detailed measurements of the flow field in vaneless and

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Detailed measurements of the flow field in vaneless andvaned diffusers of centrifugal compressors

Christian Fradin

To cite this version:Christian Fradin. Detailed measurements of the flow field in vaneless and vaned diffusers of centrifugalcompressors. Journal de Physique III, EDP Sciences, 1992, 2 (9), pp.1787-1804. �10.1051/jp3:1992213�.�jpa-00248844�

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J. Phys. iii France 2 (1992) 1787-1804 SEPTEMBER 1992, PAGE 1787

Classification

Physics Abstracts

47.90

Detailed measurements of the flow field in vaneless and vaned

diffusers of centrifugal compressors

Christian Fradin

ONERA, Direction de l'Energdtique, 29 Avenue de la Division Leclerc, 92320 Chitillon, France

(Received 28 February 1992, accepted lst June J992)

Rksumk. Le champ d'dcoulement du fluide dans les diffuseurs lisses et aubds de compresseurs

centrifuges transsoniques a £td analysd en utilisant des sondes miniatures. La premibre partie de

cet article montre une analyse exp£rimentale ddtaillde de la structure de l'dcoulement h la sortie

de deux rotors centrifuges transsoniques h aubes couch£es en ambre. L'un d'eux a des aubes

intercalaires. Ces deux rotors sont dquipds d'un diffuseur lisse k grand rapport de rayon. En

utilisant un repbre relatif lid au rotor, les mesures instationnaires faites prbs de celui-ci foumissent

le champ de vitesse du fluide h la sortie des canaux du mobile. Les effets des aubes intercalaires

sur la structure du fluide sont clairement montrds. Les mesures faites sur plusieurs rayons du

diffuseur lisse mettent en £vidence l'augmentation en amplitude de la distorsion axiale du fluide

tandis que l'hdtdrogdndit£ d'aube h aube diminue lentement. La seconde partie de l'article d£crit

le champ d'dcoulement dans un diffuseur h aubes dquipant le rotor dots d'aubes intercalaires. Les

mesures instationnaires effectudes dans la section du col montrent les fortes fluctuations

pdriodiques des angles et nombres de Mach instantands du fluide. Elles sont fonction de la

situation des aubes du rotor. La distortion du fluide s'accroit dans le canal du diffuseur. A la

sortie, une zone d'dcoulement secondaire appar#t prbs de l'intrados des aubes.

Abstract. The flow field development in vaneless and vaned diffuser of a transonic centrifugal

compressors has been investigated using miniature probes. The first part of this paper gives a

detailed experimental analysis of the flow structure at the outlet of two backswept transonic

centrifugal impellers, one with splitter blades. These two impellers are equipped with a largeradius ratio vaneless diffuser. Using a relative frame linked to the rotor, time-dependent

measurements close to the rotor provide the flow field velocity at the outlets of the impellerchannels. The effects of the splitter blade on the flow structure are clearly shown. Measurements

made at several radii of the vaneless diffuser show that the amplitude of the axial flow distortion

increases while the blade to blade heterogeneities decrease slowly. The second part of the paper

describes the flow field in a vaned diffuser facing an impeller with splitter blades. Time-dependent

measurements performed in the throat section show large periodic fluctuations of the instan-

taneous flow angles and Mach number, which depend on the impeller blade locations. The flow

distortion increases in the diffuser channel. At the outlet, a secondary flow area appears close to

the vane pressure side.

Introduction.

Small gas turbine engines with centrifugal compressor stages are coming into increasing use,

while improvements in the centrifugal compressors have been made primarily in advanced

1788 JOURNAL DE PHYSIQUE III N° 9

small gas turbine engines for aircraft, helicopters, as well as automotive and industrial

purposes. Any further improvement of the aerodynamic technology of highly loaded

turbomachinery requires an understanding of the complex flow phenomena occurring in the

impeller and in the vaned diffuser.

