity n, the harmonic considered B, the number of rotor blades ω, the angular frequency of the rotor

Assuming a rotational speed of the rotor of 2000 r.p.m, the following values are obtained: Table 1 : Acoustic compactness assumptions for the rotor Harmonic order, m

1

2

3

4

5

6

7

fm

267 Hz

533 Hz

800 Hz

1067 Hz

1333 Hz

1600 Hz

1867 Hz

0.098

0.196

0.294

0.392

0.490

0.588

0.686

0.155

0.310

0.465

0.621

0.776

0.932

1.09

chord

λm span

λm

For harmonic n°3 for the chord and n°4 for the span, the blade is not compact from the acoustic point of view. Consequently, the expressions in Equation 3.2 and Equation 3.3 need to be integrated in the span-wise and chord-wise directions.

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3.2

Description of critical input data

The critical data is the loading harmonics on the rotor blades and on the stator vanes, Fs or FmB in the previous formulas. It must be noticed that, for both sources, the harmonic which has the main impact on the radiated pressure is the (mB)th harmonic. Therefore, the loading harmonics have to be assessed with a good level of accuracy. Several means can be used to assess the loading harmonics: CFD calculation, measurements, or analytical models. In this study, it is chosen to use analytical models. 3.2.1 Assumptions of the analytical models Assuming small camber and small thickness of the airfoil (blade or vane), an analysis of a small perturbation of the steady-state equation relating the lift force per unit span force to the mean flow velocity shows that the unsteady loading on the airfoil is mainly due to the transverse velocity fluctuations.

Figure 5: Small disturbances of the flow incident on an airfoil.

Figure 6: The two-dimensional Sears’ or Amiet’s problem.

dF U0 Convection speed

w

−c / 2

c/ 2

x1

The analytical models used in the study (Sears’ or Amiet’s theory) describe the behaviour a thin plate, in a two-dimension coordinate system, for transverse harmonic excitations (a Fourier transform of the data is required). Regarding the frequency domain, Sears is devoted to low frequency problems; Amiet is more adapted for high frequency problems. The high frequency assumption to use Amiet’s formula is given by:

c

β

2

>

λm

(3.5)

4

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With: • c, the chord length • β² = 1- M², where M is the Mach number of the average flow • λm, the acoustic wavelength for harmonic m Assuming a rotational speed of the rotor of 2000 r.p.m, this high-frequency criteria is met from m = 3 for the rotor blade and from m = 5 for the stator blade. The rotational speed, the chord length and the blade number of the rotor being known, the high frequency criteria of Amiet’s theory is reached from the 3rd harmonic (880 Hz). With the frequency sensitivity of the human ear, the frequencies around 1000 Hz are more critical. Consequently, Amiet’s formulation is used in the model. 3.2.2 Amiet’s formulation To asses unsteady loads applied on the blades, high frequency Amiet’s theory is used (see [2]). This approach can be used for: • a thin plate • a compressible fluid • high frequencies Amiet’s theory allows taking account of the phenomena located on the leading edge and on the trailing edge of the blade. This model allows assessing Fourier coefficients of the pressure load p(x) along the chord knowing the Fourier coefficients of the transverse velocity. The pressure in free field, far field can then be calculated using Equation 3.2 and Equation 3.3. Notations The following notations are used:

chord 2

•

b=

• • • • •

U : velocity of the mean flow M : Mach number of the mean flow w : transverse velocity regarding the blade l : lift force per unit length (in the span-wise direction) ρ : density

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• • • •

B 2 β² = 1 − M ² ω = 2πΩ

nωb U Mk n µn = β² kn =

Formulation In Amiet’s theory, input data is the transverse velocity at the leading edge of the blade, the time dependency of which is sinusoidal. In the axial fan case, the transverse velocity is periodic: it can be written as a Fourier series. Amiet’s theory can then be used on each component. w(t) is the transverse velocity, and W n are the corresponding Fourier coefficients : +∞

=

− ω

= −∞

ω = π

π ω

ω

(3.6)

l(t) is the lift force per unit length, and Ln are the corresponding Fourier coefficients :

=

+∞

− ω

= −∞

=

ω π

π ω

ω

(3.7)

Amiet’s theory allows assessing the pressure load distribution along the chord. The parameter x is divided by b to make it unit-independent. The Amplitude of the nth harmonic of the pressure load on the blade according to Amiet’s theory is

=

−

ρ π

+

ρ π

{− −µ

+

−

µ

−

}

−µ

−

−π

(3.8)

−π

+

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where E is the Fresnel integral :

=

π

* indicates the complex conjugate operator. This Pn,Amiet allows to assess FmB and Fs which appear in Equation 3.2 and Equation 3.3. A and B respectively correspond to trailing edge boundary conditions and to leading edge boundary conditions. If the blade is acoustically compact in the chord-wise direction, this expression can be directly integrated to obtain the load. As the blade can be considered as a simple dipole, this load suffices to know the acoustic source. For the diffuser and impeller sources, the blade is not acoustically compact. A direct integration is not possible. The x-dependency of p(x) has to be taken into account. The integration over x has to be done in the last step.

