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HAL Id: jpa-00248844https://hal.archives-ouvertes.fr/jpa-00248844
Submitted on 1 Jan 1992
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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�
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).
1790 JOURNAL DE PHYSIQUE III N° 9
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.
N° 9 FLOW IN DIFFUSERS OF CENTRIFUGAL COMPRESSORS 1791
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.
1792 JOURNAL DE PHYSIQUE III N° 9
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.
N° 9 FLOW IN DIFFUSERS OF CENTRIFUGAL COMPRESSORS 1793
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
1794 JOURNAL DE PHYSIQUE III N° 9
---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
N° 9 FLOW IN DIFFUSERS OF CENTRIFUGAL COMPRESSORS 1795
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
1796 JOURNAL DE PHYSIQUE III N° 9
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
N° 9 FLOW IN DIFFUSERS OF CENTRIFUGAL COMPRESSORS 1797
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
1798 JOURNAL DE PHYSIQUE III N° 9
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
N° 9 FLOW IN DIFFUSERS OF CENTRIFUGAL COMPRESSORS 1799
~~~~~~
.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
1800 JOURNAL DE PHYSIQUE III N° 9
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
N° 9 FLOW IN DIFFUSERS OF CENTRIFUGAL COMPRESSORS 1801
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
1802 JOURNAL DE PHYSIQUE III N° 9
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
N° 9 FLOW IN DIFFUSERS OF CENTRIFUGAL COMPRESSORS 1803
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 :
1804 JOURNAL DE PHYSIQUE III N° 9
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.