High-level axial and tangential heterogeneities appear in the flow leaving the impellerchannels. These ones are due to the three-dimensional effects (curvature of the shroud and of

the blades, vortex), the viscosity of the flow, the compressibility effects and the Coriolis

forces. There are also large steady tangential distortions generated by the action of the

diffuser vanes on the flow streamlines.

The job of the vaneless and vaned diffuser is to convert the kinetic energy of the flow into

maximum static pressure over a wide range of incident flow conditions.

In the vaneless diffuser, there are high tangential flow heterogeneities. These ones are due

to the frequency at which the blades pass, combined with the steady flow field created by the

diffuser vanes located downstream of the vaneless diffuser. Moreover, the great slowdown of

the flow leaving the impeller and the high adverse pressure gradient generate a three-

dimensional boundary layer on the walls of the vaneless diffuser.

In fact, vaneless and vaned diffuser flow fields are very complicated and are difficult to

compute at the present time because of the limits of modern computers and the insufficient

accuracy of losses, secondary flows and boundary layer models. Therefore, tests are necessary

to define correct boundary conditions and to validate the computations. In order to get such

results, tests have been undertaken on transonic centrifugal compressors.

The first part of this paper describes the flow field at the outlet of two centrifugal transonic

impellers. These impellers have backward leaning blades. The blades have similar geometries.One of the impellers has mid-channel splitter blades. During this investigation the impellers

are equipped with a large vaneless diffuser.

In the second part of this paper, the impeller with splitter blades is equipped with a vaned

diffuser.

To define the flow field structure in the compressor, measurements were made at several

locations. Tests where run at the nominal rotation speed of the impeller and at the mass flow

rate corresponding to the best efficiency of the compressor.

The impellers tested.

The two transonic centrifugal impellers investigated are shown in figure I. The rotor without

splitter blades is called Rl and the one with is called R2. Their geometries are given in table I.

The splitter blades of impeller R2 are located at mid-channel and their geometries are

identical to the corresponding part of the main blades.

Description of the compressor configurations.

1. CONFIGURATION WITH VANELESS DIFFUSER. A large parallel-wall vaneless diffuser of

1.5 radius ratio was placed at the impeller outlet. Measurements were taken at three different

radii (stations Sl, 52 and 53). Table II gives their radii ratios A to the outlet radius of each

rotor tested.

A vaned diffuser is located far from the impeller outlet. The flow disturbances due to these

vanes are estimated very small in the test sections.

2. CONFiGURATiON WITH VANED DiFFUSER. In this test configuration, the impeller R2

with splitter blades is equipped with a vaned diffuser that has curved vanes and a higher

pressure recovery capability (Fig. 2). Its main characteristics are given in table III.

N° 9 FLOW IN DIFFUSERS OF CENTRIFUGAL COMPRESSORS 1789

(

Fig. I. Impeller Rl and R2.

Table I. Main characteristics of the nvo impellers.

Impeller Rl R2

Impeller inlet

o hub diameter (mm) 100 100

o tip diameter (mm) 173 173

o number of blades 16 13

' angle of the blades from 64.71 62.54

axial direction (tip diameter)

o absolute Mach number 0.331 0.324

Impeller outlet

o diameter (mm) 287 294

o number of blades 16 26

o angle of the blades from 30 30

radial direction

o diffuser width (mm) I I I I

o peripheric Mach number 1.248 1,189

Rotating speed (rpm) 10 420 9 660

The vaned diffuser throat is of rectangular cross section. Measurements were taken at three

different test sections in the diffuser (Fig. 3).

I) The throat of the vaned diffuser (station 54).

2) The channel outlet of the vaned diffuser (station 55), in an orthogonal section to the

mean streamline and containing a diffuser vane trailing edge.3) The exit radius of the vaned diffuser (station 56).


Table II. Radius ratio A corresponding to the measurement locations in the vaneless

diffuser.

Impeller Sl 52 53

RI 1.049 1.088 1.127

R2 1.024 1.062 1,1

Fig. 2. View of the vaned diffuser.

A schematic representation of the facility is presented in figure 4. The hub in front of the

impeller gives a converging channel without inlet guide vanes that simulates the effects of

upstream axial flow compressors.