Simplification This formula is heavy to use in the remainder of the study. Relative magnitudes of terms A and B (see Equation 3.8) can be compared. Term A appears negligible at the first order .The distribution of the pressure on the blade is given by the following simplified formula :

=−

ρ

−µ

π

−

−π

+

(3.9)

All parameters are known, except Fourier coefficients of the transverse velocity, wn.. These coefficients will be assessed using CFD data measured at KTH, and also assumptions on the directions of the velocities, which is usually called the velocity triangle.

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Figure 7: Wake-interaction noise mechanisms (visualization of velocity triangle)..Top: rotor-stator interaction, Bottom: stator-rotor interaction

Ω R0

Vr

IGV

Vr

3.2.3

- ΩR

0

Implementation of the formula

Description of the procedure: • • •

Each blade is divided into radial segments (for instance, 4 segments in Figure 8). The Amiet’s formula is dedicated to two-dimensional plate, thus, the blade segments do not have to be too short regarding the blade span length. For a given segment, the blade loading is assessed on the segment using Equation 3.9. The rotating dipole formula is applied to each segment. Figure 8: Sampling of the blade to assess acoustic sources.

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This procedure is implemented using Matlab software. For instance, on the rotor source, after analytical calculation for the integration over the chord, the pressure for the mth harmonic is given by:

p mB Rtip

Rroot

imB ΩR c0

imB ² Ωe = 4π c0 R U (r )ws (r )

k s (r ) × M (r )

× (− ρ 0 ) × b × e

−i

π 4

× 2×

s = +∞ i ( mB − s ) ϕ − π 2

e

× ...

s = −∞

× ...E * [2 µ s (r ) × (1 − M (r ))] × J mB − s (mBM (r )sin θ ) × cos[γ (r )]cos θ −

mB − s sin[γ (r )] mB M (r )

The infinite sum is in fact a finite one due to the quick decreasing of Bessel functions values. The integration over the span (from root to tip) is carried out with a rectangle method.

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4 First calculation results (rotating dipole model) A few calculations are carried out for artificial transverse velocity values. For instance, the noise map of the rotor/stator interaction source is given in for the 3rd harmonic. Z, the vertical axis, is the axis of the fan. Figure 9: Noise map of the rotor/stator interaction noise for the 3rd harmonic.

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5 The 3D Acoustic-Vortex Method With recent technology mature, designers need a better tool to predict the performance, efficiency and noise of ventilation systems across a wider range of operating conditions. The use of a combined CFD-CAA1 approach is an increasingly powerful tool that enables designer to better understand all features of unsteady flow in fluid machines and to find the optimal solution regarding energetic characteristics and lower pressure pulsation and noise. In particular it is hoped that this techniques will make it possible to predict the effect of both tonal generation due to disturbed inflow conditions as well as broad-band noise from e.g. inflow turbulence and turbulent boundary layers acting at the blade trailing edge. However in the present study the focus will be on the harmonic part. Because still the prediction of the broad-band part is not realistic for real cases and for practical applications the tonal content is normally dominating due to the disturbed inflow conditions. During the period 2000-2005 Prof. Timouchev2 and co-workers have completed the development of a 3D CFD-CAA method for the prediction of blade passage frequencies (BPF) in fluid machines with an axial or diagonal flow pattern. It is known that application of common CFD codes for compressible medium proves to be ineffective for design optimization problem especially regarding the acoustical part of pressure pulsation field. The developed method [5-9] is based on splitting the equations of compressible medium dynamics into vortex and acoustic modes. As a result the whole processor time for both modes of oscillations is reduced and the accuracy of prediction for the acoustical mode is improved. In the present report the method is outlined and the first results of applying the method to the axial fan studied in the SILENCE project are presented.

5.1

Definition of the Problem

The problem is the study of the general features of unsteady flow and pressure pulsations in a train axial cooling fan. The main goal of the work is to establish efficient tools and procedures for obtaining computational results on pressure pulsations and noise generated by such a fan. In particular the ability to predict the effect of in- and outflow (guide vanes) disturbances on the sound generation is of interest.