Test facility. A closed loop test facility with freon gas as a working fluid is employed for

the experiments. It includes a water-cooled heat exchanger, a calibrated nozzle for mass flow

measurements, and a throttle valve to regulate the back-pressure. The compressor is driven

by an electric motor. Uniform total pressure and swirl-free flow conditions are provided at the

compressor inlet by a set of antiturbulence screens and honeycomb meshes.

The thermodynamics of freon l14 are temperature and pressure-dependent and these

variations of the thermodynamic parameters have been taken into account.

On the other hand, the low speed of sound in freon 114 makes it more convenient than air

for low speed rotors with a comparable Reynolds number range.

Experimental measurement techniques. Considering the narrowness of the flow passage,

the probes used are as small as possible.


Table III. Vaned diffuser design parameters.

vane number 19

diffuser depth (mm) 11

throat aspect ratio 1.527

exit diffuser channel aspect ratio 3.394

ratio of diffuser leading edge radius to impeller 1.034

exit radius

ratio of diffuser trailing edge radius to impeller 1.527

exit radius

vane leading edge angle 15.3°

vane trailing edge angle 36.7°

diffuser channel divergence angle 2a =

8°

(channel outlet] T~S' S~'~'~~ ~~

~~~~ ~~~~~ ~~

(throat]Measurement points

Semi vanelessdfiluser

Suction sdej

Pressure sidedxtuser

~~,ess~~

Rotor

~

01

Splfiter blade

LongTest section 56

blade(vaned diffuser ouileJ)

Fig. 3. Location of measurement stations in the vaned diffuser.

a) Each test section is instrumented with several static pressure taps and miniaturized

piezoelectric pressure transducers to give the time-averaged and time-dependent static

pressures. Transducers are mounted flush with the walls. They have a large passband(100 kHz) and are calibrated in a shock tube.

b) A calibrated«

cobra»

probe measures the time-averaged total pressure and flow angle

versus the channel depth. The local time-averaged Mach number is calculated from a linear

interpolation of the steady static pressures in the axial direction.


Inlet flow

iHub

i Vaneless d#luser

Vaned d#fuser

lmoeller

'

' Electricdrivingmechanism coiiectingchamber

Fig. 4. Cross-section of compressor test stage.

c) Time-dependent measurements of the flow Mach numbers and flow angles are carried

out with a single hot wire anemometer, 1.5 mm long and 5 ~Lm in diameter.

There are two drawbacks to using hot wire anemometers in transonic air compressors :

the passband range is too small,

the wires themselves are not strong enough to resist the aerodynamic forces.

Furthermore, most transonic compressors operate in open circuits and the dust deposit on

the wires quickly deteriorates the probe.Using freon as a working fluid in a closed loop compressor facility remedies these problems

since transonic flows are obtained at low flow velocities and low temperature rises, with low

speed of impeller rotation.

A hot-wire anemometer inserted in the wind tunnel freon loop was calibrated to calculate

the flow Mach number versus the output voltage and the Reynolds number iii.

The same hot wire is also used to measure the instantaneous flow angles. For this purpose,

the wire is placed at a sequence of different angles to the tangential direction of the impeller.It is interesting to note that measurements of instantaneous flow angles are independent of

local Mach and Reynolds numbers. However, the wire being parallel to the diffuser walls,

velocity component orthogonal to the walls could not be measured.

For the time-dependent measurements, a photoelectric pickup detects the passage of each

long blade of the impeller and triggers the data acquisitions.Since the signal contained some noise, the unsteady data were phase averaged together.The probes are fited on a probe carrier attachment in order to make measurements at

several cross wise stations between the two diffuser walls.

Transverse measurements with pressure probe. Absolute flow Mach numbers and anglesmeasured in the vaneless diffuser are plotted in figure 5. Outside the boundary layers, test

results show that the maximum Mach number is located near the shroud side. The flow angledistribution is less disturbed at the outlet of rotor R2 than at the outlet of rotor RI.


1

, , ',

0 025 050 0.75 0 025 050 075

dab nab

'

'

,

~ '~,~ i <

Ih h

" p H0 025 0.50 0.75 0 0.25 0.50 075

lmpellerRl lmpellerR2

Sl

52

~3

Fig. 5. Time-averaged of the absolute flow Mach numbers M~~ and angles a~~ measured in the

vaneless diffuser.

As predicted by the tangential momentum law, the slow down of the flow is clearlyobserved in the vaneless diffuser. These results emphasize the increase in amplitude of the

axial disturbances as the test sections move further downstream. This is namely the case for

the radial Mach number whose value decreases rapidly in the vicinity of the hub side.

A potential flow theory shows that such an increase in axial distortion amplitude can be