5.2

Input Data Specification

The 3D geometry of the fan was contoured using geometrical input data from KTH. To validate the geometry computational tests using an operational state defined by the fan manufacturer (data supplied by Bombardier) was used: at fan RPM 2015, air temperature 312 K and ambient pressure 0.877 Bar the volume flow rate must be 8.63 m3/s. The computational test gave a value very close to this volume flow (-0.35% off). This shows the 3D geometry of the fan is well produced and should give reliable data regarding pressure pulsations and noise.

5.3

3D assembly of the fan and computational domain

The assembly geometry is defined by the test rig set up at KTH for the single fan, see Picture 1-4. The fan casing is installed into the thick wall and bounded by inlet and outlet air semi

1

“Computational Fluid Dynamics - Computational Aero-Acoustics”

2

Director General, Intellectual Reserves International Ltd., Moscow RUSSIA.

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spheres of radius that is higher than two fan diameters. At the inlet and outlet boundaries of the computational domain the specific impedance is equal unity (“normalized”). There are some simplifications in the flow part geometry aimed to reduce necessary computational resources and processing time: no disturbances at the fan inlet, no radial gap between the rotor and casing and a motionless wall boundary condition. The above simplifications will influence the prediction of absolute values of pressure pulsations. Anyway the model built can still serve as a reliable base for the study of pressure pulsations against main fan parameters as it is specified in the problem definition. The computational domain consists of three main regions, namely, inlet region, rotor and stator with the outlet region. Figure 10: View of the 3D computational domain.

The computational domains are connected to each other through the so-called “sliding-grid” interfaces that provide the data exchange between the rotor region and the motionless regions.

5.4 Governing equations In the development of the physics and mathematical model of the pulsating flow in a fan one have to account for the non-linear character of the sound generation process as well as the wave character of its spreading in the flow. The following assumptions are made: -Subsonic flow; -Isentropic flow; -Viscous diffusion is neglected for acoustical waves; -Acoustic oscillations (velocities of acoustic motion due to the fluid compressibility) are small in comparison with the vortex perturbations (velocities of swirl and translation motion of the absolutely incompressible medium); For prediction of pressure pulsations and noise in the fan computational domain, the mathematical model is based on a representation of the fluctuating flow velocity field V as a combination of vortex and acoustic modes,

V = U + ∇ϕ = U + Va

(5.1)

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Where U - Velocity of transitional and rotational motion of an incompressible medium (vortex mode)

Va

- Velocity of pure deformation (the acoustic mode)

ϕ - Acoustic potential

This gives the acoustic-vortex wave equation in terms of enthalpy oscillation i in the isentropic flow of a compressible medium (a – speed of sound assumed constant):

1 ∂ 2i − ∆i = ∇(∇( 1 2 U 2 ) − U × (∇ × U )) 2 2 a ∂t

(5.2)

The right hand side of this equation represents the source function, defined from the velocity field of vortex mode flow. It is determined from the solution of the unsteady equations of an incompressible medium with the appropriate boundary conditions. There is a simplification in eq. 5.2 with regard to convection of acoustical perturbations by the mean flow. For higher frequencies it will be more accurate to use the convective operator ∂ ∂t + U ⋅∇ or substantial derivative instead of ∂ ∂t , where U is the vector of the vortex mode velocity. By using a local complex specific acoustic impedance Z, the boundary condition for the acoustic mode can be put for each BPF harmonic in the form

∂ (ik − I k ) k ∂ (ik − I k ) =− ∂n ∂t a Zk

(5.3)

where k is the number of BPF harmonic, n – normal direction to the boundary, I – enthalpy oscillations due to the vortex pressure perturbations of pseudo-sound mode. The mathematical model of incompressible fluid flow is based on the Navier-Stokes equation

∂V ∇P 1 + ∇(V ⊗ V ) = − + ∇(( µ + µ t )(∇V + (∇V ) T ) + F ∂t ρ ρ

(5.4)

by taking into account the continuity equation for incompressible flow ∇U = 0 ; The standard k − ε model of turbulence [10] is used to determine turbulent viscosity via the relation

µt = Cµ ρ

k2

ε

(5.5)

The following equations are used for kinetic energy and dissipation:

µ ∂k 1 G + ∇(Vk ) = ∇(( µ + t )∇k ) + − ε ∂t ρ σk ρ

(5.6)

µ ∂ε 1 ε G + ∇(Vε ) = ∇(( µ + t )∇ε ) + (C1 − C 2 ε ) ∂t ρ σε k ρ Where G = µ t

(5.7)

∂ Vi ∂ Vi ∂ V j ( + ). ∂ x j ∂ x j ∂ xi

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The turbulence model constants used are

σ k =1.0; σ ε =1.3; C µ =0.09; C1 =1.44; C 2 =1.92 Initial values of kinetic energy and dissipation are calculated automatically during the first step of computation. The Navier-Stokes equations are solved by a splitting method with the implicit algorithm and high-order numerical scheme for convective transfer terms. The advantage of the method is that it is possible to reduce processing time by making computation on a rough grid, refining the grid when approaching to the convergent oscillatory solution. The iterative procedure of the vortex mode equations goes up to a convergent periodical solution and subsequent definition of the source function of the equation (5.2). Initial condition of the vortex mode flow is zero pressure and velocity in the entire computational domain. At the outer boundary freeoutlet flow condition is used with linear extrapolation of velocity from inner nodes. Boundary condition on the wall is represented by a numerical equivalent of the logarithmic

U

law for the tangential component τ of the velocity vector. It can be found by the following expression

y+, y + < y*+ Uτ = 1 ln(Ey + ) y + ≥ y*+ U*

(5.8)

κ

τw , ρ

U* =

where U * is the friction velocity, τ w is the wall shear stress and the turbulent boundary layer kinematical parameters are defined by the following formulas

u+ =

y+ =

1

κ

ln (Ey + )

ρU * h µ

(5.9) (5.10)

+ where κ = 0.41 , E = 9 , y* ≈ 11 and h -- distance from the wall to the center of the boundary grid cell.

5.5 Methodology of computations The solution of the acoustic-vortex equation is divided into two steps: 1) Computation of the incompressible flow for the determination of the source function; 2) Solution of the inhomogeneous wave equation, eq. 5.2. The pressure pulsation field can be represented like a sum of acoustic and vortex oscillation. The wave equation is solved relatively to pressure oscillation using an explicit numerical procedure. Zero pressure pulsations is an initial condition for solution of the wave equation. The local complex specific acoustic impedance is used to define boundary conditions for the acoustical part of the pressure field. The 3D numerical procedure is based on a non-staggered Cartesian grid with adaptive local refinement and a sub-grid geometry resolution method for description of curvilinear complex boundaries. Grid generation procedure produces a rectangular grid with local multilevel __________________________________________________________________________ SILENCE report: E.D16 - Simulation 23

methodologies

adaptation. The grid is automatically adapted in the casing region resulting in more accurate simulation. The original cell is divided into 8 equally size cells (1st adaptation level). Furthermore the resulting cell can be divided again (2nd adaptation level) and so up to the required level of accuracy. The sub-grid resolution method is used ‘to fit’ the Cartesian grid to the geometrical boundary in order to accurately describe the boundary conditions. It is especially important for the blade that represents a relatively thin and curved surface. When an initial “parent” rectangular cell is cut off by the curvilinear surface then it is disjoined into new volume elements formed by the facet surface and the original grid cell faces. During computation of the incompressible medium flow the BPF amplitudes of source function of equation (5.2) are accumulated in each grid cell. On the final stage the acousticvortex wave equation (5.2) is solved for each BPF harmonic. In both stages iterative procedure goes up to convergence to a “steady” periodical solution. The reference level for the pressure is defined as 101000Pa. Thus in computations and resulting data the pressure is defined from the reference level. Initial conditions are zero pressure and velocity in the whole domain.

6 Computational Tests (3D Acoustic-Vortex Method) 6.1 Computational grid Computational tests are completed on a grid of the fourth level of adaptation in the zone of rotor and third level in the zone of stator vanes. Figure 11: Computational grid. Meridian view.

Figure 12: Zoomed computational grid.

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In the vortex mode computation at the inlet and outlet boundaries the pressure is defined zero and velocity is obtained by interpolation from inner nodes. The computational grid is shown in Figure 11-13.

Figure 13: Computational grid. Plane view.

6.2

Computational Results

The first computational data presented below mainly concerns the fan operation mode at RPM 1900. A comparison of pressure pulsation and noise is made between the operation modes RPM 1900 and RPM 950. Representation of the computational data is made in a virtual domain shown in Figure 14. In the virtual domain all three parts are connected to each other by sliding interfaces giving an image of the physical domain. Figure 14: Meridian section of the virtual computational domain.