predicted when the flow is not uniform at the vaneless diffuser inlet. Moreover the three-

dimensional boundary layers thicken on both diffuser walls. For these reasons, a recirculation

zone may appear close to the walls [2].

Unsteady flow distribution in the vaneless diflkser. Instantaneous absolute flow Mach

numbers and angles are measured at several depths from the diffuser walls.

The absolute flow angle contour maps and the three-dimensional representations of the

total pressure corresponding to the pitch of one long blade are plotted in figure 6 for the two

impellers studied. The absolute angles indicated on the maps are measured between the local

flow velocity vector and the local tangential direction. At the impeller outlet (Sect. Sl), the

circumferential distortion of the main flow and the wake of the blades can be clearlydistinguished. The highest Mach numbers and smaller flow angles are measured in the vicinity


---52

~~"'~- Cl',

~"«=C--_S

Xi

3.5

3

2.5~

Shroud

,

-'~'-

" ", si ',

~,---

'"'J

,>

~+

~° /~~ n ~L-B- L-B- S-B- -

L-B- L.B.impeller Hi

Impeller R2

5 degreeslo degreeslsdegrees

20 degrees

Fig. 6. Distribution of the absolute flow angles and total pressures in the vaneless diffuser (L.B, longblade S-B- splitter blade).

of the blade suction sides [3]. The flow angle fluctuations around impeller Rl is 12.5° at mid

channel.

This variation is slightly smaller for impeller R2.

Conceming this impeller, the flow field structure is different in each of the channels

separated by the splitter blade : the main difference is in the flow angle distribution. Tests in

sections 52 and 53 determined the flow distortions within the vaneless diffuser. Due to the

shear forces between the streamlines, mainly in the wake area, and the reversible work

transfer, the blade-to-blade flow heterogeneities decay slowly, so that the blade wakes do not

disappear in the investigated part of the vaneless diffuser [4, 5]. The time-lag of the flow in the


opposite direction of rotation is clearly shown. It is due to momentum conservation law in the

circumferential direction.

Performance of the vaneless diffuser. To determine the performance of the vaneless

diffuser and understand the behaviour of the flow field heterogeneities better, it would be

desirable to represent the real flow correctly by uniform steady flow models. Many different

averaging procedures have been proposed in the literature. However, these mean values are

not consistent with all conservation laws. In fact, what averaging method is used dependsessentially on the intended application.

In order to estimate the heterogeneity level of the flow field at the impeller outlet, and its

variation in the vaneless diffuser, we can introduce a channel blockage factor B. Using a

conventional jet-wake representation, we can assume that the flow vanishes in a fraction B of

the test section and that the homogeneous equivalent flow discharges in the remaining(I B ) fraction.

Using the conservation laws for mass, momentum, energy and entropy, the balance

equations written on one test section having a surface A are :

mass flow :

p§~(l -B)A=

lpv~dAA

radial momentum :

lpU)(i B) + pi A=

I(pV) + P) dA

A

tangential momentum :

pU~@~(l -B)A= lpv~v~dA

A

total enthalpy H~ :

p@~H,(I -B)A= lpv~H~dA

A

entropy S :

p@~S(I -B)A=

lpv~sdAA

where v~ and v~ are the radial and circumferential flow velocities, p the density and p the

pressure.

The set of equations is closed with the thermodynamic gas laws, It is used to calculate the

mean total pressuref, and total temperature f;.

Figure 7 displays the measured variations of the mean flow characteristics versus the radius

ratio A of the vaneless diffuser. The values at the rotor outlet (A=

I) are obtained byextrapolation.

Despite the decay of the heterogeneities in the blade-to-blade direction, the blockage factor

B increases along the vaneless diffuser. This is due to the greater development of the axial

heterogeneities as compared to the decay of blade-to-blade heterogeneities. Results show the


if 03

'mpellerRl

fl

~f

U

I '' '' '

'

/A~

~'_A

,

,''

0.2

~impeller R2

3.4_

' '-- _,

impeller Rl-,, +

- --_ _

_

CZ

~~j

3.2 jn~j~m~n~~~~+-

pSieady measurements

~#E

~'lmpellerR2

0.8 Steady measurements

measurements

I 1.05 1, i

Radius ratio ~

Fig. 7. Evolutions of the flow parameters in the vaneless diffuser.

flow issuing from the impeller R2 with splitter blades is more homogeneous than the flow

issuing from the impeller Rl without the splitters. The total pressure ratio #t, and work

coefficient R' obtained by instantaneous measurements can be compared to the values given

by the pressure probe (steady measurements). The work coefficient R' is the ratio between

the local tangential flow velocity and the peripheral velocity at the impeller tip.Differences are due to the distribution of flow Mach numbers and angles in the blade-to-

blade direction. So, it is postulated that if instruments for steady flow measurements are used

close to the impeller outlet, the total pressure and the work coefficient may be overestimated.

The decay of the work coefficient versus the radius ratio is due to the tangential

momentum losses.

These results confirm the importance of carrying out time-dependent measurements very

close to the impeller if its performances is to be determined with accuracy.

With the averaging method used, the calculated mean static pressure f differs from the

pressure# actually measured. This is a drawback if one wants to determine the performance


of the vaneless diffuser. To overcome this problem, we use the mean static pressure

# measured on the two walls of each test section. The performance of the vaneless diffuser at

a radius ratio A can be characterized by the efficiency

~A ~(A =1)'i IA )

(P; Pj~

i)) (Pi PA )

The efficiency is constant along the vaneless diffuser. We found J~ =

0.52 for the impeller Rl

andJ~ =

0.60 for the impeller R2. Therefore, it can be stated that the flow discharged from

rotor R2 performs better than that of rotor Rl, whose outlet flow is less homogeneous.

Relative flow pattern at the outlet of the impellers. It is assumed that the nearest

measurement taken in the vaneless diffuser (station Sl) can be extrapolated to the outlet

radius of the impeller.The components of the relative velocity vector have been calculated using the impeller

outlet circumferential velocity, conservation of the local mass flow rate, and tangential

momentum law. The time-lag of the fluid versus the impeller blades and the tangential

momentum loss shown in figure 7 have also been taken into account.

Furthermore, the viscous effects are particularly large in the vicinity of both walls and of

both sides of the blades. In fact, the relative flow pattem in these channel parts can not be

calculated with accuracy.

The results are presented in figure 8. In each case, the highest relative Mach number

(M=

0.5 ) is located close to the blade pressure side. Relative Mach number fluctuations can

reach AM=

0.3 in the blade-to-blade direction. However, the relative Mach number is

almost constant in the axial direction. The relative flow angle fluctuates in the range of 30 to

60°. Large axial angle distortions are also noted. It is postulated that the absolute axial flow

heterogeneities detected using pressure probes are due to the effect of the relative angledistortions in the axial direction.

Figure 8 shows that the flow pattem at the outlet of the two passages of impeller R2 are

notably different. The flow structure emanating from the impeller Rl looks like the one from

the channel of the impeller R2 whose pressure side is a long blade.

Local efficiency and work parameter. The isentropic efficiency and work parameter of

the streamlines at the impeller outlet are computed from the known flow field pattem in the

relative frame of reference.

The distribution of these parameters is highly non-uniform as shown in figure 9.

The work coefficient R' is maximum in the area corresponding to a small relative flow

velocity.The distribution of efficiency corresponds to the usual schematic representation for the jet

flow and secondary flow. Maximum efficiency is found in the channel area close to one blade

pressure side and the hub. The flow efficiency close to the shroud is affected by the clearance

between the casing and the impeller.The distribution of the work coefficient and efficiency is quite different at the outlet of the

two passages of the impeller R2.

Measurements in the vaned dfluser throat. The flow in a vaned diffuser is characterized

by fluctuations due to the flow heterogeneities at the impeller outlet. Thus, the circumferential

locations of the impeller's long blades versus the diffuser vanes can be characterized by a

dimensionless period t/T. The period T is the time separating the passage of two long blades

of the impeller in front of one diffuser vane leading edge. The initial location of the impeller'slong blades is drawn in figure 3. Only 8 out of the 240 analysed rotor locations have been


Relative Mach

number

.5

.4

.3

.2

Shrcud

-

Hub~

i- fi°~

Impeller RlL-B-

30 degrees40degreess0degrees

60 degrees

Shroud

-~ _= ~i

/ ~~'~ >j

j /~

/~~

/ /~

-~

Hub

fi ~_ fi

L-B-~'~' °~

L_~_impeller R2

Fig. 8. Relative flow angles and Mach numbers at the impeller outlets.

selected to illustrate the complex flow pattem within the throat. These pattems are shown in

figure 10. The contour maps plotted on the left-hand side show the instantaneous flow angles.The fight-hand side represents the instantaneous three-dimensional flow Mach number

distributions.

The drawings show the impelldr blade positions in sequence and indicate the correspondingdimensionless period t/T. The vaned diffuser is exposed to the highly distorted flow coming

from the impeller. In addition, the wake leaving the impeller at different dimensionless

periods t/T does not arrive at the same time in the throat section. This is due to the flow