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Below are presented flow parameters in the meridian section. The first task is to prove that the fan computational modeling gives the correct flow parameters behavior. Figure 15: Instantaneous distribution of total pressure (Pa); RPM 1900.

Distribution of total pressure shows that this machine works as a compressor increasing the energy of air. One can note that using the relative pressure level bring negative values of static and total pressure but in such a computation only the pressure differential has an importance due to uncertainty in boundary values of pressure. At the fan outlet there is a high velocity jet going to the outlet boundary of the computational domain as is clear from Figures 16 and 17. Figure 16: Instantaneous velocity vectors colored by velocity modulus, (m/s); RPM 1900.

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Figure 17: Instantaneous axial velocity map, (m/s); RPM 1900.

There are some back flow phenomena in the outlet domain due to a non-optimal aerodynamic configuration of the fan exhaust and air ejecting along the jet boundary. Such outlet jet phenomena could bring difficulties into the noise measurements and also make accurate modeling of the BPF noise generation more difficult. The most important effect that determines the amplitudes at the BPF is non-uniform flow at the fan inlet. The inflow parameters distribution presented in plane 4, see Figure 18, shows a circular non-symmetry of the flow field that is the main cause of the existing BPF pressure pulsations. What is important to note is that near the impeller inlet there is an essential circumferential non-uniformity of flow parameters that leads to pressure pulsations as it rotates with rotor. Figure 18: Position of plane 4 (left bound of the picture) corresponding to the beginning of the rotor inlet hub cone.

Instantaneous map of the axial velocity component shows an 8th order circular symmetry linked with the 8 rotor blades. An unsteady pressure field and noise is generated at the blade passing frequencies due to the rotation of this inflow distribution with the rotor. Following the velocity non-uniformity one can estimates the amplitude of pressure pulsations near the rotor blades to be of the order 8—25 Pa.

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Figure 19: Instantaneous map of axial velocity in plane 4; RPM 1900.

Next we will look at the flow parameter distributions in plane 5 (

Figure 20).

Figure 20: Position of plane 5 is in the middle of the rotor.

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Distributions of the axial velocity component and the static pressure are outlined in Figure 21 and Figure 22. Figure 21: Instantaneous map of axial velocity in plane 5; RPM 1900.

Figure 22: Instantaneous distribution of static pressure in plane 5 (Pa); RPM 1900.

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The distribution of static pressure gives a possibility to estimate the total amplitude of unsteady pressure acting on the fan casing due to the passage of the blades. It gives a value around 1000 Pa. Figure 23: Plane and point P1 position used to register pressure pulsation data (marked by red color).

Pressure pulsations in the vortex (“non-acoustic”) flow at point P1 (shown in Figure 23) at the fan outlet are shown in Figure 24 and Figure 25. The pressure signals for the acoustic part on the inlet side of the computational domain are shown in Figure 26 and Figure 27. Figure 24: Spectrum of vortex pressure pulsations at P1, (Pa); RPM 1900.

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Figure 25: Spectrum of vortex pressure pulsations at P1, (Pa); RPM 950.

Near the zone of rotor-stator interaction the amplitude of pressure pulsations is comparable with the pressure blade loading and relatively high. Thus it is possible to use common CFD methods for evaluation of pressure pulsations with good accuracy. The main tonal component in the spectra is blade-passing frequency with the amplitude 3.7 Pa for the RPM 1900 mode and 0.8 Pa for the RPM 950. Thus the square-law for the dependence of pressure amplitude versus rotation speed is well confirmed. In the spectrum for RPM950 there is an essential low frequency peak but it does not influence the conclusion concerning the BPF amplitude. With additional computational time the low frequency trend in the pressure signal will be reduced. Far from the rotor the pressure pulsations BPF amplitude must be determined by solution of the acoustic-vortex equation (5.2). Below are presented the results of the solution for the 1st BPF harmonic for both modes of the fan operation in the inlet computational domain.

Figure 26: Acoustic pressure pulsations instantaneous map in meridian plane, (reduced by 4 ρ u 2p ); RPM 1900.

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Figure 27: Acoustic pressure pulsations instantaneous map in meridian plane, (reduced by 4 ρ u 2p ); RPM 950.

Finally, in Figure 28 and Figure 29 the acoustic pressure distributions of the 1st BPF amplitude on the boundary of the inlet domain are presented for the two RPM:s.

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Figure 28: 1st BPF amplitude at the inlet boundary, plane view, (Pa); RPM 1900.

Figure 29: 1st BPF amplitude at the inlet boundary, plane view, (Pa); RPM 950.

7 REFERENCES

[1] [2] 9

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