~~~~~~

.88.90' ~

~

Rl

~,92~ ~.94-

L-B, Hub L-B-

.90

~

~#_.__

~ ~ll~/~~~L-B, S-B- L,B,

al Local efficiency

Shmud

Rl

~~.84

0 /

-..

.~ )~L-B- Hub 68 L-B-

~~~ '~~~ ~~~~~~

L,B-~~ ~~ S-B- ~~

L,B-

b) Work coefficient

Fig. 9. Local efficiency and work coefficient of the flow at the impeller outlets.

heterogeneities at the impeller outlet, to the different lengths of the flow streamlines, and to

the various local adverse pressure gradients in the semi-vaneless space.These effects create large fluctuations in the flow Mach number in the vaned diffuser throat

as seen in figure10 [6]. For example, the Mach number flow pattem is almost flat at the

dimensionless periods t/T=

0.375 and t/T=

0.875. Then the flow pattem suddenly becomes

highly distorted, which is due to the passage of one blade (t/T=

0 and t/T=

0.5). At this

moment, the local flow Mach number can reach unity. Meanwhile, in the vicinity of the

diffuser vane suction side, the local flow Mach number corresponding to the blade wakes is

less than 0.6. However, the blade wake effects are not clearly observed in the vicinity of the

vane pressure side. It is assumed that the high positive pressure gradient in this part of the

semi-vaneless space contributes to the rapid mixing of the flow. The flow angles indicated on


f~ l

j"", '~-'~~~i

fitfl~0

fi ~"',,,," j

'~"~+tfl.0,125

ill

I

§'~" ~'ijj

'~

~~'~-j",-' ,/

it'<--~'~tfl=0,25

~-

f~~,--, ii

,'I' ~~t ~', j')

,I it l~~~~,I'--"' -Zi~_ ~~~~

~~

__

,-

~

ill1.

'', ~,'

,l.ill"f ''

-

tfl=0.625

~i-

L_ ','

ill""? , j

, I"

' fl.0.75

iq ~~~ ~

~~

~jw~i~

~~~

~ngles Mach bers

4deg0deg,4deg10deg

Fig. 10. Fluctuations of the flow structure in the diffuser throat versus the impeller blades.

the contour lines of the maps are made of the local flow velocity vector and the perpendicular

to the throat line as shown in figure 3.

Considering the 2-D diffuser vane geometry, the constant angle lines should be vertical

lines. Experimental results are in bad agreement with this ideal flow pattem. In fact, the map

of the instantaneous flow angles never corresponds to the geometrical angle distribution of

the walls of the diffuser vanes. Outside the boundary layers, the flow angles are very often

negative. Only a small part of the flow has positive angles. For this reason, the streamlines


Front wall (Shroud)

i3I

~

a j

I .l -2.~

~~~

fi ) /

~i

m ~ "

~

Back wall

h

,

37.834.5

37.8~

30 6cn

~~

34.5

.8'~dgeofthediffuservane34.5 ,,

, j~,

30.6.

fl~',

Fig.

tend to go towards the diffuser vane pressure side, In the middle of the throat, the flow anglefluctuates between 4° (t/T

=

0.375) and 5° (t/T=

0.125). Large flow angle fluctuations are

also seen close to the diffuser vane walls. It seems that the fluid does not adhere to the vane

walls. Unsteady separated zones probably occur in the diffuser throat. The very small angles(less than 15°) on both the flont and back walls are due to the three-dimensional boundarylayers. Occasionally, separations can occur with a back flow. These results show that

increasing the vaneless space radius ratio tends to increase the boundary layer thickness and

induce three-dimensional flow distortions in the diffuser throat. Therefore, it is postulatedthat a large radius ratio of the vaneless diffuser entails an excess three-dimensional boundarylayer and consequently decreases the pressure recovery of the vaned diffuser.

Measurements at the di#fkser outlet (stations 55 and 56). The contour maps of the time-

averaged flow Mach numbers and flow angles measured in the diffuser channel outlet (station55) are shown in figure I I. The maximum flow Mach number is located near the front wall

side, as in both the impeller outlet and the diffuser throat. Along the front wall, the large flow

angle is evidence of a three-dimensional boundary layer.The sharp drop in the flow Mach number in the vicinity of the vane pressure side is

remarkable. Local flow angles become smaller than the angle of the vane wall. In this area,

secondary flows and separations are expected. These are certainly due to the flow

heterogeneities and three-dimensional boundary layer observed in the throat section.

Measurements made at three locations in the vaned diffuser outlet (station 56) confirm the

large and increasing heterogeneities of the flow structure. A very low flow Mach number is

measured at the nearest location on the pressure side (less than 0.02). Tile enlargement of the

secondary flow area is due to the greater increase of the static pressure in the pressure side


streamlines than in those on the suction side. In fact, the time-averaged flow heterogeneities

are higher than in the throat section. However, time-dependent flow heterogeneities

decrease. For example, the flow angle fluctuations are only 2° at the center of station 55. The

probe used can not measure the axial velocity components of the streamlines. However, they

are certainly important in these test sections.

The flow field structure at the vaned diffuser outlet can be described schematically as a jet-

wake configuration with accumulation of losses in the vicinity of the vane pressure side. This

impairs the vaned diffuser efficiency and does not give an ideal inlet flow to the next

compressor stage whose inlet is far downstream.

Performance of the vaned diffuser. The mean total pressure ratio fi~ in the tested sections

54 and 55 is calculated from the conservation law for entropy. It is assumed that the process in

the diffuser is adiabatic. Consequently the mean flow temperature f~ is equal to the

temperature at the compressor outlet. The mean static pressure ratio if is calculated from

static pressure taps located on both walls of the tested section. The mean flow angle

& with regard to the tested section is found from momentum law on the plane parallel to the

section :

fl sin &=

M sin a dQQ

A

and the blockage factor B is determined according to the relationship :

B=

~

Apdm cos &

Table IV shows the main results. At the impeller outlet (50), the mean total pressure ratio

and mean flow Mach number were calculated from the local static pressure, the mass flow rate

and Euler's theorem. It was assumed that the blockage factor can be computed from results

obtained with large vaneless diffuser. Conceming the diffuser outlet (test Sect. 56), it is

assumed there are no losses between the test section 55 (diffuser channel outlet) and 56.

Results emphasize the decrease of the mean total pressure ratio fi~ and the slowdown of the

flow throughout the vaned diffuser. The blockage factor B increases continuously in the

vaned diffuser corresponding to an increase of the flow field heterogeneity.Results emphasize the continuous growth of the flow field heterogeneities throughout the

diffuser, and consequently the decay of the mean flow total pressure ratio.

Table IV. Mean characteristic of the flow in each section tested.

Test section w~i B

Impeller outlet (50) 3,164 0.975 0.214

Vaned diffuser throat (54) 2.985 0.622 0.047

Diffuser channel outlet (55) 2.933 0.302 0.231

Vaned diffuser outlet (56) 2.933 0.295 0.285


The performance of the vaned diffuser between two sections I and 2 can be

characterized by :

the total pressure loss coefficient K

if~ if~~

K=

"1 "1

the static pressure recovery C~

~if~ -if,

~~lf~i

-ifi

the efficiencya~

~2 ~~l

~(fi~

jit (7T~ ~ 7T ~

Results calculated for different parts of the diffuser are shown in table V. Due to the highlydistorted flow delivered by the impeller and the strong adverse pressure gradient, losses are

greater in the semi-vaneless diffuser than in the diffuser channel [7].

Table V. Performance of each part of the vaned diffuser.

Inlet section Outlet K C~ a~

section

Impeller outlet Vaned 0, 143 0. 410 0. 741

(50) diffuser

throat (54)

Impeller outlet Channel 0, 185 0.705 0.792

(50) diffuser

outlet (55)

Impeller outlet Diffuser 0. 185 0.710 0.793

(50) outlet (56)

Vaned diffuser Channel 0.093 0.660 0.876

throat (54) diffuser

outlet (55)

Conclusion.

The first part of this paper presents a detailed experimental analysis of the flow field at the

outlet of two transonic centrifugal compressors. Both impellers have backward-leaningblades, but one is equipped with splitter blades. In the second part, results of measurements

made in a vaned diffuser are described. The folkwing conclusions have been drawn :


1) Blade-to-blade flow heterogeneities along the vaneless diffuser decay slowly ; but the

transverse non uniformities become even larger as one proceeds downstream, and flow

separation can even be encountered.

2) Time-dependent measurements reveal that the flow structure is highly unsteady in the

diffuser throat. The blade-to-blade heterogeneities delivered by the impeller do not decrease

rapidly in the semi-vaneless diffuser.

3) Measurements performed at the vaned diffuser outlet show that the flow distortions

grow in the divergent channel. Flow separation on the vane pressure side is expected.

References

Ii MORKOVIN M. V.,«

Fluctuations and hot-wire anemometry in compressible flows », Agardograph(24 November 1956).

[2] BAMMERT K., RAUTENBERG M., WITTERKINDT W.,«

Vaneless diffuser flow with extremelydistorted inlet profile », ASME paper n 78 GT 47 (1978).

[3] ECKARDT D.,«

Instantaneous measurements in the jet-wake discharge flow of a centrifugal

compressor impeller », ASME paper n 74 GT 90 (1974).[4] SENOO Y., ISHIDA M.,

«Behavior of severely asymmetric flow in a vaneless diffuser », ASME paper

n 74 GT 64 (1974).[5] INOUE M.,

«Radial vaneless diffusers : a re-examination of the theories of Dean and Senoo and of

Johnston and Dean », ASME paper n 78 GT 186 (1978).[6] KRAIN H.,

«A study on centrifugal impeller and diffuser flow », ASME paper n 81 GT 9 (198 Ii-

[7] STEIN W., RAUTENBERG M., «Analysis of measurements in vaned diffusers of centrifugal

compressors », ASME paper n 87 GT 170.

detailed measurements of the flow field in vaneless and

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